EP0455761B1 - Hydraulische ventilsteuervorrichtung für brennkraftmaschinen - Google Patents

Hydraulische ventilsteuervorrichtung für brennkraftmaschinen 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
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Lifetime
Application number
EP90915296A
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German (de)
English (en)
French (fr)
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EP0455761A1 (de
Inventor
Helmut Rembold
Ernst Linder
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Robert Bosch GmbH
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Robert Bosch GmbH
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Publication date
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Publication of EP0455761A1 publication Critical patent/EP0455761A1/de
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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.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Valve Device For Special Equipments (AREA)
EP90915296A 1989-11-25 1990-10-26 Hydraulische ventilsteuervorrichtung für brennkraftmaschinen Expired - Lifetime EP0455761B1 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE3939065 1989-11-25
DE3939065A DE3939065A1 (de) 1989-11-25 1989-11-25 Hydraulische ventilsteuervorrichtung fuer brennkraftmaschinen

Publications (2)

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

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EP90915296A Expired - Lifetime EP0455761B1 (de) 1989-11-25 1990-10-26 Hydraulische ventilsteuervorrichtung für brennkraftmaschinen

Country Status (5)

Country Link
US (1) US5263441A (ja)
EP (1) EP0455761B1 (ja)
JP (1) JPH04502660A (ja)
DE (2) DE3939065A1 (ja)
WO (1) WO1991008385A1 (ja)

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DE19949514A1 (de) * 1999-10-14 2001-04-19 Bosch Gmbh Robert Vorrichtung zum schnellen Druckaufbau in einer durch eine Förderpumpe mit einem Druckmedium versorgten Einrichtung eines Kraftfahrzeugs
WO2003018965A1 (de) 2001-08-21 2003-03-06 Robert Bosch Gmbh Ventilmechanismus mit einem variablen ventilöffnungsquerschnitt
WO2003018967A1 (de) 2001-08-21 2003-03-06 Robert Bosch Gmbh Ventilmechanismus mit einem variablen ventilöffnungsquerschnitt

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DE4206696C2 (de) * 1992-03-04 2000-12-14 Bosch Gmbh Robert Hydraulische Ventilsteuervorrichtung für Motorventile
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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 (it) * 1998-01-23 1999-07-23 Fiat Ricerche Perfezionamenti ai motori a combustione intenra con valvole ad azionam ento variabile.
GB9906504D0 (en) * 1999-03-23 1999-05-12 Csa Performance Ltd Valve actuation means
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ITTO20010270A1 (it) * 2001-03-23 2002-09-23 Fiat Ricerche Motore a combustione interna con sistema idraulico di azionamento variabile delle valvole e punteria a doppio stantuffo.
ITTO20010269A1 (it) * 2001-03-23 2002-09-23 Fiat Ricerche Motore a combustione interna, con sistema idraulico di azionamento variabile delle valvole, e mezzi di compensazione delle variazioni di vol
DE10140952A1 (de) 2001-08-21 2003-03-20 Bosch Gmbh Robert Ventilmechanismus mit einem variablen Ventilöffnungsquerschnitt
LU90889B1 (en) * 2002-02-04 2003-08-05 Delphi Tech Inc Hydraulicv control system for a gas exchange valve of an internal combustion engine
DE10231143B4 (de) * 2002-07-10 2004-08-12 Siemens Ag Verfahren zum Steuern des Ventilhubes von diskret verstellbaren Einlassventilen einer Mehrzylinder-Brennkraftmaschine
JP2004197588A (ja) * 2002-12-17 2004-07-15 Mitsubishi Motors Corp 内燃機関の動弁装置
FI117348B (fi) * 2004-02-24 2006-09-15 Taimo Tapio Stenman Hydraulinen laitejärjestely polttomoottorin venttiilien toiminnan ohjaamiseksi
DE102010018209A1 (de) * 2010-04-26 2011-10-27 Schaeffler Technologies Gmbh & Co. Kg Hydraulikeinheit für einen Zylinderkopf einer Brennkraftmaschine mit hydraulisch variablem Gaswechselventiltrieb
KR20120017982A (ko) * 2010-08-20 2012-02-29 현대자동차주식회사 전기-유압 가변 밸브 리프트 장치
JP6003439B2 (ja) * 2012-09-18 2016-10-05 アイシン精機株式会社 弁開閉時期制御装置
SE540359C2 (sv) * 2013-10-16 2018-08-07 Freevalve Ab Förbränningsmotor

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DE3135650A1 (de) * 1981-09-09 1983-03-17 Robert Bosch Gmbh, 7000 Stuttgart "ventilsteuerung fuer hubkolben-brennkraftmaschinen mit mechanisch-hydraulischen bewegungsuebertragungsmitteln"
JPH0612058B2 (ja) * 1984-12-27 1994-02-16 トヨタ自動車株式会社 可変バルブタイミング・リフト装置
DE3511820A1 (de) * 1985-03-30 1986-10-02 Robert Bosch Gmbh, 7000 Stuttgart Ventilsteuervorrichtung fuer eine hubkolben-brennkraftmaschine
DE3511819A1 (de) * 1985-03-30 1986-10-09 Robert Bosch Gmbh, 7000 Stuttgart Ventilsteuervorrichtung
DE3532549A1 (de) * 1985-09-12 1987-03-19 Bosch Gmbh Robert Ventilsteuervorrichtung
DE3815668A1 (de) * 1988-05-07 1989-11-16 Bosch Gmbh Robert Ventilsteuervorrichtung mit magnetventil fuer brennkraftmaschinen
US4930465A (en) * 1989-10-03 1990-06-05 Siemens-Bendix Automotive Electronics L.P. Solenoid control of engine valves with accumulator pressure recovery

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE19949514A1 (de) * 1999-10-14 2001-04-19 Bosch Gmbh Robert Vorrichtung zum schnellen Druckaufbau in einer durch eine Förderpumpe mit einem Druckmedium versorgten Einrichtung eines Kraftfahrzeugs
DE19949514C2 (de) * 1999-10-14 2001-10-18 Bosch Gmbh Robert Vorrichtung zum schnellen Druckaufbau in einer durch eine Förderpumpe mit einem Druckmedium versorgten Einrichtung eines Kraftfahrzeugs
WO2003018965A1 (de) 2001-08-21 2003-03-06 Robert Bosch Gmbh Ventilmechanismus mit einem variablen ventilöffnungsquerschnitt
WO2003018967A1 (de) 2001-08-21 2003-03-06 Robert Bosch Gmbh Ventilmechanismus mit einem variablen ventilöffnungsquerschnitt

Also Published As

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

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