EP2386731B1 - Soupape hydraulique - Google Patents

Soupape hydraulique Download PDF

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Publication number
EP2386731B1
EP2386731B1 EP11159294.5A EP11159294A EP2386731B1 EP 2386731 B1 EP2386731 B1 EP 2386731B1 EP 11159294 A EP11159294 A EP 11159294A EP 2386731 B1 EP2386731 B1 EP 2386731B1
Authority
EP
European Patent Office
Prior art keywords
sleeve
hollow piston
pivoting
camshaft adjuster
pressure
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.)
Not-in-force
Application number
EP11159294.5A
Other languages
German (de)
English (en)
Other versions
EP2386731A1 (fr
Inventor
Marc Hohmann
Dietmar Schulze
André Selke
Patrick Gautier
Markus Todt
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.)
Hilite Germany GmbH
Original Assignee
Hilite Germany GmbH
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Hilite Germany GmbH filed Critical Hilite Germany GmbH
Publication of EP2386731A1 publication Critical patent/EP2386731A1/fr
Application granted granted Critical
Publication of EP2386731B1 publication Critical patent/EP2386731B1/fr
Not-in-force legal-status Critical Current
Anticipated expiration legal-status Critical

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Classifications

    • 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
    • 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
    • F01L2001/34426—Oil control valves
    • 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
    • F01L2001/34426—Oil control valves
    • F01L2001/34433—Location oil control valves
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
    • Y10T—TECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
    • Y10T137/00—Fluid handling
    • Y10T137/8593—Systems
    • Y10T137/86493—Multi-way valve unit
    • Y10T137/86574—Supply and exhaust
    • Y10T137/8667—Reciprocating valve
    • Y10T137/86694—Piston valve
    • Y10T137/86702—With internal flow passage

Definitions

  • the invention relates according to the patent claim 1 a Schwenkmotorversteller with a hydraulic valve.
  • the hydraulic valve has a bushing and a hollow piston axially displaceable within it by means of an actuator against the force of a helical compression spring and having a piston crown. At this piston bottom, the actuator is applied.
  • a bushing-fixed sleeve is arranged relative to this displaceable.
  • the wall of the sleeve has a breakthrough. The breakthrough leads to an opening of the hollow piston. This opening can lead to a supply pressure P applied within the sleeve, alternatively to two pressure chambers of the Schwenkmotornockenwellenverstellers.
  • the sleeve has a sleeve bottom closing the interior of the hollow piston. The sleeve is connected at its end remote from the actuator fixed to the socket.
  • a hydraulic valve for a Schwenkmotomockenwellenversteller has a bushing and a hollow piston axially displaceable within it by means of an actuator against the force of a helical compression spring. Within the hollow piston, a sleeve is provided.
  • a supply pressure P can be performed alternatively to two load ports A, B or two pressure chambers of the Schwenkmotomockenwellenverstellers.
  • There are two tank connections T1, T2 are provided.
  • the order of the radial ports is P-T1-BA. This is followed by the second tank connection T2 as an axial end-side connection.
  • the object of the invention is to provide a low-cost and small Schwenkmotornockenwellenversteller with a high control quality.
  • a sleeve is arranged relatively displaceable within the hollow piston of the hydraulic valve.
  • this sleeve can maintain its position relative to a socket within which the hollow piston is displaceable.
  • a limited axial clearance and radial play be provided which prevents jamming of the mutually movable parts or compensates for tolerances.
  • the sleeve has a sleeve bottom closing the interior of the hollow piston.
  • This sleeve base is thus firmly supported relative to the bush, so that the resulting from the pressure from the supply port P forces are supported on the sleeve bottom and the sleeve on the socket.
  • these forces do not act on the piston head of the hollow piston, which serves as the support for an actuator.
  • the hollow piston is thus free of axial forces from the supply pressure, the axial position of the hollow piston can be regulated by the actuator without the supply pressure having to be taken into account.
  • the supply pressure can vary depending on the nature of its provision. Since an oil pump mechanically driven by the internal combustion engine is usually used, the supply pressure varies depending on the engine speed and temperature or viscosity of the oil. Other consumers can also play a role.
  • the achievable according to the invention particularly high control quality offers a particular advantage when combined with a hydraulic structure that uses the camshaft alternating moments to support the angular adjustment by means of the Schwenkmotornockenwellenverstellers.
  • This use makes higher demands on the control of the hydraulic valve, since these camshaft alternating torques are non-uniform and quickly changing effect.
  • Such a function for the use of the camshaft alternating torques is from the DE 10 2006 012 733 B4 and the DE 10 2006 012 775 B4 already known.
  • the hydraulic valve according to the invention can accordingly be designed such that it makes it possible in a particularly advantageous manner to use pressure fluctuations in the first working port B associated pressure chambers of Schwenkmotomockenwellenverstellers to supply the opposite direction of rotation associated pressure chambers with sufficient fluidic volume flow.
  • These pressure variations result from the camshaft alternating torques that occur at the camshaft in response to the gas exchange valve forces.
  • the smaller the number of combustion chambers per camshaft the stronger the camshaft alternating torques, so that the advantages of using the camshaft alternating torques come to particular extent in internal combustion engines with few - for example, three - cylinders to bear.
  • influencing parameters are still the strength of the gas exchange valve springs and the camshaft speed.
  • phase adjustment of the camshaft can thus be done quickly. Moreover, due to the use of camshaft alternating torques, it is advantageously possible to enable adjustment with a relatively low oil pressure. A thus possible small dimensions of the oil pump improves the efficiency of the internal combustion engine. The saved volume flow of hydraulic fluid is available to other consumers, such as a hydraulic valve lift.
  • camshaft alternating torques can be done for both directions of rotation or only for one direction of rotation.
  • a coil spring according to DE 10 2006 036 052 A1 can be used, which then compensates for the additional adjustment forces in one direction of rotation.
  • check valve which can be designed in particular band-shaped.
  • the hydraulic valve may be embodied in a particularly preferred embodiment as a central valve, wherein the supply pressure is supplied via the camshaft.
  • a central valve has space advantages.
  • the opposite of a central valve are the external hydraulic valves for operating the Schwenkmotomockenwellenverstellers.
  • the hydraulic passages for the camshaft adjustment run from the pivoting camshaft adjuster to a separate timing drive cover with the hydraulic valve screwed in there or else to the cylinder head with the hydraulic valve screwed in there.
  • the likewise hydraulic central valve is arranged radially inside the rotor hub of the Schwenkmotomockenwellenverstellers.
  • the central valve comes in the aforementioned DE 10 2006 012 733 B4 and the DE 10 2006 012 775 B4 listed method for faster adjustment of the Schwenkmotornockenwellenverstellers especially for carrying, since the hydraulic fluid from the one direction of rotation associated chambers in the other direction of rotation associated chambers has a short path.
  • the line losses would consume the advantage with increasing line length.
  • the control engineering challenges go along, which give the pressure-balanced hollow piston according to the invention a particular advantage.
  • the sleeve of a central valve can be designed in a particularly advantageous manner with a thread for screwing the rotor to the camshaft, so that forms a so-called central screw.
  • the supply pressure does not have to be introduced axially into the socket on the front side. It is also possible to provide the supply connection radially, so that the supply pressure also takes place radially into the hydraulic valve. However, the supply of the supply pressure does not have to be done in the sleeve end face. It is also possible to introduce the pressure via a transverse bore in the socket, which then leads into the interior of the sleeve. In this case, the supply can be made in the sleeve in the frontal opening or in an opening in the said wall of the sleeve.
  • the sleeve must be according to the invention fixed relative to the socket. This means that the sleeve is firmly supported against the bushing.
  • the support of the pressure relief sleeve is preferably only in the axial direction.
  • a radial clearance so that the movement of the hydraulic piston is ensured In order to ensure a tightness despite a large radial clearance between the sleeve and the sleeve, which prevents hydraulic fluid from the supply port on the outside sleeve pass can, is provided in a particularly advantageous manner in the region of this radial clearance, a sealing ring which compensates for the radial clearance.
  • the hollow piston is completely pressure-balanced in a particularly advantageous manner.
  • the controllability a bit.
  • the assembly is simplified, since the hollow piston is preferably configured such that its first area to be inserted has a smaller diameter than its subsequently inserted area. The likelihood of damage to the functional surfaces / sealing surfaces during assembly is reduced in particular during manual assembly.
  • the Schwenkmotornockenwellenversteller 14 allows a continuous adjustment of the camshaft relative to the crankshaft.
  • the Schwenkmotornockenwellenversteller 14 has a cylindrical stator 1 which is rotatably connected to a drive wheel 2.
  • the drive wheel 2 is a sprocket over which a chain, not shown, is guided.
  • the drive wheel 2 may also be a toothed belt wheel, via which a drive belt is guided as a drive element.
  • the stator 1 is drivingly connected to the crankshaft.
  • the stator 1 comprises a cylindrical stator base body 3, on the inside of which protrude webs 4 at equal intervals radially inwardly. Between adjacent webs 4 gaps 5 are formed, in which, via an in Fig. 2 controlled hydraulic valve 12 shown controlled, pressure medium is introduced. Between adjacent webs 4 protrude wings 6, which project radially outward from a cylindrical rotor hub 7 of a rotor 8. These wings 6 divide the spaces 5 between the webs 4 each in two pressure chambers 9 and 10th
  • the webs 4 lie with their end faces sealingly against the outer circumferential surface of the rotor hub 7.
  • the wings 6 in turn lie with their end faces sealingly against the cylindrical inner wall of the stator main body 3.
  • the rotor 8 is rotatably connected to the camshaft, not shown. In order to change the angular position between the camshaft and the crankshaft, the rotor 8 is rotated relative to the stator 1. For this purpose, depending on the desired direction of rotation, the pressure medium in the pressure chambers 9 or 10 is pressurized, while the respective other pressure chambers 10 or 9 are relieved to the tank. In order to pivot the rotor 8 counter to the stator 1 in the illustrated position counterclockwise, radial hub bores 11 in the rotor hub 7 are pressurized by the hydraulic valve 12. In contrast, in order to pivot the rotor 8 in a clockwise direction, the hydraulic valve 12 pressurizes further radial hub bores 13 in the rotor hub 7.
  • These further radial hub bores 13 are arranged axially and circumferentially offset from the first-mentioned hub bores 11.
  • the hydraulic valve 12 is inserted as a so-called central valve in the rotor hub 7 and bolted to the underlying camshaft.
  • the rotor 8 is biased in a graphically unillustrated manner by means of acting as a compensation spring coil spring against the stator 1 torsionally elastic.
  • Fig. 2 This has a helical bushing 52 with an axial supply port P, from which the coming of a non-illustrated oil pump hydraulic pressure alternately to a first working port A or a second working port B can be passed.
  • These two working connections A, B lead in annular grooves 31, 32 in the rotor hub 7.
  • the first working port A leads over this assigned first annular groove 31 in the said radial hub bores 11.
  • the leads second working port B via this associated 32 annular groove in the other hub bores 13th
  • first working port A associated with the first annular groove still leads to the use of camshaft alternating torques associated port A1, which is formed by a transverse bore 21 in the sleeve 52.
  • the bushing 52 also has two radial tank connections T1, T2 and an axial tank connection T3.
  • the first two radial tank ports T1, T2 are axially adjacent to each other next to the two working ports A, B arranged.
  • the order of the radial ports from the engine to an actuator 43 is sequentially T1-T2-A-A1-B.
  • the axial or third tank connection T3, leads out of the hydraulic valve 12 at a screw head 49 of the screw-shaped bush 52.
  • the first radial tank port T1 does not serve to remove oil from the respectively to be relieved pressure chambers 9 and 10. Instead, this first tank port T1 serves as a volume compensation or for ventilation.
  • the sleeve 52 ends on the motor side with an external thread 53, which is screwed into an internal thread of the camshaft, not shown, and the rotor 8 rotatably clamped against the camshaft frictionally engaged.
  • the rotor hub 7 is on the one hand via a thin friction disc at the front end of the camshaft and on the other hand to the screw head 49 of the socket 52 at.
  • Such a friction disc - but with oil ducts - is for example the subject of DE 10 2009 050 779.5 ,
  • a hollow piston 54 is displaceable. This is only in Fig. 2 Ansatzweise illustrated plunger 48 of an electromagnetic linear actuator 43 to a piston head 51 of the hollow piston 54 at.
  • the hydraulic pressure coming from an axial supply port P is directed to a second working port B.
  • the in Fig. 1 shown pressure chambers 9 applied via the hub bores 13 with hydraulic pressure. The thus necessarily out of the oppositely directed pressure chambers 10 via the hub bores 11 to the first working port A guided hydraulic fluid can be derived from the hydraulic valve 12 to the second tank port T2.
  • a pot-shaped closed sleeve 55 is inserted into the hollow piston 54 hollowed out in the form of a blind hole 56.
  • Their sleeve bottom 50 prevents a pressure of the supply port P acts on the blind hole bottom 57 and thus acts on the hollow piston 54 in addition to a helical compression spring 58 with a force.
  • the electromagnetic linearly effective actuator 43 disengaging the plunger 48 does not have to apply a force against the varying pressure of the supply port P.
  • the control or control quality of the central valve is thus very good.
  • the sleeve bottom 50 is supported via a wall 23 of the sleeve 55 to the sleeve 52.
  • the entire sleeve 55 is female.
  • the sleeve 55 has on its side facing away from the plunger 48 a radially outwardly cantilevered collar 25 which rests on its side facing the plunger 48 side 26 on a shoulder 24 of the bushing 52.
  • On the other side 27 of the sleeve 52 is an axial securing ring 29, which is inserted into an inner annular groove of the bush 52.
  • the sleeve 55 is provided with elongated slots 59 which are so long that they allow in all axial positions of the sleeve 54 relative to the axially fixed sleeve 54 movable piston 54 to transverse bores 60 in the hollow piston.
  • a band-shaped check valve 61 is provided, which bears radially on the hollow piston 54 and thereby conceals the transverse bores 60.
  • This radially outboard check valve 61 has the function of a pump check valve.
  • the introduced by the oil pump in the annular space 62 pressure is locked by the working as a pump check valve check valve 61 and another check valve 33, so that this pressure can be relieved only via a gap 38 into the pressure chambers 9 in.
  • This pressure in the annular space 62 together with the further check valve 33, prevents the penetration of hydraulic fluid from the transverse bore 21, which communicates with the pressure chambers 10 via the port A1.
  • the hydraulic fluid from the pressure chambers 10 is therefore guided exclusively via the first working port A to the second tank port T2, as long as the internal pressure in the pressure chambers 10 and the transverse bore 21 does not rise above the pressure in the annular space 62.
  • the said annular groove 16 on the hollow piston 54 is bounded on either side by a guide web 36 and 30, respectively.
  • the first guide web 36 is thus arranged on the motor side by the central or second guide web 30.
  • the third guide web 28 is arranged on the actuator side by the central or second guide web 30.
  • the transverse bore 60 in the hollow piston 54 coming from the supply connection P is arranged axially between the second guide web 30 and the third guide web 28.
  • the function of the latter two guide webs 30, 28 is the following:
  • the hollow piston 54 is moved against the force of the helical compression spring 58 in its end position, which is also the second valve position.
  • the third guide bar 28 closes the gap 38 and correspond Fig. 3 a gap 37 on its the linear actuator 43 zugewanden side surface 39 free. With this then open gap 37 of the previously closed by a sealing gap 41 access to the annular space 62 is created.
  • the supply pressure coming from the supply port P can be conducted to the first working port A.
  • the rotor 8 pivots in the opposite direction of rotation.
  • the second working port B is relieved against the third tank port T3.
  • the third guide bar 28 at a transverse bore 35 of the second working port B, a gap 34 free.
  • the hollow piston 54 is provided at this end with a transverse bore 46 which has several functions.
  • the volume of hydraulic fluid displaced by the sleeve base 50 can flow away through the transverse bore 46 to the third tank connection T3.
  • the slight pressure of the hydraulic fluid draining from the second working port B to the third tank port T3 acts on both sides of the piston head 51, so that the hollow piston 54 is also pressure-balanced in this regard.
  • camshaft alternating torques Use of the camshaft alternating torques is in contrast to the valve position according to Fig. 2 not provided. Peak pressures due to camshaft alternating torques are conducted directly from the second working port B to the third tank port T3.
  • Fig. 4 shows a between the two illustrated extreme valve positions of the hollow piston 54 lying locking center position.
  • the two working ports A, B are closed by the two guide webs 28, 30.
  • the hydraulic fluid in the two directions of rotation associated pressure chambers 9, 10 is locked.
  • a small volume flow forces itself out of the annular space 62 on the guide webs 28, 30 over to the two working ports A, B and compensates for leakage losses and ensures a damped pivoting of the rotor 8 according to the DE 198 23 619 A1 ,
  • the hydraulic fluid flows from the tank ports T1, T2, T3 into the timing drive box.
  • this control gear is designed with a chain, lubricates the hydraulic fluid at the same time the timing drive.
  • wet belt drives There are also known wet belt drives.
  • the bushing 52 has an inside annular groove into which an axial securing ring 40 is set.
  • This Axialommesring 40 is used in de-energized actuator 43 as an axial stop for the valve position according to Fig. 2 .
  • the check valve 61 could also extend from the second guide web 30 to the third guide web 28, in which case a further axial lock 42 is not provided.
  • the annular groove 16 is sealed by a sealing gap 45 with respect to an end face 44 of the hollow piston 54 which points towards the supply connection P.
  • the hydraulic valve 12 would also function without the region of the hollow piston 54 which extends to the first guide web 36 after the central guide web 30.
  • the helical compression spring 58 would abut axially on the guide web 30.
  • the in Fig. 2 to Fig. 4 illustrated shaping of the hollow piston 54 with the first guide bar 36 a particularly high control quality. So is in particular in Fig. 2 It can be seen that the hydraulic current flowing from the first working connection A to the second tank connection T2 applies the same force to the first two guide webs 36, 30 in the axially opposite directions.
  • the hollow piston 54 is also pressure-balanced in this regard.
  • a pressure in the annular groove 16 can build up, since the second tank port T2 forms a throttle point with decreasing flow cross-section. If the first guide web 36 was dispensed with, then the end face of the hollow piston 54 flowing to the first two tank connections T1, T2 could bear against and effect a force in the direction pointing to the actuator 43.
  • the helical compression spring 58 is disposed in an annular space 64 radially outside the sleeve 55, which leads via an opening 63 in the socket 52 to a tank port T1.
  • Fig. 5 shows a hydraulic valve 112 in another embodiment.
  • the hollow piston 154 abuts in the direction pointing to the actuator to a perforated cover 76 which is fixed to the sleeve 152.
  • the sleeve 155 is designed as a rotary member with a longitudinal bore 74 and a transverse bore 75.
  • the hollow piston 154 is not made in one piece with a piston bottom 151, but axially pressed into the continuously hollow piston 154 drilled as an insert 65.
  • This use 65 is pierced outside of the central axis 122 with holes 146, which relieves analogous to the transverse bores 46 of the previous embodiment, a space 66 between the piston bottom 151 and the sleeve bottom 150 to the second tank outlet T2 out.
  • This relief is necessary because it comes due to the relative displacement between the hollow piston 154 and the sleeve 155 to the volume change of the space 66, which must be compensated by the holes 146 with oil and / or air.
  • a sealing ring 68 can be used to compensate for production-related tolerances in the previous embodiment.
  • the annular groove 69 for receiving the sealing ring 68 may be provided both in the sleeve 55 and 155 and in the sleeve 52 and 152, respectively.
  • the check valves can be designed with or without overlap. From the EP 1 703 184 B1 Already a Auspatisternativative a tape-shaped check valve is known. Instead of the unbalanced distribution of the check valves to be closed through openings claimed in this European patent, it is also possible to dispense with an overlap and provide an anti-rotation on the check valve.

Claims (11)

  1. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur, présentant une soupape hydraulique (12) dotée d'une douille (52) et d'un piston creux (54) apte à coulisser axialement à l'intérieur de cette douille au moyen d'un organe d'ajustement (43) en opposition à la force d'un ressort hélicoïdal de poussée (58),
    le piston creux étant doté d'un fond (51) de piston sur lequel repose l'organe d'ajustement (43),
    un manchon (55) fixé sur la douille étant disposé à l'intérieur du piston creux (54) de manière à pouvoir coulisser par rapport à lui et sa paroi présentant une perforation (59) qui conduit à au moins une ouverture (60) du piston creux (54) qui peut conduire une pression d'alimentation (P) appliquée à l'intérieur de l'enveloppe (55) en variante à deux chambres (9, 10) sous pression du dispositif pivotant (14) d'ajustement d'arbre à cames de moteur,
    le manchon (55) présentant un fond (50) de manchon qui ferme l'espace intérieur du piston creux (54), caractérisé en ce que
    sur son extrémité (67) non tournée vers l'organe d'ajustement, l'enveloppe (155) présente par rapport à la douille (152) un jeu radial compensé au moyen d'un joint d'étanchéité (68) élastiquement déformable, de manière à fermer hermétiquement l'interstice entre l'enveloppe (155) et la douille (152) et
    en ce que l'enveloppe (55, 155) présente une possibilité limitée de déplacement axial par rapport à la douille (52, 152).
  2. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon la revendication 1, caractérisé en ce que dans la direction de l'axe central (22) de la soupape hydraulique (12), les perforations (59) ont une longueur telle que l'ouverture (60) débouche dans les deux positions (figure 2, figure 3) du piston dans la perforation (59) pour alimenter les deux chambres (9, 10) sous pression.
  3. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications précédentes, caractérisé en ce que l'ouverture (60) est recouverte du côté radialement extérieur par une soupape anti-retour (61) qui l'entoure en forme de ruban.
  4. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon la revendication 3, caractérisé en ce que la soupape anti-retour (61) est une soupape anti-retour de pompe qui bloque le retour de l'écoulement depuis un espace annulaire (62) jusque dans le raccordement d'alimentation (P) lorsque la pression qui règne dans l'espace annulaire (62) est sensiblement égale à la pression de raccordement (P) d'alimentation, cette soupape anti-retour (61) étant bloquée axialement au moyen de nervures (42, 28) qui s'étendent radialement en partant du piston creux (54).
  5. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications précédentes, caractérisé en ce que la douille (52) présente pour l'ajustement des deux chambres (9, 10) sous pression des raccordements de travail (A, B) situés à distance mutuelle dans la direction axiale et entre lesquels un raccordement (A1) destiné à utiliser le couple alternatif de l'arbre à came est prévu.
  6. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications précédentes, caractérisé en ce que le piston creux (54) présente deux nervures de guidage (30, 28) entre lesquelles est disposée l'ouverture (60), la pression hydraulique provenant de l'ouverture (60) pouvant être amenée par les deux nervures de guidage (30, 28) sur l'un des raccordements de travail (B ou A) tandis que la pression hydraulique provenant de l'autre raccordement de travail (A ou B) peut être amenée par l'une des deux nervures de guidage (30 ou 28) sur l'un de deux raccordements (T2, T3) de réservoir.
  7. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon la revendication 6, caractérisé en ce que le raccordement de travail (A) situé le plus près du raccordement d'alimentation (P) peut être amené par une rainure annulaire (16) du piston creux (54) sur le raccordement (T2) de réservoir qui est fermé hermétiquement vis-à-vis d'un côté frontal (44), tourné vers le raccordement d'alimentation (P), du piston creux (54) par l'intermédiaire d'un interstice d'étanchéité (45).
  8. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications précédentes, caractérisé en ce que le ressort hélicoïdal de poussée (58) est disposé radialement dans un espace annulaire (64) du manchon (55) et qui conduit par une ouverture (63) ménagée dans la douille (52) à un raccordement (T1) de réservoir.
  9. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications précédentes, caractérisé en ce qu'un espace (66) qui est délesté vis-à-vis d'un raccordement (T2) de réservoir au moyen de découpes (alésages transversaux 46) est formé à l'intérieur du piston creux (54) entre le fond (50) d'enveloppe et le fond (51) de piston.
  10. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications précédentes, caractérisé en ce que la soupape hydraulique (12) est une soupape centrale.
  11. Dispositif pivotant (14) d'ajustement de l'arbre à cames d'un moteur selon l'une des revendications 1 à 9, caractérisé en ce que la soupape hydraulique (12) est une soupape dite "cartridge".
EP11159294.5A 2010-05-03 2011-03-22 Soupape hydraulique Not-in-force EP2386731B1 (fr)

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DE102010019005.5A DE102010019005B4 (de) 2010-05-03 2010-05-03 Schwenkmotorversteller

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DE102010019005B4 (de) 2017-03-23
DE102010019005A1 (de) 2011-11-24
CN102235196A (zh) 2011-11-09
JP2011241823A (ja) 2011-12-01
JP5759654B2 (ja) 2015-08-05
US20110266479A1 (en) 2011-11-03
DE102010019005A9 (de) 2013-05-29
US8505582B2 (en) 2013-08-13
EP2386731A1 (fr) 2011-11-16

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