EP2386731A1 - Hydraulic valve - Google Patents

Abstract

The valve has a hollow piston with a piston bottom on which an actuator is applied. The piston is axially moved in the bush by the actuator against force of a screw-type pressure spring. A sleeve is provided in the piston and secured in the bush for movement relative to the piston. A wall of the sleeve includes a though-opening that leads to an opening of the piston. The opening leads a supply pressure that is applied inside the sleeve to two sets of pressure chambers (9, 10) of an oscillating-motor camshaft adjuster (14). A sleeve bottom of the sleeve closes an inner space of the piston.

Description

The invention relates according to the preamble of claim 1, a hydraulic valve and according to claim 9 whose use for a Schwenkmotornockenwellenversteller.

From the DE 10 2004 038 252 A1 A hydraulic valve for a Schwenkmotornockenwellenversteller is already known. 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. Within the hollow piston, a sleeve is provided. By means of the hydraulic valve, a supply pressure P can be performed alternatively to two load ports A, B or two pressure chambers of the Schwenkmotornockenwellenverstellers. 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.

From the DE 10 2005 013 085 B3 already known as a cartridge valve hydraulic valve is known. This hydraulic valve has axially offset from each other three ports B, P, A, which constitute as openings in a socket of the hydraulic valve. Within this socket a band-shaped check valve is used.

The object of the invention is to provide a low-cost and small Schwenkmotornockenwellenversteller with a high control quality.

This object is achieved with the features of claim 1.

According to the invention, a sleeve is arranged relatively displaceable within the hollow piston of the hydraulic valve. However, this sleeve can maintain its position relative to a socket within which the hollow piston is displaceable. there may be provided a limited axial play and radial clearance, 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. Thus, these forces do not act on the piston head of the hollow piston, which serves as the support for an actuator. Since 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. This is of particular advantage, since 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 the Schwenkmotornockenwellenverstellers 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 are particularly important in internal combustion engines with few - for example, three - cylinders come into play. Furthermore, influencing parameters are still the strength of the gas exchange valve springs and the camshaft speed.

The 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.

The use of the camshaft alternating torques can be done for both directions of rotation or only for one direction of rotation. In the case of the use of the camshaft alternating torques in only 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.

The use of the camshaft alternating moments is carried out by means of a check valve that 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. Such a central valve has space advantages. The opposite of a central valve are the external hydraulic valves for operating the Schwenkmotornockenwellenverstellers. In the case of the external hydraulic valve, the hydraulic passages for the camshaft adjustment run from the pivoting motor 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, however, is arranged radially inside the rotor hub of the Schwenkmotornockenwellenverstellers. 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. On the other hand, if the hydraulic fluid had a long way from the rotor hub to an external hydraulic valve, the line losses would consume the advantage with increasing line length. However, with the direct action of the camshaft alternating torques via a central valve instead of via a damping section, 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.

However, 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 frontally into the sleeve. 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. By contrast, in an advantageous embodiment, the sleeve has a radial play, so that the mobility of the hydraulic piston is ensured. In order to ensure a tightness despite a large radial play between the bush and the sleeve, which prevents hydraulic fluid from the supply port to the socket outside pass, a sealing ring is provided in a particularly advantageous manner in the field of this radial clearance, which compensates for the radial clearance.

If the supply pressure in the normal operation of the hydraulic valve constantly applied to the sleeve bottom, then a support can be sufficient only in this direction, since a displacement of the sleeve in the direction away from the actuator direction is then prevented by the supply pressure. This is especially true when the sleeve is received in the socket with a fit, so that even frictional forces between the sleeve and the socket act. These friction forces are therefore to be kept low by means of appropriate material pairings, tolerances, component surfaces and constructive measures.

The hollow piston is completely pressure-balanced in a particularly advantageous manner. However, it is also possible to perform the hollow piston with slightly varying outer diameters. In this case, unfortunately, the controllability a bit. In return, 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 claims 9 and 10 have particularly advantageous embodiments, which compensate for production-related tolerances via a radial or axial clearance, so that it can not come to jamming of the hollow piston.

Further advantages of the invention will become apparent from the other claims, the description and the drawings.

The invention is explained in more detail below with reference to two exemplary embodiments.

Show

• Fig. 1 a swivel camshaft adjuster in a cutaway view,
• Fig. 2 in a first valve position in a half section, a hydraulic valve for adjusting a Schwenkmotornockenwellenverstellers according to Fig. 1 .
• Fig. 3 the hydraulic valve off Fig. 2 in a second valve position for adjustment in the other direction of rotation,
• Fig. 4 the hydraulic valve off Fig. 2 and Fig. 3 in a lock middle position and
• Fig. 5 in a further embodiment, a hydraulic valve for adjusting a Schwenkmotornockenwellenverstellers according to Fig. 1 ,

With a Schwenkmotornockenwellenversteller 14 according to Fig. 1 During operation of an internal combustion engine, the angular position on the camshaft is changed. By turning the camshaft, the opening and closing times of the gas exchange valves are shifted so that the internal combustion engine brings its optimum performance at the respective speed. 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. In the exemplary embodiment, 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. About this drive element and the drive wheel 2, 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. However, the leads second working port B via this associated 32 annular groove in the other hub bores 13th

In the 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.

In addition, 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. Here, 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, on the other hand, 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. For this purpose, 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 ,

Within the sleeve 52, 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. In the illustrated de-energized state of the actuator, the hydraulic pressure coming from an axial supply port P is directed to a second working port B. From this 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.

The supply connection P coming from an oil pump of the internal combustion engine, for example via the camshaft, is axial. 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. Thus, 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. As a result, 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. Thus, the sleeve 55 is secured in both directions against axial displacement along a central axis 22.

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. Radially outside of these transverse bores 60, 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. Thus, a hydraulic pressure from the axial supply port P through the slots 59 and the check valve 61 in an annular space 62 reach radially outwardly of the check valve 61. On the other hand, a return flow from this annulus 62 to the supply port P prevented by an internal pressure that is above the pressure at the supply port P is blocked by the check valve 61.

In the in Fig. 2 shown first valve position of the hollow piston 54 relative to the sleeve 52 thus hydraulic fluid from the oil pump via the supply port P and the check valve 61 to the second working port B is promoted. In this case, the pressure chambers 9 are pressed over the hub bores 13, so that the rotor 8 is pivoted relative to the stator 1 in the one direction of rotation. The pressure chambers 10, which inevitably shrink, press the hydraulic fluid via the hub bores 11 to the first working port A. From there, the hydraulic fluid passes via an annular groove 16 on the hollow piston 54 to the tank port T2. 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.

Once the camshaft due to their change moments endeavors to rotate in the direction to be adjusted, the pressure in the pressure chambers 10 and the transverse bore 21 increases abruptly and strongly. Once this pressure is increased so far above the pressure in the annular space 62, the losses at the first working port A and the preloaded further check valve 33 are overcome, the annular space 62 via the second working port B the hydraulic fluid "sucking" pressure chambers 9 sufficient flow for a Quick adjustment available, as the oil pump can not provide alone. This connection is also in the DE 10 2006 012 775 A1 explained in more detail.

The said annular groove 16 on the hollow piston 54 is bounded on either side by a guide web 36 and 30, respectively. The second guide web 30, together with another guide web 28, forms a further annular groove 47, which forms the radially inner boundary of the annular space 62. Thus, the hollow piston 54 in the direction pointing from the engine to the linear actuator 43 three axially successive guide webs 36, 30, 28, with which the hollow piston 54 is guided within the sleeve 52. The first guide web 36 is thus arranged on the motor side by the central or second guide web 30. By contrast, 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:

If the linear actuator 43 energized maximum, 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. In this case, 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. Thus, 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. For this purpose, the third guide bar 28 at a transverse bore 35 of the second working port B, a gap 34 free. From there, the hydraulic fluid flows outside along the hollow piston 54 to the third tank port T3. In this case, the hollow piston 54 is provided at this end with a transverse bore 46 which has several functions. On the one hand, with the axial displacement of the hollow piston 54, 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. On the other hand, 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.

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. In this locking middle position, the two working ports A, B are closed by the two guide webs 28, 30. Thus, the hydraulic fluid in the two directions of rotation associated pressure chambers 9, 10 is locked. At most, 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. In particular, when this control gear is designed with a chain, lubricates the hydraulic fluid at the same time the timing drive. There are also known wet belt drives.

At the actuator end, the bushing 52 has an inside annular groove into which an axial securing ring 40 is set. This Axialsicherungsring 40 is used in de-energized actuator 43 as an axial stop for the valve position according to Fig. 2 , In an alternative embodiment, 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.

Between the second guide web 30 and the third guide web 28 is an annularly from the hollow piston 54 radially outwardly extending axial securing 42 is provided. In order for this Axialsicherung 42 on the one hand and the third guide bar 28 on the other hand, the return spring 61 axially. This prevents that the return spring 61 can be moved so far that it no longer covers the transverse bore 60 sufficiently.

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. In principle, 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. In this case, the helical compression spring 58 would abut axially on the guide web 30. However, 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. That is, 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. There is no additional connection for special use of camshaft alternating torques. The hollow piston 154 abuts in the direction pointing to the actuator to a perforated cover 76 which is fixed to the sleeve 152. Instead of the deep-drawn sheet metal sleeve 55 of the previous embodiment, 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.

• dem Versorgungsanschluss P und
• dem ersten Tankabfluss T1 bzw. sogar dem ersten Arbeitsanschluss A besteht. Das relativ große radiale Spiel ermöglicht dabei neben einem axialen Spiel, dass der Hohlkolben 154 im geringen Maße gekippt und parallel zur Zentralachse 122 versetzt werden kann. Damit werden fertigungsbedingte bzw. toleranzbedingte Koaxialitätsfehler zwischen der Hülse 155 , dem Hohlkolben 154 und der Buchse 152 ausgeglichen, so dass es nicht zum Verklemmen des Hohlkolbens 154 kommen kann. Um das axiale Spiel zu ermöglichen - aber auch zu begrenzen - weist das Ende 67 der Hülse 155 analog dem Bund 25 der vorangegangenen Ausführungsform an den beiden voneinander hinfort weisenden Richtungen Anschläge 71, 72 auf. Der eine Anschlag 71 kann dabei an einem Absatz 124 der Buchse 152 zur Anlage kommen. Hingegen kann der andere Anschlag 72 an einer in die Buchse 152 fest eingesetzte Anschlaghülse 73 zur Anlage kommen.

The sleeve 155 has at its end remote from the actuator 67 an annular groove 69 in which a designed as an O-ring sealing ring 68 is added. The outer diameter of the sleeve 155 at this end 67 has a large radial clearance to the associated receiving bore 70 of the bush 152. This great game is compensated by the sealing ring 68, so that oil despite this game no hydraulic leakage connection between

• the supply connection P and
• the first tank drain T1 or even the first working port A consists. The relatively large radial clearance allows in addition to an axial clearance that the hollow piston 154 tilted to a small extent and can be offset parallel to the central axis 122. Thus, production-related or tolerance-related Koaxialitätsfehler between the sleeve 155, the hollow piston 154 and the sleeve 152 are compensated, so that it can not come to jamming of the hollow piston 154. In order to enable - but also to limit - the axial clearance, the end 67 of the sleeve 155, analogous to the collar 25 of the preceding embodiment, has stops 71, 72 in the two directions pointing away from each other. The one stop 71 can come to rest on a shoulder 124 of the bushing 152. On the other hand, the other stop 72 can come to rest on a stop sleeve 73 firmly inserted into the bushing 152.

As in this embodiment, 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 Ausgestaltungsternativative 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.

The described embodiments are only exemplary embodiments. A combination of the described features for different embodiments is also possible. Further, in particular not described features of the device parts belonging to the invention are to be taken from the geometries of the device parts shown in the drawings.

LIST OF REFERENCE NUMBERS

1

stator

2

drive wheel

3

stator base

4

Stege

5

interspaces

6

wing

7

rotor hub

8th

rotor

9

pressure chambers

10

pressure chambers

11

hub bores

12

hydraulic valve

13

hub bore

14

Oscillating motor camshaft adjuster

15

ring groove

16

ring groove

17

ring groove

18

ring groove

19

cross hole

20

cross hole

21

cross hole

22

central axis

23

wall

24

paragraph

25

Federation

26

page

27

other side

28

third guide bar

29

axial securing

30

second guide bridge

31

ring groove

32

ring groove

33

check valve

34

gap

35

cross hole

36

first guide bridge

37

gap

38

gap

39

side surface

40

axial securing

41

sealing gap

42

axial safety

43

actuator

44

front

45

sealing gap

46

cross hole

47

ring groove

48

tappet

49

screw head

50

sleeve base

51

piston crown

52

Rifle

53

external thread

54

hollow piston

55

shell

56

blind

57

Blind basis

58

Helical compression spring

59

slots

60

cross hole

61

check valve

62

annulus

63

opening

64

Rough m

65

commitment

66

Rough m

67

The End

68

seal

69

ring groove

70

location hole

71

attack

72

attack

73

stop sleeve

74

longitudinal bore

75

cross hole

76

cover

112

hydraulic valve

122

central axis

124

paragraph

146

drilling

150

sleeve base

151

piston crown

152

Rifle

154

hollow piston

155

shell

T1

tank drain

T2

tank drain

Claims (13)

Schwenkmotornockenwellenversteller (14) with a hydraulic valve (12) having a sleeve (52) and within this by means of an actuator (43) against the force of a helical compression spring (58) axially displaceable hollow piston (54) having a piston head (51) in which the actuator (43) rests, wherein within the hollow piston (54) a sleeve fixed sleeve (55) is arranged relative to this slidably, the wall of which has an opening (59) which at least one opening (60) of the hollow piston (54) leads to a pressure within the sleeve (55) applied supply pressure (P) alternatively to two pressure chambers (9, 10) of the Schwenkmotornockenwellenverstellers (14), wherein the sleeve (55) a the interior of the hollow piston (54) closing sleeve bottom (50 ) having. Schwenkmotornockenwellenversteller (14) according to claim 1, characterized in that the openings (59) in the direction of a central axis (22) of the hydraulic valve (12) are so long that the opening (60) in the two piston positions (Fig. 2, Fig. 3) for acting on the two pressure chambers (9, 10) in the opening (59) opens. Schwenkmotornockenwellenversteller (14) according to any one of the preceding claims, characterized in that the opening (60) is covered radially outwardly by a band-shaped circumferential non-return valve (61). Pivot camshaft phaser (14) according to claim 3, characterized in that the check valve (61) is a pump check valve, which blocks the return flow from an annular space (62) in the supply port (P) when the pressure in the annular space (62) is almost equal to the pressure from the supply port (P), said check valve (61) being axially by means of radially from the hollow piston (54) extending webs (42, 28) is secured. Schwenkmotornockenwellenversteller (14) according to one of the preceding claims, characterized in that the bushing (52) for adjusting the two pressure chambers (9, 10) axially spaced working connections (A, B), between which a connection (A1) for the use of Camshaft alternating torques is provided. Schwenkmotornockenwellenversteller (14) according to any one of the preceding claims, characterized in that the hollow piston (54) has two guide webs (30, 28) between which the opening (60) is arranged, wherein of the two guide webs (30, 28) of The hydraulic pressure coming from the opening (60) can be conducted to the one working port (B or A), while the hydraulic pressure coming from the other working port (A or B) can be directed from one of the two guide webs (30 or 28) to one of two tank ports (T2, T3) is conductive. Schwenkmotornockenwellenversteller (14) according to claim 6, characterized in that the supply port (P) closer working port (A) via an annular groove (16) of the hollow piston (54) on the tank port (T2) is feasible, which opposite to the supply port (P) facing end face (44) of the hollow piston (54) via a sealing gap (45) is sealed. Pivot camshaft adjuster (14) according to one of the preceding claims, characterized in that the helical compression spring (58) is arranged in an annular space (64) radially outside the sleeve (55), which via an opening (63) in the socket (52) a tank connection (T1) leads. Schwenkmotornockenwellenversteller (14) according to any one of the preceding claims, characterized in that the sleeve (155) facing away from the actuator end (67) relative to the bush (152) has a radial clearance, which is bridged by means of an elastically deformable seal (68) such that a gap between the sleeve (155) and the sleeve (152) is sealed. Schwenkmotornockenwellenversteller (14) according to claim 9, characterized in that the sleeve (55, 155) relative to the bushing (52, 152) has a limited axial displaceability. Schwenkmotornockenwellenversteller (14) according to any one of the preceding claims, characterized in that within the hollow piston (54) between the sleeve bottom (50) and the piston head (51) a space (66) is formed by means of recesses (transverse bores 46) opposite a tank connection (T2) is relieved. Schwenkmotornockenwellenversteller (14) according to any one of the preceding claims, characterized in that the hydraulic valve (12) is a central valve. Schwenkmotornockenwellenversteller (14) according to one of the claims 1 to 11, characterized in that the hydraulic valve (12) is a cartridge valve.

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