Patent Application
20110266479
This application claims the benefit of German patent application number DE 10 2010 019 005.5 filed on May 3, 2010, which is incorporated herein by reference in their entirety and for all purposes.
The invention relates to a hydraulic valve and its use for an oscillating-motor camshaft adjuster.
A hydraulic valve for an oscillating-motor camshaft adjuster is already known from DE 10 2004 038 252 A1. The hydraulic valve has a bush and a hollow piston that can be shifted axially inside this bush against the force of a screw-type pressure spring by means of an actuator. A sleeve is provided inside the hollow piston. A supply pressure P can be guided alternatively to two working ports A, B or two pressure chambers of the oscillating-motor camshaft adjuster by means of the hydraulic valve. Two tank ports T1, T2 are provided. The sequence of the radial ports is P-T1-B-A. The second tank port T2 then follows as an axial port on the front side.
A hydraulic valve designed as a cartridge valve is already known from DE 10 2005 013 085 B3. This hydraulic valve has three ports B, P, A, which are axially displaced relative to one another and which are present as openings in a bush of the hydraulic valve. A band-shaped non-return valve is inserted inside this bush.
The object of the invention is to create a cost-effective and small oscillating-motor camshaft adjuster having a high control performance.
This problem is solved according to the embodiments of the invention set forth herein.
According to one example embodiment of the invention, a hydraulic valve for an oscillating-motor camshaft adjuster is provided. A sleeve is disposed in a relatively moveable manner inside the hollow piston of the hydraulic valve. This sleeve, however, can maintain its position relative to a bush within which the hollow piston can be moved. In this way, a limited axial play and a limited radial play can be provided, which prevents a jamming of the parts moving against each other or equilibrates tolerances. The sleeve has a sleeve bottom that seals off the inside space of the hollow piston. This sleeve bottom is solidly supported relative to the bush, so that the forces arising from the pressure from a supply port P are supported at the bush via the sleeve bottom and the sleeve. Because of this, these forces do not act on the piston bottom of the hollow piston, which serves for support for an actuator. Thus, since the hollow piston is free of axial forces from the supply pressure, the axial position of the hollow piston can be controlled by the actuator, without needing to consider the supply pressure. This is of particular advantage, since the supply pressure can fluctuate depending on how it is provided. Usually, since an oil pump which is driven mechanically by the internal combustion engine is used, the supply pressure fluctuates depending on the engine speed and temperature or viscosity of the oil. In addition, other factors may play a role.
The particularly high control performance that can be achieved according to the invention offers a particular advantage, if it is combined with a hydraulic construction that utilizes the camshaft alternating torques for supporting the angle adjustment by means of the oscillating-motor camshaft adjuster. That is, this utilization establishes higher requirements for control of the hydraulic valve, since these camshaft alternating torques operate in a non-uniform and rapidly fluctuating manner. Such a function for utilizing camshaft alternating torques is already known from DE 10 2006 012 733 B4 and DE 10 2006 012 775 B4. The hydraulic valve according to the invention can consequently be configured in such a way that it makes possible, in a particularly advantageous way, the utilization of pressure fluctuations in the pressure chambers of the oscillating-motor camshaft adjuster that are assigned to the first working port B, in order to supply the pressure chambers assigned to the opposite direction of rotation with sufficiently fluid flow volume. These pressure fluctuations result from the camshaft alternating torques that are established on the camshaft in reaction to the forces of the gas exchange valves. In each case, the fewer the number of combustion chambers there is per camshaft, the larger will be the camshaft alternating torques, so that the advantages of utilizing camshaft alternating torques are particularly effective in the case of internal combustion engines with few, for example, three, cylinders. Further, the influence parameters are still the strength of the springs of the gas exchange valves and the camshaft rpm.
The phase adjustment of the camshaft can thus be produced rapidly. In addition, as a consequence of utilizing camshaft alternating torques in an advantageous manner, it is possible to make an adjustment with a relatively low oil pressure. A small dimensioning of the oil pump made possible in this way improves the efficiency of the internal combustion engine. The flow volumes of hydraulic fluid that are saved are available for other uses, such as, for example, adjusting the hydraulic valve stroke.
The camshaft alternating torques can be utilized for both directions of rotation, but they can also be utilized for only one direction of rotation. In the case of utilizing the camshaft alternating torques only in one direction of rotation, a flat spiral spring according to DE 10 2006 036 052 A1 can be used, which then compensates for the additional adjusting forces in one direction of rotation.
The camshaft alternating torques are utilized in this case by means of a non-return valve that can be designed particularly in a band shape.
The hydraulic valve in this case can be designed as a central valve in a particularly preferred embodiment, whereby the supply pressure is introduced via the camshaft. Such a central valve has advantages relative to structural space. External hydraulic valves for actuating the oscillating-motor camshaft adjuster represent the counterpart of a central valve. In the case of an external hydraulic valve, the hydraulic channels for the camshaft adjustment run from the oscillating-motor camshaft adjuster to a separate control drive cover having the hydraulic valve screwed thereon or, to the cylinder head having the hydraulic valve screwed therein. In contrast, the central valve, which is also hydraulic, is disposed radially inside the rotor hub of the oscillating-motor camshaft adjuster. In the case of the central valve, the method employed for the more rapid adjustment of the oscillating-motor camshaft adjuster, which is described in DE 10 2006 012 733 B4 and DE 10 2006 012 775 B4 named above, is particularly effective, since the hydraulic fluid from the chambers assigned to one direction of rotation has a short path into the chambers assigned to the other direction of rotation. If, in contrast, the hydraulic fluid were to have a long path from the rotor hub to an external hydraulic valve, then with increasing line length, the line losses would obliterate the advantage. Of course, challenges with respect to the control technology that create a special advantage for the pressure-equilibrated hollow piston according to the invention go hand in hand with the direct action of the camshaft alternating torques via a central valve instead of via a damping path.
The bush of a central valve can be designed in a particularly advantageous way with a thread for screwing the rotor to the camshaft, so that a so-called central screw is formed.
The supply pressure, however, need not be introduced into the bush axially on the front side. It is also possible to provide the supply port radially, so that the supply pressure is also radially introduced into the hydraulic valve. The supply pressure, however, need not be introduced into the sleeve on the front side. It is also possible to introduce the pressure into the bush via a cross bore, which then leads into the inside space of the sleeve. In this way, the introduction can be made into the sleeve in its front-side opening or, however, in an opening in said wall of the sleeve.
According to the invention, the sleeve must be fixed relative to the bush. This means that the sleeve is solidly supported relative to the bush. In this case, the support of the pressure-relieving sleeve is preferably provided only in the axial direction. In contrast, the sleeve has a radial play in an advantageous configuration, for which reason the good functioning of the hydraulic piston is assured. In order to assure a tight sealing between the bush and the sleeve despite a large radial play, hydraulic fluid from the supply port is prevented from getting outside past the bush, in a particularly advantageous manner, by providing a sealing ring, which compensates for the radial play, in the region of this radial play.
If, in the normal operation of the hydraulic valve, the supply pressure is applied continually at the bottom of the sleeve, then a support exclusively in this direction may also be sufficient, since a displacement of the sleeve in the direction pointing out from the actuator is then prevented by the supply pressure. This applies even more so, if the sleeve is fitted in the bush so that friction forces also act between the sleeve and the bush. These friction forces are consequently to be kept small by means of appropriate material pairings, tolerances, component surfaces and structural measures.
The hollow piston is completely pressure-equilibrated in a particularly advantageous manner. It is also possible, however, to design the hollow piston with slightly varying outer diameter. In this case, unfortunately, there is little controllability. In return, however, assembly is simplified, since the hollow piston is preferably configured in such a way that its region that is to be introduced first has a smaller diameter than its region that is subsequently to be introduced. The probability of damage to the working surfaces/sealing surfaces during assembly is reduced, particularly in the case of manual assembly.
Other example embodiments of the present invention discussed below have particularly advantageous configurations, which equilibrate tolerances caused in the manufacture via a radial or an axial play, so that jamming of the hollow piston cannot occur.
Additional advantages of the invention are derived from the description and the drawing.
The present invention will hereinafter be described in conjunction with the appended drawing figures, wherein like reference numerals denote like elements, and
The ensuing detailed description provides exemplary embodiments only, and is not intended to limit the scope, applicability, or configuration of the invention. Rather, the ensuing detailed description of the exemplary embodiments will provide those skilled in the art with an enabling description for implementing an embodiment of the invention. It should be understood that various changes may be made in the function and arrangement of elements without departing from the spirit and scope of the invention as set forth in the appended claims. The angular position at the camshaft is changed with an oscillating-motor camshaft adjuster 14 according to
Stator 1 comprises a cylindrical stator base 3, on the inner side of which webs 4 protrude radially toward the inside at equal distances. Intermediate spaces 5 into which pressure medium is introduced via a hydraulic valve 12, which is shown in further detail in
Webs 4 are applied tightly by their front sides to the outer jacket surface of rotor hub 7. Vanes 6 in turn are applied tightly by their front sides to the cylindrical inner wall of stator base 3.
Rotor 8 is connected in a way that is torsionally rigid relative to the camshaft, which is not shown in further detail. In order to change the angular position between the camshaft and the crankshaft, rotor 8 is rotated relative to stator 1. For this purpose, depending on the desired direction of rotation each time, the pressure medium in either pressure chambers 9 or pressure chambers 10 is pressurized, while the other pressure chambers 10 or 9 are relieved of pressure to the tank. In order to pivot rotor 8 relative to stator 1 in a counterclockwise direction in the position shown, radial hub bores 11 in rotor hub 7 are pressurized by hydraulic valve 12. In order to pivot rotor 8, in contrast, in the clockwise direction, additional radial hub bores 13 in rotor hub 7 are pressurized by hydraulic valve 12. These additional radial hub bores 13 are arranged offset axially and circumferentially to the first-named radial hub bores 11. Hydraulic valve 12 is inserted as a so-called central valve into rotor hub 7 and screwed with the camshaft lying behind it.
Rotor 8 is pre-stressed against stator 1 in a torsionally elastic manner by means of a flat spiral spring acting as a compensation spring in a way that is not shown in the drawing.
Another port A1, which is formed by a cross bore 21 in bush 52 and which is assigned for the utilization of camshaft alternating torques, leads into the first annular groove assigned to the first working port A.
In addition, bush 52 has another two radial tank ports T1, T2 and an axial tank port T3. The first two radial tank ports T1, T2 are disposed axially adjacent to one another next to the two working ports A, B. In this case, the sequence of radial ports from the internal combustion engine to an actuator 43 is T1-T2-A-A1-B, successively. The axial or third tank port T3, in contrast, leads out from hydraulic valve 12 at a screw head 49 of bush 52, which is designed in screw shape.
The first radial tank port T1 in this case does not serve for the discharge of oil from the respective pressure chambers 9 or 10 to be relieved of pressure. Instead, this first tank port T1 serves for volume equilibration or for venting.
Bush 52 terminates on the engine side with an outer thread 53, which is screwed into an inner thread of the camshaft, which is not shown in further detail, and clamps rotor 8 against the camshaft in a frictionally engaged, torsionally rigid manner. For this purpose, rotor hub 7, on the one hand, is applied to the front-side end of the camshaft via a thin friction disk, and, on the other hand, to screw head 49 of bush 52. Such a friction disk, but with oil guides, is, for example, the subject of DE 10 2009 050 779.5.
A hollow piston 54 can be displaced inside bush 52. For this purpose, a tappet 48 of an electromagnetic linear actuator 43, which is shown only in a rudimentary manner in
The supply port P coming from an oil pump of the internal combustion engine, for example, via the camshaft, is axial. A cup-shaped, closed sleeve 55 is inserted into the hollow piston 54 which is hollowed out in the form of a blind hole 56. Its sleeve bottom 50 prevents the pressure of supply port P from acting on blind hole base 57 and thus force-loading hollow piston 54 in addition to a screw-type pressure spring 58. Because of this, electromagnetic linear-acting actuator 43 disengaging tappet 48 does not introduce a force against the varying pressure of supply port P. The control or regulating performance of the central valve is thus very good. Sleeve bottom 50 is supported for this purpose at bush 52 via a wall 23 of sleeve 55. Consequently, the entire sleeve 55 is secured in the bush. Sleeve 55 has a radially outwardly projecting collar 25 on its side facing away from tappet 48, and this collar is applied to a shoulder 24 of bush 52 on its side 26 facing tappet 48. An axial locking ring 29, which is inserted in an inner annular groove of bush 52, is applied onto the other side 27 of bush 52. In this way, sleeve 55 is secured in both directions against an axial displacement along a central axis 22.
Sleeve 55 is provided with at least one through-opening 59. The through-openings may comprise lengthwise slots as shown in
In the first valve position of hollow piston 54 relative to bush 52, which is shown in
As a consequence of its alternating torques, as soon as the camshaft attempts to rotate in the direction to be adjusted, the pressure in pressure chambers 10 and cross bore 21 increases sharply and abruptly. As soon as this pressure is increased far enough above the pressure in annular space 62, the losses at the first working port A and the pre-stressed additional non-return valve 33 are overcome, annular space 62 provides sufficient flow volume to pressure chambers 9 “aspirating” the hydraulic fluid via the second working port B for a rapid adjustment, since sufficient flow volume could not be provided by the oil pump alone. This relationship is also explained in more detail in DE 10 2006 012 775 A1.
Said annular groove 16 on hollow piston 54 is bounded on both sides by a guide web 36 or 30 in each case. Together with another guide web 28, the second guide web 30 forms another annular groove 47, which forms the radial inner boundary of annular space 62. Thus, in the direction pointing from the internal combustion engine to the linear actuator 43, hollow piston 54 has three axially successive guide webs 36, 30, 28, with which hollow piston 54 is guided inside bush 52. The first guide web 36 is thus disposed on the engine side relative to the center or second guide web 30. In contrast, the third guide web 28 is disposed on the actuator side relative to the center or second guide web 30. The cross bore 60 in hollow piston 54 coming from the supply port P is axially disposed between the second guide web 30 and the third guide web 28. The function of the last-named two guide webs 30, 28 is the following:
If the linear actuator 43 is maximally powered up, then hollow piston 54 is shifted against the force of screw-type pressure spring 58 into its end position, which is also the second valve position. In this case, the third guide web 28 closes gap 38 and releases a gap 37 corresponding to
A utilization of the camshaft alternating torques is not provided here, in contrast to the valve position according to
The hydraulic fluid flows from the tank ports T1, T2, T3 into the control drive box. In particular, if this control drive is designed with a chain, the hydraulic fluid equally lubricates the control drive. Wet belt drives are also known.
At the end on the actuator side, bush 52 has an annular groove on the inside, in which an axial locking ring 40 is placed. This axial locking ring 40 serves as the axial stop for the valve position according to
An axial locking element 42, which extends radially toward the outside annularly from hollow piston 54, is provided between the second guide web 30 and the third guide web 28. Here, this axial locking element 42, on the one hand, secures the third guide web 28 and, on the other hand, the non-return valve 61 axially. This prevents the non-return valve 61 from being displaced to such an extent that it no longer sufficiently covers cross bore 60.
Annular groove 16 is sealed via a sealing gap 45 relative to a front side 44 of hollow piston 54 pointing to supply port P. Hydraulic valve 12 would function in fact basically also without the region of hollow piston 54 that, after the center guide web 30, extends up to the first guide web 36. In this case, the screw-type pressure spring 58 would be applied axially at guide web 30. However, the forming of hollow piston 54 with the first guide web 36, which is shown in
Screw-type pressure spring 58 is disposed radially outside sleeve 55 in an annular space 64, which leads to a tank port T1 via an opening 63 in bush 52.
On its end 67 facing away from the actuator, sleeve 155 has an annular groove 69, in which a sealing ring 68 designed as an O-ring is taken up. On this end 67, the outer diameter of sleeve 155 has a large radial play relative to the associated uptake bore 70 of bush 152. This large play is equilibrated by sealing ring 68, so that despite this play, no hydraulic leakage of oil occurs between:
In this embodiment, as also in the previous embodiment, a sealing ring 68 can find use for the equilibration of tolerances caused by manufacturing. The annular groove 69 for taking up sealing ring 68 can be provided both in sleeve 55 or 155 as well as in bush 52 or 152.
The non-return valves may be designed with or without overlap. An alternative configuration of a band-shaped non-return valve is known from EP 1 703 184 B1. Instead of the asymmetrical distribution of through-openings to be closed by the non-return valve, which is claimed in this European Patent Specification, it is also possible to dispense with an overlap and to provide an anti-rotating element on the non-return valve.
The described embodiments only involve exemplary configurations. A combination of the described features for the different embodiments is also possible. Additional features for the device parts belonging to the invention, particularly those which have not been described, can be derived from the geometries of the device parts shown in the drawings.
| Number | Date | Country | Kind |
|---|---|---|---|
| 10 2010 019 005 | May 2010 | DE | national |
