EP2905434A1 - Oscillating-camshaft phaser having a hydraulic valve - Google Patents

Abstract

A swing camshaft phaser ensures that software can only use camshaft alternation torques under conditions where torque is adequate and / or it is important to reduce flow consumption. If there is a two-stage stroke and the low-lift camshaft alternating torque is not adequate, the software can position the piston to utilize part of the camshaft alternation torque while also channeling oil into the tank for faster adjustment.

Description

background

The invention relates to a Schwenkmotornockenwellenversteller with a hydraulic valve having two working ports.

The DE 10 2006 012 733 B4 and the DE 10 2006 012 775 B4 already relate to a Schwenkmotornockenwellenversteller with a hydraulic valve having two working ports. These two working ports each have axially adjacent to each other a standard opening and an opening for utilizing pressure peaks due to camshaft alternating torques. To adjust the camshaft, a hydraulic pressure can be conducted from a supply connection to the work connection to be loaded, while the work connection to be relieved is guided to a tank connection. The hydraulic valve is designed as a multi-port multi-position valve in cartridge design. In the carrier or central pin check valves are used on the inside, which are designed as band-shaped rings. By means of these check valves camshaft alternating torques are used to adjust the camshaft adjuster faster or with a relatively low oil pressure can. To do this, the check valves open to utilize pressure spikes due to camshaft alternating torques and obscure the backflow preventive ports into the unloaded port.

short version

It is an object of one embodiment of the present invention to provide a swivel camshaft phaser that is controlled in a manner that is easily controlled by electronic control means.

Briefly, an embodiment of the present invention provides a swing camshaft phaser that ensures that software can utilize camshaft cycle torques only under conditions where the torque is adequate and / or it is important to reduce flow consumption. If there is a two-stage stroke, and the low-lift camshaft alternating torque is not adequate, the software can position the piston to utilize part of the camshaft alternation torque while passing oil into the tank for faster adjustment.

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

Brief description of the drawings

• Figur 1 eine beispielhafte Ausführungsform eines Schaltschemas eines proportional ansteuerbaren Hydraulikventils, das in fünf Hauptstellungen betätigt werden kann;
• Figur 2 eine perspektivische Ansicht einer Kolbenkomponente des Hydraulikventils;
• Figur 3 eine vergrößerte Querschnittsansicht eines der Stege des Kolbens; und
• Figuren 4 - 10 eine beispielhafte konstruktive Umsetzung des Hydraulikventils nach Figur 1 in verschiedenen Stellungen.

The present invention will be described below in conjunction with the appended drawing figures in which like reference numerals designate like elements; show it:

• FIG. 1 an exemplary embodiment of a circuit diagram of a proportionally controllable hydraulic valve that can be operated in five major positions;
• FIG. 2 a perspective view of a piston component of the hydraulic valve;
• FIG. 3 an enlarged cross-sectional view of one of the webs of the piston; and
• FIGS. 4 to 10 an exemplary constructive implementation of the hydraulic valve after FIG. 1 in different positions.

Detailed Description of an Exemplary Embodiment

The following detailed description provides only exemplary embodiments and is not intended to limit the scope, applicability, or configuration of the invention. Instead, by the following detailed description of the exemplary embodiments, those skilled in the art will be given a description that enables them to practice an embodiment of the invention. It will be understood that various changes in the function and arrangement of elements may be made without departing from the spirit and scope of the invention as set forth in the appended claims.

FIG. 1 shows in a circuit diagram a by means of an electromagnet 17 against a spring force of a spring 21 operable hydraulic valve 3 according to an exemplary embodiment of the present invention, which is proportionally controlled. With this hydraulic valve 3 a Schwenkmotornockenwellenversteller 4 is pivotable. With such a Schwenkmotornockenwellenversteller 4, the angular position between the crankshaft and the camshaft can be changed during the operation of an internal combustion engine. By By turning the camshaft, the opening and closing times of the gas exchange valves are shifted so that the internal combustion engine performs optimally at the respective load and speed. The Schwenkmotornockenwellenversteller 4 allows a continuous adjustment of the camshaft relative to the crankshaft.

From the hydraulic valve 3, a first working port A and a second working port B go to the Schwenkmotornockenwellenversteller 4 from. The hydraulic valve 3 has four ports and five main switch positions and can thus also be referred to as a 4/5-way valve with a locking middle position 7. The valve has in principle seven states, but the switch positions 7, 7a and 7b are used to maintain the relative position of the rotor to the stator, with the switch positions 7a and 7b leaving oil in ports B and A, respectively, as required to compensate for system leakage is. Although the oil guide changes to shift positions, the flow opening of the valve is variable by incremental positioning in a shift state.

In order to pivot the Schwenkmotornockenwellenversteller 4 in the first direction of rotation 1, the hydraulic valve 3 is in one of the two switching positions 16 or 19, which are represented by the two boxes to the right of the locking center position 7. In the drawing figure 1, the hydraulic valve 3 is moved in the switching position 19 when the hydraulic valve 3 undergoes the full stroke movement by the actuator. In this case, pressure chambers 6 associated with this direction of rotation 1 are pressurized by the first working port A (which comes from the supply port P).

By contrast, in the switching position 16 or 19 the second working port B associated pressure chambers 5 are relieved. The second working port B is performed in switch position 19 via a tank connection T to a tank 20. In the intermediate positions 7b and 16 between the locking middle position 7 to the switching position 19, the pressure chambers 6 are acted upon by the first working port A coming from the supply port P pressure, but the second working port B is blocked against the tank port T.

Conversely, analogous applies. That is, to pivot the Schwenkmotornockenwellenversteller 4 in the second direction of rotation 2, the hydraulic valve 3 is in one of the two switching positions 18 or 15, which are represented by the two boxes to the left of the locking center position 7. In drawing figure 1, the hydraulic valve 3 is fully extended by the spring 21 in switch position 18 in box. In this way, the pressure chambers 5 associated with this direction of rotation 2 are pressurized by the second working port B (with a pressure coming from the supply port P).

In the switching positions 18 or 15, however, the first working port A associated pressure chambers 6 are relieved. In switching position 18, the first working port A is led to the tank 20 via the tank connection T. In the intermediate positions 15 and 7a between the locking middle position 7 and to switch position 18, the pressure chambers 5 are acted upon by the second working port B with a pressure coming from the supply port P, but the first working port A is blocked against the tank port T.

In the blocking middle position 7 all four ports A, B, P, T are blocked. This switch position as well as the switch positions 7a and 7b (the adjacent switch positions) are used to hold the rotor in a constant position relative to the stator.

For this purpose, a connection of the supply connection P with the second working connection B is present in the switching position 7a, whereas the first working connection A is blocked against the tank connection T. In switching position 7a prevents interaction between the inner web of the piston and the web of the cartridge or the central valve pin, that the first working port A opposite to the supply port P is opened. Therefore, in the switching position 7a, it is prevented that the first working port A is opened to the tank port T as well as to the supply port P.

In switching position 7b is a connection of the supply port P with the first working port A, while the second working port B is blocked against the tank port T. In switching position 7b prevents interaction between the inner web of the piston with the web of the cartridge or the central valve pin, that the second working port B is open to the supply port P. Therefore, in the switching position 7b, it is prevented that the second working port B is opened to both the tank port T and the supply port P. The switch positions 7a and 7b offer the advantage that the adjuster remains completely filled with oil at lower pump pressures. By blocking a working port against the supply port P, the supply port P can better fill the other working port.

In the two outermost switching positions 18 and 19 of the hydraulic valve 3, the adjustment of the camshaft is accomplished by loading one side of the vanes by utilizing recirculated oil available as a result of camshaft alternating torques in conjunction with oil supplied from the supply port P. Pressure is relieved from the other side of the wings by returning oil to the loaded wings while simultaneously passing oil into the tank. For this purpose, in the outermost switching position 18, a hydraulic fluid volume flow coming from a check valve RSV-A assigned to the first working port A is made available to the supply port P and B. Furthermore, in the switching position 18, an additional A port, which does not contain a check valve, can be discharged via the tank connection T to the tank 20. In switching position 19, on the other hand, a hydraulic fluid volume flow coming from a check valve RSV-B assigned to the first working port A is made available to the supply port P and A. Furthermore, in the switching position 19, an additional B connection, which does not contain a check valve, can be discharged via the tank connection T to the tank 20.

Also, in shift positions 15 and 16 of the hydraulic valve 3, the adjustment of the camshaft is accomplished by loading one side of the vanes by utilizing recirculated oil available as a result of camshaft alternating torques, in conjunction with oil supplied from the supply port P. Unlike the switch positions 18 and 19, pressure from the other side of the vanes is relieved only by returning oil to the loaded vanes. For this purpose, one of the first working port A is assigned in shift position 15 Check valve RSV-A coming hydraulic fluid flow provided to the supply port P and B. In switching position 16, on the other hand, a hydraulic fluid volume flow coming from a check valve RSV-B assigned to the second working port B is made available to the supply port P and A. In switching positions 15 and 16, however, no connection of any connection with the tank 20 is made.

In switching positions 15, 16, 18 and 19, this additional volume flow from the working port A or B to be relieved is fed into the volumetric flow coming from an oil pump 12 at the supply port P. There is a connection from the supply port P via a pump check valve RSV-P to the oil pump 12, which applies the pressure for the adjustment assistance of the Schwenkmotornockenwellenverstellers 4. This pump check valve RSV-P blocks the pressures in the hydraulic valve 3, so that peak pressures coming from the work connection A or B to be relieved can be made available to a larger proportion of the adjustment assistance than is the case with an open oil pump line 14a, 14b would.

FIG. 4 to FIG. 10 show exemplary structural embodiments of the hydraulic valve 3 in the seven switching positions 18, 15, 7a, 7, 7b, 16, 19 according to Fig. 1 ,

FIG. 4 shows the hydraulic valve 3 in the first switching position 18, in which the electromagnet 17 according to Fig. 1 a piston 22 of the hydraulic valve 3 does not move. The stroke of the piston 22 is thus at zero. The piston 22 is within a central bolt 27 against the force of as Helical compression spring running spring 21 slidably. The electromagnet 17 facing the end 50 of the piston 22 is closed to produce a contact surface for an actuating plunger of the electromagnet 17, whereas the other end 52 of the piston 22 for receiving one end of the spring 21 is open. The piston 22 is held in the central bolt 27 via a retaining ring 54. The piston 22 has at its two ends outer webs 23, 24, which are guided relative to the central pin 27. The two outer webs 23, 24 have partially over the webs flat flow surfaces 29, 30, so that along these flow surfaces 29, 30 from the ends of the central bolt 27 access to the tank port T is present. In an alternative embodiment, it could well be provided that the piston 22 is hollow and that axial connection bores for flow to the tank connection T are contained.

Axially between the two outer webs 23, 24, two narrow ribs or webs 31, 32 are provided, which run around the piston 22. These circumferential ribs 31, 32 correspond to two annular webs 33, 34 extending radially inwardly from the central pin 27. In addition to these two annular webs 33, 34, two axially outer annular webs 35, 36 are provided. These four ring lands 33, 34, 35, 36 are formed by five inner ring grooves 37, 38, 39, 40, 41 are hollowed out of the central pin 27. In this five Innenringnuten 37, 38, 39, 40, 41 open five connecting holes 60, 62, 64, 66, 68, which are drilled through the wall of the central pin 27. Depending on the flow requirements more than one hole per ring groove are possible.

These five connection bores 60, 62, 64, 66, 68 form from the sides of the electromagnet 17 axially along the Bolt 27 comprises: a standard opening B associated with the second working port B, an opening B1 associated with the second working port B for utilizing camshaft alternating torques, the supply port P, an opening A1 associated with the first working port A, and an opening A associated with the first working port A for the use of camshaft alternating torques.

At the two working ports A, B thus two openings A, A1 and B, B1 are provided. Of these, the axially inner openings A1, B1 are provided for a camshaft alternating torque utilization. In contrast to the axially outer openings A, B, which can be locked from the inside only with the outer webs 23, 24, the axially inner openings A1, B1 have the band-shaped check valves RSV-A, RSV-B. In each case one of the band-shaped check valves RSV-A and RSV-B is inserted into an inner ring groove 40 or 38 radially inside the axially inner openings A1 and B1 of the central pin 27. With the check valves RSV-A, RSV-B it is according to the in the DE 10 2006 012 733 B4 described method possible to provide a hydraulic pressure that increases due to camshaft alternating torques briefly above the level of hydraulic pressure in the hydraulic chambers to be loaded 6 and 5, in the region of the supply port P. From this supply port P, these hydraulic pressure peaks or this additional hydraulic fluid flow together with the hydraulic pressure applied to the supply port P by the oil pump 12 are then made available to the hydraulic chambers 6 and 5 to be loaded.

In addition, the band-shaped pump check valve RSV-P is still provided in an inner ring groove 39. This The pump check valve RSV-P has basically the same structure as the two check valves RSV-A, RSV-B. However, this pump check valve RSV-P may have a different response force.

In the switch position 18 according to Fig. 4 the two central ribs 31, 32 are axially spaced from the two annular ribs 33, 34, so that hydraulic fluid can pass through the gap therebetween. Likewise, hydraulic fluid can pass through the gap between the foremost outer web 23 and the corresponding annular web 35 on the central pin 27. On the other hand, the other outer bar 24 blocks the rearmost inner annular groove 41 or the standard opening A associated with the first working port A. For this purpose, the outer bar 24 and the rearmost annular bar 36 overlap over a large sealing length.

Thus, in this switching position 18 hydraulic fluid from the supply port P via the pump check valve RSV-P to the second working port B associated standard opening B arrive. In contrast, the other two check valves RSV-A and RSV-B block the openings A1 and B1 against pressures from the supply connection P and from the second working port B associated standard opening B. However, short-term peak pressures due to the camshaft alternating moments from the first working port A associated opening A1 through the check valve RSV-A. If the pressure related to the working port A is high due to a cam torque, it is greater than the pressure P. Then, the RSV-A check valve opens and directs oil from A while the P-type check valve (RSV-P) closes. In switching position 18, pressure from the first working port A is returned from A to B (via opening A1), and the first working port A also becomes (via standard opening A and flow area 30) relieved to the tank port T.

FIG. 5 shows the piston 22 at a stroke of 0.4 mm. Shift position 15 is very similar to shift position 18, except that the piston 22 has advanced to a position in which the first working port A by the interaction of web 24 with the bolt surface 98 against the tank port T is locked, whereby no connection from A to the flow surface 30 can be made.

Between the in FIG. 5 shown switch position 15 and the in FIG. 4 shown switching position 18, the first working port A is increasingly opened to the tank port T. This allows both hydraulic fluid return from the first working port A to the second working port B and guiding the first working port A to the tank (that is, hydraulic fluid flow from the first working port A to the tank port T).

FIG. 6 shows the piston 22 at a stroke of 1.1 mm. In this case, the hydraulic valve 3 is in the switching position 7a. Shift position 7a is greatly similar in that switching position 15, as the first working port A is blocked by the interaction of web 24 with surface 98 against the tank port T. In switching position 7a, however, the first working port A is blocked via the interaction of web 32 with web 34 against the second working port B. There is a connection from the supply connection P to the second working connection B.

Between the in FIG. 6 shown switching position 7a and the in FIG. 5 shown switching position 15 allowed the supply port P is increasingly gaining access to the second working port B, and hydraulic fluid flow from the first working port A is increasingly returned to the second working port B when the pressure of the first working port A increases by cam moment pulses over that of the second working port B and the supply port P.

FIG. 7 shows the piston 22 at a stroke of 1.7 mm. In this case, the hydraulic valve 3 is in the locking middle position 7. The supply port P is closed by the two ribs 31, 32. To cover the ribs 31, 32, the corresponding annular ridges 33, 34 in a correspondingly large extent. The two working ports A, B are blocked due to the interaction of web 24 with the surface 98 and the interaction of web 23 with the surface 99 and against the tank drain T.

Although the in FIG. 7 shown locking middle position 7 is practically the holding position, the piston between this switching position and either the in FIG. 6 shown switching position 7a or in FIG. 8 move shown switching position 7b to compensate for hydraulic fluid leakage.

FIG. 8 shows the piston 22 at a stroke of 2.3 mm. In this case, the hydraulic valve 3 is in the switching position 7b, and a second working port B is locked via the interaction of web 23 with the surface 99 against the tank port T. In switching position 7b of the second working port B is blocked by the interaction of web 31 with web 33 and against the first working port A. There is a connection from the supply port P to the first working port A.

FIG. 9 shows the piston 22 at a stroke of 3.0 mm. In this case, the hydraulic valve 3 is in the switching position 16, and the second working port B is blocked by the interaction of web 23 with the surface 99 against the tank port T. Furthermore, short-term peak pressures are transmitted from the second working port B associated opening B1 through its check valve RSV-B due to the camshaft alternating torques. In switching position 16, the first working port A is pressurized by the supply port P, and pressure from the second working port B is returned (via port B1) from B to A.

Between the in FIG. 8 shown switching position 7b and the in FIG. 9 shown switching position P allows access to the supply port P increasingly to the first working port A, and hydraulic fluid flow from the second working port B is increasingly returned to the first working port A, when the pressure of the second working port B by cam moment pulses over that of the first working port A and the supply port P addition increases ,

FIG. 10 shows the piston 22 at a stroke of 3.4 mm. In this switching position 19, the two middle ribs 31, 32 axially spaced from the two annular ridges 33, 34, so that hydraulic fluid can pass through the gaps between them. Likewise, hydraulic fluid can pass through the gap between the rearmost outer web 24 and the corresponding annular web 36. On the other hand, the other outer web 23 blocks the foremost inner annular groove 37 or the standard opening B associated with the second working port B. For this purpose, the outer web 23 and the foremost annular web 35 overlap over a large sealing length. This can be done in This switching position 19 hydraulic fluid from the supply port P via the pump check valve RSV-P to the first working port A associated standard opening A arrive. In contrast, the other two check valves RSV-A and RSV-B block the openings A1 and B1 against pressures from the supply port P. On the other hand, short-term peak pressures due to the camshaft alternating torques from the second working port B associated opening B1 through its check valve RSV-B. Thus, pressure from the second working port B (via port B1) is returned from B to A, and the second working port B is also relieved (via standard port B and flow surface 29) to the tank port T.

Between the in FIG. 9 shown switching position 16 and the in FIG. 10 shown switching position 19, the second working port B is increasingly opened to the tank port T. This allows both hydraulic fluid return from the second working port B to the first working port A and guiding the second working port B to the tank (that is, hydraulic fluid flow from the second working port B to the tank port T).

One of the major advantages of the system described herein is that by software control of the hydraulic valve, the duty cycle (or current) may be limited to allowing feedback only (switch positions 15 and 16) when there is adequate cam torque to achieve desired phase rates , It may also be limited to the switch positions 15 and 16 when insufficient current is present in the engine oil system and further loading is undesirable.

If a cam torque is not adequate, such as in the low lift mode of a two-stage lift system, the software allows the use of the shift positions 18 and 19 for phasing. High revolutions per minute also do not allow enough time to make good use of cam torque pulses, thus utilizing switch positions 18 and 19 can increase phase-up speeds at high revolutions per minute, if required. The amount of flow opening to the tank port T and the valve lift positions in which the shift positions 18 and 19 begin can be tailored to the application.

In the exemplary embodiment set forth, the standard opening A or B and the opening A1 or B1 are combined to initially utilize camshaft alternating moments outside of the central pin 27 to the working port A and B, respectively. In an alternative embodiment, it is also possible to combine the standard opening A or B and the opening A1 or B1 also within the central pin 27 in order to use the camshaft alternating moments.

In another alternative embodiment, ball check valves may be used instead of tape check valves. For example, it is also possible to use within the hydraulic valve ball check valves, such as the DE 10 2007 012 967 B4 shows. However, the ball check valves need not necessarily be installed in the central valve of a cartridge valve. It is also possible, for example, to use ball check valves in a rotor and to perform the piston as a central valve which is arranged coaxially and centrally within the rotor hub slidably.

Depending on the operating conditions of the valve may be provided in the flow direction in front of one or more or even all ports and filters that protect the running surfaces between the piston and the central valve.

The use of camshaft alternating torques does not have to be provided for both directions of rotation. It is also possible to dispense with one of the two axially outermost switching positions 18 or 19. As a result, the camshaft alternating torques can then be used directly for faster adjustment only for the one direction of rotation.

In an alternative embodiment, a use of the camshaft alternating torques can be provided for both directions of rotation, in which case, however, is dispensed with one of the two bypass check valves RSV-A, RSV-B.

Any combinations of switch positions are possible. For example, it is possible to dispense with one or more positions or states or add one or more additional switch positions or states.

On the hydraulic valve can also be provided a further switching position, wherein oil is metered through a self-centering center lock A and B, with one side is relieved to centering. The pin is relieved, allowing it to fall into the locking pin hole, locking the stage in the middle locking position. Center lock is used for example in the DE 10 2004 039 800 and DE 10 2009 022 869.1-13 shown.

FIG. 2 shows a preferred piston 22 and is self-explanatory, especially in view of the above description. Preferably, the webs 31, 32 are provided in the form of a shark fin shape, as in FIG FIG. 3 showing an enlarged view of the web 32. Functionally, it is important to have a minimum stroke pressure of the webs 31, 32 in order to guide the supply port P either to the working port A or B. However, it is difficult to heat treat the very thin webs. Thus, a preferred piston provides ridges which, for example, have a thickness of only 0.3 mm at their base, but taper at least on one side 90, so that they actually engage with the ridges 33, 34 on the surface 92 the central pin 27 cooperates, only have a thickness in the range of 0.1 mm to 0.3 mm. As mentioned, presents FIG. 3 an enlarged view of the web 32 ready. An enlarged view of the other land 31 would look very similar, but would give an inverted image with the tapered surface 90 on the opposite side.

It should be pointed out that one or more tapered ridges (such as in shark fin shape) may be provided on piston 22 or on bolt 27 or both. In addition, it is possible that the web (the webs) only rejuvenated (rejuvenated) on one side or on both sides of the web.

Other advantages of providing thin webs are that it allows shorter piston strokes. In addition, this allows a better timing for the proportional control of the valve. This is due to the fact that thereby the stroke associated with the switch positions 7a to 7b can be shortened, allowing a faster transition from one direction to another.

With reference to the Figures 4-10 Now, some preferred degrees of overlap (which prevents fluid flow) and ports (allowing fluid flow) will now be described. Of course, other levels of overlap and apertures may be used while in no way departing from the scope of the present invention.

In FIG. 4 There is preferably an opening of 1.5 mm at the point P to B1, preferably an opening of 1.5 mm at the location of B1 to B, preferably an overlap of 3.0 mm at the point B to T, preferably one Opening of 1.1 mm at the point P to A1, preferably an overlap of 1.6 mm at the point A1 to A and preferably an opening of 0.4 mm at the point A to T.

In FIG. 5 there is preferably an opening of 1.1 mm at the point P to B1, preferably an opening of 1.1 mm at the point B1 to B, preferably an overlap of 2.6 mm at the point B to T, preferably an opening from 0.7 mm at the point P to A1, preferably an overlap of 1.5 mm at the point A1 to A and preferably an overlap of 0.0 mm at the point A to T.

In FIG. 6 There is preferably an opening of 0.4 mm at the point P to B1, preferably an opening of 0.4 mm at the location of B1 to B, preferably an overlap of 1.9 mm at the point B to T, preferably one Opening of 0.0 mm at the point P to A1, preferably an overlap of 0.8 mm at the Place A1 to A and preferably an overlap of 0.7 mm at location A to T.

In FIG. 7 there is preferably an overlap of 0.2 mm at the point P to B1, preferably an overlap of 0.2 mm at the point B1 to B, preferably an overlap of 1.3 mm at the point B to T, preferably an overlap from 0.2 mm at the point P to A1, preferably an overlap of 0.2 mm at the point A1 to A and preferably an overlap of 1.3 mm at the point A to T.

In FIG. 8 There is an opening of 0.0 mm at the point P to B1, preferably an overlap of 0.8 mm at the point B1 to B, preferably an overlap of 0.7 mm at the point B to T, preferably an opening of 0.4 mm at the point P to A1, preferably an opening of 0.4 mm at the point A1 to A and preferably an overlap of 1.9 mm at the point A to T.

In FIG. 9 there is preferably an opening of 0.7 mm at the point P to B1, preferably an overlap of 1.5 mm at the point B1 to B, preferably an opening of 0.0 mm at the point B to T, preferably an opening from 1.1 mm at the point P to A1, preferably an opening of 1.1 mm at the point A1 to A and preferably an overlap of 2.6 mm at the point A to T.

In FIG. 10 there is preferably an opening of 1.1 mm at the point P to B1, preferably an overlap of 1.6 mm at the point B1 to B, preferably an opening of 0.4 mm at the point B to T, preferably an opening from 1.5 mm at the point P to A1, preferably an opening of 1.5 mm at the Place A1 to A, and preferably overlap 3.0 mm at location A to T.

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.

While particular embodiments of the invention have been shown and described, it is contemplated that one skilled in the art can devise various modifications without departing from the spirit and scope of the present invention. For example, the check valves may be configured as ball or plate check valves.

Claims (15)

A swivel camshaft phaser (4) comprising a hydraulic valve (3) including two working ports (A, B), a supply port (P) and a tank port (T), the hydraulic valve (3) configured to prevent one of the working ports (A, B) to the tank port (T) is relieved, while the supply port (P) to the other working port (B, A) pressurized. A swing motor camshaft adjuster (4) according to claim 1, wherein each working port (A, B) has a standard opening (A, B) and an additional opening (A1, B1) for utilizing pressure spikes due to camshaft alternating torques. Pivot camshaft adjuster (4) according to claim 1 or 2, wherein each standard opening (A, B) is configured for selective unloading to the tank connection (T). A swing motor camshaft adjuster (4) according to any one of the preceding claims 1 to 3, wherein the hydraulic valve (3) is configured to allow recirculation from the one working port (A, B) to the other working port (B, A) while preventing that the one working port (A, B) is relieved to the tank port (T), and while the supply port (P) to the other working port (B, A) with pressure. A swivel camshaft adjuster (4) according to any of the preceding claims 1 to 3, wherein the hydraulic valve (3) is configured to return from the one working port (A, B) to prevent the other working port (B, A) while preventing one working port (A, B) from being relieved to the tank port (T) and while the supply port (P) is pressurizing the other port. A rotary camshaft phaser (4) according to any one of the preceding claims 1 to 3, wherein the hydraulic valve (3) is configured to allow in a first state a return from the one working port (A, B) to the other working port (B, A). while preventing the one working port (A, B) from being relieved to the tank port (T) and while the supply port (P) is pressurizing the other port (B, A), and the hydraulic valve (3) is configured thereto to prevent in a second state, a return from the one working port (A, B) to the other working port (B, A), while preventing the one working port (A, B) is relieved to the tank port (T), and while the supply port (P) is pressurizing the other working port (B, A). Pivot camshaft adjuster (4) according to one of the preceding claims 1 to 6, wherein the hydraulic valve (3) comprises a piston (22) and / or a bolt (27) with at least one tapered web (31, 32). Pivot camshaft adjuster (4) according to one of the preceding claims 1 to 7, wherein the hydraulic valve (3) comprises a piston (22) and / or a bolt (27) having at least one web (31, 32) with a shark fin shape on at least one side (90 ) of the web (31, 32). Pivot camshaft phaser (4) according to one of the preceding claims, wherein each working port (A, B) has a standard opening (A, B) and an additional opening (A1, B1) for utilizing pressure peaks due to camshaft alternating torques and check valves (RSV-A, RSV -B) are provided at the additional opening (A1, B1) of each working port (A, B). Pivot camshaft phaser (4) according to claim 9, wherein a check valve (RSV-P) is provided at the supply port (P). A swing motor camshaft adjuster (4) according to any one of the preceding claims 3 to 10, wherein the hydraulic valve (3) is configured to move a flow from the supply port (P) to the one working port (A) as the hydraulic valve (3) moves from one position to another position , B) increasingly allow and increasingly allow a return from the other working port (B, A) in the one working port (A, B). A swing motor camshaft adjuster (4) according to any of the preceding claims 3 to 11, wherein the hydraulic valve is configured to move a flow from the supply port (P) to the one working port (A, B) upon movement of the hydraulic valve (3) from a first position to a second position ) and to allow a return from the other working port (B, A) in the one working port (A, B) increasingly when the pressure of the one working port (A, B) by cam torque pulses on the pressure of the second working port (B , A) and the supply terminal (P) increases. A rotary camshaft phaser (4) according to any of the preceding claims 1 to 12, wherein the hydraulic valve (3) is configured to move a flow from the one working port (A, B) to the tank port as the hydraulic valve (3) moves from one position to another position (T) to allow increasingly and a return of one work connection (A, B) to the other work port (B, A) increasingly allow. A rotary camshaft phaser (4) according to any of the preceding claims 1 to 13, wherein the two working ports (A, B) comprise a first working port (A) and a second working port (B), the hydraulic valve (3) being configured to provide: a first state during which the supply port (P) pressurizes the first working port (A), while the second working port (B) allows recirculation into the first working port (A) and at the same time relieves the tank port (T); a second state during which the supply port (P) pressurizes the first working port (A) while the second working port (B) returns to the first working port (A) but is not relieved to the tank port (T); and a third state during which the supply port (P) pressurizes the first working port (A) while preventing the second working port (B) from returning to the first working port (A) and is prevented from being connected to the tank port (A) T) is relieved. Schwenkmotornockenwellenversteller (4) according to one of the preceding claims 1 to 13, wherein the both working ports (A, B) comprise a first working port (A) and a second working port (B), the hydraulic valve (3) being configured to provide seven states: a first state during which the supply port (P) pressurizes the first working port (A), while the second working port (B) allows recirculation into the first working port (A) and at the same time relieves the tank port (T); a second state during which the supply port (P) pressurizes the first working port (A) while the second working port (B) returns to the first working port (A) but is not relieved to the tank port (T); a third state during which the supply port (P) pressurizes the first working port (A) while preventing the second working port (B) from returning to the first working port (A) and being prevented from being connected to the tank port (A) T) is relieved; a fourth state, during which it is prevented that neither the first nor the second working port (A, B) are pressurized by the supply port (P), and prevented from being relieved to the tank port (T); a fifth state during which the supply port (P) pressurizes the second working port (B) while preventing the first working port (A) from returning to the second working port (B) and being prevented from being connected to the tank port (B). T) is relieved; a sixth state, during which the supply port (P) pressurizes the second working port (B), while the first working port (A) connects to the second working port (B) returns, but is not relieved to the tank port (T); and a seventh state during which the supply port (P) pressurizes the second working port (B) while the first working port (A) permits recirculation to the first working port (A) and at the same time relieves the tank port (T).

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