Pressure regulating valve in gate design with improved damping characteristics

Abstract

Pressure regulating valve (18) with a control piston (34) and a spool sleeve (28), wherein the pressure regulating valve (18) has a supply port (16), a working port (22) and a return port (20), wherein the control piston (34) is guided with a first section (44) sealingly in a guide bore (32) of the spool sleeve (28), wherein the first section (44) of the control piston (34), the guide bore (32) and a sealing cap (36) define a first compensation chamber (64), the control piston (34) is guided with a third section (48) sealingly in a guide bore (32) of the spool sleeve (28), the third section (48) of the control piston (34), the guide bore (32) and a bearing bushing (37) define a second compensation chamber (58), the first compensation chamber (64) and the second compensation chamber (58) are hydraulically connected to each other. connection and at least one damping element (60,94,) between first compensation space (64) and second compensation space (58)96, 98) is arranged, characterized in that the slide sleeve (28) in the area of ​​the supply connection (16) is tightly closed by a sealing plug (36) pressed onto the end face.

Description

State of the art

[0001] The invention relates to a pressure control valve, in particular for an automatic transmission in a motor vehicle, according to the preamble of claim 1.

[0002] Modern automatic car transmissions use hydraulically operated clutches to change gears. To ensure smooth gear shifting and imperceptible to the driver, the hydraulic pressure at the clutches must be adjusted with the utmost precision according to predefined pressure ramps. Electromagnetically operated pressure control valves are used for this purpose. These can be designed as either seat or slide valves. Both designs typically have three hydraulic connections: supply, control pressure, and return.

[0003] In spool valves, the use of damping orifices or damping throttles has proven particularly advantageous for providing the required damping. This utilizes the effect that hydraulic oil is displaced when the control piston moves. When this hydraulic oil flows through a damping orifice or damping throttle, a pressure drop occurs, which dampens the movement of the control piston depending on the speed. An orifice or throttle The shorter cross-section compared to the hydraulically effective cross-section is particularly advantageous because the damping effect is virtually independent of the oil temperature or viscosity of the hydraulic oil. This ensures sufficiently consistent damping behavior over a wide range of operating temperatures.

[0004] There are numerous examples of slide valves in which such damping orifices and / or damping throttles are used.

[0005] From DE 198 29 549 B4 a slide valve with an internal damping orifice and an external damping orifice in the working connection of the pressure control valve is known.

[0006] In the slide valve known from US 2009 / 0 000 677 A1, the damping orifice is arranged in the return connection.

[0007] From US 2008 / 0 072 978 A1 a slide valve with a damping orifice at the tank connection and a damping throttle in the working connection is known.

[0008] From DE 103 25 070 A1 a pressure control valve with two damping ends in the working pressure connection on the front side of the control piston is known.

[0009] A disadvantage of the prior art solutions, in which the damping orifice is located in the working connection, is that this requires additional sealing points on the pressure control valve. If the working pressure is routed to the damping orifices in addition to the direct consumer connection, appropriate sealing must be provided there as well, or additional leakage must be expected.

[0010] Furthermore, whenever the damping orifices are used in the pressurized area of ​​the pressure control valve, a pressure force is exerted on the piston. This force is undesirable and can be reduced by using only an annular area in the piston rather than the entire piston surface of the control piston as the displacement volume. However, this also reduces the displaced volume and thus the damping effect.

[0011] If an orifice damper is implemented in the pressure-free tank connection, the problem arises that the oil filling of this pressure-free space cannot be reliably guaranteed.

[0012] US Pat. No. 6,289,921 B1 discloses a pressure control valve for controlling a camshaft in a motor vehicle, comprising a slide sleeve and a control piston. The pressure control valve has a supply port, a working port, and a return port. A first portion of the control piston is sealingly guided in a guide bore of the slide sleeve. The first portion of the control piston, the guide bore, and a housing define a first hydraulic chamber. A third portion of the control piston is sealingly guided in a guide bore of the slide sleeve. The third portion of the control piston, the guide bore, and a bearing bush define a second hydraulic chamber. The first hydraulic chamber and the second hydraulic chamber are hydraulically connected to one another, and a damping element is arranged between the first hydraulic chamber and the second hydraulic chamber.

[0013] From MX PA00009798 A and US 2001 / 0 025 662 A1, further pressure control valves are known which have a control piston provided with various bores. Advantages of the invention

[0014] In the pressure control valve according to the invention, the control piston is designed as a stepped piston in which the two unpressurized end face volumes, hereinafter referred to as the first displacement chamber and the second displacement chamber, are used to dampen the movement of the control piston in a speed-dependent manner with each movement of the control piston. Both displacement chambers are tightly closed and require no additional seals. The hydraulic oil displaced due to an adjusting movement of the control piston flows through the control piston and the damping means according to the invention. The damping means are mounted in the immediate vicinity of the tank outlet so that the hydraulic oil flowing from the valve inlet via the working closure to the tank floods the displacement chambers according to the invention and the damping means, thus preventing the damping volumes from running dry.

[0015] Another advantage of the pressure control valve according to the invention is that it is technically very easy to implement and requires no additional components. The pressure control valve according to the invention is characterized by low overall costs.

[0016] To ensure that the supply pressure does not act on the front face of the control piston and negatively impact its pressure control function, the sliding sleeve is sealed pressure-tight in the area of ​​the supply connection by a sealing plug pressed onto the front face. The sealing plug can be made of plastic or as a stamped and bent sheet metal part. Both variants can be attached captively and tightly to the sliding sleeve using a simple press connection.

[0017] In an advantageous embodiment of the invention, it is possible to form a rib on the control piston which separates a third compensation chamber from the second compensation chamber, wherein these compensation chambers are also hydraulically connected to one another by a damping element.

[0018] The damping element can be designed as a throttle bore between the compensation chambers or as a flattening, groove or notch in the rib of the control piston or directly in the control piston.

[0019] The damping element can also be designed as a suitable clearance fit between the rib of the damping piston and the control piston and the guide bore and thus as a circular annular gap.

[0020] Of course, combinations of these damping elements are also possible. All of the damping elements described have one thing in common: they are easy to manufacture and require no additional components. Even the rib between the second and third compensation chambers is formed integrally on the control piston, eliminating any additional manufacturing and assembly effort.

[0021] Furthermore, the damping elements are preferably arranged at the end of the control piston that is located in the immediate vicinity of the return connection, so that any internal leaks that occur fill the compensation chambers with hydraulic oil without any further action. This ensures a reliable damping effect over the entire service life of the pressure control valve according to the invention.

[0022] The pressure control function is achieved by the valve having a control piston which, in an open end position, hydraulically connects a working port to the supply port and hydraulically separates it from a return port. In the non-activated state, i.e. when the coil of an electromagnetic actuator is de-energized, its armature moves in the opening direction via the spiral spring installed on the magnet side and, via a push rod, moves the control piston towards an open end position of the pressure control valve. As a result, the first section of the control piston releases the inlet pressure opening and the pressure medium flows into an annular space defined by the slide sleeve and control piston. Since the third section of the control piston simultaneously separates the return opening from the annular space, the pressure prevailing at the supply port is also present at the working port.

[0023] When the coil is energized, the electromagnetic force counteracts the restoring force of the coil spring on the solenoid side, which deflected the control piston toward the front end. The control piston is moved back to its closed end position by the return spring mounted on the piston side. In the closed end position, the control piston closes the supply pressure port and simultaneously opens the return port. Because the tank pressure in the return port is lower than the working pressure in the annular space, pressure fluid flows through the return port to the return port.

[0024] The same applies to an electromagnetic actuator (not described in detail here), which operates without a coil spring on the magnet side and whose electromagnet acts in the opposite direction. This closes the supply connection when de-energized and closes the return opening when energized.

[0025] In the pressure control valve according to the invention, the force acting on the control piston in the opposite direction of opening depends on the pressure currently prevailing at the control pressure opening. If the pressure at the control pressure opening drops, the force acting on the control piston in the opposite direction of opening also decreases, and the control piston is moved in the opening direction. If, on the other hand, the pressure at the control pressure opening increases, the force acting on the control piston in the opposite direction of opening also increases, causing it to move in the opposite direction of opening. This self-regulating function of the control piston is achieved by the hydraulic area acting in the opening direction being different from the hydraulic area acting in the opposite direction of opening.

[0026] This difference between the hydraulic surfaces acting in the opposite and opening direction is achieved by the stepped guide bore in the slide sleeve, which has a smaller diameter in the first section of the control piston than in its third section.

[0027] All in all, the present invention provides a pressure control valve that provides a precise self-regulating function while at the same time being simple in design and correspondingly low in manufacturing costs.

[0028] It is also advantageous that the control piston has a first control edge, which, when the control piston is in an intermediate position, throttles the pressure medium flow from the supply port to the working port. Furthermore, the control piston has a second control edge, which, when the control piston is in an intermediate position between the open and closed end positions, throttles the pressure medium flow from the working port to the return port. This allows for a continuously variable pressure control characteristic of the valve.

[0029] It is particularly helpful if the control piston has at least one channel connecting the first compensation chamber near the supply connection with the second compensation chamber at the opposite end of the control piston. This ensures the same hydraulic pressure at the front ends of the control piston, ensuring the pressure control function.

[0030] In addition, it is proposed that the channel be designed as a combination of a longitudinal bore and a throttle bore. This allows the control piston for the pressure control valve according to the invention to be manufactured simply and cost-effectively. Drawings Fig. 1 a schematic representation of a hydraulic circuit in which a pressure control valve according to the invention is used, Fig. 2 a partial section through a pressure control valve according to the invention; Fig. 3 an enlarged view of the pressure control valve according to the invention is used and Fig. 4 a section along the line AA Fig. 3.

[0031] Automatic transmissions, such as those used in cars, are controlled by, among other things, a hydraulic circuit 10, which includes a pressureless hydraulic oil reservoir 12 and a hydraulic pump 14. An outlet of the hydraulic pump 14 forms a supply connection 16, to which a pressure control valve 18 is connected.

[0032] A return line leads from the pressure control valve 18 to a return port 20, which leads back to the hydraulic oil reservoir 12. Furthermore, the pressure control valve 18 is connected to a working port 22, to which the pressure to be regulated by the pressure control valve 18 is applied. Furthermore, the pressure control valve 18 has an electromagnetic actuating device 24.

[0033] In Fig. Figure 2 shows the structure of a pressure control valve 18 according to the invention. The pressure control valve 18 comprises a valve connection element 26, which is preferably made of plastic. The valve connection element 26 has a concentric recess (without reference symbol) into which a slide sleeve 28 is inserted such that it forms an annular inlet channel 30 with the valve connection element 26.

[0034] The slide sleeve 28 has a continuous and stepped guide bore 32, which serves to accommodate a control piston 34. On the front side, on the left side in Fig. 2, the slide sleeve 28 is pressure-tightly closed by a closure cap 36, which can be pressed or shrunk on, for example.

[0035] The opposite side, (in Fig. 2 on the right), the slide sleeve 28 is closed by a bearing bush 37. Arranged axially next to one another on the circumference of the slide sleeve 28 are three openings 38, 40, and 42. The first opening, hereinafter referred to as the inlet pressure opening 38, in the slide sleeve 28 opens the guide bore 32 to the inlet channel 30 and thus to the supply connection 16 when the control piston 34 is actuated accordingly.

[0036] In Fig. In Figure 2, the pressure control valve 18 is shown in the equilibrium position, so that there is no hydraulic connection between the inlet channel 30 and the guide bore 32. This equilibrium position is an intermediate position between the open and closed end positions.

[0037] The second transverse bore of the slide sleeve 28, hereinafter referred to as the control pressure opening 40, connects the guide bore 32 to the working connection 22. The third transverse bore, hereinafter referred to as the return opening 42, establishes a hydraulic connection between the guide bore 32 and the return connection 22.

[0038] The control piston 34 is divided in the axial direction into four adjacent sections 44, 46, 48 and 50. The Fig. 2 The leftmost and first section 44 has a first diameter D1. This first diameter D1 guides the control piston 34 in the guide bore 32.

[0039] The approximately centrally located second section 46, which adjoins the first section 44, has a second diameter D2 that is smaller than the first diameter D1 and smaller than the diameter of the guide bore 32 in this area. This creates an annular space 52.

[0040] The third section 48, which adjoins the second section 46, has a larger diameter than the first section 44 and is guided in a sealing but axially displaceable manner in the guide bore 32 of the slide sleeve 28. This limits the annular space 52 in the axial direction.

[0041] In the fourth and final section 50, the diameter of the control piston 34 is smaller than that of the guide bore 32. This creates a second compensation chamber 58, which is bounded in the axial direction by the bearing bush 37 and the push rod 72.

[0042] Due to this special shape of the control piston 34, the second section 46 of the control piston 34 and the slide sleeve 28 form an annular space 52 that communicates with the working port 22 via the control pressure opening 40. The edge of the first section 44 facing the second section 46 forms a first control edge 54, the function of which will be discussed in more detail later.

[0043] The edge of the third section 48 facing the second section 46 forms a second control edge 56. In the fourth section, the control piston 34 has a transverse bore. Adjoining the transverse bore is a longitudinal bore 62, which penetrates the control piston 34 along its entire length. As a result, the same pressure prevails in a first compensation chamber 64 and in the second compensation chamber 58.

[0044] In the first compensation chamber 64, which is designed here as a front-side recess of the control piston 34, there is a first spiral spring 66 which is supported on the one hand against the control piston 34 and on the other hand against the closure cap 36 which closes the slide sleeve 28 on the front side.

[0045] The electromagnetic actuating device 24 is in Fig. 2 is arranged on the right side of the pressure control valve 18. It comprises, among other things, an annular coil 68 and a centrally arranged armature 70. A push rod 72 arranged coaxially to the armature 70 transmits the actuating movement of the armature 70 to the control piston 34.

[0046] The first spiral spring 66 holds the control piston 34 in contact with the push rod 72. The push rod 72 is guided in a sealing but axially displaceable manner in a through opening 76 in the bearing bush 37, which closes the slide sleeve 28.

[0047] A second coil spring 78 is pushed onto the piston rod 72 in a concentric recess 80 of the armature 70. The coil spring 78 is supported on the one hand on the armature 70 and on the other hand on a coil core 82. The coil core 82 also forms a cover for a housing 84 in which the electromagnetic actuating device 24 is arranged. A plain bearing 88 is inserted into a coaxial bore 86 of the coil core 82, which receives the end of the piston rod 72 facing away from the armature 70.

[0048] The pressure control valve 18 functions as follows: When the control piston 34 is moved into the open position by the electromagnetic actuating device 24, in the currentless state of the coil 68 by the spiral spring 78 (in Fig. 2 to the left, not shown), then hydraulic oil flows under high pressure from the supply port 16 and via the inlet pressure opening 38 into the annular space 52 and from there via the control pressure opening 40 to the working port 22. The return opening 42 is essentially covered by the second control edge 56: This essentially separates the return port 20 from the annular space 52. As a result, the same pressure prevails at the working port 22 as at the supply port 16.

[0049] If, however, the control piston 34 is in a more right-hand position, for example, when the coil is energized, the supply pressure opening 38 is covered by the first control edge 54, thus essentially separating the annular space 52 from the supply connection 16. Instead, the second control edge 56 now opens the return opening 42, so that the working connection 22 communicates with the return connection 20 via the control pressure opening 40, the annular space 52, and the return opening 42. In this way, the pressure prevailing at the working connection 22 is reduced via the return connection 20 because, to a first approximation, ambient pressure prevails there.

[0050] The various intermediate positions of the control piston 34 allow any pressure to be set in the working port 22, whereby the pressure cannot be greater than that in the supply port 16 and cannot be less than that in the return port 20.

[0051] The position of the control piston 34 results from the equilibrium of forces between, on the one hand, the hydraulic forces acting on the control piston 34 and the push rod 72 and the restoring force of the first spiral spring 66 and, on the other hand, the force exerted by the electromagnetic actuating device 24 via the push rod 72 on the control piston 34 and the force applied by the second spiral spring 78 to the armature 70.

[0052] It is important for maintaining the pressure control function that in an equilibrium state, as in Fig. 2, the sum of the hydraulic forces acting on the end faces of the control piston 34 is approximately zero. This is ensured by the fact that Fig. 2 The first compensation chamber 64, located to the left of the control piston 34 and in which the first spiral spring 66 is arranged, communicates with the return opening 42 via the longitudinal bore 62 and the transverse bore, as does the Fig. 2 second compensation chamber 58 located to the right of the control piston 34. In both pressure chambers 58 and 64, the tank pressure present at the return connection 20 or in the return opening 42 prevails.

[0053] The closure 36 of the slide sleeve 28 ensures that the inlet pressure does not act on one end face of the control piston 34.

[0054] In the Fig. 3 and Fig. 4 shows the design features of the control piston 34 that are essential to the invention in an enlarged manner.

[0055] As can be seen from the Fig. 3, a circumferential rib 92 is formed on the end of the control piston 34 facing the push rod 72, which rib divides the second compensation chamber 58 into a third compensation chamber 100.

[0056] At the Fig. 3 At the lower edge of the rib 92, there is a flattening 94, which represents a possible damping element. A second damping element is the throttle 60, which connects the longitudinal bore 62 with the third compensation chamber. Of course, it is also possible (but in Fig. 3 not shown) to dispense with the rib 92 and to use the throttle 60 as the only damping element. It is also possible to provide an orifice plate instead of the throttle 60.

[0057] It is important, however, that in the spool valve according to the invention, the volume changes of the first compensation chamber 64 and the second compensation chamber 58 that occur with each movement of the control piston 34 result in liquid hydraulic oil flowing from one chamber to the other through the throttle 60 and the longitudinal bore 62. In this case, the first compensation chamber 64 and the second compensation chamber 58 are not subjected to the pressure of the supply port 16 or the pressure of the working port 22, but are more or less pressureless.

[0058] As a result, the energy losses caused by the damping according to the invention are very low. Furthermore, it is ensured that only actuating movements of the control piston 34 exert a damping effect, regardless of the hydraulic oil flow through the supply port 16 or the working port 22.

[0059] In the Fig. Figure 4 shows a section along the line A - A, from which the shape of the flattening 94 and alternative designs of damping elements in the form of a groove 96 and a notch 98 can be clearly seen.

[0060] In the Fig. 4 shows a gap S between the outer diameter of the rib 92 and the inner diameter of the guide bore 32. If this gap is dimensioned accordingly, an annular gap (without reference symbol) is created, which also has the damping effect according to the invention.

[0061] At the same time, it is also possible to reliably remove air bubbles from the second compensation chamber 58 via the annular gap. This can also be achieved by, as in Fig. As shown in Figure 4, the groove 96 or another damping means is arranged at the highest point of the rib 92 in the installed position. Alternatively, several damping means can be arranged distributed over the circumference of the rib 92.

Claims

[1] Pressure control valve (18) with a control piston (34) and a slide sleeve (28), wherein the pressure control valve (18) has a supply connection (16), a working connection (22) and a return connection (20), wherein the control piston (34) is guided with a first section (44) in a sealing manner in a guide bore (32) of the slide sleeve (28), wherein the first section (44) of the control piston (34), the guide bore (32) and a closure cap (36) delimit a first compensation chamber (64), the control piston (34) is guided with a third section (48) in a sealing manner in a guide bore (32) of the slide sleeve (28), the third section (48) of the control piston (34), the guide bore (32) and a bearing bush (37) delimit a second compensation chamber (58),the first compensation chamber (64) and the second compensation chamber (58) are hydraulically connected to one another and at least one damping element (60, 94, 96, 98) is arranged between the first compensation chamber (64) and the second compensation chamber (58), , characterized by that the slide sleeve (28) is tightly closed in the area of ​​the supply connection (16) by a sealing plug (36) pressed onto the front side. [2] Pressure control valve (18) according to claim 1, characterized by that a rib of the control piston (34) separates a third compensation chamber (100) from the second compensation chamber (58), and that the second compensation chamber (58) and the third compensation chamber (100) are hydraulically connected to one another by a damping element (94, 96, 98). [3] Pressure control valve (18) according to claim 1 or 2, characterized by that the damping element is designed as a flattening (94), groove (96), notch (98) and / or bore (60). [4] Pressure control valve (18) according to one of claims 2 or 3, characterized by that the damping element is designed as a circular annular gap between the rib (92) of the control piston (34) and the guide bore (32). [5] Pressure control valve (18) according to one of the preceding claims, characterized by that the control piston (34) in an open end position hydraulically connects the working connection (22) to the supply connection (16) and hydraulically separates the working connection (22) from the return connection (20). [6] Pressure control valve (18) according to one of the preceding claims, characterized by that the control piston (34) in a closed end position hydraulically connects the working connection (22) to the return connection (20) and hydraulically separates the working connection (22) from the supply connection (16). [7] Pressure control valve (18) according to one of the preceding claims, characterized bythat the control piston (34) has a first control edge (54), and that the first control edge (54) throttles the pressure medium flow flowing from the supply connection (16) to the working connection (22) in an intermediate position of the control piston (34). [8] Pressure control valve (18) according to one of the preceding claims, characterized by that the control piston (34) has a second control edge (56), and that the second control edge (56) throttles the pressure medium flow flowing from the working connection (22) to the return connection (20) in an intermediate position of the control piston (34). [9] Pressure control valve (18) according to one of the preceding claims, characterized by that the control piston (34) has at least one channel which connects the first compensation chamber (64) with the second compensation chamber (58) and / or the third compensation chamber (100). [10] Pressure control valve (18) according to claim 9, characterized bythat the channel comprises a longitudinal bore (62) and a throttle (60) or an orifice.

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