EP0609233A1

EP0609233A1 - Hydraulic control device.

Info

Publication number: EP0609233A1
Application number: EP92918895A
Authority: EP
Inventors: Helmut Rembold; Martin Mueller
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 1991-10-10
Filing date: 1992-09-04
Publication date: 1994-08-10
Anticipated expiration: 2012-09-04
Also published as: DE59204625D1; US5476031A; WO1993007362A1; DE4133526A1; JPH07500163A; EP0609233B1; KR100287308B1

Abstract

L'organe de réglage hydraulique (10) possède un cylindre différentiel (11) dont les chambres de pression (14, 18) sont alimentées chacune par une pompe (17). Les conduites de refoulement (15 et 19) entre la pompe et les deux chambres de pression (14 et 18) sont reliées par des conduites de pilotage (29 et 33) à la soupape de commande (31). Par un pilotage correspondant de la soupape de commande, il est possible de régler une pression partielle dans les chambres de pression (14 et 18) par l'écoulement de fluide hydraulique.The hydraulic adjustment member (10) has a differential cylinder (11), the pressure chambers (14, 18) of which are each supplied by a pump (17). The delivery lines (15 and 19) between the pump and the two pressure chambers (14 and 18) are connected by pilot lines (29 and 33) to the control valve (31). By corresponding actuation of the control valve, it is possible to set a partial pressure in the pressure chambers (14 and 18) by the flow of hydraulic fluid.

Description

Hydraulic actuator

State of the art

The invention is based on a hydraulic actuator according to the preamble of the main claim. From US Pat. No. 3,516,331, a hydraulic actuating device with a differential cylinder is known, the pressure chamber of which is assigned to the larger effective pressure surface of the differential piston and is controlled via a 3/2-way valve. By correspondingly clocking the 3/2-way valve, a pressure difference can be generated in the two pressure spaces, which causes an adjustment movement. Such a hydraulic actuating device has the disadvantage that with stationary, ie. H. Non-moving differential pistons in the pressure chambers are subject to relatively high pressures, which are not or only slightly less than the adjustment pressures. As a result, in the holding position (stationary position) of the differential piston, a high expenditure of energy is necessary, which can lead to high costs in the operation of the hydraulic actuating device.

A hydraulic actuating device is known from DE-OS 40 37 824, in which these disadvantages are avoided. In this known hydraulic actuating device, the large one Piston surface of a differential cylinder arranged pressure chamber controlled via an electromagnetically actuated control valve. This control valve is designed so that it has negative overlap in its central position. It is controlled in such a way that the pressures in the pressure chambers of the differential piston remain approximately constant and that the holding pressures when the differential piston is in a stationary position are substantially lower than the adjustment pressures. Such a hydraulic actuating device is used, for example, to actuate a device for adjusting the camshaft relative to the crankshaft in an internal combustion engine (DE-OS 36 16 234). In order to securely seal the respective pressure-carrying connection against the return connection in the end positions of the control valve, narrow and long guide gaps are sometimes necessary for the valve member in these control valves. As a result, these control valves may be sensitive to dirt, ie if the pressure medium (engine oil of the internal combustion engine) is contaminated, the valve function can be impaired. In addition, such control valves are expensive to manufacture due to the necessary small tolerances.

Advantages of the invention

The hydraulic actuating device according to the invention with the characterizing features of the main claim has the advantage that it works with low losses if there is no adjustment movement of the differential piston, that it is simple in construction and that the sensitivity of the control valve to dirt is low .

Further advantages of the invention and advantageous developments result from the subclaims and the description. drawing

Two embodiments of the invention are explained in more detail in the following description and drawing. The latter shows a simplified illustration of a first exemplary embodiment of the hydraulic actuating device in FIG. 1. FIG. 2 shows the pump of the hydraulic actuating device in a simplified representation, and FIG. 3 shows the control valve of the hydraulic actuating device in a longitudinal section. Figure 4 shows a second embodiment of the hydraulic actuator in a simplified representation.

Description of the embodiments

In FIG. 1, 10 denotes a hydraulic actuating device which has a differential cylinder 11 with differential pistons 12, 13. The pressure chamber 14 on the large piston surface of the differential piston 12 is connected via a pressure line 15 to a pump work chamber 16 of a pump 17 - shown in more detail in FIG. 2. The pressure chamber 18 on the smaller effective piston surface of the differential piston 12 is connected via a pressure line 19 to a further pump chamber 20 of the pump 17, which acts in the opposite direction to the first pump chamber 16.

The pump work space 16 is supplied with pressure medium via a supply line 21 opening into the pressure line 15. In these

Supply line 21, a check valve 22 is used, which in the pressure medium flow from the pressure medium designated P

M source to the pump work space 16 opens. The analog is

Pump work room 20 via a mouth opening in the pressure medium line 19

Supply line 24 with check valve 25 connected to the pressure medium source P. The pressure medium source P is MM is, for example, a device for supplying pressure medium or lubricant to an internal combustion engine.

A check valve 26 and 27 is arranged in each of the pressure lines 15 and 19 between the supply line 21 and 24 and the pressure chamber 14 and 18, which opens when the pressure medium flows from the pump work chamber to the pressure chamber.

A control line 29 branches off from the pressure line 15 between the check valve 26 and the pressure chamber 14 and is connected to a connection 30 of a control valve 31 - shown in more detail in FIG. 3. In the exemplary embodiment, the control valve 31 is a 3/2 seat valve, from the second connection 32 of which a control line 33 emerges, which opens into the pressure line 19 - between the check valve 27 and the pressure chamber 18. The third connection of the control valve is designed as a return 34 and connected to a container 35.

In the exemplary embodiment, the pump 17 shown schematically in FIG. 2 is a radial piston pump with pistons operating in opposite directions, which are driven, for example, by the camshaft of an internal combustion engine via a drive shaft 36. When using the hydraulic actuator z. B. for adjusting the camshaft of a motor vehicle relative to its crankshaft (DE-OS 36 16 234), the camshaft can also be used directly as the drive shaft of the pump. The two pump work spaces 16 and 20 are arranged offset from one another by 180 °, their pistons 37 and 38 are driven by an eccentric 40 arranged on the drive shaft 36. The control valve 31 shown in FIG. 3 has an approximately cup-shaped housing 41, in the bottom 42 of which a central bore 43 is arranged. Two sleeve-shaped extensions 44, 45 extend from the bottom 42, of which the extension 44 projects into the interior of the housing 41 and the extension 45 points in the opposite direction. The extensions 44, 45 are dimensioned such that their interior spaces, together with the bore, form a cylindrical valve space 46.

A recess 48 extends from the free end face of the extension 45 and is closed on one side by a cover 49 resting on the end face. The depression 48 extends as far as a sleeve 47, which is inserted into the valve chamber 46 and is made of a non-magnetic material. The sleeve 47 and the extension

45 are penetrated by a transverse bore 50 which is connected to the return 34 of the control valve and via which the valve chamber

46 is connected to the container 35.

The housing 41 is surrounded by a cylinder jacket 51 made of non-magnetic material, which projects above the housing and is closed by a cover 52, so that an armature space 53 is formed.

A magnetic coil 54 is inserted into the interior of the housing 41, which comprises the sleeve-shaped extension 44 and whose inside diameter is larger than its outside diameter. In the annular space 55 formed between the magnet coil 54 and the extension 44, a compression spring 56 is inserted, one end of which rests on the bottom 42 of the housing and the other end of which rests on a disk-shaped flat armature 57 which is arranged in the armature space 53. This flat armature 57 interacts with an essentially cylindrical valve member 59 which is guided in the sleeve 47. The length of the valve member 59 is less than the distance between the covers 49 and 52. The valve member 59 penetrates the flat armature 57 in the center and is firmly connected to it. On the end face facing the cover 52, the valve member has a shoulder 60 of smaller diameter. The free end face 61 of the shoulder 60 interacts with a bore 62 in the cover 52 which is designed as a valve seat and is connected to the control line 33.

Around the valve member 59, the flat armature 57 is penetrated by a plurality of regularly arranged bores 65, which serve to pass the pressure medium through.

The valve member 59 projects with its end facing the cover 49 into the recess 48 and there has a shoulder 66 of smaller diameter, the end face 67 of which cooperates with a bore 68 in the cover 49 designed as a valve seat. The bore 68 serves as the first connection 30 of the control valve 31 and is connected to the control line 29.

The valve member 59 has a section 70 of smaller diameter located within the sleeve 47, so that an annular space 71 is formed between the latter and the sleeve 47. The outer sections 72 and 73 of larger diameter guide the valve member 59 in the sleeve 47 and have flattened areas 74 and 75, respectively, for the passage of the pressure medium, which pressure medium can flow past.

The hydraulic actuating device 10 is used, for example, in a device for continuously adjusting the camshaft of an internal combustion engine relative to its crankshaft, as a result of which a phase shift is generated between these two shafts. A displacement of the differential piston 12, 13 to the left (FIG. 1) produces an adjustment of the camshaft to "late" in this device, ie to a late rotational position or later valve actuation. An adjustment of the differential piston to the right consequently produces an adjustment according to "early" or earlier rotational position and earlier valve actuation.

In the switching position of the currentless control valve 31 shown in FIG. 1, the control line 29 is connected to the container 34, while the control line 33 is closed on one side by the control valve 31. In this switching position, the shoulder 60 of the valve member 59 shown in FIG. 3 lies against the bore 62 connected to the control line 33 due to the action of the compression spring 56 and closes it. At the same time, pressure medium can get from the control line 29 through the bore 68 into the depression 48. From there, there is a connection to the transverse bore 50 and thus to the container 35, specifically via the space between the sleeve 47 and the flattened regions 74 of the section 72 and via the annular space 71. The pressure line 15 and thus also the pressure space 14 of the Differential cylinder 11 to the container 34 is relieved, while the pressure chamber 18 is pressurized by the pump 17 via the pump work chamber 20. The differential piston 12, 13 is moved to the left.

If the differential piston 12, 13 is to be moved to the right, the control valve 31 is switched into the second switching position by appropriate actuation of the solenoid 54, so that the shoulder 66 of the valve member 59 closes the bore 68 and thus the control line 29 on one side. As a result, the opposite bore 62 is then connected to the transverse bore 50 or the return 34 and the container 35 via the armature space 53, the space between the sleeve 47 and the flats 75 of the section 73 and the annular space 71. In this switching position, the pressure line 19 and thus the pressure chamber 18 to the container 35 are relieved, while the pressure chamber 14 is pressurized by the pump 17 via the pump working chamber 16 and the pressure line 15.

A stationary position of the differential piston 12, 13 is achieved by correspondingly clocked or proportional activation of the control valve 31, a pressure being set in the control line 29 and thus in the pressure chamber 14 which is just sufficient for the restoring force (acting on the adjusting device) of the differential piston. The holding pressures in this stationary position of the differential piston are thus very much lower than the pressures required for a (fast) adjustment movement.

Appropriate control of the solenoid coil also ensures that these holding pressures are maintained at a level which is just sufficient to absorb the restoring forces from the device for adjusting the camshaft, even when the camshaft speeds change.

The described design of the hydraulic actuating device and the control valve 31 ensures that the internal combustion engine runs smoothly even if the control valve or the hydraulic supply fails. In the non-activated state, the control valve 31 assumes the switch position shown in FIG. 1 due to the action of the spring 56. As a result, as previously described, the pressure chamber 18 is pressurized, while the pressure chamber 14 is relieved of pressure to the container 35. This moves the differential piston to the left ("late"). If the hydraulic supply fails, the differential piston 12, 13 is moved to the left due to the mechanical restoring force from the device for adjusting the camshaft. In both cases, an engine emergency run is ensured due to this reset to the late rotational position of the camshaft. Due to the described design of the control valve 31 or the valve member 59, only the pressure of the return 34 prevails in the region of the guide of the valve member. On the one hand, the effective guide length for the valve member can be kept small, since no sealing function against higher pressures is necessary. On the other hand, there is no great pressure difference in the guide area of the valve member, through which dirt could be conveyed into the guide gap. In order to prevent metallic abrasion from accumulating due to the effect of magnetic fields in the area of the valve member, the sleeve used for guiding is made of non-magnetic material. The control valve 31 can, as shown in FIG. 3, be designed with a flat armature or as a proportional solenoid valve with a suitably designed magnetic circuit.

In order to keep the adjustment speed of the differential piston constant over the entire speed range of the internal combustion engine or the drive shaft 36, the control valve 31 can be used as a pressure control valve. By appropriate metering of the magnetic force (corresponding control of the magnetic coil), the valve member is then only pressed against the valve seat as strongly as is necessary for the corresponding generation of pressure. It is advantageous to design the control valve with a suitably designed magnetic circuit as a proportional solenoid valve, since the adjusting forces differ depending on the speed of the drive shaft. In order to keep the differential piston 12, 13 in a stationary position, the solenoid 54 is then actuated so that the pressure in the control line 29 and thus in the pressure chamber 14 maintains the equilibrium with the restoring forces acting on the differential piston. In an advantageous embodiment of the first exemplary embodiment of the hydraulic actuating device, the pump 17 is throttled on the suction side. B. via a slot-controlled intake throttling. This makes it possible to implement a flow rate curve that is constant over the entire speed range of the internal combustion engine or the drive shaft. The pump or the intake throttling is designed so that the beginning of the constant delivery range (constant flow of conveying medium) coincides with the lower limit speed of the working range (e.g. idle speed of the internal combustion engine). The pump delivery rate is matched to the required adjustment speed of the differential cylinder. The control valve 31 is designed as a simple solenoid valve with a flat armature, since a pressure control function - as described above - is not required. Irrespective of this, the speed of adjustment of the differential piston can be influenced by clocking the solenoid valve. The holding function (stationary position of the differential cylinder) can also be realized by correspondingly timed control of the control valve. In order to avoid different fillings of the pump work spaces in the constant delivery area of the pump due to fluctuations in the pressure medium supply, the pressure medium supply of the pump is expediently from a reservoir (container).

In the second exemplary embodiment of the hydraulic actuating device shown in FIG. 4, a pump 17a feeds into a common delivery line 80, from which two pressure lines 81, 82 originate. A pressure-controlled check valve 83, 84 is connected to each of the two pressure lines 81, 82 for reversing the pump delivery flow. The check valve 83 is connected on the outlet side to the pressure line 15 and the check valve 84 to the pressure line 19. The check valves 83, 84 are designed so that they open when there is a pressure medium flow from the pump 16a to the differential cylinder 11. For this purpose, the respective valve members 85, 86 are each acted upon by a compression spring 87 or 88 and additionally by the pressure in a control line 89 or 90. The control line 89 on the check valve 83, on the other hand, is connected to the pressure line 19, while the control line 90 on the check valve 84 leads to the pressure line 15.

A throttle 91 is inserted between the control line 90 and the check valve 83 in the pressure line 15. A throttle 92 is also inserted into the pressure line 19, namely between the check valve 84 and the control line 89.

The control valve 31 is connected via the control lines 29 and 33 to the pressure lines 15 and 19, respectively between the control lines 90 and 89 leading to the check valves and the differential cylinder 11.

In the illustrated position of the control valve 31 and the check valves 83, 84, both check valves 83, 84 can open when the actuating device is initially depressurized and the pump 17a starts. A certain pressure builds up in front of the throttles 91 and 92 in the pressure lines 81, 82 and in the pressure lines 15 and 19, which pressure also prevails in the pressure line 19 behind the throttle 92. Due to the relief of the pressure line 15 to the container via the control valve 31, no such pressure can build up behind the throttle 91.

The pressure in the pressure line 19 also acts on the check valve 83 via the control line 89, so that the check valve 83 is closed due to the additional force of the compression spring 87. Due to the pressure building up in the pressure chamber 18 and due to the relief of the pressure chamber 14 to the container 35 the differential piston is moved to the left ("late").

In order to produce an adjustment of the differential piston to the right ("early"), the control valve 31 is moved into the second switching position by appropriate excitation of the solenoid coil, so that the pressure line 19 to the container is relieved. In accordance with the switching position described above, the check valve 84 is then moved into the closed position, so that when the pressure chamber 18 and pressure chamber 14 are relieved of pressure, the differential piston moves to the right.

The stop position (stationary position of the differential piston) can be achieved either by appropriately clocked control of the control valve or by pressure control with a partially excited solenoid. In order to limit the power loss of the hydraulic actuating device, the pressure drop at the throttles 91 and 92 should be limited to, for example, 5 to 10 bar.

Claims

Expectations

1. Hydraulic actuating device (10) with at least one hydraulic pump (17, 17a) and with a differential cylinder (11), in the pressure chambers (14 and 18) of which partial pressure is set by partial discharge of pressure medium via an electromagnetically actuated control valve (31) that certain by corresponding activation of the control valve continuously or discontinuously in dependence criteria is changeable, wherein holding pressures are set at stationary position of the ^• differential piston, which are much smaller in than the adjustment pressures, characterized in that each of the pressure chambers in each case via a pressure line (15 , 81; 19, 82) is connected to the pump, and that the control valve (31) is connected between the two pressure lines (15, 19).

2. Hydraulic actuating device according to claim 1, characterized gekenn¬ characterized in that the control valve (31) is a 3/2-way valve in Sitzbau¬ way.

3. Hydraulic actuating device according to claim 1 or 2, characterized ge indicates that the control valve (31) is connected to a return (34), and that in the region of the guide of the valve member (59) essentially the pressure present in the return prevails.

4. Hydraulic actuating device according to one of claims 1 to 3, characterized in that each pressure line (15, 19) is each connected to at least one pressure chamber (16, 20) of the pump (17).

5. Hydraulic actuating device according to one of claims 1 to 3, characterized in that in each pressure line (15, 19) a check valve (83, 84) controlled by the pressure in the other line (19, 15) is arranged.

6. Hydraulic actuating device according to claim 5, characterized gekenn¬ characterized in that the check valve (83, 84) is a check valve, the valve member (85, 86) is acted upon by a compression spring and the pressure in a control line (89, 90).

7. Hydraulic actuating device according to one of claims 1 to 6, characterized in that a throttle (91, 92) is arranged in each pressure line (15, 81; 19, 82).

8. Hydraulic actuating device according to claim 7, characterized gekenn¬ characterized in that the throttle (91, 92) between the check valve (83, 84) and the check valve (84, 83) in the other Druck¬ line control line (90, 89th ) is arranged.