EP3557013A1 - Hydraulic valve drive for a cylinder valve of a combustion engine - Google Patents

Abstract

A hydraulic valve train for a cylinder valve of an internal combustion engine has a hydraulic brake which brakes a movement of a control piston from its rest position into its working position by increasing a flow resistance of a hydraulic fluid displaced by the control piston in the movement end phase. Providing the hydraulic brake, which brakes the movement of the control piston from the rest position into the working position, can minimize a noise which the control piston generates when the working position is reached, for example as a result of hitting a stop, a housing, etc.

Description

The present invention relates to a hydraulic valve drive for a cylinder valve of an internal combustion engine.

STATE OF THE ART

WO 2016/000048 A1 discloses a hydraulic valve train for a cylinder valve of an internal combustion engine. The valve drive has a control piston and a working piston, which are reciprocable in a housing and forth. The hydraulic valve train has brake mechanisms to decelerate movement of the control piston and the working piston, respectively, so that noise generation is minimized and operating accuracy is ensured.

Based on this prior art, it is the object of the invention to provide a hydraulic valve train for a cylinder valve of an internal combustion engine, which has a simple structure and further minimizes noise generation.

The object of the invention is achieved with a hydraulic valve drive for a cylinder valve of an internal combustion engine according to claim 1. Advantageous developments of the invention are the subject of the dependent claims.

• ein Gehäuse (einteilig oder mehrteilig) mit einem Einlasskanal und einem Auslasskanal,
• einen Steuerkolben, der in dem Gehäuse angeordnet ist und zwischen einer Ruhestellung, in der er den Einlasskanal schließt, während der Auslasskanal geöffnet ist, und einer Arbeitsstellung hin- und herbewegbar ist, in der er den Auslasskanal schließt, während der Einlasskanal geöffnet ist, wobei eine Bewegung des Steuerkolbens in die Arbeitsstellung einen Druckanstieg im Inneren des Gehäuses aufgrund eines Einströmens von Hydraulikfluid durch den Einlasskanal bewirkt und eine Bewegung des Steuerkolbens in die Ruhestellung einen Druckabfall im Inneren des Gehäuses aufgrund eines Ausströmens von Hydraulikfluid aus dem Auslasskanal bewirkt,
• einen Arbeitskolben, der in dem Gehäuse angeordnet ist, der mit dem Zylinderventil wirkverbunden ist und der sich in Abhängigkeit des Druckanstiegs und des Druckabfalls im Inneren des Gehäuses hin- und herbewegt, um das Zylinderventil zu öffnen und zu schließen, und
• eine hydraulische Bremse, die eine Bewegung des Steuerkolbens von der Ruhestellung in die Arbeitsstellung durch Erhöhen eines Strömungswiderstands eines von dem Steuerkolben verdrängten Hydraulikfluids in der Bewegungsendphase abbremst.

The valve drive according to the invention for a cylinder valve of an internal combustion engine has the following:

• a housing (one-piece or multi-part) with an inlet channel and an outlet channel,
• a control piston which is disposed in the housing and between a rest position in which it closes the inlet channel while the outlet channel is opened, and a working position is reciprocated, in which it closes the outlet channel, while the inlet channel is opened, wherein a movement causes the control piston in the working position, a pressure increase inside the housing due to an inflow of hydraulic fluid through the inlet channel and causes a movement of the control piston in the rest position, a pressure drop inside the housing due to a leakage of hydraulic fluid from the outlet channel,
• a working piston which is disposed in the housing, which is operatively connected to the cylinder valve and which reciprocates in response to the pressure rise and the pressure drop inside the housing to open and close the cylinder valve, and
• a hydraulic brake that decelerates a movement of the control piston from the rest position to the working position by increasing a flow resistance of a displaced from the control piston hydraulic fluid in the movement end phase.

By providing the hydraulic brake which decelerates the movement of the control piston from the rest position to the working position, a noise can be minimized, which generates the control piston when reaching the working position, for example. As a result of hitting a stop, the housing, etc. Moreover, since the hydraulic brake is active only in the final movement phase (i.e., just before reaching the working position), i. slows down the movement of the control piston, a quick and reliable operation of the control piston can be achieved. Since the brake operates on the principle of increasing a flow resistance, a structure of the hydraulic valve train is simplified and the number of wearing parts is minimized. In addition, a uniformly increasing braking effect is achieved.

The flow resistance in the hydraulic brake is preferably increased by virtue of the fact that in the movement end phase a part of the displaced hydraulic fluid is pressed by the control piston through a brake channel.

This represents an extremely simple, reliable and cost-effective design of the hydraulic brake.

Preferably, the hydraulic brake is essentially formed by a trained on the control piston circumferential step and the brake channel.

A surface of the revolving stage extending in the axial direction of the control piston preferably coincides with a wall surface, extending in the axial direction of the control piston, of a section in which the brake channel is formed, thereby separating in cooperation with a piston extending in the radial direction of the control piston Surface of the orbiting stage, a in the axial direction of the control piston extending wall surface of the housing, and extending in the radial direction of the control piston brake channel surface into which opens the brake channel, the part of the displaced hydraulic fluid from the remaining displaced hydraulic fluid from, and the separated part of the displaced Hydraulic fluid is forced through the brake channel.

This embodiment provides a simple and cost-effective way to realize the hydraulic brake.

Preferably, the brake channel surface is a stop surface for the control piston, against which the control piston rests in the working position.

Through this functional integration of stop surface and brake channel surface of the space is effectively used and a structure of the hydraulic valve train is simplified.

The brake channel is preferably arranged in a damping ring (for example an elastic component such as a rubber component), which is arranged between housing parts of the housing, or is formed in the housing.

The provision of the brake channel in the damping ring or in the housing, which are already components of the hydraulic valve train, allows a simple, space- and component-saving design of the hydraulic valve train. If the brake channel is arranged in the damping ring and the damping ring forms the stop surface for the control piston, a noise be further minimized due to the damping properties of the damping ring.

Preferably, the hydraulic valve train further comprises a second hydraulic brake which decelerates a movement of the control piston from the working position to the rest position by increasing a flow resistance of a displaced from the control piston hydraulic fluid in the movement end phase.

By providing the second hydraulic brake, which decelerates the movement of the control piston from the working position to the rest position, a noise can be minimized, which generates the control piston upon reaching the rest position, for example. As a result of hitting a stop, the housing, etc. Moreover, since the second hydraulic brake is active only in the final movement phase (i.e., shortly before reaching the rest position), i. slows down the movement of the control piston, a quick and reliable operation of the control piston can be achieved. Since the second hydraulic brake operates on the principle of increasing a flow resistance, a structure of the hydraulic valve train is simplified and the number of wearing parts is minimized. In addition, a uniformly increasing braking effect is achieved.

In the case of the second hydraulic brake, the flow resistance is preferably increased by virtue of the fact that, in the movement end phase, part of the displaced hydraulic fluid is pressed by the control piston through a second brake channel.

This represents an extremely simple, reliable and inexpensive embodiment of the second hydraulic brake.

Preferably, the second hydraulic brake is provided essentially by a circumferential shoulder of the control piston and the second brake channel.

A surface of the shoulder which extends in the radial direction of the control piston preferably comes into register in the movement end phase with a wall surface of the housing extending in the axial direction of the control piston separates it in cooperation with an axially extending in the axial direction of the control piston surface of the control piston and an inner wall surface of the housing, in which opens the second brake channel, the part of the displaced hydraulic fluid from the remaining displaced hydraulic fluid, and the separated part of the displaced hydraulic fluid is through the pressed second brake channel.

This embodiment is a simple and cost-effective way to realize the second hydraulic brake.

Preferably, the second brake channel is formed in the housing. But it can also be formed in another component. The provision of the brake channel in the housing, which is already part of the hydraulic valve train, allows a simple, space- and component-saving design of the hydraulic valve train.

Preferably, the hydraulic valve train further includes a third hydraulic brake that decelerates movement of the working piston due to a pressure increase inside the housing by increasing a flow resistance of a hydraulic fluid displaced from the working piston in the working piston movement end phase.

By providing the third hydraulic brake which decelerates the movement of the working piston due to a pressure increase inside the housing, a noise can be minimized, which generates the working piston upon reaching its end position, for example. As a result of hitting a stop, the housing, etc. Moreover, since the third hydraulic brake is active only in the working piston movement end phase (ie, shortly before reaching the corresponding piston end position), ie, decelerates the movement of the working piston, a quick and reliable operation of the working piston can be achieved. Since the third hydraulic brake operates on the principle of increasing a flow resistance, a structure of the hydraulic valve train is simplified and the number of wearing parts is minimized. In addition, a uniformly increasing braking effect is achieved.

Preferably, the movement of the working piston due to a pressure increase inside the housing is a movement that leads to the opening of the cylinder valve. Alternatively, it can also be a movement that leads to the closing of the cylinder valve.

In the third hydraulic brake, the flow resistance is preferably increased by virtue of the fact that, in the working piston movement end phase, part of the displaced hydraulic fluid is pressed by the working piston through a third brake channel.

This represents an extremely simple, reliable and cost-effective design of the hydraulic brake.

Preferably, the third hydraulic brake is formed substantially by a rotating on the working piston stage and the third brake channel.

Preferably, a surface of the revolving stage extending in the axial direction of the working piston in the working piston movement end phase coincides with a wall surface of a section in the axial direction of the working piston in which the third braking channel is formed, thereby separating in cooperation with a working piston in the radial direction extending surface of the orbiting stage, extending in the axial direction of the working piston wall surface of the housing, and extending in the radial direction of the working piston brake channel surface into which opens the brake channel, the part of the displaced hydraulic fluid from the remaining displaced hydraulic fluid from, and the separated part of the displaced hydraulic fluid is forced through the third brake channel.

This embodiment is a simple and inexpensive way to realize the third hydraulic brake.

Preferably, the third brake channel in a damping ring, via which the valve train is to be attached to a cylinder head of the internal combustion engine, or formed in the housing.

The provision of the third brake channel in the damping ring or in the housing, which are already components of the hydraulic valve train, allows a simple, space- and component-saving design of the hydraulic valve train. If the third brake channel is arranged in the damping ring and the damping ring forms a stop surface for the working piston, a noise due to the damping properties of the damping ring can be further minimized.

Preferably, the hydraulic valve train further includes a fourth hydraulic brake that decelerates movement of the working piston due to a pressure drop within the housing by increasing a flow resistance of a hydraulic fluid displaced from the working piston in the working piston travel end phase.

By providing the fourth hydraulic brake, which slows down the movement of the working piston due to a pressure drop inside the housing, a noise can be minimized which the working piston upon reaching its end position, for example. As a result of hitting a stop, the housing, another component etc . generated. Moreover, since the fourth hydraulic brake is active only in the working piston movement end phase (i.e., shortly before reaching the corresponding piston end position), i. slows down the movement of the working piston, a faster and more reliable operation of the working piston can be achieved. Since the fourth hydraulic brake operates on the principle of increasing a flow resistance, a structure of the hydraulic valve train is simplified and the number of wearing parts is minimized. In addition, a uniformly increasing braking effect is achieved.

Preferably, the movement of the working piston due to a pressure drop in the interior of the housing is a movement for closing the cylinder valve leads. Alternatively, it may also be a movement that leads to the opening of the cylinder valve.

Preferably, in the fourth hydraulic brake, the flow resistance is increased by reducing a maximum possible hydraulic fluid discharge amount from a fluid chamber forming the working piston with a component fixed to the housing and decreasing its volume with movement of the working piston due to a pressure drop in the working piston movement end phase becomes.

This represents an extremely simple, space-saving, reliable and cost-effective embodiment of the fourth hydraulic brake.

Preferably, the component has at least one opening, through which a hydraulic fluid can flow out of the fluid chamber, and the working piston has at least one opening through which a hydraulic fluid can flow out of the fluid chamber, and the valve drive is designed so that in a Movement of the working piston due to a pressure drop, a hydraulic fluid from the fluid chamber initially flows through both the opening of the component and through the opening of the working piston, and then, in the Arbeitskolbenbewegungsendphase, only flows through the opening of the component.

This embodiment represents a simple and inexpensive way to realize the fourth hydraulic brake.

Preferably, the hydraulic valve train has an electromagnetic drive system configured to reciprocate the control piston between the rest position and the working position. The electromagnetic drive system allows accurate and quick adjustment and reliable operation of the control piston.

Preferably, the hydraulic valve train on a compensation component, which is an upwardly open cylinder and which bears against the brake cylinder, wherein an end sleeve of the control piston is received axially displaceable in the compensation component.

BRIEF DESCRIPTION OF THE DRAWINGS

• Fig. 1 ist eine Längsschnittansicht des erfindungsgemäßen hydraulischen Ventiltriebs gemäß einem ersten Ausführungsbeispiel sowie eines Teils eines Zylinderkopfs einer Brennkraftmaschine, an dem der hydraulische Ventiltrieb montiert ist, wobei Fig. 1 einen Zustand zeigt, in dem ein in dem Zylinderkopf aufgenommenes Zylinderventil durch den Ventiltrieb geschlossen ist.
• Fig. 2A ist eine Längsschnittansicht des erfindungsgemäßen hydraulischen Ventiltriebs und Fig. 2B ist eine Draufsicht des erfindungsgemäßen hydraulischen Ventiltriebs, wobei Fig. 2A und 2B einen Zustand zeigen, in dem das in dem Zylinderkopf aufgenommene Zylinderventil durch den Ventiltrieb geöffnet ist.
• Fig. 3 ist eine Querschnittsansicht entlang Linie B-B in Fig. 2.
• Fig. 4A ist eine Teilansicht des hydraulischen Ventiltriebs und stellt einen ersten Betriebszustand des hydraulischen Ventiltriebs dar.
• Fig. 4B ist eine Teilansicht des hydraulischen Ventiltriebs und stellt einen zweiten Betriebszustand des hydraulischen Ventiltriebs dar.
• Fig. 4C ist eine Teilansicht des hydraulischen Ventiltriebs und stellt einen dritten Betriebszustand des hydraulischen Ventiltriebs dar.
• Fig. 4D ist eine Teilansicht des hydraulischen Ventiltriebs und stellt einen vierten Betriebszustand des hydraulischen Ventiltriebs dar.
• Fig. 5A zeigt vergrößert einen Bereich des hydraulischen Ventiltriebs, in dem eine hydraulische Bremse, die in einem noch nicht aktivierten Zustand ist, vorgesehen ist.
• Fig. 5B zeigt vergrößert einen Bereich des hydraulischen Ventiltriebs, in dem die hydraulische Bremse, die in einem aktivierten Zustand ist, vorgesehen ist.
• Fig. 5C zeigt vergrößert einen Bereich des hydraulischen Ventiltriebs, in dem eine zweite hydraulische Bremse, die in einem noch nicht aktivierten Zustand ist, vorgesehen ist.
• Fig. 5D zeigt vergrößert einen Bereich des hydraulischen Ventiltriebs, in dem die zweite hydraulische Bremse, die in einem aktivierten Zustand ist, vorgesehen ist.
• Fig. 5E zeigt vergrößert einen Bereich des hydraulischen Ventiltriebs, in dem eine dritte hydraulische Bremse, die in einem noch nicht aktivierten Zustand ist, vorgesehen ist.
• Fig. 5F zeigt vergrößert einen Bereich des hydraulischen Ventiltriebs, in dem die dritte hydraulische Bremse, die in einem aktivierten Zustand ist, vorgesehen ist.
• Fig. 6 ist eine Längsschnittansicht des erfindungsgemäßen hydraulischen Ventiltriebs gemäß einem zweiten Ausführungsbeispiel sowie eines Teils eines Zylinderkopfs einer Brennkraftmaschine, an dem der hydraulische Ventiltrieb montiert ist.
• Fig. 7A ist eine Längsschnittansicht einer Modifikation gemäß einem dritten Ausführungsbeispiel.
• Fig. 7B ist eine Teilansicht einer Modifikation des hydraulischen Ventiltriebs gemäß einem vierten Ausführungsbeispiel.

Hereinafter, preferred embodiments of the invention will be described with reference to the accompanying drawings.

• Fig. 1 is a longitudinal sectional view of the hydraulic valve train according to the invention according to a first embodiment and a part of a cylinder head of an internal combustion engine on which the hydraulic valve train is mounted, wherein Fig. 1 shows a state in which a cylinder valve received in the cylinder head is closed by the valve train.
• Fig. 2A is a longitudinal sectional view of the hydraulic valve gear according to the invention and Fig. 2B is a plan view of the hydraulic valve train according to the invention, wherein Fig. 2A and 2 B show a state in which the cylinder valve received in the cylinder head is opened by the valvetrain.
• Fig. 3 is a cross-sectional view taken along line BB in FIG Fig. 2 ,
• Fig. 4A is a partial view of the hydraulic valve train and represents a first operating state of the hydraulic valve train.
• Fig. 4B is a partial view of the hydraulic valve train and represents a second operating state of the hydraulic valve train.
• Fig. 4C is a partial view of the hydraulic valve train and represents a third operating state of the hydraulic valve train.
• Fig. 4D is a partial view of the hydraulic valve train and represents a fourth operating state of the hydraulic valve train.
• Fig. 5A shows enlarged a portion of the hydraulic valve train in which a hydraulic brake, which is in a not yet activated state, is provided.
• Fig. 5B Fig. 10 shows an enlarged portion of the hydraulic valve train in which the hydraulic brake which is in an activated state is provided.
• Fig. 5C shows enlarged a portion of the hydraulic valve train in which a second hydraulic brake, which is in a not yet activated state, is provided.
• Fig. 5D Fig. 15 shows an enlarged portion of the hydraulic valve train in which the second hydraulic brake, which is in an activated state, is provided.
• Fig. 5E Fig. 11 shows an enlarged portion of the hydraulic valve train in which a third hydraulic brake, which is in a not yet activated state, is provided.
• Fig. 5F Fig. 11 shows an enlarged portion of the hydraulic valve train in which the third hydraulic brake, which is in an activated state, is provided.
• Fig. 6 is a longitudinal sectional view of the hydraulic valve train according to the invention according to a second embodiment and a part of a cylinder head of an internal combustion engine, on which the hydraulic valve train is mounted.
• Fig. 7A is a longitudinal sectional view of a modification according to a third embodiment.
• Fig. 7B is a partial view of a modification of the hydraulic valve train according to a fourth embodiment.

DESCRIPTION OF THE PREFERRED EMBODIMENTS First embodiment

The structural design of the hydraulic valve drive according to the invention for a cylinder valve of an internal combustion engine according to a first embodiment will first with reference to FIGS. 1 to 3 explained.

The hydraulic valve drive has a housing which is designed in several parts in the present case. In particular, the housing is essentially formed of a working piston housing 2, a control piston housing 3 and a cover 5.

The working piston housing 2 defines a working area 32 in the radial direction and accommodates therein a working piston 10, 11, which is constructed in two parts in this embodiment, ie, a first piston part 10 and a second piston part 11, which are positively connected with each other. The working piston 10, 11 is radially supported in the working piston housing 2 and received axially displaceable. The end position of the working piston 10, 11 down in Fig. 1 and 2 is determined by an axial abutment of the working piston 10, 11 on a damping ring 7, via which the working piston housing 2 is flanged or fastened to a cylinder head 1 (alternatively, the working piston housing 2 can also be attached directly to the cylinder head 1, in which case axial end position of the working piston 10, 11 may be determined by the cylinder head 1 or a component connected thereto). The end position of the working piston 10, 11 upward remains indefinite in order to compensate for a thermal expansion of a valve 16 received in the cylinder head 1 can. The second piston part 11 of the working piston 10, 11 has radial openings 36 near the bottom.

Furthermore, in the working piston housing 2, an upwardly open brake cylinder 12 is arranged, which is fixed by means of a securing ring on the working piston housing 2 in the axial direction. The brake cylinder 12 is arranged coaxially with the working piston 10, 11 and the brake cylinder 12 is partially arranged in the second piston part 11 such that the brake cylinder 12 is fixed by the second piston part 11 in the radial direction and the working piston 10, 11 is displaceable in the axial direction relative to the brake cylinder 12. The brake cylinder 12 is provided in its bottom surface with a throttle valve 21. The throttle valve 21 may be designed as a replaceable component or may be integrated in the brake cylinder 12. Instead of the throttle valve 21, only an opening or a bore can be provided. Furthermore, the brake cylinder 12 is provided with radial openings 35 in the region of its axial center and its upper end. The outer lower surface of the brake cylinder 12 defines, together with the inner surface of the second piston member 11, a fluid chamber 40.

Above the brake cylinder 12 define an inner peripheral surface of the damping ring 8, the control piston housing 3 and the end cover 5 a cavity, which is referred to as the control region 33. In the control region 33, a control piston is received axially displaceable.

The control piston consists in this embodiment of a piston ring 14 which is connected via an end sleeve 15 with a control piston shaft 13. The control piston may have any other structure and, for example, be integrally formed. The control piston is biased by a control piston spring 22 against the brake cylinder 12, that is, the control piston spring 22 is partially received in the brake cylinder 12 and supported on this. At rest, the control piston is urged by the control piston spring 22 in its upper end position, which is defined by an abutment against the end cap 5, as in Fig. 1 is shown. The lower end position of the control piston is fixed by a stop on the damping ring 8. The control spool shaft 13 is guided by the end cover 5 and another end cover 6 and is operatively coupled to an electromagnetic drive system 23 (eg, a solenoid, a magnetic vibration system, etc.) which is in Fig. 1 is indicated only schematically and is flanged to the end cap 6 in this embodiment. By means of the electromagnetic drive system 23, the control piston between its two end positions can be moved axially back and forth.

A connection ring 4 extends around the control piston housing 3 around. An intake passage 30 and an exhaust passage 31 are provided in the control piston housing 3 In particular, the inlet channel 30 consists of a radial connection between the control region 33 and an intermediate volume 37, which is bounded radially by the connecting ring 4 and connected via this to a not shown fluid supply system. The radial connection is designed here as a circumferential groove. Topologically identical, the outlet channel 31 arranged below the inlet channel 30 is constructed with its intermediate volume 38. Fig. 3 shows a section through the outlet channel 31 and the intermediate volume 38 (in this embodiment, eight intermediate volumes 38 are provided).

A hydraulic fluid (eg, a hydraulic incompressible fluid) supplied from the fluid supply system may be supplied to the work area 32 and the control area 33 via the inlet passage 30, with fluid communication between the control area 33 and the work area 32 via axial passages in the control piston (shown in FIG Piston ring 14) takes place in Fig. 3 are shown. Via the outlet channel 31, the fluid can be discharged from the working area 32 and the control area 33.

The control piston is closed by means of the electromagnetic drive system 23 between a rest position in which the control piston closes (radially seals) the inlet channel 30, while the outlet channel 31 is open or while it releases the outlet channel 31 (this state is in FIG Fig. 1 shown), and a working position to reciprocate, in which the control piston closes the outlet channel 31 (radially seals), while the inlet channel 30 is opened or while it releases the inlet channel 30 (this state is, for example, in Fig. 4B shown later). Movement of the spool to the working position causes a pressure increase inside the housing due to inflow of hydraulic fluid through the inlet passage 30, and movement of the spool to the rest position causes a pressure drop within the housing due to leakage of hydraulic fluid from the exhaust passage.

In operation, the working area 32 and the control area 33 are filled with a hydraulic fluid, the pressure of which depends on the position of the control piston varied. In particular, the pressure of the fluid supply system is higher than a residual pressure in the working area 32 and control area 33 at standstill of the valve train or the system. The residual pressure is hereinafter referred to as the outlet pressure, the pressure in the fluid supply system as the working pressure. Output pressure and working pressure correspondingly form the minimum value and the maximum value of the system pressure.

As described above, the valvetrain is attached to a cylinder head 1 which accommodates a valve mechanism which essentially comprises a charge exchange passage 34, a valve 16 with which the power piston 10, 11 is in contact (in this embodiment, the power piston 10, 11 on the valve 16, ie, it is non-positively connected to the valve 16, but it may also be positively connected to the valve 16), a valve seat 17, a valve guide 18, a valve spring 19 and a valve plate 20 has. Such embodiments are known from the prior art and therefore a structure is not explained in detail.

The valve train according to the invention is provided in this embodiment with four hydraulic brakes, which serve to avoid shocks and noise and will be described below. It should be noted here that it is sufficient if at least the first hydraulic brake is provided.

The first hydraulic brake decelerates a movement of the control piston from the rest position to the working position by increasing a flow resistance of a displaced from the control piston hydraulic fluid in the movement end phase.

In this embodiment, as in particular in Figs. 5A and 5B is shown, the first hydraulic brake substantially by a formed on the control piston circumferential step (formed by the in Fig. 5B shown surfaces 14a and 14b) and a brake channel 42 formed in the damping ring 8 is formed. The brake channel 42 consists of an axially extending bore which opens upwards into the control region 33, as well a radially extending bore communicating therewith which opens radially inwardly into the control region 33 and which is sealed radially outwardly by a plug screw 9. The surface into which the radially extending bore opens (brake channel surface) is a stop surface for the control piston, against which the control piston rests in its working position. Although in this embodiment, the brake passage 42 is formed in the damper ring 8, the brake passage 42 may be formed in the housing or a member connected to the housing. Although only one brake passage 42 is illustrated, a plurality of brake passages 42 may be provided along the circumferential direction.

The second hydraulic brake brakes movement of the control piston from the working position to the rest position by increasing a flow resistance of a hydraulic fluid displaced by the control piston in the movement end phase.

In this embodiment, as in particular in Figs. 5C and 5D is shown, the second hydraulic brake substantially by a circumferential shoulder of the control piston (the shoulder, the in Fig. 5D formed surface 13 a) and a second brake channel 43 is formed, which is formed in the housing. The brake channel 43 consists of an axially extending bore, which opens on an inner wall surface 5b of the end cover 5 down into the control portion 33, and a radially extending bore connected thereto, the radially extending at an axially extending wall surface 5a of the end cover 5 inside in the control portion 33 opens and which is sealed radially outward by a screw plug 9. Although in this embodiment, the brake passage 43 in the housing (in particular the end cover 5) is formed, the brake passage 43 may also be formed in a component connected to the housing. Although only one brake passage 43 is shown, a plurality of brake passages 43 may be provided along the circumferential direction.

The third hydraulic brake brakes movement of the working piston 10, 11 due to a pressure increase inside the housing by increasing a flow resistance one of the working piston displaced hydraulic fluid in the Arbeitskolbenbewegungsendphase from.

In this embodiment, as in particular in Figs. 5E and 5F is shown, the third hydraulic brake substantially by a on the working piston 10, 11 rotating stage (formed by the in Fig. 5F shown surfaces 11a and 11b) and a third brake channel 41 is formed. The third brake channel 41 is formed in the damping ring 7. The brake channel 41 consists of an axially extending bore which opens upwardly into a fluid-filled space below the working piston 10, 11 and a radially extending bore radially inwardly thereof, which opens radially inwardly into the space and radially outwardly is sealed externally by a screw plug 9. A surface 7 a of the damping ring 7, in which the axially extending bore opens, is a stop surface for the working piston 10, 11. Although in this embodiment, the brake channel 41 is formed in the damping ring 7, the brake channel 41 in the housing or a be formed with the housing connected component. Although only one brake passage 41 is illustrated, a plurality of brake passages 41 may be provided along the circumferential direction.

The fourth hydraulic brake decelerates movement of the working piston 10, 11 due to a pressure drop inside the housing by increasing a flow resistance of a hydraulic fluid displaced by the working piston 10, 11 in the working piston movement end phase. In this embodiment, the fourth hydraulic brake is basically constituted by the first piston part 11 and the brake cylinder 12 defining the fluid chamber 40 (see FIG Fig. 1 ).

Hereinafter, an operation of the valve gear according to the invention with reference to FIGS. 4A to 4D described.

In the neutral initial state ( Fig. 4A ), the electromagnetic drive system 23 is inactive. The control piston is held by its biased control piston spring 22 in its rest position in which it closes the inlet channel 30. In this state, the control area 33 and the work area 32 are filled with hydraulic fluid, the system pressure (pressure in the control area 33 and work area 32) corresponds to the outlet pressure. The valve 16 is urged by the prestressed valve spring 19 in its closed position in which the valve 16 rests against the valve seat 17. The working piston 10 is due to the force of gravity and the residual pressure or output pressure in the working area 32 to the valve 16, without this to operate or open.

If the electromagnetic drive system 23 is actuated in this state, it exerts an axial force on the control piston shaft 13. By this force, the control piston is axially displaced to its end position, in Fig. 4B is shown and which characterizes his working position. Before the spool reaches its end position, a braking effect is exerted by the first hydraulic brake.

The operation of the first hydraulic brake will be described with reference to figures Figs. 5A and 5B described, wherein Fig. 5A the not yet activated state and Fig. 5B shows the activated state of the first hydraulic brake. In fact, in the final phase of its movement, the control piston comes from the position which is in Fig. 5A is shown in the position in Fig. 5B is shown. Here, in the axial direction of the control piston (the control piston ring 14) extending surface 14a of the peripheral stage of the control piston with an axially extending wall surface of a portion in which the brake passage 42 is formed (wall surface of the damping ring 8 in this embodiment), in overlap (please refer Fig. 5B ). As a result, the axially extending surface 14a of the orbiting step separates together with the circumferential plane 14b extending in the radial direction of the control piston, an axially extending wall surface 3a of the control piston housing 3, and a radially extending brake channel surface into which the brake passage 42 penetrates opens (upper surface 8 a of the damping ring 8 in this embodiment), a part of the displaced from the control piston hydraulic fluid from the rest of the displaced hydraulic fluid from (the separated part is in Fig. 5B - compared to Fig. 5A - shown darker); ie a kind of closed volume is formed. This separated hydraulic fluid part is then pushed in the further course of the downward movement of the control piston through the brake passage 42, which is a throttle, whereby the braking effect is generated. The braking effect is thus achieved by increasing the flow resistance of the displaced from the control piston hydraulic fluid.

By the movement of the control piston from its rest position to its working position, the inlet channel 30 is released and the outlet channel 31 is closed, and hydraulic fluid with the working pressure flows through the inlet channel 30 a. As a result, the system pressure in the control area 33 and the work area 32 increases to the level of the working pressure. The now higher system pressure generates a force normal to the bottom of the working piston 10, 11 (a force directed in the axial direction). As soon as this normal force exceeds the pretensioning force of the valve spring 19, an axial movement of the valve assembly, consisting of valve 16 and valve disk 20, begins downward. Thus, the valve opening begins and the available volume in the working area 32 becomes larger, whereby fluid flows through the inlet channel 30 and subsequently through the axial openings of the control piston.

At the beginning of the movement of the working piston 10, 11 down the lateral radial openings 36 of the second piston member 11 are covered or closed by the brake cylinder 12 and a hydraulic fluid passes only through the throttle valve 21 into the fluid chamber 40, whose volume increases. This leads to a throttling effect and slows down the movement of the working piston 10, 11. As a result of a further movement of the working piston 10, 11 downwards, the radial openings 36 are released from an intermediate position, so that hydraulic fluid through both the throttle valve 21 and the radial openings 36 flows into the fluid chamber 40. Thus, the movement of the working piston 10, 11 can be continued with minimized resistance. The end position of the movement of the working piston 10, 11 and thus the valve assembly is determined downwards by the axial contact with the damping ring 7. Before this is achieved, a braking effect is exerted by the third hydraulic brake.

The operation of the third hydraulic brake will be described with reference to figures Figs. 5E and 5F described, wherein Fig. 5E the not yet activated state and Fig. 5F shows the activated state of the third hydraulic brake. In fact, the working piston 10, 11 arrives in the final phase of its movement from the position which is in Fig. 5E is shown in the position in Fig. 5F is shown. Here, in the axial direction of the working piston 10, 11 extending surface 11a of the peripheral step reaches with an axially extending wall surface of the damping ring, in which the third brake channel 41 is formed, in registration (see Fig. 5F ). Thereby, the circumferential surface 11a of the revolving stage extending in the axial direction of the working piston 10, 11 separates together with the peripheral surface 11b of the revolving stage extending in the radial direction of the working piston 10, 11, an axially extending wall surface 2a of the working piston housing 2 and one in the radial direction the working piston 10, 11 extending brake channel surface into which the brake passage 41 opens (surface 7a of the damping ring 7 in this embodiment), a part of the displaced from the working piston 10, 11 hydraulic fluid from the remaining displaced hydraulic fluid from (the separated part is in Fig. 5F - compared to Fig. 5E - shown darker); ie a kind of closed volume is formed. This separated hydraulic fluid part is then pushed through in the course of the downward movement through the brake passage 41, which is a throttle, whereby the braking effect is generated. The braking effect is thus achieved by increasing the flow resistance of the displaced by the working piston 10, 11 hydraulic fluid. The hydraulic fluid forced through the brake passage 41 flows back through a center hole formed in the cylinder head 1 and a drain hole connected thereto to the fluid supply system.

Subsequently, the working piston 10, 11 comes with the damping ring in abutment (see Fig. 4C ). In this position, the valve 16 is open to the maximum.

To close the valve 16 again, the electromagnetic drive system 23 is controlled such that the axial force on the control piston shaft 13 is eliminated. On the control piston thus acts only the force of the control piston spring 22, the control piston from its working position to its rest position moved back. Before the control piston reaches its rest position, a braking effect is exerted by the second hydraulic brake.

The operation of the second hydraulic brake will be described with reference to figures Figs. 5C and 5D described, wherein Fig. 5C the not yet activated state and Fig. 5D shows the activated state of the second hydraulic brake. In fact, in the final phase of its movement, the control piston comes from the position which is in Fig. 5C is shown in the position in Fig. 5D is shown. In this case, the radially extending surface 13a of the shoulder reaches the axially extending wall surface 5b of the end cover 5 in registration (see Fig. 5D ). As a result, the radially extending surface 13a of the shoulder, together with the axially extending surface 13b of the control piston (the control spool shaft 13) and the inner wall surface 5a of the end cover 5 into which the second brake channel 43 opens radially inwardly, separates the part of the displaced one Hydraulic fluid from the remaining displaced hydraulic fluid (the separated part is in Fig. 5D - compared to Fig. 5C - marked darker); ie a kind of closed volume is formed. This separated hydraulic fluid part is then pushed through in the further course of the downward movement of the control piston through the brake passage 43, which is a throttle, whereby the braking effect is generated. The braking effect is thus achieved by increasing the flow resistance of the displaced from the control piston hydraulic fluid.

By the movement of the control piston from its working position to its rest position, the control piston closes the inlet channel 30 and releases the outlet channel 31 (this state is in Fig. 4D shown), whereby the hydraulic fluid can flow through the outlet channel 31 back to the fluid supply system. As a result, the system pressure in the control area 33 and the work area 32 decreases to the level of the outlet pressure. The normal force on the bottom of the working piston 10, 11 is eliminated and the force of the valve spring 19 initiates an axial movement of the valve assembly and the working piston 10, 11 upwards, whereby the valve 16 is closed.

In the final phase of this closing movement, the fourth hydraulic brake is activated in order to decelerate the movement of the working piston 10, 11 and of the valve 16, so that a gentle or low-noise closing of the valve 16 is achieved. Namely, the working piston 10, 11 passes during the closing operation from the position in Fig. 4D is shown in the position in Fig. 4A is shown. By this movement of the working piston 10, 11, the volume of the fluid chamber 40 decreases. At the beginning of the movement, hydraulic fluid can flow out of the hydraulic chamber through both the radial openings 36 of the second piston member 11 and through the throttle valve 21. In the final phase of the movement, however, the radial openings 36 are closed by the brake cylinder 12, so that the hydraulic fluid from the fluid chamber 40 can only escape through the throttle valve 21. In this way, the fourth hydraulic brake develops its braking effect. The throttling effect can be adjusted as desired by appropriately selecting the cross section and the arrangement of the throttle valve 21 and the radial openings 35 and 36. With the fourth hydraulic brake, the impact velocity of the valve 16 on the valve seat 17 and an associated noise and wear can be minimized. A secure sealing of a combustion chamber of the internal combustion engine with respect to the charge exchange channel 34 is ensured.

After completion of this movement, the valve 16 is closed again (it is applied to the valve seat 17). This condition is in Fig. 4A (When used in an internal combustion engine, this corresponds to the time between two charge changes or the complete engine standstill). The above cycle can be arbitrarily repeated by re-driving the electromagnetic drive system 23.

The valve train described above is preferably used in internal combustion engines for cars and trucks. However, it is also suitable for use in stationary and marine internal combustion engines.

Second embodiment

The structural design of the hydraulic valve drive according to the invention for a cylinder valve of an internal combustion engine according to a second embodiment will now be with reference to Fig. 6 explained. Only changes compared to the first embodiment will be explained. Components that are not described and in Fig. 6 are not identified are identical to the first embodiment.

In this embodiment, a balancing member 151 is provided. The balancing member 151 is an upwardly opened cylinder and has a bottom surface above the brake cylinder 112 abutting the brake cylinder 112. The bottom surface projects radially further than an outer wall of the cylinder of the balance member 151. A control piston spring 122 is supported on this projecting portion of the bottom surface.

An internal volume of the compensation component 151 is bounded radially by its inner wall surface and axially by its bottom surface and by an end sleeve 115. The end sleeve 115 substantially corresponds to the end sleeve 15 of the first embodiment, but has a continuous axial bore. The end sleeve 115 is received axially displaceable in the compensation component 151. A seal 150 is provided at the upper axial end of the balance member 151 to seal an air-filled internal volume of the balance member 151 against the fluid-filled control portion 33.

A control piston shaft 113, which is formed with a continuous axial bore, is received axially displaceably in the compensation component 151, otherwise corresponds to the control piston shaft 13 of the first embodiment. The continuous axial bore makes it possible to discharge the air in the inner volume of the compensation component 151, with a movement of the control piston shaft 113 with little resistance from the hydraulic valve train. Thus, it is avoided that the air in the internal volume affects the effect of the first hydraulic brake.

Furthermore, the brake cylinder 112 is provided with radial openings 135 in the region of its axial center and its upper end. In contrast to the Radial openings 35 of the first embodiment are in the second embodiment, the radial openings 135 at the upper end of the brake cylinder 12 is greater than the radial openings in the axial center of the brake cylinder 12. By this construction, an increase in the flow resistance of the hydraulic fluid is avoided, the the fluid chamber 40 escapes through the throttle valve 21. In this way, the braking effect of the fourth hydraulic brake is not influenced by the bottom surface of the balancing component 151.

On the control piston act the axial force generated by the electromagnetic drive system 23, and the force of the control piston spring 22. In addition, a compressive force acting on an upper side and on an underside of belonging to the control piston piston ring 14. This pressure force thus affects the total force of the control piston.

An effective area at the top of the piston ring 14 (surface on which a fluid pressure inside the housing acts) can be determined by the cross section of the control piston shaft 113 at the level of a sealing member, above which the pressure inside the housing can not act the entire surface of the piston ring 14 is withdrawn. Similarly, an effective area on the underside of the piston ring 14 (surface on which the fluid pressure acts from below) can be determined by the cross section of the ferrule 115 at the level of the seal 150, below which the pressure inside the housing can not act is subtracted from the entire surface of the piston ring 14.

The provision of the compensation component 151 represents a possibility that the active surface at the top of the piston ring 14 substantially corresponds to the active surface on the underside of the piston ring 14. Thus, a reaction force could be avoided, which would be caused by a different sized effective area.

In addition, in this embodiment, a connection ring 104 is provided, which is also like the connection ring 4 of the first embodiment extends around the control piston housing 3 around. However, in this embodiment, the connection ring 104 is formed in several parts. As a result, an assembly of the connection ring is simplified.

Third and fourth embodiment

The structural design of the hydraulic valve train according to the invention for a cylinder valve of an internal combustion engine according to a third embodiment will be with reference to Fig. 7A explained. The structural design of the hydraulic valve drive according to the invention for a cylinder valve of an internal combustion engine according to a fourth embodiment will be with reference to Fig. 7B explained. In this case, the third and the fourth embodiment serve as examples of possible modifications with respect to the first or the second embodiment. The following modifications may thus be combined with the teachings of the first and second embodiments.

While the working piston 10, 11 is constructed in two parts in the first and the second embodiment of a first piston part 10 and a second piston part 11, which are positively connected to each other, the working piston 310, 311 in the third embodiment of Fig. 7A composed of a first piston part 310 and a second piston part 311, which are not positively connected to each other, but positively. The second piston member 311 is biased by a piston member spring 350 against the outer lower surface of the brake cylinder 12, whereby an investment in the first piston member 310 is ensured during operation. By this configuration, a structure of the hydraulic valve train is simplified.

While the end position of the working piston 10, 11 in the first and the second embodiment remains upward indefinite in order to compensate for thermal expansion of a valve 16 received in the cylinder head 1 is in the fourth embodiment of Fig. 7B a hydraulic lash adjuster 452 is provided between a valve 416 and a power piston 410,411. The hydraulic lash adjuster 452 is, for example, a hydraulic bucket tappet, which is conventionally an automatic tappet Balancing the valve clearance caused hydraulically. By virtue of this configuration, thermal expansion of a valve 416 accommodated in the cylinder head 1 can be compensated, although the end position of the working piston 410, 411 is determined. As a result, the valve 416 is not biased by the valve spring 22 against the valve seat 17. Thus, a load on the valve 416 and the valve seat 17 is minimized.

Claims (15)

Hydraulic valve train for a cylinder valve (16) of an internal combustion engine, comprising: a housing (2, 3, 5, 6) having an inlet channel (30) and an outlet channel (31), a control piston (13, 14, 15) which is arranged in the housing (2, 3, 5, 6) and between a rest position in which it closes the inlet channel (30), while the outlet channel (31) is opened, and a working position in which it closes the outlet channel (31), while the inlet channel (30) is opened, wherein a movement of the control piston (13, 14, 15) in the working position, a pressure increase inside the housing (2 , 3, 5, 6) due to inflow of hydraulic fluid through the inlet channel (30) and movement of the control piston (13, 14, 15) to the rest position causes a pressure drop in the interior of the housing (2, 3, 5, 6) causing outflow of hydraulic fluid from the outlet channel (31), a working piston (10, 11) disposed in the housing (2, 3, 5, 6), which is operatively connected to the cylinder valve (16) and the back and forth in dependence on the pressure rise and the pressure drop inside the housing moved to open and close the cylinder valve (16), and a hydraulic brake (13, 14, 15, 42), the movement of the control piston (13, 14, 15) from the rest position to the working position by increasing a flow resistance of a of the control piston (13, 14, 15) displaced hydraulic fluid in the Motor end phase decelerates. The hydraulic valve train according to claim 1, wherein the flow resistance is increased by pushing a part of the displaced hydraulic fluid from the spool (13, 14, 15) through a brake passage (42) in the end of travel phase. Hydraulic valve train according to claim 2, wherein a peripheral step is formed on the control piston (13, 14, 15), an axis of the control piston (13, 14, 15) extending surface of the revolving stage in the movement end phase with a in the axial direction the control piston (13, 14, 15) extending wall surface of a portion in which the brake passage (42) is formed, in overlap and thereby in cooperation with a in the radial direction of the control piston (13, 14, 15) extending surface of the peripheral step , a in the axial direction of the control piston (13, 14, 15) extending wall surface of the housing (2, 3, 5, 6), and a in the radial direction of the control piston (13, 14, 15) extending the brake channel surface into which the brake channel ( 42), separates the part of the displaced hydraulic fluid from the remaining displaced hydraulic fluid, and the separated part of the displaced hydraulic fluid is pushed through the brake passage (42). Hydraulic valve train according to claim 3, wherein the brake channel surface is a stop surface for the control piston (13, 14, 15) against which the control piston (13, 14, 15) rests in the working position. Hydraulic valve drive according to one of claims 2 to 4, wherein the brake channel in a damping ring (8) which is arranged between housing parts (2, 3) of the housing (2, 3, 5, 6), or in the housing (2, 3 , 5, 6) is formed. Hydraulic valve train according to one of claims 1 to 5, wherein the hydraulic valve train further comprises a second hydraulic brake (13, 14, 15, 43), the movement of the control piston (13, 14, 15) from the working position to the rest position by Increasing a flow resistance of a of the control piston (13, 14, 15) displaced hydraulic fluid in the movement end phase decelerates. Hydraulic valve train according to claim 6, wherein in the second hydraulic brake, the flow resistance is increased by the fact that in the end of motion portion of the displaced hydraulic fluid from the control piston (13, 14, 15) by a second brake channel (43) is pressed. Hydraulic valve train according to claim 7, wherein the control piston (13, 14, 15) is provided with a circumferential shoulder, a surface of the shoulder extending in the radial direction of the control piston (13, 14, 15) in the movement end phase coincides with a wall surface of the housing (2, 3, 5, 6) extending in the axial direction of the control piston (13, 14, 15) and thereby in cooperation with a in the axial direction of the control piston (13, 14, 15) extending surface of the control piston (13, 14, 15) and an inner wall surface of the housing (2, 3, 5, 6), in which the second brake channel ( 43) opens, part of the displaced hydraulic fluid separates from the remaining displaced hydraulic fluid, and the separated part of the displaced hydraulic fluid is forced through the second brake passage (43). Hydraulic valve train according to claim 7 or 8, wherein the second brake channel (43) in the housing (2, 3, 5, 6) is formed. Hydraulic valve train according to one of claims 1 to 9, wherein the hydraulic valve train further comprises a third hydraulic brake (10, 41), the movement of the working piston (10, 11) due to a pressure increase inside the housing (2, 3, 5 , 6) decelerates by increasing a flow resistance of a hydraulic fluid displaced by the working piston (10, 11) in the working piston movement end phase. The hydraulic valve train according to claim 10, wherein in the third hydraulic brake (10, 41), the flow resistance is increased by pushing a part of the displaced hydraulic fluid from the working piston (10, 11) through a third brake passage (41) in the working piston movement end phase. Hydraulic valve train according to claim 11, wherein a circumferential step is formed on the working piston (10, 11), an area of the revolving stage in the working piston movement end phase extending in the axial direction of the working piston (10, 11) with a wall surface of a section extending in the axial direction of the working piston (10, 11), in which the third brake channel (41) is formed, passes into coincidence and thereby in cooperation with a in the radial direction of the working piston (10, 11) extending surface of the peripheral step, in the axial direction of the working piston (10, 11) extending wall surface of Housing (2, 3, 5, 6), and a in the radial direction of the working piston (10, 11) extending the brake channel surface into which opens the third brake passage (41), the part of the displaced hydraulic fluid from the remaining displaced hydraulic fluid separates, and the separated part of the displaced hydraulic fluid is pressed by the third brake passage (41). Hydraulic valve train according to claim 11 or 12, wherein the third brake channel (41) in a damping ring (7) via which the valve train is to be attached to a cylinder head (1) of the internal combustion engine, or in the housing (2, 3, 5, 6 ) is trained. Hydraulic valve train according to one of claims 1 to 13, wherein the hydraulic valve train further comprises a fourth hydraulic brake (11, 12), which movement of the working piston (10, 11) due to a pressure drop inside the housing (2, 3, 5 , 6) decelerates by increasing a flow resistance of a hydraulic fluid displaced by the working piston (10, 11) in the working piston movement end phase. Hydraulic valve train according to claim 14, wherein in the fourth hydraulic brake (11, 12) the flow resistance is increased by a maximum possible Hydraulikfluidausströmmenge from a fluid chamber (40), the working piston (10, 11) with a on the housing (2 , 3, 5, 6) and their volume decreases with movement of the working piston (10, 11) due to a pressure drop, is reduced in the Arbeitskolbenbewegungsendphase.

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