EP3406866A1 - Hydraulic drive for accelerating and braking components to be dynamically moved - Google Patents

Abstract

In order to ensure a simple, reliable and recuperative drive in a hydraulic drive (10) for accelerating and decelerating a gas exchange valve (20) of internal combustion engines or other reciprocating engines, it is proposed that a first pressure reservoir (41) for providing a first pressure p 1 , a resetting, preferably as a spring (25) formed restoring energy storage and at least one hydraulic base pressure reservoir (40) having a lower pressure p 0 than the first pressure reservoir (41) to provide. In a connecting line (48) between the first hydraulic pressure reservoir (41) and the working cylinder (22) a controllable opening (49) of a first valve (46) with at least one upstream in the flow path or behind check valve (47) is arranged, which the pressure medium (30) allows flow in the direction of the working cylinder (22), but prevents backflow toward the pressure reservoir (41). In order to initiate the closing movement or to enable the gas exchange valve to be hydraulically simply and reliably decelerated, a controllable opening (59) of a second valve (56) is provided in a second connecting line (58) between the first pressure reservoir (41) and the working cylinder (22) ) with non-return valve (57), which prevents flow in the direction of the working cylinder (22), but allows a return flow in the direction of the pressure reservoir (41).

Description

Technical area

The invention relates to a hydraulic drive for accelerating and decelerating dynamically moving components, in particular of valves in gas exchange controls of internal combustion engines and other piston engines.

State of the art

Variable valve timing on internal combustion engines is known as a suitable means of both improving torque versus speed performance, as well as improving overall engine efficiency and reducing emissions. The variety of optimization possibilities is described in the literature.

Today, a large variety of mechanical, electromechanical, pneumatic and hydraulic construction options for partially or fully variable valve controls are known, but mostly due to their high own energy consumption or due to high technical complexity and the associated manufacturing costs were only selectively enforce. Furthermore, many such systems do not have full variability; For example, opening time and opening duration or opening duration and opening stroke can be firmly coupled together, which can considerably limit the possibilities for optimizing the internal combustion engine or another piston engine. In particular, hydraulic systems can be built to save space due to their high energy density (SAE-1996-0581) and are therefore particularly suitable for variable valve timing on internal combustion engines, if it succeeds, both a low own energy consumption and a low system cost and to achieve high reliability.

• Freies, nämlich unabhängiges Einstellen von Öffnungs- und Schliesszeitpunkten, d.h. der sogenannten Steuerzeiten, von Einlass- und Auslassventilen, bei Bedarf auch zylinderselektiv. Beispielsweise kann über die Öffnungsdauer der Einlassventile die Luft- oder Gemischmenge gesteuert werden.
• Schnelles Öffnen und Schliessen der Ventile auch bei niedrigen Motordrehzahlen, also geringe Drosselverluste beim Gaswechsel.
• Von der Öffnungsdauer unbeeinflusste Steuerung bzw. Variationsmöglichkeit des Öffnungshubes, beim Einlassventil zum Beispiel zur gewollten Turbulenzerzeugung in der Frischgasmenge, beim Auslassventil zum Beispiel zur Erhöhung der Motorbremswirkung sowie bei beiden zum Beispiel zur Minimierung des Eigenenergie- bzw. Gesamtenergieverbrauchs.
• Unabhängiges und sicheres Schliessen zwecks Vermeidung von Verlusten und Vermeidung von Schäden durch ungeplantes Durchströmen heisser Gase, aber auch zur Vermeidung von Kollisionen der Gaswechselventile untereinander oder mit dem Kolben.
• Sichere Maximalhubbegrenzung zwecks Vermeidung von Kollisionen der Gaswechselventile untereinander oder mit dem Kolben.
• Elektronische Ansteuerbarkeit mit hoher Robustheit und geringem Aufwand an Sensorik und Aktorik.
• Sanftes Aufsetzen der Ventile beim Schliessvorgang.
• Abschalten einzelner Ventile oder Ventilgruppen, beispielsweise zwecks Drallerzeugung oder Zylinderabschaltung.

Depending on the task, the following control tasks can be performed on a fully variable valve control system on an internal combustion engine today:

• Free, namely independent setting of opening and closing times, ie the so-called timing of intake and exhaust valves, if necessary, also cylinder-selective. For example, the amount of air or mixture can be controlled via the opening duration of the intake valves.
• Quick opening and closing of the valves even at low engine speeds, so low throttle losses during gas exchange.
• From the opening duration uninfluenced control or variation possibility of the opening stroke, the intake valve, for example, the desired turbulence generation in the fresh gas, the exhaust valve, for example, to increase the engine braking effect and both, for example, to minimize the own energy or total energy consumption.
• Independent and safe closure to prevent losses and avoid damage due to unplanned flow of hot gases, but also to avoid collisions of the gas exchange valves with each other or with the piston.
• Safe maximum stroke limitation in order to avoid collisions of the gas exchange valves with each other or with the piston.
• Electronic controllability with high robustness and low cost of sensors and actuators.
• Gentle placing of the valves during the closing process.
• Switching off individual valves or valve groups, for example for the purpose of swirl generation or cylinder deactivation.

Hydraulic valve actuators, in particular for gas exchange valves in the working space of an internal combustion engine, are known per se for a long time, for example from the German Offenlegungsschrift 1'940'177 A. They were used as a replacement for the camshaft-controlled opening of a gas exchange valve, while the Close the valve was still provided by a spring mechanism. The provision of the gas exchange valves by means of spring means, usually in the form of helical compression springs, is still by far the most widely used closing method, since it ensures safe closure.

The aim of these systems was the optimization of the timing of the gas exchange valve and a steeper / faster opening and closing of the valves, with an optimization of the own energy consumption was usually not explicitly provided. A stroke adjustment was not provided in DE 1'940'177 A, but it was thought to damp hard abutment of the mechanical stroke limiter as well as the touchdown on the valve seat of the gas exchange valve by displacing the medium through a throttle cross-section.

To optimize the self-consumption of hydraulic valve actuators various "symmetrical pendulum systems" have been proposed, in which spring means are used for energy storage. DE 38 36 725 A shows a solution with mechanical spiral compression springs.
Typically, in such systems a valve mass clamped symmetrically between two springs oscillates about a central position. In the end (holding) positions, the energy is stored in the form of spring energy. This is converted into kinetic energy when the movement is generated, in order to be temporarily stored in the other end position in the form of spring energy.
In the end positions, respectively a capture or capture of the moving component must take place. Such symmetrical pendulum systems are inter alia complicated by the fact that the gas exchange valve to be driven must be brought into one of the respective end positions before starting. In addition, occur during engine operation by the gas pressures in particular at exhaust valves partly high unidirectional forces, which require unbalanced drive forces. Friction-related energy losses must be restored by the capture devices be supplemented.

In the WO 93/01399 A1 It is shown that even in systems with simple, one-sided spring return as in DE 1'940'177 A, a minimization of the own energy consumption is possible. The kinetic kinetic energy that results from the hydraulic drive is temporarily stored in compression work of the one-sided, resetting spring accumulator before it is used again for the closing movement.
Therefore, this principle can also be called an "asymmetric pendulum system". A disadvantage of the proposal WO 93/01399 A1 is, for example, that in each case one of the actuating movements of the controlling hydraulic valve takes place in the middle of the movement phase, namely while the drive piston of the gas exchange valve moves at high speed and a high volume flow flows through the hydraulic valve. In order to avoid high throttle losses in this situation, the controlling valve must be very fast. Likewise, in the opening endpoint of the gas exchange valve movement, for example, it must switch precisely and reliably, so that the kinetic energy can be captured to the full extent and retained in the spring. These requirements thus require very expensive, high-speed control valves and a complex control electronics.

Another such asymmetric pendulum system is described in SAE 2007-24-008. About the height of the hydraulic operating pressure of the opening stroke can be adjusted independently of the control period. In contrast to WO93 / 01399 A1 The system dispenses with high-speed switching operations of the hydraulic control valve in the middle of the movement phase. However, the actuating movement of the control valve as a whole must also be precisely coordinated with the movement of the gas exchange valve. The flow path for opening must close precisely when the gas exchange valve has delivered its kinetic energy to the return spring. Closes the control valve cross-section too early, the movement of the gas exchange valve is braked loss, closes it is too late, the gas exchange valve is already pushed back by the spring, is not held in the desired position, and is then decelerated in the return movement again with losses. For this high-precision, time-accurate motion control of the hydraulic control valve, a main spool is acted upon by a precisely defined volume flow of a pilot valve. For example, the pilot valve is fed by a separate constant pressure system to provide the defined volume flow for controlling the main valve. Deviations of the pilot volumetric flow due to wear or clogging of the pilot valve ports, however, have an effect on the speed of the main valve and thus on the quality of the time coordination with the drive piston or the gas valve movement.

US 4 009 695 A shows, among other things, the construction of a hydraulic valve drive by means of rotary valve control valves. The spool shafts run continuously with camshaft speed (half engine speed) in rotary valve sleeves; In this case, in the exemplary embodiment, the phase angles are adjusted with simple, comparatively slow worm drives in the angular phase, while the fast processes are automatically clocked by means of the rotating slide shaft. The motor can be operated in stationary operating points completely without control intervention; Adjustments are only displayed when changing an operating point. Such simple adjustment mechanisms can be carried out in principle even without control electronics. Unfortunately allows US 4 009 695 A No control of the Gaswechselventilhubs and does not recognize any possibility of recovery of hydraulically fed energy.

Presentation of the invention

The object of the invention is thus to provide a hydraulic drive for accelerating and decelerating dynamically moving components available, in which the above-mentioned disadvantages of the prior art need not be taken into account. The invention solves this problem by means of a hydraulic drive according to claim 1. It is clear that the present invention is particularly applicable to gas exchange controls of internal combustion engines and other piston engines. However, it results from the elements used that the drive according to the invention is quite generally advantageous, ie also in other applications in which highly dynamic masses have to be moved.

The invention presented here works - like the other aforementioned "asymmetric pendulum systems" - also with simple, one-sided restoring energy storage or spring means and with the described energy conversions. In this case, the control is designed so advantageous that scatters in speed, precision and uniformity of the control valves hardly affect the hydraulic losses of the drive, so that it can be constructed in return of simple and robust elements.
Therefore, a true fully variable hydraulic drive system for gas exchange valves or other highly dynamically moving masses is shown, which keeps the own energy consumption minimal and yet simple and reliable.

The invention is also well suited for triggering with rotary valves US 4 009 695 A , The full variability of the opening and closing times of the gas exchange valves is maintained, a stroke control is possible via pressure level and the own energy consumption is minimized due to energy recovery.

The advantageous embodiments of the present invention are already partially named above, partially defined in the dependent claims.

The above-mentioned as well as the claimed and described in the following embodiments, according to the invention to be used elements are subject to their size, shape design, material usage and their technical conception no special conditions of exception, so that the well-known in the respective field of application selection criteria can apply without restriction.

Brief description of the drawings

Fig. 1

eine Ventilanordnung zu einem ersten Ausführungsbeispiel der vorliegenden Erfindung mit zwei 2/2-Wegeventilen, zwei Hochdruckniveaus und einem dritten 2/2-Wegeventil mit einer aktiv geschalteten Bremsdrossel;

Fig. 2

eine Ventilanordnung zu einem zweiten Ausführungsbeispiel der vorliegenden Erfindung mit einem Hochdruckniveau, einem 3/2-Wegeventil und einer automatischen hydraulisch zeitgesteuerten Bremsdrossel;

Fig. 3

eine Ventilanordnung zu einem dritten Ausführungsbeispiel der vorliegenden Erfindung mit einem 4/2-Wegeventil, zwei Hochdruckniveaus und einer automatischen druckgesteuerten Bremsdrossel;

Fig. 4

Schematische zeitliche Darstellung der Gaswechselventil-Bewegungsphasen und der Öffnungsverläufe der hydraulischen Steuerventile.

Fig. 5

eine Variante zum Ausführungsbeispiel 1 in ausschnitthafter Darstellung

Fig. 6

eine weitere Variante zum Ausführungsbeispiel 1 in ausschnitthafter Darstellung

Further details, advantages and features of the subject matter of the present invention will become apparent from the following description of the accompanying drawings, in which - exemplary - inventive devices will be explained. In the drawings shows:

Fig. 1

a valve assembly to a first embodiment of the present invention with two 2/2-way valves, two high pressure levels and a third 2/2-way valve with an actively switched brake throttle;

Fig. 2

a valve assembly to a second embodiment of the present invention having a high pressure level, a 3/2-way valve and an automatic hydraulically timed brake throttle;

Fig. 3

a valve assembly to a third embodiment of the present invention with a 4/2-way valve, two high pressure levels and an automatic pressure-controlled brake throttle;

Fig. 4

Schematic timing of the gas exchange valve movement phases and the opening curves of the hydraulic control valves.

Fig. 5

a variant of the embodiment 1 in a sectional representation

Fig. 6

a further variant of the embodiment 1 in a fragmentary representation

Ways to carry out the invention

In a first embodiment of the present invention is - as in Fig. 1 a gas exchange valve 20 for a motor both for opening and closing by means of a hydraulic drive 10 with a working cylinder 22 and a drive piston 23 and a force acting against the force of the drive piston spring 25 operated.
The hydraulic drive 10 may be divided into a core part 11 and a supply part 90 for ease of understanding. In the supply part, the pressure supply for the proposed pressure reservoir, in a conventional manner preferably with controllable pumps 91, 92, which can adjust the flow rate to the flow and pressure requirements. The control takes place in this example via pressure sensors 96 and control electronics 97. The control electronics also takes over the control of the actively electrically switching valves 46, 56 and 66. These valves are designed in this embodiment as a direct-operated, solenoid-operated 2/2-way valves, wherein the electrical connection lines are not shown for better clarity. The supply unit also includes a pressure limiting valve 99, which protects the system against overpressure and at the same time, as explained below, ensures that the gas exchange stroke does not reach a critical value. In the exemplary embodiment, a slightly raised base pressure p _{0 was} selected, for which reason a small pump 95 from a collecting tank 98 recirculates the leakage quantities of the pressure medium 30 fed into the self-contained system via a leak collecting line 94 from the spring chamber 93. An embodiment of the base pressure reservoir as a normal, ambient-ventilated tank is basically also possible, but the slightly elevated pressure has various advantages. For example, no pressure spring is required to bring the working piston in contact with the gas exchange valve 20. So you have an inherent valve clearance compensation.

Die hohe Federkonstante c bewirkt, dass die Federkraft F_F mit zunehmendem Öffnungshub h markant ansteigt. Sobald die hydraulische Kraft p₁xA auf den Antriebskolben 23 durch die Federkraft (und etwaige Zusatzkräfte) ausgeglichen ist (statischer Gleichgewichtspunkt), ist die Bewegung - statisch betrachtet - beendet, wobei das System aus bekannten physikalischen Gründen - in der bewegten Masse m gespeicherte kinetische Energie - zu einem Überschwingen tendiert, welches das Zweifache des statischen Hubes erreichen kann.
Für den statischen Hub h_stat gilt: $${\mathrm{h}}_{\mathrm{stat}}\left({\mathrm{p}}_{\mathrm{1}}\right)=\left({\mathrm{p}}_{\mathrm{1}}\times \mathrm{A}-{\mathrm{F}}_{\mathrm{Fv}}\right)/\mathrm{c}$$

Dynamisch kann der zweifache statische Hub erreicht werden: $${\mathrm{h}}_{\max }\left(\mathrm{p}\mathrm{1}\right)=\mathrm{2}\times {\mathrm{h}}_{\mathrm{stat}}\left({\mathrm{p}}_{\mathrm{1}}\right)$$

bzw. $${\mathrm{h}}_{\max }\left({\mathrm{p}}_{\mathrm{1}}\right)=2\times \left({\mathrm{p}}_{\mathrm{1}}\times \mathrm{A}-{\mathrm{F}}_{\mathrm{Fv}}\right)/\mathrm{c}$$

erreicht werden.The phases of the movement sequence and the associated valve openings are in Fig. 4 shown.
In the idle state - phase 0, gas exchange valve closed - the so-called third valve 66 is open and the working cylinder 22, in which the drive piston 23 is arranged with pressure acting surface 24 of the surface A movably connected to the base pressure reservoir 40 at the pressure level p ₀ . The biasing force F _{Fv of} the spring 25 in the idle state (drive or Gaswechselventilhub h = 0) is so dimensioned that the gas exchange valve with respect to the opening force of the product p ₀ x A, but also with respect to other opening forces, for example on the plate 21 of the gas exchange valve 20 attacking by negative pressure in the engine cylinder 15 or pressure in the gas exchange channel 16, safely in the closed rest position remains or can move back reliably there, even with expected frictional forces, such as valve stem seal 17 or valve guide 19th
It should be noted here that the said forces acting vary depending on the operating point and use (type of internal combustion engine or the piston engine, intake or exhaust valve) and can also change direction. A short time before the planned opening of the gas exchange valve, the relieving valve 66 is closed.
To open the gas exchange valve 20 (phase I), the hydraulic pressure force here from a first pressure reservoir with the pressure p ₁ , via a first 2/2-way valve 46 and a first check valve 47, the drive piston 23 and its Druckwirkfläche 24 of the surface A acted upon , The gas exchange valve 20 begins to open as soon as the hydraulic pressure force p ₁ x A exceeds the biasing spring force F _{Fv of} the spring 25.
It will be appreciated that the actual force at which opening occurs may vary according to the aforementioned additional forces involved. In the case of a small proportion, the additional forces in the following formulas are neglected or a corresponding substitute force can be used instead of F _Fv . Also, in the actual design due to flow losses and wave processes in the working cylinder, an effective pressure is set, which is not exactly the Pressure p ₁ corresponds. This too can be taken into account by correction values. In the exemplary embodiment, the spring 25 used as energy storage is designed with a high spring constant c, so that a rapid movement of the mass is achieved. The time for full opening corresponds approximately to half the period T _{1/2 of} a vibration of the mass-spring oscillator, formed from the effective mass m, namely the mass of gas exchange valve 20, spring plate, drive piston 22, possibly valve bridge, mass fraction of the spring 25 and of resonant pressure medium 30, and the spring 25 with spring constant c: ie: $${\mathrm{<figure-callout\; id="1"\; label="T"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">T</figure-callout>}}_{\mathrm{1}/\mathrm{2}}=\mathrm{\pi }\times \mathrm{square\; root}\left(\mathrm{m}/\mathrm{c}\right)$$

The high spring constant c causes the spring force F _F increases markedly with increasing opening stroke h. Once the hydraulic force p ₁ xA on the drive piston 23 by the spring force (and any additional forces) is balanced (static equilibrium point), the movement - statically considered - is completed, the system for known physical reasons - stored in the moving mass m kinetic Energy - tends to overshoot, which can reach twice the static lift.
For the static stroke h _stat : $${\mathrm{<figure-callout\; id="1"\; label="H\; stat\; p"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">H\; </figure-callout>}}_{\mathrm{stat}}\left({\mathrm{p}}_{}\right)\left({\mathrm{}}_{\mathrm{1}}\right)=\left({\mathrm{<figure-callout\; id="1"\; label="p"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{1}}\times \mathrm{A}-{\mathrm{F}}_{\mathrm{Fv}}\right)/\mathrm{c}$$

Dynamically the double static stroke can be achieved: $${\mathrm{<figure-callout\; id="1"\; label="H\; Max\; p"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">H\; </figure-callout>}}_{\mathrm{Max}}\left(\mathrm{p}\right)\left(\mathrm{}\mathrm{1}\right)=\mathrm{2}\times {\mathrm{<figure-callout\; id="1"\; label="H\; stat\; p"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">H\; </figure-callout>}}_{\mathrm{stat}}\left({\mathrm{p}}_{}\right)\left({\mathrm{}}_{\mathrm{1}}\right)$$

respectively. $${\mathrm{<figure-callout\; id="1"\; label="H\; Max\; p"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">H\; </figure-callout>}}_{\mathrm{Max}}\left({\mathrm{p}}_{}\right)\left({\mathrm{}}_{\mathrm{1}}\right)=2\times \left({\mathrm{<figure-callout\; id="1"\; label="p"\; filenames="imgf0004.png,imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{1}}\times \mathrm{A}-{\mathrm{F}}_{\mathrm{Fv}}\right)/\mathrm{c}$$

be achieved.

It can easily be seen from the formula that a desired stroke h _max can also be controlled via the magnitude of the pressure p ₁ via the magnitude of the force F _Fv . So even a stroke control is possible in two ways.

Thus, for example, collisions of the gas exchange valve with the piston or with other valves are avoided, the maximum desired stroke on the maximum pressure p ₁ in a known and reliable manner by means of a Pressure limiting valve to be ensured, provided in the embodiment with the pressure relief valve 99th

With the use of a spring 25 with progressive spring characteristic, the stroke control can be refined in the small stroke range, while the protection against too large stroke is accordingly robust.
The skilled artisan also recognizes that such a progressive spring can also be represented very well as a pneumatic spring. He also recognizes that the setting of the biasing force F _Fv in a pneumatic spring by adjusting its pneumatic biasing pressure is possible in a particularly simple manner. the pneumatic preload pressure is adjusted accordingly. It is clear that equations 1 to 4 must be adapted appropriately if a progressive spring is used instead of a linear spring with fixed spring constant c.

By the first check valve 47, which prevents backflow of pressure medium in the direction of the pressure reservoir, the gas exchange valve 20 now remains in its open position, even if the 2/2-way valve has not yet closed. This starts the holding phase (phase II) of the gas exchange valve. Only a minimal backward movement (closing movement) of the gas exchange valve due to the compression of the pressure medium itself - caused mainly by its low compressibility - will be observed. Thus, the gas exchange of the engine can now continue with the desired stroke.
As a precaution, it should be noted that any other flow branches or leakage paths must be prohibited or closed on the flow path between the working cylinder 22 and the check valve since these would impair the holding function. Since the non-return valve has taken over the blocking function, the 2/2-way valve 46 can now be closed in a comparatively wide time range, without it coming to the exact closing time. Fig. 4 shows three exemplary for the valve opening 49 Cross-sectional curves of the valve opening 49: A _1a , A _1b and A _1c , which are all possible in the embodiment. The opening of the flow cross-section of the switching valve 46 must only take place about as fast as the movement of the gas exchange valve expires. So no expensive and expensive valve principle is needed. In addition, the check valve 47 automatically ensures that the kinetic energy of the moving mass is almost completely converted into spring energy and also cached in the spring 25 - which would be achievable both with an active control intervention of valve 46 only at great expense.
It should be noted that in this phase in the working cylinder 22 sets a pressure which - is higher than the pressure p ₁ - as a result of the overshoot and the stored spring energy - usually

Out Fig. 1 also the closing process of the gas exchange valve 20, phase 3, by means of another part of the hydraulic drive can be seen. For this purpose, the second 2/2-way valve 56 is opened. It should be pointed out to those skilled in the art that this second 2/2-way valve has hitherto been closed (in phases I and II) ( Fig. 4 , Course A ₂ ). Valve 56 is connected to a second pressure reservoir 42 with a pressure p ₂ , which is generally lower than the pressure p ₁ , but higher than p ₀ . A hydraulic flow takes place in the pressure reservoir 42, while the drive piston 23 performs the closing movement ( Fig. 4 , Lift diagram, phase III). If now the pressure in the working cylinder 22 falls below the pressure p ₂ , the hydraulic reflux is terminated, by the second check valve 57, which is of course arranged in the other direction as the first check valve and a return flow from the pressure reservoir 42 into the working cylinder prevented. It thereby causes - similar to the check valve 47 when opening the gas exchange valve - that the gas exchange valve stops in the reached position, and that the 2/2-way valve must be closed later and at any time before the next gas exchange valve opening cycle ( Fig. 4 , A _2a , A _2b ). Above all, this automatism recuperates a maximum of energy.

Due to the lack of a need for a precise closure, the valve 56 can be easily constructed and the cost of the electronic control is reduced considerably. Also, the control valve 56 may in turn turn comparatively slowly, which can be dispensed with in many cases to elaborate design using eg vortex current-inhibiting magnet special materials.
Finally, it should be mentioned that the late closing of the use of rotary vane technology is very accommodating, since a different length of open the cross section is not disturbing.

In principle, it would be possible to dimension the pressure level p _{2 in} such a way that the gas exchange valve is closed exactly at this operating point, that is to say it has seated virtually at a speed close to zero on its seat. However, this is not very easy and especially in the case of an exhaust valve of an internal combustion engine, this operating point is not the same for all operating conditions. For this reason, in the in Fig. 1 illustrated embodiment, the pressure p ₂ designed so that the process of reflux through the second 2/2 way valve 56 is terminated in pressure reservoir 42 at a certain distance before the touchdown of the gas exchange valve 20 ( Fig. 4 , Transition phase III-IV).

The placement of the gas exchange valve 20 - ie the closing of the "stop" from the valve seat (phase V) - is in the in Fig. 1 illustrated embodiment, in that a third 2/2-way valve 66 opens a flow path from the working cylinder 22 to the base pressure reservoir by means of a connecting line 68 out. In series with this is a brake throttle 67, by means of which the speed of Aufsetzvorgangs can be controlled. The force for the safe closing and placing the gas exchange valve is obtained from the remaining energy of the spring 25, which is designed so that the closing force at the touchdown point, which is equal to the spring biasing force F _Fv is greater than the product of the pressure p ₀ x A and other opening forces, as described above.

The switching time of the third 2/2-way valve 66 ( Fig. 4 , A _V3 , beginning of phase V) determines the residence time in the holding phase near the valve seat (phase IV). Often, in internal combustion engines and other piston engines here no lingering desired; The closing process of a gas exchange valve should be completed quickly. Since the system represents a vibration system, the time duration of phase III (beginning of the closing movement of the gas exchange valve to the stop point) corresponds to approximately half the period T _{1/2 of} the spring-mass oscillator according to equation 1.
The electronic control can now be programmed so that the start of opening of 2/2-way valve 66 by T _{1/2 is} later than the beginning of opening of the 2/2 way valve 56. In many cases, the person skilled in the art will choose the duration slightly longer in order to be on the safe side with regard to maximum energy recovery.

Often, a particularly gentle placement of the gas exchange valves on the valve seats is desired for noise and wear reasons. The embodiment according to Fig. 1 For this purpose, it can be equipped with a travel-controlled braking device, such as cut-out in FIG. 5 shown. The connecting line 68 must be performed for this task separately from the other connecting lines 48 and 58 in the working cylinder 22, so that thus the transitional cross section 61 from the working cylinder into the connecting line 68 when approaching the working piston 23 to the position h = 0 or approach of the gas exchange valve 20 is closed to the valve seat 18 by the control edge 26 of the working piston so far that the gas exchange valve braked strong and gently moves into the seat. It is clear to the person skilled in the art that the transitional cross-section can be suitably shaped, for example with a notch-like contouring in the wall of the working cylinder, or as a bore or groove in the drive piston.

In FIG. 6 is shown in detail how the soft braking can be performed alernatively. Here, the connecting line 68 is divided into two ports 62 and 63, wherein the first port 62 is shut off at the latest in the vicinity of stroke zero, so shortly before placing the gas exchange valve 20 on the valve seat 18 by the control edge 26 of the drive piston 23, so that the Pressure medium only via port 63 and the throttle 64 can flow. This can also be arranged in the working piston.

Finally, the embodiment according to Fig. 1 advantageous to be performed with rotary valves. The 2/2-way valves 46, 56 and 66 are replaced by one rotary valve each. The adjustment takes place by means of the adjustment of the phase angle. Since, in the control of the flow paths 49 and 59, due to the inventive automatic holding function of the check valves 47 and 57 for each direction of movement in each case mainly arrives only at the opening time, while the closing time may be in a relatively wide range, it plays - at least within a certain limits - No matter if the closing time is also shifted as a result of the phase rotation. Thus, the invention allows to build a fully variable and energy-efficient hydraulic gas exchange valve drive with - the internal combustion engine - cycle synchronously running rotary valves.

• Beim Abschalten des Stellaktors 88, also beim Einleiten der Schliessphase des Gaswechselventils, wird durch Zurückziehen des Mitnehmers neben dem 3/2-Wegeventil auch die Rückstellung des Ventils 82 freigegeben.

Die Bewegung durch die Rückstellfeder 73 erfolgt jedoch langsam, da über eine Druckwirkfläche 71 des Ventils Druckmedium durch die Drossel 72 gepresst werden muss. Das hier parallel zur Drossel 72 angeordnete Rückschlagventil 74 sperrt in dieser Situation. Drossel, Druckwirkfläche und Federkraft sind so abgestimmt, dass erst nach der gewünschten Zeitverzögerung der Querschnitt 69 zum Basisdruckreservoir hin öffnet. Wiederum wird die Zeitverzögerung gegenüber der halben Periodendauer des Feder-Masse-Schwingers etwas grosszügiger gewählt. Dadurch liegt man bzgl. optimaler Energierückgewinnung auf der sicheren Seite, was durch die automatische Haltefunktion des Rückschlagventils 57 sichergestellt ist.
Wenn der Stellaktor ausgeschaltet wird, vollzieht das 3/2-Wegeventil 84, gesteuert durch seine Rückstellfeder, eine schnelle Bewegung in seine Ruhestellung 0. Das parallel geschaltete 2/2 Wegeventil 82 stellt sich jedoch langsam zurück, weil seine Rückstellbewegung durch die Drossel 72 gebremst ist. Die Öffnungsbewegung erfolgt aufgrund eines Rückschlagventils 74 ungebremst.According to the second embodiment Fig. 2 is only worked with a high pressure reservoir, namely pressure reservoir 41 with pressure p ₁ . As a result, p ₂ = p ₁ . This embodiment variant can be used to advantage with sufficient cross-sectional design of all hydraulic valves and connecting lines and friction-optimized design of the movable elements (drive piston 23 in the drive cylinder 22 and gas exchange valve 20 in the valve guide 19 with valve stem seal 17), as at low energy losses Return swinging occurs close to the valve seat. The construction cost is thereby smaller overall.
As a further simplification, the 3/2-way valve 84 is used, wherein the check valves 47 and 57 are arranged in this case between the 3/2-way valve and the pressure reservoir 41. The opening of the gas exchange valve (phase I) is initiated by switching on the Stellaktors 88, the keeping open (phase II) in a known manner by the check valve 47, the closing of the gas exchange valve is initiated by switching off the Stellaktors 88. Finally, the second holding phase takes place in the vicinity of the seat in a known manner by means of the check valve 57.
In another embodiment, the third valve 66 is formed as a hydraulically timed valve 86. In this case, it is mitbetätigt by a driver 87 of Stellaktors 88. This driver is designed so that when energizing the Stellaktors 88 first the valve cross-section 69 of the valve 82 is closed before the 3/2-way valve is significantly moved so that when opening the cross section 49 no unnecessary short circuit from the pressure reservoir 41 to the base pressure reservoir 40 is formed. This is achieved by the game 83 between driver and valve part of the 3/2-way valve.
The timing of the valve 82 now works as follows:

• When switching off the Stellaktors 88, ie when initiating the closing phase of the gas exchange valve, the provision of the valve 82 is released by retracting the driver in addition to the 3/2-way valve.

However, the movement by the return spring 73 is slow, since pressure medium must be pressed through the throttle 72 via a pressure-acting surface 71 of the valve. The check valve 74 arranged here parallel to the throttle 72 blocks in this situation. Throttle, pressure acting surface and spring force are adjusted so that only after the desired time delay, the cross section 69 opens to the base pressure reservoir. Again, the time delay compared to half the period of the spring-mass oscillator is chosen slightly more generous. As a result, one is in terms of optimal energy recovery on the safe side, which is ensured by the automatic holding function of the check valve 57.
When the Stellaktor is turned off, performs the 3/2-way valve 84, controlled by its return spring, a quick movement to its rest position 0. However, the parallel connected 2/2 way valve 82 is slowly back because its return movement braked by the throttle 72 is. The opening movement is unbraked due to a check valve 74.

In the third embodiment according to Fig. 3 the 4/2-way valve 86 is used. This is suitable for the use of two high-pressure levels. Furthermore, the third valve 66 is arranged in pressure-controlled design 80 in the connecting line 68 between the working cylinder and the base pressure reservoir. The valve 80 uses the effect that the gas exchange valve 20 during the transition from phase III to phase IV similar to the transition from phase I to II springs back slightly, that tries to reopen, whereby in the working cylinder 22, a negative pressure is generated. This opens the pressure-controlled valve 80 and establishes the desired connection to the base pressure reservoir via the throttle 67 integrated in the cross-section 69.

LIST OF REFERENCE NUMBERS

10

Hydraulic drive

11

Core part of the drive

15

engine cylinder

16

Gas exchange channel

17

Valve stem seal

18

valve seat

19

valve guide

20

Gas exchange valve

21

Plate of the gas exchange valve

22

working cylinder

23

drive piston

24

Pressure acting surface of the drive piston 23

25

feather

26

Control edge of the drive piston

30

print media

40

Basic pressure reservoir with pressure level p ₀

41

first pressure reservoir with pressure level p ₁

42

second pressure reservoir with pressure level p ₂

46

first valve

47

first check valve

48

first connection line

49

controllable opening of the first valve 46

56

second valve

57

second check valve

58

second connection line

59

controllable opening of the second valve 56

61

Transition cross-section v. Working cylinder 22 in the connecting line 68th

62

first connection of the connecting line 68 on the working cylinder 22nd

63

second connection of the connecting line 68 on the working cylinder 22nd

64

Throttle in the second port 63

66

third valve

67

throttle

68

Connection line d. Working cylinder 22 with the base pressure reservoir 40th

69

controllable opening of the third valve 66

70

closed intermediate position of the third valve 66th

71

Pressure acting surface of the third valve 66th

72

Throttle of the third valve 66

73

Spring for returning the third valve 66

74

check valve

80

Execution of the third valve 66 as a pressure-controlled valve

82

Execution of the third valve 66 as a hydraulically timed valve

83

Play between driver 87 and valve part of the 3/2-way valve 84th

84

3/2-way valve

86

4/2-way valve

87

Driver of Stellaktors

88

common actuator

90

Pressure medium supply unit

91

Pump for first pressure reservoir

92

Pump for second pressure reservoir

93

spring chamber

94

Leak-manifold

95

Pump for re-feeding the leak

96

pressure sensor

97

electronics

98

Clippings

99

Pressure relief valve

A

Area of the pressure-acting surface 24 of the drive piston 23rd

p ₀

Pressure of the base pressure reservoir 40

p ₁

Pressure of the first pressure reservoir 41

p ₂

Pressure of the second pressure reservoir 42

Comment:

all pressures are understood relative to the ambient pressure.

H

Stroke of gas exchange valve 20 or drive piston 23

h _max

maximum opening stroke

_stat

theoretical static opening stroke

m

Effective mass of the moving component
(= Sum of the masses of:

• Gas exchange valve with spring plate, if necessary valve bridge etc.
• Mass of the drive piston 23
• Mass fraction of the spring 25th
• Mass fraction with moving pressure medium 30
• other moving parts such as valve bridge etc.)

F _F

Spring force of spring 25, depending on deflection

F _Fv

Biasing force of the spring 25
(in the closed position of the gas exchange valve, h = 0)

c

Spring constant of spring 25 (for linear characteristic)

t

Time

T _1/2

Half period of the spring mass oscillator from m and c

phases:

O

dormancy

I

Opening the gas exchange valve

II

First holding phase in the open state

III

Close the gas exchange valve

IV

Second holding phase in front of valve seat

V

Final closing of the gas exchange valve

VI

dormancy

A _1a , A _1b . A _1c

Cross-sectional course variants a, b, c of the first valve

A _2a , A _2b

Cross-sectional variations of the second valve

A ₃

Cross-sectional third valve

Claims (14)

Hydraulic drive (10) for accelerating and decelerating dynamically moving components, in particular valves in gas exchange controls of internal combustion engines and other piston engines, the hydraulic drive comprising: at least one component to be driven, in particular a valve, preferably a gas exchange valve (20) or a plurality of gas exchange valves of an internal combustion engine or of another piston engine which can be actuated jointly via a valve bridge, - A working cylinder (22) having a pressure acting surface (24) of a drive piston (23), at least one first pressure reservoir (41), for providing a first pressure p _{1 of} a hydraulic pressure medium (30), - at least one, preferably as a spring (25) formed on the component or on the gas exchange valve (20) acting, restoring energy storage with a biasing force F _Fv , at least one hydraulic base pressure reservoir (40) having a lower pressure p ₀ than the first pressure reservoir (41), characterized in that
in a first connecting line (48) between the first hydraulic pressure reservoir (41) and the working cylinder (22) a controllable opening (49) of a first valve (46) with at least one upstream in the flow path, in or behind series connected, preferably spring-loaded check valve (47) is arranged, which allows the pressure medium (30) to flow in the direction of working cylinder (22), but prevents a backflow in the direction of the pressure reservoir (41). Hydraulic drive (10) according to claim 1, characterized in that in a second connecting line (58) between the first pressure reservoir (41) and the working cylinder (22) a controllable opening (59) of a second valve (56) with at least one in the flow path in front of, in or behind the controllable opening of the second valve in series, preferably spring-loaded check valve (57) is arranged, which prevents flow in the direction of the working cylinder (22), but allows a return flow in the direction of the pressure reservoir (41). Hydraulic drive according to claim 2, characterized in that the drive has at least a second pressure reservoir (42) with a pressure p ₂ , and the controllable opening (59) of the second valve (56) with this second pressure reservoir (42) instead of the first pressure reservoir (41) is connected, wherein the pressure of the second pressure reservoir (p ₂ ) is preferably between and the pressure of the hydraulic base pressure reservoir p ₀ and the first pressure p ₁ and is preferably chosen so low that the gas exchange valve during closing reliably up to the valve seat can swing back. Hydraulic drive according to one of claims 1 to 3, characterized in that the biasing force F _{Fv of} the restoring energy storage device is adjustable. Hydraulic drive according to claim 1 to 4, characterized in that the restoring spring accumulator is formed with a progressive spring characteristic . Hydraulic drive according to one of the preceding claims, characterized in that at least the controllable opening (49) of the first valve (46) and the controllable opening (59) of the second valve (56) combined to form a valve unit having a common Stellaktor (88) are, wherein the combined valve unit is preferably designed as a 3/2-way valve (84) or as a 4/2-way valve (86). Hydraulic drive (10) according to one of the preceding claims, characterized in that in a connecting line (68) between the working cylinder (22) and the base pressure reservoir (40) a third controllable opening (69) of a third valve (66) is arranged, in the flow path before, in or behind the third valve (66) a - preferably adjustable - throttle (67) is arranged. Hydraulic drive according to claim 7, characterized in that the controllable opening (69) of the third valve (66) opens time-controlled by a predetermined time after opening the second valve (56), which is preferably dimensioned such that the second check valve (57 ) has already closed again at this time and keeps the gas exchange valve (20) fixed in this position. Hydraulic drive according to claim 8, characterized in that the third valve (66, 82) has a closed intermediate position (70) and with the opening of the second valve (56) preferably by a spring (73) driven return movement of the third valve (66 ) is released and started, wherein pressure medium via a pressure acting surface (71) of the valve displaced and by a throttle (72) is pressed, so that the intermediate position (70) of the valve is traversed only slowly and the cross section (69) only after the desired Delay time opens. Hydraulic drive according to claim 6, characterized in that the third valve (66) is pressure-controlled alone or in addition to another actuation pressure controllable, by the pressure in the working cylinder (22) in such a way that it is below a switching pressure level opens and closes above this pressure level, this pressure level is preferably a small amount above the pressure in the base pressure reservoir and significantly lower than the pressures in the first or second pressure reservoir. Hydraulic drive according to claim 7, characterized in that the transition cross section (61) from the working cylinder in the connecting line (68) is formed so that this approaching the gas exchange valve (20) to the valve seat (18) by a control edge (26) of the drive piston (23) is reduced so that the gas exchange valve brakes and gently touches the valve seat. Hydraulic drive according to claim 7, characterized in that the connecting line (68) divides into two connections on the working cylinder, the first connection (62) approaching the gas exchange valve (20) to the valve seat (18) by the control edge (26) of the Drive piston (23) is cut off and the second port (63) via a fixed or adjustable throttle (64) is guided so that the gas exchange valve brakes and gently touches the valve seat. Hydraulic drive according to one of the preceding claims, characterized in that the first valve (46) and / or the second valve (56) and / or the third valve (66) is designed as a rotary slide valve, wherein the rotary valve or the Drehschieberbventile synchronously, be driven in a fixed speed ratio to the duty cycle frequency of the reciprocating engine or the internal combustion engine or be. Hydraulic drive according to claim 13, characterized in that the phase angle at which opens the rotary valve, with respect to a reference point in the working cycle of the reciprocating engine or the internal combustion engine is adjustable.

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