The present invention relates to a valve drive with a first actuator for opening and/or closing a gas-exchange valve, in particular of an internal combustion engine, and a separate second actuator for the valve-play compensation of the gas-exchange valve.
Valve drives with a device for valve-play compensation are known in principle from the state of the art. There are many reasons for valve-clearance. Thus, during the overall life cycle of a gas-exchange valve, mechanical wear occurs at the valve seat. Above all in large engines, because of the long valve stems of the gas-exchange valves used there, considerable thermal expansions occur which also require a valve-play compensation. Thus it is the rule in large engines to compensate for a valve-clearance of up to 4 mm over the life cycle of a gas-exchange valve.
It is known in the state of the art to provide a hydraulically operated plunger cylinder with a ram as a second actuator for the valve-play compensation of the gas-exchange valve, while the opening and closing of the gas-exchange valve is carried out by a first actuator e.g. in the form of a cam, acting on the cylinder, of a camshaft. In principle it is preferable that the second actuator for the valve-play compensation remains as rigid or firm as possible in terms of length during the actuation which means opening and/or closing of the gas-exchange valve, in order to guarantee a relatively rigid transfer of the lifting movement caused by the first actuator and thus a slight sinking of the second actuator. On the other hand the valve-clearance is to be compensated in enclosed engine valves. In the valve drives known in the state of the art the rigidity of the second actuator during the opening and closing of the gas-exchange valve leaves something to be desired. This is frequently attributable to problems with a very high air content in the hydraulic fluid.
It is thus the object of the invention to improve a valve drive according to the preamble such that these problems are avoided.
This is achieved according to the invention by developing the second actuator as a double-action, preferably hydraulic, piston cylinder arrangement with at least two working surfaces that can be subjected to pressure.
Thus according to the invention it is provided, in a system with a first actuator for opening and closing the gas-exchange valve and a separate second actuator for valve-play compensation, to develop the second actuator as a double-action cylinder. This has at least two, preferably opposed, working surfaces to which pressure can be applied. It is now possible, through the measure according to the invention, to operate with higher system pressures for valve-play compensation. Thus valve drives according to the invention provide, in favourable variants, impaction with a pressure of at least 100 bar, preferably at least 200 bar. Overall a valve drive is thereby created, the second actuator or which is very rigid for valve-play compensation upon the lifting movement for the opening and closing of the gas valve. Additionally, problems with air bubbles in hydraulic fluids at correspondingly high pressures are avoided. Moreover, air bubbles initially present are quickly displaced at high pressures in the cylinder space. Operational readiness can therefore be quickly reached with high pressures.
In order to guarantee a compact design it is preferably provided that the first actuator is arranged essentially in the piston of the piston cylinder arrangement of the second actuator. The first actuator can be able to be actuated hydraulically or pneumatically or electrically. Alternatively a compact design is also achieved if the first actuator has a bush, preferably a bush-shaped piston, the piston cylinder arrangement of the second actuator being arranged essentially in the bush, preferably in the bush-shaped piston. Here the first actuator can be actuated hydraulically or pneumatically or electrically or by means of a cam or a camshaft.
Valve drives according to the invention can be used for both small and large engines. They generally have a high dynamic during adjustment. The sinking of the second actuator during a period is very well settable.
Further advantages and details of the present invention arise from the following description of figures. There are shown in:
FIG. 1 a valve-play compensation as per the preamble according to the state of the art,
FIG. 2 a first variant according to the invention,
FIG. 3 a hydraulic schematic representation,
FIG. 4 a second embodiment according to the invention and
FIG. 5 a third embodiment according to the invention.
In the state of the art it is customary, as shown in FIG. 1, to use a single-action cylinder 30 with only one working surface able to be subjected to pressure 29 for the valve-play compensation. The ram 27, the cylinder 30 and the reset spring 16 for the second actuator for the valve-play compensation. The opening and closing of the gas-exchange valve 2 results from the interaction of the cam 21 and the reset spring 17. These thus form the first actuator 1. Between the second actuator 3 and the cam 21 the play-free compensation takes place by topping up oil via the non-return valve 28. The driving force for this compensation movement comes from the compensating spring 16 of the second actuator. If the gas-exchange valve 2 is opened by means of the cam 21 the return valve closes because or the build-up of pressure in the working area 31 of the cylinder 30. The volume of oil in the second actuator 3 must be kept very small in order to achieve an at least somewhat rigid transfer and a slight sinking of the second actuator 3.
A first variant according to the invention of a valve drive is represented in FIG. 2. The gas-exchange valve 2 is driven by means of the first actuator 1. In this version this operates on a hydraulic principle and comprises a cylinder in which a ram 32 is guided. The reset spring 17 of the first actuator 1 is integrated in the cylinder. Feed pipe 8 is provided to supply the first actuator with hydraulic fluid. The whole of the first actuator 1 is arranged to achieve as compact as possible a design inside the piston 25 of the second actuator 3. In the embodiment according to FIG. 2 the piston-cylinder arrangement according to the invention for valve-play compensation is developed as a synchronous cylinder, i.e. the working surfaces A1 and A2, which each at least partially border a compensation chamber, are the same size. A central principle of valve-play compensation rests in the shown embodiment on two compensation chambers—the lower compensation chamber 5 and the upper compensation chamber 4. These are supplied with hydraulic fluid via the feed bore 6 and via the inlet throttles 14 and 15. A pressure of preferably at least 100 bar, particularly preferably at least 200 bar, applies at the feed line 6. If, because of the acceleration of the gas-exchange valve 2, a resulting force now acts upwards on the second actuator 3, the piston 25 moves upwards and the pressure in the upper compensation chamber 4 increases, as a result of which the movement is opposed. Thus there is a small sinking of the whole system. If force is oppositely directed, the pressure rises analogously in the lower compensation chamber 5. The short constrictions 14 and 15 represent the inlet throttles. They lie between the feed line 6 and the compensation chambers 4 and 5. The throttles 14, 15 can he very finely set by means of the diameter play. In order to be able to better influence the behaviour of the system and above all influence the consumption of hydraulic fluid, additional constrictions or outlet throttles 12 and 13 are provided. The hydraulic fluid flows through these to the two drain lines 7 and 11, to then flow away into a tank not further represented here. The transmittivity of the throttle coefficient of the outlet throttles 12 and 13 can again be set by means of the diameter play. This arrangement corresponds to the hydraulic circuit represented in FIG. 3, pS representing the system pressure of preferably at least 100 bar or at least 200 bar or optionally at least 300 bar and pT representing the outflow tank. In order to be able to compensate the resulting valve play with the embodiment represented in FIG. 2, a force must push the piston 25 with the first actuator 1 located in it upwards during the lifting phase. This resulting force can be achieved by means of a spring 16. This is in turn preferably developed as a helical spring. Alternatively, however, an additional elastic element, also known in the state of the art, can be provided instead of the helical spring.
In a further variant, not explicitly represented here, the spring 16 can also be omitted in the embodiment according to FIG. 2, if it is ensured via a corresponding setting of the diameter play of the outlet throttles 12 and 13 and optionally of the inlet throttles 14 and 15 that a higher pressure prevails in the lower compensation chamber 5 than in the upper compensation chamber 4. In this case the throttles have correspondingly different throttle coefficients, a higher pressure being established in the lower compensation chamber 5 than in the upper compensation chamber 4. This difference in pressure leads to a resulting force which forces the pistons 25 upwards.
In both variants the piston 25 is drawn back down by the actuator force upon closing of the gas-exchange valve. A corresponding valve-play compensation takes place through the described alternating play upon each opening and closing stroke of the first actuator 1. This movement is damped by the throttle action of the inlet throttles 14 and l5. The first actuator 1 is thus housed floating essentially inside the second actuator 3. The cylinder of the second actuator 3 is made up, for ease of dismantling and assembly, of two e.g. cylinder elements 9 and 10 which can be screwed to each other. In order to make possible an assembly and dismantling of the first actuator 1, a double-shell structure, not shown in detail here, is also recommended for the piston 25 in all variant versions with the first actuator 1 located in the piston 25.
Another version of the invention is represented in FIG. 4. Here, the double-action piston-cylinder arrangement of the valve-play compensation is developed as a differential piston cylinder arrangement with two working surfaces A1 and A2 of different sizes. Through this the additional elastic element 16 can likewise be avoided. Even if the throttle coefficients of the inlet throttles 14 and 15 are of equal size and the pressures in the compensation chambers 4 and 5 are thus of equal size, forces of different sizes act on the differently-sized surface areas A1 and A2. A resulting upward force arises therefrom in the shown embodiment, which ensures the valve-play compensation. The first actuator 1 is also arranged inside the piston 25 of the second actuator 3 in this embodiment. However, in this variant it is designed as an electric drive. The opening and closing of the in- and outlet valve 2 results from the cooperation of the reset springs 17 and 18 and the magnetic coils 19 and 20. The electric feed lines and the controls of the coils 19 and 20 can be designed as in the state of the art. The use of flexible lines or electric slip rings is recommended for the electric feed lines. The gas-exchange valve 2 is opened and closed by alternately switching the coils 19 and 20 on and off. The essentially central location of the feed line 6 and the development of the outlet throttles 12 and 13 and of the two drain lines 7 and 11 assigned to them corresponds to the embodiment from FIG. 2.
A further embodiment is shown in FIG. 5 in which the piston 25 of the second actuator is located in a cylinder formed by the two elements 9 and 10, this cylinder being part of the first actuator 1. In this embodiment, the first actuator 1 has the cam 21 and the reset spring 17 and also the bush, formed from the elements 9 and 10, in which the second actuator 3 is housed. Here, the valve-play compensation piston 25 is connected by way of example to the gas-exchange valve 2 via the conical seat 22, a split ring 23 and a securing ring 24. The cylinder housing formed from the two elements 9 and 10 is axially guided in the engine block 26. The permanent supply via the feed line 6 and the corresponding disposal via the drain lines 7 and 11 is effected via corresponding annuli 32. The second actuator 3 can, as shown, be developed as a synchronous cylinder but also, as shown in FIG. 4, as a differential cylinder. In this variant, different pressures can also be set in the work areas 4 and 5 by corresponding throttle coefficients of throttles 12 and 13 or 12, 13, 14 and 15. But in the shown embodiment, the throttles 14 and 15 have identical throttle coefficients. The relative displacement between the cylinder formed from the elements 9 and 10 and the valve-play compensation piston 25 is ensured by the elastic element 16, again preferably a helical spring. In this embodiment also, another hydraulically or electrically powered first actuator 1 can be provided instead of the cam 1. In general, the variants shown in the various embodiments can be combined in different ways with the result that the invention is not limited to the shown embodiments.
In order on the one hand to be able to set, uncoupled from each other, the slight sinking of the valve-play compensation upon the lifting movement of the gas-exchange valve 2 and on the other hand a low power requirement of the system, it is provided to use the outlet throttles 12 and 13 in order to set the total consumption of the system. For this purpose, they preferably have a much greater throttle coefficient than the inlet throttles 14 and 15. Thus something approaching the system pressure pS is established in the compensation chambers 4 and 5. The influence of the two inlet throttles 14 and 15, which advantageously have the smaller throttle coefficient, ensure sinking during the lifting phase. As a result, the two described requirements are uncoupled from each other, an overall low output and fluid consumption respectively of the second actuator being able to be ensured by the outlet throttles 12 and 13. In further versions, not explicitly shown here in more detail, the throttles 12, 13, 14 and 15 can also be realized as separate throttles. The second actuator is preferably developed hydraulically and is operated with hydraulic fluid. In the case of the shown embodiments, however, it is also possible to operate the second actuator 3 pneumatically. In this case the driving fluid is a gas, preferably air.
Patent Application
20050279305
