WO2023143682A1

WO2023143682A1 - A hydraulic valve lifter system and an engine therewith and a method of its operation

Info

Publication number: WO2023143682A1
Application number: PCT/DK2023/050009
Authority: WO
Inventors: Christian BOËL VOLLERTSEN
Original assignee: VOLLERTSEN, Troels; VOLLERTSEN, Tobias
Priority date: 2022-01-31
Filing date: 2023-01-19
Publication date: 2023-08-03
Also published as: DK202200092A1

Abstract

A hydraulic valve lift system and an engine therewith, and a method of its operation, the hydraulic valve lift system comprising a lift piston (12) sliding reciprocatingly inside a pump cylinder (11) to pump a portion of a hydraulic fluid to a valve system, where the size of the part can be regulated by turning the pump cylinder (11) in relation to the lifting piston (12).

Description

A HYDRAULIC VALVE LIFTER SYSTEM AND AN ENGINE THEREWITH AND A METHOD OF ITS OPERATION

Field of invention

The present invention relates to a hydraulic valve lifter system and an engine with such system. The invention also relates to a method of operating such engine.

Background of the invention

For internal combustion engines, performance and emission characteristics are regulated by timing of the opening and closing of the cylinders' gas intake valves. The timing includes the opening timing as well as the duration of the gas flow into the cylinder through the intake port in the cylinder head. Timing variations corresponding to even small amounts relatively to the engine crank angle have significant effects on the performance of the engine as well as the emission characteristics. In modern engines, the opening timing and duration is varied in dependence on the load of the engine and the overall desired performance characteristics. Additionally, some modern automobiles can switch between an economic mode and a sports mode, which affects the valve operation.

Hydraulic actuated intake valves are regulated by adjusting the amount of hydraulic liquid, typically oil, that is supplied from a camshaft-driven pump to the engine valve system, for example to a slave piston that presses onto the stem of the valve. For regulation of the opening distance and time, various systems in the prior art use combinations of computer-controlled solenoid drain valves and drain reservoirs that take up a predetermined portion of the pumped hydraulic liquid in dependence on the desired amount remaining for the actuation of the intake valve. Examples are disclosed in US patent applications US2013/047942 and US2019/0162085.

GB224892A discloses a hydraulic mechanism for actuating the distributing valves of reciprocating engines with an active piston, formed as a hollow cylinder, open at the pressure side, the wall of said piston having one or more openings collaborating with openings in the cylinder, wherein said active piston being operated from the controlling shaft of the engine and one or more passive pistons connected to the valves. Said passive pistons under the influence of mechanical or liquid pressure being returned into the normal' position during the working stroke of the active piston as soon as openings in the wall of_ said piston are juxta- posed with regard to the openings in the cylinder wall. During that part of the working stroke of the active piston, at which the passive pistons are at rest, or are returned into the normal position, liquid is pressed towards a compensating vessel. During the return stroke-of the active piston liquid is sucked from said compensating vessel and during a part of the return stroke along the pressure space of the cylinder of the passive pistons.

However, such draining function needs precise valves that also have to be precisely timed in order to open correctly for each valve lifting, not only for the volume control but also for the valve opening and closing timing. These requirements make such systems relatively complex as the proper functioning thereof needs precise dimensioning and computer steering, which is more difficult the faster an engine is running. Especially for modern high- performance engines that run at high revolution speed with up to 12 cylinders, this is a challenging task.

A more robust hydraulic valve actuation system that has less stringent requirements in one of its functional modes with respect to timing of a solenoid is disclosed in US patent No. 4716863 and will be discussed in more detail in the following. A drawing from this disclosure is reproduced in FIG. 1. A piston 2 at the end of a connecting rod 3 slides in the engine's combustion cylinder 1. An intake valve 8 in the intake port 10 of the cylinder head is held in its closed position by the force of a valve spring 9. When the valve 8 is opened, a mixture of air and fuel can pass into the cylinder 1 for combustion between the cylinder head and the top of the piston 2. The valve 8 is actuated hydraulically by a piston pump comprising a lifter piston 12 driven by a camshaft 4 and reciprocating inside a pump cylinder 11. When the camshaft 4 rotates, its cam lobe moves the front end of the lifter piston 12 in a forward pumping direction and passes a spill port 23, which leads to pressurization of oil in the pump cylinder 11 and in the hydraulic line 15, the latter being fluid-flow connected to the valve system. The pressurized oil in the hydraulic line 15 flows into a slave cylinder 13, activating a dowel pin 14 and pressing on the stem of the valve 8 and pushing the valve 8 into an open state against the spring 9 force. In order to adjust the timing and duration of the opening of the valve 8, a control cylinder 16 is hydraulically coupled to the hydraulic line 15. Inside the control cylinder 16, a control piston 17 is pre-stressed by a spring 18 that bears against a core 21 of a solenoid 19, which is controlled by a microcomputer. By adjusting the position of the core 21 and correspondingly the position of the control piston 17, the control cylinder 16 can take up more or less of the oil from the hydraulic line 15, which affects the opening duration and distance of the valve 8. When the core 21 and the control piston 17 are retracted, more oil is taken up by the control cylinder 16 and less oil is available from the slave cylinder 13 for lifting the valve 8, so that the valve 8 will open later and close earlier with a reduced valve lift. A port with a check valve 22 is attached to the hydraulic line 15 to continuously recirculate the engine oil in the system.

In one of the modes explained in US4716863, the solenoid is not moving for each revolution of the engine and can thus be placed in a certain position, until the valve actuation shall be changed. This is a robust and simple mode of function. However, the hydraulic valve lifter system of US patent No. 4716863 has a number of disadvantages. Firstly, the control cylinder 16 with the spring 18 loaded control piston 17 is a relatively complex technical solution, which adds costs and which can be difficult to implement in engines with tight space around the cylinder head 1. Furthermore, at high camshaft speed, control of the timing of the valve 8 is subject to a slight uncertainty due to hysteresis in the reaction of the control spring 18 relatively to the valve spring 9, especially at high speed, which causes small but unwanted delays in the movement of the control piston 17, especially during adjustment of the solenoid core 21, which, in turn, translates into uncertainties in the timing of the valve 8. This is unfortunate, as high precision is needed particularly at high speed in order to provide high power characteristics in the engine. As it appears, the prior art discloses solenoid solutions which are either complex with high requirements to timing or are more robust but lack precision.

Accordingly, for performance optimization of internal combustion engines, it would be desirable to have a system for hydraulically operated valves, which is robust, precise, reliable and simple.

Summary of the invention

It is an objective of the invention to provide an improvement in the art. In particular, it is an objective to provide an improved hydraulic actuation system for intake valves in a cylinder head of an internal combustion engine, especially having improvements with respect to speed, reliability and simplicity. Other objectives will become apparent in the following.

One or more of these objectives are achieved with a technical solution as defined in the claims and as described in the following. In particular, the objective is achieved with a hydraulic valve lifter system and an internal combustion engine comprising such system as set forth in the following as well as a method of operation of the system and the engine.

In short, the hydraulic valve lifter system comprises a lifter piston slidingly reciprocating inside a pump cylinder for pumping a portion of hydraulic fluid to a valve system at each reciprocation of the engine, wherein the size of the pumped portion can be regulated by rotating the pump cylinder relatively to the lifter piston or vice versa. Details and practical embodiments are explained in the following.

The type of engine in which the hydraulic valve lifter system is useful to integrate has a combustion cylinder that is closed at one end by a cylinder head. The cylinder head comprises an intake valve system for intake of gas into the cylinder for internal combustion. For example, the intake gas is a mix of gaseous fuel and air. For the case that fuel is injected directly into the cylinder, the intake gas is air. Additionally, the cylinder head also comprises an outlet valve system for outlet of combustion gases from the combustion.

A hydraulic valve lifter system, as explained in the following, will typically be used for the intake valves but can alternatively or in addition be used for the outlet valves.

The valves are provided with a valve head for opening and closing a gas port with a corresponding valve seat in the cylinder head. Typically, the valve head is provided at one end of a valve stem that is arranged for reciprocal movement in a valve bore. Advantageously, such valves with valve stem are pre-stressed by a valve spring towards a closed valve position.

The hydraulic valve lifter system is operated by a rotational camshaft of the engine, the rotation of which is synchronized with revolutions of the engine.

The hydraulic valve lifter system comprises a hydraulic pump for periodically actuating a single engine valve, for example intake valve, or a group of engine valves by hydraulic pressure. If the valves are pre-stressed towards the closed position by a spring, the hydraulic pump is pressing the hydraulic liquid, typically oil, against the force of the valve spring in order to open the valves.

The hydraulic pump comprises a lifter piston, arranged tightly fitting inside a pump cylinder for linear reciprocation along a reciprocation axis. The tight fit prevents leak of the hydraulic fluid, typically oil, to pass the lifter piston, at least to the extent that a reliably sufficient quality of the hydraulic pumping can be obtained. It is pointed out that the outer side of the lifter piston and the inner tube of the pump cylinder are coaxial and circular in cross section with a common central axis.

The lifter piston is coupled, or configured for being coupled, to a cam lobe of the camshaft, typically by providing a tappet at a first end of the piston, the tappet being in contact with the cam lobe, so that, in operation, the lifter piston reciprocates due to the rotation of the camshaft. Potentially, the coupling of the lifter piston to the cam is achieved through a gearing, which in some cases optimizes performance.

The pump cylinder comprises a reservoir with a volume of hydraulic liquid. When the pump is installed in an engine, the reservoir is in fluid-flow connection with the valve system for flow of the hydraulic liquid from the reservoir to the valve system through a corresponding hydraulic line for actuation of the valve from a closed state into an open state when the lifter piston in the cylinder by a forward movement towards a top dead centre pumps a portion of the hydraulic liquid out of the reservoir and through the hydraulic line for opening the corresponding valve.

In contrast to the prior art, the hydraulic pump is peculiar in that it is configured for relative rotation between the pump cylinder and the lifter piston about the central axis, and the portion pumped from the reservoir to the valve, during a forward movement of the lifter piston, is variable and determined by the relative rotational orientation between the pump cylinder and the lifter piston. In other words, by rotating the pump cylinder relatively to the lifter piston, or the lifter piston relatively to the pump cylinder, or by rotating both, relatively to each other, in order to change the relative orientation between these two, the size of the portion that is pumped to the valve system is changed. In some practical embodiments, only the pump cylinder is rotational, whereas the lifter piston is rotationally fixed.

Optionally, the hydraulic valve lifter system comprises a ball bearing or roller bearing that is holding the pump cylinder and configured for smooth rotation of the pump cylinder in the bearing. This is advantageous when implementing the hydraulic pump in an engine

Advantageously, a sensor is used for determining the actual relative rotational orientation and changes thereof.

It is preferred that the size of the pumped portion is variable by the relative rotation and determined by the relative rotational orientation between the pump cylinder and the lifter piston. In practical embodiments, the hydraulic lifter system comprises an actuator which is functionally connected, typically mechanically connected, to the hydraulic pump for causing the relative rotation in a controlled way and which is configured for automatically varying and adjusting the relative rotational orientation between the pump cylinder and the lifter piston, on the basis of electronic instructions from a correspondingly programmed computer. For example, the actuator is functionally connected only to the pump cylinder for causing rotation thereof but not to the lifter piston.

Typically, the adjustment of the relative rotational orientation between the lifter piston and the pump cylinder is controlled by a computer which receives operational parameters from the engine, such as speed and load, and potentially exhaust gas data, and effectuates the adjustment of the relative rotational orientation on the basis of an evaluation of the data, and a computer program for optimization of the engine performance by adjustment of the actuation characteristics of the valves. For example, the computer control for the adjustment of the rotation has implemented a feedback loop for continuously adjusting the size of the portion of the hydraulic fluid on the basis of the actual engine parameters for optimised engine performance.

The adjustment of the relative rotational orientation between the lifter piston and the pump cylinder is a simple way of adjusting the size of the portion without the need of solenoid valves and buffer reservoirs. Also, the timing is not so critical, as the rotation can be done gradually even during the reciprocation without the necessity of quick changes for each revolution of the engine. Nevertheless, it is practically possible to change the portion size quickly by such rotation, especially when the sensitivity of the adjustment is high due to specific shaping of a front end of the lifter piston. Such features are described in more detail in the following with some concrete examples.

In some practical embodiments, the lifter piston is reciprocated by the lobe of the camshaft between a bottom dead centre (BDC) position and a top dead centre (TDC) position. Similar to the prior art, as described in the introduction, the pump cylinder comprises a spill port for flow of hydraulic liquid through the spill port. Such spill port is in fluid-flow connection with the hydraulic liquid reservoir in the pump cylinder when the lifter piston is in the BDC position, and the spill port is disconnected from reservoir when the lifter piston is in the TDC position. The lifter piston has a front end, opposite to the first end, and is arranged for disconnecting the spill port from the reservoir by its front end when the front end passes the spill port during the forward motion of the lifter piston towards the TDC position. During the forward motion of the lifter piston, after the front end has passed the spill port, the hydraulic liquid in the reservoir is pressurised sufficiently by the lifter piston to act against the return spring force of the valves in the engine valve system, and hydraulic liquid is pumped to the valve system for lifting the valve.

In a practical embodiment, the front end of the lifter piston is profiled with a first part and a second part, where the second part is offset towards the first end relatively to the first part. In this configuration, the lifter piston is arranged for passing and closing the spill port during its forward motion only by the first part for a first rotational orientation and only by the second part for a second rotational orientation between the pump cylinder and the lifter piston. The second orientation is different from the first orientation, for example offset rotationally by an angle of 90 or 180 degrees.

Due to the second part being offset towards the first end of the lifter piston relatively to the first part in a direction parallel with the central axis, the lifter piston will, during forward movement towards the TDC position, advance relatively more until the spill port is covered and accordingly press relatively more hydraulic fluid out of the reservoir 35 (see FIG. 5) through the spill port first, before the spill port is covered by the second part. Only after that, the reservoir subsequently gets pressurized and a portion of the hydraulic liquid from the reservoir gets pumped through a hydraulic line to the valves. Accordingly, the portion that is pumped to the valve is smaller in the case where the second part is passing the spill port, relatively to the first part passing the spill port.

In summary, the portion of hydraulic liquid pumped to the valve system is smaller in the second rotational orientation than in the first orientation, and the valves are correspondingly lifted less in the second orientation than in the first orientation. By the features of the profiled front end of the lifter piston in combination with the rotational adjustment between the lifter piston and the pump cylinder, the size of the portion of hydraulic liquid that is pumped to the valve system is reduced by changing the relative orientation from the first to the second position.

Accordingly, the effective length of the lifter piston is reduced by the relative rotation from the first to the second orientation. The size of the pumped portion is smaller for a smaller effective length.

Optionally, the front end of the lifter piston is profiled with step-wise transition, for example multi-step transition, from the first part to the second part for step-wise adjustment of the portion that is pumped to the valve system in dependence on a relative orientation in between the first and the second orientation.

Alternatively, the front end of the lifter piston is profiled with a curved or linear slope from the first part to the second part for smooth and continuous variable adjustment of the portion that is pumped to the valve system in dependence on a relative orientation in between the first and the second orientation. For example, for continuous adjustment, the front end is skew shaped and sloping linearly from the first part to the second part, the latter being offset towards the first end relatively to the first part.

In operation, the portion of hydraulic liquid can be adjusted continuously or in steps by variably adjusting the relative orientation in between the first and the second position.

For clarification, it is pointed out that the described shape of the front end relates to its shape at the outer peripheral rim of the lifter piston, which is abutting the inner surface of the pump cylinder and which is also decisive for closing the spill port. Accordingly, the first part and second part as well as the slope in between are to be understood as at the rim of the lifter piston.

Short description of the Drawings

The invention will be explained in more detail with reference to the drawings, which are given for illustration and not delimiting the present invention, although they are showing aspects useful for the invention. In the accompanying drawings:

FIG. 1 is a reproduction of a prior art drawing from US4716863;

FIG. 2A is a drawing with a partial cross-sectional perspective view of a test rig illustrating the principles of an embodiment of the invention;

FIG. 2B shows a different partial cross-sectional perspective view of the test rig;

FIG. 3 shows a perspective view in partial cross section of the hydraulic pump system;

FIG. 4 shows a perspective view in partial cross section of the hydraulic pump system in a different embodiment;

FIG. 5 shows a close-up illustration of the region around the spill port in the pump cylinder of FIG. 3;

FIG. 6 shows an example of an outer side of the pump cylinder;

FIG. 7 shows an example of a sloped front end of the lifter piston;

FIG. 8A shows another example of a double-sloped front end of the lifter piston;

FIG. 8B shows an example of the lifter piston;

FIG. 9 shows another example of a pump cylinder with an elongated spill port.

Detailed description of the invention

Referring now to the drawings for the purpose of explaining embodiments and further details of the present invention:

FIG. 1 is a reproduction of prior art drawing from US4716863 and has been explained in the introduction.

FIG. 2A is an illustrative drawing of a test rig, illustrating the main components and the principles of the hydraulic valve lifter system for an engine. In case that the illustrated valve system would have been part of an internal combustion engine, the combustion chamber of the engine would have been to the right of the valve heads 27.

Each of the valves 8 comprises a valve head 27 which when resting against a valve seat 37 closes a port 10, for example a gas intake port through a cylinder head of an engine, as explained in relation to the prior art drawing in FIG. 1. A valve spring 9 pre-stresses the valve 8 into a closed position. The valve head 27 is carried by a valve stem 24, which extends through bearings 26 for guided reciprocal movement of the valve 8.

In the shown illustration, the movement of the valve stem 24 is caused by hydraulic pressure acting on the back of the stem 24. However, it is also possible to use a pressure transfer block between the hydraulic liquid and the valve stem for transfer of the pressure from the hydraulic liquid to the valve, for example similarly in principle to the dowel pin 14 in slave cylinder 13 illustrated in FIG 1.

The force on the valve stem 24 is caused by pressurized hydraulic liquid, typically oil, in valve pressure canal 15C, which is seen in greater detail in FIG. 2B, which is an enlarged perspective view of a partial cross section of the test rig in FIG. 2A. As part of a hydraulic line 15, valve pressure canal 15C receives pressurized hydraulic liquid through opening 15A and transverse canal 15B, which is best seen in FIG. 2B. In order for the hydraulic line 15 to be filled with hydraulic liquid at all times, hydraulic liquid is supplied through check valve 22.

As shown in FIG. 2A and in enlarged partial cross-sectional view in FIG. 3, a hydraulic pump 25 is used for pumping the pressurized hydraulic liquid through hydraulic line 15 to the valves 8. The hydraulic pump 25 comprises a lifter piston 12 with an outer circular cylindrical shape arranged for linear reciprocation inside a pump cylinder 11 which has a corresponding inner circular cylindrical canal, tightly fitting to the lifter piston 12 for prevention of leaking hydraulic liquid. The lifter piston 12 has at its first end 12A a tappet 6 in contact with a cam lobe 5 of a rotational camshaft 4. The tappet 6 illustrated is a roller tappet, although a flat tappet could be used as well. Alternatively, the first end 12A could be coupled to the camshaft 4 differently, for example through a gearing or lever. In operation, the rotation of the camshaft 4 is synchronized with the revolution of the engine.

A lifter spring 7 presses the lifter piston 12 and the tappet 6 against the lobe 5 of the camshaft 4, so that the rotation of the camshaft 4 causes the lifter piston 12 into a reciprocal movement inside the pump cylinder 11. The outer side of the lifter piston 12 is tightly abutting the inner side of the pump cylinder, so that a forward movement of the lifter piston 12 towards its TDC position results in pressurizing hydraulic liquid in the hydraulic line 15 and pumping of a predetermined portion of the hydraulic liquid through the hydraulic line 15 to the valves 8, which causes opening of the valves 8 against the force of the valve spring 9.

The term forward movement of the lifter piston 12 is defined as a direction causing flow of the hydraulic liquid towards the valve. In the illustrated embodiment, the term forward movement of the lifter piston 12 implies a direction away from the camshaft 4 and towards the TDC of the lifter piston and a rearward direction is directed towards the camshaft 4.

As already mentioned, FIG. 3 shows the hydraulic pump 25 as illustrated in the test rig of FIG. 2A. Grooves 30 are provided with O-rings 36 for tightening the outer side of the pump cylinder 11 against the boring in the test rig block of FIG. 2A.

FIG. 4 illustrates a modified hydraulic pump 25 for use in an engine. The pump cylinder 12 is embedded in bearings 32, for tightening against unwanted flow of hydraulic liquid, typically oil, to pass the outer side of the pump cylinder. Further functions of the bearings will become apparent in the following after further explanation of the adjustment features of the pump 25. FIG. 5 is an enlarged part of FIG. 3 and illustrates in greater detail the lifter piston 12 inside the pump cylinder 11 with a first part 31A of the front end 31 of the lifter piston 12 covering about half of a spill port 23 of the pump cylinder 11 in the present intermediate position of the lifter piston 12 between the BDC and the TDC. When the lifter piston 12 is moved further forward, which is upwards in the present illustration, hydraulic fluid is pressed through spill port 23 out of a hydraulic reservoir 35 inside the pump cylinder 11, until the first part 31A at the rim of the lifter piston 12 fully covers the spill port 23. Once the first part 31A of the front end 31 has passed the spill port 23, the reservoir 35 inside the pump cylinder 11 is no longer connected to the spill port 23, so that further advance of the lifter piston 12 pressurizes the hydraulic liquid in the reservoir 35 and pumps it through the hydraulic line 15 to the valves 8 for actuating the valves 8 into an open state. By retraction of the lifter piston 12, which is done by the lifter spring 7 when the camshaft continues turning and gives way for the lifter piston to return to its BDC position, the pressure on the valves 8 is released again, and the valves 8 close by the pre-stressed return force valve spring 9, and spill port 23 is opened again.

In the shown system, the pump cylinder 11 can be rotated relatively to the lifter piston 12 about the central axis 38 of the lifter piston 12, which is also the axis for reciprocation of the lifter piston 12 inside the pump cylinder 11. However, in some alternative technical solutions, it is possible in principle to rotate the lifter piston 12 instead of or in addition to rotating the pump cylinder 11.

By rotating the pump cylinder 11 relatively to the lifter piston 12 by 180 degrees about the reciprocal axis 38, or equivalently rotating the lifter piston 12 180 degrees relatively to the pump cylinder 11, it is not the first part 31A of the front end 31 of the lifter piston 12 that is going to cover the spill port 23 during advance of the lifter piston 12 but the second part 31B. Due to the second part 31B being offset towards the first end 12A of the lifter piston 12 relatively to the first part 31A in a direction parallel with the central axis 38, the lifter 12 piston will, during forward movement towards the TDC position, advance relatively more until the spill port 23 is covered, and accordingly press relatively more hydraulic fluid out of the reservoir 35 through the spill port 23 before the spill port 23 is covered by the second part 31 B. Only after that, the reservoir 35 subsequently gets pressurized and a portion of the hydraulic liquid from the reservoir 35 gets pumped through the hydraulic line 15 to the valves 8. Accordingly, the portion that is pumped to the valve 8 is smaller in the case where the second part 31 B is passing the spill port, relatively to the first part 31A passing the spill port 23.

As a result, when the orientation of the pump cylinder 11 is changed so that not the first part 31A but the second part 31B is passing and covering the spill port 23, the timing for pumping of the hydraulic fluid to the valves 8 during forward motion of the lifter piston 11 becomes later, and the opening of the spill port 23 during retraction of the lifter piston 12 becomes earlier, which affects the timing of the valves as well as the amount of hydraulic liquid being pumped to the valves 8, so that the valves are opened less and for a shorter duration.

As there is a sloping transition 31C at the front end 31 from the first part 31A to the second part 31B, a relative rotation of less than 180 degrees adjusts the pump timing and liquid amount in between the two extremes of the first orientation where the first part 31A passes the spill port 23, and the second orientation where the second part 31B passes the spill port 23.

It is, however, also possible to arranged shape the lifter piston 12 in such a manner that the second part 31 B is arranged in a lower position, so that no lift is achieved.

By providing an actuator 40, stylistically shown in FIG. 2B, which is configured for adjusting the rotational orientation at any position between the first and second orientation, the amount of pumped hydraulic fluid, and correspondingly the timing and lifting of the valves 8, can be regulated to any value in between these two extremes. For the sake of completeness, it is emphasized that the first part 31A and the second part 31B and the transition 31C in order to cover the spill port 23 are at the periphery of the lifter piston 12, especially the front end 31, where it abuts the inner surface of the pump cylinder 11.

In comparison with the prior art system in FIG. 1, it is pointed out that the relative rotation between the lifter piston 12 and the pump cylinder 11 has several advantages. The system according to the invention is quick, reliable, precise, robust, and a relatively simple technical solution that can be provided at relatively low cost.

With reference to FIG. 4, a rotational bearing 33, for example ball bearing or roller bearing, is advantageous for smooth rotation of the pump cylinder 11.

The spill port 23 can be shaped according to preferences. FIG. 4 and FIG. 9, as seen from outside and inside, respectively, illustrate one embodiment with an elongated spill port 23. In comparison, FIG. 6 illustrates an embodiment where the spill port 23 is circular, corresponding to the one illustrated in FIG. 5.

With reference to FIG. 4, the spill port 23 fluid-flow communicates with a groove 34, which is turn is in fluid-flow communication with a spill canal 20, illustrated in FIG. 2B, fluid-flow communicating with spill port 23 for drain of hydraulic fluid out through the spill port 23 for recycling in the system.

The illustration in FIG. 7 shows a linearly sloped transition 31C at the front end 31 between a single first part 31A and a single second part 31B. The second part 31B is at a location on the outer periphery of the front end 31 where the distance "d" from the periphery at the second part 31 B to a lateral plane 39 at or near the first end 12A is shortest. In comparison, the first part 31A is at a location on the outer periphery of the front end 31 where the distance "D" from the periphery of the front part 31 to the same lateral plane 39 at or near the first end 12A is longest. The term lateral plane 39 is used for a plane perpendicular to the central axis 38 which is closer to the first end 12A than the front end 31. The location of the lateral plane 38 is not important as long as it is remote from the second part 31B towards the first end 12A.

Alternatively, in the present configuration the distance could be measured to the centre 6A of the tappet 6, which is illustrated in FIG. 9, as this is a suitable stationary point relative to the central axis.

However, alternative profiles for the front end 31 are possible, and an example is illustrated in FIG. 8A, where two identical first parts 31A are arranged oppositely to each other on the front end 31 with an angular offset of 180 degrees between the two first parts 31A. Further, two second parts 31B are provided oppositely on the front end 31 with an offset to each other of 180 degrees. Similar to the embodiment in FIG. 7, the two second parts 31B are at locations of the front end 31 on the periphery of the lifter piston 12 where the distance "d" to the lateral plane 39 or to the centre 6A is shortest, and the two first parts 31A are at locations of the front end 31 on the periphery of the lifter piston 12 where the distance "D" to the lateral plane 39 is longest. Four transition regions 31C are provided from the two second parts 31A to the two first parts 31 B. An advantage of this profile is a higher sensitivity with respect to adjustment of timing and amount relatively to a given change of rotation. For such profile of the front end 31, the pump cylinder 11 could be provided with two spill ports 23 (not shown) arranged oppositely to each other with a 180 degrees offset about the central axis 38.

FIG. 8B illustrates a lifter piston 12 provided with holes 42 at the top portion of the piston 12. These holes 42 are formed as through-going bores. Accordingly, the holes 42 are configured to provide fluid communication with spill port (as shown and explained with reference to FIG. 5 and FIG. 5). Accordingly, the lifter piston 12 is configured to be arranged for passing and disconnecting the spill port from the reservoir ((as shown and explained with reference to FIG. 5 and FIG. 5) depending on the rotational orientation of the piston 12. It is possible to apply rectifier valves, wherein several holes and oil channels are provided in the cylinder wall in order to ensure that the oil has a single flow direction. In one embodiment, there is a single pressure port and one spill port.

FIG. 2B shows an example of a point of connection 29, where an actuator 40 (illustrated stylistically only) can potentially be connected for rotating the pump cylinder 11. When the hydraulic lifter system is implemented in an engine, other practical options exist and can be selected in dependence on how the engine and the pump system is configured.

It is possible to rotate the pump cylinder 11 relatively to the lifter piston 12, while the lifter piston 12 is kept rotationally fixed. Alternatively, a rotational mechanism can be provided for the lifter piston, for example between the tappet 6 and the lifter piston 12 so that the lifter piston 12 is rotated while the pump cylinder 11 is rotationally fixed. As a further alternative, both the lifter piston 12 and the pump cylinder 11 can be mounted rotationally, although this is a more complex solution and often not preferred.

List of reference numerals

1 - Combustion cylinder of engine

2 - Piston in combustion chamber of cylinder

3 - Connecting rod for piston 2

4 - Camshaft

5 - Cam lobe

6 - Tappet

7 - Lifter spring for return of lifter piston 11

8 - Valve (intake valve or outlet valve) in cylinder head

9 - Valve spring for valve 8

10 - Intake port (or outlet port)

11 - Pump cylinder

12 - Lifter piston

12A - First end of lifter piston 12

13 - Slave cylinder

14 - Dowel pin

15 - Hydraulic line for hydraulic liquid

15A - Throughput hole for hydraulic line 15

15B - Transverse canal of hydraulic line 15

15C - Valve pressure canal of hydraulic line 15

16 - Control cylinder

17 - Control piston in control cylinder 16

18 - Control spring for control piston 17

19 - Solenoid

20 - Spill canal

21 - Solenoid core

22 - Check valve

23 - Spill port

24 - Valve stem

25 - Hydraulic pump

26 - Bearing for valve stem 24

27 - Valve head

28 - Valve guide

29 - Actuator attachment location

30 - Groves for O-rings in pump cylinder 11 31 - Front end of lifter piston 12

31A - First part of front end 31

31 B - Second part of front end 31

32 - Bushing for pump cylinder 11 33 - Roller bearing for pump cylinder 11

34 - Groove connected to spill canal 20

35 - Reservoir in pump cylinder 11

36 - O-ring

37 - Valve seat 38 - Central axis of lifter piston 12

39 - Lateral plane near the first end 12A

40 - Actuator

42 - Hole

Claims

1. An internal combustion engine having a combustion cylinder (1) that is closed at one end by a cylinder head, the cylinder head comprising a valve system, which comprises a valve (8) that is an intake valve at a gas intake port for intake of gas for internal combustion or a gas outlet valve for outlet of combustion gases; wherein the engine comprises a camshaft-driven hydraulic valve lifter system for hydraulic actuation of the valve (8), wherein the rotation of the camshaft (4) is synchronized with revolutions of the engine; wherein the hydraulic valve lifter system comprises a hydraulic pump (25), which comprises a pump cylinder (11) and comprises a lifter piston (12) arranged for linear reciprocation inside the pump cylinder (11); wherein the lifter piston (12) is coupled to a cam lobe (5) of the camshaft (4) for causing the reciprocation by rotation of the cam-shaft (4); wherein the lifter piston (12) is configured for pressurizing hydraulic liquid in the pump cylinder (11) and pumping a portion thereof to the valve system for hydraulic actuation of the valve (8); wherein the hydraulic valve lifter system comprises an actuator (40) which is functionally connected to the hydraulic pump (25) for relative rotation between the pump cylinder (11) and the lifter piston (12) and configured for variably adjusting a relative rotational orientation between the pump cylinder (11) and the lifter piston (12), and wherein the size of the portion is variable by the relative rotation and determined by the relative rotational orientation, characterised in that the lifter piston (12) has a first end (12A) coupled to the cam lobe (5) of the camshaft (4) and an opposite front end (31) reciprocal inside the pump cylinder (11); wherein the pump cylinder

(11) comprises a reservoir (35) of the hydraulic liquid and a spill port (23) for flow of hydraulic liquid from the reservoir (35) through the spill port (23); wherein the lifter piston (12) is configured for disconnecting the spill port (23) from the reservoir (35) when the front end (31) passes the spill port (23) during advance of the lifter piston (12) for pressurizing the hydraulic liquid in the reservoir (35) and for pumping the portion of hydraulic liquid to the valve system; wherein the front end (31) of the lifter piston

(12) is profiled with a first part (31A) and a second part (31B), where the second part (31B) is offset towards the first end (12A) relatively to the first part (31A), and wherein the lifter piston (12) is arranged for passing and disconnecting the spill port (23) from the reservoir (35) only by the first part (31A) for a first rotational orientation and only by the second part (3 IB) for a second rotational orientation between the pump cylinder (11) and the lifter piston (12), and wherein the portion of hydraulic liquid to the valve system is smaller in the second rotational orientation than in the first rotational orientation for actuating the valve (8) less in the second than in the first rotational orientation.

2. An engine according to claim 1, wherein the front end (31) of the lifter piston (12) is sloping curvedly or linearly from the first part (31A) to the second part (31 B) for smooth and continuous variable adjustment of the size of the portion that is to be pumped to the valve system in dependence on a relative rotational orientation in between the first and the second rotational orientation.

3. An engine according to claim 1 or claim 2, wherein only the pump cylinder

(11) is rotational, and the actuator (40) is functionally connected only to the pump cylinder (11) for causing rotation thereof, whereas the lifter piston

(12) is rotationally fixed.

4. An engine according to claim 3, wherein the hydraulic valve lifter system comprises a ball bearing or roller bearing (33) that is holding the pump cylinder (11) and configured for smooth rotation of the pump cylinder (11) in the bearing (33).

5. A method for operating an internal combustion engine having a combustion cylinder (1) that is closed at one end by a cylinder head, the cylinder head comprising a valve system, which comprises a valve (8) that is an intake valve at a gas intake port for intake of gas for internal combustion or a gas outlet valve for outlet of combustion gases; wherein the engine comprises a camshaft-driven hydraulic valve lifter system for hydraulic actuation of the valve (8), wherein the rotation of the camshaft (4) is synchronized with revolutions of the engine; wherein the hydraulic valve lifter system comprises a hydraulic pump (25), which comprises a pump cylinder (11) and comprises a lifter piston (12) linearly reciprocating inside the pump cylinder (11); wherein the lifter piston (12) is coupled to a cam lobe (5) of the camshaft (4) causing the reciprocation by rotation of the camshaft (4); wherein the lifter piston (12) is pressurizing hydraulic liquid in the pump cylinder (11) and pumping a portion thereof to the valve system for hydraulic actuation of the valve (8); wherein the hydraulic pump (25) is configured for relative rotation between the pump cylinder (11) and the lifter piston (12) and for variable adjustment of a relative rotational orientation between the pump cylinder (11) and the lifter piston (12), and wherein the size of the portion is variable by the relative rotation and determined by the relative rotational orientation; wherein the method comprises pumping a portion of hydraulic fluid by the pump system (25) to the valve system for each reciprocation of the lifter piston (12) during operation of the engine and changing the size of the portion by changing the relative rotational orientation between the pump cylinder (11) and the lifter piston (12), characterised in that the lifter piston (12) has a first end (12A) coupled to the cam lobe (5) of the camshaft (4) and an opposite front end (31) reciprocally inside the pump cylinder (11); wherein the pump cylinder (11) comprises a reservoir (35) of the hydraulic liquid and a spill port (23) for flow of hydraulic liquid from the reservoir (35) through the spill port (23); wherein the lifter piston (12) is configured for disconnecting the spill port (23) from the reservoir (35) when the front end (31) passes the spill port (23) during advance of the lifter piston (12) for pressurizing the hydraulic liquid in the reservoir (35) and for pumping the portion of hydraulic liquid to the valve system; wherein the front end (31) of the lifter piston (12) is profiled with a first part (31A) and a second part (31B), where the second part (31B) is offset towards the first end (12A) relatively to the first part (31A), and wherein the lifter piston (12) is arranged for passing and disconnecting the spill port (23) from the reservoir (35) only by the first part (31A) for a first rotational orientation and only by the second part (31B) for a second rotational orientation between the pump cylinder (11) and the lifter piston (12), and wherein the portion of hydraulic liquid to the valve system is smaller in the second rotational orientation than in the first rotational orientation for actuating the valve (8) less in the second than in the first rotational orientation, wherein the method comprises reducing the size of the portion of hydraulic liquid by changing the relative orientation from the first to the second position.

6. A method according to claim 5, wherein the front end (31) of the lifter piston (12) is sloping curvedly or linearly from the first part (31A) to the second part (31 B) for smooth and continuous variable adjustment of the size of the portion that is pumped to the valve system in dependence on a relative rotational orientation in between the first and the second rotational orientation, wherein the method comprises adjusting the portion of hydraulic liquid continuously by variably adjusting the relative orientation in between the first to the second position.

7. A method according to claim 5 or 6, wherein the hydraulic valve lifter system comprises an actuator (40) which is functionally connected to the hydraulic pump (25) for the relative rotation between the pump cylinder

(11) and the lifter piston (12) and wherein the method comprises variably adjusting the size of the portion by controlled adjustment of the relative rotational orientation between the pump cylinder (11) and the lifter piston

(12) by the actuator (40).

8. A method according to claim 7, wherein only the pump cylinder (11) is rotational, and the actuator (40) is functionally connected only to the pump cylinder (11) for causing rotation thereof, whereas the lifter piston (12) is rotationally fixed, wherein the method comprises rotating only the pump cylinder for adjustment of the size of the portion.

9. A hydraulic valve lifter system for operating a valve system of an internal combustion engine according to anyone of the claims 1-4, wherein the hydraulic valve lifter system is camshaft-driven for hydraulic actuation of gas intake or outlet valves in a cylinder head; wherein the hydraulic valve lifter system comprises a hydraulic pump (25), which comprises a pump cylinder (11) and comprises a lifter piston (12) arranged for linear reciprocation inside the pump cylinder (11); wherein the lifter piston (12) is coupled to a cam lobe (5) of the camshaft (4) for causing the reciprocation by rotation of the camshaft (4); wherein the lifter piston (12) is configured for pressurizing hydraulic liquid in the pump cylinder (11) and pumping a portion thereof to the valve system for hydraulic actuation of the valve (8); characterised in that the hydraulic pump (25) is configured for relative rotation between the pump cylinder (11) and the lifter piston (12) and for variable adjustment of a relative rotational orientation between the pump cylinder (11) and the lifter piston (12), and wherein the size of the portion is variable by the relative rotation and determined by the relative rotational orientation.

10. A hydraulic valve lifter system according to claim 9, wherein the hydraulic valve lifter system comprises an actuator (40) which is functionally connected to the hydraulic pump (25) for the relative rotation between the pump cylinder (11) and the lifter piston (12) and configured for variably adjusting the relative rotational orientation between the pump cylinder (11) and the lifter piston (12).

11. A hydraulic valve lifter system according to claim 10, wherein only the pump cylinder (11) is rotational, and the actuator (40) is functionally connected only to the pump cylinder (11) for causing rotation thereof, whereas the lifter piston (12) is rotationally fixed.

12. A hydraulic valve lifter system according to anyone of the claims 9-11, wherein the lifter piston (12) has a first end (12A) coupled to the cam lobe (5) of the camshaft (4) and an opposite front end (31) inside the pump cylinder (11); wherein the lifter piston (12) is reciprocal between a bottom dead centre, BDC, position and a top dead centre, TDC, position, wherein the cylinder (11) comprises a reservoir (35) of the hydraulic liquid and a spill port (23) for flow of hydraulic liquid from the reservoir (35) through the spill port (23), wherein the spill port (23) is in fluid-flow connection with the reservoir (35) when the lifter piston (12) is in the BDC and disconnected from the reservoir (35) when the lifter piston (12) is in the TDC; wherein the lifter piston (12) is configured for disconnecting the spill port (23) from the reservoir (35) by its front end (31) when the front end (31) passes the spill port (23) during advance of the lifter piston (12) towards the TDC, and configured for subsequent pressurization of the hydraulic liquid in the reservoir (35) and for pumping the portion of hydraulic liquid to the valve system; wherein the front end (31) of the lifter piston (12) is profiled with a first part (31A) and a second part (31B), where the second part (31B) is offset towards the first end (12A) relatively to the first part (31A), and wherein the lifter piston (12) is arranged for passing and disconnecting the spill port (23) from the reservoir (35) during its forward advance towards the TDC only by the first part (31A) for a first rotational orientation and only by the second part (3 IB) for a second rotational orientation between the pump cylinder (11) and the lifter piston (12), and wherein the portion of hydraulic liquid to the valve system is smaller in the second rotational orientation than in the first rotational orientation for actuating the valve (8) less in the second than in the first rotational orientation.

13. A hydraulic valve lifter system according to claim 12, wherein the front end (31) of the lifter piston (12) is sloping curvedly or linearly from the first part (31A) to the second part (31B) for smooth and continuous variable adjustment of the size of the portion that is to be pumped to the valve system in dependence on a relative rotational orientation in between the first and the second rotational orientation.