WO2018083141A1

WO2018083141A1 - Cylinder head for an internal combustion engine

Info

Publication number: WO2018083141A1
Application number: PCT/EP2017/077997
Authority: WO
Inventors: Ennio Codan
Original assignee: Abb Turbo Systems Ag
Priority date: 2016-11-03
Filing date: 2017-11-02
Publication date: 2018-05-11
Also published as: DE102016120958A1

Abstract

A cylinder head (5) for a four-stroke internal combustion engine with supercharging is proposed. It comprises a cylinder head main body (10), at least one inlet valve (14) and an inlet duct (16) which is connected to it, wherein the inlet valve (14) has a valve head (15) and is provided in the stem region with a valve spring (18) which is supported on the cylinder head main body (10); a first pressure equalization duct (22) between the inlet duct (16) and a chamber (17) in the region of the valve stem (24) of the inlet valve (14), as a result of which boost pressure is guided into the chamber (17) during operation, which boost pressure exerts a force on the inlet valve (14) in the closing direction; wherein the inlet duct (16) has, on the valve seat, a cylindrical opening (23) which at least partially receives the valve head (15) of the inlet valve (14) in the closed state of the inlet valve (14), and wherein the inlet duct (16) remains closed during a first part of an opening movement of the inlet valve (14).

Description

CYLINDER HEAD FOR AN INTERNAL COMBUSTION ENGINE

Technical area

The present disclosure relates to an improved cylinder head for four-stroke internal combustion engines, in particular for supercharged internal combustion engines with valve control times according to the Miller method, as well as corresponding internal combustion engines.

Background of the invention

Most internal combustion engines use valves to control the gas exchange process. Typically, internal combustion engines employ poppet valves that are fully mechanically actuated, typically using camshafts. The cams of this shaft determine the opening or closing position of the valve as a function of the crank angle of the affected cylinder via its profile geometry. A kinematic chain ensures that the motion of the cam profile is transmitted to the valve. The permissible maximum acceleration in the kinematic chain, which depends on the cam profile or shape, engine speed and opening time, determines the maximum valve lift and the times for opening and closing the valves.

In the closed state, that is with no or with only minimal force of the cam, the valve is held closed by one or more springs. In general, the kinematic chain has a game, ie the parts remain unloaded. When opened, the spring should ensure that the entire kinematic chain does not lose contact as this would result in vibration, shock and uncontrolled valve movement. Critical here is the maximum negative acceleration of the valve, which determines the required spring force. The cam determines the valve movement, the spring force causes the contact is not lost, especially in negative acceleration. When closed, only the gas and spring forces act on the valve. The gas pressure in the inlet passage above the inlet valve acts on the projected annular area between valve seat and valve stem and generates a force which acts to open the valve. In contrast, the pressure in the cylinder acts on the circular surface of the valve disk, it generates a force in the direction of the valve closing. If the pressure in the inlet channel is higher than the pressure in the cylinder, there is a resulting gas force in the direction of "opening." Apart from a typically small difference between the two surfaces on the top and bottom of the valve disk, it can be assumed that this Force is proportional to the pressure difference between the intake port and cylinder. The spring force in the closed state of the valve must be greater than the resulting gas force, otherwise the valve tends to an unplanned or uncontrolled opening movement. The spring force normally corresponds to a pressure difference of 1 to 2 bar and is sufficient for conventional engines to keep the valve closed.

In order to achieve good engine efficiencies and low NO _x emissions in conjunction with knock-resistant operation, the Miller process has shown significant potential in recent years and gained in importance. In order to enable the control times for the optimal operation with extreme Miller process, the opening time of the intake valve compared to the standard opening characteristic is considerably reduced, which has the consequence, in particular with high-speed engines, that to maintain the acceleration limit and the maximum valve lift must be reduced. However, this has the consequence that pressure losses occur in the intake process, which limit the advantage of the Millerprozesses. In addition, in engines with extreme Millersteuerzeiten a negative pressure is generated in the cylinder, which can be up to 4 to 5 bar lower than the pressure in front of the intake valve in supercharged engines. This large pressure difference causes the spring force would then tend to be no longer sufficient to keep the valve closed, which can cause significant problems.

This problem could be solved simply by using stronger springs to increase the holding or closing force. But this would have several negative consequences. Thus, in general, the spring force is minimal when the valve is closed, but by increasing the spring stiffness, the spring force increases at fully open valve also at least by the same amount, although this increase in force would not be necessary with the valve open. At the same time, the surface pressure in all contact points of the kinematic chain increases, thereby increasing the friction losses of the system and also the mechanical stress on the parts. These negative effects are also present at all engine load conditions, although the larger closed-valve holding force would only be needed in the upper load range, where charging provides the highest pressures.

For the above and other reasons, there is a need for the present invention. Summary of the invention

The object of the invention is achieved by a cylinder head according to claim 1 and an internal combustion engine according to claim 10.

In one aspect of the invention, a cylinder head for a four-cycle internal combustion engine with charging is proposed. It comprises a cylinder head main body, at least one inlet valve and an associated inlet channel, wherein the inlet valve has a valve disk and is provided in the shaft region with a valve spring which is supported on the cylinder head main body; a first pressure equalization channel between the inlet channel and a chamber in the region of the valve stem of the inlet valve, whereby in operation boost pressure is conducted into the chamber, which exerts a force in the closing direction on the inlet valve. In this case, the inlet channel has a cylindrical opening on the valve seat, which at least partially receives the valve disk of the inlet valve in the closed state of the inlet valve, wherein the inlet channel remains closed during a first part of an opening movement of the inlet valve. Embodiments of the invention thus provide a system for actuating the valves which eliminates the need to increase the closed-state holding force that would otherwise be required on turbocharged, Miller-operated engines. On the one hand this is realized by the special design of the valve seat, whereby - in contrast to conventional valve trains - in the first phase of the valve movement still no opening of the valve channel takes place. For this purpose, the valve seat has a cylindrical part in which the valve disk can move without first giving a significant flow area between the cylinder and channel free, which goes beyond the permissible game. The phase of valve acceleration, which takes a relatively long time in a conventional system, is thus removed from the effective valve opening curve, so that the valve can effectively open and close at high speed.

On the other hand, according to the invention, in the shaft region of the valve, a surface, for example in the form of a disc, a plate or a piston, which is connected to the valve, is acted upon by a pressure equalization channel with the boost pressure in the inlet channel, so that the pressure generates a force closing the valve, which is opposite to the tendency of the charging pressure in the inlet channel to open the valve and largely compensates or overcompensates for this. Further features and advantages of the present invention are presented in the following detailed description of preferred embodiments of the system.

Brief description of the figures

Other features and advantages of the present invention will become apparent to those skilled in the art upon reference to the detailed description taken in conjunction with the appended drawings. Showing:

Fig. 1 shows a cylinder head according to embodiments of the invention, and an enlarged partial view thereof;

FIG. 2 schematically shows the graphs of intake and exhaust timing for cylinder heads according to embodiments in comparison with standard control times and Miller process timing; FIG.

3 is a detail view of a valve stem portion of a cylinder head according to embodiments of the invention;

Fig. 4 shows a side view of a cylindrical bush, which in the device of

Fig. 3 can be used.

Detailed description

Although preferred embodiments are described, the scope of the invention is not limited to the illustrated embodiments, but also includes embodiments obvious to those skilled in the art.

As used in this disclosure, the term "Miller process" or "Miller method" stands for an operating mode of a four-stroke internal combustion engine in which the opening time, in particular of the intake valve, is significantly shortened, ie the integrated area under the opening curve (plotted above the crank angle ) is at least 30% less than an engine with standard timing, preferably 40% or more lower. This is typically accomplished by reducing the valve lift and prematurely closing the intake valve. That is, the opening curve is no longer, as standard, largely symmetrical in relation to the duration of the intake stroke, but the inlet valve is closed again at the latest about 70%> the path length of the intake stroke. Fig. 1 shows a cylinder head 5 for a four-stroke internal combustion engine. The engine typically has a charge mechanism for the make-up gas, such as one or more exhaust gas turbochargers, a mechanical compressor, or the like. The cylinder head 5 comprises a cylinder head base body 10. In this at least one inlet valve 14 per cylinder is provided. The inlet valve 14 has a valve disk 15 and a valve stem 24. In the shaft region, the inlet valve 14 is supported on the cylinder head base body 10 via a disk 25 and a valve spring 18 attached to it, the bottom end of the spring 18 on the cylinder head body being shown in FIG. Base 10 rests. The spring is located in a chamber 17 whose upper end is formed by the disc 25. The cylindrical side walls of the chamber 17 may be formed approximately by a bush 27. The side walls of the chamber 17 may also be - except for their upper limit - part of the cylinder head body 10, that is to be integrated in this. The disk 25 can be designed differently in embodiments, in particular as part of a piston, which is explained below. The disk 25 is movable relative to the walls of the chamber 17, ie, it can move up and down in the chamber 17 together with the valve stem 24, but closes off the chamber 17 in a gastight manner. For this purpose, a seal 21 is typically attached to the outer surface of the disc 25 and the piston.

The chamber 17 is connected by a first pressure equalization channel 22 with the inlet channel 16 in fluid communication. During operation, boost pressure is conducted from the inlet channel 16 into the chamber 17 through the pressure equalization channel 22. That is, the chamber 17 is typically at the same pressure level as the inlet channel 16. In the presence of a boost pressure, this therefore also prevails in the chamber 17 and exerts a force on the inlet valve 14 via the disc 25 or over its surface. This force acts in the closing direction of the inlet valve 14 or tends to cause a closure. It thus counteracts the opening force exerted by the boost pressure in the inlet channel 16 on the upper surface 30 of the valve disk 15. Typically, the surface of the disc 25 is designed to correspond approximately to the (pressure) effective area of the upper surface 30 of the valve disk 15. Thus, the force effects of the boost pressure on the two surfaces are comparable, and it is pressed with the same amount of the valve plate 15 down and pressed the disc 25 upwards. Therefore, since the active surfaces and the respectively prevailing pressure are largely identical, it results for the intake valve 14 almost complete compensation of the two opposing gas forces caused by the boost pressure. By the embodiment described above, the necessary holding or closing force is significantly reduced by the valve springs - compared to a conventional supercharged engine without the pressure compensation. It can therefore be used smaller-sized valve springs. The reduced pressure in the kinematic chain of the valve train also reduces the frictional forces and, consequently, the drive power of the camshaft. This has a positive influence on the engine efficiency, especially in the partial load range. This improvement is even more pronounced, even if the components of the kinematic chain are dimensioned smaller. The lower surface pressures between the camshaft and valves also tend to increase the life expectancy of the moving parts in the valve train.

The described pressure compensation can be used for both intake and exhaust valves. In the case of the outlet valve 34 (see FIG. 1), it is expedient to realize the equalization as well as the inlet valve 14 by means of a second pressure equalization channel 32 which directs the pressure in the inlet channel 16 into the second chamber 37. Although compensation with gas from the exhaust duct would theoretically be possible, the gas is hot, aggressive or corrosive and can lead to deposits, which would affect the overall durability of the solution. The construction of the valvetrain at the exhaust valve 34 is otherwise practically equal to that already described for the intake valve, with a valve spring 37, a disc 45, a bushing 47, a valve stem 44, and a valve plate 35. The described reduction of the holding force is particular in embodiments advantageous in connection with the design of the valve seat of the inlet valve 14 in the cylinder head 5. The valve seat is realized such that the end of the inlet channel 16, ie whose confluence with the combustion chamber has a cylindrical opening 23. That is, the last portion of the intake passage at the transition to the cylinder is cylindrical, with a diameter corresponding to the outer diameter of the valve disk 15. The height of the cylindrical opening 23 corresponds approximately to the height of the cylindrical edge of the inlet valve. This is shown enlarged in the detail in FIG. 1. Above, that is upstream in the inlet channel, the inlet channel tapers, with a bevel, which provides a seat or stop for the outer part of the upper surface 30 of the valve disk 15. This design is applied to the exhaust valve 34 in embodiments in the same form.

The design described causes no opening of the valve channel takes place in a first phase of the valve movement. That is, the valve is moving, but the interaction The valve disk 15 and the cylindrical opening 23 initially prevents a significant flow area between the cylinder and the inlet channel, that is, beyond the existing mechanical clearance, being released. The phase of valve acceleration, which takes considerable time in a conventional system, is thus removed from the effective valve opening curve. As a result, the effective opening and closing of the valve are made at a higher speed. This is shown in FIG. 2. Curve A shows a typical course of the lift of the exhaust valve according to embodiments of the invention. Curve B, on the other hand, shows the course of the exhaust valve lift in a standard engine or a Miller engine. Curve C illustrates a typical intake valve lift according to the Miller method, while curve D shows a simplified intake valve lift according to embodiments of the invention. Curve E shows the course in a standard engine. The dashed parts of the curves A and D show the valve movement during the phase, as long as the valves according to the invention are still closed, as described above. By means of these optimized courses of the valve lift or the valve surfaces, the following advantages are expected in a non-limiting example, for example with a gas engine with premix: After opening the exhaust valve, the pressure in the cylinder is reduced faster, which reduces the piston work in the subsequent push-out phase , The valve overlap can be reduced, which leads to a reduction of the over-flushed mixing quantity. This reduces fuel consumption and HC emissions. At the end of the charge cycle, pressure and temperature are lower than in the reference case, because more work is done to the piston. One approaches the ideal Miller process, where the sucked charge expands from the boost pressure and is not throttled.

With the embodiments of the valve seat according to the invention described above, the volumes between the valve disk and the seat, which have the fluidically unfavorable shape of a circumferential groove or slot in a conventional combustion chamber, are reduced to virtually zero. As a result, better combustion efficiencies are to be expected. It is readily apparent that in embodiments, a design can be realized in which the valves in the closed state with the combustion chamber surface are practically aligned.

Fig. 3 shows a valve train according to embodiments similar to FIG. 1, wherein the connection between the pressure equalization channel 22 and the chamber 17 at a defined Valve lift of the intake valve 14 is closed. As the valve lift continues to increase, the volume of air trapped in the chamber 17 from closing acts as an air spring. For this purpose, the disk 25 can be embodied, for example, as part of an auxiliary piston 28, which closes the connection to the pressure compensation channel 22 at a certain valve stroke, the pressure equalization channel opening as a slot or circular opening in the side wall of the chamber 17. From this moment acts in the chamber 17 trapped air volume as an air spring, since the pressure with the further movement of the auxiliary piston 28 - ie further increasing valve lift - also increases. The advantage of this design is that the force of the air suspension is strongest at maximum valve lift. Especially in this situation, the highest force is required to keep the valve in contact with the kinematic drive chain. The air suspension can thus be used to dimension the actual valve springs even softer, as they can already be designed by the structure described above with pressure equalization according to embodiments anyway.

With a fixed slot geometry, the maximum force of the air suspension would depend on the charge pressure, since the pressure compensation channel is always separated from the chamber 17 at the same valve lift. Since the boost pressure generally varies with engine load, the spring force exerted by the air in this case may be too small or insufficient at low or minimum engine load. This situation can be improved if the timing at which the auxiliary piston 28 closes the pressure-equalizing passage 22 is made variable depending on the load condition. It is understood that the design of this variability can be realized in various ways. One of several possible variants, which will not be discussed in detail here, comprises a movable cover of the slot, that is the end of the pressure compensation channel. In particular, a bushing 47 may comprise a slot 48 made in the circumferential direction of the variable height bushing, as shown in FIG. 4 as a side view. The bushing is typically designed to be rotatable in this case and can be rotated depending on the load, for example by an electric or hydraulic actuator (not shown). In this way it is ensured that the time at which the auxiliary piston 28 closes the pressure equalization channel 22, is changed depending on the load condition.

According to exemplary embodiments of the invention, internal combustion engines with, among other things, the following advantages are possible: Due to the optimized course of the valve opening times, the pressure in the cylinder can be reduced more rapidly after opening the exhaust valve, which reduces the piston work in the subsequent exhaust / exhaust phase. The valve overlap can be reduced in size, which leads to a reduction of the amount of mixture flushed over and thus to a reduction in specific fuel consumption and HC emissions.

At the end of a charge cycle, pressure and temperature are lower than in the reference case of a standard engine because more work is done to the piston. It approaches the ideal Miller process, that is, the sucked charge is expanded by the boost pressure and not throttled. The standard existing volumes between valve disk and seat in the form of slots, which give the combustion chamber an unfavorable shape, are reduced to virtually zero. This leads to better combustion and also higher efficiency. Due to the design by means of pressure compensation in the shaft area, the necessary holding force is significantly reduced by the valve springs and the valve springs can be made smaller. The following reduced pressure also reduces the frictional forces and, as a result, the drive power of the camshaft, which has a positive influence on the engine efficiency, in particular in the part-load range. This influence is even more pronounced, although the components of the kinematic chain, ie cam (shaft), rocker arm, possibly rod ram or bucket tappet, can be made smaller.

Claims

claims

A cylinder head (5) for a four-cycle internal combustion engine with charge, comprising: a cylinder head base (10),

at least one inlet valve (14) and an associated inlet channel (16), wherein the inlet valve (14) has a valve plate (15) and is provided in the shaft region with a valve spring (18) which is supported on the cylinder head base (10);

a first pressure equalization channel (22) between the inlet channel (16) and a chamber (17) in the region of the valve stem (24) of the inlet valve (14), whereby in operation boost pressure in the chamber (17) is passed, the force in the closing direction the inlet valve (14) exerts;

wherein the inlet channel (16) on the valve seat has a cylindrical opening (23) which at least partially houses the valve plate (15) of the inlet valve (14) in the closed state of the inlet valve (14), and wherein the inlet channel (16) during a first part an opening movement of the inlet valve (14) remains closed.

A cylinder head according to claim 1, wherein the boost pressure in the chamber (17) exerts a closing force on the intake valve (14) via a disc (25) secured to the valve stem (24) forming the upper end of the chamber (17) is designed to at least partially compensate for a force opening the inlet valve (14) by the boost pressure in the inlet channel (16).

3. Cylinder head according to one of the preceding claims, wherein the cylinder head (5) has shortened valve inlet times by the Miller method.

4. Cylinder head according to one of the preceding claims, wherein the connection between the pressure equalization channel (22) and the chamber (17) at a defined valve lift of the inlet valve (14) is closed, and with further increasing valve lift in the chamber (17) trapped air volume acts as an air spring.

5. Cylinder head according to claim 4, wherein the connection through an opening (48) in the peripheral wall of the chamber (17) is controlled.

6. Cylinder head according to claim 5, wherein the opening (48) in the peripheral wall with an auxiliary piston (28) cooperates, which is fixed to the valve stem (24).

7. Cylinder head according to claim 5 or 6, wherein the opening (28) has a variable geometry.

8. Cylinder head according to one of claims 4 to 7, wherein the time of closing of the connection is load-dependent variable designed.

9. Cylinder head according to one of the preceding claims, further comprising an outlet valve (34) with a valve plate (35), and an outlet channel (34), and a second pressure equalization channel (32) between the inlet channel (16) and the shaft region of the outlet valve (34 ).

10. Internal combustion engine with a cylinder head according to one of claims 1 to 9.