WO2006074497A2

WO2006074497A2 - Method for operating an internal combustion engine

Info

Publication number: WO2006074497A2
Application number: PCT/AT2006/000021
Authority: WO
Inventors: Eduard Unger; Horst Schaffer; Gregor Schnoell; Hans Felix Seitz; Gabor Hrauda
Original assignee: Avl List Gmbh
Priority date: 2005-01-17
Filing date: 2006-01-17
Publication date: 2006-07-20
Also published as: DE112006000194A5; WO2006074497A3; DE112006000194B4

Abstract

The invention relates to a method for operating an internal combustion engine, in particular a multi-cylinder internal combustion engine with an engine brake unit, at least one brake valve being provided for each cylinder, preferably in addition to intake and exhaust valves. The brake valve opens into a common pressure container (brake rail), said brake valve being opened at least once during the engine operation, before, at the start of and/or during the compression phase of the cylinder. The aim of the invention is to regulate the braking power in the simplest possible manner. To achieve this, the braking power of the engine brake unit is controlled by modifying the control times of the brake valve, preferably the closing time of the brake valve.

Description

Method for operating an internal combustion engine

The invention relates to a method for operating an internal combustion engine, in particular a multi-cylinder internal combustion engine with an engine braking device, with per cylinder at least one, preferably in addition to intake and exhaust valves provided brake valve, which opens into a common pressure vessel (brake rail), wherein the brake valve during engine braking before, at the beginning and / or during the compression phase of the cylinder is opened at least once. Furthermore, the invention relates to a method for operating an internal combustion engine, in particular a multi-cylinder internal combustion engine, with per cylinder at least one, preferably additional to inlet and outlet valves provided, opening into the combustion chamber auxiliary valve which controls a leading to a common cylinder for several pressure vessel flow path , wherein preferably the additional valve, the flow path and the pressure vessel are part of an engine braking device, and wherein at least one additional valve in at least one engine operating region before, at the beginning and / or during the compression phase is opened at least once, and an internal combustion engine for performing the method. Furthermore, the invention relates to a valve actuating device, in particular for a reciprocating piston engine, with at least one actuator actuating at least one lifting valve, wherein an actuating piston of the actuator is connected to a valve stem of the lifting valve.

In vehicle engines, in particular commercial vehicle engines, integrated braking systems are becoming increasingly important, since these systems are cost-effective and space-saving additional braking systems. However, increasing the specific power of modern commercial vehicle engines also requires increasing the braking power to be achieved.

An engine brake is known for example from DE 34 28 626 A. Therein a four-stroke internal combustion engine is described, which comprises two cylinder groups, each with four cylinders. Each cylinder has charge exchange valves and an additional exhaust valve, wherein in the brake operation, the additional exhaust valves are open during the entire braking process. Furthermore, in the common exhaust port of the two cylinder groups arranged on a shaft rotatably mounted throttle valve whose position via a control rod by an actuator can be influenced. A disadvantage of this known system is the dependence on the speed, in particular a relatively low braking power in the lower speed range. Furthermore, DE 25 02 650 A shows a valve-controlled reciprocating internal combustion engine, in which during the braking process compressed air is conveyed via a compressed air valve into a storage tank and returned to the work performance when starting on the same compressed air valve.

From EP 0 898 059 A, a decompression valve engine brake is known in this context with which a compressed air generator can be realized for all operating states of the internal combustion engine. In this case, a compressed air tank of a compressed air system is filled via a bypass line with compressed gas from the combustion chamber of the cylinder. One or more cylinders can be used to supply the compressed air system.

From EP 0 828 061 A an engine brake is known in which a gas exchange between the individual cylinders is made possible via the common exhaust gas collecting pipe. The gas exchange takes place via the exhaust valves of the six-cylinder internal combustion engine. A disadvantage of this engine brake is, among other things, the relatively low recoverable brake pressure.

From AT 004 963 Ul a multi-cylinder internal combustion engine is known, which in addition to the intake and exhaust valves per cylinder has a brake valve. All brake valves of the internal combustion engine open into a common, tubular pressure vessel, so that upon actuation of the brake valves, a gas exchange between the individual cylinders of the internal combustion engine is possible. The tubular pressure vessel has a pressure control valve, which can be acted upon by control signals in dependence on the position of a brake switch or brake pedal.

From DE 25 02 650 A a valve-controlled reciprocating internal combustion engine is known in which compressed air is conveyed via a compressed air valve into a storage tank during braking and returned to the work performance when starting on the same compressed air valve.

From DE 1 401 223 Al it is known to convert a reciprocating internal combustion engine for vehicles in the braking periods in an outside air-sucking compressor and to promote the compressed air in a storage vessel to then, with appropriate savings of fuel to use to drive the engine.

Furthermore, DE 199 773 C discloses a reciprocating internal combustion engine in which compressed air is conveyed into a storage vessel during the braking process and is used partly for starting, partly for increased braking of the internal combustion engine. The control of the compressed air valve is effected by an additional camshaft, which operates the compressed air valve via rods and levers.

In internal combustion engines with an exhaust gas turbocharger, in the event of a sudden increase in load, the fuel injection quantity must be reduced until sufficient charge pressure has built up to prevent inadmissible soot emissions. Depending on the tuning of the turbocharger system, it may therefore take a few seconds for the fuel injection quantity requested by the driver to be released by the engine electronics due to the inertia of the turbocharger rotor ("turbo-hole").

From JP 2004-084670 A a valve actuator for a lift valve of an internal combustion engine with a hydraulic double-acting actuator is known. The valve stem of the lift valve is connected to a piston adjacent to pressure chambers, which is displaceable by pressurizing the pressure chambers between two end positions.

For manufacturing reasons, it is advantageous to perform the actuating piston and valve stem as separate parts. The problem is, however, the connection of the actuator on the valve stem, as it can lead to tension in the binding area as a result of manufacturing tolerances.

The object of the invention is to carry out a regulation of the braking power in the simplest possible way. Furthermore, it is an object of the invention to improve the response of the internal combustion engine at sudden load increase. It is another object of the invention to propose a less production-sensitive valve actuator.

According to the invention this is achieved in that the braking power of the engine braking device is controlled by changing the timing of the brake valve, preferably the closing time of the brake valve. If the brake valve can be variably actuated, a separate pressure regulating valve for controlling the pressure in the pressure vessel can be dispensed with.

A particularly simple regulation of the braking power can be achieved if the closing time is determined as a function of the engine speed. The closing time can also be determined depending on the pressure and / or the temperature in the pressure vessel. The inclusion of the pressure and / or the temperature of the pressure vessel is therefore advantageous because by closing the brake valve at a wrong time, the temperature and pressure in the pressure vessel could rise too high and thus lead to a mechanical destruction of the brake system. Therefore, a limitation of the closing time in the direction of the top dead center of the ignition is necessary. This limitation can most easily be realized by maps which contain the corresponding limit values as a function of the engine speed, the pressure and / or the temperature in the pressure vessel. Another possibility is to measure the pressure and / or the temperature in the pressure vessel and supply regulators, which change the closing time in the direction of lower braking power when exceeding a certain limit for the pressure or temperature and thus bring about a reduction in pressure and temperature stress , If you want to increase the braking power at a speed, so only the closing time must be adjusted in the direction of early. A regulation or increase of the braking power is also possible if the closing time - starting from early - is adjusted to late. Thus, by modulating the closing time of the brake valve in a very simple manner, an increase or a reduction of the braking power possible.

In order to obtain sufficient braking power over longer braking distances, it is advantageous if the pressure vessel is cooled by a cooling device.

The control can be significantly simplified if individual cylinders are combined into cylinder groups.

The response of the engine in case of sudden load increase can be improved if the additional valve is opened at least one fired cylinder at sudden increase in load of the internal combustion engine, so that pre-compressed air or gas is blown into the combustion chamber. Preferably, it is provided that the additional valve is opened in a range between about 520 ° crank angle to 580 ° crank angle.

By this measure, the "turbo-hole" can be bridged, wherein at least one cylinder under elevated pressure gas or compressed air from the pressure vessel is supplied until the exhaust gas turbocharger has built up sufficient boost pressure to avoid inadmissible soot emissions.

If the storage volume of the pressure vessel is insufficient to bridge the "turbo-hole", it can be provided in an advantageous further development of the invention that compressed air is applied to the pressure vessel by means of an external compressor. This is of particular interest in vehicles, where an external compressor, for example, for a compressed air assisted braking system, is already provided.

A simple filling of the pressure vessel is possible if the additional valve of at least one cylinder in engine operating ranges with lower load request or in deceleration operation in a range between about 540 ° crank angle to 720 ° crank angle, preferably between 570 ° crank angle to 690 ° crank angle is opened to the pressure vessel laden with charge air.

The additional valve for controlling the flow connection to the pressure vessel is advantageously controlled by an electronic control unit, which processes signals via the charge pressure built up by the exhaust gas turbocharger. For this purpose, it can be provided that the charge pressure is determined via a pressure sensor in the intake line. Alternatively or additionally, the speed of the exhaust gas turbocharger can be determined via a corresponding speed sensor.

A little production-sensitive valve actuator can be achieved when the actuating piston acts on at least one valve member with at least one valve plug on the valve stem, wherein the actuating piston acts on at least one conical outer surface of at least one conical outer surface of at least one valve plug, and wherein the valve plug at least in the direction of the actuating piston facing valve stem end tapers.

Valve taper pieces are used in conventional mechanically actuated lift valves to support the valve spring plate against the force of the valve spring on the valve stem. The valve plug pieces are arranged in one or more circumferential grooves in the region of one end of the valve stem so that the conical surface widens towards the shaft end. Tapered inner surfaces of the valve spring plate are pressed by the valve spring against the conical outer surfaces of the valve plug pieces, so that they are pressed captive by the valve spring force into the annular groove of the valve stem.

In the subject invention, the valve cone assembly is arranged in the opposite direction compared to mechanical valve actuators, so that the conical surfaces taper to the shaft end. The valve cone pieces are pressed, in particular by the opening force of the actuating piston, and not or less by a closing force of a valve spring, in at least one receiving groove of the valve stem. The link is centered on the actuating piston, with any tolerances compensated become. The tolerance chain between valve stem and actuator is thus reduced.

Preferably, it is provided that with the actuating piston a pushed onto the valve stem mounting sleeve is detachably connected, wherein the inner diameter of the mounting sleeve is larger than the Ventilschaftdurch- diameter and smaller than the largest diameter of the connecting member, so that the mounting sleeve facing away from the actuating piston Side of the connecting member engages positively. The actuating piston is screwed into the mounting sleeve, wherein the mounting sleeve is pressed from the side facing away from the actuating piston against the connecting member. Thus, the actuating piston is firmly connected to the valve stem in the closing direction of the lift valve.

In order to prevent the latter from rotating when the actuating piston is screwed into the fastening sleeve, it is advantageous if the fastening sleeve has at least one engagement surface preferably formed by at least one recess or opening for at least one assembly tool at the edge or on an outer surface. Furthermore, it is favorable to facilitate assembly when the mounting sleeve has an annular receiving groove for the connecting member on the side facing the connecting member. Thus, the valve plug pieces can be inserted into the receiving groove of the pushed over the valve stem mounting sleeve. A step in the edge region of the groove ensures the correct arrangement of the valve plug pieces in the predetermined position and prevents their slippage during assembly. Thereafter, the actuating piston is pushed over the connecting member on the valve stem, wherein the valve plug pieces are pressed against the valve stem.

The actuating piston can be designed as a double piston, which acts on the valve stem in both the opening, as well as in the closing direction. As a result, the valve spring can be omitted. In order to force the return of the lift valve in a safe closed position in the event of a fault in the valve operating system, however, a valve spring acting on the lift valve in the closing direction is advantageous. It can be provided that the fastening sleeve is formed by a valve spring plate on which preferably engages a valve spring in the closing direction of the lift valve.

The stroke of the actuating piston and the lifting valve is expediently limited, wherein preferably the stroke limiter is formed by the fastening sleeve, which rests at maximum stroke on a stop preferably formed by the cylinder head. In the case of a valve spring retainer, the stroke limitation may be formed by the edge of the valve spring retainer. The elasticity of the valve spring plate can be optimized in terms of optimal damping and desired acoustic behavior.

To allow easy mounting of the valve plug pieces, the valve plug pieces and / or the valve stem can be magnetised in the area of the valve plug pieces.

The invention will be explained in more detail with reference to FIGS. Show it:

Fig. 1 is a schematic representation of an internal combustion engine with an engine braking device;

2 shows the pressure curve of the internal combustion engine in a p, V diagram for the braking operation.

3 shows the pressure curve of the internal combustion engine in a p, V diagram for the fired operation;

Fig. 4 is a cylinder pressure-crank angle diagram for the braking operation;

FIG. 5 shows characteristic parameters as a function of the closing time for an operation; FIG.

FIG. 6 shows characteristic parameters as a function of the closing time of the brake valve for another operating point of the internal combustion engine; FIG.

7 shows the strategy for the activation of the engine brake;

8 shows a schematic representation of an internal combustion engine according to the invention;

9 is a diagram with the valve lift of the additional valve in millimeters,

10 shows the valve actuating device in a longitudinal section in a first embodiment variant;

FIG. 11 shows the valve actuating device in a side view according to the arrow XI in FIG. 10; FIG.

FIG. 12 the valve actuating device in a longitudinal section in a second embodiment variant; FIG. and

13 shows the valve actuating device in a side view according to the arrow XII in FIG. 12. The invention will be explained using the example of a 6-cylinder internal combustion engine. It should be noted, however, that the method according to the invention is independent of the number of cylinders. The structure of the engine brake system for the internal combustion engine 1 is shown in FIG. Reference numeral 2 denotes the injection system, which will not be discussed further here.

Per each cylinder C ₁ , C ₂ , C ₃ , C ₄ , C ₅ , C ₆ each opening into the combustion chamber brake valve 10 is provided. So that the internal combustion engine 1 can also be used in braking mode, the brake valves 10 arranged in addition to conventional intake and exhaust valves (not shown) can be operated via a control unit 4. The intake and exhaust valves of the internal combustion engine 1 are conventionally controlled via camshafts. The brake valves 10 in the combustion chamber are hydraulically operated, that is, there is a hydraulic intermediate circuit 12, with oil tank 12 a, pump 12 b, oil distribution line 12 c, pressure sensor 12 d and shut-off valve 12 e, which is responsible for the actuation of the brake valves 10. For each cylinder C ₁ , C ₂ , C ₃ , C ₄ , C ₅ , C ₆ , a respective hydraulic line 14 opens to the respective brake valve 10, wherein a hydraulic valve 16 is arranged in each hydraulic line 14. Each hydraulic valve 16 is controlled by the combined engine brake control unit 4, whereby the braking power P _B can be adjusted continuously as a function of the desired braking power.

In order to change with the internal combustion engine 1 from the fired operation in the braking mode, the injection of the injection system 2 must first be deactivated. Subsequently, a pressure in the pressure vessel 18 (brake rail) is established by the hydraulic valves 16. In steady state braking, that is, after a few engine cycles, a certain gas pressure in the pressure vessel 18 is set. This gas pressure is mainly determined by the start of control, the drive time, as well as by the drive end of the hydraulic valves 16. During braking operation, the additional brake valve 10 is opened in the compression stroke of the internal combustion engine 1, as shown in FIG. 4 can be seen. As a result, the air or the gas flows from the pressure vessel 18 into the respective cylinder Ci, C ₂ , C ₃ , C ₄ , C ₅ , C ₆ . This causes that already at the beginning of the compression phase, when a connection of the combustion chamber with the pressure vessel 18 prevails, a higher pressure in the cylinder Ci, C ₂ , C ₃ , C ₄ , C ₅ , C ₆ prevails. If there were no connection to the pressure vessel 18, the boost pressure in the intake manifold of the internal combustion engine 1 would determine the pressure level in the cylinder C 1, C ₂ , C ₃ , C ₄ , C ₅ , C ₆ at the beginning of the compression phase. Due to the increased initial pressure or the increased filling in the pressure vessel 18, a higher compression work is necessary in the compression stroke. This increased compression work can be used, for example, for braking a vehicle or a moving mass. At 20, the brake pedal is indicated by 21 the accelerator pedal of the vehicle.

Fig. 2 and Fig. 3 show p, V plots with the cylinder pressure p and the cylinder volume V for the towed operation (Fig. 2) and the fired operation (Fig. 3) to show the pressure ratios and the work done.

In fired operation (FIG. 3), the high pressure phase provides a positive, the low pressure phase a negative contribution to the charge cycle. This sum of the two surfaces, which result in the high-pressure and low-pressure part corresponds to the piston work. In fired operation, this results in a positive torque after the deduction of all losses, which is available at the crankshaft.

In braking mode or in trailing mode (FIG. 2), the cylinder C 1, C 2, C 3, C 4, C 5, C 6 has a negative contribution to the charge cycle loop both in the high-pressure part and in the charge change part. The sum of the two surfaces corresponds to a piston work, which is available as a negative torque on the crankshaft. This braking torque is determined primarily by the parameters of the control strategy of the additional brake valves 10 in the combustion chamber. The braking torque can thus be applied directly from a fired cycle to a braking cycle upon actuation of the engine brake.

Fig. 4 shows a typical control strategy for a braking operation. The curve p describes the pressure in the cylinder Ci, C ₂ , C ₃ , C ₄ , C ₅ , C ₆ , the curve P _r the pressure in the pressure vessel 18 for a 6-cylinder internal combustion engine. Clearly recognizes the pulsations of the pressure p _r in the pressure vessel 18 with a firing interval of 120 ° crank angle KW. Bars O and I show the timing for the exhaust and intake valves, respectively. The bar for the brake valve 10 is designated B. The brake valve 10 opens in this specific embodiment at about 550 ° crank angle KW to the top dead center of the ignition ZOT and closes at about 30 ° crank angle KW to the top dead center of the ignition ZOT.

In order to be able to influence the braking torque or the braking power in the case of a braking request, the relationship between the braking power P _B and the closing angle or closing time α in ° crank angle KW of the brake valve 10 at different rotational speeds ni and n ₂ , wherein the rotational speed ni in the operating point shown in Fig. 5, for example, is smaller than the rotational speed n _{2 of} the driven in Fig. 6 operating point. It can be seen that a maximum braking power P _B or a maximum pressure P _r in the pressure vessel 18 in FIG. 5 is established at a closing time α of approximately 38 ° crank angle KW after the top dead center of the ignition ZOT. If the closing time is retarded, the brake performance according to the relationship shown with the closing time α (also decreases in the specific case, the braking power P _B at a shift in the direction of "early"). This relationship can be used in the brake control to adjust the braking power P _B according to the driver's request. As can be seen in FIG. 6, however, certain boundary regions must be maintained, especially at low speeds, so that there are no impermissibly high pressures p _r or temperatures T _r in the pressure vessel 18.

Figure 7 shows a simple structure for the realization of the brake control. The driver passes by means of brake pedal 20 his brake request a _b to the controller. By way of the position of the brake pedal, the desired braking power P _B or the desired braking torque M _b dependent on the respective engine rotational speed n through the characteristic map KFM _{b is} read into the brake control device 4. The opening time α ₀ is determined via a characteristic Ka ₀ , which is plotted against the engine speed n. The pilot control of the closing of the brake valve 10 can be determined via the relationships shown in FIGS. 5 and 6. The parameters for the map KFα _c can thus be parameterized in this context. Since the closing at an incorrect time, the temperature T ₁ - and the pressure p _r in the pressure vessel 18 increase too much and thus cause mechanical destruction of the brake system, the closing time α _c must be adjusted to the speed n of the engine 1. In order to avoid destruction of the brake system, it is expedient to limit the closing time α _c in the direction of the top dead center of the ignition ZOT. This limitation can most easily be realized by maps which include the corresponding limit values as a function of the engine speed n, the brake pressure p _r in the pressure vessel 18 and the temperature T ₁ - in the pressure vessel 18. Alternatively, it can also be provided that the pressure p _r and / or the temperature T ₁ - measured and compared with a maximum value p _rm ax, T _rmax , as shown in Fig. 7. If these values exceed a limit value, the regulators PCTRL and TCTRL reliably ensure that the closing time point α _{c is} shifted in the direction of lower braking power, which leads to a reduction in the pressure and temperature load. Due to the controller PCTRL or TCTRL a determined by the map KFα _c base value α _c0 for the closing time of the brake valve 10 by a size Δα _cp or Δα _cT changed, resulting in the final closing time α _{c of} the brake valve 10.

If the braking power P _{B is to be increased} at a certain speed n, then only the closing time α _{c has} to be adjusted in the direction of early. Thus, by modulating the closing time α _{c of} the brake valve 10, an increase or a decrease in the braking power P _{B is} possible. The prerequisite for the correct functioning of the system is that the resulting heat loss in the cylinders C ₁ , C ₂ , C ₃ , C ₄ , C ₅ , C ₆ and in the pressure vessel 18 can be discharged accordingly. In the concrete embodiment, the power loss in the cylinder C ₁ , C ₂ , C ₃ , C ₄ , C ₅ , C _{6 was} discharged via the cooling water and in the pressure vessel 18 via an additional heat exchanger (not shown). If the cooling is not sufficient, the braking power P _{B is} automatically reduced in the manner described above, to avoid any overheating of the system.

In FIG. 8, the method according to the invention is explained in more detail with reference to a 6-cylinder turbocharged engine, wherein it should be noted that the method is independent of the number of cylinders and is also used, for example, in 5, 8 or 12-cylinder engines can come.

The six cylinders C 1, C 2, C 3, C 4, C 5, C 6 are connected via inlet ducts (not shown) to an intake manifold 102, which is supplied with charge air from the air filter 103 via the compressor part C of the turbocharger 104 and via the intercooler 105. The exhaust valves of the internal combustion engine 101 open into the exhaust system 106, wherein the exhaust gases are routed in a conventional manner via the turbine part T of the turbocharger 104 and exit via a silencer 107.

Internal combustion engine 101 has an engine brake device 108 with a tubular pressure vessel 109 (brake rail) into which flow paths 111 formed by auxiliary valves 110 pass through channels, so that a gas exchange between the individual cylinders Cl, C2, C3, C4, C5 , C6 is possible at a relatively high pressure level. Constructively, the pressure vessel 109 can be integrated directly into the cylinder head of the commercial vehicle engine or designed as an external connecting pipe similar to an inlet or outlet container. The actuation of the additional valves 110 preferably takes place electro-hydraulically and can therefore be varied as desired. During braking operation of the internal combustion engine, the brake valves 110 are actuated several times per operating cycle of the engine, for example, two braking strokes per working cycle, wherein the first brake stroke is close to the top dead center of the high-pressure cycle. During this braking stroke, highly compressed air from one of the cylinders C1, C2, C3, C4, C5, C6 enters the pressure vessel 109. As a result, on the one hand the pressure vessel 109 is filled with compressed air (up to about 30 bar operating pressure), on the other hand reduces the expansion of the cylinder, whereby braking power is produced. Shortly after closing the inlet valve, the additional valve 110 opens again, whereby compressed air flows from the pressure vessel 109 into the combustion chamber. As a result of the second brake stroke, the cylinder pressure at the beginning of the compression phase of the high-pressure cycle increases to the pressure level of the pressure vessel 109. This increases the compression work to be applied and thus again the braking power of the engine.

In order to improve the response of the internal combustion engine 101 in the event of a sudden increase in load or a load jump, the additional valves 110 are actuated not only by the brake pedal during a deceleration request, but also by the control electronics 116 during a sudden load request. It is provided that the additional valve 110 is at least one cylinder Cl, C2, C3, C4, C5, C6 opened after the intake stroke at the beginning of the compression phase, so that pre-compressed air or gas can flow from the pressure vessel into the combustion chamber. The additional valve 110 is opened in a range between 520 ° crank angle to 580 ° crank angle after the top dead center of the ignition. About pressure and / or speed sensors 120, 121 is monitored whether the exhaust gas turbocharger 104 can build the required for the sudden load request increased boost pressure. As soon as this is the case, the additional valves 110 are closed by the control unit 116.

In order to build up the pressure in the pressure vessel 109, one or more auxiliary valves 110 can be opened by the control unit 116 toward the end of a compression stroke, preferably in a range between 540 ° crank angle KW to 720 ° CA, as in the unloaded state of the internal combustion engine 102 is shown by the dashed line V ₂ in Fig. 9. The solid line Vi shows the opening stroke of the additional valves 110 during the bypass of the turbo-hole.

If the volume of the pressure vessel 109 is insufficient for bridging the turbo-hole of the exhaust-gas turbocharger 104, the pressure vessel 109 can be connected via a pressure line 122, in which an additional valve 123 is arranged, to a pressure vessel 124, which is fed by an external compressor 125 is. This is particularly advantageous when the vehicle is already em Dϊuckϊüft-Bbrdnetz has for other purposes ^".

According to the present proposal, an existing engine brake device with additional valves 110 and pressure vessel 109 can be used for bridging the turbo-hole of an exhaust gas turbocharger 104 with extremely little effort. The type of such synergistic effect results in a better utilization of the engine brake device, without other complex design changes would be required.

A valve actuating device 201 has an actuator 202 which is hydraulically actuated in the exemplary embodiment and has an actuating piston 204 which is displaceably mounted in an actuator cylinder 203. The actuating piston 204 is releasably connected to the valve stem 205 of a lift valve 206, such as a gas exchange valve of an internal combustion engine, connected. The connection is made by a connecting member 208 having at least two valve plug pieces 207 and a fastening sleeve 209.

In this case, the actuating piston 204 has, on its end face 210 facing the valve stem 205, a conical inner surface 211 which widens toward the end face 210. The inner surfaces 211 abut against the outer surfaces 212 of the link 208. The valve plug pieces 207 are arranged with their correspondingly shaped inside in the exemplary embodiment in three annular grooves 213 of the valve stem 205 so that the conical outer surfaces 212 to the actuating piston 204 facing the end 214 of the valve stem 205 tapers. By the conical surfaces 211, 212, the valve plug pieces

207 thus pressed into the annular grooves 213 when the actuating piston 204 acts in the opening direction on the lift valve 206.

The actuating piston 204 has in the region of its end face 210 an external thread 215 which is screwed into a corresponding internal thread of the fastening sleeve 209. The pushed onto the valve stem 205 mounting sleeve 209 has an inner diameter di, which is greater than the valve stem diameter d ₂ , but smaller than the maximum diameter D of the connecting member 208. A shoulder 216 of the fastening sleeve 209 in this case comprises the connecting member 208 in a region remote from the actuating piston 204. As a result, the actuating piston 204 via the mounting sleeve 209 and the connecting member 208 in the closing direction of the lift valve 206 with the valve stem 205 is firmly connected.

The actuating piston 204 is advantageously designed to be double-acting and is adjacent to at least two working spaces 217, 218, which can be acted upon alternately by pressure of a hydraulic system not shown further. As a result, both an opening force value and a closing force can be applied to the lift valve 206. A valve spring 219 for closing the lift valve 206 would thus not necessarily be required. FIGS. 12 and 13 show in this respect a valve springless variant embodiment. However, a valve spring 219 has the advantage that in case of malfunction, such as failure or non-operation of the hydraulic system, the lift valve 206 is brought into a defined closed position, so that consequential damage can be avoided in the internal combustion engine. In the embodiment shown in FIGS. 10 and 11, a valve spring 219 acts via a valve spring retainer 220 and the connecting member

208 on the valve stem 205 in the closing direction. In the exemplary embodiment, the valve spring retainer 220 also simultaneously forms the fastening sleeve 209 connected to the actuating piston 204. In order to limit the maximum opening stroke of the lift valve 206, the attachment sleeve 209 forms a stroke limiter 221, which rests against a stop 222 formed by the cylinder head 223 at maximum lift. In Fig. 10, the stroke limiter 221 is formed by the edge 224 of the valve spring retainer 220.

In order to facilitate assembly, the fastening sleeve 209 has a recess 225 formed by an annular groove in a region facing the connecting member 208 and a recess 226 in the region of the edge 224 of the valve spring retainer 220. The assembly of the valve actuator 201 is carried out as follows:

The fastening sleeve 209 is placed on the valve stem 205 counter to the force of the valve spring 219 on the valve stem 205 so that the valve plug pieces 207 can be inserted into the annular grooves 213. Upon relieving the valve spring 219, the valve plug pieces 207 come to lie in the recess 225 and are held by the edge of the recess 225 in position, so that an accidental falling out of the fixed position during assembly is prevented. Thereafter, the actuating piston 204 is screwed onto the mounting sleeve 209, wherein the inner surfaces 211 of the actuating piston 204 are pressed against the outer surface 212 of the connecting member 208. This results in a firm connection between the valve stem 205 and the actuating piston 204. By the valve plug pieces 207 tolerances between the valve stem 205 and the actuating piston 204 are compensated, so that tolerance-related mechanical stresses can be largely avoided. As a result, an unobstructed rotational movement of the lift valve 206 is ensured. To prevent inadvertent rotation of the mounting sleeve 209 during the screwing of the actuating piston 204, the mounting sleeve 209 is secured against rotation during assembly. This happens because a hook-like tool is introduced through an opening 227 in the space 228 and hooked in the recess 226. After installation, the tool is removed again. After screwing on the actuating piston 204, a closing part 229 is mounted in the housing 231 in the area of the end face 230 of the actuating piston 204 facing away from the valve stem 205.

In order to facilitate the insertion of the valve plug pieces 207, these and / or the valve stem 205 may be magnetized in the region of the annular grooves 213.

If necessary, the damping can be influenced via the elasticity of the valve spring retainer 220 in order, for example, to optimize the acoustic behavior of the valve actuating device 201. The actuator may be hydraulic, pneumatic, electromagnetic or piezo-magnetic or mechanical. The invention is particularly suitable for the actuation of gas exchange valves of internal combustion engines, but is also advantageous for other Betätigungsseinrichtrungen applicable for globe valves.

Claims

PATENT APPLICATIONS

1. A method for operating an internal combustion engine, in particular a multi-cylinder internal combustion engine with an engine braking device, with per cylinder at least one, preferably in addition to intake and exhaust valves provided brake valve which opens into a common pressure vessel (brake rail), wherein the brake valve in engine braking operation before, at the beginning and / or during the compression phase of the cylinder is opened at least once, characterized in that the braking power of the engine braking device by changing the timing of the brake valve, preferably the closing time of the brake valve is controlled.

2. The method according to claim 1, characterized in that the closing time is determined in dependence on a predetermined desired braking power and / or a predetermined desired braking torque.

3. The method according to claim 1 or 2, characterized in that the closing time is determined in dependence of the engine speed.

4. The method according to any one of claims 1 to 3, characterized in that the closing time is determined in dependence on the pressure and / or the temperature in the pressure vessel.

5. The method according to any one of claims 1 to 4, characterized in that is adjusted for an increase in the braking power of the closing time to early.

6. The method according to any one of claims 1 to 4, characterized in that is adjusted to control and / or increase the braking power of the closing time starting from early to late.

7. The method according to any one of claims 1 to 6, characterized in that the opening time of the brake valve is determined in dependence on the engine speed.

8. The method according to any one of claims 1 to 7, characterized in that based on the braking power and / or the desired torque merten torkennfeldabhängig a base value for the closing time is determined.

9. The method according to claim 8, characterized in that the base value for the closing time in dependence on the engine speed, as well as in Dependence of the pressure and / or the temperature in the pressure vessel in the direction of the top dead center of the ignition is limited.

10. The method according to claim 8, characterized in that the base value is increased or reduced in dependence on the pressure and / or the temperature in the pressure vessel.

11. The method according to any one of claims 1 to 10, characterized in that individual cylinders are combined into cylinder groups.

12. Control device for controlling the braking power (P _B ) of an engine braking device of an internal combustion engine (1), which per cylinder (Ci, C ₂ , C ₃ , C ₄ , C ₅ , C ₅ ) at least one, preferably in addition to intake and exhaust valves provided brake valve (10), which opens into a common for all cylinders (C ₁ , C ₂ , C ₃ , C ₄ , C ₅ , C ₆ ) pressure vessel (18) (brake rail), wherein the brake valve (10) in the engine brake operation before, at the beginning and / or during the compression phase of the cylinder (Ci, C ₂ , C ₃ , C ₄ , C ₅ , C ₆ ) is opened at least once, characterized in that the control means comprises a means for determining the closing time ( α _c ) of the brake valve (10) in dependence on a predetermined desired braking power (P ₆ ) and / or a desired braking torque (M _B ), and as a function of the engine speed (n).

13. Control device according to claim 13, characterized in that it has a means for determining the closing time of the brake valve (10) as a function of the pressure (p _r ) and / or temperature (T _r ) in the pressure vessel (18).

14. A method for operating an internal combustion engine, in particular a multi-cylinder internal combustion engine, with per cylinder at least one, vorzugsswei _. se _.. additionally, to inlet and AuslassventilerL provided, in ._ the combustion chamber opening additional valve which controls a leading to a common cylinder for a plurality of pressure vessel flow path, preferably the additional valve, the flow path and the pressure vessel are part of an engine braking device, and wherein at least one additional valve in at least one engine operating range before, at the beginning and / or during the compression phase is opened at least once, characterized in that the additional valve is opened at least one fired cylinder at sudden load increase of the internal combustion engine, so that pre-compressed air or gas from the Pressure vessel is blown into the combustion chamber.

15. The method according to claim 14, characterized in that the additional valve is opened in a range between about 520 ° crank angle to 580 ° crank angle.

16. The method according to claim 14 or 15, characterized in that the pressure vessel is acted upon by means of an external compressor with compressed air.

17. The method according to any one of claims 14 to 16, characterized in that the additional valve of at least one cylinder in engine operating ranges with lower load request or in deceleration operation in a range between about 540 ° crank angle to 720 ° crank angle, preferably between 570 ° crank angle to 690 ° crank angle is opened to load the pressure vessel with charge air.

18. internal combustion engine (101), in particular multi-cylinder internal combustion engine, with per cylinder (C1, C2, C3, C4, C5, C6) at least one, preferably provided to intake and exhaust valves, opening into the combustion chamber additional valve (110), which a to a for several cylinders (Cl, C2, C3, C4, C5, C6) common pressure vessel (109) leading flow path (111) controls, preferably the auxiliary valve (110), the flow path (111) and the pressure vessel (109) part an engine brake device (108), and wherein at least one additional valve (110) can be opened at least once in at least one engine operating region, characterized in that the additional valve (110) at least one cylinder (Cl, C2, C3, C4, C5, C6) is activated at a sudden increase in load of the internal combustion engine, so that pre-compressed air from the pressure vessel (109) is inflatable into the combustion chamber.

19. Internal combustion engine (101) according to claim 18, characterized in that the additional valve (110) in a range between 520 ° to 580 ° crank angle (KW) can be activated.

20. Internal combustion engine (101) according to claim 18 or 19, characterized in that the pressure vessel (109) via a pressure line (122) with a compressor (125) is connected.

The internal combustion engine (101) according to any one of claims 18 to 20, characterized in that the auxiliary valve (110) of at least one cylinder (Cl, C2, C3, C4, C5, C6) in an engine operating region with a lower load in a range between 540th ° crank angle (KW) to 720 ° crank angle (KW), preferably in a range between 570 ° crank angle (KW) to 690 ° crank angle (KW) can be activated to load the pressure vessel (109) with charge air.

22. Valve actuating device (201), in particular for a reciprocating piston engine, with at least one at least one lift valve (206) actuated actuator (202), wherein an actuating piston (204) of the actuator (202) with a valve stem (205) of the lift valve (206) characterized in that the actuating piston (204) via at least one connecting member (208) with at least one valve plug (207) acts on the valve stem (205), wherein the actuating piston (204) via at least one conical inner surface (211) on at least one conical outer surface (212) of at least one valve plug piece (207) engages, and wherein the valve plug (207) tapers at least in the direction of the actuating piston (204) facing the valve stem end (214).

23 valve actuating device (201) according to claim 22, characterized in that with the actuating piston (204) on the valve stem (205) deferred mounting sleeve (209) is detachably connected, wherein the inner diameter (di) of the mounting sleeve (209) is larger as the valve stem diameter (d ₂ ) and smaller than the largest diameter (D) of the connecting member (208), so that the fastening sleeve (209) on the actuating piston (204) facing away from the form-fitting engagement on the connecting member (208).

24. Valve actuating device (201) according to claim 23, characterized in that the fastening sleeve (209) by a valve spring plate (220) is formed on which preferably a valve spring (220), particularly preferably in the closing direction of the lift valve (206) engages.

25. Valve actuating device (201) according to one of claims 22 to 24, characterized in that the actuator (202) has a stroke limiter.

26. Valve actuating device (201) according to claim 25, characterized in that the stroke limiter is formed by the fastening sleeve (209), which rests at maximum opening stroke on a stop preferably formed by the cylinder head (223).

27. Valve actuating device (201) according to claim 26, characterized in that the stroke limiter is formed by the edge (224) of the valve spring plate (220).

28. Valve actuating device (201) according to one of claims 22 to 27, characterized in that the connecting member (208) is rotatably connected to the valve stem (205), wherein preferably a valve plug piece (207) in a plurality of annular grooves (213) of the valve stem (205 ) intervenes.

29 valve actuating device (201) according to one of claims 22 to 28, characterized in that the fastening sleeve (209) at the edge (224) or on an outer surface at least one preferably formed by a recess (226) or opening engagement surface for at least one assembly tool ,

30. Valve actuating device (201) according to any one of claims 22 to 29, characterized in that the fastening sleeve (209) on the said connecting member (208) facing side has a preferably circular recess (225) for the connecting member (208).

31. Valve actuating device (201) according to any one of claims 22 to 30, characterized in that the valve plug pieces (207) and / or the valve stem (205) are magnetized at least in the region of the connecting member (208).

32. Valve actuating device (201) according to any one of claims 22 to 31, characterized in that the elasticity of the valve spring plate (220) is adapted to the desired damping and acoustic behavior of the valve actuator (201).

33. Valve actuating device (201) according to any one of claims 22 to 32, characterized in that the actuating piston (204) is designed as a double piston, which acts both in the opening and in the closing direction on the lifting valve (206).

34. Valve actuating device (201) according to any one of claims 22 to 33, characterized in that the lifting valve (206) is designed valve springless.