WO2007056784A2

WO2007056784A2 - Internal combustion engine

Info

Publication number: WO2007056784A2
Application number: PCT/AT2006/000468
Authority: WO
Inventors: Paul Kapus; Reinhard Glanz
Original assignee: Avl List Gmbh
Priority date: 2005-11-15
Filing date: 2006-11-15
Publication date: 2007-05-24
Also published as: WO2007056784A3; DE112006001575A5

Abstract

The invention relates to an internal combustion engine, especially a charged internal combustion engine, comprising at least one cylinder (6) with a reciprocating piston (7) and at least two intake valves (4) per cylinder (6) as well as a roof-shaped combustion chamber cover area (8) in the cylinder head (1). The intake ducts (2) extending to the intake valves (4) generate a tumble flow (T) in the combustion chamber (5), the main directions of flow (40) of the partial flows (E) that are sucked into the combustion chamber (5) via the intake valves (4) respectively enclosing an acute angle (φ) together with a longitudinal plane (6b) determined by the cylinder axes (6a) of a cylinder bank. The inventive internal combustion engine further comprises at least one discharge duct (3) per cylinder (1). In order to reduce fuel consumption, the tumble rate ranges between 1.2 and 2, preferably between 1.4 and 1.9.

Description

Internal combustion engine

The invention relates to an internal combustion engine, in particular a supercharged internal combustion engine, having at least one cylinder with a reciprocating piston and at least two inlet valves per cylinder and a roof-shaped combustion chamber top surface in the cylinder head, the inlet ducts leading to the inlet valves generating a tumble flow in the combustion chamber, the main flow directions the sub-streams sucked into the combustion chamber via the inlet valves each include an acute angle with a longitudinal plane determined by the cylinder axes of a row of cylinders, and at least one outlet channel per cylinder. Furthermore, the invention relates to a method for operating an internal combustion engine with exhaust gas turbocharger, an intake and an exhaust line, an internal exhaust gas recirculation system and an external exhaust gas recirculation system with at least one exhaust gas recirculation line, wherein in at least one low and / or medium load range of the internal combustion engine, an internal exhaust gas recirculation and in at least a load range at the same time an internal and external exhaust gas recirculation is performed. The invention also relates to an internal combustion engine having a cylinder head with at least one inlet valve and at least one outlet valve per cylinder, wherein at least one outlet opening is at least partially surrounded by a masking. Furthermore, the invention relates to a method for producing a cylinder head with at least one inlet channel opening into a combustion chamber roof via an inlet opening, which is produced at least partially by casting technology. Furthermore, the invention relates to a cylinder head for an internal combustion engine with at least one at least partially cast technology produced via an inlet opening into a combustion chamber roof opening inlet channel.

EP 0 444 018 A1 discloses an internal combustion engine having at least two inlet ducts per engine cylinder and roof-shaped boundary surfaces, wherein the main flow directions of the partial flows sucked into the combustion chamber via the inlet ducts form an acute angle with the longitudinal engine plane. The angle of a partial flow of an inlet channel is greater by 10 ° to 40 ° than the angle of a partial flow of the other inlet channel. Both inlet channels are designed for good air flow performance. The different channel design causes a combination of tumbling and swirling motion in the combustion chamber, whereby good swirl numbers are achieved at high throughputs. From the 3P 2003-206827 A an internal combustion engine with two intake valves and two exhaust valves, a central spark plug, and a laterally arranged injection device is known. The injector has seven injection ports on the Injektorstirnseite. The injection of the inclined with respect to the cylinder axis arranged injection device is - in the region of the top dead center of the piston - directed into a piston recess of the piston. The atomized fuel is passed through the wall surface of the piston bowl around the spark plug, forming a charge stratification. The seven injection openings are arranged elliptically on the Injektorstirnseite. The fuel cloud includes an angle between 25 ° and 35 ° in a side view of the injector and an angle between 35 ° and 45 ° in a floor plan. As a result, ignition stability and good combustion up to high loads can be achieved with small quantities of fuel and smoke can be avoided in the high-speed and high-load range.

Current turbocharged internal combustion engines have relatively high Volilast range consumptions. Reason is a relatively low compression ratio (knocking), as well as a high Anfettungsbedarf in the range of full load to lower the exhaust gas temperatures upstream of the turbine to a technically feasible level.

Since the introduction of direct injection, the problem of oil dilution has increased. The injected fuel wets the cylinder wall and gets into the oil sump.

Another problem with turbo engines is the light-off behavior of the catalyst due to the thermal mass of the turbine.

From US 2003/0196646 Al an internal combustion engine with a plurality of intake and exhaust valves per cylinder is known, which has an external and an internal exhaust gas recirculation system. The exhaust gas recirculation line of the external exhaust gas recirculation system branches off the exhaust gas line upstream of the turbine of the exhaust gas turbocharger and enters the inlet line downstream of the compressor of the exhaust gas turbocharger. Depending on the load range, internal and / or external exhaust gas recirculation is carried out, internal exhaust gas recirculation being possible over the entire operating range of the internal combustion engine. Due to the intended high-pressure exhaust gas recirculation external exhaust gas recirculation is possible only in the low and medium operating range of the internal combustion engine. Also, JP 2001-214812 A discloses an internal combustion engine in which external exhaust gas recirculation is performed in the lower and middle load range and internal exhaust gas recirculation in the upper load range. At low to medium speeds results in good tuning in an internal combustion engine with turbocharging a positive pressure gradient. This means that the intake pressure in the intake manifold is higher than the exhaust back pressure upstream of the turbine. Thus, at full load but no external exhaust gas recirculation in the high pressure part is possible.

An externally cooled exhaust gas recirculation would be desirable at high loads, especially at full load, since an improvement of the knocking behavior by inert gas would lead to earlier combustion position. This would reduce the oiling requirement and thus the fuel consumption.

In order to carry out external exhaust gas recirculation at full load, a low-pressure exhaust gas recirculation system is necessary. The exhaust gas is removed after the turbine, cooled and fed before the compressor. The mixture of exhaust gas and fresh air can also be cooled via the intercooler. An internal combustion engine with a low-pressure exhaust gas recirculation system and cooling of the recirculated exhaust gas is known, for example, from EP 0 596 855 A1.

However, especially in low-pressure exhaust gas recirculation systems, it is disadvantageous that a relatively large volume is filled with air / exhaust gas mixture. This adversely affects the transient behavior of the internal combustion engine, in particular in the event of a rapid lowering of the load, for example to almost zero load. The internal combustion engine tolerates only a small amount of exhaust gas recirculation at low loads, but due to the large volume filled with air / exhaust gas mixture, it still receives too much exhaust gas over a number of combustion cycles.

Efficient Otto engine combustion is characterized in the partial load range by high stability at the highest possible exhaust gas recirculation rates, as well as at full load by compact combustion time and optimal combustion focal points, especially during charging and low speeds.

The prior art satisfies these requirements essentially by the appropriate design of the sometimes switchable intake ducts to achieve a high charge movement, which in conjunction with suitable valve timing at part load a high degree of charge dilution with exhaust gas, and at full load good residual gas purging and high charge turbulence is achieved ,

An effective and structurally simple method for low partial load consumption is the late positions of both the intake and exhaust camshafts known from the AT 3.134 Ul and the AT 4.786 U1: here, the charge-diluting residual gas flows from the exhaust during the downstroke of the piston until the exhaust valve closes back to the cylinder. By flowing back the cylinder charge into the intake tract until the inlet valve closes causes a further de-throttling of the engine during the compression stroke. The maximum beneficial residual gas quantities are limited by the intensity of the charge movement and adjusted by the amount of late setting of at least the exhaust camshaft.

The requirement for the highest possible residual gas acceptance at partial load, however, drives the required level of charge movement more than the need for compact and stable combustion at full load. Since high charge movement always requires a proportional loss of cylinder filling, conventional designs are always based on a compromise between the demands for low partial load consumption and for high performance.

From US Pat. No. 6,502,541 B2, an internal combustion engine is known in which exhaust gas can be recirculated via the outlet valves into the combustion chamber. The internal exhaust gas recirculation is carried out especially at partial load to improve fuel consumption. In order to achieve a swirl of the recirculated exhaust gas in the combustion chamber, the exhaust valves are closed at different times. The different closing times are achieved by a phase shift of the timing of the two exhaust valves. The exhaust valves are thus opened at different times. In order to adjust the exhaust valves independently, however, a high technical effort is required.

An internal combustion engine with two intake and exhaust valves per cylinder is also known from US Pat. No. 5,870,993, wherein an internal exhaust gas recirculation can be carried out from the exhaust ducts into the combustion chamber by shifting the exhaust lift curves. By means of masking in the region of both outlet openings, a swirl of the recirculated exhaust gases in the combustion chamber can be achieved. This masking of both outlet openings adversely affects the full load.

In particular, working by the Otto process suction engines require filling channels with little or no tumble flow to realize high power at high speeds. For supercharged internal combustion engines, on the other hand, a strong tumble flow in the combustion chamber is advantageous for achieving rapid combustion. The flow resistance plays a minor role. So far, so had to be used for naturally aspirated engines and supercharged engines different cylinder head shapes to produce either filling channels or tumbleerzeugende inlet channels. However, this has the disadvantage of a relatively high production costs. The object of the invention is to avoid these disadvantages and to further reduce the fuel consumption in the full load range in an internal combustion engine with turbocharger. It is also an object of the invention to lower emissions and fuel consumption, in particular in the high load range, without impairing the transient behavior with a negative load step. Another object of the invention is to enable fuel consumption improvement at part load without degrading full load performance. It is another object of the invention to reduce the production cost.

According to the invention, this is achieved in that the tumble number is between 1.2 and 2, preferably between 1.4 and 1.9.

Known internal combustion engines have significantly lower tumble numbers. In known naturally aspirated engines, tumbling numbers between about 0.3 and 0.8 are achieved, with supercharged internal combustion engines tumbling numbers between about 0.5 to 1.1. The provided in the internal combustion engine according to the invention high tumble numbers allow high turbulence in the combustion chamber and a significant reduction in fuel consumption.

Preferably, it is provided that the angles of the main flow directions of the partial flows to the longitudinal plane are determined by the shape of the inlet channels, wherein in each case the axis of the inlet channel, at least just before the inlet valve, with the axis of the inlet valve at an acute angle, preferably greater than 10 ° , The main flow directions of the two partial flows may include the same acute angle with the longitudinal plane, preferably greater than 10 °, wherein advantageously the inlet channels - viewed in the direction of the longitudinal plane - may be formed congruent.

The Tumbleströmung flows in an outlet-side cylinder segment of the combustion chamber cover surface in the direction of the piston, which extends over an angle δ smaller than 180 °, preferably between 60 ° and 120 °.

To enhance the charge movement, it is advantageous if each inlet channel immediately upstream of the valve seat ring has a molded-flow rupture edge. Each flow-off edge can be formed by an intersection of the substantially continuously extending inlet part of the inlet channel with a widening in the direction of flow, preferably conical wall portion in the region of the valve seat ring. The conical wall section is incorporated, for example, by a control cutter provided coaxially with the valve seat ring axis, wherein preferably the control router has at least one conical, cylindrical and / or curved jacket region. Furthermore, it is advantageous for enhancing the tumbling movement, if the cross section of the inlet channel is narrowed in a nozzle-like manner, at least in one direction, into the area of the flow-breaking edge.

It can be provided that the theoretical jet shape of the injection jets from the injection openings arranged on the second side intersects the valve disk of at least one open inlet valve, wherein preferably the theoretical jet shape the valve disk of the inlet valve from a valve opening between the 0.5 to 0.9 times, preferably between 0.6 to 0.8 times the maximum Ventilhubes cuts. The theoretical jet shape preferably intersects the inlet valve in a region of the circumferential line of the valve disk that is one third to one sixth of the circumference. To realize a fine atomization and to reduce fuel consumption, it is advantageous if the imaginary outer beam center lines of the fuel injection jets - viewed in a longitudinal section through the cylinder containing the injector axis - have a first beam angle between 30 ° and 50 °, preferably between 35 ° and 40 °, span, wherein preferably the outer imaginary beam center lines of the injection jets - viewed in a front view of the injector - span a second injection angle between about 40 ° and 70 °, preferably between 50 ° and 60 °.

The internal combustion engine is preferably charged by at least one exhaust gas turbocharger.

A significant reduction in fuel consumption can be achieved by storing a setpoint value for the recirculated exhaust gas quantity in a control device for the exhaust gas recirculation for each operating range, determining or calculating an actual value for the recirculated exhaust gas quantity and reducing the load from an operating range. in which internal and external exhaust gas recirculation is performed simultaneously, the internal exhaust gas recirculation is reduced or stopped until the actual value of the recirculated exhaust gas amount is less than or equal to the set value of the recirculated exhaust gas quantity at the corresponding operating point.

In particular, when using a low-pressure exhaust gas recirculation system, an internal and external exhaust gas recirculation can be performed simultaneously or only an external exhaust gas recirculation in the upper load range and / or Volliastbereich. It is preferably provided that the internally and / or externally recirculated exhaust gas is cooled between the removal from the combustion chamber and the return into the combustion chamber. In order to achieve an external exhaust gas recirculation particularly in the full load range, it is advantageous if the exhaust gas for the external exhaust gas recirculation downstream of the turbine of the exhaust gas turbocharger, preferably downstream of an exhaust aftertreatment device taken from the exhaust line and upstream of the compressor of the exhaust gas turbocharger is fed to the inlet line.

The reduction of the internal EGR amounts is most easily done by retarding the intake timing and / or by advancing the exhaust timing. Alternatively, it can also be provided that the valve lift of at least one inlet and / or outlet valve in the valve overlap area is reduced in order to reduce the internal exhaust gas recirculation quantity.

An increase in the external or internal exhaust gas recirculation rate in the upper load range can be done by additionally throttling the intake air or the exhaust gas. The control of the external exhaust gas recirculation rate can be done by a control valve and / or by throttling on the suction or exhaust side.

The regulation of the internal and external exhaust gas recirculation via a motor control unit such that the total exhaust gas recirculation rate in the internal combustion engine, which is composed of internal exhaust gas recirculation rate, empty suction of a filled with exhaust gas / air mixture volume and external exhaust gas recirculation rate, regardless of the transient operating state, the in the map stored setpoint values, or the exhaust gas recirculation amounts required to achieve the best consumption, or the best emission corresponds.

In a particularly preferred embodiment of the invention can be provided that in at least one part load operating range, preferably after a negative load step and / or below a lower load limit for the implementation of external exhaust gas recirculation of the compressor and / or the intercooler is bypassed. By bypassing the compressor, the filling and emptying behavior of the external exhaust gas recirculation system can be compensated. In the bypass line, a first obturator can be arranged. Preferably, it is provided that the bypass line upstream of the mouth of the exhaust gas recirculation line branches off from the inlet branch and re-opens into the inlet branch downstream of the compressor, preferably downstream of the charge air cooler, particularly preferably downstream of a second obturator arranged in a main flow path through the compressor. When the bypass line branches off the intake manifold downstream of the compressor, the compressor can be used to purge the exhaust gas / air mixture from the charge air cooler. For this purpose, the charge air flow can be controlled via switching devices in the region of the branch and / or the junction of the bypass line. The durchzusaugende after a negative load jump volume with air / exhaust gas mixture is thereby drastically reduced. On the other hand, an air / exhaust gas mixture is thus available again for the next high-load phase.

In this way, optimal results for consumption and emissions can be achieved without the transient behavior of the internal combustion engine is adversely affected.

A further improvement of the transient behavior (in particular of the positive load change from low load) can be carried out by flushing the main flow path of the inlet line through the compressor between the branch and the mouth of the bypass line of the compressor with fresh air during the bypassing of the compressor. In order to also free the charge air cooler and the pipes in the intake line from recirculated exhaust gas (especially with exhaust gas recirculation rates above 10%), it may be helpful to flush the intercooler and the pipes. For this purpose, it may be provided that upstream of the second obturator and downstream of the charge air cooler, a purge line branches off from the inlet branch and preferably opens downstream of the turbine in the exhaust line, wherein preferably in the purge line, a third obturator is arranged.

The internal combustion engine draws in fresh air via the bypass line of the compressor. The intercooler and the pipes of the inlet pipe are purged during this time. The air / exhaust gas mixture is fed to the internal combustion engine upstream of the catalyst fed to the exhaust line.

A further improvement in fuel consumption can be achieved if the masking is arranged in each case on the side of the outlet opening facing the inlet valve and shades the outlet valve plate of the outlet valve against the inlet valve. In contrast to US Pat. No. 5,870,993 A, the masking of the outlet opening is arranged on the side facing the inlet opening. As a result, due to the amount of exhaust gas recirculated through the outlet openings, a tumble flow arises which rotates in the same direction as the later following tumble of the inlet flow.

In order to allow a sufficiently high tumbling effect, it is advantageous if the mask encloses a wrap angle of about 120 ° to 180 ° about the center of the outlet opening. The height of the mask should be about 1 mm to 4 mm. It is preferably provided that the distance of the masking from the valve disk edge is approximately 0.3 mm to 0.7 mm.

It is preferably provided that the masking is formed by a projection in the combustion chamber ceiling. Alternatively, it can also be provided that the masking is formed by a depression in the combustion chamber cover surface, the outlet opening being arranged in the depression. The depression is arranged essentially concentrically with respect to the outlet opening and preferably runs out into the combustion chamber cover surface on the side opposite the masking. The depth of the recess can be uniform. Alternatively, it is also possible that the depression has a different depth relative to the surrounding combustion chamber cover surface, wherein preferably the depth is greatest in the region of the masking.

In an internal combustion engine having at least two inlet valves and at least two outlet valves, it is advantageous if each outlet opening is surrounded by a masking in the region of the side facing the adjacent inlet valve. The masking of the individual outlet valves is preferably the same design. But it is also a different design possible.

The two exhaust valves are preferably opened at the same time during the exhaust stroke. The opening time of the exhaust valves can be adjusted synchronously via phasing.

By simultaneously opening the exhaust valves, a relatively large amount of the exhaust gas can be directed into the exhaust ports at the beginning of the exhaust stroke. As a result, favorable emptying behavior of the combustion chamber can be achieved with minimal throttling losses. Especially at full load thus results in a high power output.

The exhaust valves are also closed synchronously.

A significant improvement in fuel consumption at partial load is achieved when in the partial load range in the region of the top dead center of the charge exchange, preferably immediately after the top dead center of the charge exchange, exhaust gas sucked back through at least one outlet in the combustion chamber and at the same time or directly through at least one inlet air or charge in the combustion chamber is introduced, wherein a rectified Tumbleströmung is initiated by the sucked back exhaust gas and the air or charge in the combustion chamber. The closing time of the exhaust valve and the opening time of the intake valve, in the partial load range is retarded. It is preferably provided that in the full load range for speeds below the maximum speed, preferably below half the maximum speed, the closing time of the exhaust valve is retarded and / or the opening timing of the intake valve is advanced.

The invention achieves the following advantageous effects: 1) Additional increase of the charge movement in the partial load range in order to further increase the residual gas acceptance of the combustion. Actuators for switching off parts of the inlet channel are not required to achieve increased charge movement.

2) Increasing the rinsing quality at full load, especially at low speeds and large valve overlaps. The improved residual gas purge with lower purge air efficiencies improves knocking performance, torque, and the tendency for catalyst exothermic reactions in suction and supercharging engines at lower full load speeds.

To reduce the production costs can be provided that a raw channel is produced by casting technology, the channel contour in at least a first wall area geometric features of a filling channel and in at least one second wall area geometric features of a tumbleerzeugenden inlet channel, and that in at least one material-removing processing step optionally a filling channel or a tumble-generating inlet channel is formed from the raw channel. With a single cylinder head shape thus suction internal combustion engines, as well as supercharged internal combustion engines can be operated.

Thus, it is possible that the second wall area is machined to remove material for forming a filling channel. Alternatively, it can be provided that, in order to form a tumble-generating inlet channel, the first wall region is processed in a material-removing manner.

In this case, the first wall region is formed by the cylinder head plane closest to the bottom region of the raw channel. The second wall region is formed by the ceiling region of the raw channel located furthest away from the cylinder head plane and / or by the side walls of the raw channel.

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

1 shows a cylinder head of an internal combustion engine according to the invention in a cross section through a cylinder.

2 shows a cylinder with schematically indicated injection jets;

FIG. 3 shows the injection jets from the direction III-III in FIG. 2; FIG.

Fig. 4 is an end view of the injection jets;

4a shows an end view of an injector with five injection openings; Fig. 5 is a side view of an injector and the injection jets;

6 shows an inlet channel of the internal combustion engine according to the invention in a section;

7 shows a diagram with geometric parameters of the inlet channel;

8 shows a cylinder of the internal combustion engine according to the invention during a first injection;

9 shows the cylinder during a second injection;

10 shows the arrangement of the gas exchange channels and the injection device in a longitudinal section.

11 shows the arrangement of the gas exchange channels and the injection device in a plan view;

12 shows an injection with wetted inlet valve in a side view;

Fig. 13 this injection in a plan view;

14 shows a longitudinal section through a cylinder chamber with determined Tumbleströmung.

15 shows a cross section through the cylinder chamber with registered flow field;

16 shows the internal combustion engine according to the invention in a first embodiment;

17 shows the internal combustion engine according to the invention in a second embodiment variant;

18 shows the internal combustion engine according to the invention in a third embodiment;

FIG. 19 shows the internal combustion engine according to the invention in a fourth embodiment variant; FIG.

FIG. 20 shows the internal combustion engine according to the invention in a fifth embodiment variant; FIG.

Fig. 21 is a load-speed diagram; FIG. 22 shows a control strategy for the exhaust gas recirculation according to the method according to the invention; FIG.

FIG. 23 is a valve diagram of a cylinder head of the internal combustion engine according to the invention; FIG.

FIG. 24 shows a cylinder head in a section according to the line XXIV-XXIV in FIG. 23 with the intake and exhaust valves open in a first embodiment according to the invention; FIG.

FIG. 25 shows a cylinder head analogous to FIG. 24 in a second embodiment according to the invention; FIG.

FIG. 26 shows a cylinder head in a third variant according to the invention; FIG.

FIG. 27 shows the cylinder head in a section according to the line XXVII-XXVII in FIG. 23; FIG.

FIG. 28 is a part-load valve lift crank angle diagram; FIG.

FIG. 29 is a full-load valve lift crank diagram; FIG.

FIG. 30 shows the flows in the area of top dead center of the charge cycle at full load; FIG.

FIG. 31 shows a cylinder head according to the invention with a raw channel in a longitudinal section; FIG.

FIG. 32 shows a cylinder head with a tumble-generating inlet channel; FIG. and

33 shows a cylinder head with an inlet channel designed as a filling channel.

Intake channels 2 and outlet channels 3 are arranged in the cylinder head 1 of the internal combustion engine. The inlet channels 2 open via two inlet valves 4 into the combustion chamber 5, which is formed by a cylinder 6, a reciprocating piston 7 and a roof-shaped combustion chamber cover surface 8 formed by the cylinder head 1. In the cylinder head 1 is further per cylinder 6 an ignition 9 approximately centrally, and an injection device 10 arranged laterally. The longitudinal axis 10a of the injection device 10 encloses an angle ε between 20 ° and 35 ° with the cylinder head density plane 1a. The injection device 10 has at its Injektorstirnseite 11 at least five injection openings 12 which are sized and / or arranged such that an area ratio of the total area of all injection openings of at least 65% and / or a mass fraction of at least 65% of the injected fuel on the piston 7 facing first side 13 of a reference plane 30 is assigned by the Gesamtstrahlmittelachse 31. The reference plane 30 is normal to a plane defined by the Gesamtstrahlmittelachse 31 and the longitudinal axis 10a of the injector 10 level 32. The dissolution of the individual injection jets 15 is such that a contiguous fuel cloud 16 without visible separation into individual jets from a distance a of about 10 mm 20 mm - seen from the beam root 17 - is formed. The distribution of the fuel density in the jet cloud 16 substantially corresponds to the geometric arrangement of the injection openings 12 of the injection device 10. The injection openings 12 are preferably arranged asymmetrically with respect to the reference plane. In the case of an injector with seven injection openings 12, for example, three on the first side 13, two on the second side 14 and also two injection openings 12 in the reference plane 31 are arranged (FIG. 4). FIG. 4 a shows an injector end 11 of an injector with five injection openings 12, three injection openings being arranged on the first side 13 and two injection openings 12 being arranged on the second side 14 of the reference plane 30. The injection jets 15a through the injection openings 12 on the first side 13 may be bundled.

The imaginary beam center lines 15b ^{1 of} the injection jets 15b, which are injected into the combustion chamber 5 through the second side 14 of the reference plane 30 facing away from the piston 7, cut the generatrix 6a of the cylinder 6 into a range of approximately 30% to 45% of the stroke length L of the piston 7, measured from top dead center of the piston 7. The projection range of the upper fuel jets 15 on the cylinder jacket 6 'is designated by reference numeral 18.

The imaginary beam center lines 15a ^{1 of} the lower injection jets 15a, which flow through injection openings 12 arranged on the first side 13 of the reference plane 30 facing the piston 7, intersect a normal plane 19 assumed at half the stroke of the piston 7 on the cylinder axis 6a in an inlet side Range 20, which corresponds to about 30% to 50% of the cylinder diameter D.

The theoretical jet shape of the upper injection jets 15a sweeps each sectors of the valve head 4a of the intake valves 4 in the order of one third and one sixth of the circumference, but only at valve strokes greater than 0.6 to 0.8 times the maximum valve lift, from the closed valve measured starting. The inlet channels 2 have a fixed geometry, which is designed so that a high level of charge movement in the combustion chamber 5 is made possible.

Preferably, the tumble number is between 1.2 and 2, more preferably 1.4 to 1.9.

The determination of the tumble number is described for example in AT 004,097 Ul.

A tumble characteristic ω _FK is calculated from ω _{mol from the} flow field w _LDA , which is measured, for example, by a differential measurement method by means of laser Doppler anemometry. First, an average axial flow velocity w is calculated from the flow field w _WA , ie the measured axial velocities in the direction of the cylinder axis 6 a. By reducing the flow field w _WA by the mean velocity w, the reduced, quasi-rotating flow field W ₁ follows, assuming a tumble axis 111 which intersects the cylinder axis 6a and is perpendicular to the cylinder axis 6a of the cylinder 6. It thus applies:

From this reduced flow field W ₁ , the angular velocity ω, at each measuring point i, is calculated as follows: w

where T _{1 is} the distance of the measurement point i from the tumble axis 111.

If the surface element f _{t is} assigned to each measuring point i, the angular velocity ω _{FK of} the total tumble motion is given by the equation

From this, the tumble characteristic can be calculated as follows:

Tumble - characteristic = - ^, (4) ^ω mol where ω _{mol is} the engine angular velocity of the internal combustion engine.

If the tumble characteristic values for a plurality of valve lifts of the inlet valve 4 of the inlet channel 2 are determined, the tumble number can be calculated by integration over the crank angle a, the tumble characteristics being weighted with the piston speed c:

Tumble number (5)

FIG. 13 shows a typical measuring result of the measuring method, indicated in a cross section through the cylinder 6. It shows a flow structure generated by inlet channels 2 in a cylinder 6 as normalized flow fields. The intake valves 4 are shown in their position relative to the cylinder 6. Furthermore, the engine longitudinal axis 114 'is located. The tumble axis 111 'is indicated as a vector arrow with an indicated direction of rotation. The measuring points i are arranged in a hexagonal lattice, so that the surface elements f _{f are} all the same. The direction and magnitude of the angular velocities ω _t in each surface element f _t . is indicated by the slope and density of the hatching. Right-inclined hatching means downward flow, left-inclined hatching symbolizes an upward flow. The density of hatching is proportional to the size of the velocity. In the region of the longitudinal axis 114 ', the lowest flow velocity prevails, which is represented by scarfless hexagonal surface elements /. The calculated flow profile shown in FIG. 13 corresponds to the tumble flow T shown in FIG. 12 according to the intake flow E.

To illustrate, only 37 measuring points per measuring level are entered in FIG. In order to obtain reproducible measurement results, however, in practice at least 200 measuring points per measuring level should be used.

To reach the mentioned high tumble numbers between 1.2 and 2 several factors are decisive:

Both inlet channels 2 are formed tumbleerzeugend and have such a channel design that the angle φ of the main flow directions 40 of the flowing through the inlet valves 4 into the combustion chamber 5 partial flows E for defined by the cylinder axes 6a longitudinal plane 6b determined by the shape of the inlet channels 2. The axis 2 'of each inlet channel 2 includes, at least just in front of the inlet valve 4, an acute angle p with the axis 4b of the inlet valve 4. The main flow directions 40 of the two partial flows E close with the longitudinal plane 6b the same acute angle v, wherein the inlet channels 2 - viewed in the direction of the longitudinal plane 6b - may be formed congruent. The angles p and v are preferably greater than 10 °. The inlet channels can thus be essentially mirrored (equal).

The tumble motion T flows on the side opposite the inlet valves 4 side within a cylinder segment of the combustion chamber cover surface 8 in the direction of the piston 7. The cylinder segment extends over an angle δ <180 °, which is preferably between 60 ° and 120 ° (Fig. 13). A further increase in the charge movement can be achieved by a conical design of the inlet channel 2, as shown in Figures 6 and 7. In this case, the inlet channel 2 is formed between a first cross-section 21 in the region of the inlet channel flange and a second cross-section in the region of a tear-off edge 23 with a tapering cross-sectional area A. The cross-sectional area A is plotted along the inlet channel axis x in FIG. It can be seen from the diagram that in the region 22 of the narrowest cross-section the flow cross-section A is only between 65% and 90% of the cross-sectional area A in the region of the inlet flange surface 21. The region 22 results from the intersection of a cutting plane formed normally on the channel centerline at the location of the tear-off edge 23 with the channel centerline. Up to the area 22, the inlet channel 2 is thus narrowed like a nozzle.

The flow separation edge 23 is formed by an intersection of the substantially continuously extending inlet part 2 a of the inlet channel 2 with a conical wall section 2 b expanding in the flow direction in the region of the valve seat ring 24. The conical wall region 2b is formed by a conical or stepped conical control cutter 25. The control cutter 25 can have a cylindrical and conical region. It may also have a rounded skirt area which merges into the conical area. It is also possible that the inspection cutter 25 - viewed in profile - has an area with a small radius, which merges into an area with a larger radius. The area with the larger radius thus forms a cone with curved lateral surfaces.

The spanned by the outermost beam center lines 15 'first beam angle ß is - viewed in a front view according to the arrow III in Fig. 2 - approximately between 50 ° to 60 °. In a side view according to FIG. 5, the second beam angle γ defined by the center lines 15a 'and 15b' of the outermost upper and lower injection jets 15a, 15b is approximately 30 ° to 40 °.

By the described design of the injector 10 prevents individual fuel jets 15 impinge directly on the cylinder 6 and dilute the lubricating oil. Furthermore, an excessive loading of the surface of the piston 7 is prevented, which would lead to the formation of smoke. Due to the special combination of charge movement, injector geometry and optimization of control parameters, such as injection pressure and time of injection, the oil dilution can be reduced to the level of intake manifold-injecting internal combustion engines.

Furthermore, it can be achieved by combining the injector 10 with a suitable piston 7 that during the heating phase of the catalyst with double Einein- spraying can be driven. A suitable piston 7 has, for example, a teardrop-shaped piston recess, as described for example in EP 1 362 996 A1, which is incorporated by reference into the application.

Figures 8 and 9 show the two-time injection. Fig. 8 shows a first injection during an intake stroke, Fig. 9 shows a second injection during the compression stroke 20 ° to 70 ° before the top dead center of the ignition.

By partially wetting the inlet valve disk during an injection, an interaction of inlet valve 4 and fuel jet 15 can be achieved, as indicated in FIGS. 12 and 13. The theoretical jet shape of the injection jets 15 of the second side 14 intersect the valve disk 4 a of at least one open inlet valve 4 from a valve opening between 0.5 to 0.9 times, preferably between 0.6 and 0.8 times the maximum Ventilhubes in a region 43 of the circumferential line of the valve disk 4a, which is one-third to one-sixth of the circumference.

As a result of the measures described, a slightly rich mixture arises in the area of the ignition device 9, while a slightly lean air-fuel ratio is present in the entire combustion chamber. By optimizing operating parameters such as timing, injection timing and injection sharing, hydrocarbon raw emissions can be reduced by up to 50%. This, in turn, makes it possible to use a less expensive catalyst which can be positioned farther away from the internal combustion engine, thereby reducing the need for greasing. This makes it possible to achieve extremely low emission levels without further measures, such as variable charge movement.

The high charge movement allows operation at high partial load with high internal or external exhaust gas recirculation. This reduces the specific fuel consumption. Again, no additional measure such as variable charge movement is necessary. Exhaust masking allows even a variable charge motion to be integrated without additional components.

The internal combustion engine 101 has an intake line 102, an exhaust line 103 and an exhaust gas turbocharger 104 with a turbine 105 in the exhaust line 103 and a compressor 106 in the intake line 102.

In order to be able to recirculate exhaust gas in the entire operating range of the internal combustion engine 101, an external exhaust gas recirculation system 107 is provided with an exhaust gas recirculation line 108 between the exhaust line 103 and the intake line 102, which branches off from the exhaust line 103 upstream of the turbine 105 and downstream of an exhaust aftertreatment device 109 and upstream of the exhaust line 103 Turbine 106 opens into the intake manifold 102. In the exhaust gas recirculation tion 108, an exhaust gas cooler 110 and an exhaust gas recirculation valve 111 is arranged. The exhaust gas recirculation line 108 enters the intake line 102 downstream of an air filter 112.

Upstream of the compressor 106, an intercooler 113 and a throttle 114 is provided in the intake line 102.

The internal combustion engine 101 further includes an internal exhaust system, not further shown, which may be formed by a variable valve actuation device. The variable valve actuation device can, for example, have phase shifters arranged on the intake and / or exhaust camshaft which adjust the timing of the intake and exhaust valves. Alternatively or additionally, a change in the valve lift profile, for example, by a multi-stage, continuous, electro-mechanical or electrohydraulic valve lift can be done.

In FIG. 21, the load L 'is plotted against the rotational speed n. In the lower and middle operating range A internal exhaust gas recirculation is carried out. In the middle and upper operating range B, both internal and external exhaust gas recirculation is carried out in order to allow high exhaust gas recirculation quantities. The reference character C denotes a full load range with external exhaust gas recirculation.

By means of the external exhaust gas recirculation system 107 designed as a low pressure system, external exhaust gas recirculation can also be carried out in the upper load range B and C. In this case, the exhaust gas is removed after the turbine 105 and after an exhaust gas aftertreatment device 109 formed, for example, by a catalytic converter or a particle filter, cooled and fed to the inlet line 102 before the compressor 106. Due to the external exhaust gas recirculation, the exhaust gas recirculation quantity can be increased by up to 15% in addition to the internal exhaust gas recirculation rate. The mixture of exhaust gas and fresh air is additionally guided via the intercooler 113.

Since, however, a large volume is filled with air / exhaust gas mixture, the transient behavior of the internal combustion engine 101 is very poor, in particular with a rapid lowering of the load L to almost zero load. The internal combustion engine 101 tolerates only a small amount of recirculated exhaust gas at low loads L, but due to the large volume filled with air / exhaust gas mixture in the intake line 102 is still supplied with a relatively large amount of exhaust gas for a few work cycles.

A control strategy for internal combustion engines 101 with turbocharger 104 is shown for example in FIG. 22, where load L ¹ , setpoint for external exhaust gas feedback AGR _ext , s, the actual value for external exhaust gas recirculation amount AGR _ext , i and the internal exhaust gas recirculation amount AGR _int are plotted against the time t.

With a negative load jump from high-load to very low load L, the external exhaust gas recirculation rate AGR _ext , i, which is too high for a few cycles, is compensated for by adjusting the control times in the direction of lower internal exhaust gas recirculation rate AGR _int . This can be done, for example, for a few cycles of large valve overlap on small valve overlap. When the volume filled with the exhaust gas / air mixture is used up, internal exhaust gas recirculation is again carried out. By changing the internal exhaust gas recirculation quantities AGRj _nt , a compensation of the filling and emptying behavior of the external exhaust gas recirculation system 107 can thus be achieved.

The regulation of the internal and external exhaust gas recirculation via a not further shown engine control unit such that the total exhaust gas recirculation rate in the internal combustion engine 101, which is composed of internal exhaust gas recirculation rate, empty suction filled with exhaust gas / air mixture volume and external exhaust gas recirculation rate, regardless of the operating state corresponds to a value stored in a map or calculated by a model, or corresponds to the exhaust gas recirculation quantities corresponding to the best consumption or the best emissions.

The engine control unit stores the optimum exhaust gas recirculation rates for each operating point. With suitable sensors, the exhaust gas recirculation quantity is constantly monitored. In the case of a sudden negative load step, the actual value of the external recirculated exhaust gas quantity AGR _ext , i is too large. In order to compensate for this excess in the dynamic case, the internal exhaust gas amount AGR _{int is} reduced until the actual value of the external exhaust gas quantity AGR _ext , i corresponds to the setpoint AGR _ex t, s, as shown in FIG.

Internal combustion engines with exhaust gas turbocharger 104 often have variable valve actuation devices in order to be able to control the residual gas content in the cylinder.

The compensation of the filling and emptying behavior of the intercooler and the pipes of the intake manifold 102 by adjusting the timing to lower internal exhaust gas recirculation rate can be achieved by the following measures:

- turning the inlet-side phase shifter in the direction of late;

- Turning the exhaust side phase shifter in the direction of early; - Twisting of both phase shifters so that smaller overlaps arise.

Furthermore, the compensation of the filling and emptying behavior can be achieved by such changes of the valve lift profile that smaller internal exhaust gas recirculation rates occur. This can be done by:

- two or more stages valve lift switching;

- Continuous valve lift adjustment;

- electromechanical valve lift control;

- electrohydraulic valve lift adjustment.

An increase in the external or internal exhaust gas recirculation rate can be achieved by additionally throttling the intake air or the exhaust gas through throttle elements 114, 115.

The control of the external exhaust gas recirculation rate is performed by the exhaust gas recirculation valve 111 and / or by throttling on the suction or exhaust side.

Further compensation for the filling and emptying behavior of the external cooled exhaust gas recirculation system 107 may be accomplished by means of a bypass line 116 for the compressor 106, as shown in FIG. 17. At the partial load (after a negative load step) or below that threshold at which external exhaust gas recirculation is operated, the bypass line 116 is opened via a first shut-off device 117. As a result, the large volume of the compressor 106 and of the charge air cooler 113 filled with air / exhaust gas mixture is bypassed.

The durchzusaugende after a negative load jump volume with air / exhaust gas mixture is thereby drastically reduced. Air / exhaust gas mixture is already available for the next high-load phase.

If appropriate, a second shut-off device 118 may be arranged in the inlet line 102 downstream of the charge air cooler 113. The shut-off devices 117, 118 may be pressure or vacuum controlled or electrically operated.

Another possibility for improving the transient behavior (in particular with positive load changes from low load) is shown in FIG. In particular, at exhaust gas recirculation rates well above 10%, it may be helpful to flush the charge air cooler 113 and the main flow path 102a between the compressor 106 and the charge air cooler 113 with fresh air. For this purpose, a purge line 119 is provided between inlet branch 102 and exhaust line 103, which downstream of the charge air cooler 113 and upstream of the obturator 118 branches off from the intake manifold 102 and downstream of the turbine 105 and upstream of the exhaust aftertreatment device 109 into the exhaust line 103 ^■ opens. In the purge line 119, a further obturator 120 is provided.

The engine 101 draws fresh air via the bypass line 116 of the compressor 106. The charge air cooler 113 and the main flow path 102a are purged during this time. The air / exhaust gas mixture is guided past the internal combustion engine 101 in front of the exhaust gas aftertreatment device 109.

FIGS. 19 and 20 show embodiments, in particular for diesel internal combustion engines, in which only the intercooler 113 with the bypass line 116 is bypassed. Denoted by reference numerals 116a and 116b are switching members at the branch and the mouth of the bypass line 116, which divide the charge air flow accordingly. The compressor 106 supports the flushing of the exhaust gas / air mixture from the intercooler 113th

20, a high-pressure exhaust gas recirculation system 107a is provided with a high-pressure exhaust gas recirculation line 108a and an exhaust gas recirculation valve lil to return exhaust gas from the exhaust header 103a directly to the inlet header 102b during warm-up for a short time.

FIG. 23 shows a cylinder head 201 in a view on the combustion chamber side of the cylinder head bottom 202 of a cylinder 203. For each cylinder 203, two inlet openings 204, 205 and two outlet openings 206, 207 open into the combustion chamber 210. Via the inlet openings 204, 205, the combustion chamber is connected to inlet channels 204b, 205b and via the outlet openings 206, 207 to outlet channels 206b, 207b. The inlet openings 204, 205 and the outlet openings 206, 207 are respectively controlled by an inlet valve 204a, 205a and outlet valve 206a, 207a. The combustion chamber 210 is bounded by the roof-shaped combustion chamber cover surface 211 and the piston 212, as well as the cylinder 203.

In particular, at partial load operation of the internal combustion engine can be carried out to improve the consumption of an internal exhaust gas recirculation. This internal exhaust gas recirculation is realized by exhaust gas from the outlet channels 206b, 207b is sucked back into the combustion chamber 210 subsequent to the exhaust stroke in the region of top dead center OTW of the charge cycle. To improve the firing conditions and emissions, a high tumble flow in the combustion chamber 210 is desirable. This tumble flow S _A is replaced by the same closing timings of the two exhaust valves 206a, 207a zieit. A substantial increase in the tumble flow can be achieved by a mask 208 around the two outlet openings 206, 207, on the side facing the inlet openings 204, 205. The masks 208 each have a wrapping angle α of about 120 ° to 180 ° about the center 206 ', 207' of the respective outlet opening 206, 207. The height of the mask 208 is about 1 mm to 4 mm - measured to half the height h of the valve plate edge 206b. The distance a between the mask 208 and the valve disc rim 206b is about 0.3 mm to 0.7 mm.

It is essential that the Tumbleströmung S _A according to the recirculated exhaust gas has the same orientation as the Tumbleströmung S _E in the combustion chamber 210 through the inlet openings 204, 205 incoming air or the incoming air-fuel mixture. The masking 208 may be formed by a projection V (FIG. 24) or a recess T of the combustion chamber cover surface 211 (FIGS. 25, 26). The depression T is in each case formed concentrically with the inlet opening 204, 205.

FIG. 25 shows a depression T with a constant depth HM, FIG. 26 a depression T, which runs out from the masking 208 into the combustion chamber cover surface 211 and whose deepest point in the region of the mask 208 'is pronounced.

FIG. 28 shows a valve lift H-crank angle KW diagram, wherein the valve lift curve of the intake valves 204a, 205a is designated by E. A denotes the valve lift curves of the exhaust valves 206, 207.

Due to the fact that the masking 208 is provided on the side facing the inlet valves 204, 205, a tumble E _{A is produced} by the backflowing exhaust gases, which are in the same direction as the later following inlet tumble (at the same control time of the outlet valves 206, 207) symmetrically designed inlet channels 204b, 205b) rotates.

In order to achieve higher amounts of recirculated exhaust gas, the closing times A _{s of} the two exhaust valves 206, 207 and the opening times Eö are postponed so late in the partial load range that is sucked through the outlet openings 206a, 207a first.

In the described embodiments, at partial load, the gas inflow phases obtained from the control times which are advantageous for charge dilution and de-throttling are consistently used to increase the charge movement, in particular also the fresh charge upstream of the exhaust gas from the outlet duct 206b, 207b. This is accomplished by the partial masking 208 of the exhaust valves 206a, 207a on portions of the enclosure facing the inlet valves 204a, 205a, whereby the exhaust ports 206b, 207b incoming exhaust gas masses a unique flow direction S _A; which follows the basic flow direction S _{E of} the charge movement, granted. Since the backflow of exhaust gas occurs only in the region of low valve strokes H, as can be seen from the valve timing (FIG. 28), a limitation of the height HM of the mask 208 can be made so as not to restrict the flow capacity of the exhaust passage 206b, 207b under full load operation.

Since the proportion of the exhaust gas to the total charge can readily be more than half, resulting from the use of this partial flow S _A a significant additional impulse for the charge movement S _E. This difference has the effect that switching elements for influencing the charge movement of the inlet channel 204b, 205b can be dispensed with.

Furthermore, the charge cycle is also favorably influenced even at full load, especially for suction and supercharged engines with adjustable exhaust and / or intake camshafts. By adjustable large valve overlaps as far as possible rinsing of the combustion chamber 210 is achieved by residual gas. The amount of air required for rinsing should not be too great in order to avoid an excessive oxygen supply or cooling of the exhaust gas in the turbo engine. In conventional internal combustion engines due to the typical design of the combustion chamber 210 when flushing to a relatively large short circuit content of purge air or, in engines with intake manifold injection, to mixture which passes directly into the Auslasstrakt, without affecting the purge quality. On the other hand, in the internal combustion engine according to the present invention, the targeted masking 208 of the exhaust valves 206a, 207a reduces the short-circuiting ratio in purging, and the scavenging efficiency is improved, see FIGS. 29 and 30.

The design of the masking 208 can take place with an unchanged arrangement of the valves by application to an existing combustion chamber contour (FIG. 24) or by resetting the outlet valves 206a, 207a and a fluid-free release of the valve disk on its side facing away from the inlet valves 204a, 205a by means of locally modified combustion chamber contour (FIG. FIG. 25), as well as, for example, by pivoting a changed arrangement of the outlet valves 206a, 207a, whereby a corresponding residue is achieved on the side facing the inlet valves (FIG. 26).

A cylinder head 301 has at least one inlet channel 302, which opens into a combustion chamber roof 304 via an inlet opening 303. In order to use a single cylinder head shape both for suction and for supercharged internal combustion engines, a raw channel 305 is minimally cast with the cylinder head 301 with a minimum cross section. The cylinder head density plane 306 next located, formed by the bottom portion 305a of the raw channel 305 first wall portion A 'is designed so that it corresponds to the geometric requirements of a filling channel 307. On the other hand, the second wall portion B 'formed by the ceiling portion 305b remote from the cylinder head sealing plane 306 and the side walls of the raw passage 305 is designed to meet the requirements of a high tumble channel 308.

Depending on whether the cylinder head 301 is intended for a suction internal combustion engine or for a supercharged internal combustion engine, the first wall portion A 'or the second wall portion B' of the Rohkanals 305, for example by CNC machining, removed, according to the dashed, or 32 shows the form of a tumble-generating inlet channel 308 obtained after material removal of the first wall region 305a. FIG. 33, on the other hand, shows a cylinder head 301 with a filling channel 307 whose contour is formed by material-removing machining of the second wall region B 'of the raw channel 305 was obtained.

Reference numerals 308b and 307b denote control section for a tumble-generating channel 308 and a filling channel 307, respectively.

Claims

PATENT CLAIMS

1. internal combustion engine, in particular turbocharged internal combustion engine, with at least one cylinder (6) with a reciprocating piston (7) and at least two inlet valves (4) per cylinder (6) and a roof-shaped combustion chamber cover surface (8) in the cylinder head (1), wherein the inlet ducts (2) leading to the intake valves (4) generate a tumble flow (T) in the combustion chamber (5), the main flow directions (40) of the partial flows (E) sucked into the combustion chamber (5) via the intake valves (4) a respective longitudinal plane (6b) defined by the cylinder axes (6a) of a row of cylinders, each having an acute angle (φ), and at least one outlet channel (3) per cylinder (1), characterized in that the tumble number is between 1.2 and 2, preferably between 1.4 and 1.9.

2. Internal combustion engine according to claim 1, characterized in that the angles (φ) of the main flow directions (40) of the partial flows (E) to the longitudinal plane (6b) by the shape of the inlet channels (2) are determined, in each case the axis (2 ') of the inlet channel (2), at least just before the inlet valve (4), with the axis (4b) of the inlet valve (4) includes an acute angle (p), wherein preferably the angle (p) is greater than 10 °.

3. Internal combustion engine according to claim 1 or 2, characterized in that the main flow directions (40) include the two partial flows (E) with the longitudinal plane (6b) the same acute angle (v), wherein preferably the inlet channels (2) - in the direction of the longitudinal plane (6b) - are formed congruent, and particularly preferably the angle (v) is greater than 10 °.

4. Internal combustion engine according to one of claims 1 to 3, characterized in that each inlet channel (2) immediately upstream of the valve seat ring (41) has a molded-flow rupture edge (23).

5. Internal combustion engine according to one of claims 1 to 4, characterized in that the flow separation edge (23) by an intersection of the substantially continuously extending inlet part (2a) of the inlet channel (2) with a widening in the flow direction, preferably conical wall portion (2b) is formed in the region of the valve seat ring (41).

6. Internal combustion engine according to one of claims 1 to 5, characterized in that the conical wall section (2 b) by a coaxial with Valve seat ring axis (41a) zuzustunden control cutter is incorporated, wherein preferably the control cutter has at least one cylindrical, conical and / or curved cladding region.

7. Internal combustion engine according to one of claims 1 to 6, characterized in that the cross section of the inlet channel (2) in the region of the flow separation edge (23), at least in one direction, is narrowed like a nozzle.

8. Internal combustion engine according to one of claims 1 to 7, characterized in that the Tumbleströmung (T) flows in an exhaust-side cylinder segment of the combustion chamber cover surface (8) in the direction of the piston (7), which over an angle (δ) is smaller than 180 ° , preferably between 60 ° and 120 °, extends.

9. Internal combustion engine according to one of claims 1 to 8, at least one of the cylinder axis (6 a) spaced and inclined to this arranged injector (10) with at least five injection openings (12), characterized in that the theoretical beam shape of the injection jets (15) of second side (14) intersects the valve plate (4a) of at least one opened inlet valve (4).

10. Internal combustion engine according to claim 9, characterized in that the theoretical jet form the valve disk (4a) of the inlet valve (4) from a valve opening between 0.5 to 0.9 times, preferably between the 0.6 to 0, 8 times the maximum valve lift cuts.

11. Internal combustion engine according to claim 9 or 10, characterized in that the theoretical beam shape, the inlet valve (4) in a region (43) of the circumferential line of the valve disc (4a) intersects, which is one third to one sixth of the circumference.

12. Internal combustion engine according to one of claims 1 to 11, characterized in that the imaginary outer beam centerlines (15a ¹ , 15b ¹ ) of the fuel injection jets (15) - in a the injector (10a) containing longitudinal section through the cylinder (6) considered - a first beam angle (ß) between 30 ° to 50 °, preferably between 35 ° and 40 ° span.

13. Internal combustion engine according to one of claims 1 to 12, characterized in that the outer imaginary beam center lines (15 ') of the injection jets - viewed in a front view of the injector (10) - a second injection angle (γ) between about 40 ° and 70 ° , preferably span between 50 ° and 60 °.

14. Internal combustion engine according to one of claims 1 to 13, characterized in that the internal combustion engine is preferably charged via at least one exhaust gas turbocharger.

15. A method for operating an internal combustion engine with Abgasturboiader, an intake and an exhaust line, an internal exhaust gas recirculation system and an external exhaust gas recirculation system with at least one exhaust gas recirculation line, wherein in at least one low and / or medium load range of the internal combustion engine, an internal exhaust gas recirculation and in at least one load range at the same time an internal and external exhaust gas recirculation is carried out, characterized in that for each operating range a setpoint for the recirculated exhaust gas is stored or calculated in a control device for the exhaust gas recirculation, that an actual value for the recirculated exhaust gas amount is determined and that when reducing the Load from an operating range in which internal and external exhaust gas recirculation is carried out simultaneously, the internal exhaust gas recirculation is reduced or stopped until the actual value of the recirculated exhaust gas quantity is lower or lower is equal to the set point of the recirculated exhaust gas quantity at the corresponding operating point.

16. The method according to claim 15, characterized in that in the upper load range and / or in the full load range, an internal and external exhaust gas recirculation is performed simultaneously.

17. The method according to claim 15 or 16, characterized in that in the upper load range and / or in the full load range, an external exhaust gas recirculation is performed alone.

18. The method according to any one of claims 15 to 17, characterized in that the internally and / or externally recirculated exhaust gas is cooled between the removal from the combustion chamber and the return to the combustion chamber.

19. The method according to any one of claims 15 to 18, characterized in that the exhaust gas for the external exhaust gas recirculation downstream of the turbine of the exhaust gas turbocharger, preferably downstream of an exhaust aftertreatment device taken from the exhaust line and upstream of the compressor of the exhaust gas turbocharger is supplied to the intake manifold.

20. The method according to any one of claims 15 to 19, characterized in that to reduce the internal exhaust gas recirculation, the control time of at least one inlet valve is retarded.

21. The method according to any one of claims 15 to 20, characterized in that to reduce the internal exhaust gas recirculation, the control time of at least one exhaust valve is adjusted to early.

22. The method according to any one of claims 15 to 21, characterized in that is reduced to reduce the internal exhaust gas recirculation amount of the valve lift at least one inlet and / or outlet valve in the valve overlap region.

23. The method according to any one of claims 15 to 22, characterized in that to increase the external and / or internal exhaust gas recirculation amount, the intake air in the intake manifold and / or the exhaust gas is throttled in the exhaust line.

24. The method according to any one of claims 15 to 23, characterized in that in at least one partial load operating range, preferably after a negative load step and / or below a lower load limit for the implementation of external exhaust gas recirculation of the compressor and / or the intercooler is bypassed.

25. The method according to claim 24, characterized in that the main flow path of the inlet strand is closed by the compressor between the branch and the mouth of a bypass line bypassing the compressor by at least one obturator.

26. The method of claim 24 or 25, characterized in that during the bypass of the compressor and / or the intercooler of the main flow path of the inlet pipe between the junction and the mouth of the bypass line of the compressor is purged with fresh air.

27. Internal combustion engine (101) with exhaust gas turbocharger (104), an input and an exhaust line (102, 103), an internal exhaust gas recirculation system and an external exhaust gas recirculation system (107) with at least one exhaust gas recirculation line (108) and at least one cooler (110, 113) for the externally recirculated exhaust gas, characterized in that the exhaust gas recirculation line (108) downstream of the turbine (105) of the exhaust gas turbocharger (104), preferably downstream of an exhaust aftertreatment device (109) branches off from the exhaust line (103) and upstream of the compressor (105 ) of the exhaust gas turbocharger (104) into the inlet branch (102) opens.

28. internal combustion engine (101) according to claim 27, characterized in that the internal exhaust gas recirculation system comprises an adjusting device for the timing of the intake and exhaust valves.

29. Internal combustion engine (101) according to claim 28, characterized in that the adjusting device has at least one phase shifter for rotating the intake or exhaust camshaft.

30. Internal combustion engine (101) according to claim 28, characterized in that the adjusting device has a Ventilhubumschaltvorrichtung.

31. Internal combustion engine (101) according to claim 28, characterized in that the adjusting device has electromechanical or electro-hydraulic actuators for the intake or exhaust valves.

32. Internal combustion engine (101) according to any one of claims 27 to 31, characterized in that the compressor (106) of the exhaust gas turbocharger (104) and / or the intercooler (113) via a bypass line (116) is bypassable, preferably in the bypass line (116) a first obturator (117) is arranged.

33. Internal combustion engine (101) according to claim 32, characterized in that downstream of the compressor (106), preferably downstream of the charge air cooler (110), particularly preferably downstream of a in a main flow path (102 a) through the compressor (106) arranged second obturator (118 ) again in the inlet strand (102) opens.

34. Internal combustion engine (101) according to claim 32 or 33, characterized in that the bypass line (116) upstream of the mouth of the exhaust gas recirculation line (108) branches off from the inlet branch (102).

35. Internal combustion engine (101) according to claim 32 or 33, characterized in that the bypass line (116) branches off from the inlet branch (102) downstream of the compressor (106).

36. Internal combustion engine (101) according to any one of claims 33 to 35, characterized in that upstream of the second obturator (118), and preferably downstream of a charge air cooler (113), a purge line (119) from the main flow path (102a) of the inlet strand (102). branches off and preferably downstream of the turbine (105) into the exhaust line (103) opens, wherein preferably in the purge line (119), a third obturator (120) is arranged.

37. An internal combustion engine having a cylinder head (201) with at least one inlet valve (204a, 205a) and at least one outlet valve (206a, 207a) per cylinder (203), at least one outlet opening (206, 207) being at least at least partially surrounded by a masking (208), characterized in that the masking (208) respectively on the inlet valve (204a, 205a) facing side of the outlet opening (206, 207) is arranged and the outlet valve plate of the outlet valve (206a, 207a ) shades off against the inlet valve (204a, 205a).

38. Internal combustion engine according to claim 37, characterized in that the mask (208) around the center of the outlet opening (206, 207) includes a wrap angle (α) of about 120 ° to 180 °.

39. Internal combustion engine according to claim 37 or 38, characterized in that the height (HM) of the mask (208) - measured up to half the height of the valve head edge of the corresponding exhaust valve (206a, 207a) - is about 1 mm to 4 mm.

40. Internal combustion engine according to one of claims 37 to 39, characterized in that the distance (a) of the masking (208) from the valve disk edge of the nearest outlet valve (206a, 207a) is about 0.3 mm to 0.7 mm.

41. Internal combustion engine according to one of claims 37 to 40, characterized in that the masking (208) by a projection (V) in the combustion chamber cover surface (211) is formed.

42. Internal combustion engine according to one of claims 37 to 41, characterized in that the masking (208) by a recess (T) in the combustion chamber cover surface (211) is formed, wherein in the recess (T), the outlet opening (206, 207) is.

43. Internal combustion engine according to claim 42, characterized in that the depression (T) is arranged substantially concentrically with respect to the outlet opening (206, 207), the depression (T) preferably being located on the side opposite the masking (208) into the combustion chamber cover surface ( 211) runs out.

44. Internal combustion engine according to claim 42 or 43, characterized in that the recess (T) - based on the surrounding combustion chamber cover surface (211) has a uniform depth (HM).

45. Internal combustion engine according to claim 42 or 43, characterized in that the recess (T) - with respect to the surrounding combustion chamber cover surface (211) - a different depth (HM), wherein preferably the depth (HM) in the mask (208) is greatest.

46. Internal combustion engine according to one of claims 37 to 45, having at least two inlet and at least two outlet valves (204a, 205a, 206a, 207a) per cylinder (203), characterized in that each outlet opening (206, 207) in the region of the adjacent Inlet valve (204a, 205a) facing side of a mask (208) is surrounded.

47. Internal combustion engine according to claim 46, characterized in that the masks (208) of the individual exhaust valves (206a, 207a) are designed the same.

48. Internal combustion engine according to claim 46, characterized in that the masks (208) of the individual exhaust valves (206a, 207a) are designed differently.

49. Internal combustion engine according to one of claims 38 to 42, characterized in that the exhaust valves are synchronously actuated.

50. A method for describing an internal combustion engine according to one of claims 37 to 50, characterized in that in the partial load range in the region of the top dead center of the charge exchange (OTW), preferably immediately after the top dead center of the charge exchange (OTW), exhaust gas through at least one outlet port ( 206, 207) is sucked back into the combustion chamber (210) and at the same time or directly through at least one inlet opening (204, 205) air or charge is introduced into the combustion chamber (210), wherein the sucked back exhaust gas and the air or charge in the combustion chamber ( 210) a rectified Tumbleströmung (S _E , S _A ) is initiated.

51. The method according to claim 50, characterized in that in the partial load range of the closing time (A ₅ ) of the exhaust valve (206a, 207a) and the opening time (E _Ö ) of the inlet valve (204a, 205a) is retarded.

52. The method of claim 50 or 51, characterized in that in the full load range for speeds below the maximum speed, preferably below half the maximum speed, the closing time (A ₅ ) of the exhaust valve (206a, 207a) late and / or the opening time (E _Ö ) of the inlet valve (204a, 205a) is advanced.

53. The method according to any one of claims 50 to 52, comprising at least two exhaust valves (206a, 207a) per cylinder (203), characterized in that the exhaust valves (206a, 207a) are actuated synchronously.

54. A method for producing a cylinder head (301) having at least one inlet channel (302) opening into a combustion chamber roof (304) via an inlet opening (30), which is produced at least partially by casting, characterized in that a raw channel (305) is produced by casting whose channel contour has geometric features of a filling channel (307) in at least one first wall region (A ') and geometric features of a tumble-producing inlet channel (308) in at least one second wall region (B'), and optionally in at least one material-removing machining step Fill passage (307) or a tumbleerzeugender inlet channel (308) from the raw channel (305) is formed.

55. The method according to claim 54, characterized in that, in order to form a filling channel (307), the second wall region (B ') is processed in a material-removing manner.

56. The method according to claim 54, characterized in that, in order to form a tumble-producing inlet channel (308), the first wall region (A ') is processed in a material-removing manner.

57. Method according to claim 54, characterized in that the first wall region (A ') is assigned a bottom region (305a) of the raw channel (305) which is preferably adjacent to the inlet opening (303) and closest to a cylinder head sealing plane (306).

58. The method according to any one of claims 54 to 57, characterized in that the second wall region (B ¹ ) is assigned to a side region and / or ceiling region (305b) of the raw channel (305), preferably adjoining the inlet opening (303), wherein the Ceiling portion (305b) from the cylinder head density plane (306) is further spaced than the bottom portion (305a).

59. Cylinder head (301) for an internal combustion engine having at least one inlet channel (302) which opens at least partially by casting, via an inlet opening (303) into a combustion chamber roof (304), characterized in that the inlet channel (302) consists of a raw channel ( 305) whose channel contour has at least one first wall region (A ') with geometric features of a filling channel (307) and at least one second wall region (B') with geometric features of a tumble-generating inlet channel (308), and in that at least one wall region (A ', B ¹ ) for forming a filling channel (307) or a tumble-generating inlet channel (308) is formed by a material querying processing from the raw channel (305).

60. Cylinder head (301) according to claim 59, characterized in that for forming a filling channel (307) the second wall region (B ¹ ) is machined to remove material, and the first wall region (A ') of the raw channel (305) forms part of the filling channel (307) trains.

61. Cylinder head (301) according to claim 59, characterized in that for forming a tumbleerzeugenden inlet channel (308) of the first wall portion (A ¹ ) is machined, and the second wall portion (B ') of the Rohkanals (305) part of the tumbleerzeugenden Inlet channel (308) forms.

62. Cylinder head (301) according to one of claims 59 to 61, characterized in that the first wall region (A ') adjoins a bottom region (305a) of the raw passage (305a), which is preferably adjacent to the inlet opening (303) and faces a cylinder head sealing plane (306) ) is formed.

63. Cylinder head (301) according to one of claims 59 to 62, characterized in that the second wall region (B ') is formed by a preferably adjacent to the inlet opening (303) side region and / or ceiling region (305b) of the raw channel (305) wherein the ceiling portion (305b) is spaced further from the cylinder head sealing plane (306) than the bottom portion (305a).