CN100404814C - Method for operating a direct injection diesel internal combustion engine - Google Patents

Abstract

The invention relates to a method for operating a direct injection diesel internal combustion engine with at least one reciprocating piston (27) in a cylinder (24), wherein the internal combustion engine is operated in such a way that the fuel is kept at a local temperature below the NOx-forming temperature Combustion at temperatures and with local air ratios above the soot formation limit, where fuel injection starts before top dead center of the compression phase in the crank angle range between 50° and 5° and exhaust gas recovery, where exhaust gas recovery The rate is about 50%-70%. In order to achieve particularly low nitrogen oxide and soot emissions, at least one piston (27) is provided with an extrusion face (34) and an annular piston basin-shaped combustion chamber (28) and a ) and the constriction (29) in the transition zone between the piston bowl-shaped combustion chamber (28); when the piston (27) moves upwards, a kind of extrusion is produced against the piston bowl-shaped combustion chamber (28) from the outside to the inside flow (43); the fuel is at least mainly injected into the annular piston bowl-shaped combustion chamber (28), and is passed through the squeeze flow (43) along the piston bowl-shaped combustion chamber side wall (31) and/or the piston bottom (32) at least Transported under partial vaporization.

Description

Method for operating a direct injection diesel internal combustion engine

technical field

The invention relates to a method for operating a direct injection diesel internal combustion engine with at least one piston reciprocating in a cylinder, wherein the internal combustion engine is operated such that the fuel is at a local temperature below the NOx-forming temperature and at a temperature above Soot formation Combustion at extreme local air ratios where fuel injection starts before top dead center in the compression phase and exhaust gas recovery in the range between 50° and 5° of crank angle where exhaust gas recovery is about 50%- 70%. Furthermore, the invention relates to an internal combustion engine for carrying out the method.

Background technique

The most important parameters of the combustion process in an internal combustion engine with internal combustion are the phase of the combustion process or the start of combustion, the maximum rate of rise of the cylinder pressure and the peak pressure.

In internal combustion engines in which the combustion takes place essentially by self-ignition of the direct injection of the fuel quantity, these parameters are determined crucially by the timing of the injection, the charge composition and the ignition delay. These parameters are themselves determined by a large number of influencing factors such as rotational speed, fuel quantity, intake air temperature, boost pressure, effective compression ratio, exhaust gas content of the cylinder charge and component temperatures.

Strict legal regulations make it necessary to constantly find new ways to reduce the emissions of soot particles and NOx-emissions from diesel internal combustion engines when designing combustion methods.

A known way to reduce NOx and soot emissions in the exhaust gas is to increase the ignition delay by advancing the timing of the fuel injection, so that combustion takes place by self-ignition of a lean fuel-air mixture. A possible solution here is called the HCLI-method (Homogenous Charge Late Injection). If the fuel injection takes place far enough before the top dead center of the compression phase, this air-fuel mixture takes place, whereby a somewhat premixed fuel-air mixture is produced. By means of exhaust gas recovery it is possible to keep the combustion temperature below the minimum temperature required for NOx formation.

US patent US 6.338.245 B1 describes a diesel internal combustion engine operating according to the HCLI-method, in which the combustion temperature and ignition delay are adjusted in such a way that the combustion temperature is lower than the NOx-forming temperature and the local air in the low and middle part load range The ratio is higher than the value critical for carbon black formation. The combustion temperature is controlled in this respect by varying the exhaust gas recovery ratio, and the ignition delay is controlled by the timing of the fuel injection. At medium and high loads, the combustion temperature is kept as low as possible to reduce both NOx- and soot formation.

Furthermore, the pistons of known diesel internal combustion engines are formed with essentially annular piston bowl-shaped combustion chambers. In this respect, a constriction is provided in the transition region between the piston end face and the piston bowl-shaped combustion chamber, forming a relatively narrow passage cross section. Due to the narrow delivery cross-section, a high mixture formation energy is provided, whereby the fuel handling is significantly improved. Pistons with such an annular piston basin-shaped combustion chamber are basically covered by published patents EP 0 383 001 A1, DE 1 122 325 AS, AT 380 311 B, DE 21 36 594 A1, DE 974 449 C or JP 60-206960 A is known. In a conventionally operating internal combustion engine, the use of such a piston has the following advantageous effects on the operating characteristics of the internal combustion engine: the full load of the smoke restriction can be increased; a high compression can be achieved, resulting in lower combustion noise due to a smaller ignition delay , lower hydrocarbon emissions, more favorable starting behavior of the engine and increased efficiency of the internal combustion engine; in addition, the possibility arises to delay the ignition time in the direction and to keep the energy of the mixture formation high over a longer time interval level this fact, there is no noticeable rise in smoke, consumption and HC. This possibility means mainly reduced NOx, combustion noise and peak cylinder pressures.

Furthermore, laid-open patent DE 11 22 325 C1 has known a piston with a piston bowl-shaped combustion chamber and a constriction, wherein a molding is provided between the extrusion surface and the constriction.

In internal combustion engines operating according to the HCLI method, such a piston configuration with a deeply constricted piston bowl combustion chamber has not been used until now, because it was thought until now that due to the deep piston bowl combustion chamber and the strong squeeze flow starting Capacity and thermodynamic efficiency can become too poor. The US patent US 6.158.413 A therefore proposes to firstly suppress the squeeze flow, wherein a piston with a very flat piston-cone-shaped combustion chamber is used.

Contents of the invention

The object of the present invention is to improve the HCLI method for operating an internal combustion engine in such a way that on the one hand the nitrogen oxide and soot emissions can be further reduced and on the other hand a widening of the load range that can be driven in HCLI operation can be achieved.

This object is achieved according to the invention in that at least one piston is provided with at least one pressing surface and an annular piston bowl-shaped combustion chamber and a constriction in the transition region between the pressing surface and the piston bowl-shaped combustion chamber. position; when the piston moves upwards, a squeeze flow is generated from the outside to the inside of the piston bowl-shaped combustion chamber; the fuel is at least mainly injected into the annular piston bowl-shaped combustion chamber, and through the squeeze flow along the side of the piston bowl-shaped combustion chamber The wall and/or the piston bottom are transported under at least partial vaporization. In this case, the fuel jet is injected into the squeeze flow flowing into the piston bowl-shaped combustion chamber. Squeeze flow delivers most of the fuel into the piston bowl, where it vaporizes and mixes nearly uniformly with incoming air. The flow in the piston-cone combustion chamber depends on whether there is an intake air flow with or without swirl.

Therefore, in an embodiment according to the invention, a kind of swirl flow with a swirl number ≥ 1 is generated in the cylinder, and the fuel is squeezed along the side wall of the piston bowl-shaped combustion chamber, under at least partial vaporization To the direction of the bottom of the piston and continue along the bottom of the piston to the center of the basin-shaped combustion chamber. The swirl is maintained inside the piston bowl-shaped combustion chamber during the compression phase.

In another embodiment, on the contrary, a swirl-free intake flow with a swirl number <1 is generated in the cylinder, and the fuel flows from the center of the basin-shaped combustion chamber, along the bottom of the piston to the piston through the extrusion flow, at least partially vaporized. The side wall of the basin-shaped combustion chamber and continue to deliver to the contraction part.

Surprisingly, it has been shown that the startability of an internal combustion engine operating according to the HCLI method is not significantly impaired by the installation of a piston-cone-shaped combustion chamber. The loss of thermodynamic efficiency due to squeeze flow can be significantly offset by the effect of high swirl through the improved mixture handling in the piston bowl-shaped combustion chamber.

In this respect, the fuel is preferably injected in the direction of the constriction of the piston, wherein at the start of the injection the majority of the fuel quantity is at least one point of intersection of the jet axis of the fuel jet between the side wall of the bowl-shaped combustion chamber and the squeeze surface In an area between, the area includes the hanging wall area, the constriction and the air intake area between the extrusion surface and the constriction.

In conventional diesel internal combustion engines, the engagement point and the injection timing of the fuel are usually chosen such that the fuel hits the suspension wall region below the constriction at the start of the injection—independent of the load. In the subject matter of the invention, the cut-in point is set at low loads by a region of the hanging wall area inside the piston bowl-shaped combustion chamber and shifts in the direction of the constriction as the load increases. This can be achieved by advancing the injection timing. Part of the fuel is thus injected—in contrast to the squeeze flow—into the gap between the piston and cylinder head. Most of the fuel injected into the gap between the piston surface and the cylinder head is carried by the squeeze flow into the piston bowl-shaped combustion chamber. This advantageously improves air distribution and mixture treatment while reducing HC- and CO-emissions. The fuel-air mixture is combusted both in the piston cone-shaped combustion chamber and in the gap between the piston surface and the cylinder head.

Since the internal combustion engine is operated with a relatively high exhaust gas recovery rate of between 50% and 70%, the local combustion temperature is below the NOx formation temperature. The local excess air ratio remains above the carbon black formation limit. Exhaust gas recuperation can be carried out with external or internal exhaust gas recuperation, or in combination with external and internal exhaust gas recuperation with variable valve distribution. Also, external exhaust gas recovery takes place by opening the intake valve during the exhaust phase and/or by opening the exhaust valve during the intake phase. Fuel injection takes place at injection pressures between 500 and 2500 bar. The focus of combustion is between top dead center and up to 10° before the crank angle of 10°, resulting in a very high efficiency. Internal combustion engines operate with a non-local air ratio of about 1.0-2.0.

An internal combustion engine is used for carrying out the method, which has at least one injection device for direct fuel injection, an exhaust gas recuperation device and at least one reciprocating piston in the cylinder, which piston has a distinct pressing surface and a Annular piston bowl shaped combustion chamber. In this respect, the piston has a circular constriction in the transition region between the pressure surface and the piston bowl-shaped combustion chamber. On the one hand, this results in a sharp squeeze flow, and on the other hand, a relatively high velocity of the liquid flow into the bowl-shaped combustion chamber is achieved. The relatively high swirl level in the piston bowl-shaped combustion chamber favorably influences the complete combustion conditions, whereby HC and CO emissions can be significantly reduced. It is particularly advantageous that the piston bowl-shaped combustion chamber is dimensioned such that the ratio of the maximum bowl-shaped combustion chamber diameter D _B to the piston diameter D is: 0.5< _DB /D<0.7, and the piston bowl-shaped combustion chamber is determined in this way Size, so that it is suitable for the ratio of the depth H _B of the largest basin-shaped combustion chamber to the diameter D of the piston: 0.12<H _B /D<0.22. As a result, the free fuel jet length can be kept as large as possible. In order to form a distinct squeeze flow, it is best to determine the size of the piston basin-shaped combustion chamber so that the ratio of the diameter D _T of the suitable contraction part to the diameter D _B of the largest basin-shaped combustion chamber is: 0.7<D _T /D _B <0.95 .

An annular shape with a flat bottom and cylindrical walls is provided as an inlet region between the extrusion surface and the constriction. Preferably the shape has a depth of between 5% and 15% of the maximum basin-shaped combustion chamber depth, the shape has an at least partially cylindrical wall, and the shape has a diameter in the region of the wall between 10% and 20% greater than the diameter of the constriction. diameter between. As a result of this shaping, the radial outflow velocity from the piston bowl-shaped combustion chamber is reduced during the downward stroke of the piston. The fuel portion is thus conveyed to the cylinder head not along the piston end face, but axially.

Description of drawings

The invention will be described in detail below with the aid of the drawings. in:

Figure 1 shows an internal combustion engine for implementing the method according to the invention;

Figures 2a and 2b show a longitudinal section of the cylinder of the internal combustion engine;

Figure 3 shows a partial enlargement III of Figure 2a, and

FIG. 4 shows an enlarged view of this part according to the state of the art.

Detailed ways

FIG. 1 shows an internal combustion engine 1 with an intake air collector 2 and an exhaust gas collector 3 . The internal combustion engine 1 is supercharged by an exhaust gas turbocharger 4 , which has an exhaust gas-driven turbine 5 and a compressor 6 driven by the turbine 5 . A charge air cooler 7 is arranged on the intake side downstream of the compressor 6 .

Furthermore, the high-pressure exhaust gas recovery system 8 has a first exhaust gas recovery line 9 between the exhaust gas line 10 and the intake line 11 . The exhaust gas recovery system 8 has an exhaust gas recovery cooler 12 and an exhaust gas recovery valve 13 . According to the pressure difference between the exhaust line 10 and the intake pipe 11, an exhaust gas pump 14 may also be provided in the first exhaust gas recovery pipe 9, so as to control or increase the exhaust gas recovery rate.

In addition to the high-pressure exhaust gas recovery system 8, there is also a low-pressure exhaust gas recovery system 15 downstream of the turbine 5 and countercurrent to the compressor 6, wherein in the exhaust gas line 16 it is branched along the downstream flow of the particle filter 17 The second exhaust gas recovery pipe 18 flows into the intake pipe 19 along the reverse flow of the compressor 6 . Furthermore, an exhaust gas recovery cooler 20 and an exhaust gas recovery valve 21 are arranged in the second exhaust gas recovery pipe 18 . A waste gas valve 22 is set in the downstream of the branch line in the waste gas pipe 16 for controlling the waste gas recovery rate.

An oxidation catalytic converter 23 is provided in the exhaust gas line 10 along the counterflow of the branch line of the first exhaust gas recovery pipe 9 to remove HC, CO and the volatile part of particulate emissions. A side effect is that at the same time the exhaust gas temperature increases and thus additional energy is delivered to the turbine 5 . In this respect, it is also possible in principle to arrange the oxidation catalytic converter 23 downstream along the branch of the exhaust gas recovery line 9 . The arrangement shown in FIG. 1 with a branch line downstream of the oxidation catalyst 23 has the advantage that the exhaust gas cooler 12 is less polluted, but has the disadvantage that a higher exhaust gas recovery cooler 12 is required due to the increased temperature of the exhaust gas. cooling capacity.

Each cylinder 24 of the internal combustion engine 1 has at least one injection valve 25 that injects diesel directly into the combustion chamber 26 , the injection start of which can be varied in the range between 50° and 5° crank angle before top dead center. The injection pressure is here between 500 and 2500 bar.

The reciprocating piston 27 in the cylinder 24 has a substantially rotationally symmetrical annular piston basin-shaped combustion chamber 28 designed as a constriction 29 of a hanging wall region 30 . The side wall of the piston basin-shaped combustion chamber 28 adopts 31, the bottom of the piston adopts 32 and the center of the raised basin-shaped combustion chamber adopts 44 to mark.

The outside of the constriction 29 is designed as a pressing surface 34 on the piston end face 33 . The geometry of the piston 27, the timing of the injection and the injection geometry of the injection valve 25 are determined in such a way that the axis 35 of the injection jet faces a region surrounding the constriction 29 between the side wall 31 and the pressure surface 34 36 (Fig. 3). This region 36 includes the hanging wall region 30 , the constriction 29 itself and the inlet region 37 formed between the pressing surface 34 and the constriction 29 by means of an annular profile 37 a. The profile 37a has a flat bottom 37b and a cylindrical wall 37c, wherein the transition radius r is designed to be between approximately 1 mm and 50% of the piston bowl-shaped combustion chamber depth H _B . The depth h of the profile 37a is approximately 5%-15% of the depth H _B of the largest bowl-shaped combustion chamber. The diameter D ₁ of the profile 37 a is 10%-20% larger than the diameter _DT of the constriction 29 . The first cut-off point 38 specific to the first injection jet axis 35 for the majority of the injected fuel quantity is located within the region 36 and is dependent on the load. At low loads, the engagement point 38 is in the region of the hanging wall region 30 . Reference numeral 39 designates the lowermost cut-in point 38 at very low loads. As the load increases, the engagement point 38 moves toward the pressing surface 34, as shown by the arrow _P2 in FIG. 3 . Reference numeral 40 in FIG. 3 indicates the uppermost extreme position of the juncture point 38 . At higher loads, therefore, a part of the injected fuel is injected counter to the direction of the squeeze flow 43 or 43 a into the squeeze space 41 between the squeeze surface 34 and the cylinder head 42 . In FIG. 2 b , reference numeral 43 designates a squeeze flow in an intake flow with swirl, while reference numeral 43 a designates a squeeze flow in an intake flow without swirl. As a result of the upward movement of the piston 27 , the engagement point 38 is displaced during the fuel injection in the direction of the piston bowl-shaped combustion chamber 28 , as indicated by the arrow P ₁ . When the piston 27 moves upwards, the effect of the squeeze flow 43, 43a produced by the squeeze surface 34 is to partly enter the fuel in the squeeze space 41 formed between the piston end face 33 and the cylinder head 42, and squeeze it out. The streams 43 , 43 a are carried in the direction of the piston bowl-shaped combustion chamber 28 and vaporized there. This results in a particularly good thorough mixing with the air, so that on the one hand the achievable maximum load in HCLI operation is increased and on the other hand the HC and CO emissions can be further reduced. Combustion takes place both within the piston bowl-shaped combustion chamber 28 and also in the region of the compression space 41 .

The shaping 37 a significantly reduces the radial outflow velocity during the downward movement of the piston 27 , so that the fraction of fuel delivered to the piston end face 33 and onward to the cylinder wall is considerably reduced. As a result, only small amounts of combustion residues get into the engine oil.

For comparison, FIG. 4 shows the region 36' of the first cut-off point of the injection jet at the start of the fuel injection in the top dead center region of a diesel internal combustion engine in conventional stratified operation. This area 36' of the fuel - regardless of the load state - is generally always in the area of the hanging wall area 30. So the tangency point does not move.

The start of the fuel injection is relatively early in the compression stroke, in particular in the lower partial load range, that is, at a crank angle of about 50°-5° before top dead center, resulting in a relatively long ignition delay for the formation of a locally homogeneous mixture , available for premixed combustion. Very low carbon black- and NOx-emission values can be achieved through sharp premixing and dilution. In this respect, the local excess air factor is always above the critical limit for soot formation. Achieving a high exhaust gas recovery rate between 50% and 70%, the local combustion temperature is always below the minimum nitrogen oxide formation temperature. The recovery of external exhaust gases takes place by opening the intake valve during the exhaust phase and/or by opening the exhaust valve during the intake phase. In this respect, the injection takes place at a pressure between 500-2500 bar. The effect of the long ignition delay is to delay the combustion phase to the efficiency sweet spot around top dead center. The combustion focus is between top dead center and up to 10° before the crank angle of 10°, whereby a high efficiency can be achieved. A high exhaust gas recovery can be achieved either solely by means of an external exhaust gas recovery or also by means of a combination of external and internal exhaust gas recovery via a variable valve control. In order to achieve a high swirl during mixture formation, it is advantageous if the swirl-generating inlet channel is used to generate up to about 5 high swirl numbers.

Piston bowl combustion chamber 28 has a relatively large maximum diameter _DB , wherein the ratio of _DB to D is in the range of 0.5-0.7. The ratio of the maximum piston depth H _B to the piston diameter D is advantageously between 0.12 and 0.22. This results in a long free jet length, which has advantages for mixture formation. In order to form a strong squeezing flow 43, the ratio of the diameter _DT of the constriction 29 to the maximum piston diameter _DB is between 0.7-0.95. This results in a high intake velocity into the piston bowl-shaped combustion chamber 28 , which facilitates the homogenization of the fuel-air mixture.

The geometry of the injection jet 35 and the geometry of the piston bowl-shaped combustion chamber 28 can be optimized for a conventional diesel internal combustion engine at full load.

Claims (18)

1. Method for operating a direct injection diesel internal combustion engine with at least one piston (27) reciprocating in a cylinder (24), wherein the internal combustion engine is operated such that the fuel is at a local temperature below the NOx formation temperature and at Combustion with a local air ratio above the soot formation limit, where fuel injection starts before top dead center in the compression phase and exhaust gas recovery in the crank angle range between 50° and 5°, where the exhaust gas recovery is about 50 %-70%, it is characterized in that at least one piston (27) is provided with at least one extrusion surface (34) and an annular piston basin-shaped combustion chamber (28) and an extrusion surface (34) and piston basin The constriction (29) in the transition zone between the shaped combustion chambers (28); when the piston (27) moves upwards, a squeeze flow (43) is generated from the outside to the inside of the piston bowl shaped combustion chamber (28) ; the fuel is injected at least mainly into the annular piston basin combustion chamber (28) and is conveyed under at least partial vaporization along the piston basin combustion chamber side walls (31) and/or the piston bottom (32) by the squeeze flow (43) , and the fuel is injected in the direction of the constriction (29) of the piston (27), wherein at the beginning of the fuel injection the tangent point (38) of the jet axis (35) of at least one fuel jet of the majority of the fuel quantity is in the basin In an area (36) between the side wall (31) and the extrusion surface (34) of the shaped combustion chamber, this area includes the hanging wall area (30), the constriction (29) and the extrusion surface and the constriction (29) Between the intake area (37), the cut-off point (38) is adjusted according to an area of the hanging wall area (30) inside the piston bowl-shaped combustion chamber (28) under low load, and the cut-off point (38) is adjusted as the load increases ) moves in the direction of the constriction (29). 2. by the described method of claim 1, it is characterized in that, in cylinder (24), produce a kind of intake flow with vortex number ≥ 1 with vortex, and fuel oil burns along piston basin shape by extrusion flow (43) The chamber side walls ( 31 ) are conveyed under at least partial vaporization in the direction of the piston bottom ( 32 ) and continue along the piston bottom ( 32 ) towards the center ( 44 ) of the basin-shaped combustion chamber. 3. The method according to claim 1, characterized in that, in the cylinder (24), a swirl-free intake air flow with a swirl number<1 is produced, and the fuel oil passes through the squeeze flow (43) under at least partial vaporization from The center of the basin-shaped combustion chamber (44) is conveyed to the side wall (31) of the piston basin-shaped combustion chamber along the bottom of the piston (32) and continues to the contraction part (29). 4. The method according to any one of claims 1-3, characterized in that as the load increases, the fuel injection starts from an area of about 5°-15° allocated to the low partial load before the top dead center until the top dead center About 50° before the crank angle starts early. 5. The method as claimed in one of claims 1-3, characterized in that the fuel injection takes place at an injection pressure of 500-2500 bar. 6. The method according to any one of claims 1-3, characterized in that the combustion emphasis is within the crank angle range between 10° before top dead center and 10° after top dead center. 7. The method as claimed in claim 1, characterized in that the non-local air ratio is set between 1.0 and 2.0. 8. The method as claimed in one of claims 1 to 3, characterized in that the waste gas recovery takes place externally and/or internally. 9. The method as claimed in claim 8, characterized in that the external exhaust gas recuperation takes place by opening the intake valve during the exhaust phase and/or by opening the exhaust valve during the intake phase. 10. Direct injection diesel internal combustion engine, be used for carrying out by the described method of one of claim 1-12, utilize this method to be able to start fuel injection between 50 ° to 5 ° of crank angle before top dead center of compression stage range, and with an exhaust gas recovery system with an exhaust gas recovery rate between 50% and 70%, with at least one piston (27) reciprocating in the cylinder (24), characterized in that the piston (27 ) has on its end face (33) at least one extrusion surface (34) and an annular piston basin-shaped combustion chamber (28) with a constriction (29), substantially concavely curved side walls (31) and Bottom (32), and a hanging wall region (30) between a side wall (31) and the constriction, wherein at least one jet axis (35) of the fuel jet of the fuel injection device (25) for the majority of the fuel quantity Facing a region (36) between side wall (31) and extrusion surface (34), this region (36) comprises hanging wall area (30), contraction part (29) and extrusion surface (34) and contraction The inlet area (37) between the points (29), the point of intersection (38) of the jet axis (35) is carried out at low loads by means of an area of the hanging wall area (30) inside the piston bowl-shaped combustion chamber (28) Adjustment, along with the rising of load, the cutting point (38) moves to the direction of the contraction part (29). 11. Internal combustion engine according to claim 10, characterized in that the piston bowl-shaped combustion chamber (28) is dimensioned such that the ratio of the maximum bowl-shaped combustion chamber diameter ( _DB ) to the piston diameter (D) is suitable for: 0.5< _DB /D<0.7. 12. Internal combustion engine according to one of claims 10 or 11, characterized in that the piston bowl (28) is dimensioned such that it is suitable for the maximum bowl depth (H _B ) and piston diameter (D) The ratio is: 0.12<H _B /D<0.22. 13. Internal combustion engine according to one of claims 10 or 11, characterized in that the piston pot combustion chamber (28) is dimensioned such that it is suitable for the diameter (D _T ) of the constriction (29) and the maximum pot combustion The ratio of chamber diameter (D _B ) is: 0.7<D _T /D _B <0.95. . 14. The internal combustion engine as claimed in claim 10, characterized in that the intake region (37) has an annular shape (37a) between the pressing surface (34) and the constriction (29). 15. The internal combustion engine as claimed in claim 14, characterized in that the molding (37a) has a flat base (37b) which opens into the piston bowl-shaped combustion chamber (28). 16. The internal combustion engine as claimed in claim 14, characterized in that the shaping (37a) has a depth (h) of between 5% and 15% of the maximum bowl-shaped combustion chamber depth (H _B ). 17. The internal combustion engine as claimed in claim 14, characterized in that the molding (37a) has an at least partially cylindrical wall (37c). 18. Internal combustion engine according to claim 17, characterized in that the molding (37a) has a diameter ( _{D 1} ).

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