WO2004057167A1 - Method for operating a direct-injection diesel engine - Google Patents

Abstract

The invention relates to a method for operating a direct-injection Diesel engine that comprises at least one piston (27), reciprocating in a cylinder (24), whereby the internal combustion engine is operated in such a manner that the fuel is combusted at a local temperature below the Nox formation temperature and at a local air ratio above the soot formation limit. Fuel injection is started in a crank angle range of between 50 DEG and 5 DEG before the upper dead center of the compression phase and exhaust gas is returned, the exhaust gas return rate amounting to approximately 50 % to 70 %. The aim of the invention is to achieve particularly low nitrogen oxide and soot emissions. For this purpose, at least one piston (27) is provided with a squish area (34) and a toroid piston recess (28) and a necking (29) in the transition area between the squish area (34) and the piston recess (28). When the piston (27) moves upward a squish flow (43) is produced which flows from the outside to the inside. At least the major portion of the fuel is injected into the toroid piston recess (28) and is transported by the squish flow (43) along the piston recess wall (31) and/or the piston head (32) while at least partially evaporating.

Description

Method for operating a direct-injection diesel internal combustion engine

The invention relates to a method for operating a direct-injection diesel internal combustion engine with at least one piston reciprocating in a cylinder, the internal combustion engine being operated such that the combustion of the fuel at a local temperature below the NOx formation temperature and with a local air ratio occurs above the soot formation limit, the fuel injection is started in a range between 50 ° and 5 ° crank angle before top dead center of the compression phase and exhaust gas is recirculated, and the exhaust gas recirculation rate is approximately 50% to 70%. The invention further relates to an internal combustion engine for carrying out the method.

The most important determinants for the combustion process in an internal combustion engine are the phase position of the combustion process or the start of combustion, the maximum rate of increase of the cylinder pressure and the peak pressure.

In the case of an internal combustion engine in which the combustion occurs essentially by means of auto-ignition of a directly injected fuel quantity, the determinants are largely determined by the time of injection, by the charge composition and by the ignition delay. These parameters are in turn determined by a large number of influencing variables, such as speed, fuel quantity, intake temperature, boost pressure, effective compression ratio, exhaust gas content of the cylinder charge and component temperature.

Strict legal framework conditions mean that new approaches have to be taken in the design of combustion processes in order to reduce the emission of soot particles and NOx emissions in diesel engines.

It is known to reduce NOx and soot emissions in the exhaust gas by increasing the ignition delay by bringing the injection timing forward, so that the combustion takes place by auto-ignition of a lean fuel / air mixture. One possible variant is referred to here as the HCLI method (Homogeneous Charge Late Injection). If the fuel injection is carried out sufficiently far before the top dead center of the compression phase, such a mixture formation takes place, as a result of which a largely premixed fuel / air mixture is produced. Exhaust gas recirculation can achieve that the combustion temperature remains below the minimum temperature required for the formation of NOx.

No. 6,338,245 B1 describes a diesel internal combustion engine working according to the HCLI process, in which the combustion temperature and ignition delay are set so that in the lower and middle part-load range the combustion temperature is below the NOx formation temperature and the air ratio is above the value relevant for soot formation. The combustion temperature is controlled by changing the exhaust gas recirculation rate, the ignition delay is controlled by the fuel injection time. At medium and high loads, the combustion temperature is reduced to such an extent that both NOx and soot formation are reduced.

Furthermore, it is known to design pistons for diesel engines with an essentially toroidal piston recess. A constriction is arranged in the transition area between the piston end face and the piston recess, which forms a relatively narrow overflow cross section. The narrow overflow cross section provides a high mixture formation energy, which significantly improves fuel processing. Pistons with such a toroidal piston recess are known for example from the publications EP 0 383 001 AI, DE 1 122 325 AS, AT 380 311 B, DE 21 36 594 AI, DE 974 449 C or JP 60-206960 A. In conventionally operated internal combustion engines, such pistons have the following advantageous effects on the operating behavior of the internal combustion engine: The smoke-limiting full load can be increased; it is possible to achieve high compression, which results in a lower combustion noise due to a smaller ignition delay, lower hydrocarbon emissions, a more favorable starting behavior of the engine and an improvement in the efficiency of the internal combustion engine; Furthermore, there is the possibility of moving the ignition point late, without significant increase in smoke, consumption and HC, due to the fact that the mixture formation energy remains high over a longer period of time. This option means above all a reduction in nitrogen oxides, combustion noise and cylinder tip pressure.

Furthermore, a piston with a piston recess and a constriction is known from the publication DE 11 22 325 Cl, a recess being provided between the squeezing surface and the constriction.

In internal combustion engines operating according to the HCLI process, such piston shapes with a deep, constricted piston bowl have not been used, since it was previously assumed that the deep piston bowl and the strong squeezing flow made the starting ability and thermodynamic efficiency too strong would deteriorate. US Pat. No. 6,158,413 A therefore proposes to suppress the squeezing flow at all, using a piston with a very flat piston recess.

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

According to the invention, this is achieved in that at least one piston with at least one squeeze surface and a toroidal piston recess and a constriction in the transition region between the squeeze surface and the piston recess is provided in that when the piston moves upward, a squeezing flow is generated from the outside inwards into the piston recess, that the Fuel is at least predominantly injected into the toroidal piston recess and is transported by the squeezing flow along the side of the piston recess and / or the piston crown with at least partial evaporation. The fuel jet is injected into the squeezing flow flowing into the piston bowl. The squeezed flow directs most of the fuel into the piston bowl, where it vaporizes and is mixed almost homogeneously with the incoming air. The flow in the piston bowl depends on whether there is a swirled or swirlless inlet flow.

In an embodiment variant according to the invention, it is provided that a swirling inlet flow with a swirl number> 1 is generated in the cylinder and that the fuel is transported by the squeezing flow along the side of the piston recess with at least partial evaporation in the direction of the piston crown and further along the piston crown to the bowl center. The twist is maintained within the piston bowl during the compression phase.

In another embodiment, on the other hand, it is provided that a swirl-free inlet flow with a swirl number <1 is generated in the cylinder and that the fuel is transported by the squeeze flow with at least partial evaporation from the bowl center along the piston crown to the piston bowl side wall and further to the constriction.

Surprisingly, it has been shown that the ability to start in internal combustion engines operating according to the HCLI process is not significantly impaired by the retracted piston recess. The loss of thermodynamic efficiency due to the squeezing flow can be Mix preparation in the piston bowl due to the high turbulence more than made up for.

It is preferably provided that the fuel is injected in the direction of the constriction of the piston, the intersection of the jet axis of at least one injection jet for a large part of the fuel quantity being in the area between the trough side wall and the squeezing area, the overhanging wall area, the constriction, at the start of injection as well as an inlet area between the squeeze surface and the constriction.

In conventional diesel internal combustion engines, the point of intersection and the time of injection of the fuel are usually selected so that the fuel hits the overhanging wall area below the constriction at the start of injection, regardless of the load. The object of the present invention provides that the intersection point is set at a low load to a region of the overhanging wall area within the piston recess, and that the intersection point is shifted in the direction of the constriction as the load increases. This can be achieved by moving the injection point forward. As a result, part of the fuel is injected into the gap between the piston and the cylinder head - against the squeezing flow. A large part of the fuel injected into the space between the piston surface and the cylinder head is entrained by the squeezing flow into the piston bowl. This improved air distribution and mixture preparation with an advantageous reduction in HC and CO emissions. The combustion of the fuel-air mixture takes place both in the piston bowl and in the space between the piston surface and the cylinder head.

Since the internal combustion engine is operated with relatively high exhaust gas recirculation rates between 50 ° and 70 °, the local combustion temperature is below the NOx formation temperature. The local air ratio remains above the soot formation limit. Exhaust gas recirculation can be achieved by external or internal exhaust gas recirculation or by a combination of external and internal exhaust gas recirculation with variable valve control. The fuel is injected at an injection pressure between 500 and 2500 bar. The center of combustion is between 10 ° before and 10 ° crank angle after top dead center, which results in a very high efficiency. The internal combustion engine is operated with a global air ratio of approximately 1.0 to 2.0.

An internal combustion engine with at least one injection device for direct fuel injection, with an exhaust gas recirculation device and at least one piston reciprocating in a cylinder, which has a pronounced squeeze surface and a toroidal shape, is suitable for carrying out the method Has piston bowl. The piston has a circular constriction in the transition area between the squeeze surface and the piston recess. On the one hand, this creates a pronounced squeezing flow and, on the other hand, ensures that the flow flows into the trough at a relatively high speed. The relatively high level of turbulence within the piston bowl has an advantageous effect on the blow-through behavior, as a result of which HC and CO emissions can be significantly reduced. It is particularly advantageous if the piston bowl is dimensioned such that the following applies to the ratio of the largest bowl diameter D _B TO piston diameter D: 0.5 <D _B / D <0.7 and if the piston bowl is dimensioned such that the ratio is largest Trough depth H _B to piston diameter D applies: 0.12 <H _B / D <0.22. This allows the free fuel jet length to be kept as large as possible. To form a pronounced squeezing flow, it is preferably provided that the piston recess is dimensioned such that the following applies to the ratio of the diameter D _{τ of} the constriction to the largest recess diameter D _B : 0.7 <D _T / D _B <0.95.

A circumferential annular recess with a flat bottom and a cylindrical wall is arranged between the squeeze surface and the constriction as the inlet area. It is preferably provided that the indentation has a depth between 5% and 15% of the greatest trough depth, that the indentation has an at least partially cylindrical wall and that the indentation in the region of the wall has a diameter which is between 10% and 20% greater than the diameter of the constriction. The shaping reduces the radial outflow velocity from the piston recess when the piston descends. As a result, fuel components are not directed along the piston face, but in the axial direction to the cylinder head.

The invention is explained in more detail below with reference to the figures. They show schematically

1 shows an internal combustion engine for carrying out the method according to the invention,

2a and 2b a cylinder of this internal combustion engine in longitudinal section,

Fig. 3 shows the detail III of Fig. 2a and

Fig. 4 this detail according to the prior art.

1 shows an internal combustion engine 1 with an intake manifold 2 and an exhaust manifold 3. The internal combustion engine 1 is powered by an exhaust gas turbocharger 4, which drives an exhaust gas-powered turbine 5 and a turbine 5. benen compressor 6, charged. A charge air cooler 7 is arranged upstream of the compressor 6 on the inlet side.

Furthermore, a high-pressure exhaust gas recirculation system 8 with a first exhaust gas recirculation line 9 is provided between the exhaust line 10 and the inlet line 11. The exhaust gas recirculation system 8 has an exhaust gas recirculation cooler 12 and an exhaust gas recirculation valve 13. Depending on the pressure difference between the outlet line 10 and the inlet line 11, an exhaust gas pump 14 can also be provided in the first exhaust gas recirculation line 9 in order to control or increase the exhaust gas recirculation rate.

In addition to this high-pressure exhaust gas recirculation system 8, a low-pressure exhaust gas recirculation system 15 is provided downstream of the turbine 5 and upstream of the compressor 6, a second exhaust gas recirculation line 18 branching off in the exhaust line 16 downstream of a particle filter 17 and opening into the intake line 19 upstream of the compressor 6. An exhaust gas recirculation cooler 20 and an exhaust gas recirculation valve 21 are also arranged in the second exhaust gas recirculation line 18. To control the exhaust gas recirculation rate, an exhaust valve 22 is arranged in the exhaust line 16 downstream of the branch.

An upstream of the branch of the first exhaust gas recirculation line 9, an oxidation catalytic converter 23 is arranged in the exhaust line 10, which removes HC, CO and volatile parts of the particle emissions. A side effect is that the exhaust gas temperature is increased and additional energy is supplied to the turbine 5. In principle, the oxidation catalytic converter 23 can also be arranged downstream of the branch of the exhaust gas recirculation line 9. The arrangement shown in FIG. 1 with the branch downstream of the oxidation catalytic converter 23 has the advantage that the exhaust gas cooler 12 is exposed to less contamination, but the disadvantage that the exhaust gas recirculation cooler 12 requires a higher cooling capacity due to the higher exhaust gas temperatures.

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

The piston 27 reciprocating in the cylinder 24 has an essentially rotationally symmetrical toroidal piston recess 28 with a constriction 29, which forms an overhanging wall region 30. The side wall of the piston bowl 28 is designated 31, the piston crown 32, and the raised bowl center 44. A pinch surface 34 is formed on the piston face 33 outside the constriction 29. The geometric shape of the piston 27, the injection timing and the injection geometry of the injection valve 25 are dimensioned such that the axes 35 of the injection jets are directed to an area 36 (FIG. 3) around the constriction 29 between the side wall 31 and the squeeze surface 34. The region 36 includes the overhanging wall region 30, the constriction 29 itself, and an inlet region 37, formed by a circumferential annular indentation 37a, between the squeeze surface 34 and the constriction 29. The indentation 37a has a flat bottom 37b and a cylindrical wall 37c, wherein a transition radius r between approximately 1 mm and 50% of the piston bowl depth H _{B is} formed. The depth h of the indentation 37a is approximately 5% to 15% of the greatest trough depth H _B. The diameter Di of the indentation 37a is 10% to 20% larger than the diameter D _{τ of} the constriction 29. The actual first intersections 38 of the axes 35 of the first injection jets of the majority of the injected fuel quantity lie within the range 36 and become dependent on the load changed. At a low load, the intersection 38 lies in the region of the overhanging wall area 30. The lowest intersection 38 is indicated at 40 with a very low load. As the load increases, the intersection 38 is shifted in the direction of the squeeze surface 34, as indicated in FIG. 3 by arrows Pi. Reference numeral 40 marks the uppermost extreme position for the intersection 38 in FIG. 3. At higher loads, part of the injected fuel is thus injected into the squeeze space 41 between the squeeze surface 34 and the cylinder head 42 against the direction of the squeeze flow 43 or 43a. In FIG. 2b, the squeezing flow with swirled inlet flow is shown with reference number 43 and the squeezing flow with swirlless inlet flow with reference number 43a. As a result of the upward movement of the piston 27, the intersection 38 moves in the direction of the piston recess 28 during an injection, as indicated by arrow P ₂ . The squeezing flow 43, 43a generated by the squeezing surface 34 when the piston 27 moves upward causes part of the fuel squeezed into the squeezing space 41 formed between the piston end face 33 and cylinder head 42 to be entrained by the squeezing flow 43, 43a in the direction of the piston recess 28 and evaporates there , This results in a particularly good mixing with the air, which on the one hand increases the maximum achievable load in HCLI operation and on the other hand further reduces HC and CO emissions. The combustion takes place both within the piston recess 28 and in the area of the pinch chamber 41.

The indentation 37a significantly reduces the radial outflow speed when the piston 27 moves downward, as a result of which substantially less fuel is carried to the piston top 33 and further to the cylinder wall. be changed. As a result, only a few combustion residues get into the engine oil.

For comparison, FIG. 4 shows the region 36 'of the first intersection of the injection jet at the start of fuel injection in the region of the top dead center of a conventional stratified diesel internal combustion engine. The region 36 'of the fuel - regardless of the load state - usually always remains in the region of the overhanging wall region 30. The intersection is therefore not shifted.

The start of the injection, particularly in the lower part of the load range, is relatively early in the compression cycle, i.e. at a crank angle of 50 ° to 5 ° before top dead center, as a result of which a long ignition delay is available for the formation of a partially homogeneous mixture for premixed combustion. Due to the pronounced premixing and dilution, extremely low soot and NOx emission values can be achieved. The local air ratio always remains well above the limit for the formation of soot. A high exhaust gas recirculation rate between 50% and 70% ensures that the local combustion temperature always remains below the minimum nitrogen oxide formation temperature. The injection takes place at a pressure between 500 and 2500 bar. The long ignition delay means that the combustion phase is pushed into the most efficient position around top dead center. The center of combustion is in a range between about 10 ° crank angle before to about 10 ° crank angle after top dead center, whereby a high efficiency can be achieved. The high exhaust gas recirculation rate can be achieved either by external exhaust gas recirculation alone or by combining external with internal exhaust gas recirculation through variable valve control. In order to achieve a high level of turbulence in the mixture formation, swirl-generating inlet channels are advantageous for generating a high swirl number up to about 5.

The piston recess 28 has a relatively large maximum diameter D _B , the ratio D _B to D being in the range between 0.5 to 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 allows a long free jet length to be generated, which is advantageous for the mixture formation. In order to form a strong squeezing flow 43, the ratio of the diameter D _{τ of} the constriction 29 to the maximum piston diameter D _{B is} between 0.7 and 0.95. As a result, high entry speeds into the piston recess 28 are achieved, which has a favorable effect on the homogenization of the fuel-air mixture. The geometry of the injection jets 35 and the geometry of the piston recess 28 can be optimized for a conventional diesel engine at full load.

Claims

PATENT CLAIMS 1. A method for operating a direct-injection diesel internal combustion engine with at least one piston (27) reciprocating in a cylinder (24), the internal combustion engine being operated such that the combustion of the fuel at a local temperature below the NOx formation temperature and with a local air ratio above the soot formation limit, the fuel injection being started in a range between 50 ° and 5 ° crank angle before the top dead center of the compression phase and exhaust gas being recirculated, and the exhaust gas recirculation rate being approximately 50% to 70%, characterized in that that at least one piston (27) with at least one squeeze surface (34) and a toroidal piston recess (28) and a constriction (29) is provided in the transition area between the squeeze surface (34) and piston recess (28), that when the piston (27) moves upwards a squeezing stream directed from the outside inwards into the piston recess (28) generated (43) that the fuel is at least predominantly injected into the toroidal piston recess (28) and is transported by the squeezing flow (43) along the piston recess side wall (31) and / or the piston head (32) with at least partial evaporation. 2. The method according to claim 1, characterized in that a swirled inlet flow with a swirl number> 1 is generated in the cylinder (24) and that the fuel by the squeezing flow (43) along the piston bowl side wall (31) with at least partial evaporation towards the piston crown ( 32) and further along the piston crown (32) to the trough center (44). 3. The method according to claim 1, characterized in that a swirl-free inlet flow with a swirl number <1 is generated in the cylinder (24) and that the fuel through the squeeze flow (43) with at least partial evaporation from the trough center (44) along the piston crown (32 ) is transported to the piston recess side wall (31) and further to the constriction (29). 4. The method according to any one of claims 1 to 3, characterized in that the fuel is injected in the direction of the constriction (29) of the piston (27), the intersection (38) of the jet axis (35) of at least one injection jet for one at the start of injection Most of the amount of fuel lies in an area (36) between the trough side wall (31) and the squeeze surface (34), which is an overhanging wall area (30) Contains constriction (29) and an inlet area (37) between the squeeze surface and constriction (29). 5. The method according to claim 4, characterized in that the intersection (38) is set at low load to a region of the overhanging wall region (30) within the piston recess (28). 6. The method according to claim 4 or 5, characterized in that with increasing load, the intersection (38) is shifted in the direction of the constriction (29). 7. The method according to any one of claims 1 to 6, characterized in that with increasing load, the start of injection from a range assigned to the low partial load from about 5 ° to 15 ° before top dead center to about 50 ° crank angle before top dead center is brought forward. 8. The method according to any one of claims 1 to 7, characterized in that the fuel injection takes place at an injection pressure between 500 to 2500 bar. 9. The method according to any one of claims 1 to 8, characterized in that the center of combustion is placed in a crank angle range between 10 ° before top dead center to 10 ° after top dead center. 10. The method according to any one of claims 1 to 9, characterized in that the global air ratio is set between 1.0 and 2.0. 11. The method according to any one of claims 1 to 10, characterized in that the exhaust gas recirculation is carried out externally and / or internally. 12. The method according to claim 11, characterized in that an internal exhaust gas recirculation is carried out by opening the intake valve during the exhaust phase and / or by opening the exhaust valve during the intake phase. 13. Direct-injection diesel internal combustion engine for performing the method according to one of claims 1 to 12, with which the start of fuel injection can be set in a range between 50 ° to 5 ° crank angle before top dead center of the compression phase, and with an exhaust gas recirculation system for exhaust gas recirculation rates between 50% and 70%, with at least one piston (27) reciprocating in a cylinder (24), characterized in that the piston (27) on its Front side (33) at least one squeezing surface (34) and a toroidal piston recess (28) with a constriction (29), essentially concave curved side walls (31) and bottom (32), and an overhanging wall area (30) between side walls (31) and constriction, wherein at least one jet axis (35) of a fuel injection jet of the injection device (25) for a large part of the fuel quantity at the start of injection is directed to an area (36) between the side wall (31) and the squeeze surface (34), which area (36 ) contains the overhanging wall area (30), the constriction (29), and an inlet area (37) between the squeezing surface (34) and the constriction (29). 14. Internal combustion engine according to claim 13, characterized in that the intersection (38) of the at least one jet axis (35) of the fuel jet at the start of injection can be varied at least between the overhanging wall area (30) and the constriction (29). 15. Internal combustion engine according to one of claims 13 or 14, characterized in that the piston bowl (28) is dimensioned such that the following applies to the ratio of the largest bowl diameter (D _B ) to piston diameter (D): 0.5 <D _B / D < 0.7. 16. Internal combustion engine according to one of claims 13 to 15, characterized in that the piston recess (28) is dimensioned such that the following applies to the ratio of the greatest recess depth (H _B ) to piston diameter (D): 0.12 <H _B / D < 0.22. 17. Internal combustion engine according to one of claims 13 to 16, characterized in that the piston recess (28) is dimensioned such that for the ratio diameter (D _τ ) of the constriction (29) to the largest trough diameter (D _B ) applies: 0.7 <D _T / D _B <0.95. 18. Internal combustion engine according to one of claims 13 to 17, characterized in that the inlet region (37) has a circumferential annular recess (37a) between the squeeze surface (34) and the constriction (38). 19. Internal combustion engine according to claim 18, characterized in that the indentation (37a) has a flat bottom (37b) which runs out to the piston recess (28). 20. Internal combustion engine according to claim 18 or 19, characterized in that the indentation (37a) has a depth (h) between 5% and 15% of the greatest trough depth (H _B ). 21. Internal combustion engine according to one of claims 18 to 20, characterized in that the indentation (37a) has an at least partially cylindrical wall (37c). 22. Internal combustion engine according to one of claims 18 to 21, characterized in that the indentation (37a) in the region of the wall (37c) has a diameter (Di) which is between 10% to 20% larger than the diameter (D _τ ) the constriction (29).