WO2004088120A2

WO2004088120A2 - Internal combustion engine comprising a piston injection pump

Info

Publication number: WO2004088120A2
Application number: PCT/AT2004/000116
Authority: WO
Inventors: Martin Rieckh; Michael Weissbäck; Janos Csato; Michael Glensvig; Theodor Sams; Peter Herzog
Original assignee: Avl List Gmbh
Priority date: 2003-04-01
Filing date: 2004-04-01
Publication date: 2004-10-14
Also published as: WO2004088120A3; DE112004000500A5

Abstract

The invention relates to an internal combustion engine comprising a piston injection pump for transporting fuel, said pump comprising a cylinder for receiving a reciprocating pressure piston, the front face of the pressure piston defining a pressure chamber. The edge of the front face forms a peripheral first control edge. The cylinder comprises at least one radial spill port which is located in a piston guiding section of the cylinder and can be covered at least by a second control edge of the pressure piston, formed by shaping the envelope surface of the pressure piston.

Description

Internal combustion engine

The invention relates to an internal combustion engine, in particular a diesel internal combustion engine, with at least one piston injection pump for delivering fuel with a cylinder for receiving a reciprocating pressure piston, the end face of which borders a pressure chamber, the edge of the end face forming a circumferential first control edge, and the Cylinder has at least one radial control bore in a piston guide section of the cylinder, which can be run over by at least one second control edge of the pressure piston formed by a shape in the outer surface of the pressure piston. Furthermore, the invention relates to a method for operating an internal combustion engine, in particular a diesel internal combustion engine, and a device for carrying out the method.

The conventional design of a piston injection pump has a smooth cylinder surface with a radial control bore and a pressure piston with an upper control edge, usually running in a normal plane to the cylinder axis and a lower control edge running inclined to the normal plane, and a zero-delivery groove running in the axial direction. Piston injection pumps of this type are known from US 5,396,871 A, US 4,964,789 A or US 4,824,341 A. When the control bore is covered by the control surface of the pressure piston, fuel is delivered. Due to the sloping lower control edge, the injection quantity can be controlled by turning the piston. In the case of injection pumps for very high pressures, the features described can also be provided twice, offset by 180 °, in order to create conditions that are as symmetrical as possible.

Due to a lateral deflection of the pressure piston in a direction normal to the pilot bores, the resulting pressure distribution in the sealing gap can result in a lateral force on the piston, which acts in the direction of the deflection and thus reinforces it. As a result, the pressure piston may rub.

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 an internal combustion engine in which the combustion takes place essentially by means of auto-ignition of a directly injected fuel quantity, the Determinants are largely determined by the time of injection, the charge composition and 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, inert gas content of the cylinder charge and component temperature.

Stricter legal framework conditions mean that new approaches have to be taken in the design of combustion processes in order to reduce emissions 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 advancing the injection timing, so that the combustion takes place through self-ignition of a lean fuel / air mixture. One possible variant is referred to here as the HCLI method (Homogeneous Charge Late Injection). If such a mixture combustion is carried out, the fuel is thus injected sufficiently far before the top dead center of the compression phase, which results in a largely homogeneous fuel-air mixture. Exhaust gas recirculation can ensure that the combustion temperature remains below the minimum temperature required for the formation of NOx. However, since the homogenization of fuel and air is time-dependent, the implementation of this method is speed-dependent and load-dependent, since the particle emissions increase with insufficient homogenization.

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 so much that both NOx and soot formation are avoided. It is disadvantageous that, especially in the middle part-load range, a low air ratio combined with low combustion temperatures occurs and therefore a poor efficiency has to be accepted.

US Pat. No. 6,158,413 A describes a direct-injection diesel internal combustion engine in which the fuel injection is not started before the top dead center of the compression, and in which the oxygen concentration in the combustion space is reduced by exhaust gas recirculation. This operating method is also referred to here as HPLI (Highly Premixed Late Injection). Because of the - compared to conventional injection before top dead center - the falling temperature level after top dead center and the increased amount of recirculated exhaust gas compared to conventional operation, the ignition delay is longer than with so-called diffusion combustion. The low temperature level controlled by the exhaust gas recirculation rate causes the combustion temperature to remain below the value that is decisive for the formation of NOx. Due to the large ignition delay caused by the later injection time, a good mixture formation is achieved, whereby the local lack of oxygen is significantly reduced when the mixture is burned, thereby reducing the formation of particles. The late shift in the burning process causes a lowering of the maximum temperature, but at the same time leads to an increase in the average temperature at a given late crank angle, which increases the soot burn-off. The shift of the combustion into the expansion stroke also leads, in conjunction with the high exhaust gas recirculation rate, despite the larger premixed fuel quantity due to the long ignition delay and consequently higher maximum combustion rate, to a pressure rise rate in the cylinder which does not exceed the permissible level. The disadvantage is the poor efficiency in the lower part-load range.

The object of the invention is to avoid these disadvantages and to reduce wear on a piston injection pump of the type mentioned, in particular to prevent the pressure piston from rubbing against the cylinder.

Another object of the invention is to develop a method for operating an internal combustion engine with which, on the one hand, minimal nitrogen oxide and soot emissions can be achieved from the lower partial load range to the full load range, and on the other hand a high degree of efficiency.

According to the invention, this is done in that the cylinder has at least two, preferably diametrically arranged, pressure compensation channels for pressure compensation between two different peripheral regions of the piston guide section of the cylinder. Due to the pressure compensation channels, the lateral forces can be reduced as a result of a lateral deflection. It is essential that the pressure compensation channels in the region of the piston guide section of the cylinder are spaced from the control bore, so that a sealing surface that can be covered by the pressure piston is formed around the control bore of the cylinder. The sealing function of the pressure piston in the area around the pilot hole is thus still fully guaranteed. The lateral forces can be significantly reduced if the pressure compensation channels at least in a conveying position of the pressure piston fluidly connect the pressure chamber to an area of the piston guide section of the cylinder which is approximately at the level of the control bore.

In a very advantageous embodiment variant of the invention it is provided that at least one pressure compensation channel is formed by at least one recess in the cylinder, preferably by a longitudinal groove running parallel to the cylinder axis. Several narrow axial grooves or a few wide axial grooves can be formed in the cylinder wall.

As an alternative to this, it can also be provided that at least one pressure compensation channel has at least two transverse bores which open approximately radially into the cylinder and are connected to one another by flow through at least one longitudinal bore running approximately parallel to the cylinder axis. At least two radial transverse bores can be arranged directly one above the other in the piston guide section in the axial direction of the cylinder, preferably at the level of the control bore. The cross bores can in particular be arranged one above the other in at least one axial row.

Unwanted lateral forces can be largely prevented if at least two pressure compensation channels are arranged in the circumferential direction on both sides and preferably symmetrically to the control bore. In particular, it is advantageous if at least two pressure compensation channels are arranged point-symmetrically with respect to the cylinder axis.

The method according to the invention provides the following steps:

Operating the internal combustion engine in a first operating range assigned to the low partial load with largely homogeneous mixture combustion and later fuel injection, the fuel injection being started in a range between approximately 50 ° and 5 ° crank angle before top dead center of the compression phase;

Operating the internal combustion engine in a second operating range associated with the medium partial load with low-temperature mixture combustion and even later injection than in the first operating range, fuel injection in a range between 2 ° crank angle before top dead center and about 20 ° crank angle after top dead center Compression phase is started

the fuel in the first operating area via first injection openings and in the second operating area at least via second injection ports injection openings of an injection valve, preferably designed as a double needle nozzle, is injected into the combustion chamber.

In the first operating range, the internal combustion engine works according to the HCLI method, in which the injection time is relatively early in the compression stroke, that is to say in the range of approximately 50 ° to 5 ° crank angle before top dead center. The fuel is preferably injected in the first operating range in a range between 400 and 1000 bar. The center of combustion is between 10 ° before and 10 ° crank angle after top dead center, which means that very high efficiency can be achieved. Due to the relatively high exhaust gas recirculation rate between 50% and 70%, the local combustion temperature is below the NOx formation temperature. The local air ratio thus 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.

In the second operating area, the internal combustion engine is operated using the HPLI method. The main part of the injection phase lies after the top dead center of the compression. Because of the - compared to conventional injection before top dead center - the temperature level falling after top dead center and the increased amount of recirculated exhaust gas compared to conventional operation, the ignition delay is longer here. If necessary, other means such as lowering the effective compression ratio and / or the inlet temperature and increasing the injection pressure and / or increasing the injection hole cross-sections of the injection nozzle can be used to extend the ignition delay. The injection duration is designed in such a way that the end of injection is before the start of combustion. In this case, soot emissions can be kept at a very low level. This can be explained by the fact that the simultaneous occurrence of liquid fuel in the fuel jet on the one hand and the flame conventionally enveloping the jet on the other hand is avoided, as a result of which the oxidation reactions near the jet, which otherwise lead to soot formation and run off due to lack of air, are also prevented. Injection pressures of at least 1000 bar are required for the combustion process in the second operating range. The advantage of this method is that very low NOx and particle emissions arise and that a relatively high exhaust gas temperature is reached, which in turn is advantageous in the regeneration of particle exhaust gas aftertreatment devices.

It is preferably provided that the fuel is injected in the first operating range with a lower flow rate than in the second operating range. Particularly low nitrogen oxide and soot emissions can be achieved if the fuel is injected in the first and in the second operating area in fuel jets arranged along a conical surface, the fuel being injected in the first operating area with a different, preferably smaller, cone opening angle than in the second operating area.

In a further embodiment of the invention it is provided that in a third operating range assigned to the upper partial load and the full load, the main part of the fuel injection takes place in a range between 10 ° before and 10 ° crank angle after the top dead center, it being preferably provided that in the third Operating range a multiple injection is carried out. The exhaust gas recirculation rate in the third operating range is up to 30%, preferably about 10% to 20%. This enables high performance on the one hand, and low NOx emissions and low particle emissions on the other.

In the third operating range, the fuel can be injected through the first and / or through the second injection openings.

The internal combustion engine is operated in the first, second and / or third operating range with a global air ratio of approximately 1.0 to 2.0.

It is advantageously provided that the exhaust gas recirculation is carried out externally and / or internally and the swirl is variable at least in one area, preferably in all three areas. Favorable exhaust gas values with low fuel consumption can be achieved with swirl numbers between 0 and 5.

It is also advantageous if the geometric compression ratio is variable. The geometric compression ratio can be changed in a range between 13 and 19. A high compression ratio is beneficial for the cold start phase. A reduction in the compression ratio during the load increase increases the maximum achievable load both in the first and in the second operating range and reduces the soot emissions due to a longer ignition delay.

It can be provided that the effective compression ratio is changed by the closing time of at least one inlet valve. Delaying the intake closure or very early intake closure can reduce the effective compression ratio, thereby reducing the exhaust gas recirculation rate required for low NOx rates and soot emissions. Both the time of opening and closing of the inlet or only the time of closing the inlet can be shifted. In a further embodiment of the invention it is provided that the change from the first to the second operating range or from the second to the first operating range is initiated by reducing or increasing the exhaust gas recirculation rate. As an alternative to this, it is also possible for the transition from the first to the second engine operating range or vice versa to be initiated by reducing the internal or external exhaust gas recirculation rate and by delaying the start of injection or by increasing the exhaust gas recirculation rate and by advancing the start of injection.

It is preferably provided that the exhaust gas recirculation rate is reduced during the transition between the first and second engine operating range by controlling the opening and / or closing time of the intake valve.

The effective mean pressure is preferably in the first operating range between approximately 0 and 6 bar, particularly preferably up to 5.5 bar, in the second operating range between approximately 3.5 to 8 bar, particularly preferably between 4 and 7 bar, and in the third operating range at least approximately 5 , 5 bar, particularly preferably at least about 6 bar.

To carry out the method, a direct-injection diesel internal combustion engine with at least one cylinder for a reciprocating piston is required, in which the start of fuel injection is at least between 50 ° crank angle before top dead center and 20 ° after top dead center, preferably up to 50 ° after top dead center, and the exhaust gas recirculation rate is variable between about 0% to 70%. It is further provided that the fuel injection pressure can be varied at least between a first and a second pressure level, the first pressure level preferably covering a range up to approximately 1000 bar and the second pressure level covering a range of at least 1000 bar, and a device for changing the Swirl levels can be provided.

Furthermore, it is advantageous if the inlet opening time and inlet closing time can be varied. To achieve this, it is advantageous if the timing of the inlet valve or the outlet valve can be shifted by means of a phase shifter device. It is very advantageous if at least one inlet valve can be activated in the outlet phase. Additionally or alternatively, it can be provided that at least one exhaust valve can be activated in the intake phase.

A double needle nozzle with first and second injection openings, which can be controlled separately, is best suited for carrying out the injection. In order to achieve different flows in the first and second operating range, it can be provided that the first injection openings have a smaller flow cross-section than the second injection openings.

Since different combustion strategies are used in the first and second operating range, it is advantageous if the central axes of the first injection openings are arranged along a first conical surface and the central axes of the second injection openings are arranged along a second conical surface, wherein the cone opening angle of the first conical surface can be smaller than that Cone opening angle of the second surface of the cone.

In a particularly preferred embodiment of the invention it is provided that the first and the second nozzle needle are arranged coaxially, the first nozzle needle preferably being guided in the second nozzle needle designed as a hollow needle. As an alternative to this, it is also possible for the first and the second nozzle needle to be arranged parallel to one another in a nozzle holder.

Double needle nozzles with coaxial or parallel nozzle needles are known from DE 100 40 738 AI.

The invention is explained in more detail below with reference to the figures. Show it

1 shows the cylinder of a conventional piston injection pump in an oblique view,

2 shows a development of the pressure piston and cylinder of this conventional piston injection pump,

3 shows a cylinder of a piston injection pump according to the invention in a first embodiment in an oblique view,

4 pressure piston and cylinder of this piston injection pump in a development,

5 shows the cylinder of this piston injection pump in a cross section,

6 this cylinder in a longitudinally cut oblique view,

7 pressure piston and cylinder of a piston injection pump according to the invention in a second embodiment in a development,

8 the cylinder of this embodiment variant in a cross section, 9 this cylinder in a longitudinally cut oblique view,

10 shows a cylinder of a piston injection pump according to the invention in a third embodiment variant in a cross section,

11 this cylinder in a longitudinal section in an oblique view,

12 shows a cylinder of a piston injection pump according to the invention in a fourth embodiment variant in a longitudinally sectioned oblique view,

13 shows an internal combustion engine for carrying out the method according to the invention, in a schematic view,

14 is a diagram in which the local air ratio λ is plotted against the local temperature T _L ,

15 is a load speed diagram,

16 to 19 valve lift diagrams with different variable control times,

20 is an injection time EGR rate load diagram;

21 shows a measurement diagram for the first operating range A,

22 shows a measurement diagram for the second operating range B,

23 shows a double needle nozzle with a coaxial nozzle needle and

24 shows a double needle nozzle with nozzle needles arranged in parallel next to one another.

1 and 2 show the cylinder 1 and the pressure piston 2 of a conventional piston injection pump. The cylinder 1 has an essentially smooth cylinder surface 3, with two pilot bores 4 opening radially into the cylinder 1. The pressure piston 2 has a first control edge 7 arranged essentially in a normal plane 5 on the cylinder axis 6 and in the lateral surface 2a a helical second control edge 8 which is inclined to the normal plane 5, and a zero-conveying groove 9 running in the axial direction. Cylinder 1 and pressure piston 2 have two second control edges 8 and control bores 4 offset by 180 ° in order to create symmetrical conditions. When the control bores 4 are covered by the control surface 10 of the piston 2, fuel is delivered. Through the oblique second control edge 8, the injection quantity can be controlled by rotating the pressure piston 2. Through a lateral deflection of the pressure piston 2 in the direction normal to the pilot bores 4, the resulting pressure distribution in the sealing gap can result in a resulting side force on the pressure piston 2, which acts in the direction of the deflection and thus increases it, which increases the wear and, in extreme cases, leads to one Rubbing the pressure piston 2 on the cylinder 1 can result.

Cylinders 1, 101, 201 and pressure pistons 2, 102, 202 are indicated in FIGS. 2, 4 and 7 by different hatching directions.

In order to avoid an asymmetrical pressure distribution in the event of a lateral deflection of the pressure piston 102, 202, pressure compensation channels 150, 250, 350, 450 are in the cylinder in the embodiment variants described below

101, 201, 301, 401 are provided, which enable pressure compensation in the circumferential direction in the sealing gap between the cylinders 101, 201, 301, 401 and the pressure piston 102, 202. The pressure equalization channels 150, 250, 350, 450 advantageously extend between a pressure space 112, 212, 312, 412 spanned by the end face 111, 211 of the pressure piston 102, 202 and the cylinders 101, 201, 301, 401 and extend into one Piston guide section 113, 213, 313, 413 of the cylinder 101, 201, 301, 401, which in the region of the pilot bores 104, 204, 304, 404 and / or on the side of the pilot bores 104 facing away from the pressure chamber 112, 212, 312, 412, 204, 304, 404. The edge of the end face 111, 211 forms a first control edge 107, 207. In the axial direction, the outer surface 102a, 202a of the pressure piston 102, 202 has a zero conveying groove 109, 209. The piston guide section 113, 213, 313, 413 is covered in at least one conveying position of the pressure piston 102, 202 by an annular space 114, 214, which borders on the second control edge 108, 208 of the pressure piston 102, 202 and between the pressure piston 102, 202 and the cylinder 101, 201, 301, 401 extends.

The pressure equalization channels 150, 250, 350, 450 can be designed as open or as closed channels. It is essential that the pressure compensation channels 150, 250, 350, 450 are spaced apart from each control bore 104, 204, 304, 404, around the control bores 104, 204, 304, 404 from the pressure piston

102, 202 paintable sealing surfaces 118, 218, 318, 418.

In the first exemplary embodiment shown in FIGS. 3 to 6, the pressure compensation channels 150 are formed by narrow, axial longitudinal grooves 151 in the surface 103 of the cylinder 101 on both sides of the control bore 104. The longitudinal grooves 151 may be formed in the surface 103 of the cylinder 101 by milling or erosion machining. Reference numeral 120 is the pressure equalization between the individual ones via the pressure chamber 112 Longitudinal grooves 151 indicated. As can be seen from FIG. 5, the longitudinal grooves 151 are arranged point-symmetrically with respect to the cylinder axis 106 and symmetrically with respect to a longitudinal plane 115 spanned by the axes 104a of the control bore 104 and the cylinder axis 106.

7 to 9 show a second exemplary embodiment of a cylinder 201 for a reciprocating pressure piston 202 of a piston injection pump, the pressure compensation channels 250 being formed by wide longitudinal grooves 251 between the compensation bores 204. The two longitudinal grooves 251 are arranged point-symmetrically with respect to the cylinder axis 206, but also symmetrically with respect to a longitudinal plane 215 spanned by the axes 204a of the pilot bores 204 and the cylinder axis 206, and a plane 216 normal thereto, through the cylinder axis 206.

10 and 11 show a third exemplary embodiment of a cylinder 301 for receiving a pressure piston, not shown, corresponding to FIGS. 2, 4 and 7, for a piston injection pump, in which the pressure compensation channels 350 are arranged one above the other at an axial distance horizontal transverse bores 351, 352 are formed, which are fluidly connected to one another by a longitudinal bore 353 running approximately parallel to the cylinder axis 306. The cross bores 351, 352 and longitudinal bores 353 are closed to the outside after manufacture. As can be seen from FIG. 10, the transverse bores 351, 352 are arranged point-symmetrically with respect to the cylinder axis 307, as well as symmetrically to the longitudinal plane 315 spanned by the axis 304a of the control bores 304 and the cylinder axis 306 and a plane 316 normal to this. The lower transverse bores 350 lie in the piston guide section 313 approximately in the region of the control bores 304.

FIG. 12 shows a fourth exemplary embodiment, which differs from the exemplary embodiment illustrated in FIGS. 10 and 11 in that several lower transverse bores 451 are arranged one above the other in a row. The lower transverse bores 451 are connected via the longitudinal bore 453 to an upper transverse bore 452, which opens into the pressure chamber 512 located above the pressure piston. The lower transverse bores 451 also open here in the piston guide section 413 in the region of the control bore 404 in the cylinder 401.

The transverse bores 352, 452 ensure pressure equalization with the upper pressure chamber 312, 412. The lower transverse bores 351, 451 serve to reduce the sealing surface 318, 418. 13 shows an internal combustion engine 1001 with an intake manifold 1002 and an exhaust manifold 1003. The internal combustion engine 1001 is charged via an exhaust gas turbocharger 1004, which has an exhaust gas-driven turbine 1005 and a compressor 1006 driven by the turbine 1005. A charge air cooler 1007 is arranged upstream of the compressor 1006 on the inlet side.

Furthermore, there is a high-pressure exhaust gas recirculation system 1008 with a first exhaust gas recirculation line 1009 between the exhaust line 1010 and the inlet line

1011 provided. The exhaust gas recirculation system 1008 has an exhaust gas recirculation cooler 1012 and an exhaust gas recirculation valve 1013. Depending on the pressure difference between the outlet line 1010 and the inlet line 1011, an exhaust gas pump 1014 can also be provided in the first exhaust gas recirculation line 1009 in order to control or increase the exhaust gas recirculation rate.

In addition to this high-pressure exhaust gas recirculation system 1008, a low-pressure exhaust gas recirculation system 1015 is provided downstream of the turbine 1005 and upstream of the compressor 1006, a second exhaust gas recirculation line 1018 branching off in the exhaust line 1016 downstream of a particle filter 1017 and opening into the intake line 1019 upstream of the compressor 1006. An exhaust gas recirculation cooler 1020 and an exhaust gas recirculation valve 1021 are also arranged in the second exhaust gas recirculation line 1018. To control the exhaust gas recirculation rate, an exhaust valve 1022 is arranged in the exhaust line 1016 downstream of the branch.

An upstream of the branch of the first exhaust gas recirculation line 1009, an oxidation catalytic converter 1023 is arranged in the exhaust line 1010, 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 1005. In principle, the oxidation catalytic converter 1023 can also be arranged downstream of the branch of the exhaust gas recirculation line. The arrangement shown in FIG. 13 with the branch downstream of the oxidation catalytic converter 1023 has the advantage that the exhaust gas cooler 1012 is exposed to less pollution, but the disadvantage that, due to the higher exhaust gas temperatures, the exhaust gas recirculation cooler has a higher cooling capacity

1012 becomes necessary.

For each cylinder 1024, the internal combustion engine 1001 has at least one injection valve 1025 that directly injects diesel fuel into the combustion chamber, which is able to carry out several injections per working cycle and start their respective injection in a range between 50 ° crank angle CA before top dead center TDC to 50 ° crank angle CA after top dead center TDC can be changed. The maximum injection pressure should be at least 1000 bar.

The shape of the combustion chamber and the fuel injection configuration must be designed for conventional full-load diesel combustion.

14 shows a diagram in which the local air ratio λ is plotted against the local combustion temperature T _L. Heavy soot formation occurs in the area labeled SOOT, and NOx denotes the area of high nitrogen oxide formation. A, B, C are the first, second and third operating areas of the method described here.

The first operating area A is assigned to the lower to middle part-load area L _L , the second engine operating area B to the middle to upper part-load area L _M and the third engine operating area C to the high-load and full-load area L _H , as shown in FIG. Speed n diagram can be seen.

In the first operating area A, which is also referred to as the HCLI area (Homogeneous Charge Late Injection), the start of injection is relatively early in the compression stroke, i.e. at about 50 ° to 5 ° crank angle CA before top dead center TDC after the compression stroke, which means that a long ignition delay to form a partially homogeneous mixture for premixed combustion is available. Due to the pronounced premixing and dilution, extremely low soot and NOx emission values can be achieved. As can be seen from FIG. 14, the first operating range A is clearly above the limit for the local air ratio λ _LS which is decisive for the formation of soot. A high exhaust gas recirculation rate EGR of between 50% and 70% ensures that the local combustion temperature T _L always remains below the minimum nitrogen oxide formation temperature T _N0 . The injection takes place at a pressure between 400 and 1000 bar. The long ignition delay causes the combustion phase to be shifted to the most efficient position around top dead center TDC. The center of combustion is in a range between -10 ° to 10 ° crank angle CA after the top dead center TDC, whereby a high efficiency can be achieved. The high exhaust gas recirculation rate EGR, which is required for the first operating range A, can be achieved either by external exhaust gas recirculation alone, or by combining external with internal exhaust gas recirculation through variable valve control.

In the second operating area B, the internal combustion engine is operated according to the so-called HPLI process (Highly Premixed Late Injection). The main part of the injection phase is after TDC. In the second loading Drive range B, the internal combustion engine is operated with an exhaust gas recirculation rate between 20% and 40%, the start of injection being in a range between 2 ° crank angle CA before top dead center and 20 ° crank angle CA after top dead center. By completely separating the end of the injection and the start of the combustion, a partial homogenization of the mixture with premixed combustion is achieved. Because of the decreasing temperature level compared to conventional injection before top dead center and the increased amount of recirculated exhaust gases compared to conventional operation, the ignition delay is longer. To extend the ignition delay, other means, such as a lowering of the effective compression ratio ε and / or the inlet temperature, as well as a shortening of the injection period, an increase in the injection pressure and / or an increase in the cross-sectional area of the injection nozzle can be used. The short injection period is necessary so that the end of injection is before the start of combustion. In this case, soot emissions can be kept at a very low level. This can be explained by the fact that the simultaneous occurrence of liquid fuel in the fuel jet and the flame which conventionally envelops the jet is avoided, as a result of which the oxidation reactions in the vicinity of the jet, which otherwise lead to soot formation and occur due to lack of air, are also prevented. The late position of the injection time, together with the relatively long ignition delay, leads to a late shift in the entire combustion process, as a result of which the cylinder pressure curve is also shifted late and the maximum temperature is lowered, which leads to low NOx emissions.

The late shift in the combustion process causes a lowering of the maximum temperature, but at the same time leads to an increase in the temperature at a given later crank angle CA, which in turn increases the soot burnup.

The shift of the combustion in the expansion stroke, in addition to the high exhaust gas recirculation rate EGR, leads to a pressure increase rate in cylinders which does not exceed the permissible level, despite the larger amount of premixed fuel due to the long ignition delay and consequently higher maximum combustion rate. The high maximum burning rate, which leads to a high degree of uniformity, is able to partially compensate for the loss of efficiency due to late shifting of the combustion phase. In order to achieve a high degree of efficiency, the center of combustion should be as close as possible to the top dead center TDC.

The advantage of the HPLI process used in the second operating area B is that very low NOx and particle emissions occur and that a high exhaust gas temperature can be reached, which is necessary for the regeneration of a part. kelfilters is an advantage. As can be seen from FIG. 14, the local combustion temperature T _L in the second operating area B is to a small extent above the lower NOx formation _temperature T _N0x . The local air ratio λ is largely above the soot formation limit λ s Soot is formed at the beginning of the combustion process, but due to the strong turbulence resulting from the high-pressure injection and high temperatures, the soot is oxidized towards the end of the combustion process, resulting in very low soot emissions overall.

In the third operating area C, the internal combustion engine is operated conventionally with exhaust gas recirculation rates EGR between 0% and 30%, multiple injections being possible. This enables premixed and diffusion combustion to be carried out. A combination of external and internal exhaust gas recirculation can also be used for exhaust gas recirculation.

For comparison purposes, the operating range D is shown in dotted lines in FIG. 14. This operating range D is operated, for example, in US Pat. No. 6,338,245 B1 in the medium to high partial load range. However, this has the disadvantage that the efficiency is poor as a result of low temperatures. In the present method according to the invention, this area can generally be avoided.

Options can also be generated in the combustion chamber in the first, second and / or third operating range A, B, C. The swirl formation is advantageous in order to further reduce the soot formation. Swirl and high efficiency must be coordinated.

It is particularly advantageous if the valve timing in the internal combustion engine 1001 can be variably adjusted. As a result, the EGR rate between the operating ranges A, B, C can be set precisely and quickly in the event of load changes. The combination of external and internal exhaust gas recirculation enables particularly fast and precise control of the EGR exhaust gas recirculation rate. Finally, the effective compression ratio ε can also be regulated by means of variable valve control, as a result of which lower nitrogen oxide and soot emissions can be achieved with a reduced exhaust gas recirculation rate EGR.

16 shows a valve lift diagram, in which the valve lifts l of at least one exhaust valve A and at least one intake valve E are plotted against the crank angle CA. By shifting the intake valve curve E late, for example with a phase shifter, the effective compression ratio ε and the required EGR rate can be reduced. This can be done in all three operating areas A, B and C. With E ₀ or E _c are the opening and closing times of the intake valve. E ₀ s and E _oc indicate the opening and closing times of the shifted intake valve lifting curve E _s .

Alternatively, only the closing flank of the intake valve lift curve E can be changed, as a result of which the closing time is sooner or later, as indicated by the lines E _s ' and E _s "in FIG. 17. Essentially the same effect can be achieved achieve, as by shifting the entire valve lift curve (Fig. 16).

Internal exhaust gas recirculation can be accomplished by reopening the exhaust valve during the intake stroke, as line A 'in Fig. 18 shows, or by reopening the intake valve during the exhaust stroke, see line E' in Fig. 19. This allows quick control of the Exhaust gas recirculation rate EGR can be effected in all operating areas A, B, C. It is possible to carry out the transition between the second operating area B with a 20% to 40% EGR rate to the first operating area A with a 50% to 70% EGR rate only by internal exhaust gas recirculation and by advancing the start αi of fuel injection I. The opposite change from the first operating area A to the second operating area B is also possible.

If no variable valve control is used, the transition between the first and the second operating range A, B can be carried out by reducing the external exhaust gas recirculation rate EGR and at the same time advancing the start α _{∑ of} the fuel injection I, as can be seen from FIG. 20. By simultaneously reducing the EGR rate and advancing the start of fuel injection I, misfires can be avoided. Conversely, a transition between the second operating area B and the first operating area A can be carried out by simultaneously increasing the internal exhaust gas recirculation rate EGR and by bringing the start αi of the injection I forward.

FIG. 21 shows a measurement diagram of an exemplary embodiment for the first operating range A, with injection I, heat release rate Q, cumulative heat release rate ΣQ and cylinder pressure p being plotted against the crank angle CA. 22 shows an analog measurement diagram for the second operating range B. Thin and thick lines represent different parameter configurations. The relatively long ignition delay between injection I and combustion can be clearly seen.

23 and 24 show injection valves 1025 with double needle nozzles 1100, 1200 having nozzle holders 1110, 1210. The double needle nozzle 1100 from FIG. 23 has a first nozzle needle 1101, which is arranged displaceably in a hollow second nozzle needle 1102. The first nozzle needle 1101 controls first injection openings 1103, and the second nozzle needle 1102 controls second injection openings 1104, which are arranged in the nozzle tip 1106. The sum of the diameters di of the first injection openings 1103 is smaller than the sum of the diameters d _{2 of} the second injection openings 1104. The central axes 1103a, 1104a of the first injection openings 1103 and the second injection openings 1104 are each arranged on a conical surface 1107, 1108, the cone opening angle of which is designated by αi and α ₂ . Here, the cone angle is .alpha..sub.i the first injection ports 1 103 α smaller slightly than the cone angle ₂ of the second injection ports 1104th

The injection valve 1025 shown in FIG. 24 has a double needle nozzle 1200 with a first nozzle needle 1201 and a second nozzle needle 1202, wherein both nozzle needles 1201, 1202 are arranged side by side in parallel. The nozzle axes 1201 'and 1202' are spaced apart. The first nozzle needle 1201 controls first injection openings 1203 and the second nozzle needle 1202 controls second injection openings 1204, which are each arranged in a nozzle tip 1206a, 1206b. The first and second injection openings 1203, 1204 are arranged along a conical surface 1207, 1208, the cone opening angle of which is denoted by αi and α _2, respectively. The diameters of the first and second injection openings 1203, 1204 are denoted by di and by d ₂ . The sum of the passage cross sections of the first injection openings 1203 is smaller than the sum of the passage cross sections of the second injection openings 1204. The openings of the injection openings 1203 into the combustion chamber of the first nozzle tip 1206a and the openings of the injection openings 1204 of the second nozzle tip 1206b are in normal planes 1209a, spaced apart from one another. 1209b arranged on the nozzle needle axes 1201 ', 1202'. The distance between the normal planes 1209a, 1209b is denoted by a. This distance has the effect that the jets of the first and second injection openings 1203, 1204 do not interfere with one another at full load, that is to say do not strike one another. Both nozzle tips 1206a, 1206b are advantageously designed with the same number of holes, preferably three.

The first and second nozzle needles 1101, 1102, 1201, 1202 can be controlled separately in a known manner, as described for example in DE 100 40 738 AI. In the first engine operating area A, the first nozzle needle 1101, 1201 is actuated and the first injection openings 1103, 1203 are opened, while the second injection openings 1104, 1204 remain closed. In the second operating area B, the second nozzle needle 1102, 1202 actuated, whereby the second injection openings 1104, 1204 are opened, while in turn the first injection openings 1103, 1203 remain closed. Because separate injection openings 1103, 1104, 1203, 1204 are used in operating areas A, B, the injection characteristics for HCLI operation in first operating area A and HPLI operation in second engine operating area B can be optimally implemented. In the third operating area C, both nozzle needles 1101, 1102; 1201, 1202 actuated, whereby the injection through all injection openings 1103, 1104; 1203, 1204.

The described method allows the internal combustion engine to be operated with high efficiency and low NOx and soot emissions in both the first, second and third operating ranges A, B, C.

The patent claims submitted with the application are proposals for formulation without prejudice for the achievement of further patent protection. The applicant reserves the right to claim further features previously only disclosed in the description and / or drawings.

References used in the subclaims indicate the further development of the subject matter of the skin claim through the features of the respective subclaim; they are not to be understood as a waiver of the achievement of independent, objective protection for the characteristics of the related subclaims.

However, the subjects of these subclaims also form independent inventions which have a design which is independent of the subjects of the preceding subclaims.

The invention is also not restricted to the exemplary embodiment (s) of the description. Rather, numerous changes and modifications are possible within the scope of the invention, in particular such variants, elements and combinations and / or materials, for example by combining or modifying individual ones in conjunction with the embodiments described in the general description and the claims and in the drawings Features or elements or process steps contained therein are inventive and, by means of combinable features, lead to a new object or to new process steps or process step sequences, also insofar as they relate to manufacturing, testing and working processes.

Claims

PATENT CLAIMS

1. Internal combustion engine, in particular a diesel internal combustion engine, with at least one piston injection pump for fuel delivery with a cylinder for receiving a reciprocating pressure piston, the end face of which borders a pressure chamber, the edge of the end face forming a circumferential first control edge and the cylinder having at least one has radial control bore in a piston guide section of the cylinder, which can be run over at least by a second control edge of the pressure piston formed by a shape in the outer surface of the pressure piston, characterized in that the cylinder has at least two, preferably diametrically arranged, pressure compensation channels for pressure compensation between two different circumferential areas - Chen of the piston guide portion of the cylinder.

2. Internal combustion engine, in particular according to claim 1, characterized in that the pressure compensation channels at least in a conveying position of the pressure piston fluidly connect the pressure chamber with a region of the piston guide section of the cylinder which is approximately at the level of the control bore.

3. Internal combustion engine, in particular according to claim 1 or 2, characterized in that at least one pressure compensation channel is formed by at least one recess in the cylinder, preferably by a longitudinal groove running parallel to the cylinder axis.

4. Internal combustion engine, in particular according to one of claims 1 to 3, characterized in that at least one pressure equalization channel has at least two transverse bores opening at an axial distance approximately radially into the cylinder, which are fluidically connected to one another by at least one longitudinal bore which runs approximately parallel to the cylinder axis are.

5. Internal combustion engine, in particular according to one of claims 1 to 4, characterized in that at least two radial transverse bores in the axial direction of the cylinder are arranged directly one above the other in the piston guide section, preferably at the level of the control bore.

6. Internal combustion engine, in particular according to one of claims 1 to 5, characterized in that at least two pressure compensation channels are arranged point-symmetrically with respect to the cylinder axis.

7. Internal combustion engine, in particular according to one of claims 1 to 6, characterized in that at least two pressure compensation channels in the circumferential direction are arranged on both sides and preferably symmetrically to the control bore.

8. Internal combustion engine, in particular according to one of claims 1 to 7, characterized in that the pressure compensation channels in the region of the piston guide section of the cylinder are spaced from the control bore, so that a sealing surface which can be painted over by the pressure piston is formed around the control bore of the cylinder.

9. Method for operating an internal combustion engine, in particular a direct injection engine, in particular a diesel internal combustion engine, with the following steps:

Operating the internal combustion engine in a second operating range associated with the medium partial load with low-temperature mixture combustion and even later injection than in the first operating range, the fuel injection in a range between 2 ° crank angle before top dead center and about 20 ° crank angle after top dead center of the compression phase is started

wherein the fuel is injected into the combustion chamber in the first operating region via first injection openings and in the second operating region at least via second injection openings of an injection valve, which is preferably designed as a double needle nozzle.

10. The method, in particular according to claim 9, characterized in that in the first operating area the fuel is injected at a lower flow rate than in the second operating area.

11. The method, in particular according to claim 9 or 10, characterized in that the fuel is injected in the first and in the second operating area in fuel jets arranged along a conical surface, the fuel being injected in the first operating area with a different, preferably smaller, cone opening angle than in the second operating area ,

12. The method, in particular according to one of claims 9 to 11, characterized in that in a third operating range assigned to the upper partial load and the full load, the main part of the fuel injection takes place in a range between 10 ° before and 10 ° crank angle after the top dead center.

13. The method, in particular according to one of claims 9 to 12, characterized in that in the third operating range the fuel is injected through the first and / or through the second injection openings.

14. The device, in particular for performing the method according to one of claims 9 to 13, with an injection valve for direct fuel injection into the combustion chamber, characterized in that the injection valve is designed as a double needle nozzle which has first and second injection openings, wherein first and second injection openings can be controlled separately.

15. The device, in particular according to claim 14, characterized in that the first injection openings have a smaller flow cross-section than the second injection openings.

16. The device, in particular according to claim 14 or 15, characterized in that the central axes of the first injection openings are arranged along a first conical surface and / or the central axes of the second injection openings are arranged along a second conical surface, preferably the cone opening angle of the first conical surface being smaller than that Cone opening angle of the second surface of the cone.

17. The device, in particular according to one of claims 14 to 16, characterized in that the first and the second nozzle needle are arranged coaxially, the first nozzle needle preferably being guided in the second nozzle needle designed as a hollow needle.

18. The device, in particular according to one of claims 14 to 16, characterized in that the first and the second nozzle needle are arranged parallel next to one another in a nozzle holder.