DE102007053890B4 - Internal combustion engine with auto-ignition - Google Patents

Abstract

An internal combustion engine (101) with auto-ignition having an engine block (102) with at least one cylinder (306, 402) and a piston associated therewith, an injection system (104) for direct fuel injection into the cylinder (306, 402) and an intake device (103, 310). for charge air with at least one inlet channel (404) and an inlet valve associated therewith and exhaust gas outlet means (105) having at least one outlet channel and an outlet valve associated therewith and a downstream exhaust system (106) with particulate filter (107) the cylinder (306, 402) a cylinder head and the piston has a piston head (201) and the cylinder head and piston head (201) inner surfaces of a combustion chamber and in the surface of the piston head (202) a piston recess (206) is arranged with a trough edge (208), wherein at one on a circle Normalized bowl rim (208) the ratio of bowl rim diameter (203) to piston diameter (204) greater than 0.55 i or with a 90% filled piston recess (206) with a filling surface normalized to a circle, the ratio of the filling surface to the piston cross-sectional area> 0.30 is an inlet channel (404) designed as a swirl channel (301) with an associated swirl channel valve, the inlet flow in the combustion chamber receives a swirl, characterized in that for at least one selected cylinder (306, 402) the intake flow at at least one valve lift Vh between 2 mm and Vhh = 0.5Vhm a swirl increase with a Ventilhubabhängigen Tippelmann swirl number D a value> 0, 3, where Vhm denotes the maximum valve lift or, in particular, a value Vhm = 8mm, the intake flow in a valve lift range between Vh = 2 mm and Vhh = 0.5Vhm a swirl increase with a normalized integral Tippelmann swirl Di numbering> 0.40 / mm or the inlet flow has a swirl lift in a valve lift range between 2 mm and Vhh = 0.5 Vhm, in a measurement of the valve lift-dependent Tippelmann swirl number D at intervals of 0.5 mm, the mean Tippelmann swirl number Dm has a value ≥ 0.4 for at least one valve lift interval DeltaV = 1 mm.

Description

The present invention relates to an internal combustion engine with auto-ignition according to the preamble of the independent claim.

To meet today's and future stringent emission limits for diesel passenger cars, the use of various technologies is known, which include a vote of relatively many internal and external engine measures to reduce emissions. Known engine measures to reduce the raw emissions related to the combustion process (combustion design, channel shape, swirl of the inlet flow ), Adjustment of the lambda value, injection timing and exhaust gas recirculation, injection system (adaptation of the injection curve), exhaust gas recirculation, charging, intercooling, water injection. Non-exhaust aftertreatment measures include the use of particulate filters, oxidation catalysts, and NOx catalysts.

It is also known that in diesel engines, measures that reduce soot emissions tend to increase the levels of knit oxide emissions. While it was still possible to comply with the Euro4 exhaust limits to tolerate relatively high levels of soot emissions at relatively low NOx emissions, since the latter could be absorbed by arranged in the exhaust system particulate filter, this is no longer possible with the stricter emission limits according to Euro5, as the Particle filter after too short a time already have a high load, which has a frequent regeneration with correspondingly high fuel consumption and increased carbon dioxide emissions result.

From the DE 3612029 A1 is already an air-compressing reciprocating internal combustion engine with an approximately centrally arranged in the piston arc rotationally symmetric combustion bowl, an injection device having a injecting into the combustion bowl injection nozzle with at least four spray holes and with a rotational movement of the entering into the combustion bowl charge air inlet device known. The diameter of the combustion bowl amounts to more than 55% of the cylinder diameter. The rotational movement of the charge air generated by the inlet device has a swirl number determined by the impeller method of about 1.5 to 2.5, wherein the injection device for a delivery rate of 5 to 15 mm ³ / degree crank angle and in the pressure chamber of the injector prevailing injection pressures from 500 to 2000 bar is designed. The aim of the specified measures is to achieve an increase in power output, a reduction in raw emissions and emissions of gaseous exhaust gas components, a reduction in noise and a relief of the mechanically highly stressed machine parts. However, with these known measures it is not possible to achieve the limit values of the Euro5 emission standard for nitrogen oxides and particulates (PM) for diesel passenger cars.

From the DE 103 48 366 B4 A method for operating a direct-injection diesel internal combustion engine is already known in which exhaust gas is recirculated in an operating region b, the exhaust gas recirculation rate being about 20% to 40% and the fuel injection being carried out at an injection pressure of at least 1000 bar. And further, the fuel injection in an operating range a at an injection pressure of between 400 - 1000 bar takes place.

From the DE 691 02 839 T2 A direct-injection diesel engine is already known in which a direct injection diesel engine is proposed for reducing the amount of smoke having a swirl control valve that controls the opening of an intake port and a control means that adjusts the opening of the swirl control valve in response to the injection pressure of a fuel injector ,

From the DE 196 21 635 B4 A diesel engine is already known, with two intake ports, two exhaust ports, four valves, one injector and one piston per cylinder, the axes of the valves being substantially parallel to the axis of the cylinder, the injector substantially parallel to and parallel to the cylinder axis Cross-sectional area of the cylinder is arranged substantially centrally and the one inlet channel is designed as a swirl channel and the other inlet channel as a filling channel. The piston in this case has a combustion bowl with recessed bowl edge, wherein the injection nozzle is formed with an injection angle of 140 ° to 155 °, so that the spray jets of the injection nozzle impinge tangentially on the inside of the bowl rim.

From the DE 100 48 238 A1 a method for operating a diesel internal combustion engine is known, wherein with directly injected fuel and separately supplied combustion air in the combustion chamber at the latest at the time of auto-ignition, a homogeneous mixture is formed. In this case, it is provided that the combustion air is admitted into the combustion chamber in a swirling flow that can be adjusted by means of an actuator with respect to the swirl intensity about a longitudinal axis of the cylinder. A control unit adjusts the swirl intensity in coordination with one another in the injection parameters for the process of metering the fuel.

From the DE 698 33 526 T2 For example, there is known a diesel engine having a swirling device configured to generates a swirling flow in a combustion chamber of the engine and includes a swirl intensity compensating device configured to compensate for a swirling intensity of the controlled swirling flow depending on the engine temperature.

The object of the present invention is therefore to achieve a reduction of the raw emissions of nitrogen oxides and soot particles in an internal combustion engine with auto-ignition, so that they are as close as possible to those specified by the Euro standard 5 In the case of a motor vehicle with a particulate filter arranged in the exhaust system, the said EU5 limit values of the end-pipe emissions can be reached or undershot.

The object is achieved with the features of the independent claim.

• Einlasseinrichtung für Ladeluft mit zumindest einem Einlasskanal und einem zugeordneten Einlassventil und eine
• Auslasseinrichtung für Abgas mit zumindest einem Auslasskanal und einem diesem zugeordneten Auslassventil und einem anschließendem Abgassystem mit Partikelfilter versehen,
• wobei der Zylinder einen Zylinderkopf und der Kolben einen Kolbenkopf aufweist und Zylinderkopf und Kolbenkopf Innenflächen eines Brennraumes bilden
• und in der Oberfläche des Kolbenkopfs eine Kolbenmulde mit einem Muldenrand angeordnet ist, wobei bei einem auf einen Kreis normierten Muldenrand das Verhältnis von Muldenranddurchmesser zu Kolbendurchmesser größer 0,55 ist oder bei einer zu 90% gefüllten Kolbenmulde bei einer auf einen Kreis normierten Füllfläche das Verhältnis der Füllfläche zur Kolbenquerschnittsfläche > 0,30 ist
• ein Einlasskanal als Drallkanal mit einem zugeordneten Drallkanalventil ausgebildet ist, wobei die Einlassströmung in den Brennraum einen Drall erhält.

The internal combustion engine according to the invention with auto-ignition with an engine block having at least one cylinder and a piston associated therewith, an injection system for fuel injection into the cylinder comprises a

• Inlet device for charge air with at least one inlet channel and an associated inlet valve and a
• Exhaust gas outlet device with at least one outlet channel and an outlet valve associated therewith and a downstream exhaust system with particle filter,
• wherein the cylinder has a cylinder head and the piston has a piston head and the cylinder head and piston head form inner surfaces of a combustion chamber
• and in the surface of the piston head, a piston recess is arranged with a trough edge, wherein in a normalized to a circle trough edge, the ratio of bowl rim diameter to piston diameter is greater than 0.55 or in a 90% filled piston bowl with a normalized to a circle filling surface, the ratio the filling surface to the piston cross-sectional area> 0.30
• an intake passage is formed as a swirl passage with an associated swirl passage valve, wherein the intake flow into the combustion chamber receives a swirl.

• die Einlassströmung bei zumindest einem Ventilhub Vh zwischen 2 mm und Vhh = 0.5Vhm eine Drallanhebung mit einer ventilhubabhängigen Tippelmann-Drallzahl D mit einen Wert > 0,3 aufweist, wobei Vhm den maximalen Ventilhub, insbesondere mit einem Wert Vhm=8mm bezeichnet,
• die Einlassströmung in einem Ventilhubbereich zwischen Vh = 2 mm und Vhh = 0.5 Vhm eine Drallanhebung mit einer normierten integralen Tippelmann-Drallzahl Di mit einen Wert > 0,40/mm aufweist oder
• die Einlassströmung in einem Ventilhubbereich zwischen Vh= 2 mm und Vhh=0,5 Vhm eine Drallanhebung aufweist, derart dass bei einer Vermessung der ventilhubabhängigen Tippelmann-Drallzahl D in Abständen von 0,5 mm die gemittelte Tippelmann - Drallzahl Dm für mindestens ein Ventilhubintervall DeltaV= 1mm einen Wert > 0,4 aufweist.

The internal combustion engine is designed to operate for at least one selected cylinder at single port opening

• the inlet flow at at least one valve lift Vh between 2 mm and Vhh = 0.5Vhm has a swirl increase with a valve lift-dependent Tippelmann swirl number D having a value> 0.3, where Vhm denotes the maximum valve lift, in particular with a value Vhm = 8 mm,
• the inlet flow in a valve lift range between Vh = 2 mm and Vhh = 0.5 Vhm has a swirl increase with a normalized integral Tippelmann swirl number Di with a value> 0.40 / mm or
• the inlet flow in a Ventilhubbereich between Vh = 2 mm and Vhh = 0.5 Vhm has a swirl increase, so that in a measurement of valve lift-dependent Tippelmann swirl number D at intervals of 0.5 mm, the average Tippelmann - Dwell number Dm for at least one valve lift interval DeltaV = 1mm has a value> 0.4.

In the case of a single-channel opening, a supply of charge air into the combustion chamber takes place only via the swirl duct.

With the internal combustion engine according to the invention, it is possible at operating point 2000 l / min and an effective mean pressure of 8 bar to achieve a NOx emission of <1.55 g / kWh with a soot emission of <0.08 g / kWh. Preferably, the internal combustion engine according to the invention can also achieve emissions of <1.5, 1.43, 1.38 and 1.3 g / kWh for soot emissions of <0.075, 0.071, 0.065 or 0.006 g / kWh in the operating point mentioned above.

A motor vehicle equipped with the internal combustion engine according to the invention, when operated in the new European driving cycle, has NOx raw emissions with a value of <300 mg / km and raw particle emissions with a value of <35 mg / km. Preferably, such a motor vehicle NOx emissions with a value <280 mg / km, 235 mg / km, 200 mg / km, 180 mg / km or 160 mg / km and at the same time raw particle emissions, which have a value of 40 mg / km , 45 mg / km, 50 mg / km, 55 mg / km or 60 mg / km.

Such a vehicle reaches or falls below the exhaust emission limits of the new European standard EU5 when using a particulate filter.

In a preferred embodiment, it is provided that the inlet flow at a single-channel opening and at maximum valve lift Vhm, in particular with Vhm = 8 mm, a Ventilhubabhängige Tippelmann - spin number D with a value> 0.4 and / or the intake valve Ventilhubabhängigen flow coefficient αK> 0, 04 shows: This can be achieve improved mixture preparation with good cylinder filling at the same time.

Advantageous developments of the invention can be found in the dependent claims.

The invention will be described in more detail below independently of the summary in the claims with reference to embodiments illustrated in drawings from which further advantages and aspects of the invention.

• 1 :einen erfindungsgemäßen Verbrennungsmotor mit Selbstzündung,
• Fig, 2: einen Kolbenkopf mit Kolbenmulde,
• 3 a und 3 b: Drall- und Füllkanäle einer Einlasseinrichtung mit Drallklappe,
• 3 c: Ventilsitze mit Bearbeitung,
• 4: einen Tippelmann-Prüfstand,
• 5: eine Darstellung von ventilhubabhängigen Drall und Durchfluss im Einkanalbetrieb,
• 6: eine Darstellung von ventilhubabhängigen Drall und Durchfluss im Zweikanalbetrieb,
• 7: Rußemissionen in Abhängigkeit von Öffnungsstellungen einer Drallklappe,
• 8: Rußemissionen in Abhängigkeit von Öffnungsstellungen einer Drallklappe,
• 9: Rußemissionen in Abhängigkeit von Öffnungsstellungen einer Drallklappe,
• 10-11: Rußemissionen in Abhängigkeit von Öffnungsstellungen einer Drallklappe.

Show it

• 1 : an internal combustion engine according to the invention with auto-ignition,
• 2, a piston head with piston recess,
• 3 a and 3 b : Swirl and fill channels of an inlet device with swirl flap,
• 3 c Photos: Valve seats with machining,
• 4 : a Tippelmann test bench,
• 5 : a representation of valve lift-dependent swirl and flow in single-channel operation,
• 6 : a representation of valve lift-dependent swirl and flow in two-channel operation,
• 7 : Soot emissions depending on open positions of a swirl flap,
• 8th : Soot emissions depending on open positions of a swirl flap,
• 9 : Soot emissions depending on open positions of a swirl flap,
• 10-11 : Soot emissions depending on the opening positions of a swirl flap.

In 1 is a passenger car internal combustion engine according to the invention 101 shown with auto-ignition with which it is possible to achieve both low soot (particulate matter) and NOx raw emission values. The engine according to the invention is preferably operated with fuel according to DIN EN 590.

It is understood that in others, in the 1 not shown embodiments, certain components that are shown in Figure, missing or other components may be added without departing from the scope of the invention.

Of the 1 shown self-igniting internal combustion engine 101 has an engine block 102 , four cylinders (not provided with reference numerals), an inlet direction 103 for charge air, designed as a common rail injection system 104 for fuel injection into the cylinders as well as an outlet device 105 for exhaust on. To the outlet 105 is an exhaust system 106 with a diesel particulate filter 107 connected. Upstream of the diesel particulate filter 107 is an oxidation catalyst 107a arranged, which may be missing in other embodiments of the invention.

The diesel particulate filter 107 keeps back the soot particles produced during combustion of the diesel fuel. The particle filter preferably works 107 without the addition of additives; However, it is also the use of particulate filters that require additives conceivable. The particle filter consists of a porous ceramic component with a precious metal-containing coating. To ensure the continuity of the filter in the long term, the particles deposited in the filter must be removed by passive and / or active regeneration. With wall-flow filters today high deposition rates of 95% and more of the total particle mass can be achieved with a small increase in fuel consumption. Similarly high values are achieved with flow filters in which a more than 90% reduction in the number of particles can be achieved.

The exhaust system is equipped with an exhaust gas recirculation (EGR) system with an EGR valve 109 and EGR cooling 110 connected, the exhaust gas recirculation cooling 110 provided with a low temperature oil circuit. With the exhaust system is also a turbocharger 115 connected. Downstream of the compressor of the exhaust gas turbocharger 115 is arranged a charge air cooler.

It is understood that to meet strict emission standards, the various in 1 shown components must be coordinated.

In the following important components of an internal combustion engine according to the invention are shown in more detail. Preferably, this is designed for fast-running cars or light commercial vehicles.

In 2 is a piston head 201 of an internal combustion engine according to the invention with auto-ignition shown schematically, wherein in the surface 202 of the piston head 201 a piston recess formed by a recess 206 is arranged. The hollow 206 is formed by a trough contour forming the wall area 207 defined, located in the piston surface 202 starting from the piston edge 208 extends. As a trough edge 208 becomes the outermost part of the piston recess 206 understood, while as trough edge diameter of the cross - sectional diameter of the trough 206 at the point where the trough contour 207 a slope 1 is designated. At this point is the trough contour 207 parallel to the piston axis 209 ,

The in 2 illustrated piston head 201 is arranged in the operation of the internal combustion engine in a cylinder, not shown in the figure, in which the piston reciprocates and is connected to a crankshaft, not shown, that the piston performs a movement with a reversal at an upper and lower dead center. The cylinder has a cylinder head, which forms with the piston head inner surfaces of a combustion chamber, is injected into the fuel. The injected into the cylinder fuel forms with compressed gas, a fuel-gas mixture that is ignited by generated during operation of the internal combustion engine heat of compression.

Typically, the engine has four to six cylinders.

The piston recess 206 indicates an area of the contour 207 with less depth and a peripheral area of the contour 207 with a greater depth, in each case with respect to the piston surface, wherein the invention comprises embodiments with larger and smaller differences in depth.

The piston recess 206 can also have a non-circular piston edge 208 exhibit. In this case, the bowl rim diameter becomes 203 defined at a bowl edge nominated for a circle.

The ratio of bowl rim diameter to piston rim diameter is> 0.55, preferably> 0.58, 0.60, 0.62, 0.65, 0.70 or 0.80, resulting in a greater free injection length and improved mixture preparation.

In a further embodiment of the invention is in a 90% filled piston recess 206 in the case of a filling surface nominated for a circle, the ratio of the filling surface to the piston cross-sectional area> 0.30. Also in this embodiment of the invention is a relatively large piston recess with the corresponding advantages presented above.

In the 2 shown piston recess 206 has in its peripheral wall region on an undercut, that is, the maximum piston inner diameter is greater than the piston rim diameter 203 , In other embodiments of the invention, the piston recess 206 no undercut.

The maximum bowl internal diameter is preferably at most 15%, 10%, 8%, 6%, 4%, 3% or 1% larger than the bowl rim diameter 203 , Thus, a larger free jet length of the injected fuel can be achieved and the order of liquid fuel can be reduced to the peripheral wall area.

The cylinder or cylinders have a bore of less than 110 mm. Preferably, the bore can also be ≤ than 100 mm, 95 mm or 90 mm. The single cylinder volume is less than 0.8 l, preferably <0.7 l or 0.6 l.

Further, the cylinder is connected to an injection system for direct fuel injection into the cylinder. The injection system is designed in common-rail technology with which fuel is injected into the cylinder at up to 1800 bar. It is understood that the invention also works with other injection systems and / or higher injection pressures.

The injection system includes an injector with a nozzle seat and a nozzle element with more than 7 holes to ensure high homogeneity of the mixture and low soot formation during the diesel combustion process. Preferably, the injector has a number of holes of> = 8, 9, 10, 12 or 14. The injector is preferably arranged centrically to the piston recess.

For an 8-hole nozzle injector, a hole diameter of 0.123 mm is preferred. The larger number of holes allows a more even distribution of the fuel in the combustion chamber and thus a emission reduction with the lowest possible fuel consumption.

The injector flow rate is adjusted to the number of holes. For a hole number> 7, the flow rate DDF at the fuel through the injector is preferably determined within 30 seconds at an injection pressure of 100 bar by DDF = (-230 + 70 * L) * (1 + -0.20), the number of holes is denoted by L.

When measuring the flow rate of the injector, only the nozzle seat is measured without a nozzle element.

The cylinder is connected to an inlet device for charge air with at least one inlet channel with an associated inlet valve and an outlet device for exhaust gas with at least one outlet valve and an associated outlet valve, as will be shown in more detail below. The charge air is the charge of fresh air supplied to the engine - even in the case of an engine without turbocharging or the like.

In the engine according to the invention, the inlet device has at least one inlet channel, which is designed as a swirl channel with an associated swirl channel valve, the inlet flow into the combustion chamber receives a swirl - a turbulence of the charge air with a substantially parallel to the cylinder axis oriented axis of rotation. When Inlet flow is the charge air entering the cylinder during the intake phase.

The swirl of intake flow into the combustion chamber aids in mixture preparation and can increase the kinetic energy in the cylinder, which can contribute to accelerated combustion. As the swirl extends the ignition delay during combustion, it reduces the application of fuel to the cylinder walls and thus the formation of hydrocarbons and soot. On the other hand, low swirl results in lower heat losses, better pre-ignition efficiency, and lower NOx emissions. Therefore, more or less swirl can be beneficial for particulate and NOx emissions. The invention aims at optimizing the inlet flow din to reduce soot formation with low NOx formation.

The swirl of the intake flow is determined by the geometry of the intake port and the intake valve; in the case of more than one inlet channel, the geometry of the inlet channels and of the inlet valves, through which charge air flows in each case. To achieve a given twist, the geometry of inlet channel and associated inlet valve are determined by means of known optimization methods.

In a preferred embodiment, the internal combustion engine according to the invention has an inlet device with a swirl channel and a filling channel, wherein the filling channel for intercooling can be opened or closed operating point-dependent.

In 3a . 3b are a spin channel 301 and a filling channel 302 an inlet device for charge air of the internal combustion engine according to the invention shown.

swirl channel 301 and filling channel 302 each have a swirl channel valve 307 or filling channel valve 308 on. The filling channel 302 can by a switchable swirl flap 303 be opened and closed. Preferably, the swirl flap 303 at least two open positions, preferably three and more open positions. The open position is understood to mean the closed position. Furthermore, the swirl flap can also be adjusted continuously.

In 3a is the swirl flap 303 closed, in 3b is the swirl flap 303 open against it. When the swirl flap is closed, charge air can only pass through the swirl channel 301 be supplied to the combustion chamber.

In the in the 3a . 3b shown swirl channel 301 it is a tangential channel with direct inflow. However, the invention also includes swirl channels of other geometries, such as a baffle tangential channel or spiral or flat helical ramp spiral channels.

The filling channel 302 is traversed as straight as possible, to a higher cylinder filling than through the swirl duct 301 to allow in which the air flow obstructs the flow. By means of the swirl channel 302 Therefore, depending on demand, the air flow can be increased. However, the filling channel can also be designed as a spiral channel.

The swirl passage valve is designed according to the invention to increase the swirl at low valve strokes, as exemplified in 3 c is illustrated.

To increase the twist at low valve lift, the valve seat has a, preferably asymmetric with respect to a vertical valve axis, machining; For example, in the manner of a masking or shielding, whereby the air mass flow is directed at the beginning of the valve lift in a defined direction.

In the case of a valve with shielding, the defined flow direction is predetermined by a shielding plate on the valve; compare left side of 3 c , At a small valve lift is in an area of the periphery of the valve seat 315 between 30 ° and 210 ° a gap 317 educated.

The right side of 3 c shows a valve seat, wherein in the cylinder head a bevel is incorporated, which is referred to as Sitzdrallfase (SDF). At a small valve lift is in an area of the periphery of the valve seat 315 between 30 ° and 210 ° a gap 318 educated.

In both cases, the air mass flow at small valve strokes (in a range of 2mm to 4mm or 50% of the maximum valve lift Vhm) through the gap 317 respectively. 318 stream. Preferably, the air mass flow is deflected at small valve strokes of the cylinder wall and thus causes a rotation of the cylinder charge.

In the context of the invention, the swirl of the inlet flow through the valve-dependent Tippelmann swirl number, which can be measured on a Tippelmann test bench; see. JB Heywood Internal Combustion Engine Fundamentals, 1998 ,

• • M_z -Drehmoment an der Wabe
• • R_Zyl -Zylinderradius
• • V-Volumenstrom durch den Zylinder
• • p_L -Dichte der Luft im Zylinder

In 4 is the schematic structure of a Tippelmann test bench 401 for measuring the valve lift-dependent Tippelmann swirl number D shown, wherein through an inlet channel 404 with an inlet valve 405 incoming charge air to the interior of the cylinder 402 is supplied. A swirling flow 406 is in the cylinder 402 over a honeycomb 403 in the direction of the cylinder axis 407 diverted. That from the angular momentum of the swirling flow 406 resulting torque M _Z gets on the honeycomb 403 depending on the valve lift of the intake valve 405 measured. From the torque Mz, the valve lift-dependent Tippelmann swirl number D can be calculated using the following equation. $$\text{D}=\frac{Mz\cdot R_{Zyl}}{{\mathrm{<figure-callout\; id="2"\; label="V"\; filenames="DE102007053890B4\_0010.png,DE102007053890B4\_0011.png"\; state="\{\{state\}\}">V</figure-callout>}}_2\cdot p_L}$$

• • M _z torque at the honeycomb
• • R _Cyl cylinder radius
• • V-volume flow through the cylinder
• • p _L density of the air in the cylinder

In contrast to the twist moment Mz the swirl number D is directly comparable for different cylinder heads of similar engines and is therefore used to characterize the swirl according to the invention.

The relatively large piston recess has the advantage of a better charge with charge air and a lower thermal load of the piston, but causes a decrease in the twist compared to a relatively smaller piston recess. This decrease in the twist is counteracted according to the invention.

• die Einlassströmung bei zumindest einem Ventilhub Vh zwischen 2 mm und Vhh = 0.5*Vhm eine Drallanhebung mit einer ventilhubabhängigen Tippelmann-Drallzahl D mit einen Wert > 0,3; 0,4; 0,5; 0,6; 0,7 aufweist, wobei Vhm den maximalen Ventilhub, insbesondere mit einem Wert Vhm=8mm bezeichnet,
• die Einlassströmung in einem Ventilhubbereich zwischen Vh = 2 mm und Vhh = 0.5*Vhm eine Drallanhebung mit einer normierten integralen Tippelmann-Drallzahl Di mit einen Wert > 0,40/mm, 0,50/mm oder 0,60/mm aufweist oder
• die Einlassströmung in einem Ventilhubbereich zwischen Vh= 2 mm und Vhh=0,5Vhm eine Drallanhebung aufweist, derart dass bei einer Vermessung der ventilhubabhängigen Tippelmann-Drallzahl D in Abständen von 0,5 mm die gemittelte Tippelmann - Drallzahl Dm für mindestens ein Ventilhubintervall DeltaV= 1mm einen Wert ≥ 0,4; 0,5; 0,6; 0,7 aufweist.

According to the invention, the internal combustion engine has an inlet flow with swirl lift at low valve lifts, wherein it is provided that

• the inlet flow at at least one valve lift Vh between 2 mm and Vhh = 0.5 * Vhm a swirl increase with a valve lift-dependent Tippelmann swirl number D with a value>0.3;0.4;0.5;0.6; 0.7, wherein Vhm denotes the maximum valve lift, in particular with a value Vhm = 8 mm,
• the inlet flow in a valve lift range between Vh = 2 mm and Vhh = 0.5 * Vhm has a swirl lift with a normalized integral Tippelmann swirl Di having a value> 0.40 / mm, 0.50 / mm or 0.60 / mm or
• the inlet flow in a Ventilhubbereich between Vh = 2 mm and Vhh = 0.5Vhm has a swirl increase, so that in a measurement of valve lift-dependent Tippelmann swirl number D at intervals of 0.5 mm, the average Tippelmann swirl number Dm for at least one Ventilhubintervall DeltaV = 1mm a value ≥ 0.4; 0.5; 0.6; 0.7.

Inlet flows of compression ignition internal combustion engines having a relatively large piston recess in the prior art have a valve lift dependent Tippelmann swirl number D with a value around 0.2 for small valve strokes (range of 2mm to 5mm or range of 2mm to 50% of Vhm).

Advantageously, in the internal combustion engine according to the invention, the inlet flow in a range of small valve strokes on a swirl lift, so that in this valve lift a definite high swirl orientation of the inlet flow is maintained, in a piston with a relatively large piston bowl with a ratio of bowl rim diameter to piston rim diameter> 0, It allows both low particulate and raw NOx emissions to be achieved.

The 5 shows measurements of the swirl number in an inlet device with a swirl channel and a filling channel, wherein the filling channel is closed by a swirl flap (single-channel operation), and thus an increased swirl of the inlet flow is effected. The swirl increase at small valve strokes is preferably achieved by machining the valve seat of the swirl passage valve. Particularly preferred is the mentioned Sitzdrallfase SDF.

In 5 is with the curves 501 respectively. 503 the valve lift-dependent Tippelmann swirl number D is plotted against the valve lift, wherein 501 the swirl number according to the invention with a Sitzdrallfase and the curve 503 describes the swirl number with otherwise the same swirl passage and swirl passage valve without swirl increase by a Sitzdrallfase.

The maximum valve lift has a value of 8mm, but can of course also have a different value.

It can be seen that the inlet flow according to curve 501 at low valve strokes of 2mm, 3mm or 4mm, each have high values D of 0.7, 0.4 and 0.3.

Values of the swirl number> 2; 3; 4 in the valve lift range 2mm to 5mm.

Further, in a valve lift range between Vh = 2 mm and Vhh = 0.5Vhm, the intake flow may have a swirl increase with a normalized integral Tippelmann swirl number Di having a value> 0.40 / mm. The integral integral Tippelmann swirl number is the integral of a valve lift interval divided by the interval length.

Further, the inlet flow in a Ventilhubbereich between Vh = 2 mm and Vhh = 0.5Vhm have a swirl increase, such that in a measurement of Ventilhubabhängigen Tippelmann swirl number D at intervals of 0.5 mm the average Tippelmann swirl number Dm has a value ≥ 0.4 for at least one valve lift interval DeltaV = 1 mm.

In 5 It can be seen that the swirl number plotted against the valve lift between a value at Vh = 2mm and a value at 8mm approximately at 4mm valve lift has a minimum. The swirl number decreases from its value at 2 mm valve lift, reaches a minimum in a range of 0.5 V and then increases again. With this "bathtub shape" it is possible to achieve high swirl numbers with low and high valve lift.

The inlet flow according to curve 503 on the other hand has smaller swirl numbers D in the range between 0.15 and 2 for small valve strokes in the range between 2 and 4 mm.

According to the invention, a high twist D of more than 0.4 is also provided at maximum valve lift in order to further improve the mixture preparation. The high swirl at maximum valve lift is preferably achieved by designing the swirl duct.

In addition to the swirl number, the flow through the inlet valve (s) is an important parameter for characterizing the inlet flow.

The valve lift-dependent flow coefficient αk characterizes the inflow into the cylinder through the intake valve (s), that is, the permeability of the valve, where an actual air mass flow is related to a theoretically possible air mass flow, as for example in the publication C. Kopp, Variable Valve Timing for Cars Diesel engines with direct injection, Magdeburg 2006 is shown. A high value of α _k has a positive effect on the cylinder filling.

• • m_tats -tatsächlicher Luftmassenstrom [kg/s]
• • m_theor -theoretischer Luftmassenstrom [kg/s]
• • A_zyl -Querschnittsfläche des Messzylinders [m²]
• • c_s -Strömungsgeschwindigkeit bei isentroper Durchströmung [m/s]
• • p_s -Luftdichte im Zylinder bei isentroper Durchströmung [kg/m³]

wobei $${\text{c}}_{\text{s}}=\sqrt{\frac{2\cdot \kappa \cdot R_L\cdot T}{\kappa -l}\cdot \left[1-{\left(\frac{p_{nach}}{p_{vor}}\right)}^{\frac{K-1}{K}}\right]}$$

• • R_L -spezielle Gaskonstante für Luft [J/(kg·K)]
• • T -Temperatur [K]
• • κ -Isentropenexponent [-]
• • p_nach -Druck nach Ventil [Pa]
• • p_vor -Druck vor Ventil [Pa]

$${\text{p}}_{\text{s}}=\frac{p_{vor}}{R_L\cdot T}\cdot {\left(\frac{p_{nach}}{p_{vor}}\right)}^{\frac{1}{K}}$$

α _k for a cylinder can be determined according to the following formula. $${\mathrm{\alpha }}_{\text{k}}=\frac{m_{tats}}{m_{tHeOr}}=\frac{m_{tats}}{A_{Zyl}\cdot c_s\cdot p_s}$$

• • m _tats -tatsächlicher air mass flow [kg / s]
• • _Theoretical mass air flow [kg / s]
• • A _{cyl cross-} sectional area of the measuring cylinder [m ² ]
• • c _s flow velocity at isentropic flow [m / s]
• • p _s- air density in the cylinder during isentropic flow [kg / m ³ ]

in which $${\text{c}}_{\text{s}}=\sqrt{\frac{2\cdot \kappa \cdot R_L\cdot T}{\kappa -l}\cdot \left[1-{\left(\frac{p_{nacH}}{p_{vOr}}\right)}^{\frac{K-1}{K}}\right]}$$

• • R _L -special gas constant for air [J / (kg · K)]
• • T temperature [K]
• • κ-isotropic exponent [-]
• • p _after pressure after valve [Pa]
• • _pre- pressure before valve [Pa]

$${\text{p}}_{\text{s}}=\frac{p_{vOr}}{R_L\cdot T}\cdot {\left(\frac{p_{nacH}}{p_{vOr}}\right)}^{\frac{1}{K}}$$

The pressure before the valve p _before corresponds to the ambient pressure in the above equations. After the valve the pressure prevailing in the measuring cylinder pressure p is given _by by the compressor of the test stand. During the measurement a constant pressure difference of 50hPa is set.

When designing the intake ports and the intake valves, it is preferable to take into account a trade-off between high swirl for good mixture preparation and high valve permeability for good cylinder charge. This is achieved by providing a swirl increase at low valve lifts and a high flow at maximum valve lift.

In 5 turns curve 503 the course of the valve lift-dependent flow coefficient αk in a range of the valve lift between 2 mm and 8 mm.

Preferably, with the engine according to the invention at 8 mm valve lift, the flow coefficient is> 0.04. In further embodiments of the invention, the flow coefficient αk is> 0.045 or 0.005 at 8 mm valve lift. Thus, in the invention, a high swirl is combined with small valve strokes with a high flow coefficient at large or maximum valve lift.

The 6 shows measurements of the swirl number in an inlet device with a swirl channel and a fully open filling channel (two-channel operation), wherein the filling channel is not closed by a swirl cap and thus a higher throughput of charge air can be achieved. For a valve lift in a range between 2 and 4 mm, the valve lift-dependent Tippelmann swirl number has a value> 0.035. With a 2 mm valve lift, the swirl number is> 0.45. With a valve lift of 8 mm, the swirl number is> 0.35.

In 6 is analogous to 5 recognizable that the swirl number plotted against the valve lift between a value at Vh = 2mm and a value at 8mm approximately at 4mm valve lift has a minimum. The swirl number decreases from its value at 2mm valve lift, reaches its minimum within a range of 0.5 Vhm (8mm) and then increases again. With this "bathtub shape" it is possible to achieve high swirl numbers with low and high valve lift.

In a further embodiment, the swirl number at 8 mm valve lift has a maximum value of 0.6.

The curve 602 represents the course of the valve lift-dependent flow coefficient ak and points opposite to the curve 502 of the 5 an increase in the valve lift-dependent flow coefficient ak.

The invention makes it possible to adjust the soot-carbon emissions as a function of an opening position of a swirl flap depending on the operating point with a constant NOx crude emission.

In 7 are exemplary for a supercharged 4-cylinder diesel engine with 1,968cm ³ of displacement and a bore of 81mm at 95,5mm stroke, a depression diameter of 53mm, a bowl rim diameter of 51,88mm, a maximum power of 103kW at a speed of 4200 l / Min., A maximum torque of 320Nm at a speed of 1750 to 2500 l / min., A compression of 16.5 to 1 soot - raw emissions in units of filter smoke number (FSN) at a constant NOx raw emission of 0.5g / kWh shown. The operating point is at a speed of 1500 l / min. and an effective mean pressure of 2 bar. It is understood that the invention is not limited to this specific engine, but includes similar engines, if necessary, even without charging.

The inlet device used here has a swirl channel and a filling channel, wherein the filling channel is provided with a swirl flap. The swirl flap can, preferably continuously be switched to different open positions, wherein at a position of 90 °, the swirl flap is closed and fully open at an open position of 0 °. Intermediate positions of 60 ° and 40 ° are preferred.

The curve 701 indicates the course of the soot emissions in an inventive engine with swirl lift. For comparison, the curve shows 702 the course of the soot emission as a function of the opening position of the swirl flap in a same engine without swirl increase. It can be seen that the soot emissions according to curve 701 for the engine according to the invention at all opening positions by at least 0.1 FSN below the values of the comparison curve 702 lie. It can also be seen that the soot emissions with a closed swirl flap, ie a single-channel operation via the swirl duct, are the lowest and increase with increasing opening of the swirl flap.

In 8th For the same engine at an operating point of 2000 l / min and an effective mean pressure of 8 bar, the soot emissions at a constant NOx raw emissions of 1.3 g / kWh are shown for different opening positions of the swirl flap. The curve 801 indicates the course of the soot emission as a function of the open position of the swirl flap according to the invention, while the curve 802 Comparative data without swirl increase according to the invention shows. It can be seen that the soot emission in the case according to the invention is at least 0.1 FSN lower than in the case of conventional technology. In contrast to the situation according to 7 At this operating point, the two-channel operation with fully open swirl flap is more favorable than the single-channel operation.

In 9 are for the same engine at an operating point at a speed of 2000 l / min and an effective mean pressure of 5 bar, the soot raw emissions at a constant NOx raw emissions of 1 g / kwh different opening positions of the swirl flap shown. The curve 901 indicates the course of the soot emission dependency from the opening position of the swirl flap according to the invention, while the curve 902 Comparative data without swirl increase according to the invention shows. A minimum value of the soot emission is here at an intermediate position of the swirl flap with an open position of about 60 °.

The 10-11 show the 7-9 corresponding representations, wherein the soot emissions are plotted not in units FSN but in units g / kWh.

The internal combustion engine with auto-ignition can therefore, if the swirl duct valve has a machining for swirl increase at small valve lifts operated by means of different opening positions of the swirl flap so that at an operating point of 1500 l / min. and an effective mean pressure of 2 bar and a constant NOx raw emissions of <than 0.7 g / kWh a soot number of <0.2 FSN (this corresponds to soot emissions of <than 0.2 g / kWh) is set. The swirl increase is preferably in a range between 2mm and 4mm, if necessary. With spin values and flow coefficients as in connection with the 5 and 6 and the associated description has been presented.

It is also envisaged that at an operating point of 2000 l / min. and 5 bar effective mean pressure and a raw NOx emission of <than 1.2 g / kWh soot a soot number FSN of <0.8 (corresponding to soot emissions of <than 0.1 g / kWh) is set and / or in one Operating point at 2000 1 / min. and 8 bar effective mean pressure and a constant NOx raw emission of less than 1.5 g / kWh a soot number <1FSN (corresponding to soot emissions of less than 0.2 g / kWh) is set.

LIST OF REFERENCE NUMBERS

101

internal combustion engine

102

block

103

intake means

104

injection

105

outlet

106

exhaust system

107

particulate Filter

107a

catalyst

108

Exhaust gas recirculation system

109

Exhaust gas recirculation valve

110

Exhaust gas recirculation cooling

111

Cooling circuit

115

turbocharger

116

Intercooler

201

piston head

202

Surface of the piston head

203

Bowl rim diameter

204

Piston diameter

205

cooling channel

206

piston bowl

207

depression contour

208

bowl edge

209

piston axis

301

swirl channel

302

filling channel

303

swirl flap

304

Airflow swirl channel

305

Air flow filling channel

306

cylinder

307

Valve

308

Valve

310

intake means

315

valve seat

316

valve seat

317

gap

401

Tippelmann Measuring Stand

402

cylinder

403

honeycomb

404

inlet channel

405

Valve

406

swirl flow

407

cylinder axis

501

Curve

502

Curve

601

Curve

602

Curve

701

Curve

702

Curve

801

Curve

802

Curve

901

Curve

902

Curve

Claims (21)

A self-ignition internal combustion engine (101) having an engine block (102) having at least one cylinder (306, 402) and a piston associated therewith, an injection system (104) for direct fuel injection into the cylinder (306, 402) and an intake device (103, 310 ) for charge air with at least one inlet channel (404) and an inlet valve (110) for exhaust gas with at least one outlet channel and an exhaust valve associated therewith and a subsequent exhaust system (106) with particulate filter (107), wherein the cylinder (306, 402) comprises a cylinder head and the piston has a piston head (201) and cylinder head and piston head (201) inner surfaces of a combustion chamber and in the surface of the piston head (202) has a piston recess (206) is arranged with a trough edge (208), wherein at a on a circle standardized trough edge (208) the ratio of bowl edge diameter (203) to piston diameter (204) greater than 0, 55 or in a 90% filled piston bowl (206) at one on a circle normalized filling surface, the ratio of the filling surface to the piston cross-sectional area> 0.30, an inlet channel (404) is formed as a swirl channel (301) with an associated swirl channel valve, wherein the inlet flow into the combustion chamber receives a twist, characterized in that for at least one selected cylinder ( 306, 402) the inlet flow at at least one valve lift Vh between 2 mm and Vhh = 0.5Vhm a swirl increase with a valve lift-dependent Tippelmann swirl D has a value> 0.3, where Vhm the maximum valve lift or in particular a value Vhm = 8mm , the inlet flow in a valve lift range between Vh = 2 mm and Vhh = 0.5Vhm a swirl increase with a normalized integral Tippelmann swirl Di has a value> 0.40 / mm or the inlet flow in a valve lift between 2 mm and Vhh = 0.5Vhm having a swirl increase, such that during a measurement of the valve lift-dependent Tippelmann swirl number D at intervals of 0.5 mm, the mean Tippelmann swirl number Dm has a value ≥ 0.4 for at least one valve lift interval DeltaV = 1 mm. Internal combustion engine (101) after Claim 1 , characterized in that the inlet flow at maximum valve lift Vhm, in particular with Vhm = 8 mm, a valve lift-dependent Tippelmann swirl D has a value> 0.4 and / or the inlet valve a valve lift-dependent flow coefficient αK> 0.04. Internal combustion engine (101) after Claim 1 or 2 , characterized in that the inlet device (103) has as a further inlet channel a filling channel (302) with an associated filling channel valve. Internal combustion engine (101) after Claim 3 , characterized in that the filling channel (302) can be opened and closed by a switchable swirl flap (303), wherein the swirl flap (303) preferably has at least two open positions. Internal combustion engine (101) after Claim 3 or 4 , characterized in that the inlet flow at fully open Füllkanal (302) at least one valve lift Vh between 2 mm and Vhh = 0.5Vhm a Ventilhubabhängige Tippelmann swirl number D has a value> 0.3. Internal combustion engine (101) after Claim 5 , characterized in that at maximum valve lift Vhm and fully open Füllkanal (302) the valve lift-dependent Tippelmann swirl number D has a value> 0.6 and / or the valve lift-dependent passage coefficient αK of swirl passage valve and Füllkanalventil has a value> 0.06. Internal combustion engine (101) according to one of claims 1 to 6, characterized in that up to an effective mean pressure pme of less than 4 bar operation is possible only with the charge air supplied via the swirl duct (301). Internal combustion engine (101) according to one of Claims 3 to 7 , characterized in that after a range between 3 and 7 bar operation with a partially open filling channel (302) is possible. Internal combustion engine (101) according to one of Claims 3 to 8th , characterized in that for an effective mean pressure pme> 5 bar operation with a fully open filling channel (302) is possible. Internal combustion engine (101) according to one of the preceding Claims 4 to 9 , characterized in that the swirl flap (303) is pivotable about an axis and can assume at least two open positions from the group of the open positions of 90 °, 60 °, 40 ° or 0 °. Internal combustion engine (101) after Claim 10 , characterized in that the internal combustion engine (101) at a speed of 1500 l / min. and an effective mean pressure of 2 bar with constant NOx emission has soot emissions that are at least 0.1 units below the value of the soot emissions without swirl increase at the same speed, the same mean effective pressure and the same NOx emission. Internal combustion engine (101) after Claim 10 or 11 , characterized in that at an operating point with a speed of 2000 1 / min. and an effective mean pressure of 8 bar, the soot emissions are at least 0.1 units below the value without swirl elevation at the same operating point. Internal combustion engine (101) according to one of the preceding Claims 1 to 12 , characterized in that an exhaust gas recirculation system (108) is provided. Internal combustion engine (101) according to one of the preceding Claims 1 to 13 , characterized in that an exhaust gas turbocharger is provided. Internal combustion engine (101) according to one of the preceding Claims 1 to 14 , characterized in that the injection system (104) is designed as a common rail system. Internal combustion engine (101) according to one of the preceding Claims 1 to 15 , characterized in that the piston recess (206) has a central region with a smaller depth and a peripheral region with a greater depth, in each case with respect to the flat region of the piston surface. Internal combustion engine (101) according to one of the preceding Claims 1 to 16 , characterized in that the piston recess (206) has an undercut in its peripheral wall region. Internal combustion engine (101) according to one of the preceding Claims 1 to 17 , characterized in that the selected cylinder (306, 402) has a bore <= 100 mm. Internal combustion engine (101) according to one of the preceding Claims 1 to 18 , characterized in that the single cylinder volume is <= 0.8 L. Internal combustion engine (101) according to one of the preceding Claims 1 to 19 , characterized in that the rated speed> than 3100 l / min. is. Internal combustion engine (101) according to one of the preceding Claims 1 to 20 , characterized in that the specific power is> 30 kW.

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