WO2011110314A1

WO2011110314A1 - Internal combustion engine having two-stage supercharging

Info

Publication number: WO2011110314A1
Application number: PCT/EP2011/001074
Authority: WO
Inventors: Ekkehard Pott; Bernd Hahne; Robin Vanhaelst
Original assignee: Volkswagen Aktiengesellschaft
Priority date: 2010-03-06
Filing date: 2011-03-04
Publication date: 2011-09-15
Also published as: EP2545265A1; DE102010010480A1

Abstract

The invention relates to an internal combustion engine, in particular for a motor vehicle, comprising a fresh air tract (12) for feeding fresh air to working cylinders (10) of the internal combustion engine, an exhaust gas tract (14) for discharging exhaust gas (21) from the working cylinders (10), a first exhaust gas turbocharger (16) of a low-pressure stage (low-pressure exhaust gas turbocharger), wherein said first exhaust gas turbocharger has a first turbine (26) (low-pressure turbine) disposed in the exhaust gas tract (14) and a compressor (36) (low-pressure compressor) disposed in the fresh air tract (12), and at least one second exhaust gas turbocharger (18) of a high-pressure stage (high-pressure exhaust gas turbocharger), wherein said second exhaust gas turbocharger has a second turbine (24) (high-pressure turbine) disposed in the exhaust gas tract (14) upstream of the first turbine (26) and a second compressor (38) (high-pressure compressor) disposed in the fresh air tract (12) downstream of the first compressor (36). At least one of the turbines (24, 26) has a variable turbine geometry (68) (VTG).

Description

description

Internal combustion engine with two-stage supercharging

The invention relates to an internal combustion engine, in particular of a motor vehicle, with a fresh air tract for supplying fresh air to working cylinder of the internal combustion engine, an exhaust tract for discharging exhaust gas from the working cylinders, a first exhaust gas turbocharger of a low-pressure stage (LP exhaust gas turbocharger), which arranged in the exhaust tract first Turbine (ND turbine) and arranged in the fresh air duct compressor (LP compressor), and at least a second exhaust gas turbocharger a high-pressure (HD exhaust gas turbocharger), which arranged in the exhaust system upstream of the first turbine second turbine (HP turbine) and a second compressor (HP compressor) arranged in the fresh air tract downstream of the first compressor, both turbines having an adjustable turbine geometry (VTG), according to the preamble of patent claim 1.

Known diesel engines with controlled two-stage supercharging have a circuit consisting of a wastegate or VTG high-pressure loader (VTG - variable turbine geometry), a bypass and a wastegate downstream wastegate loader. Such engines always have external exhaust gas recirculation (EGR), which has a separate EGR cooler. The reduction in emissions is usually the highest possible EGR rate

desirable. The maximum EGR rate is influenced by the pressure ratio between the intake and exhaust side. For this reason, double-charged diesel engines usually have a throttle valve integrated in the intake manifold, with which the pressure gradient can be raised to the intake side. However, this increases the charge exchange work and the consumption increases. By means of exhaust-side throttling, the pressure gradient can also be increased as an alternative or in addition to intake manifold throttling. When VTG loaders with adjustable VTG blades are used, this is particularly easy because the VTG blades are not optimized for best turbine performance but optimized for the highest exhaust backpressure. In parallel with the VTG blades, an in-turbine bypass is possible, so that a limitation of the turbine speed is possible even without external bypass.

From DE 10 2004 056 894 A1 a generic internal combustion engine with two-stage supercharging is known, wherein the turbines of both stages each have a variable turbine geometry.

CONFIRMATION COPY From DE 198 51 028 C2 is a method for operating a charged

Internal combustion engine is known, which has two exhaust gas turbochargers arranged in parallel, wherein in each case only one exhaust gas turbocharger is operated and is switched after predetermined time intervals between the two exhaust gas turbochargers. Both arranged in parallel

Exhaust gas turbochargers are each formed with a variable turbine geometry.

The invention is based on the object, an internal combustion engine og. To improve the type of emission and consumption as well as the system complexity

simplify.

This object is achieved by an internal combustion engine o.g. Art solved with the features characterized in claim 1. Advantageous embodiments of the invention are described in the further claims.

For this it is in an internal combustion engine o.g. Art provided according to the invention that the exhaust gas downstream of the first turbine via the first turbine and in addition via a first wastegate with first wastegate valve with the exhaust gas upstream of the first turbine and downstream of the second turbine fluidly connected, the exhaust tract downstream of the second turbine and upstream of the first turbine via the second turbine and additionally via a second turbine bypass duct with second turbine bypass duct valve and a second wastegate with second wastegate valve with the exhaust gas upstream of the second turbine fluidly connected, wherein the fresh air tract upstream of the first compressor

exclusively via the first compressor with the fresh air tract downstream of the first

Compressor and upstream of the second compressor is fluidly connected, wherein the

Fresh air tract upstream of the second compressor and downstream of the first compressor via the second compressor and in addition via a second compressor bypass passage with the second compressor bypass passage valve fluidly connected to the fresh air tract downstream of the second compressor, wherein downstream of the second compressor, a second charge air cooler is arranged, wherein upstream of the second turbine a high-pressure exhaust gas recirculation (HP-EGR) line branches off from the exhaust tract and opens downstream of the second compressor in the fresh air tract, the HD EGR line downstream of the second intercooler in the

Fresh air tract opens, wherein in the HP-EGR line, a cooler for recirculated exhaust gas (HP-EGR cooler) is arranged, wherein in the HP-EGR line, a valve for recirculated exhaust gas (HP-EGR valve) is arranged.

This has the advantage that without additional exhaust flap a maximum possible

Exhaust gas recirculation rate (EGR rate) can be realized because by means of the VTG exhaust back pressure can be increased accordingly, so that there is an extended map range, in which the pressure in the exhaust system at the removal point for recirculating exhaust gas is greater than the pressure in the fresh air tract at the point of introduction of recirculating exhaust gas. The controllability of the exhaust side counterpressure is maximized. In addition, an engine braking effect can be realized by adjusting the VTG without additional equipment. It is achieved an extension of the control ability, in particular with regard to the boost pressure of the two-stage supercharging system. The fact that the first and the second turbine each have a VTG results in addition to an enlarged control spectrum with regard to the exhaust backpressure. A simultaneous reduction of pollutant emissions of the internal combustion engine is achieved in that upstream of the second turbine, a high-pressure exhaust gas recirculation (HP-AGR) line branches off from the exhaust tract and downstream of the second compressor in the

Fresh air tract opens. In order to avoid contamination of the second intercooler by recirculated exhaust gas, the HP-EGR line opens downstream of the second

Intercooler in the fresh air tract.

To increase the efficiency of the charge, a third charge air cooler is arranged in the second compressor bypass duct.

To increase the efficiency of the charge, a first charge air cooler is arranged downstream of the first compressor and upstream of the second compressor.

A reduction of pollutant emissions of the internal combustion engine is achieved in that downstream of the first turbine, a low-pressure exhaust gas recirculation (ND-EGR) line branches off from the exhaust tract and upstream of the first compressor in the fresh air tract or in the HD EGR cooler of the HD-EGR -Leitung opens.

For cooling the recirculated exhaust gas and / or for influencing the amount of

recirculated exhaust gas, a recirculated exhaust gas cooler (LP EGR cooler) and / or a recirculated exhaust gas valve (LP EGR valve) are arranged in the LP EGR passage.

A reduction of pollutant emissions of the internal combustion engine is achieved in that upstream of the first turbine and downstream of the second turbine, a medium-pressure exhaust gas recirculation (MD-EGR line) branches off the exhaust tract and opens downstream of the first compressor and upstream of the second compressor in the fresh air tract ,

For cooling the recirculated exhaust gas and / or for influencing the amount of

recirculated exhaust gas is arranged in the MD-EGR line a recirculated exhaust gas cooler (MD-EGR cooler) and / or a valve for recirculated exhaust gas (MD-EGR valve). To adapt the internal combustion engine in the respective desired performance class, this has at least 3, in particular 4, 5, 6, 8, 10 or 12 working cylinder.

A particularly functionally reliable fuel supply and a specifically controllable combustion are achieved in that the internal combustion engine has a direct fuel injection into at least one working cylinder, in particular according to the common rail system.

To adapt the internal combustion engine in the respective desired performance class, this has a rated speed of at least 3,000 revolutions per minute, in particular 3,500, 4,000, 4500, 5000 or 5,000 revolutions per minute, on.

One adapted to a desired level of performance and torque level

Internal combustion engine is achieved in that at least one working cylinder has a displacement of less than or equal to 800 cc, in particular less than or equal to 700 cc, 600 cc, 500 cc, 400 cc or 350 cc.

For discharging exhaust gas from the working cylinders into the exhaust tract, at least one working cylinder is assigned at least one, in particular two or more, exhaust valves.

For supplying combustion air to the working cylinders from the fresh air tract at least one working cylinder at least one, in particular two or more, inlet valves are assigned.

A saving in space is achieved in that at least one first and second compressor are arranged in a common housing.

The invention will be explained in more detail below with reference to the drawing. This shows in

Fig. 1 shows a first known embodiment of an internal combustion engine in a schematic

Presentation,

Fig. 2 shows a second known embodiment of an internal combustion engine in a schematic

Presentation,

Fig. 3 shows a third known embodiment of an internal combustion engine in a schematic

Presentation, 4 shows a fourth known embodiment of an internal combustion engine in a schematic representation,

Fig. 5 shows a fifth known embodiment of an internal combustion engine in a schematic

Presentation,

Fig. 6 shows a sixth known embodiment of an inventive

Internal combustion engine in a schematic representation,

Fig. 7 shows a seventh known embodiment of an internal combustion engine in a schematic

Presentation,

Fig. 8 shows an eighth known embodiment of an internal combustion engine in a schematic

Presentation,

9 shows a preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

10 shows a tenth known embodiment of an internal combustion engine in a schematic

Presentation,

11 is a graphical representation of emission regions of the internal combustion engine,

Fig. 12 is a graph of soot flow rate and NO _x flow rate at a

Operating condition of the internal combustion engine with 1,500 min ^-1 and with a mean pressure pme = 3 bar,

FIG. 13 is a graph showing a boost pressure variation in an operating condition of FIG

Internal combustion engine with 1,500 rpm ^"1 and with a mean pressure pme = 3 bar,

Fig. 14 is a graph of exhaust gas pressure before turbine without EGR in a

Fig. 15 is a graph of exhaust gas pressure before turbine with EGR at a

Operating condition of the internal combustion engine with 1,500 min ^-1 and with a mean pressure pme = 3 bar, 16 is a graph of a maximum achievable EGR rate at a

17 is a graphical representation of a minimum achievable NO _x emission at a

FIG. 18 is a graph showing a boost pressure variation in an operating condition of FIG

Internal combustion engine with 2,000 rpm ^"1 and with a mean pressure pme = 8 bar,

FIG. 19 is a graph of soot flow rate and NO _x flow rate in FIG

Operating state of the internal combustion engine with 2,000 min ^"1 and with a medium pressure pme = 8 bar,

FIG. 20 is a graph showing a boost pressure variation in an operating condition of FIG

Internal combustion engine with 2,000 min ^"1 and with a mean pressure pme = 8,9 bar and

FIG. 21 is a graph of soot flow rate and NO _x flow rate in FIG

Operating state of the internal combustion engine with 2,000 min ^"1 and with a mean pressure pme = 8.9 bar.

The illustrated in Fig. 1, known embodiment of an inventive

Internal combustion engine comprises working cylinder 10, a fresh air tract 12, an exhaust tract 14, a first exhaust gas turbocharger 16 (ND-ATL) of a first charging stage (low-pressure stage) and a second exhaust gas turbocharger 18 (HD-ATL) of a second charging stage (high-pressure stage). The exhaust tract 14 includes an exhaust manifold 20 for collecting exhaust gas 21 discharged from the working cylinders 10, and an exhaust passage 22. In the exhaust passage 22, a second turbine 24 (HP turbine) of the second exhaust gas turbocharger 18 and a first turbine 26 (ND turbine ) of the first exhaust gas turbocharger 16.

The fresh air tract 12 comprises a fresh air duct 44 in which, as seen in the flow direction, a first compressor 36 (LP compressor) of the first exhaust gas turbocharger 16, a second compressor 38 (HP compressor) of the second exhaust gas turbocharger 18, a second charge air cooler 40, a throttle valve 41 and a suction pipe 42, which opens into the working cylinder 10 via inlet valves (not shown). Furthermore, the fresh air tract 12 has a second compressor bypass passage 46, which the second compressor 38 of the second

Exhaust gas turbocharger 18 bridged. The bypass channel 46 branches downstream of an outlet 48 of the first compressor 36 of the first exhaust gas turbocharger 16 and upstream of an inlet 50 of the second compressor 38 of the second exhaust gas turbocharger 18 from the fresh air duct 44 and opens downstream of an outlet 52 of the second compressor 38 of the second exhaust gas turbocharger 18 and upstream of the second charge air cooler 40 back into the fresh air duct 44th one. In the second compressor bypass passage 46, a second compressor bypass valve 54 is arranged. This second compressor bypass valve 54 is passively mechanically or actively multi-stage or infinitely controllable. For example, the second compressor bypass valve 54 is designed as a throttle.

Optionally, a third charge air cooler 56 is additionally arranged in the second compressor bypass duct 46. The third intercooler 56 is arranged and designed so that it is only flowed through by that part of the charge air from the first compressor 36, which flows via the second compressor bypass passage 46. Furthermore, optionally downstream of the first compressor 36 and upstream of the second compressor 38, a first charge air cooler 57 is arranged. The first charge air cooler 57 is arranged and configured such that the entire charge air coming from the first compressor 36 flows through this first charge air cooler 57. For this purpose, the first charge air cooler 57 in the fresh air passage 44 between the first compressor 36 and the second compressor 38 upstream or downstream of the branch of the second compressor bypass passage 46 is arranged.

Upstream of the second turbine 24 branches off from the exhaust passage 22, an exhaust gas recirculation line 58 for high pressure recirculating exhaust gas (HD-EGR line) and opens downstream of the second intercooler 40 and downstream of the throttle valve 41 and upstream of

Intake manifold 42 in the fresh air duct 44 of the fresh air tract 12 a. In the HP EGR passage 58, there is disposed a recirculated exhaust gas cooler 60 (HD EGR cooler) and a recirculated exhaust gas valve 62 (HP EGR valve).

Optionally, an exhaust gas recirculation line 64 for low pressure recirculating exhaust gas (ND-EGR line) is provided, which branches off from the exhaust passage 22 of the exhaust tract 14 downstream of the first turbine 26 and flows upstream of the first compressor 36 into the fresh air duct 44 of the fresh air tract 12. This LP EGR line 64 may also have an exhaust gas cooler (LP EGR cooler, not shown) and a valve (LP EGR valve, not shown) for the exhaust gas recirculated under low pressure. Alternatively, the LP EGR line 64 enters the HP EGR cooler 60 of the HD EGR line 58.

Still optionally, an exhaust gas recirculation line 66 for be recirculated under medium ^"pressure exhaust (MD-EGR pipe) is provided which is upstream of the exhaust passage 22 of the exhaust section 14 of the first turbine 26, and branches off downstream of the second turbine 24 and downstream of the first compressor 36 and upstream of the second compressor 38 in the fresh air line 44th the fresh air tract 12 opens. This MD EGR line 64 may also have an exhaust gas cooler (MD EGR cooler, not shown) and a valve (MD EGR valve, not shown) for the exhaust gas recirculated under medium pressure.

The AGR-F * fade 58, 64, 66 are optionally provided with an optionally switchable air or water cooling, the water cooling, for example, from a separate

Low-temperature circuit or is supplied with coolant from the engine circuit.

1, both the first turbine 26 and the second turbine 24 are each equipped with an adjustable turbine geometry (VTG). Simultaneous is not at both turbines 24, 26

Bypass channel and no wastegate provided. This at least simplifies the system

Exhaust side to only two stops for the VTGs. The function of a bypass channel or

A wastegate is taken over by setting the VTG accordingly. It is merely a compressor-side bypass in the form of the second compressor bypass passage 46 is provided, which is passively or alternatively actively controlled with respect to the second compressor bypass valve 54. Due to the missing in particular at the HD turbine 24

Turbine bypass channel is to avoid overspeed of the HD-ATL 18

For example, only a specific power of less than or equal to 70 kW / dm ³ , 65 kW / dm ³ , 60 kW / dm ³ or 55 kW / dm ³ realized. Thus, the system comes off the exhaust side with 2 actuators for the two VTG's.

The pressure gradient between the exhaust tract 14 and the fresh air tract 12 is increased alternatively or additionally to the intake throttle by means of the throttle valve 41 by adjusting the VTG of the HP turbine 24 optimally by adjusting the exhaust gas back pressure and additionally adjusting the VTG of the LP turbine 26 by adjusting the exhaust back pressure. Thus, the rate or the mass flow for recirculating exhaust gas (EGR rate) can be significantly increased. In the prior art engines of 1, 85 I to 2.05 I cubic capacity today in the map range 1500 min ^"1 to 2000 min ^{" 1} and 50 Nm to 150 Nm usually to> 60% of the range exhaust gas recirculation rates of> 35% usual. The latter value is improved in the internal combustion engine according to the invention to> 40%, in particular> 45 to 50%. To fully exploit the full potential of the system, the power of the EGR cooler 40, 56, 57 is increased. At an operating point of 2,000 min ^-1 and 150 Nm, an EGR cooler typically has one

Cooling capacity of about 5 kW. This corresponds to approximately 3.5% to 4% of the rated motor power or approximately 6% to 6.5% of the maximum motor power at 2,000 rpm ^{"1. In this} case, the EGR cooling capacity at this operating point is set to greater than or equal to 4%. , 4.6%, 5%, 6%, 8%, 10%, 13%, 18% or 25% of the rated engine power, which is technically achieved, for example, by a corresponding volume adjustment of the EGR cooler in downhill grades support of the service brake by the engine braking effect of importance. With the inventive two-stage charging with VTG can also in

Brake be set to maximize the exhaust back pressure, so that can be dispensed with a separate exhaust flap to support the engine brake.

Preferably, the two-stage charging is used in vehicles that the

Emission level EU5, EU6 or EURO-V, EURO-VI. As a result of the high possible EGR rates, a very effective reduction of the NO _x raw emissions is achieved, the charging is preferably used in vehicles in which the cumulative NO _x tailpipe emissions in the dynamometer test by less than 40%, 30%, 20% or 10% are less than the cumulative NO _x raw emissions of the engine, ie they do not have a highly active NO _x exhaust aftertreatment and in particular no SCR system (SCR: Selective

Catalytic reduction of nitrogen oxides).

Optionally, the turbines 24, 26 are arranged in a common housing. Likewise, it is optionally provided that the compressors 36, 38 are arranged in a common housing, if no intercooling (intercooler 57) is provided or the turbines 24, 26 are not already arranged in a common housing. This saves costs and reduces the space requirement of the charging system.

The intercoolers 40, 56, 57 described above and the exhaust gas recirculation lines 58, 64, 66 are preferred developments of an internal combustion engine according to the invention, as described below with reference to FIG. 9.

In Fig. 2, a second known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1, so that reference is made to the explanation of the above description of FIG. With 70 of the working cylinder, the intake manifold and the exhaust manifold having engine block is designated. In contrast to the first known embodiment according to FIG. 1, in the second known embodiment according to FIG. 2 only the HP turbine 24 has a VTG, as indicated by arrow 68, but not the LP turbine 26. Furthermore, a second turbine bypass channel 72 (HD turbine bypass duct) with second turbine bypass valve 74 (HP turbine bypass valve) and a first wastegate 76 (LP wastegate) with first wastegate valve 78 (LP wastegate valve) provided. The HD turbine bypass passage 72 optionally bypasses the first turbine 26 of the first exhaust gas turbocharger 16. The HP turbine bypass valve 74 is actively pneumatically driven, for example, so that it selectively opens or closes the first exhaust gas bypass passage 28. The ND wastegate 76 optionally bridges the first turbine. The LP wastegate valve 78 is, for example, pneumatically operated and with a bearing feedback The ND wastegate valve 78 selectively opens or closes the second exhaust bypass passage 32. The second compressor bypass valve 54 is controlled, for example mechanically or pneumatically.

In Fig. 3, a third known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in FIGS. 1 and 2, so that reference is made to the explanation of the above description of FIGS. 1 and 2. in the

In contrast to the second known embodiment according to FIG. 2, the third known embodiment according to FIG. 3 has no HD turbine bypass duct 72 and no second compressor bypass duct 46. This results in a cost-effective variant, which nevertheless complies with all the required pollutant limits with respect to the exhaust gas. This third known embodiment has the same below a motor speed of 3,000 min ^"1

Torque curve as the second known embodiment. Since the compressor and turbine bypass damper are missing, the nominal power range is driven purely in two stages. This provides a very cost-effective variant for internal combustion engines with a rated power of less than or equal to 150 kW available. The exhaust gas-relevant area does not change with respect to the second preferred embodiment. This system is inexpensive, almost application-neutral and EU6-capable.

4, a fourth known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1, 2 and 3, so that reference is made to the explanation of the above description of FIGS. 1, 2 and 3 becomes. In contrast to the third known embodiment according to FIG. 3, the fourth known embodiment according to FIG. 4 has no LP wastegate 76 on the LP turbine 26. This system has the same characteristics as the third known embodiment. Since the boost pressure can not be reduced due to the lack of a wastegate, the power potential is below 140 kW. The torque curve below 3.000 min ^"1 , as well as the EU6 capability corresponds to the second embodiment.

In Fig. 5, a fifth known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1 to 4, so that reference is made to the explanation of the above description of FIGS. 1 to 4. in the

In contrast to the second known embodiment according to FIG. 2, the fifth embodiment according to FIG. 5 also has a VTG in the LP turbine 26, as indicated by arrow 68. The fifth embodiment provides the most thermodynamic

Degrees of freedom through the VTG 68 at the ND turbine 26 and the HD turbine 24. With this system, the rated power compared to the known embodiments, two to four be increased significantly. In addition, the exhaust gas potential is higher than in the second known embodiment.

In Fig. 6, a sixth known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1 to 5, so that reference is made to the explanation of the above description of FIGS. 1 to 5. In contrast to the third known embodiment according to FIG. 3, the sixth known embodiment according to FIG. 6 also has a VTG in the LP turbine 26, as indicated by arrow 68. The sixth known embodiment achieves a cost advantage over the fifth known embodiment because there is no second compressor bypass passage 46 and no second turbine bypass passage 72 (HD turbine bypass passage).

In Fig. 7, a seventh known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1 to 6, so that reference is made to the explanation of the above description of FIGS. 1 to 6. in the

In contrast to the fifth known embodiment according to FIG. 5, the seventh known embodiment according to FIG. 7 has a VTG only in the LP turbine 26, as indicated by arrow 68. The HD turbine 24, however, is formed without VTG. This achieves a cost advantage over the fifth known embodiment.

In Fig. 8, an eighth known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1 to 7, so that reference is made to the explanation of the above description of FIGS. 1 to 7. in the

In contrast to the sixth known embodiment according to FIG. 6, the eighth known embodiment according to FIG. 8 has a VTG only in the LP turbine 26, as indicated by arrow 68. The HD turbine 24, however, is formed without VTG. This achieves a cost advantage over the sixth embodiment.

In Fig. 9, a preferred embodiment of an internal combustion engine according to the invention is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1 to 8, so that reference is made to the explanation of the above description of FIGS. 1 to 8. In contrast to the fifth known embodiment according to FIG. 5, the ninth preferred embodiment of the internal combustion engine according to the invention according to FIG. 9 additionally has a second wastegate 80 with a second wastegate valve 82.

In Fig. 10, a tenth known embodiment of an internal combustion engine is shown, wherein functionally identical parts are designated by the same reference numerals, as in Fig. 1 to 9, so that reference is made to the explanation of the above description of FIGS. 1 to 9. in the In contrast to the fourth known embodiment according to FIG. 4, the tenth known embodiment according to FIG. 10 has a VTG only at the LP turbine 26, as indicated by arrow 68. The HD turbine 24, however, is formed without VTG. This achieves a cost advantage over the fourth known embodiment.

In the previously described embodiments without a bypass channel or wastegate on the HP turbine 24, a different choice of material is advantageous in the case of separate design of the turbine housing. At very high specific power of the internal combustion engine, for example, greater than or equal to 60 kW / dm ³ , 70 kW / dm ³ , 75 kW / dm ³ , 80 kW / dm ³ , 90 kW / dm ³ or 100 kW / dm ³ is at the HD turbine 24 a high-temperature resistant material, such as

For example, D-5S (spheroidal graphite austenitic cast iron, EN-GJSA-XNiSiCr35-5-2, EN-JS3061; microstructure: austenitic matrix with chromium carbide and nodular graphite) is preferred, while ND turbine 26 is always at a lower temperature is applied, with a less high temperature resistant, cost-effective material, such as SiMoCr, can be equipped.

Vehicles with dual-supercharged, self-igniting engine according to the present invention, when operating in the New European Driving Cycle, have NO _x gross emissions of less than 300 mg / km, 280 mg / km, 235 mg / km, 200 mg / km, 180 mg / km or

160 mg / km and at the same time the particulate matter emissions exceed 35 mg / km, 40 mg / km, 45 mg / km, 50 mg / km, 55 mg / km, 60 mg / km, 80 mg / km or 100 mg / km not (fulfillment EU5).

Motor vehicles with double-charged, self-igniting engine according to the present

When operating in the New European Driving Cycle, NO _x emissions of less than 140 mg / km, 120 mg / km, 100 mg / km, 80 mg / km, 60 mg / km or 40 mg / km exceed and simultaneously exceed Particulate matter emissions not exceeding 35 mg / km, 40 mg / km, 45 mg / km, 50 mg / km, 55 mg / km, 60 mg / km, 80 mg / km or 100 mg / km (EU6 compliance).

11, a rotational speed in [min ^"1 ] is plotted on a horizontal axis 110, and a mean pressure pme in [bar] is plotted on a vertical axis 112. In the graph, a first emission region 114 (MVEG region I) and a second emission region 116 (MVEG region II) .The first emission region 114 extends for all rotational speeds 110 to about 6 bar pme 112 and the second emission region 116 extends for all rotational speeds 110 from about 6 bar pme 12 and higher explained first to tenth embodiment with a first conventional variant of an internal combustion engine with only one exhaust gas turbocharger with VTG (monoturbo ATL with VTG, not shown) and with a second conventional variant of an internal combustion engine with two-stage supercharging but without VTG on both turbines (Biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate, not shown) compared.

In FIG. 12, a NO _x flow rate in [gib] is plotted on a horizontal axis 1 18, and a soot flow rate in [g / h] is plotted on a vertical axis 120. 122 is an EU 6 area and 124 is an EU 5 area. A dashed line 126 shows a curve of the ΝΟχ-throughput 1 18 for various EGR rates and in an operating condition of the internal combustion engine in the first emission range 1 14 with 1,500 min ^"1 and with a

Medium pressure 1 12 pme = 3 bar. As the EGR rate increases, the NO _x flow rate decreases 1 18. In the first emission range 14, there is no conventional NO _x soot trade. The

Emission potential of the internal combustion engine, i. minimum NOx emissions at compatible particulate emissions is only dependent on the EGR rate 126. The maximum possible EGR rate 126 is largely determined by the charging system.

In FIG. 13, a fresh air mass in [g / stroke] is plotted on a horizontal axis 128 and a boost pressure in [mbar] is plotted on a vertical axis 130. A first graph 132 shows a plot of the boost pressure 130 over the fresh air mass 128 without EGR for all variants, i. for the first and second conventional embodiment and for the first to tenth known embodiment of FIGS. 1 to 8 and 10 and the preferred embodiment of the internal combustion engine according to the invention shown in FIG. 9. Furthermore, a second graph 134 illustrates the course of the boost pressure 130 via the fresh air mass 128th with EGR for the first conventional embodiment of the internal combustion engine (monoturbo ATL with VTG), a third graph 136 illustrates the course of the boost pressure 130 via the fresh air mass 128 with EGR for the second conventional embodiment of the internal combustion engine (biturbo with HD-ATL as a fixed loader and ND ATL as wastegate) and a fourth graph 138

FIG. 13 illustrates an operating state of the respective internal combustion engine within the first emission region 14 (see FIG. 11) 1,500 rpm ^"1 and with a mean pressure of 1 12 pme = 3 bar. To generate the graphs 132, 134, 136 and 138, in the first conventional embodiment of the internal combustion engine and the preferred one

Embodiment of the internal combustion engine according to the invention shown in FIG. 9, the VTG initially closed without EGR stepwise. In the second conventional embodiment of the internal combustion engine (biturbo without VTG), the wastegate valve of the LP-ATL is gradually closed. This results in all embodiments, first, the first graph 132, depending on the embodiment with this mode of operation different high boost pressures can be achieved. The maximum boost pressure of the second conventional embodiment of the Internal combustion engine is 1,350 mbar at point 140. The maximum boost pressure of the preferred embodiment of the internal combustion engine according to the invention according to FIG. 9 is 1,620 mbar at point 142. The maximum charge pressure of the first conventional

Embodiment of the internal combustion engine is 1,350 mbar at point 144. In the

The respective maximum boost pressure of the respective embodiment, the EGR valve is then opened stepwise, so that the graphs 134 (monoturbo with VTG), 136 (Biturbo) and 138 (internal combustion engine according to FIG. 9: Biturbo with VTG) result. With the same fresh air mass, in the preferred embodiment of the internal combustion engine according to the invention according to FIG. 9, a higher charge pressure (graph 138) is achieved in each case than in the two conventional embodiments (graphs 134 and 136). At 145, the maximum EGR rate is reached. This result will be described below with reference to FIG.

14 to 17 analyzed in more detail.

In FIGS. 14 to 17, similarly to FIG. 13, reference numeral 34 denotes the respective result for the first conventional internal combustion engine embodiment (VTG monoturbo ATL), and reference numeral 136 denotes the respective result for the second conventional one

Embodiment of the internal combustion engine (Biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate) and the reference numeral 138, the respective result for the preferred

Embodiment of the internal combustion engine according to the invention shown in FIG. 9. In Figs. 14 and

15 is plotted on a vertical axis 146, an exhaust back pressure before turbine in [mbar]. In Fig. 16, on a vertical axis 148, an EGR rate in [%] and in Fig. 17, NO _x flow rate in [g / h] is plotted on a vertical axis 150. Fig. 14 illustrates the exhaust pressure before turbine without EGR, Fig. 15 illustrates the exhaust pressure before turbine with EGR, Fig. 16 illustrates a maximum EGR rate and Fig. 17 illustrates an achievable minimum NO _x emission. As can be seen from FIG. 14, the preferred embodiment of the internal combustion engine according to the invention according to FIG

Series connection of the turbocharger and the VTG technology higher back pressure than the first conventional embodiment of the internal combustion engine (monoturbo ATL with VTG) and the second conventional embodiment of the internal combustion engine (biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate). When the EGR valve is almost fully opened, pressure equalization occurs between the exhaust manifold and the intake manifold. This is the limiting amount for the maximum EGR rate as shown in FIG. The maximum EGR rate for the second conventional embodiment of the internal combustion engine (bar graph 136, biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate) is 45%. The maximum EGR rate for the preferred embodiment of the internal combustion engine according to the invention according to FIG. 9 (bar graph 138) is 68%. The maximum EGR rate for the first conventional embodiment of the internal combustion engine (bar graph 134, monoturbo ATL with VTG) is 58%. 12 results in the NO _x potential, which is shown in Fig. 17, The NO _x emission for the second conventional embodiment of the

Internal combustion engine (bar graph 136, biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate) is about 6 g / h. The NO _x emission for the preferred embodiment of the internal combustion engine according to the invention as shown in FIG. 9 (bar graph 138) is about 1 g / h. The NO _x emission for the first conventional embodiment of the internal combustion engine

(Bar graph 134, monoturbo ATL with VTG) is about 4 g / h.

Fig. 18 shows a boost pressure variation analogous to Fig. 13, but with the respective

Internal combustion engine (first conventional embodiment, second conventional

Embodiment or preferred embodiment of FIG. 9) in the second

Emission range 16 (see Fig. 1 1) with 2,000 min ^"1 and medium pressure 1 12 pme = 8 bar. The same reference numerals have the same meaning as in Fig. 13, so that their explanation to the above description of FIG 13. The mode of operation for generating graphs 132, 134, 136 and 138 is analogous to that described above with reference to Fig. 13. The maximum boost pressure of the second conventional embodiment of the internal combustion engine is 1850 mbar at point 140. The maximum charge pressure of Preferred embodiment of the internal combustion engine according to the invention according to FIG. 9 is 2,150 mbar at point 142. The maximum boost pressure of the first conventional

Embodiment of the internal combustion engine is 1.50 mbar at point 144. At maximum boost pressure, then, as in Fig. 13, the EGR valve is then opened stepwise. With the preferred embodiment of the internal combustion engine according to the invention according to FIG. 9, a higher charge pressure is achieved with a simultaneously higher EGR rate. This result will be further analyzed below with reference to FIG.

19 shows a representation of values for the soot flow rate and NOx throughput analogous to FIG. 12. The same reference numbers have the same meaning, so that reference is made to the above description of FIG. 12 for explanation thereof. The results for the first conventional embodiment of the internal combustion engine (monoturbo ATL with VTG) are shown in the second graph 134. The results for the second conventional embodiment of the internal combustion engine (biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate) are shown in the third graph 136. The results for the preferred embodiment of the internal combustion engine according to the invention shown in FIG. 9, the fourth graph 38. From Fig. 19 it can be seen that for all three

compared systems is a NOx / soot trade. At the second

conventional embodiment (Biturbo with HD-ATL as a fixed loader and ND-ATL as

Wastegate) high EGR rates are achieved only with low boost pressure, which leads to a high soot emissions with the same ΝΟχ-emission. The first conventional embodiment of the internal combustion engine (monoturbo ATL with VTG) has a better NO _x rate trade because of the higher boost pressures. The preferred embodiment of the invention Internal combustion engine according to FIG. 9 has the highest emission potential in each case because of the step charging in combination with the VTG technology. Here very high EGR rates can be driven with high boost pressures. The high boost pressure ensures a high air mass in the cylinder. The injected fuel mass in this case must heat a higher air mass than in the first and second conventional embodiment of the

Internal combustion engine. This results in a lower maximum combustion temperature and lower NO _x emissions. Despite the high EGR rate, there is more oxygen available for soot oxidation compared to the first and second conventional embodiments. This leads to NO _x and soot emissions at EU6 level.

FIG. 20 again shows a boost pressure variation analogous to FIGS. 13 and 8, but with the respective internal combustion engine (first conventional embodiment, second conventional embodiment or preferred embodiment according to FIG. 9) in the second

112 pme = 8.9 is operated bar-emitting region 116 (see FIG. FIG. 11) of 2.000 min ^"1 and medium pressure. The same reference numerals have the same meaning as in Fig. 13 and 18, so that to the explanation on the above description 13 and 18. Reference is made to FIGS

Operation for generating the graphs 132, 134, 136 and 138 is analogous to that described above with reference to Figs. 13 and 8, respectively. In Fig. 20, in addition, lines 152 of equal EGR rate for EGR rates of 0%, 5%, 10%, 15%, 20%, 25%, 30%, 35%, 40%, 45%, 50% and 55% drawn. The results shown in FIG. 20 clearly show that with the preferred embodiment of the internal combustion engine according to the invention according to FIG. 9, higher EGR rates can be represented in this operating range. It can also be seen from Fig. 20 that at the same EGR rate, a higher fresh air content, i. more oxygen is present in the combustion chamber.

21 shows a representation of values for the flow rate of soot and NOx throughput analogous to FIGS. 12 and 19. The same reference numerals have the same meaning, so that their

Explanation is made to the description of Figs. 12 and 19. The results for the first conventional embodiment of the internal combustion engine (monoturbo ATL with VTG) are shown in the second graph 134. The results for the second conventional embodiment of the internal combustion engine (biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate) are shown in the third graph 136. The results for the preferred embodiment of the internal combustion engine according to the invention according to FIG. 9 are shown in the fourth graph 138. In FIG. 21, in addition, lines 152 of equal EGR rate for EGR rates of 0%, 5%, 10%, 15%, 20%, 25%, 30%, 35%, 40%, 45%, 50% and 55% respectively. FIG. 21 graphically illustrates the physical effects explained with reference to FIG. 20.

Claims

claims

1. internal combustion engine, in particular of a motor vehicle, with a fresh air tract (12) for supplying fresh air to working cylinder (10) of the internal combustion engine, an exhaust tract (14) for discharging exhaust gas (21) from the working cylinders (10), a first exhaust gas turbocharger (16 ) a low pressure stage (LP exhaust gas turbocharger), which in the exhaust tract (14) arranged first turbine (26) (LP turbine) and in the fresh air tract (12) arranged compressor (ND compressor), and

at least one second exhaust gas turbocharger (18) of a high-pressure (HD) exhaust gas turbocharger, which in the exhaust tract (14) upstream of the first turbine (26) arranged second turbine (HD turbine) and one in the fresh air tract (12) downstream of the first Compressor (36) arranged second compressor (38) (HP compressor), wherein both turbines (24, 26) have an adjustable turbine geometry (68) (VTG), characterized in that the exhaust tract (14, 22) downstream of the first Turbine (26) via the first turbine (26) and in addition via a first wastegate (76) with first wastegate valve (78) with the exhaust gas tract (14, 22) upstream of the first turbine (26) and downstream of the second turbine (24) fluid-conducting the exhaust tract (14, 22) downstream of the second turbine (24) and upstream of the first turbine (26) via the second turbine (24) and additionally via a second turbine bypass passage (72) with the second turbine bypass passage valve (74) and a second Wastegate e (80) with second wastegate valve (82) with the exhaust tract (14, 22) upstream of the second turbine (24) fluidly connected, wherein the fresh air tract (12, 44) upstream of the first

Compressor (36) via the first compressor (36) with the fresh air tract (12, 44) downstream of the first compressor (36) and upstream of the second compressor (38) is fluidly connected, wherein the fresh air tract (12, 44) upstream of the second

Compressor (38) and downstream of the first compressor (36) via the second compressor (38) and additionally via a second compressor bypass passage (46) with second

Compressor bypass duct valve (54) to the fresh air duct (12, 44) downstream of the second compressor (38) is connected fluidly, wherein downstream of the second compressor (38) a second charge air cooler (40) is arranged upstream of the second turbine (24) has a high pressure Off-gas return line (58) (HD-EGR line) branches off from the exhaust tract (14, 22) and flows downstream of the second compressor (38) in the fresh air tract (12, 44), wherein the HP-EGR line (58) downstream of the second charge air cooler (40) into the fresh air tract (12, 44), wherein in the HP-EGR line (58) a cooler (60) for recirculated exhaust gas (HP-EGR cooler) is arranged, wherein in the HP-EGR line (58) a valve (62) for recirculated exhaust gas (HP-EGR valve) is arranged.

2. Internal combustion engine according to claim 1, characterized in that in the second

Compressor bypass passage (46) a third intercooler (56) arrangedjst.

3. Internal combustion engine according to at least one of the preceding claims, characterized in that downstream of the first compressor (36) and upstream of the second compressor (38), a first charge air cooler (57) is arranged.

4. Internal combustion engine according to claim 1, characterized in that downstream of the first turbine (26) a low-pressure exhaust gas recirculation line (64) (ND-EGR line) branches off from the exhaust tract (14, 22) and into the HP-EGR cooler ( 60) opens.

5. Internal combustion engine according to at least one of the preceding claims, characterized in that downstream of the first turbine (26) a low-pressure exhaust gas recirculation line (64) (ND-EGR line) branches off from the exhaust gas tract (14, 22) and upstream of the first compressor ( 36) into the fresh air tract (12, 44) opens.

6. Internal combustion engine according to claim 5, characterized in that in the LP EGR line (64), a recirculated exhaust gas cooler (LP EGR cooler) and / or a valve for recirculated exhaust gas (LP EGR valve) is arranged ,

7. Internal combustion engine according to at least one of the preceding claims, characterized in that upstream of the first turbine (26) and downstream of the second turbine (24) has a medium-pressure exhaust gas recirculation line (66) (MD-EGR line) of the exhaust gas tract (124, 22 ) branches off and downstream of the first compressor (36) and upstream of the second compressor (38) opens into the fresh air tract (12, 44).

8. Internal combustion engine according to claim 7, characterized in that in the MD-AGR line (66) a recirculated exhaust gas cooler (MD-EGR cooler) and / or a valve for recirculated exhaust gas (MD-EGR valve) is arranged ,

9. Internal combustion engine according to at least one of the preceding claims, characterized in that the internal combustion engine has at least 3, in particular 4, 5, 6, 8, 10 or 12 working cylinder (10).

10. Internal combustion engine according to at least one of the preceding claims, characterized in that the internal combustion engine has a direct fuel injection into at least one working cylinder (10), in particular according to the common rail system.

1 1. Internal combustion engine according to at least one of the preceding claims, characterized in that the internal combustion engine has a rated speed of at least 3,000 revolutions per minute, in particular 3,500, 4,000, 4,500, 5,000 or 5,000

Revolutions per minute.

12. Internal combustion engine according to at least one of the preceding claims, characterized in that at least one working cylinder (10) has a displacement of less than or equal to 800 cc, in particular less than or equal to 700 cc, 600 cc, 500 cc, 400 cc or 350 cc.

13. Internal combustion engine according to at least one of the preceding claims, characterized in that at least one working cylinder (10) at least one, in particular two or more, exhaust valves are assigned.

14. Internal combustion engine according to at least one of the preceding claims, characterized in that at least one working cylinder (10) at least one, in particular two or more, inlet valves are assigned.

15. Internal combustion engine according to at least one of the preceding claims, characterized in that at least one first and second compressor (36, 38) are arranged in a common housing.