DE102010010480A1 - Internal combustion engine with two-stage supercharging - Google Patents

Abstract

The invention relates to an internal combustion engine, in particular a motor vehicle, with a fresh air duct (12) for supplying fresh air to the working cylinder (10) of the internal combustion engine, an exhaust system (14) for discharging exhaust gas (21) from the working cylinders (10), and a first exhaust gas turbocharger (16) a low pressure stage (LP exhaust gas turbocharger), which has a first turbine (26) (LP turbine) arranged in the exhaust gas tract (14) and a compressor (36) (LP compressor) arranged in the fresh air tract (12), and at least one second exhaust gas turbocharger (18) of a high pressure stage (HP exhaust gas turbocharger), which has a second turbine (24) (HP turbine) arranged in the exhaust gas tract (14) upstream of the first turbine (26) and one in the fresh air tract (12) downstream of the first compressor (36) arranged second compressor (38) (HP compressor). At least one of the turbines (24, 26) has an adjustable turbine geometry (68) (VTG).

Description

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, 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. In order to reduce emissions, the highest possible EGR rate is generally 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 the DE 10 2007 062 366 A1 is a generic internal combustion engine with two-stage charging known.

From the DE 198 51 028 C2 a method for operating a supercharged internal combustion engine is known which comprises 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 parallel turbochargers are each formed with a variable turbine geometry.

The invention is based on the object, an internal combustion engine o. G. To improve the type of emission and consumption and to simplify the system complexity.

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 according to the invention provided that at least one of the turbines has an adjustable turbine geometry (VTG).

This has the advantage that an internal combustion engine with a simplified supercharging system with a reduced number of actuators is available, as can be dispensed with that turbine with VTG on a bypass channel with bypass valve and / or on a waiting gate with wastegate valve, as this function by a corresponding adjustment VTG can be realized. At the same time and without an additional exhaust gas flap, a maximum possible exhaust gas recirculation rate (EGR rate) can be realized since an exhaust back pressure can be increased correspondingly by means of the VTG, so that an extended characteristic field range results, in which the pressure in the exhaust gas tract at the removal point for the exhaust gas to be recirculated is greater as the pressure in the fresh air tract at the discharge point for the 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.

A particularly effective reduction of the number of actuators in the two-stage supercharging system is achieved in that the exhaust tract downstream of the first turbine exclusively via the first turbine with the exhaust gas upstream of the first turbine and downstream of the second turbine fluidly connected.

An extension of the controllability of the two-stage supercharging system, in particular with regard to the boost pressure is achieved in that the exhaust tract upstream of the first turbine via the first turbine and additionally via a first turbine bypass passage with first turbine bypass passage valve and / or a first wastegate with first wastegate valve with the exhaust gas upstream of the first Turbine and downstream of the second turbine fluidly connected.

A particularly effective reduction in the number of actuators in the two-stage supercharging system is achieved by virtue of the fact that the exhaust tract downstream of the second turbine and upstream of the first turbine is fluid-conductively connected via the second turbine to the exhaust tract upstream of the second turbine.

An extension of the controllability of the two-stage supercharging system, in particular with regard to the boost pressure is achieved in that 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 passage with second turbine bypass passage valve and / or a second wastegate with second wastegate valve with the exhaust tract upstream of the second turbine is fluidly connected.

A particularly effective reduction of the number of actuators in the two-stage supercharging system is achieved by virtue of the fact that the fresh air tract upstream of the first compressor is fluid-conductively connected via the first compressor to the fresh air tract downstream of the first compressor and upstream of the second compressor.

An extension of the controllability of the two-stage supercharging system, in particular with regard to the boost pressure achieved by the fresh air tract upstream of the first compressor via the first compressor and in addition via a first compressor bypass passage with first compressor bypass passage valve fluidly connected to the fresh air tract downstream of the first compressor and upstream of the second compressor ,

A particularly effective reduction of the number of actuators in the two-stage supercharging system is achieved in that the fresh air tract upstream of the second compressor and downstream of the first compressor is fluidly connected exclusively via the second compressor with the fresh air tract downstream of the second compressor.

An extension of the controllability of the two-stage supercharging system, in particular with regard to the boost pressure achieved by the fact that the fresh air duct upstream of the second compressor and downstream of the first compressor via the second compressor and additionally via a second compressor bypass passage with second compressor bypass passage valve is fluidly connected downstream of the second compressor with the fresh air ,

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

A particularly good and direct regulation of the exhaust backpressure as well as an engine braking effect by means of the VTG is achieved by the fact that only the second turbine has a VTG.

An inexpensive charging system is achieved by the fact that only the first turbine has a VTG.

A further enlarged range of rules with regard to the exhaust backpressure is achieved in that the first and the second turbine each have a VTG.

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

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

To avoid contamination of the second intercooler by recirculated exhaust gas, the HD-EGR line flows into the fresh air tract downstream of the second intercooler.

For cooling the recirculated exhaust gas, a recirculated exhaust gas cooler (HP-EGR cooler) is arranged in the HP-EGR passage.

To influence the amount of recirculated exhaust gas, a valve for recirculated exhaust gas (HP-EGR valve) is arranged in the HP-EGR line.

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

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 line (MD-EGR line) branches off from 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, a recirculated exhaust gas cooler (MD-EGR cooler) and / or a recirculated exhaust gas valve (MD-EGR valve) are provided in the MD-EGR passage. arranged.

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.

An internal combustion engine adapted to a desired level of performance and torque level 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

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

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

3 A third preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

4 A fourth preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

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

6 A sixth preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

7 A seventh preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

8th an eighth preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

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

10 A tenth preferred embodiment of an internal combustion engine according to the invention in a schematic representation,

11 a graphical representation of emission ranges of the internal combustion engine,

12 a graphical representation of soot flow rate and NO _x flow rate at an operating condition of the internal combustion engine with 1500 min ^-1 and with a mean pressure pme = 3 bar,

13 a graphical representation of boost pressure variation in an operating state of the internal combustion engine with 1,500 min ^-1 and with a medium pressure pme = 3 bar,

14 a graphical representation of an exhaust gas pressure before turbine without EGR in an operating condition of the internal combustion engine with 1500 min ^-1 and with a mean pressure pme = 3 bar,

15 a graphical representation of an exhaust gas pressure before turbine with EGR in an operating condition of the internal combustion engine with 1,500, min ^-1 and with a mean pressure pme = 3 bar,

16 a graphical representation of a maximum achievable EGR rate in an operating condition of the internal combustion engine with 1,500 min ^-1 and with a mean pressure pme = 3 bar,

17 a graphical representation of a minimum achievable NO _x emission at an operating condition of the internal combustion engine with 1500 min ^-1 and with a mean pressure pme = 3 bar,

18 a graphical representation of a boost pressure variation in an operating condition of the internal combustion engine with 2,000 min ^-1 and with a mean pressure pme = 8 bar, 19 a graphical representation of soot flow rate and NO _x throughput in an operating condition of the internal combustion engine with 2,000 min ^-1 and with a mean pressure pme = 8 bar,

20 a graphical representation of a boost pressure variation in an operating condition of the internal combustion engine with 2,000 min ^-1 and with a mean pressure pme = 8.9 bar and

21 a graphical representation of soot flow and NO _x throughput at an operating condition of the internal combustion engine with 2,000 min ^-1 and with a mean pressure pme = 8.9 bar.

In the 1 shown, the first preferred embodiment of an internal combustion engine according to the invention comprises working cylinder 10 , a fresh air tract 12 , an exhaust tract 14 , a first exhaust gas turbocharger 16 (ND-ATL) 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 from the working cylinders 10 discharged exhaust gas 21 , and an exhaust duct 22 , In the exhaust duct 22 is 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 arranged.

The fresh air tract 12 includes a fresh air duct 44 in which, viewed in the direction of flow, the following is arranged: A first compressor 36 (LP compressor) of the first exhaust gas turbocharger 16 , a second compressor 38 (LP compressor) of the second exhaust gas turbocharger 18 , a second intercooler 40 , a throttle 41 and a suction tube 42 which is in the working cylinder 10 via inlet valves (not shown) opens. Furthermore, the fresh air tract 12 a second compressor bypass channel 46 on which the second compressor 38 the second exhaust gas turbocharger 18 bridged. The bypass channel 46 branches downstream of an exit 48 of the first compressor 36 of the first exhaust gas turbocharger 16 and upstream of an entrance 50 of the second compressor 38 the second exhaust gas turbocharger 18 from the fresh air duct 44 from and opens downstream of an outlet 52 of the second compressor 38 the second exhaust gas turbocharger 18 and upstream of the second intercooler 40 back into the fresh air duct 44 one. In the second compressor bypass channel 46 is a second compressor bypass valve 54 arranged. This second compressor bypass valve 54 is passively mechanically or actively multi-stage or steplessly controllable. For example, the second compressor bypass valve 54 designed as a throttle.

Optionally, in the second compressor bypass passage 46 in addition a third intercooler 56 arranged. The third intercooler 56 is arranged and formed so that this only from that part of the charge air from the first compressor 36 is flowed through, which via the second compressor bypass passage 46 flows. Further optional is downstream of the first compressor 36 and upstream of the second compressor 38 a first intercooler 57 arranged. The first intercooler 57 is arranged and formed such that the entire of the first compressor 36 Coming charge air through this first intercooler 57 flows. This is the first intercooler 57 in the fresh air duct 44 between the first compressor 36 and the second compressor 38 upstream or downstream of the branch of the second compressor bypass passage 46 arranged.

Upstream of the second turbine 24 branches from the exhaust duct 22 an exhaust gas recirculation line 58 for exhaust gas to be recirculated under high pressure (HP-EGR line) and discharges downstream of the second intercooler 40 and downstream of the throttle 41 and upstream of the suction pipe 42 in the fresh air channel 44 of the fresh air tract 12 one. In the HD AGR line 58 is a cooler 60 for the recirculated exhaust gas (HD-EGR cooler) and a valve 62 arranged for the recirculated exhaust gas (HP-EGR valve).

Optionally, an exhaust gas recirculation line 64 for low pressure recirculating exhaust (ND-EGR line) provided by the exhaust duct 22 the exhaust tract 14 downstream of the first turbine 26 branches off and upstream of the first compressor 36 in the fresh air channel 44 of the fresh air tract 12 empties. This LP EGR line 64 optionally also has 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 ND-AGR line leads 64 in the HD-EGR cooler 60 the HD-AGR line 58 one.

Further optional is an exhaust gas recirculation line 66 provided for under medium pressure exhaust gas (MD-EGR line), which from the exhaust duct 22 the exhaust tract 14 upstream of the first turbine 26 as well as downstream of the second turbine 24 branches off and downstream of the first compressor 36 and upstream of the second compressor 38 in the fresh air line 44 of 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 EGR paths 58 . 64 . 66 are optionally provided with an optionally switchable air or water cooling, wherein the water cooling is fed, for example, from a separate low-temperature circuit or with coolant from the engine circuit.

In the first embodiment of an internal combustion engine according to the invention 1 are both the first turbine 26 as well as the second turbine 24 each equipped with an adjustable turbine geometry (VTG). Simultaneous is with both turbines 24 . 26 No bypass channel and no wastegate provided. This simplifies the system at least on the exhaust side to only two actuators for the VTGs. The function of a bypass channel or a wastegate is adopted by setting the VTG accordingly. It is merely a compressor-side bypass in the form of the second compressor bypass duct 46 provided with respect to the second compressor bypass valve 54 passive or alternatively actively controllable executed. Due in particular to the HD turbine 24 Missing turbine bypass channel is used 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 fresh air tract 12 is used as an alternative or in addition to the suction-side throttling by means of the throttle valve 41 by adjustment of the VTG of the HP turbine optimized by exhaust backpressure 24 and additional exhaust gas back pressure optimized adjustment of the VTG of the LP turbine 26 increased. 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 l to 2.05 l displacement 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 complete the full potential of the system. exploit the performance of the or the EGR cooler 40 . 56 . 57 raised. At an operating point of 2.000 min ^-1 and 150 Nm has an EGR cooler usually to a cooling power of about 5 kW. This corresponds to approximately 3.5% to 4% of motor nominal power and about 6% to 6.5% of the maximum engine power at 2000 min ^-1. In the present case, the EGR cooling capacity is increased at this operating point to greater than or equal to 4%, 4.6%, 5%, 6%, 8%, 10%, 13%, 18% or 25% of the rated engine power, which technically for example by a corresponding volume adjustment of the EGR cooler takes place. For commercial vehicles, a support of the service brake by the engine braking effect is also important for downhill driving. With the two-stage turbocharger with VTG according to the invention, a maximization of the exhaust backpressure can also be set in the event of a brake so that a separate exhaust flap for supporting the engine brake can be dispensed with.

Preference is given to two-stage turbocharging on vehicles that meet the emission levels 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 preferred for 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 have no SCR system (SCR: Selective Catalytic Reduction of Nitrogen Oxides).

Optional are the turbines 24 . 26 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 not already arranged in a common housing. This saves costs and reduces the space requirement of the charging system.

The intercooler described above 40 . 56 . 57 and the exhaust gas recirculation lines 58 . 64 . 66 are not according to the first embodiment according to 1 limited, but may also be present in the other, described in more detail below alternative embodiment of an internal combustion engine according to the invention. Merely for the sake of clarity of presentation, these are not shown in the following figures and accordingly no longer mentioned.

In 2 a second 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 1 so that their explanation to the above description of 1 is referenced. With 70 is the engine block, the intake manifold and the exhaust manifold having designated engine block. Unlike the first embodiment according to 1 indicates in the second embodiment according to 2 only the HD turbine 24 a VTG on, as with arrow 68 is indicated, 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 duct 72 optionally bridges the first turbine 26 of the first exhaust gas turbocharger 16 , The HD turbine bypass valve 74 For example, it is actively pneumatically actuated so that it is the first exhaust bypass channel 28 optionally opens or closes. The LP wastegate 76 optionally bridges the first turbine. The LP wastegate valve 78 For example, is pneumatically operated and formed with a bearing feedback, the LP wastegate valve 78 the second exhaust bypass channel 32 optionally opens or closes. The second compressor bypass valve 54 is for example controlled mechanically or pneumatically.

In 3 a third 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 1 and 2 so that their explanation to the above description of 1 and 2 is referenced. In contrast to the second embodiment according to 2 shows the third embodiment according to 3 no HD turbine bypass duct 72 and no second compressor bypass channel 46 on. This results in a cost-effective variant, which nevertheless complies with all the required pollutant limits with respect to the exhaust gas. This third preferred embodiment has the same torque characteristics as the second preferred embodiment below an engine speed of 3,000 min ^-1. 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.

In 4 a fourth 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 1 . 2 and 3 so that their explanation to the above description of 1 . 2 and 3 is referenced. In contrast to the third embodiment according to 3 shows the fourth embodiment according to 4 no LP wastegate 76 at the LP turbine 26 on. This system has the same characteristics as the third 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 5 a fifth 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 1 to 4 so that their explanation to the above description of 1 to 4 is referenced. In contrast to the second embodiment according to 2 shows the fifth preferred embodiment according to 5 also with the ND turbine 26 a VTG on, as with arrow 68 is indicated. The fifth embodiment provides most thermodynamic degrees of freedom through the VTG 68 at the LP turbine 26 and the HD turbine 24 , With this system, the rated power can be significantly increased compared to the embodiments two to four. In addition, the exhaust gas potential is higher than in the second embodiment.

In 6 a sixth 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 1 to 5 so that their explanation to the above description of 1 to 5 is referenced. In contrast to the third embodiment according to 3 shows the sixth preferred embodiment according to 6 also with the ND turbine 26 a VTG on, as with arrow 68 is indicated. The sixth embodiment achieves a cost advantage over the fifth embodiment in that no second compressor bypass passage 46 and no second turbine bypass channel 72 (HD turbine bypass channel) is present.

In 7 a seventh 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 1 to 6 so that their explanation to the above description of 1 to 6 is referenced. In contrast to the fifth embodiment according to 5 shows the seventh preferred embodiment according to 7 only at the LP turbine 26 a VTG on, as with arrow 68 is indicated. The HD turbine 24 is trained without VTG. This achieves a cost advantage over the fifth embodiment.

In 8th an eighth 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 1 to 7 so that their explanation to the above description of 1 to 7 is referenced. In contrast to the sixth embodiment according to 6 shows the eighth preferred embodiment according to 8th only at the LP turbine 26 a VTG on, as with arrow 68 is indicated. The HD turbine 24 is trained without VTG. This achieves a cost advantage over the sixth embodiment.

In 9 a ninth 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 1 to 8th so that their explanation to the above description of 1 to 8th is referenced. In contrast to the fifth embodiment according to 5 indicates the ninth preferred embodiment according to 9 in addition a second wastegate 80 with second wastegate valve 82 on.

In 10 shows a tenth preferred embodiment of an internal combustion engine according to the invention, wherein functionally identical parts are designated by the same reference numerals, as in 1 to 9 so that their explanation to the above description of 1 to 9 is referenced. In contrast to the fourth embodiment according to 4 shows the tenth preferred embodiment according to 10 only at the LP turbine 26 a VTG on, as with arrow 68 is indicated. The HD turbine 24 is trained without VTG. This achieves a cost advantage over the fourth embodiment.

In the previously described embodiments without bypass channel or wastegate on the HP turbine 24 is a separate choice of the turbine housing a different choice of material advantageous. 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 HP turbine 24 a high-temperature resistant material such as D-5S (austenitic ductile iron; EN-GJSA-XNiSiCr35-5-2 ; No. EN-JS3061 ; Microstructure: austenitic matrix with chromium carbides and spheroidal graphite) is preferred, while the LP turbine 26 , which is always subjected to a lower temperature, can be equipped with a less high temperature resistant, cost-effective material, such as SiMoCr.

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).

Dual-loaded, self-igniting engine vehicles according to the present invention, when operated in the New European Driving Cycle, have NO _x gross emissions of less than 140 mg / km, 120 mg / km, 100 mg / km, 80 mg / km, 60 mg / km or 40 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 EU6).

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

In 12 is on a horizontal axis 118 NO _x flow rate in [g / h] and on a vertical axis 120 a carbon black throughput in [g / h] applied. With 122 is an EU 6 area and 124 is an EU 5 area. A dashed line 126 shows a course of NO _x throughput 118 for various EGR rates and in an operating state of the internal combustion engine in the first emission range 114 1,500 min ^-1 and with a medium pressure 112 pme = 3 bar. As the EGR rate increases, the NO _x flow rate decreases 118 , In the first emission area 114 There is no conventional NO _x soot trade. The emission potential of the internal combustion engine, ie minimum NOx emissions at compatible particulate emissions, is only of the EGR rate 126 dependent. The maximum possible EGR rate 126 is largely determined by the charging system.

In 13 is on a horizontal axis 128 a fresh air mass in [g / stroke] and on a vertical axis 130 a charge pressure in [mbar] applied. A first graph 132 shows one Course of the boost pressure 130 over the fresh air mass 128 without EGR for all variants, ie for the first and second conventional embodiment and for the first to tenth embodiment of the internal combustion engine according to the invention 1 to 10 , Furthermore, a second graph illustrates 134 the course of the boost pressure 130 over the fresh air mass 128 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 over 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 a wastegate) and a fourth graph 138 illustrates the course of the boost pressure 130 over the fresh air mass 128 with EGR for the first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10 , Also in 13 is an operating state of the respective internal combustion engine within the first emission range 114 (see. 11 ) with 1500 min ^-1 and with a medium pressure 112 pme = 3 bar. To generate the graphs 132 . 134 . 136 and 138 For example, in the first conventional embodiment of the internal combustion engine and the first to tenth embodiments of the internal combustion engine according to the invention, the VTG without EGR is initially closed step by step. 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, the first graph first 132 , Depending on the embodiment with this mode of operation different levels of charge pressures can be achieved. The maximum boost pressure of the second conventional embodiment of the internal combustion engine is 1,350 mbar at the point 140 , The maximum boost pressure of the first to tenth embodiments of the internal combustion engine according to the invention is 1,620 mbar at the point 142 , The maximum boost pressure of the first conventional embodiment of the internal combustion engine is 1,350 mbar at the point 144 , At the respective maximum boost pressure of the particular 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 the invention: biturbo with VTG) result. With the same fresh air mass is in each case a higher boost pressure with the first to tenth embodiment of the internal combustion engine according to the invention (graph 138 ) than in the two conventional embodiments (Graphs 134 and 136 ). at 145 in each case the maximum EGR rate is reached. This result will be described below with reference to FIGS 14 to 17 analyzed in more detail.

In the 14 to 17 refers analogously as in 13 the reference number 134 the respective result for the first conventional embodiment of the internal combustion engine (monoturbo ATL with VTG), the reference numeral 136 the respective result for the second conventional embodiment of the internal combustion engine (biturbo with HD-ATL as a fixed loader and ND-ATL as wastegate) and the reference numeral 138 the respective result for the first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10 , In the 14 and 15 is on a vertical axis 146 an exhaust gas back pressure before turbine in [mbar] is plotted. In 16 is on a vertical axis 148 an EGR rate in [%] and in 17 is on a vertical axis 150 NO _x flow rate is plotted in [g / h]. 14 illustrates the exhaust pressure before turbine without EGR, 15 illustrates the exhaust pressure before turbine with EGR, 16 illustrates a maximum EGR rate and 17 illustrates an achievable minimum NO _x emission. How out 14 As can be seen, reach the first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10 through the series connection of the turbocharger and the VTG technology, a 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 size for the maximum EGR rate, as in 16 shown. The maximum EGR rate for the second conventional embodiment of the internal combustion engine (bar graph 136 ; Biturbo with HD-ATL as fixed loader and ND-ATL as wastegate) is 45%. The maximum EGR rate for the first to tenth preferred embodiments of the internal combustion engine according to the invention 1 to 10 (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%. According to 12 results in the NO _x potential, which in 17 is shown as follows: 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 wastegate) is approx. 6 g / h. The NO _x emission for the first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10 (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.

18 shows a boost pressure variation analogous to 13 However, wherein the respective internal combustion engine (first conventional embodiment, second conventional embodiment or first to tenth embodiment according to 1 to 10 ) in the second emission area 116 (see. 11 ) Of 2.000 min ^-1 and medium pressure 112 pme = 8 bar is operated. Like reference numerals have the same meaning as in FIG 13 , so for their explanation on the above description of 13 is referenced. The mode of operation for generating the graphs 132 . 134 . 136 and 138 is analogous as above regarding 13 described. The maximum boost pressure of the second conventional embodiment of the internal combustion engine is 1,850 mbar at the point 140 , The maximum boost pressure of the first to tenth embodiments of the internal combustion engine according to the invention is 2,150 mbar at the point 142 , The maximum boost pressure of the first conventional embodiment of the internal combustion engine is 1.50 mbar at the point 144 , At maximum boost pressure is then, as at 13 , gradually open the EGR valve. With the first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10, a higher charge pressure is achieved with a higher EGR rate. This result will be described below with reference to 19 analyzed in more detail.

19 shows a representation of values for the soot flow rate and NOx flow rate analogous to 12 , The same reference numerals have the same meaning, so that for explanation thereof to the above description of 12 is referenced. 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 first to tenth preferred embodiment of the internal combustion engine according to the invention 1 10 shows the fourth graph 138 , Out 19 It can be seen that there is a NOx / soot trade for all three compared systems. In the second conventional embodiment (Biturbo with HD-ATL as a fixed loader and ND-ATL as a wastegate) high EGR rates are achieved only with low boost pressure, resulting in the same NO _x emissions to a high soot emissions. 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 first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10 has the highest emission potential due to 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 must in this case 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.

20 again shows a charge pressure variation analogous to 13 and 18 However, wherein the respective internal combustion engine (first conventional embodiment, second conventional embodiment or first to tenth embodiment according to 1 to 10 ) in the second emission area 116 (see. 11 ) Of 2.000 min ^-1 and medium pressure 112 pme = 8.9 bar is operated. Like reference numerals have the same meaning as in FIG 13 and 18 so that their explanation to the above description of 13 and 18 is referenced. The mode of operation for generating the graphs 132 . 134 . 136 and 138 is analogous to above 13 respectively. 18 described. In 20 are additional lines 152 same EGR rate for EGR rates of 0%, 5%, 10%, 15%, 20%, 25%, 30%, 35%, 40%, 45%, 50% and 55% plotted. In the 20 shown results clearly show that with the first to tenth preferred embodiment of the internal combustion engine according to the invention

1 to 10 higher EGR rates can be displayed in this operating range. Of the 20 can also be seen that at the same EGR rate, a higher fresh air content, ie more oxygen, is present in the combustion chamber.

21 shows a representation of values for the soot flow rate and NOx flow rate analogous to 12 and 19 , The same reference numerals have the same meaning, so that their explanation to the description of 12 and 19 is referenced. 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 first to tenth preferred embodiment of the internal combustion engine according to the invention 1 to 10 shows the fourth graph 138 , In 21 are additional lines 152 same EGR rate for EGR rates of 0%, 5%, 10%, 15%, 20%, 25%, 30%, 35%, 40%, 45%, 50% and 55% plotted. 21 puts that in relation to 20 illustrated physical effects graphically.

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 102007062366 A1 [0003]
• DE 19851028 C2 [0004]

Cited non-patent literature

• EN-GJSA-XNiSiCr35-5-2 [0079]
• EN-JS3061 [0079]

Claims (31)

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 the removal of exhaust gas ( 21 ) from the working cylinders ( 10 ), a first exhaust gas turbocharger ( 16 ) a low-pressure stage (LP exhaust gas turbocharger), which one in the exhaust tract ( 14 ) arranged first turbine ( 26 ) (ND turbine) and one in the fresh air tract ( 12 ) arranged compressors ( 36 ) (LP compressor), and at least one second exhaust gas turbocharger ( 18 ) a high-pressure (HD exhaust gas turbocharger), which in the exhaust tract ( 14 ) upstream of the first turbine ( 26 ) arranged second turbine ( 24 ) (HD turbine) and one in the fresh air tract ( 12 ) downstream of the first compressor ( 36 ) arranged second compressor ( 38 ) (HD compressor), characterized in that at least one of the turbines ( 24 . 26 ) an adjustable turbine geometry ( 68 ) (VTG). Internal combustion engine according to claim 1, characterized in that the exhaust tract ( 14 . 22 ) downstream of the first turbine ( 26 ) exclusively via the first turbine ( 26 ) with the exhaust tract ( 14 . 22 ) upstream of the first turbine ( 26 ) and downstream of the second turbine ( 24 ) is fluid-conductively connected. Internal combustion engine according to claim 1, characterized in that the exhaust tract ( 14 . 22 ) downstream of the first turbine ( 26 ) over the first turbine ( 26 ) and in addition via a first turbine bypass duct with first turbine bypass duct valve and / or a first wastegate ( 76 ) with first wastegate valve ( 78 ) with the exhaust tract ( 14 . 22 ) upstream of the first turbine ( 26 ) and downstream of the second turbine ( 24 ) is fluid-conductively connected. Internal combustion engine according to at least one of the preceding claims, characterized in that the exhaust gas tract ( 14 . 22 ) downstream of the second turbine ( 24 ) and upstream of the first turbine ( 26 ) exclusively via the second turbine ( 24 ) with the exhaust tract ( 14 . 22 ) upstream of the second turbine ( 24 ) is fluid-conductively connected. Internal combustion engine according to at least one of claims 1 to 3, characterized in that the exhaust gas 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 duct ( 72 ) with second turbine bypass channel valve ( 74 ) and / or a second wastegate ( 80 ) with second wastegate valve ( 82 ) with the exhaust tract ( 14 . 22 ) upstream of the second turbine ( 24 ) is fluid-conductively connected. Internal combustion engine according to at least one of the preceding claims, characterized in that the fresh air tract ( 12 . 44 ) upstream of the first compressor ( 36 ) exclusively 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 fluid-conductively connected. Internal combustion engine according to at least one of claims 1 to 5, characterized in that the fresh air tract ( 12 . 44 ) upstream of the first compressor ( 36 ) over the first compressor ( 36 ) and in addition via a first compressor bypass passage with the first compressor bypass valve with the fresh air tract ( 12 . 44 ) downstream of the first compressor ( 36 ) and upstream of the second compressor ( 38 ) is fluid-conductively connected. Internal combustion engine according to at least one of the preceding claims, characterized in that the fresh air tract ( 12 . 44 ) upstream of the second compressor ( 38 ) and downstream of the first compressor ( 36 ) exclusively via the second compressor ( 38 ) with the fresh air tract ( 12 . 44 ) downstream of the second compressor ( 38 ) is fluid-conductively connected. Internal combustion engine according to at least one of claims 1 to 7, characterized in that 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 channel ( 46 ) with second compressor bypass channel valve ( 54 ) with the fresh air tract ( 12 . 44 ) downstream of the second compressor ( 38 ) is fluid-conductively connected. Internal combustion engine according to claim 9, characterized in that in the second compressor bypass duct ( 46 ) a third intercooler ( 56 ) is arranged. Internal combustion engine according to at least one of the preceding claims, characterized in that only the second turbine ( 24 ) a VTG ( 68 ) having. Internal combustion engine according to at least one of claims 1 to 10, characterized in that only the first turbine ( 26 ) a VTG ( 68 ) having. Internal combustion engine according to at least one of claims 1 to 10, characterized in that the first and the second turbine ( 26 . 24 ) one VTG each ( 68 ) having. Internal combustion engine according to at least one of the preceding claims, characterized characterized in that downstream of the second compressor ( 38 ) a second intercooler ( 40 ) is arranged. 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 intercooler ( 57 ) is arranged. Internal combustion engine according to at least one of the preceding claims, characterized in that upstream of the second turbine ( 24 ) a high-pressure exhaust gas recirculation line ( 58 ) (HD EGR line) from the exhaust tract ( 14 . 22 ) branches off and downstream of the second compressor ( 38 ) into the fresh air tract ( 12 . 44 ). Internal combustion engine according to claim 16 and 14, characterized in that the HP-EGR line ( 58 ) downstream of the second intercooler ( 40 ) into the fresh air tract ( 12 . 44 ). Internal combustion engine according to claim 16 or 17, characterized in that in the HP-EGR line ( 58 ) a cooler ( 60 ) is arranged for recirculated exhaust gas (HD-EGR cooler). Internal combustion engine according to at least one of claims 16 to 18, characterized in that in the HP-EGR line ( 58 ) a valve ( 62 ) is arranged for recirculated exhaust gas (HP-EGR valve). Internal combustion engine according to at least one of claims 18 or 19, characterized in that downstream of the first turbine ( 26 ) a low-pressure exhaust gas recirculation line ( 64 ) (LP EGR line) from the exhaust tract ( 14 . 22 ) and into the HD EGR cooler ( 60 ). 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 ) (LP EGR line) from the exhaust tract ( 14 . 22 ) branches off and upstream of the first compressor ( 36 ) into the fresh air tract ( 12 . 44 ). Internal combustion engine according to claim 21, characterized in that in the LP EGR line ( 64 ) is arranged a recirculated exhaust gas cooler (LP EGR cooler) and / or a recirculated exhaust gas valve (LP EGR valve). 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 ) a medium-pressure exhaust gas recirculation line ( 66 ) (MD-EGR line) from the exhaust tract ( 124 . 22 ) branches off and downstream of the first compressor ( 36 ) and upstream of the second compressor ( 38 ) into the fresh air tract ( 12 . 44 ). Internal combustion engine according to claim 23, characterized in that in the MD-EGR line ( 66 ) a recirculated exhaust gas cooler (MD-EGR cooler) and / or a recirculated exhaust gas valve (MD-EGR valve) is arranged. 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 cylinders ( 10 ) having. Internal combustion engine according to at least one of the preceding claims, characterized in that the internal combustion engine direct fuel injection into at least one working cylinder ( 10 ), in particular according to the common rail system. Internal combustion engine according to at least one of the preceding claims, characterized in that the internal combustion engine has a nominal 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. 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. 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. 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. 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.

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