WO2013083211A1

WO2013083211A1 - Internal combustion engine, in particular for a motor vehicle

Info

Publication number: WO2013083211A1
Application number: PCT/EP2012/003967
Authority: WO
Inventors: Siegfried Sumser; Thomas Kuhn; Michael Stiller
Original assignee: Daimler Ag
Priority date: 2011-12-06
Filing date: 2012-09-21
Publication date: 2013-06-13
Also published as: DE102011120337A1

Abstract

The invention relates to an internal combustion engine (10), in particular for a motor vehicle, comprising an intake duct (16) via which air can be delivered to the internal combustion engine (10) and in which, in the flow direction of the air through the intake duct (16), a first compressor (20) of a first exhaust gas turbocharger (22) is disposed and a second compressor (26) of a second exhaust gas turbocharger (28) is arranged downstream of the first compressor (20), wherein said second compressor is series-connected to said first compressor (20) and can compress the air. The internal combustion engine further comprises an exhaust gas duct (40) through which exhaust gas from the internal combustion engine (10) can flow and in which a first turbine (42) of the first exhaust gas turbocharger (22) by means of which the first compressor (20) can be driven and at least one second turbine (46) of the second exhaust gas turbocharger (28) by means of which the second compressor (26) can be driven are arranged, wherein the turbines (42, 46) are connected in parallel to one another with regard to the flow of the exhaust gas through the exhaust gas duct (40).

Description

Internal combustion engine, in particular for a motor vehicle

The invention relates to an internal combustion engine, in particular for a

Motor vehicle, according to the preamble of claim 1.

DE 10 2009 006 359 A1 discloses a device for variable

Exhaust gas turbocharging and exhaust gas recirculation of an internal combustion engine. The internal combustion engine comprises an intake tract over which the

Internal combustion engine air can be supplied. In the flow direction of the air through the intake tract, a first compressor of a first exhaust-gas turbocharger and downstream of the first compressor, a second compressor of a second exhaust-gas supercharger serially connected to the first compressor are arranged. By means of the compressor is the

To compress internal combustion engine supplied air.

The internal combustion engine further includes an exhaust gas of the

Internal combustion engine through-flow exhaust tract, in which a first turbine of the first exhaust gas turbocharger and a second turbine of the second

Exhaust gas turbocharger are arranged. The first compressor can be driven by means of the first turbine. Furthermore, the second compressor can be driven by means of the second turbine.

The turbines are associated with respective bypass devices via which the corresponding turbine is to bypass the exhaust gas. The bypass devices are also referred to as blow-off devices, by means of which exhaust gas is blown off upstream of the respective turbine. The exhaust gas blown off can bypass the corresponding turbine and does not drive it.

Due to high demands on an agile driving behavior as well as the demand for very high stationary engine torques turbines of exhaust gas turbochargers used such devices with relatively small inlet nozzle surfaces and Radausströmungsflächen. At relatively high engine speeds, the blow-off devices are used for a flow area enlargement of the turbines. The blow-off devices have a negative effect on the fuel consumption, since the exhaust gas is passed unused to the corresponding turbine.

It has also been found that wheel inlet variabilities of the turbines, i. Variable turbine geometries, not sufficient to show sufficiently large flow areas for the entire full load line can not be sufficient, which is why a relatively large map range of the map of the internal combustion engine is also provided due to a limited throughput capability of the corresponding turbine of the turbine, in which exhaust gas is blown off.

In particular, in internal combustion engine, which are designed according to the so-called downsizing principle, two-stage superchargers are used as the known device, in particular in the flow direction of the exhaust gas through the exhaust gas downstream of the second turbine arranged first turbine already at relatively low engine speeds, based on the engine speed of the maximum engine torque, very high Abblaseraten be provided in order to realize a satisfactory performance of the highly-supercharged internal combustion engine.

The negative consequences are, in addition to transient discontinuities, in particular of the second compressor, relatively poor charge changes of the internal combustion engine with high fuel consumption, in particular in an upper half of the engine speed range at high load. In addition, a not inconvenient regulation or control is needed to counter these problems at least rudimentary.

It is therefore an object of the present invention to further develop an internal combustion engine of the type mentioned above in such a way that it has an improved design

Has driveability.

This object is achieved by an internal combustion engine, in particular for a motor vehicle, with the features of patent claim 1. advantageous

Embodiments with expedient and non-trivial developments of the invention are specified in the remaining claims. Such an internal combustion engine, in particular for a motor vehicle, has an intake tract, via which air can be supplied to the internal combustion engine. In the flow direction of the air through the intake tract, a first compressor of a first exhaust gas turbocharger is arranged in the intake tract. Downstream of the first compressor is a second compressor, serially connected to the first compressor, of a second compressor

Exhaust gas turbocharger arranged in the intake. The air is to be compressed by means of the compressors.

The internal combustion engine further includes one of exhaust gas of

Internal combustion engine throughflowable exhaust tract. A first turbine of the first exhaust gas turbocharger and a second turbine of the second exhaust gas turbocharger are arranged in the exhaust gas tract. By means of the first turbine, the first compressor is drivable. By means of the second turbine, the second compressor is drivable.

According to the invention, it is provided that the turbines are arranged parallel to one another with respect to the flow of the exhaust gas through the exhaust tract. As a result, a relatively simple connection of the exhaust gas turbocharger with the associated

Internal combustion engine realized that allows at least substantially continuous switching on and off of the exhaust gas turbocharger. This creates in

Switchover phases, in which between a first operation, in which both

Are exhaust turbocharger operated, and a second operation in which only one of the exhaust gas turbocharger is operated, is switched, no or very little and imperceptible discontinuities in the provision of engine torque by means of the internal combustion engine. The internal combustion engine thus has an improved driving behavior.

The internal combustion engine can be designed as a gasoline engine, diesel engine, diesel engine or other internal combustion engine. The inventive circuit of the exhaust gas turbocharger with the internal combustion engine is particularly advantageous in a gasoline engine, which preferably with a

Combustion air ratio λ of at least substantially 1 is operated. The inventive circuit of the exhaust gas turbocharger and the internal combustion engine avoids unwanted fluctuations in the switching phases

Combustion air ratio.

The internal combustion engine according to the invention thus has a particularly advantageous, agile and efficient operation with a pleasant driving behavior. In addition, the cost of the internal combustion engine according to the invention can be kept low due to the simple circuit. In addition, the

Internal combustion engine according to the invention are particularly efficient and thus operated with only low fuel consumption, since the above-mentioned blow-off of exhaust gas can be avoided or can be kept within very narrow limits. This is associated with low C0 ₂ emissions.

In a further advantageous embodiment, the air by means of the second compressor from a first pressure level upstream of the second compressor and downstream of the first compressor to a second pressure level downstream of the second compressor more compressible than the air by means of the first compressor from a third pressure level upstream of the first compressor a fourth pressure level downstream of the first

Compressor and upstream of the second compressor is compressible. In other words, the second compressor arranged downstream of the first compressor and connected in series therewith is the high-pressure compressor in comparison to the first compressor. The first compressor is thus the low-pressure compressor. This leads to a particularly advantageous operation of the internal combustion engine.

In a further advantageous embodiment of the invention, at least one of the turbines, in particular the first turbine, a variable turbine geometry. By means of the turbine geometry, the associated turbine at different operating points of the internal combustion engine and thus to different mass or

Volume flow of the exhaust gas can be adjusted. This benefits the efficient operation of the corresponding turbine and thus the internal combustion engine.

The variable turbine geometry is in particular a

Swirl generator, which ensures an advantageous flow of an associated turbine wheel. By means of the variable turbine geometry, an effective flow cross-section of the associated turbine can also be variably adjusted by adjusting it.

It need not necessarily be provided that the effective

Flow cross-section using the variable turbine geometry is to narrow to very small values. Advantageously, leaks of the variable turbine geometry having turbine in a closed position of the variable turbine geometry, in which the effective flow cross-section is maximally narrowed, very low and substantially zero, so that in the closed position of the variable turbine geometry, the associated turbine of the exhaust gas at least substantially not driven and thus turned off. Then there is only the other turbine, which has no variable turbine geometry or whose variable turbine geometry is in an open position, is driven by exhaust gas.

In a further particularly advantageous embodiment, at least one of the compressors has an adjusting device, by means of which flow conditions for the air flowing through these at least one compressor can be variably adjusted. The adjusting device thus provides a variable flow geometry for the associated compressor, by means of which the associated compressor can be adapted as required to different operating points of the internal combustion engine and thus to different mass or volume flows of the air flowing through this compressor. This leads to a

particularly efficient operation of the internal combustion engine.

In a further advantageous embodiment of the invention, at least one of the compressors, in particular the first compressor, a bypass device

assigned, via which this at least one compressor is bypassed by the air. The at least one compressor bypassing air is thus not compressed by this at least one compressor. The bypass device provides a particularly simple and cost-effective way to minimize flow losses and to create a variability on the compressor side.

In a particularly advantageous embodiment of the invention, the second turbine at least two partially fluidly separated from each other floods. This makes it possible to carry out a particularly advantageous exhaust gas recirculation in the internal combustion engine according to the invention, in order to keep in particular their nitrogen oxide and particulate emissions low.

By means of the two floods is a further degree of freedom in the inventive

Internal combustion engine shown by means of which can be recycled to high loads exhaust gas from the exhaust system to the intake and introduced into the intake without experiencing charge changes of the internal combustion engine an undesirably large disadvantage in terms of increased fuel consumption.

Exhaust gas recirculation is particularly advantageous when the floods

are formed asymmetrically to each other. In this case, the floods are asymmetrical with respect to one another, in particular with regard to their respective flow cross-sections through which exhaust gas can flow, that is to say they are formed differently from one another. Preferably, one of the floods has a smaller flow cross-section than the other flood. By means of this small flow cross-section can be an advantageous

Aufstauverhalten be presented so that high amounts of exhaust gas are traceable.

In a further advantageous embodiment, the second turbine has a first number of floods, wherein the first number is at least two. The first turbine is only fluidically connected to a second number of flows of the second turbine, which number is smaller than the first number. In other words, if the second turbine has two floods, then the first turbine is advantageously only fluidically connected to one of the flows and connected in parallel with the first turbine with respect to this. This represents a very advantageous shading of the two turbines, resulting in a very good

Operability and thus a very good driving behavior of

Internal combustion engine leads.

Further advantages, features and details of the invention will become apparent from the following description of preferred embodiments and from the drawing. The features mentioned above in the description and

Feature combinations as well as the features and feature combinations mentioned below in the description of the figures and / or shown alone in the figures are not only in the respectively indicated combination but also in others

Combinations or alone, without departing from the scope of the invention.

The drawing shows in:

Fig. 1 is a schematic diagram of an internal combustion engine for a

Motor vehicle, which can be charged by means of two exhaust gas turbochargers;

Fig. 2 is a schematic diagram of another embodiment of the

Internal combustion engine according to FIG. 1;

Fig. 3 is a schematic diagram of another embodiment of the

Internal combustion engine according to Figures 1 and 2.

4 is a schematic diagram of another embodiment of the

Internal combustion engine according to FIGS. 1 to; 5a shows a detail of a schematic longitudinal sectional view of a compressor of one of the exhaust gas turbocharger according to FIGS. 1 to;

FIG. 5b a detail of a further schematic longitudinal sectional view of the

Compressor according to Fig. 5a;

6 is a schematic diagram for illustrating the functions of

Internal combustion engine with the two exhaust gas turbocharger according to FIGS. 1 to 4.

7 is a schematic diagram of another embodiment of the

Internal combustion engine according to FIG. 1;

8 is a schematic diagram of another embodiment of the

Internal combustion engine according to FIG. 7; and

9 is a schematic diagram of another embodiment of the

Internal combustion engine according to FIG. 7;

1 shows an internal combustion engine 10 with four cylinders 12 present. The internal combustion engine 10 comprises a crankshaft 14, via which a

Torque can be dissipated from the internal combustion engine 10.

The internal combustion engine 10 further comprises an intake tract 16, via which the internal combustion engine 10 can be supplied with air. In the intake duct 16, an air filter 18 is arranged, which filters the air sucked by the internal combustion engine. In addition, a first compressor 20 of a first exhaust gas turbocharger 22 is arranged in the intake tract 16. The first compressor 20 includes a first one

Compressor 24 for compressing the intake air.

In the flow direction of the air through the intake tract 16 is downstream of the first

Compressor 20, a second compressor 26 of a second exhaust gas turbocharger 28 is arranged. The second compressor 26 comprises a second compressor wheel 30, by means of which the air is to be compressed. The second compressor 26 is connected in series to the first compressor 20. By the serially connected compressors 20, 26 a two-stage, serial compression of the air is shown. Compared to the first compressor 20, the second compressor 26 can more compress the air. This means that the second compressor 26 is the high-pressure compressor, while the first compressor 20 is the low-pressure compressor.

Thus, a first pressure ratio TT _ND = P ₂ N PI of the low-pressure compressor is less than a second pressure ratio TT _H D = Ρ2Η Ρ2Ν · Here, P denotes a first pressure level of the air upstream of the first compressor 20. P ₂ N denotes a second one

Pressure level downstream of the first compressor 20 and upstream of the second compressor 26, to which the air is compressed from the first pressure level P by means of the first compressor 20. P ₂ H denotes a third pressure level to which the air is compressed by means of the second compressor 26, starting from the second pressure level P _2N .

During compression, the air is heated. To display particularly high levels of charge, an intercooler 32 is arranged between the compressors 20, 26, by means of which the compressed air is cooled. Downstream of the second compressor 26 is a

Intercooler 34 is arranged, by means of which the air is also cooled. Downstream of the intercooler 34, an optional flap 36 is provided, by means of which the amount of air can be adjusted. Downstream of the flap 36, the air has a boost pressure P _2S , with which it flows to a charge air manifold 38 in the intake manifold 16. By means of the charge air manifold 38, the compressed air is split on the cylinder 12.

The internal combustion engine 10 also includes an exhaust tract 40, in which a first turbine 42 of the first exhaust gas turbocharger 22 is arranged. The first turbine 42 comprises a first turbine wheel 44, which can be driven by exhaust gas of the internal combustion engine 10. In the exhaust tract 40, a second turbine 46 is further arranged, which comprises a second turbine wheel 48. The second turbine wheel 48 can also be driven by the exhaust gas of the internal combustion engine 10. To do this, appropriate

Exhaust piping of the exhaust tract 40, the exhaust gas to the turbines 42, 46th

The first exhaust gas turbocharger 22 includes a first shaft 50, with which the first

Turbine 44 and the first compressor 24 are rotatably connected. As a result, the first compressor 20 can be driven by the first turbine 42. Correspondingly, the second exhaust gas turbocharger 28 comprises a second shaft 52, with which the second turbine wheel 48 and the second compressor wheel 30 are connected in a rotationally fixed manner.

Thereby, the second compressor 26 can be driven by the second turbine 46. As can be seen from FIG. 1, the turbines 22, 28 in the exhaust tract 40 are connected in parallel with one another. This means that the turbines 42, 46 are acted upon by the inlet pressure P _{3 of} the exhaust gas prevailing in the flow direction of the exhaust gas through the exhaust tract 40 upstream of the turbines 42, 46, at least substantially the same. By means of the turbines 42, 46, the exhaust gas is expanded starting from the inlet pressure P ₃ , so that it has an outlet pressure P ₄ downstream of the turbine 42, 46.

Downstream of the turbines 42, 46 at least one exhaust aftertreatment device 54 is arranged in the exhaust tract 40, by means of which the exhaust gas is purified before it is discharged to the environment.

To illustrate a high degree of agility and thus a very good response especially of the second turbine 46, this is designed to be relatively small in terms of their flow geometry. The first turbine 42 is designed in particular to cover a high throughput requirement and to represent a low fuel consumption and can be made larger in terms of their flow geometry than the second turbine 46th

With Φ _{ΤΙ, ΑΤ} - the first flow rate _{parameter of} the first exhaust gas turbocharger 22 and the first turbine 42 is characterized. With Φ _Τ2 . _ΑΤ . the second flow rate parameter of the second exhaust gas turbocharger 28 or the second turbine 46 is characterized.

The internal combustion engine 10 further includes an exhaust gas recirculation device 56 with at least one exhaust gas recirculation line 58, which is fluidically connected to the exhaust gas tract 40 at a branch point 60. The exhaust gas recirculation line 58 is fluidically connected to the intake manifold 16 at an introduction point 62. By means of the exhaust gas recirculation line 58, exhaust gas can be branched off at the branching point 60, returned to the intake tract 16 and introduced into the intake tract 16 at the introduction point 62.

The introduction point 62 is also referred to as a mixing point M, since there takes place a mixing of the air with the exhaust gas. The exhaust gas in the air acts as an inert gas in combustion processes in the cylinders 12, which keeps the temperatures low. So can

Nitrogen oxide and particulate emissions of the internal combustion engine 10 are kept low.

To adjust the amount of recirculated exhaust gas includes the

Exhaust gas recirculation device 56, an exhaust gas recirculation valve 64. Downstream of Exhaust gas recirculation valve 64, an exhaust gas recirculation cooler 66 is provided, by means of which the recirculated exhaust gas is to be cooled.

In the exhaust gas recirculation line 58, a pressure modulator 68 is further provided, by means of which pressure oscillation excitations caused by Auslaßpulsationen the internal combustion engine 10, are modulated in the exhaust gas recirculation line 58 so far that in the region of the discharge point 62 no or only very small effective

Stimulation intensities are more present. The pressure modulator 68 is arranged downstream of the exhaust gas recirculation cooler 66 and upstream of the discharge point 62 and comprises a damping volume 70 and a matched, effective and by a

corresponding component, such as a diaphragm, formed Zuströmquerschnitt 72 and an adapted, effective and by a corresponding component,

For example, a diaphragm, formed Abströmquerschnitt 74, which cause in conjunction with the size of the damping volume 70, a strong damping of the pressure oscillation and thus of pressure pulsations at the discharge point 62.

The first turbine 42 further includes a first variable turbine geometry 77, by means of which flow conditions for the exhaust gas flowing through the first turbine 42 are variably adjustable. In particular, it is possible to have an effective

Flow cross-section of the first turbine 42 by means of the first variable

Adjust turbine geometry 77, so as to adapt the first turbine 42 to different operating points of the internal combustion engine 10 and thus to different volume and mass flows of the exhaust gas. The first variable turbine geometry 77 may be, for example, a so-called tongue slide, an axial slide or another adjusting device.

The internal combustion engine 10 further comprises a control device 76, by means of which the exhaust gas recirculation valve 64 and the first variable turbine geometry 77 are regulated.

By moving or adjusting the first variable turbine geometry 77, the effective flow cross section is optionally continuously opened from at least substantially zero or at least near zero to a predefinable setpoint value greater than zero, whereby the first turbine 42 continuously creates an additional flow cross section to the second turbine 46. The additional flow cross-section means a continuous power connection of the first turbine 42. The second increases Pressure level P _2N upstream of the second compressor 26, in this phase the

Main pressure ratio delivers, continuously. The variable effective

Flow cross-section of the first turbine 42 acts like a blow-off valve upstream of the second turbine 46, but with the advantage that the charge cycle of the

Internal combustion engine 10 is favored via the use of exhaust gas energy by means of the first turbine 42 and an increase in air supply with relatively low

Entry pressure P _3, however, can take place with exhaust gas mass increase. By combining respective operating characteristics of the two turbines 42, 46 and the two compressors 20, 26, the compressor pressure ratio distribution and the

Turbine flow rate distribution up to the rated motor nominal exerting influence on the design side.

Also in the nominal point of the internal combustion engine 10, both exhaust gas turbochargers 22, 28 run with the same turbine pressure ratio, but not necessarily the same

Speeds and same turbine throughputs. Equally, there need not be an equal distribution of the compressor pressure ratios of the two compressors 20, 26 and equal throughputs.

The advantages of the internal combustion engine 10 can be seen in particular with reference to FIG. 6. Fig. 6 shows a diagram 78, on the abscissa 80, the relative

Engine speed n _mo t / n _mo t_ _{D is} plotted. With n _mot a currently present speed of the internal combustion engine 10 is designated, while with n _{mot D} the

Rated speed of the internal combustion engine 10 is designated.

The ordinate 82 of the diagram 78 is the quotient

applied. The diagram 78 shows a design range A, as well as the effect of the exhaust gas turbochargers 22, 28 on the internal combustion engine 10 along the full load line. In a lower portion 84 of the design range A, there are relatively low agility requirements, with the high pressure compressor and associated second turbine 46 (high pressure turbine) operating at high throughput capacities. This is one

realized economical design. In an upper area 68 of the

Design area A have high agility requirements. The high-pressure compressor and the high-pressure turbine are operated with relatively low throughput capacities.

At relatively low speeds of the internal combustion engine 10, the low-pressure compressor, although it is traversed by air, provide a very small to vanishingly small first pressure ratio TT _nd , since the first variable turbine geometry 77 of the associated first turbine 42 (low-pressure turbine) generally only very low exhaust gas flow rates or no exhaust gas flow rates in their closed position lets through. However, according to the agility requirement, the high-pressure compressor may deliver very high second pressure ratios TT _H D even at low rotational speeds of the internal combustion engine 10. With increasing engine speeds, the first variable turbine geometry 77 is opened, so that the effective flow cross section of the first turbine 42 is opened. The second pressure level P ₂ N of the high-pressure compressor or the first pressure ratio TT _{nd of} the low-pressure compressor increases. Depending on the throughput and efficiency characteristics of the four turbomachines (compressors 20, 26 and turbine 42, 46), a nearly constant upper boost pressure along the engine speed will be set in a course of the quotient shown in FIG. 6 and resulting in the design range A, wherein the first pressure ratio TT _nd (low pressure ratio) increases and the second pressure ratio TT _H D (high pressure ratio) decreases due to the increasing power requirement of the high pressure compressor with increasing air flow rate.

Depending on the design, a quotient in a range from at least substantially 0.7 to 0.9 could be set in the nominal point of the internal combustion engine which is denoted by N. These relatively high quotients in the nominal point of the two compressors 20, 26 to each other are influenced by the demands of the full throughput, which must flow through the high pressure compressor. To the nominal flow rate through the geometrically smaller than the low-pressure compressor high-pressure compressor without

To let plug flow below the speed of sound in the narrowest flow cross sections of the compressor, advantageously, the volume flow of the high-pressure compressor by adjusting the second pressure level P ₂ N or a

Pressure increase of the low-pressure compressor and the intercooling by the intercooler 32 correspondingly reduced, which about the power control means of the first variable turbine geometry 77 of the low-pressure turbine, the quasi the

Abblasestrom the high-pressure turbine processed, is significantly affected.

FIG. 2 shows a further embodiment of the internal combustion engine 10 according to FIG. 1. The first compressor 20 is assigned a bypass device 88, which comprises a bypass line 90 and a valve element 92. The bypass line 90 is connected on the one hand upstream of the first compressor 20 to the intake manifold 16. On the other hand, the bypass line 90 is fluidly connected to the intake manifold 16 downstream of the first compressor 20. Air can thus bypass the first compressor wheel 24 via the bypass line 90, so that the air is not compressed by the first compressor wheel 24. By means of the arranged in the bypass line 90 valve member 92, the amount of the first compressor 24 bypassing air can be variably adjusted as needed. By means of the bypass device 88 can be easily in at least substantially completely closed, first variable turbine geometry 77 and thus at least substantially deactivated first turbine 42nd

Flow rate and flow losses through the first compressor 20 via a relatively large mass flow of the first compressor 20 bypassing air are kept low.

FIG. 3 shows a further embodiment of the internal combustion engine 10, in which the first compressor 20 has a first adjusting device 94. By means of the first adjusting device 94, flow conditions of the first compressor 20 for the air flowing through the first compressor 20 can be variably set. The first adjusting device 94 thus represents a variability by means of which the first compressor 20 can be adapted to different mass flows of the air. Thus, the fact is taken into account that the entire air mass flow on low compressor activity of the background compressor through its flow channels

flows through. In these inactivity phases, it is advantageous to open the flow channels as far as possible and to move into a closed position of variability with increasing activation of the low-pressure compressor. As can be seen from FIG. 3, the first adjusting device 94 can be regulated by means of the regulating device 76.

4 shows a further embodiment of the internal combustion engine 10, in which now the second compressor 26 has a second adjusting device 96, by means of which flow conditions of the second compressor 26 for the second

Compressor 26 can be adjusted by flowing air.

Conventionally, to satisfy high demands on the transient behavior of the internal combustion engine 10, the compressors and the turbines of exhaust gas turbochargers with regard to their flow geometry to design particularly small, in short times to provide high boost pressures and air volumes of the internal combustion engine.

As a result, however, the problem arises conventionally even at average rotational speeds of the internal combustion engine 10 that the high-pressure compressor has too low throughput capability, whereby a compressor plug occurs. To this To counteract problems with the internal combustion engine 10, are the

Parting devices 94, 96 provided by means of which flow cross-sections upstream and / or downstream of the compressor wheels 24, 30 are variably adjustable.

FIGS. 5a-b show possibilities for the representation of the adjusting devices 94, 96 of the compressors 20, 26. The adjusting devices 94, 96 comprise a cone slide 98 which is displaceable at least substantially in the axial direction, by means of which the flow surface downstream of the compressor wheels 24, 30 is variable is adjustable.

In addition, an axial slide element 100 is provided, by means of which the

Flow area upstream of the compressor wheels 24, 30 is adjustable. The

Adjusting means 94, 96 with the cone slide 98 and the axial slide element 100 thereby represent wheel variability by means of which, in particular, the limits of the quotient of the low-pressure ratio and the high-pressure ratio can be expanded.

Preferably, the cone slide 98 and the axial slide element 100 are coupled together by means of a suitable coupling device, so that both the

Taper slider 98 and the axial slide element 100 moves by means of a common actuator and so the corresponding flow cross sections can be adjusted.

FIG. 5 a shows the cone slide 98 and the axial slide element 100 in their respective maximum, which show the corresponding flow cross sections

Closed positions, while Fig. 5b, the cone 98 and the axial slide element 100 in their respective, the respective flow cross-sections maximum releasing

Open positions shows.

If the second turbine 46 of the internal combustion engine 10 according to FIGS. 1 to 4 is of single-flow design, then it is double-walled in the internal combustion engines 10 according to FIGS. 7 to 9 and has a first flow 102 and a second flow 104. As can be seen from FIG. 7, the first flow 102 is fluidically connected in fluid communication with two first of the four cylinders 12 of the internal combustion engine 10. The second flood 104 is fluidly connected to the two other, second of the four cylinders 12. The first flow 102 is a first flow throughput parameter associated with _Φ Τ 2.ιο2, while the second flow 104 is a second flood throughput parameter _Φ Τ 2, ιο assigned.

Where: ^φ Γ2, ιο2 =. In this case, m ₁₀₂ denotes the

Exhaust gas mass flow rate through the first flood 102. With m ₁₀₄ , the exhaust gas mass flow rate through the second flood 104 is designated. Further, Τ _102, the total temperature at the turbine inlet of the first flood 102, T ₁₀₄ denote the total temperature of the exhaust gas at

Turbine inlet of the second flood 104, P _102, the total pressure at the turbine inlet of the first flood 102 and P _04, the total pressure at the turbine inlet of the second flood 104th

The exhaust gas recirculation line 58 is fluidically connected to the first flow 102 at the branch point 60. Thus, the first flood 102 is also referred to as EGR (EGR) exhaust and the first cylinders as EGR. The first flood 102 thus serves in particular to represent a desired Aufstauverhalten to represent an advantageous exhaust gas recirculation with relatively high exhaust gas recirculation rates.

The second flow 104 serves, in particular, to represent a desired combustion air ratio λ, for which reason the second flow 104 is also referred to as the λ flow and the second cylinders as λ cylinders.

In general, the flow rate parameters differ Φ _{τ2 102} and Φ _Γ2 , ιο4 - where it is in the advantageous combination circuit of the turbines 42, 46 for

Using a dual-flow, asymmetric turbine than the second turbine 46 comes with which an optimal vote on fuel consumption and NO _x - emissions (nitrogen oxide emissions) is possible over the dimensions. In other words, the floods 102, 104 are their respective ones

Flow cross sections designed asymmetrically to each other.

By moving the first variable turbine geometry 77, the turbines 42, 46 of the internal combustion engine 10 according to FIG. 7 corresponding to the turbines 42, 46 of the internal combustion engine 10 according to FIGS. 1 to 4 can be regulated and adapted. Thus, both exhaust gas turbochargers 22, 28 are operated at the same turbine pressure ratio of the λ cylinder in the internal combustion engine 10 according to FIGS. 7 to 9 at the nominal point of the internal combustion engine 10, but not necessarily with the same exhaust gas turbocharger speeds and turbine throughputs. Nor does it have to be a division of the compressor pressure ratios of the two compressors 20, 26 result. A particularly important role for exhaust gas recirculation plays the twin-flow

asymmetric second turbine 46, whose generally limited absorption capacity of the λ-wave with fixed geometry, that is without variable turbine geometry, by the connection by means of the first variable turbine geometry 77 in the

Flow cross-sections is optimally expandable. The λ cylinders and the EGR cylinders are summarized according to the firing order 2-4 and 1-3.

The internal combustion engine 10 according to FIG. 8 has six cylinders 12, wherein first of the cylinders 12 are fluidically connected to the first flow 102 and thus are the EGR cylinders. The three other second of the cylinders 12 are fluidly connected to the second flood 104 and represent the λ-cylinder. In other words, in the

Internal combustion engine 10 shown in FIG. 8, a cylinder summary 1-2-3 for the λ-cylinder and a cylinder summary 4-5-6 for the EGR cylinder shown. Here, by means of the first turbine 42 and its first variable turbine geometry 77, which is connected in parallel with the λ-flood, a so-called λ-variability is shown, although the used, double-flow and asymmetrical second turbine 46 a

Solid geometry turbine is so has no variable turbine geometry.

There may be applications in which a high spread of the

Exhaust gas recirculation rates in the overall map of the internal combustion engine 10 may be required. These requirements for the exhaust gas recirculation can be handled with the circuit shown in FIG. 9.

According to FIG. 9, the variable first turbine 42 is connected in parallel to the EGR flood of the second turbine 46, whereby a so-called AGR variability can be realized. The absorption capacity of the asymmetrical second turbine 46 on the side of its EGR flood is generally very small in the second turbine 46 designed as a solid geometry turbine, since the parallel connection of the variable first turbine 42 results in a permanent optimum flow cross-sectional adaptation to the EGR system. requirement

is feasible. The λ-flood of the second turbine 46 can be simplified as

Be formed solid geometry turbine.

Alternatively, however, it is possible that the second turbine 46 has an adjusting device, which is arranged, for example, in the λ-flood at least substantially directly upstream of the second turbine wheel 48. As a result, a functional separation between the first variable turbine geometry 77 of the first turbine 42 and the adjusting device of the second turbine 46 is shown. The adjusting device of the second Turbine 46 thus represents a second variable turbine geometry, by means of which the second turbine 46 can be variably adjusted, in particular with regard to its effective flow cross section through which exhaust gas can flow, and adapted to different operating points. The two variable turbine geometries thus represent variabilities which can be distributed in each case to the turbines 42, 46, which leads to a high potential in the optimization of the supercharged internal combustion engine 10.

Claims

claims

1. internal combustion engine (10), in particular for a motor vehicle, with a

Intake tract (16) via which the internal combustion engine (10) air can be supplied and in which in the flow direction of the air through the intake tract (16), a first compressor (20) of a first exhaust gas turbocharger (22) and downstream of the first compressor (20) on the first compressor (20) serially connected second compressor (26) of a second exhaust gas turbocharger (28) are arranged, by means of which the air is compressible, and with a exhaust gas from the internal combustion engine (10) through-flowable exhaust tract (40), in which a first turbine (42) of the first exhaust-gas turbocharger (22), by means of which the first compressor (20) can be driven, and at least one second turbine (46) of the second exhaust-gas turbocharger (28), by means of which the second compressor (26) can be driven, are arranged,

characterized in that

the turbines (42, 46) are connected in parallel with respect to the flow of the exhaust gas through the exhaust tract (40).

2. internal combustion engine (10) according to claim 1,

characterized in that

the air by means of the second compressor (26) from a first pressure level (P2N) upstream of the second compressor (26) to a second pressure level (P2H) downstream of the second compressor (26) is more compressible than the air by means of the first compressor (20) from a third pressure level (P1) upstream of the first compressor (20) to a fourth pressure level (P2N) downstream of the first compressor (20) and upstream of the second compressor (26) is compressible.

3. internal combustion engine (10) according to any one of claims 1 or 2,

characterized in that

at least one of the turbines (42, 46), in particular the first turbine (42), a variable turbine geometry (77).

4. Internal combustion engine (10) according to one of the preceding claims,

characterized in that

at least one of the compressors (20, 26) has an adjusting device (94, 96) by means of which flow conditions for the air flowing through these at least one compressor (20, 26) are variably adjustable.

5. Internal combustion engine (10) according to one of the preceding claims,

characterized in that

at least one of the compressor (20, 26), in particular the first compressor (20), a bypass device (88) is assigned, via which said at least one compressor (20, 26) is bypassed by the air.

6. Internal combustion engine (10) according to one of the preceding claims,

characterized in that

the second turbine (46) has at least two at least partially fluidly separated flows.

7. internal combustion engine (10) according to claim 6,

characterized in that

the floods are asymmetrical to each other.

8. internal combustion engine (10) according to any one of claims 6 or 7,

characterized in that

the second turbine (46) has a first number of floods, wherein the first turbine (42) is only fluidically connected to a second number of flows of the second turbine (46) which is smaller than the first number.