EP1507967A2

EP1507967A2 - Method for determining the composition of a gas mixture in a combustion chamber of an internal combustion engine with re-circulation of exhaust gas and a correspondingly embodied control system for an internal combustion engine

Info

Publication number: EP1507967A2
Application number: EP02790353A
Authority: EP
Inventors: Hans-Georg Nitzke; Thorsten Rebohl; Jens Jeschke
Original assignee: Volkswagen AG
Current assignee: Volkswagen AG
Priority date: 2001-11-28
Filing date: 2002-11-11
Publication date: 2005-02-23
Also published as: US7174713B2; WO2003046356A3; EP1701022A3; EP1701025A3; WO2003046356A2; EP1701025A2; EP1701022A2; EP1715163A1; EP1715163A8; US20070012040A1; EP1701025B1

Abstract

The invention relates to an engine management system wherein physically based models (16-21) are used to determine the composition and mass of the fresh air/ exhaust gas mixture suctioned by an internal combustion engine (1). Said models respectively simulate the behavior of the internal combustion engine or corresponding engine system in relation to specific state variables. The individual physically based models (16-21) are closely coupled to each other in a partial manner and are used, for instance, to simulate the filling of the combustion chamber of the internal combustion engine(1) with the suctioned fresh air/waste gas mixture in order to simulate the flow of the mass of re-circulating exhaust gas, in order to simulate the behavior of the exhaust gas manifold of the internal combustion engine (1) upstream and downstream from a turbine (2), in order to simulate the storage behavior of the intake manifold of the internal combustion engine, and to simulate the behavior of the intake pipe or inlet manifold whereby the fresh air/exhaust gas mixture is fed to the combustion engine (1) from a corresponding mixing point (10) where the suctioned fresh air is mixed with the exhaust gas re-circulated via the exhaust gas re-circulation line. As a result, a plurality of additional state variables can be determined without additional sensors.

Description

Method for determining the composition of the gas mixture in one

Combustion chamber of an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine

The present invention relates to a method for determining the composition of the gas mixture in a combustion chamber of an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine, for example a diesel engine.

For the emission-optimal control of a turbocharged diesel engine with exhaust gas recirculation, precise knowledge of the mass and the composition of the intake gas mixture, i.e. of the fresh air / exhaust gas mixture, of crucial importance in the engine combustion chamber.

In known concepts, the fresh air mass in the combustion chamber of the internal combustion engine is determined from the measurement of the fresh air mass flow via a hot film air mass sensor very far forward in the intake tract of the internal combustion engine. Due to the fresh air storage behavior of the intake tract, this determination of the fresh air mass is subject to errors in the dynamic engine operating phases. In addition, such a sensor signal cannot be used to calculate the exhaust gas mass located in the combustion chamber, which has been returned via the exhaust gas recirculation of the internal combustion engine and mixed with the fresh air drawn in at an exhaust gas recirculation mixing point. With conventional concepts, this size cannot be determined.

The present invention is therefore based on the object of proposing a method for determining the composition of the gas mixture in a combustion chamber of an internal combustion engine with exhaust gas recirculation as well as a correspondingly designed control system for an internal combustion engine, with which the exact determination of the composition of the gas mixture in the combustion chamber of the Internal combustion engine, ie an exact determination of the fresh air and exhaust gas mass, especially in the dynamic engine operating phases is possible. This object is achieved according to the invention by a method with the features of claim 1 or a control system with the features of claim 64. The subclaims each define preferred and advantageous embodiments of the present invention.

The invention provides for determining the composition of the gas mixture in a combustion chamber of an internal combustion engine, i.e. to determine the fresh air and exhaust gas mass in this combustion chamber, to determine corresponding state variables of the internal combustion engine by using corresponding physically based models, the individual physically based models simulating the behavior of the internal combustion engine or the engine system in relation to the state variable to be calculated in each case. In this regard, the status variables can include, for example, the fresh air mass flow in the so-called intake manifold of the internal combustion engine, taking into account the storage behavior of the intake tract, the exhaust gas recirculation mass flow, the pressure and the temperature of the intake gas upstream of the intake valves of the internal combustion engine, the pressure and the temperature of the exhaust gas upstream of the turbine, etc. be calculated. The physically based models can also be partially replaced by empirical models if no real-time capable physical model can be determined for the respective model.

With the help of a first model, the filling of the combustion chamber with the fresh air / exhaust gas mixture from the so-called intake manifold, i.e. the connection between the exhaust gas recirculation mixing point and the intake valves of the respective engine, to be able to determine both the fresh air mass flow and the exhaust gas mass flow as well as the fresh air mass in the combustion chamber and the exhaust gas mass in the combustion chamber of the internal combustion engine.

Exhaust gas from the exhaust tract is returned to the intake tract via the exhaust gas recirculation line of the internal combustion engine. In this regard, another model can be used to determine the exhaust gas recirculation mass flow through the exhaust gas recirculation line and the temperature of the recirculated exhaust gas upstream of the exhaust gas recirculation mixing point. A model approach for a throttle point can be used for this.

Another model can be used to simulate the behavior of the exhaust tract before and after the turbine of the engine system. As the most important output variable, this model can be used to determine the exhaust gas back pressure upstream of the turbine, again by calculating the corresponding output or intermediate variables. The storage behavior of the intake tract between the compressor and the exhaust gas recirculation mixing point can also be simulated with the aid of a corresponding model. In this regard, a model can be used which simulates a storage volume for the fresh air and a subsequent throttle point. The essential output variables of this model are the fresh air mass flow through the aforementioned throttle in the intake manifold, ie into the engine intake, the stored fresh air mass in the aforementioned storage volume and the modeled boost pressure etc. can be determined.

The incoming exhaust gas recirculation mass flow mixes in the so-called intake manifold of the internal combustion engine, i.e. the recirculated exhaust gas mass flow to the fresh air / exhaust gas mixture from which the engine draws its fill. The behavior of this intake manifold can also be simulated using a corresponding model, for example to calculate the exhaust gas recirculation mass and the fresh air mass in the intake manifold, which can be done, for example, by evaluating the mass flow balances for the fresh air and the recirculated exhaust gas mass.

Another model can be used to determine the intake manifold temperature, i.e. the temperature of the fresh air / exhaust gas mixture in the intake manifold. In this regard, the intake manifold temperature can be determined in particular as a function of the exhaust gas recirculation mass in the intake manifold and the temperature of the exhaust gas recirculation mass flow as well as the fresh air mass in the intake manifold and the temperature of the inflowing fresh air mass flow.

The individual previously determined models are in some cases closely related, so that the results of another model are preferably accessed for the calculation of a certain state variable in one model. It should be noted that the dependencies of the individual quantities described in the form of this patent application are generally only intended to illustrate the proportional relationships, so that depending on the respective application or implementation, standardization or correction factors (not specified) may be used to further convert the corresponding ones Sizes have to be considered.

Overall, with the help of the present invention, the mass and composition of the gas mixture drawn in by the combustion chamber of the internal combustion engine, for example a diesel engine, can be exactly determined, so that an emission-optimal one Regulation of the internal combustion engine is possible based on the precise knowledge of the mass and the composition of the fresh air mass and the exhaust gas mass in the combustion chamber of the internal combustion engine. A large number of state variables can be obtained from information that is already known without the use of separate or additional sensors, so that new control strategies are made possible.

In addition to the previously explained inventive aspects, the present invention also relates to the following inventive aspects, which are in principle independent of one another and independent of the previously explained inventive aspects. However, it is of course also possible to combine individual or all of the inventive aspects described herein.

I. Method for controlling an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine

Exhaust gas from the exhaust tract is returned to the intake tract via the exhaust gas recirculation line of an internal combustion engine. For the emission-optimal control or regulation of, for example, a turbocharged diesel engine with exhaust gas recirculation, precise knowledge of the largest possible number of state variables or operating parameters is important.

In conventional engine management systems, however, the number of detected or known state variables is relatively small, or separate sensors are required for the determination of the state variables. This applies in particular to the state variables associated with the exhaust tract of internal combustion engines, such as, for example, the exhaust gas back pressure or the exhaust gas temperature before or after the turbine of the respective internal combustion engine, etc. Until now, these state variables could at best only be detected with separate sensors.

A physically based model can be used to implement a method for controlling an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine, with which control of the internal combustion engine is possible depending on state variables of the exhaust gas tract that are as effective as possible and without the need for additional sensors which simulates the behavior of the exhaust tract of the internal combustion engine before and after a turbine assigned to the internal combustion engine. With the help of this model, various state variables associated with the exhaust tract can be extracted from others State variables (already known or already detected) are determined, so that depending on them certain operating parameters of the internal combustion engine, such as the fuel injection quantity or the fuel injection time etc., can be controlled or regulated in an emission-optimized manner.

With the help of the model, for example, the exhaust gas back pressure before or after the turbine, the exhaust gas temperature before or after the turbine or the exhaust gas back pressure of the exhaust gas returned via the exhaust gas recirculation line of the internal combustion engine can be determined from known state variables without the use of additional sensors.

Preferably, several physically based (or also empirically determined) models are used, some of which are closely related, so that the results of another model are preferably used to calculate a certain state variable in one model. It should be noted that the dependencies of the individual quantities described here in terms of formulas are generally only intended to illustrate the proportional relationships, so that depending on the respective application or implementation, standardization or correction factors (not specified) may need to be taken into account for further conversion of the corresponding quantities are.

Overall, state variables of the exhaust tract of an internal combustion engine, for example a diesel engine, can thus be determined precisely and with simple means by evaluating state variables which are already known. The use of additional sensors is not necessary for this. The thus easily possible determination of the state variables of the exhaust tract makes new control and diagnostic methods possible within the respective engine management system, which, for example, permits emission-optimized control of the internal combustion engine.

This aspect of the invention thus includes in particular the following features:

1 . A method for controlling an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas of the internal combustion engine recirculated via the exhaust gas recirculation at a mixing point and the resulting gas mixture is fed to the combustion chamber of the internal combustion engine, the behavior of an exhaust gas tract of the internal combustion engine before and modeled after a turbine assigned to the internal combustion engine and with the aid of the model at least one state variable connected to the exhaust tract at least one further state variable is determined in order to control the internal combustion engine as a function thereof.

2. A method according to FIG. 1, wherein an exhaust gas back pressure in front of the turbine is determined with the aid of the model, for which purpose the blade travel of the turbine is measured or derived from a control pulse duty factor of an actuator provided for adjusting the blades of the turbine.

3. A method according to 1 or 2, wherein the model is used to determine an exhaust gas temperature upstream of the turbine as a function of a fuel mass injected into the combustion chamber of the internal combustion engine and a rotational speed of the internal combustion engine.

4. A method according to 3, wherein to determine the exhaust gas temperature upstream of the turbine, a temperature change based on the temperature of the gas mixture between the mixing point and the internal combustion engine is determined as a function of the injected fuel mass and the speed of the internal combustion engine.

5. A method according to FIG. 4, wherein the temperature change is corrected as a function of a start of delivery of the fuel to be injected in the combustion chamber of the internal combustion engine.

6. A method according to any one of items 1-5, wherein an exhaust gas mass flow emitted by the internal combustion engine is derived using the model from a total mass flow of the gas mixture supplied to the combustion chamber of the internal combustion engine and a fuel mass injected into the combustion chamber of the internal combustion engine.

7. A method according to FIG. 6, wherein an exhaust gas mass flow flowing through the turbine is determined using the model from the exhaust gas mass flow emitted by the internal combustion engine and an exhaust gas recirculation mass flow flowing via the exhaust gas recirculation.

8. A method according to any one of items 1-7, wherein the model is used to determine a rotational speed of an exhaust gas turbocharger shaft coupled to the turbine and to a compressor assigned to the internal combustion engine as a function of a fresh air mass flow flowing through the compressor and a pressure ratio across the compressor. 9. A method according to 8, wherein to determine the pressure ratio above the compressor, a pressure before the compressor from the atmospheric pressure, a measured fresh air mass flow and the atmospheric temperature and a pressure after the compressor from a boost pressure with which the fresh air is supplied from the compressor to the mixing point is determined, the measured fresh air mass flow and a charging temperature at which the fresh air is supplied from the compressor to the mixing point.

10. A method according to 9, wherein the fresh air mass flow flowing through the compressor is determined from the measured fresh air mass flow and the pressure upstream of the compressor with the aid of a standardization based on the atmospheric temperature and a reference temperature of the compressor.

11. A method according to one of items 1-10, wherein with the aid of the model an exhaust gas temperature after the turbine as a function of an exhaust gas temperature before the

Turbine, a temperature change over the turbine and a turbine efficiency is determined.

12. A method according to FIG. 11, wherein the temperature change over the turbine is determined as a function of a pressure ratio over the turbine.

13. A method according to 11 or 12, wherein the efficiency of the turbine is determined as a function of a blade travel of the turbine.

14. A method according to one of items 11-13, wherein the exhaust gas temperature T _nT after the turbine is determined as a function of the exhaust gas temperature T _vT before the turbine, the temperature change ΔT _T over the turbine and the efficiency η _{τ of} the turbine as follows: T _nT = T _vT ^• (1 - ΔT _T ^• ηγ).

15. A method according to one of items 1-14, wherein an exhaust gas back pressure behind the turbine is derived using the model from a pressure difference value, which denotes the difference between the exhaust gas back pressure behind the turbine and the atmospheric pressure.

16. A method according to 15, wherein the pressure difference value is determined as a function of an exhaust gas mass flow through the turbine. 17. A method according to FIG. 16, wherein the pressure difference value is determined as a function of the exhaust gas mass flow through the turbine after the exhaust gas mass flow through the turbine has been corrected multiplicatively using a factor which corresponds to the root of an exhaust gas temperature after the turbine.

18. A method according to any one of items 15-17, wherein the exhaust gas back pressure behind the turbine is determined by adding the pressure difference value to the atmospheric pressure.

19. A method according to 18, wherein a pressure difference in the exhaust tract behind the turbine is measured with the aid of an exhaust gas back pressure sensor, the

Exhaust gas back pressure behind the turbine is determined by adding the pressure difference measured by the exhaust gas back pressure sensor to the first-mentioned pressure difference value and the atmospheric pressure.

20. A method according to any one of items 1-19, wherein the exhaust gas back pressure before the turbine is determined from an exhaust gas back pressure after the turbine, an exhaust gas mass flow flowing through the turbine, a blade travel of the turbine and a rotational speed of an exhaust gas turbocharger shaft coupled to the turbine.

21. A method according to 20, wherein the exhaust gas back pressure before the turbine is determined from the exhaust gas back pressure p _nT after the turbine, the exhaust gas mass flow dm _τ through the turbine, the blade travel s of the turbine and the speed Drehzahl _AT L of the exhaust gas turbocharger shaft as follows:

p _vT = Z ^• p _nT with

Z = b ₀ + bi • dm _τ + b ₂ • (s - 0.5)

+ b ₃ • s ² + b ₄ ^■ (n _ATL - 0.5) ²

+ b ₅ ^■ (dm _τ + 0.5) ^• (s + 0.5)

+ b ₆ - (dm _τ - 0.5) - s ² + b ₇ - (s-1) - (n _ATL - 0.5) ²

+ b ₈ - (s-1) - (s - 0.5) ²

+ b ₉ - (dm _τ - 1) ²

+ bio ^• (dm _τ - 1) ^• (dm _τ + 0.5) ^{2 •} dm _τ

+ b ₁₁ - [(dm _τ - 1) - (s - 0.5) ³ - 0.5] - b ₁₂ + b ₁₃ ,

where b ₀ -b denote ₁₃ coefficients. 22. A method according to any one of items 1-21, wherein with the aid of the model, an exhaust gas back pressure of the exhaust gas recirculated via the exhaust gas recirculation is dependent on an exhaust gas recirculation mass flow flowing via the exhaust gas recirculation, an exhaust gas temperature upstream of the turbine and an exhaust gas back pressure of the exhaust gas emitted by the internal combustion engine the turbine is determined.

23. A method according to 22, wherein the exhaust gas back pressure P _EGR in the exhaust gas recirculation line as a function of the exhaust gas back pressure p _{vT upstream} of the turbine, the exhaust gas recirculation _{mass flow} dm _A GR in the exhaust gas recirculation line and the exhaust gas temperature

T _vT before the turbine is determined as follows:

where PF denotes an exhaust gas back pressure constant.

24. A method according to 23, wherein the exhaust gas back pressure constant is determined as a function of an effective cross-sectional area of the exhaust gas recirculation line.

25. A method according to FIG. 24, wherein the exhaust gas back pressure constant PF is derived from the gas constant R and the effective cross-sectional area A _{eff of} the exhaust gas recirculation line as follows: PF =

A eff

26. A method according to one of items 1-25, the method being carried out automatically by a control unit which is part of an engine management system of the internal combustion engine.

27. A control system for an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine in a mixing point and the resulting gas mixture is fed to a combustion chamber of the internal combustion engine, the control system being designed in such a way that it is physically based Model the behavior of an exhaust tract of the

Internal combustion engine before and after a turbine assigned to the internal combustion engine and with the help of the model automatically at least one with the exhaust tract

^ The associated state variable is determined from at least one further state variable and controls the internal combustion engine automatically as a function thereof. 28. A control system according to FIG. 27, the control system being designed to carry out the method according to one of items 1-26.

II. Another method for controlling an internal combustion engine with exhaust gas recirculation and a correspondingly designed further control system for an internal combustion engine

Via the exhaust gas recirculation line of an internal combustion engine, exhaust gas is returned from the exhaust tract into the intake tract and mixed there with fresh air drawn in via the intake tract. The resulting gas mixture is then fed to the combustion chambers of the internal combustion engine. For the emission-optimal control or regulation of, for example, a turbocharged diesel engine with exhaust gas recirculation, precise knowledge of the largest possible number of state variables or operating parameters is important.

In conventional engine management systems, however, the number of detected or known state variables is relatively small, or separate sensors are required for the detection of the individual state variables. This also applies, for example, to the state variables associated with the intake tract of the respective internal combustion engine and, in particular, state variables there, which are associated with the so-called intake manifold or intake manifold of the internal combustion engine, i.e. the connection between the exhaust gas / fresh air mixing point and the engine intake valves, such as the fresh air or exhaust gas mass or the gas temperature in this connection section. Up to now, these state variables could at best only be recorded with separate sensors.

To implement a method for controlling an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine, with which the internal combustion engine can be controlled as a function of state variables of this connection between the mixing point and the intake valves of the internal combustion engine that are as effective as possible and without the need for additional sensors a physical model based on the connection or the connecting section between the mixing point, at which recirculated exhaust gas is mixed with fresh air drawn in, and the intake valves of the internal combustion engine, which simulates the behavior of this connection, can be used to determine under Using this model to be able to automatically control or regulate various operating parameters of the internal combustion engine in an emission-optimal manner. With the help of the model, for example, the fresh air or Exhaust gas mass in this connection or also the gas temperature in this connection can be determined from known state variables without the use of additional sensors, so that depending on it certain operating parameters of the internal combustion engine, such as the fuel injection quantity or the fuel injection time etc., are controlled or regulated in an emission-optimized manner can be.

The fresh air mass or the exhaust gas mass in the connection can be integrated in time by a difference between a fresh air mass flow supplied to the connection and a fresh air mass flow supplied by the connection to the internal combustion engine or by time integrating a difference between an exhaust gas mass flow supplied to the connection via the exhaust gas recirculation and one of the Connection to the exhaust gas mass flow supplied to the internal combustion engine can be determined.

The total gas mass in the connection can then be determined simply by adding the fresh air mass in the connection and the exhaust gas mass in the connection.

The temperature of the gas mixture supplied to the combustion chambers of the internal combustion engine via the connection can be determined as a function of the exhaust gas mass in the connection, a temperature of the exhaust gas recirculation mass flow returned via the exhaust gas recirculation, the fresh air mass in the connection and a temperature of the fresh air mass flow supplied to the connection, the thus determined temperature value is preferably corrected by an amount which depends on a difference between a wall temperature of the connection and the temperature of the gas mixture in the connection and a factor multiplied by it, the factor in turn on the speed of the internal combustion engine and a via the connection the combustion chambers of the Combustion engine supplied fresh air mass flow depends. For this purpose, the wall temperature of the connection can be derived from a cooling water temperature of the internal combustion engine and a wall heat factor of the connection.

Preferably, several physically based (or also empirically determined) models are used, some of which are closely related, so that the results of another model are preferably used to calculate a certain state variable in one model. It should be noted that the dependencies of the individual quantities described here in terms of formulas are generally only intended to illustrate the proportional relationships, so that depending on the respective application or implementation, standardization or (not specified) Correction factors for the further conversion of the corresponding quantities must be taken into account.

Overall, state variables of the intake tract or the connection between the mixing point and the inlet valves of an internal combustion engine, for example a diesel engine, can thus be determined precisely and with simple means by evaluating state variables that are already known. The use of additional sensors is not necessary for this. As a result of the simple determination of the corresponding state variables, new control and diagnostic methods are possible within the respective engine management system, which allows, for example, emission-optimized control of the internal combustion engine.

1. A method for controlling an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine at a mixing point and the resulting gas mixture is fed to the combustion chamber of the internal combustion engine, the behavior of one of the mixing point through a physically based model to the connection leading to the internal combustion engine, which feeds the gas mixture to the combustion chamber of the internal combustion engine, simulated and with the help of the model at least one state variable related to this connection is determined from at least one further state variable in order to control the internal combustion engine as a function thereof.

2. A method according to 1, wherein a fresh air mass and / or an exhaust gas mass in the connection by time integrating a difference between a fresh air mass flow supplied to the connection and a fresh air mass flow supplied by the connection or by integrating a difference between one of the connections over time the exhaust gas recirculation supplied exhaust gas mass flow and an exhaust gas mass flow supplied by the connection to the internal combustion engine is determined.

3. A method according to FIG. 2, wherein the fresh air mass m _L in the connection and the exhaust gas mass m _EGR in the connection as a function of that of the

Connection supplied fresh air mass flow dm _L , the exhaust gas mass flow dm _EGR supplied to the connection, that supplied from the connection to the internal combustion engine Fresh air mass flow dm _Lmot and the exhaust gas mass flow dm _AGRm ot supplied by the connection to the internal combustion engine can be determined as follows: l + T "m τ = J ( ^{dm ~ dm} Ln, a,) ^{dτ rd} ™ AGR ^{≥ 0}

1

™ L = - ^dm Lmo, + ^dm ΛGR) for ^{ÜV dm} EGR < ⁰ and ™ AGR = ) ^dτ f ^{Ur dm} EGR ≥ ⁰ l + T ₀ ^m EGR = \ ~ ^dm AGRmot ^dτ for ^dm EGR < ⁰ 'where t denotes an integration time and T ₀ an integration ^interval .

4. A method according to 2 or 3, wherein the total gas mass in the connection is determined by adding the fresh air mass in the connection and the exhaust gas mass in the connection.

5. A method according to FIG. 4, wherein the model is used to determine the pressure prevailing in the connection from the total gas mass in the connection, a temperature in the connection and a volume of the connection.

6. A method according to 4 or 5, wherein an exhaust gas recirculation rate is determined by relating the exhaust gas mass in the connection to the total gas mass in the connection.

7. A method according to one of items 1-6, wherein the model is used to determine a temperature of the gas mixture supplied to the combustion chamber of the internal combustion engine via the connection.

8. A method according to 7 and one of items 2-6, wherein with the aid of the model the temperature of the gas mixture supplied via the connection to the combustion chamber of the internal combustion engine as a function of the exhaust gas mass in the connection, a temperature of the exhaust gas recirculation mass flow returned via the exhaust gas recirculation, the Fresh air mass in the connection and a temperature of the fresh air mass flow supplied to the connection is determined. 9. A method according to FIG. 8, wherein the temperature T _{sr of} the gas mixture in the connection as a function of the exhaust gas recirculation mass m _EGR in the connection, the temperature T _AG R of the exhaust gas recirculation _{mass flow,} the fresh air _mass m in the connection and the temperature T _{iad of} the supplied fresh air _{mass flow} is determined as follows:

_{Tir =} T _AGR - m _AGR + T _Iad - m _L j ^ _{+ ≠ Q}

™ AGR + ^m L and

^T sr = ^T lad for ^m EGR + m _L = 0.

10. A method according to any one of items 7-9, wherein the temperature of the gas mixture in the connection is corrected by an amount which depends on a difference between a wall temperature of the connection and the temperature of the gas mixture in the connection and a factor multiplied by it, the factor in turn depends on the speed of the internal combustion engine and a fresh air mass flow supplied to the combustion chamber of the internal combustion engine via the connection.

11. A method according to 10, wherein the wall temperature of the connection is derived from a cooling water temperature of the internal combustion engine and a wall heat factor of the connection.

12. A method according to one of items 1-11, the method being carried out automatically by a control unit which is part of an engine management system of the

Internal combustion engine is running.

13. A control system for an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine in a mixing point and the resulting gas mixture is fed to a combustion chamber of the internal combustion engine, the control system being designed in such a way that it is controlled by a physically based one Model simulates the behavior of a connection leading from the mixing point to the internal combustion engine, which feeds the gas mixture to the combustion chamber of the internal combustion engine, and automatically uses the model to determine at least one state variable related to this connection from at least one further state variable and automatically controls the internal combustion engine as a function thereof. 14. A control system according to 13, wherein the control system is designed to carry out the method according to one of items 1-12.

III. Method for determining the fresh air mass flow of an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine

For the emission-optimal control of, for example, a turbocharged diesel engine with exhaust gas recirculation, precise knowledge of a large number of state variables or operating parameters is important. Via the exhaust gas recirculation line of an internal combustion engine, exhaust gas is returned from the exhaust tract into the intake tract and mixed there with fresh air drawn in, in order finally to supply the fresh air / exhaust gas mixture to the internal combustion engine.

In conventional engine management systems, however, the number of detected or known state variables is relatively small, or separate sensors are required for the detection of the individual state variables. This also applies, for example, to various state variables associated with the intake tract of the internal combustion engine, such as the fresh air mass flow.

In order to implement a method for determining the fresh air mass flow of an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine, with which simple determination of the fresh air mass flow without the use of a separate sensor is possible, a physically based model can be used to determine the fresh air mass flow of the storage behavior of the intake tract of the internal combustion engine between the compressor and the mixing point, at which the intake fresh air is mixed with recirculated exhaust gas, by modeling a storage volume for the fresh air drawn in from the intake tract with a subsequent throttle point, the control of the internal combustion engine automatically using the Model of the fresh air mass flow flowing through the intake tract to the mixing point is determined.

The fresh air mass flow flowing through the intake tract to the mixing point can be determined as a function of the temperature and pressure of the fresh air and the effective cross-sectional area of the throttle point. The pressure of the fresh air can in turn be determined as a function of the fresh air mass located between the compressor and the mixing point and the temperature of the fresh air.

The fresh air mass can be determined by integrating the fresh air mass flow difference between the fresh air mass flow flowing into the compressor and the fresh air mass flow flowing from the compressor to the mixing point.

According to a further exemplary embodiment, the fresh air mass flow can also be determined as a function of the exhaust gas turbocharger speed of the internal combustion engine. In addition to the exhaust gas turbocharger speed, the boost pressure, the atmospheric or ambient pressure and the atmospheric or ambient temperature are also included in the determination of the fresh air mass flow.

Overall, the fresh air mass flow of an internal combustion engine, for example a diesel engine, can thus be determined precisely and with simple means by evaluating already known state variables. The use of additional sensors, in particular a hot-film air-mass meter that is usually required to determine the fresh air mass flow in the intake tract, is not necessary for this. The thus easily possible determination of the exhaust gas recirculation mass flow enables new control and diagnostic methods within the respective engine management system, which allows, for example, emission-optimized control of the internal combustion engine.

1. A method for determining the fresh air mass flow of an internal combustion engine with exhaust gas recirculation, wherein fresh air with a via the exhaust gas recirculation recirculated exhaust gas from the internal combustion engine is mixed at a mixing point and the resulting gas mixture is fed to the combustion chamber of the internal combustion engine, with the aid of a physically based model the storage behavior of an intake tract of the internal combustion engine between a compressor assigned to the internal combustion engine and the mixing point by modeling a storage volume for that of the Intake tract fresh air sucked in with a subsequent throttle point and is determined with the help of the model of the fresh air mass flow flowing through the intake tract to the mixing point.

2. A method according to FIG. 1, wherein the fresh air mass flow flowing via the intake tract to the mixing point is determined as a function of a temperature and a pressure of the fresh air and an effective cross-sectional area of the throttle point.

3. A method according to FIG. 2, wherein the fresh air mass flow dm _{L is determined} as a function of the temperature T _{! A} d of the fresh air, the pressure pia _d od of the fresh air, the effective cross-sectional area A _{dr of} the throttle point and a flow characteristic DF:

4. A method according to 2 or 3, wherein the pressure of the fresh air is determined as a function of a fresh air mass located between the compressor and the mixing point and the temperature of the fresh air.

5. A method according to 4, wherein the pressure pi _{admod of} the fresh air as a function of the fresh air _mass mia _d , the storage volume V and the temperature Tι _{ad of} the fresh air is determined as follows:

Piadmod ~~ ^m iad ^' 2 _v L r ^l lad <

where R denotes a gas constant.

6. A method according to 4 or 5, wherein the fresh air mass is determined by integrating a fresh air mass flow difference between a fresh air mass flow flowing into the compressor and the fresh air mass flow flowing from the compressor to the mixing point. 7. A method according to one of items 1-6, the method being carried out automatically by a control unit which is part of an engine management system of the internal combustion engine.

8. A method for determining the fresh air mass flow of an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas of the internal combustion engine recirculated via the exhaust gas recirculation at a mixing point and the resulting gas mixture is fed to the combustion chamber of the internal combustion engine, a rotational speed of an exhaust gas turbocharger assigned to the internal combustion engine being recorded and the fresh air mass flow flowing to the mixing point is determined from the speed.

9. A method according to 8, wherein the fresh air mass flow is determined as a function of the speed of the exhaust gas turbocharger, a pressure at which the fresh air is fed to the mixing point, an atmospheric pressure and an atmospheric temperature.

10. A control system for an internal combustion engine with exhaust gas recirculation, where fresh air is mixed with an exhaust gas recirculated exhaust gas from the internal combustion engine in a mixing point and the resulting gas mixture is fed to a combustion chamber of the internal combustion engine, the control system being designed in such a way that it is physically based model simulates the storage behavior of an intake tract of the internal combustion engine between a compressor assigned to the internal combustion engine and the mixing point by modeling a storage volume for the fresh air drawn in by the intake tract with a subsequent throttle point and for controlling the internal combustion engine automatically with the aid of the model to the mixing point via the intake tract flowing fresh air mass flow determined in order to control the internal combustion engine.

11. A control system according to 10, wherein the control system is designed to carry out the method according to one of items 1-7.

12. A control system for an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine in a mixing point and the resulting gas mixture is fed to a combustion chamber of the internal combustion engine, the control system being designed such that it leads to the mixing point flowing fresh air mass flow determined from a speed of an exhaust gas turbocharger assigned to the internal combustion engine in order to control the internal combustion engine as a function thereof. 13. Control system according to FIG. 12, the control system being designed such that it determines the fresh air mass flow as a function of the speed of the exhaust gas turbocharger, a pressure at which the fresh air is fed to the mixing point, an atmospheric pressure and an atmospheric temperature.

IV. Method for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine

For the emission-optimal control of, for example, a turbocharged diesel engine with exhaust gas recirculation, precise knowledge of a large number of state variables or operating parameters is important.

In conventional engine management systems, however, the number of detected or known state variables is relatively small, or separate sensors are required for the determination of the state variables. This also applies, for example, to the exhaust gas recirculation mass flow flowing through the exhaust gas recirculation line of an internal combustion engine with exhaust gas recirculation. Exhaust gas from the exhaust tract is returned to the intake tract via the exhaust gas recirculation line of an internal combustion engine.

To implement a method for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation and a correspondingly designed control system for an internal combustion engine, with which simple determination of the exhaust gas recirculation mass flow without the use of a separate sensor is possible, a physically based model can be used, which uses a the exhaust gas recirculation of the internal combustion engine arranged replica exhaust valve as a Drossseistelle replicates. As a result, the exhaust gas recirculation mass flow can be determined as a function of the exhaust gas back pressure and the temperature of the recirculated exhaust gas upstream of the exhaust gas recirculation valve, with a flow characteristic, an effective cross-sectional area of the exhaust gas recirculation valve and the gas constant in particular being taken into account in the physically based model for determining the exhaust gas recirculation mass flow.

The course of the temperature of the exhaust gas recirculated via the exhaust gas recirculation is preferably simulated with the aid of the model in order to derive the current temperature of the recirculated exhaust gas upstream of the exhaust gas recirculation valve. The effective cross-sectional area of the exhaust gas recirculation valve can be adapted as a function of a comparison between the charge pressure measured in each case, with which the fresh air is supplied to the mixing point, and a charge pressure modeled with the aid of a further model by using a corresponding correction factor.

The aforementioned flow characteristic can be derived, for example, from the pressure ratio across the exhaust gas recirculation valve.

Overall, the exhaust gas recirculation mass flow of an internal combustion engine, for example a diesel engine, can thus be determined precisely and with simple means by evaluating already known state variables. The use of additional sensors is not necessary for this. The thus easily possible determination of the exhaust gas recirculation mass flow enables new control and diagnostic methods within the respective engine management system, which allows, for example, emission-optimized control of the internal combustion engine.

1. A method for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated through the exhaust gas recirculation of the internal combustion engine at a mixing point and the resulting gas mixture is fed to the combustion chamber of the internal combustion engine, an exhaust gas recirculation valve arranged in the exhaust gas recirculation using a modeled physically based model for a throttle point and the exhaust gas recirculation mass flow flowing via the exhaust gas recirculation to the mixing point is determined with the help of this model. 2. A method according to 1, wherein the exhaust gas recirculation mass flow is determined as a function of an exhaust gas back pressure and a temperature of the recirculated exhaust gas upstream of the exhaust gas recirculation valve.

3. A method according to FIG. 2, wherein the profile of the temperature of the exhaust gas recirculated via the exhaust gas recirculation is simulated with the aid of the model and the temperature of the recirculated exhaust gas upstream of the exhaust gas recirculation valve is derived therefrom.

4. A method according to 2 or 3, wherein the exhaust gas recirculation mass flow is determined as a function of a flow parameter, an effective cross-sectional area of the exhaust gas recirculation valve and a gas constant.

5. A method according to FIG. 4, wherein the exhaust gas recirculation mass _flow dm _EGR from the effective cross-sectional area A _AG R of the exhaust gas recirculation valve, the exhaust gas back pressure p _EGR , the

Temperature T _{EGR of} the recirculated exhaust gas, the gas constant R and the

Flow parameter DF is determined as follows:

2 ^άra ΑGR = ^A AGR ^• PAGR ^' ι _ ^{' DF} ■

"V ^' EGR where, in the event that the exhaust gas recirculation mass flow flows from an intake tract of the internal combustion engine into an exhaust tract of the internal combustion engine, the charge pressure of the fresh air in the intake tract is used as a value for the exhaust gas counterpressure and the charge temperature of the fresh air in the intake tract as a value for the temperature becomes.

6. A method according to 4 or 5, wherein the effective cross-sectional area of the exhaust gas recirculation valve is dependent on a comparison between a measured one

Boost pressure, with which the fresh air is fed to the mixing point, and a boost pressure modeled with the help of a further model is adjusted by using a corresponding correction factor.

7. A method according to any one of items 4-6, wherein the flow characteristic is derived from a pressure ratio across the exhaust gas recirculation valve.

8. A method according to one of items 1-7, the method being carried out automatically by a control unit which is part of an engine management system of the internal combustion engine. 9. A control system for an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine in a mixing point and the resulting gas mixture is fed to a combustion chamber of the internal combustion engine, the control system being designed such that it is used to control the internal combustion engine automatically determines the exhaust gas recirculation mass flow flowing via the exhaust gas recirculation to the mixing point with the aid of a physically based model which simulates an exhaust gas recirculation valve arranged in the exhaust gas recirculation as a throttle point.

10. A control system according to FIG. 9, the control system being designed to carry out the method according to one of items 1-8.

The present invention is explained in more detail below with reference to the accompanying drawing using a preferred exemplary embodiment.

FIG. 1 shows a simplified representation of a simulation model for simulating the gas flow in a motor vehicle or a corresponding internal combustion engine according to the present invention,

FIG. 2 shows an illustration to explain an engine filling model,

FIG. 3 shows an illustration to explain an exhaust gas recirculation mass flow model,

FIG. 4 shows an illustration to explain a turbine model,

FIG. 5 shows a schematic representation of the intake tract of the internal combustion engine shown in FIG. 1,

FIG. 6 shows an illustration to explain a fresh air mass flow model,

FIG. 7 shows a schematic representation of the intake tract of the internal combustion engine shown in FIG. 1,

FIG. 8 shows an illustration to explain a suction pipe model,

FIG. 9 shows a representation to explain an intake manifold temperature model, and FIG. 10 shows the course of a flow characteristic as a function of a pressure ratio.

1 shows an internal combustion engine 1 with four combustion chambers or cylinders. The internal combustion engine 1 is coupled to an exhaust gas turbocharger (ATL), which comprises a turbine 2 and a compressor 7, the turbine and the compressor 7 being mounted on a common shaft, the so-called turbocharger shaft 14. The turbine 2 uses the energy contained in the exhaust gas of the internal combustion engine 1 to drive the compressor 7, which draws in fresh air via an air filter 6 and presses pre-compressed air into the individual combustion chambers of the internal combustion engine 1. The exhaust gas turbocharger formed by the turbine 2, the compressor 7 and the turbocharger shaft 14 is fluidly coupled to the internal combustion engine 1 only by the air and exhaust gas mass flow.

The air that is sucked in and precompressed by the compressor 7 via the air filter 6 is fed to a so-called replacement volume (ERS) 9 via a charge air cooler (LLK) 8, which reduces the exhaust gas temperature and thus the NO _x emission and the fuel consumption. An inlet manifold (ELS) 10 is connected upstream of the individual combustion chambers of the internal combustion engine 1. The exhaust gas generated in the combustion chambers of the internal combustion engine 1 is collected by an exhaust gas collector (ASA) 11 and fed to the turbine 2. The turbine 2 is followed by the exhaust system (APU) 12 of the motor vehicle in the exhaust gas flow direction, which breaks down the pollutant components of the exhaust gases generated during operation of the internal combustion engine 1 and discharges the remaining exhaust gases as quietly as possible. Part of the exhaust gas generated in the combustion chambers of the internal combustion engine 1 is returned from the exhaust manifold 11 via an exhaust gas recirculation (EGR) to the intake manifold 10 and mixed there with the fresh air drawn in. The reference number 13 denotes valves arranged in corresponding air or gas paths. The reference numeral 15 denotes an actuator for adjusting the vane of the turbine 2.

Furthermore, a control device 4 is shown in FIG. 1, which is a component of a corresponding engine management system of the motor vehicle. Various variables or parameters of the motor system shown are monitored by the control unit 4 and converted into different intermediate and output variables by using corresponding stored physically based models, the variables or parameters monitored by the control unit 4 being fed to the control unit 4 via an interface 3 , The individual variables evaluated by the control device 4 are explained in more detail below with reference to the individual physically based models explained in detail. In particular, the control device 4 determines the mass and composition of the gas mixture in the combustion chambers of the internal combustion engine 1, ie the fresh air and exhaust gas mass therein, and converts them into corresponding control signals for the engine system to achieve emission-optimized control, which - like is indicated in Figure 1 - can be applied to various components of the engine system via the interface 3.

For a stable calculation of the overall model formed by the individual physically based models by the control unit 4, a certain minimum effective computing time, for example in the order of 2 ms, is required for some parts of the overall model. Since this cannot be achieved with conventional control unit concepts, an existing time-synchronous grid is preferably used as the basis and the overall model is calculated several times in this grid (oversampling). For example, in order to achieve an effective computing time of 2 ms for an existing 20 ms grid, the overall model must be calculated ten times within the specified grid. Since the overall model, which is composed of the individual physically based sub-models mentioned above, is used for filling detection of internal combustion engines, i.e. serves for the exact determination of the fresh air and exhaust gas mass in the combustion chambers of the respective internal combustion engine, the overall model can also be referred to as a filling model.

One of these physically based partial models executed by the control device 4 serves to emulate the filling of the respective combustion chamber of the internal combustion engine 1 with the fresh air / exhaust gas mixture from the so-called intake manifold. The intake manifold is the connection between the mixing point 10 shown in FIG. 1, from which the fresh air sucked in via the compressor 7 is mixed with the exhaust gas returned via the exhaust gas recirculation line, and the intake valves of the internal combustion engine 1. This model can therefore also be referred to as an engine filling model.

With the help of this engine filling model, the intake gas mass in the combustion chamber depending on the pressure p _sr and the temperature T _{sr of} the intake gas, which, taking into account the gas constant R, define the density of the intake gas, can be determined in front of the engine intake valves, i.e. in the intake manifold a linear approach depending on the density of the intake gas is chosen:

(1) m _tot = d.dio) + d ₂ (n ₀₎ ^■ - £ Si_. CORR

R Here, m _ge s denotes the gas mass sucked into the combustion chamber, ie the mass of the fresh air / exhaust gas mixture sucked in, n ₀ the (normalized) engine speed and KORR a correction factor, which will be discussed in more detail below. The filling behavior of the internal combustion engine 1 is dependent on the engine speed n ₀ . The coefficients d- and d ₂ are therefore a function of the engine speed n ₀ . This dependence on the engine speed can be represented by quadratic polynomials:

(2) di = ai + a ₂ • n ₀ + a ₃ • n ₀ ² d ₂ = a ₄ + a ₅ ^• n ₀ + a ₆ ^• n ₀ ² .

A aβ denote coefficients of these quadratic polynomials. Optionally, the above-described dependency on the engine speed can also be realized by speed-dependent characteristic curves, it being possible to switch between these alternatives in the control unit 4, for example depending on the current value of a corresponding variable.

The filling of the combustion chamber of the internal combustion engine 1 is composed of fractions of fresh air and recirculated exhaust gas. The gas mass flow dm _{ges drawn in} can be calculated from the previously determined gas mass m _tot in the combustion chamber and the current engine speed n of the internal combustion engine 1. The fresh air mass flow dm _Lm o _t in the internal combustion engine 1 results as a function of the gas mass flow dm _ges and the current exhaust gas recirculation rate Γ _AG R as follows:

(3) drriLmot = (1 - Γ _AG R) ^• dm _tot

The sucked-in gas mass flow dm _ge s or the sucked-in air mass flow dm _Lm o _t are preferably calculated in the unit kg / s. Of course, a conversion to kg / h is also possible.

As a further intermediate _variable , which can serve as the basis for calculating the air ratio in the combustion chamber, the fresh air _mass m _Lmot in the combustion chamber of the internal combustion engine 1 can be determined as follows:

(4) m _Lmo , = (1 - r _EGR ) • m _tot Similarly, the air sucked into the engine 1 exhaust gas recirculation mass flow dm _AG RMOI from the aspirated gas mass flow dm _ges and the actual exhaust gas recirculation rate Γ _AG R can be calculated as follows:

(5) _A dm GR _m ot = IAGR ^• dm _ges

Analogous to the fresh air mass in the combustion chamber of the internal combustion engine, the exhaust gas mass m _AGRm ot in the combustion chamber of the internal combustion engine can also be determined from the already known intake gas mass m _tot :

(6) iriAGRmot - Γ _A QR ^• m tot

The air ratio R _L in the combustion chamber of the internal combustion engine 1 is determined from the fresh air mass m _Lm o _t and the injected fuel mass m _kr as a further output variable from the engine filling model:

(7) R _L = ^ E2∑

^L 14.5 • m kr

With the aid of the aforementioned correction factor KORR, the engine charge model can be adapted to the actual behavior of the internal combustion engine, a comparison being made between a modeled boost pressure pi _admod and a measured actual boost pressure p _{) ad} . This comparison can be carried out in a further partial model, which can be referred to as a correction model. The difference between these two quantities can feed an integrator whose output value gives the proportional correction factor KORR for the total filling of the internal combustion engine 1. For this adaptation process, preferably defined conditions, such as stationary engine operation without exhaust gas recirculation, must exist. For this purpose, the control device 4 can contain a separate function block, which controls the adaptation release, ie the integrator, and for this purpose evaluates certain input variables, which, for example, determine the permitted adaptation range with regard to the injection quantity and speed or monitor the change over time of these variables. In addition, this function block can be supplied with additional parameters, with the aid of which the maximum dynamic range of the fresh air mass flow and the boost pressure can be set, in which case an on and off behavior with hysteresis can preferably be implemented. The output variable KORR of this function block of the control unit 4 corrects the slope of the filling line according to formula (1) and thus adapts the engine filling model to the actual behavior of the internal combustion engine 1.

The engine filling model 16 described above, which is implemented in the control unit 4, is shown schematically in FIG. 2 with regard to its input and output variables.

Exhaust gas recirculation line indicated in FIG. 1 leads, as has already been mentioned, exhaust gas from the exhaust tract back into the intake tract. A further physically based model is therefore provided which calculates the exhaust gas recirculation mass flow through the exhaust gas recirculation line and the temperature of the recirculated exhaust gases upstream of the exhaust gas recirculation mixing point 10, so that this model is also referred to below as an exhaust gas recirculation mass flow model.

The exhaust gas recirculation mass flow dm _EGR is determined with the aid of a model approach for a throttle point of the exhaust gas recirculation valve 13 present in the exhaust gas recirculation line (see FIG. 1) as a function of a flow characteristic DF, an effective cross-sectional area A _{EGR of} the exhaust gas recirculation valve 13, the gas constant R and the exhaust gas counter pressure p _AGR and the temperature T _{EGR upstream} of the exhaust gas recirculation valve 13:

(8) dm EGR = A AGR 'AGR DF

J * - " ^• EGR

To calculate the exhaust gas recirculation mass flow dm _EGR , a case distinction must be made depending on whether the exhaust gas recirculation mass flow flows from the exhaust tract into the intake tract (dm _EGR > 0) or from the intake tract into the exhaust tract (dm _EGR <0). The above formula (8) therefore only applies to the case dm _EGR > 0, while for the case dm _EGR <0 the exhaust gas recirculation mass _flow dm _EGR can be determined as follows:

(9) dm _EGR = A _EGR • p _sr • _ ^• DF sr

The root functions contained in formulas (8) and (9) can preferably be approximated by a quadratic polynomial, which is valid, for example, in the temperature range of 200-1200K of interest here. In order to take into account the inertia of the exhaust gas recirculation in the overall system, the Exhaust gas recirculation mass flow in the exhaust gas recirculation mass flow model of control unit 4 is preferably delayed by a PT1 element.

As already mentioned, not only the exhaust gas recirculation mass _flow dm _EGR , but also the temperature T _AG R of the recirculated exhaust gases upstream of the mixing point with the fresh air is calculated using this model. The temperature T _EGR is required in particular for calculating the exhaust gas recirculation mass _flow dm _EGR (see formula (8)). For the calculation of the temperature T _EGR of the recirculated exhaust gases upstream of the exhaust gas recirculation valve, a distinction must also be made between the forward flow and the reverse flow. The following applies:

(10) T _AGR = T _AG - RF ^■ (T _A G - T _κ ) for dm _EGR > 0

(1 1) T _A GR = T _sr for dmAGR <0

In the case of the forward flow (dm _EGR > 0), hot exhaust gases are led through the exhaust gas recirculation line, while in the case of the backward flow fresh air flows through the exhaust gas recirculation line. The cooling of the hot gases over the exhaust gas recirculation line is simulated according to formula (10) by subtracting 2 RF • (T _AG - T _κ ) from the exhaust gas temperature T _{AG in} front of the turbine, where RF denotes a pipe factor of the exhaust gas recirculation line with its help the cooling can be adapted to the type of exhaust gas recirculation section (for example, differentiation between cooled and uncooled exhaust gas recirculation), while T «corresponds to the cooling water temperature of the internal combustion engine 1 and is thus a measure of the cooling of the exhaust gas temperature T _A G. The exhaust gas temperature T _A G in front of the turbine 2 is generated by a further physically based model, which will be explained in more detail below.

The flow parameter DF required according to the formulas (8) and (9) is a function of the pressure ratio via the throttle point simulated by this exhaust gas recirculation mass flow model, ie above the exhaust gas recirculation valve. Since the flow parameter DF is also used in other models of the overall system, it is preferably also implemented as a separate method that can be called up by the other models. The corresponding method evaluates the pressure in front of the corresponding throttle point and the pressure behind the corresponding throttle point and, depending on this, returns a specific value for the flow characteristic DF. It must be between a so-called supercritical flow case, in which the pressure ratio across the throttle point is less than a predetermined critical pressure ratio, and one subcritical case, in which the pressure ratio is greater than the critical pressure ratio, are distinguished.

The course of the flow _parameter DF as a function of the pressure ratio between the pressure p _vdr before the throttle point and the pressure p _ndr after the throttle point is shown in FIG. 10. It can be seen from FIG. 10 that in the supercritical flow case, which according to FIG. 10 is separated from the subcritical flow case by a dashed line, the flow parameter DF can be equated to a certain maximum value. In the subcritical case, on the other hand, the flow characteristic DF is calculated according to an equivalent function, which corresponds to the curve shape for the subcritical case that continuously decreases as a function of the pressure ratio. A distinction is made in particular between the case of the forward flow and the case of the backward flow. The forward flow can be distinguished from the backward flow, for example by setting a corresponding bit in a corresponding variable.

The effective cross-sectional area A _{EGR of} the exhaust gas recirculation valve is determined with the aid of a characteristic field corrected by a correction factor AKORR, with the control valve 4 optionally using the measured valve lift or the control pulse duty factor of this valve as the input variable of this characteristic field depending on the instantaneous value of a corresponding bit becomes. Which of these input variables is used to determine the effective cross-sectional area of the exhaust gas recirculation valve depends on the type of actuator used. In the case of an electrical exhaust gas recirculation controller, the control pulse duty factor of the control device 4 is used as an input variable for the corresponding characteristic diagram, while in the case of an actuator with charge feedback, the measured valve lift is used as the input variable. In order to take into account the inertia of the exhaust gas recirculation valve during an adjustment, the effective cross-sectional area of the exhaust gas recirculation valve calculated in this way can be delayed by a PT1 element.

The previously mentioned correction factor AKORR can be used to correct the calculated valve cross-sectional area of the exhaust gas recirculation valve as a function of a comparison between the measured and the modeled boost pressure in the stationary operating phases of the internal combustion engine 1, similarly to the previously described engine filling model. In this regard too, an integrator can be used for this purpose, which evaluates the difference between the measured and modeled boost pressure and as Output value provides the correction value AKORR for the calculated cross-sectional area of the exhaust gas recirculation valve.

FIG. 3 schematically shows the exhaust gas recirculation mass flow model 17 described above with its input and output variables.

The behavior of the exhaust gas tract before and after the turbine 2 shown in FIG. 1 can be simulated with the aid of a further model, which is also referred to below as a turbine model. The turbine model determines the exhaust gas back pressure upstream of the turbine 2 as the most important output variable. In addition, further output and intermediate variables are calculated, which will be discussed in more detail below.

Within the turbine model, the blade travel s of the turbine 2 is an important variable for determining the exhaust gas back pressure in front of the turbine 2. The blade travel s can either be measured directly in combination with a corresponding analog / digital conversion or via the control duty cycle of the one shown in FIG Actuator 15 are determined. The undelayed blade travel s can be determined via this control pulse duty cycle by accessing a corresponding characteristic curve which assigns each value of the control pulse duty factor to a corresponding value of the blade stroke s of the turbine 2. The dynamics of the blade movement of the turbine 2 are preferably taken into account by a PT1 element in order to be able to emulate the time behavior of the blade path s as well as possible.

The exhaust gas temperature T _{AG in} front of the turbine 2 is dependent on the injection quantity m _kr and the engine speed n ₀ (normalized engine speed) or n (non-normalized engine speed) via a differential temperature approach between the exhaust gas temperature in front of the turbine 2 and the intake manifold temperature, i.e. the temperature in the intake tract. The

Differential temperature, i.e. the temperature increase due to the combustion in front of the turbine

2, determined via a map as a function of the engine speed and the injection quantity or injected fuel mass. The one obtained in this way

Differential temperature value ΔT1 _ASA can depend on the start of delivery, ie

Start of fuel injection into the corresponding combustion chamber of the

Internal combustion engine 1, multiplicatively corrected to a final value for the

Differential temperature ΔT _ASA . ie for the temperature increase due to the combustion in front of the turbine 2:

(12) ΔTASA = ΔT1ASA - ΔT2 _AS A Alternatively, an additive correction can also be made:

(13) ΔT _AS A = ΔT1ASA + ΔT2 _ASA

The differential temperature correction value ΔT2 _ASA is determined with the aid of a further characteristic curve as a function of the start of delivery FB. Switching between the two alternatives mentioned above (compare formulas (12) and (13)) can take place depending on the position of a corresponding switch or bit.

The exhausted exhaust gas mass flow dm _A s _{A of} the internal combustion engine 1 is calculated from the gas mass flow dm _{ges drawn in} by the internal combustion engine 1 or the corresponding combustion chamber and the injected fuel mass flow dm _kr or a proportion dependent on the injected fuel mass m _kr and the engine speed n:

(14) _A dm = sA dm _ges + _kr = dm dm _ges + f (n, m _kr)

The gas mass flow dm _τ through the turbine 2 can be determined from the exhaust gas mass flow dm _A s _A emitted by the internal combustion engine 1 and the exhaust gas recirculation mass flow dm _EGR :

(15) dmτ = dm _A sA - dm _A GR

Furthermore, an exhaust gas turbocharger or compressor speed n _v related to the compressor 7 can be determined with the aid of a map as a function of the fresh air mass flow dm through the compressor 7 and the pressure ratio above the compressor 7. To calculate the pressure ratio above the compressor 7, the pressure downstream of the compressor 7 and the pressure upstream of the compressor 7 are determined in order to subsequently calculate the pressure ratio above the compressor 7. The pressure p _vV before the compressor 7 or p _n after the compressor 7 can be as follows from the atmospheric pressure p _A , the fresh air mass flow dm _H F _M> measured by the hot-film air mass sensor and _supplied on the _{inlet side} in the model shown in FIG. 1, the atmospheric temperature T _A , the charge pressure p _ad and the charge temperature T _{ad are} determined: p _vV = p _A - ^dm »™ ^'T *. VFAKl

(16)

_PnV = P _lad - ^{dmHlFM 'Tl} ° ^d - VFAK2

According to equation (16), a loss factor VFAK1 or VFAK2 is used in order to take into account the pressure loss before and after the compressor 7, which in each case is formed by forming the quotient from the gas constant R and the square of a corresponding one

_{Replacement area} A ² _vV or A _nV can be determined.

The fresh air mass flow dm _v through the compressor 7 is defined as follows:

Here K denotes a constant and T _o a reference or reference temperature of the compressor 7, which is used in the measurement of the compressor maps. The exhaust gas turbocharger speed n _{AT is} calculated from the exhaust gas turbocharger speed n _v related to the compressor 7 as a function of the ambient or

Atmospheric temperature T _A and the reference temperature T _{o of} the compressor 7 as follows:

For reasons of computing time, the root function contained in formulas (17) and (18) can be calculated using a quadratic polynomial as a function of T _A / T ₀ v.

As a further output variable, the temperature T _nT in the exhaust tract behind the turbine 2 is calculated using the turbine model. This takes place as a function of the temperature T _{vT of} the turbine 2 by simulating the temperature drop above the turbine 2, the turbine efficiency ηr also being taken into account as follows:

(19) T _nT = T _vT - (1 - ΔT _τ - ητ)

The temperature change ΔT _T across the turbine 2 is determined with the aid of a corresponding characteristic curve as a function of the pressure ratio across the turbine 2, ie the ratio between the pressure p _vT before the turbine and the pressure p _nT after the turbine, while the efficiency η _{τ of} the turbine 2 is applied with the aid of a corresponding characteristic curve as a function of the blade travel s of the turbine 2. The temperature T _{vT upstream} of the turbine 2 corresponds to the previously determined value T _AG , ie the exhaust gas temperature _{upstream of} the turbine 2. Likewise, the pressure p _{vT upstream} of the turbine 2 corresponds to the modeled exhaust gas back pressure p _AG _{upstream of} the turbine 2.

The exhaust gas back pressure p _n τ behind the turbine is calculated as a further variable, a pressure difference between the exhaust gas tract behind the turbine 2 and the atmospheric pressure p _{A being} determined for this purpose. This can also be done via a corresponding characteristic curve, the gas mass flow dm _τ through the turbine 2 being used as the input variable for this characteristic curve, which is corrected multiplicatively as follows by the root from the exhaust gas temperature T _nT after the turbine 2:

(20) dm ^* = dm _τ ^• / ϊ ^

Depending on the corrected gas mass flow dm * _τ through the turbine 2, the pressure difference Δp _n τ between the exhaust tract behind the turbine 2 and the atmospheric pressure p _A can be determined using a quadratic equation as a function of Δp _nT , the coefficients of this quadratic equation are applicable. The exhaust gas back pressure p „τ after the turbine 2 (in bar) results for the case that there is no exhaust gas back pressure sensor in the exhaust tract after the turbine 2, as follows from the addition of the atmospheric pressure p _A and the calculated pressure difference Δp _n τ- '

(21) p _nT = (Δp _nT + P _A ) / 10 ⁵

If, on the other hand, an exhaust gas back pressure sensor is provided in the exhaust tract or a differential pressure sensor is located behind the turbine 2, the pressure difference Δp _AG measured by this exhaust gas back pressure sensor is added to the modulated exhaust gas back pressure behind the turbine 2:

(22) p _nT = (Δp _nT + p _A + Δp _AG ) / 10 ⁵

The exhaust gas back pressure p _vT before the turbine 2 can be determined from the exhaust gas back pressure p _nT after the turbine 2 with the aid of a polynomial with 13 coefficients depending on the

Input variables turbine mass flow dm _τ , blade travel s and exhaust gas turbocharger speed n _A τ _{L are} calculated, the three last-mentioned variables preferably using corresponding applicable parameters can be used standardized. An exemplary and preferred calculation _rule for determining the exhaust gas back pressure p _{vT upstream} of the turbine 2 is given below, but in principle any combinations of the input variables are possible:

(23) p _vT = Z • p _nT

Z = b ₀ + bi ^■ dm _τ + b ₂ • (s - 0.5) + b ₃ ^■ s ² + b ₄ • (n _ATL - 0.5) ² + b ₅ • (dm _τ + 0.5 ) • (s + 0.5) + b ₆ - (dm-r - 0.5) • s

+ b ₇ - (s-1) - (n _ATL - 0.5) ²

+ b ₈ - (s-1) (s - 0.5) ²

+ b ₉ - (dm _τ - 1) ²

+ b ₁₀ ^• (dm _τ - 1) ^■ (dm _τ + 0.5) ^{2 ■} dm _τ + bn • [(dm _τ - 1) • (s - 0.5) ³ - 0.5] • b ₁₂

+ b ₁₃

The coefficients b ₀ - b ₁₃ are preferably variable.

The exhaust gas back pressure p _{EGR upstream} of the exhaust gas recirculation valve 13 shown in FIG. 1 is calculated as a further output variable. It results as follows depending on the exhaust gas back pressure in front of the turbine p _vT , the exhaust gas recirculation _{mass flow} dm _EGR , the exhaust gas temperature in front of the turbine T _vT and a constant PF:

(24) p EGR = f PvT dm EGR ^l vT

In the formula (24), the exhaust gas back pressure p _v τ _{upstream of} the turbine and the exhaust gas temperature T _{vT upstream} of the turbine are used, preferably with the aid of a PT1 element, delayed or filtered.

With this approach, a pressure drop in the exhaust gas recirculation line upstream and downstream of the exhaust gas recirculation valve is taken into account. The pressure drop can be applied via the effective cross-sectional area A _{eff of} the exhaust gas recirculation line (without exhaust gas recirculation valve). During an initialization phase of the control unit 4, the parameter PF can be calculated from this as follows, where R denotes the gas constant: R

(25) PF =

A eff

The turbine model 18 explained in detail above is shown schematically in FIG. 4 with regard to its input and output variables.

Another physically based model serves to simulate the storage behavior of the intake tract between the compressor 7 shown in FIG. 1 and the exhaust gas recirculation fresh air mixing point 10 also shown in FIG. 1. This model is also referred to below as the fresh air mass flow model and consists of the simulation of a storage volume V _L for the sucked-in fresh air and a subsequent throttle point with the effective cross-sectional area A _dr , as shown in FIG. 5.

The output quantities of this fresh air mass flow model include, in particular, the fresh air mass flow dm _L through the aforementioned throttling point in the intake manifold, ie into the engine intake, the stored fresh air mass m _Ls in the storage volume between the compressor 7 and the exhaust gas recirculation fresh air mixing point 10 and the modeled boost pressure p ^ _mc - _d determined. In addition, the difference Δdm between the measured fresh air mass flow dm _{HF of} the hot-film air mass sensor and the fresh air mass flow dm _L flowing into the internal combustion engine 1 is determined.

The modeled boost pressure pia _d mo _d can be as follows from the fresh air mass m _ad in the volume between the compressor 7 and the intake manifold or the engine intake and a measured charging temperature T | be calculated _ad of fresh air:

(26) p _{lad raod} = m _lad • - • T _lad

The charge air temperature Tι _a is preferably used PT1 -filtered.

The fresh air mass flow dm _L in the intake manifold can be as follows depending on the PT1 -filtered, measured charge air temperature Tι _ad , the modeled charge pressure Pia _d mo _d , the gas constant R, the modeled intake manifold pressure p _sr , i.e. the pressure of the intake gas in front of the intake valves Internal combustion engine 1, and the effective cross-sectional area A _{dr of} the throttle valve in front of the exhaust gas recirculation fresh air mixing point can be determined: (27) dm _L = A dr lad raod DF

R '-lad

The fresh air mass flow dm determined in this way can also be filtered with the aid of a corresponding PT1 element in order to emulate the inertia of the fresh air mass flow. The time constants used in PT1 filtering, which simulate the inertia of the fresh air mass flow for a positive or negative change, should be as small as possible (e.g. <20 ms). The root in formula (27) can again be approximated by a third-order polynomial. As has already been described on the basis of the previously explained engine filling model, the flow parameter DF is again determined by a corresponding function call.

The effective cross-sectional area A _{dr of} the throttle point is a function of the control pulse duty factor of the control device 4, which is likewise delayed by a PT1 element, in which case the time constants of the PT1 element should be selected such that they largely correspond to the time constants for opening and closing correspond to the throttle valve.

From the mass flow balance of the volume between the compressor 7 and the intake manifold or engine inlet of the internal combustion engine 1, the fresh air mass mi _ad results from the integration of the differential mass flow Δdm between the inflowing, measured fresh air mass flow dm _HFM and the outflowed, modeled fresh air mass flow dm _L into the intake manifold:

t + τ ₀ _{ml art} = | Δdm _L dτ (28) ^J

Δdm _L = dm _Hm - dm _L

T ₀ denotes the time integration interval selected in each case. The fresh air mass m ^ _d obtained in this way between the compressor and the intake manifold of the internal combustion engine is preferably limited to a minimum value and a maximum value via the corresponding integrator output. The integrator time constant can preferably be set variably with the aid of a corresponding parameter.

The fresh air mass m ^ _d determined in this way forms - as has been described above - the basis for determining the modeled boost pressure Pia _d mo _d according to formula (26) by applying the ideal gas law. The fresh air mass flow model 19 explained in detail above is shown schematically in FIG. 6 with regard to its input and output variables.

The behavior of the intake manifold, ie the connection between the exhaust gas recirculation / fresh air mixing point and the engine intake valves, is simulated with the aid of another model, the intake manifold also being modeled by a container with a volume V _sr . This container is referred to below as the intake manifold, so that the corresponding model can be referred to as the intake manifold model. A schematic representation of the intake tract intake pipe is shown building on the schematic representation of the intake tract shown in FIG. 5 in FIG.

In the intake manifold, the incoming exhaust gas recirculation mass flow dm _EGR and the fresh air mass flow dm _L mix to form a fresh air / exhaust gas mixture from which the internal combustion engine 1 draws its charge. The exhaust gas recirculation mass and the fresh air mass in the intake manifold can be calculated from the mass flow balances for the fresh air and the recirculated exhaust gas mass by integration.

The fresh air mass m results from the integration of the difference between the incoming and outgoing fresh air mass flow into the intake manifold:

t + τ- m _L J (dm _L - dm ^ dτ for dm _EGR > 0

(29) ^t t ι-T ₀ m _L = J (dm _L - dm ^ + dm _EGR ) for dm _EGR <0

As can be seen from formula (29), a case distinction is made for the calculation of the fresh air mass m _L depending on whether exhaust gas in the exhaust pipe (dm _EGR ≥ 0) or fresh air from the intake pipe into the exhaust tract (dm _EGR <) 0) flows. Within the intake manifold model, the integrator outputs and thus the calculated fresh air mass rm _{L are} preferably limited to a minimum value and to a maximum value.

The calculation of the returned exhaust gas mass ITI _A GR in the intake manifold is carried out analogously to the calculation of the fresh air mass. By integrating the difference between the incoming exhaust gas recirculation mass flow dm _A G _R and the exhaust gas recirculation mass flow dm _EGR ot flowing into the engine, the mass m _{EGR of} the recirculated exhaust gases in the intake manifold results: ^m _EGR = - dm _AGanol ) dτ ßr dm _EGR ≥ O

< ³⁰ >

'AGR = \ - ^dm _AGRmo , ^dτ far ^d _EGR <0 and m _EGR > 0

In the event of a negative exhaust gas recirculation mass flow, it is assumed in a simplistic manner that only the fresh air flows into the exhaust tract via the exhaust gas recirculation line, i.e. it is assumed: dm _EGR = 0. The mass of the recirculated exhaust gases is in turn limited to a minimum value and a maximum value via the integrator output ,

The time constants of the integrators used for the calculation of the air mass and the recirculated exhaust gas mass in the intake manifold and their ranges of validity can preferably be changed via corresponding parameters.

The total gas mass m _sr then results from the addition of the fresh air mass m and the exhaust gas mass m _EGR in the intake manifold. In an initialization phase of the control device 4, an initial value can be calculated in each case for the fresh air mass m _L and the exhaust gas mass m _{EGR as} a function of a predeterminable temperature and a predefinable pressure.

The pressure p _sr in the intake manifold depends on the volume V _sr from the ideal gas law:

(31) p "- - ^m "^{'R' T} "

V.

Finally, as another output variable of the intake manifold model, the exhaust gas recirculation rate Γ _EGR is calculated from the mass fractions in the intake manifold as follows:

m

(32) r EGR

EGR m "

The intake pipe model 20 explained in detail above is shown schematically in FIG. 8 with regard to its input and output variables. In a further physically based model, which is implemented in the control unit 4, the intake manifold temperature T _{sr of} the fresh air / exhaust gas mixture in the intake manifold is determined.

The intake manifold temperature T _sr is determined as a function of the exhaust gas recirculation mass m _EGR in the intake manifold and the temperature T _{EGR of} the exhaust gas recirculation _mass flow as well as the fresh air mass m _L in the intake manifold and the (PT1-delayed) temperature T _{lad of} the inflowing fresh air _mass flow:

T

(33) γ _ AGR m AGR + T l, ad m, for m _AGR + m _L ≠ O AGR ^{+ m} L

In the event that the denominator of formula (33), ie the total mass in the intake manifold, corresponds to the value 0, the intake manifold temperature T _{sr is set} to the value of the temperature T _{! Ad of} the incoming fresh air mass flow:

(34) T _sr = Ta _d for m _EGR + m _L = 0

The wall heat transfers in the intake manifold lead to cooling or heating of the fresh air / exhaust gas mixture. This temperature change can be taken into account by adding an additional term ΔT _sr .

(35) T _sr = T _sr + ΔT sr

The temperature change ΔT _sr depends on the wall temperature T _{w of} the intake manifold and the temperature T _{sr of} the fresh air / exhaust gas mixture in the intake manifold. The heating or cooling effect due to the wall heat transfer also depends on the current engine operating point. This can be taken into account with the aid of a map in the calculation of the temperature change ΔT _sr , in which a factor can be adapted as a function of the speed n and the fresh air mass _flow dm _Lm o _{t drawn} into the combustion chamber of the internal combustion engine 1:

(36) ΔT _sr = (T _w - T _sr ) • f (n, dm _Lmot )

From the formula (36) it can be seen that the temperature change ΔT _sr from a

Differential value of the wall temperature T _{w of} the intake manifold and the temperature T _{sr of} the fresh air / exhaust gas mixture in the intake manifold is calculated, this differential value being multiplied by a factor which is the engine speed n and the intake Fresh air mass flow dm _L mo _{t of} the internal combustion engine 1 is dependent. The wall temperature T _{w of} the intake manifold can be calculated from the PT1 -filtered cooling water temperature T _κ and an applicable wall heat factor WF:

(37) Tw = T _κ ^■ WF

The output variable, ie the intake manifold temperature T _sr , of the intake manifold temperature model is in turn preferably determined by a PT1 element with a time delay.

The previously explained intake manifold temperature model 21 is shown schematically in FIG. 9 with regard to its input and output variables.

As has already been explained above, further physically based models or functions or methods can be provided, which can be called up from the individual models described in detail above, in order to be able to determine specific sizes depending on the parameters transferred in each case. Such a function or method can - as has already been explained - be provided for determining the flow parameter DF.

In addition, a function for converting certain quantities and for providing certain constants and parameters can be provided, which the individual models can access. To improve clarity, sizes in SI units are mainly used in the overall model. For this reason, preprocessing or conversions of the sizes used in each case are required for the individual partial or sub-models, which can be performed by this function block. Separate sections or methods can be provided for the processing of speed-synchronous variables and time-synchronous variables. An example of such a conversion is the conversion of the input signals for position measuring systems. If displacement measuring systems are used for the exhaust gas recirculation valve and vane position, which require signal processing in the control unit 4, additional characteristic curves must be provided, which allow the voltage signals to be converted into the corresponding path. Another example of such a conversion is the consideration of the hysteresis μ and loose of the blade adjustment system of the turbine 2. Due to the design, the blade adjustment system has a hysteresis behavior which is reduced by the loose between the guide pins of the respective control linkage and the guide blades on the adjustment ring of the turbine 2. This can result in dead paths when the direction is reversed Result in control linkage in which there is no blade adjustment. In order to take this behavior into account, the blade path determined by a path system is preferably shifted on one side depending on the direction, the dead path being able to be set by application using a corresponding parameter.

Another centrally provided method or a further centrally provided function block can be provided to implement the PT1 filtering of various sizes which has already been explained above. For this purpose, this method is implemented in such a way that it is called with two time constants, which are switched depending on the input signal direction (ascending or descending). In addition to specifying the respective input signal, this method is also preferably called up with a parameter which describes the respective time grid. The return value of this method or function block is then the PT1 filtered input signal.

REFERENCES IC H EN L ISTE

1 internal combustion engine

2 turbine

3 interface

4 control unit

5 injection system

6 air filters

7 compressors

8 intercoolers

9 replacement volume

10 inlet collectors

11 exhaust gas collectors

12 exhaust system

13 valve

14 turbocharger shaft

15 Turbine guide vane adjustment

16 Engine filling model

17 Exhaust gas recirculation mass flow model

18 turbine model

19 Fresh air mass flow model

20 intake manifold model

21 Intake manifold temperature model

Psr pressure of the fresh air / exhaust gas mixture in the intake manifold

Tsr temperature of the fresh air / exhaust gas mixture in the intake manifold n ₀ normalized engine speed

TAGR exhaust gas recirculation rate a, coefficient m _kr injected fuel mass

KORR correction factor m _g _total suctioned _gas mass in the combustion chamber of the internal combustion engine dm _total suctioned _gas mass flow in the combustion chamber of the internal combustion engine dm _Lmot air mass flow sucked in the combustion chamber of the internal combustion engine mι_mot fresh air mass in the combustion chamber of the internal combustion engine dm _A GRmot exhaust gas recirculation mass flow drawn into the combustion chamber of the internal combustion engine m _A GRmot exhaust gas mass in the combustion chamber of the internal combustion engine

R Air ratio PAGR exhaust back pressure

AAGR cross-sectional area of the exhaust gas recirculation valve

DF flow characteristic

DAY exhaust gas temperature

RF pipe factor T _κ cooling water temperature

AKORR correction factor dm _AG R exhaust gas recirculation mass flow

TAGR Temperature of the exhaust gas recirculation mass flow n Engine speed FB Start of delivery

T _A atmospheric temperature p _A atmospheric pressure

Tia _d charging temperature of the fresh air mass flow

Pia _d boost pressure of the fresh air _{mass flow} A _eff effective cross-sectional area of the exhaust gas recirculation line s blade path of the turbine

T _vT exhaust gas temperature in front of the turbine

T _nT exhaust gas temperature behind the turbine dm _A s _A exhaust gas mass flow dm-r exhaust gas mass flow through the turbine p _vV pressure _upstream of the compressor p _nV pressure downstream of the compressor

TL _A TL exhaust gas turbocharger speed dm _v fresh air _{mass flow} through the compressor p _nT pressure behind the turbine p _vT pressure _upstream of the turbine

Δp _AG pressure difference in the exhaust tract dm _H F _M fresh air mass flow measured by a hot film air mass sensor

V _L storage volume of the fresh air mass flow A _dr cross-sectional area of a throttle point m _Ls stored fresh air mass in the storage volume

Pia _d mo _d modeled boost pressure Δdm difference between the measured fresh air mass flow and the

Fresh air mass flow from the storage volume mia _d fresh air mass between the compressor and the intake manifold of the

Internal combustion engine m _sr mass of the fresh air / exhaust gas mixture in the intake manifold

V _sr intake manifold volume m _L fresh air mass in the intake manifold

ΓΠ _A GR exhaust gas mass in the intake manifold

WF wall heat factor of the intake manifold p _ndr pressure behind a throttle point p _vdr pressure before a throttle point

Claims

PATE N TA NSPRÜ CHE

1. Method for determining the composition of the gas mixture in a combustion chamber of an internal combustion engine with exhaust gas recirculation, fresh air being mixed with an exhaust gas of the internal combustion engine (1) returned via the exhaust gas recirculation at a mixing point (10) and the resulting gas mixture being fed to the combustion chamber of the internal combustion engine (1) is supplied, characterized in that the composition of the gas mixture in the combustion chamber of the internal combustion engine (1) while determining corresponding state variables of the internal combustion engine (1) by using corresponding physically based models (16-21), which in relation to the state variable to be determined in each case Emulate behavior of the internal combustion engine (1) is determined.

2. The method according to claim 1, characterized in that the composition and mass of the gas mixture in the combustion chamber of the internal combustion engine (1) is determined by using the physically based models (16-21).

3. The method according to claim 1 or 2, characterized in that the filling of the combustion chamber of the internal combustion engine (1) with the gas mixture supplied by the mixing point (10) is simulated with a model (16).

4. The method according to claim 3, characterized in that using the model, the total mass (m _tot ) of the gas mixture in the

Combustion chamber of the internal combustion engine (1) is determined as a function of the pressure (p _sr ) and the temperature (T _sr ) of the gas mixture upstream of the internal combustion engine (1).

5. The method according to claim 4, characterized in that the total mass (m _tot ) of the gas mixture in the combustion chamber of the internal combustion engine (1) using a linear approach depending on the Pressure (p _sr ) and the temperature (T _sr ) of the gas mixture in front of the internal combustion engine (1) are determined using coefficients which are dependent on the speed (n) of the internal combustion engine (1).

6. The method according to claim 5, characterized in that the coefficients are determined by quadratic polynomials as a function of the speed (n) of the internal combustion engine (1).

7. The method according to claim 5, characterized in that the coefficients are determined with the aid of characteristic curves which are dependent on the speed (s) of the internal combustion engine (1).

8. The method according to claim 6 or 7, characterized in that when determining the total mass (m _tot ) depending on the pressure (p _sr ) and the temperature (T _sr ) upstream of the internal combustion engine (1) in a correction factor (KORR) in the linear approach for adapting the model (16) to the actual behavior of the internal combustion engine (1) is used.

9. The method according to claim 8, characterized in that the correction factor (KORR) as a function of a difference between a measured pressure (pι _ad ) of the fresh air and a pressure determined using a further physically based model (19) (pι _adm0 d) the fresh air is set.

10. The method according to any one of claims 4-9, characterized in that from the total mass (m _tot ) with the help of the model (16)

Total gas mass flow (dm _tot ) of the gas mixture in the combustion chamber of the internal combustion engine (1) is determined.

11. The method according to claim 10, characterized in that, depending on the total gas mass flow (dm _total ) in the combustion chamber of the internal combustion engine (1), the fresh air mass _flow (dm _Lm o _t ) and Exhaust gas recirculation _{mass flow} (dm _EGRmot ) in the combustion chamber of the internal combustion engine (1) can be determined taking into account an exhaust gas recirculation rate (Γ _EGR ).

12. The method according to any one of claims 4-11, characterized in that depending on the located in the combustion chamber of the internal combustion engine (1) the total mass (m _{ges) of} the fresh air mass (m _Lmo ι) and the exhaust gas mass (m _AGRmot) in the combustion chamber of the internal combustion engine (1) can be determined taking into account an exhaust gas recirculation rate (r _EGR ).

13. The method according to claim 12, characterized in that from the fresh air mass (mo _t ) in the combustion chamber of the internal combustion engine (1) and in the combustion chamber of the internal combustion engine (1) injected fuel mass (m _kr ) an air ratio (R) in the combustion chamber of the Internal combustion engine (1) is determined.

14. The method according to any one of the preceding claims, characterized in that with the aid of a physically based model (17) the exhaust gas recirculation mass flow (dm _EGR ) flowing via the exhaust gas recirculation to the mixing point (10) is simulated.

15. The method according to claim 14, characterized in that the exhaust gas recirculation mass flow (dm _EGR ) as a function of a flow characteristic (DF), an effective cross-sectional area (A _EGR ) of an exhaust gas recirculation valve (13) contained in the exhaust gas recirculation, a gas constant (R) and one Exhaust gas back pressure (P _AG R) and a temperature of the recirculated exhaust gas (T _EGR ) _upstream of the exhaust gas recirculation valve (13) is determined.

16. The method according to claim 15, characterized in that the course of the temperature of the exhaust gas recirculated via the exhaust gas recirculation is simulated with the aid of the model (17) and from this the temperature (T _A G _R ) of the recirculated exhaust gas upstream of the exhaust gas recirculation valve (13) is derived ,

17. The method according to claim 15 or 16, characterized in that the exhaust gas recirculation mass flow dm _EGR from the effective cross-sectional area A _{EGR of} the exhaust gas recirculation valve (13), the exhaust gas back pressure P _A G _R , the temperature T _EGR , the gas constant R and the flow characteristic DF as follows is determined:

in the event that the exhaust gas recirculation mass flow flows from an intake tract of the internal combustion engine (1) into an exhaust tract of the internal combustion engine (1) as

Value for the exhaust gas counterpressure is the boost pressure of the fresh air in the intake tract and the value for the temperature is the charge temperature of the fresh air in the intake tract.

18. The method according to any one of claims 15-17, characterized in that the effective cross-sectional area (A _AG R) of the exhaust gas recirculation valve (13) as a function of a comparison between a measured boost pressure (pι _ad ) with which the fresh air of the mixing point (10 ) is supplied, and a boost pressure (p ^ _d m _od ) modeled with the aid of a further model (19) is adapted by using a corresponding correction factor (AKORR).

19. The method according to any one of claims 15-18, characterized in that the flow characteristic (DF) from the pressure ratio above the

Exhaust gas recirculation valve (13) is derived.

20. The method according to any one of the preceding claims, characterized in that with the help of a physically based model (18) the behavior of a

Exhaust tract of the internal combustion engine (1) is simulated before and after a turbine (2) assigned to the internal combustion engine (1).

21. The method according to claim 20, characterized in that to determine an exhaust gas back pressure (P _A G) upstream of the turbine (2), the blade travel (s) of the turbine (2) is measured or derived from a control pulse duty factor of an actuator (15) provided for adjusting the blades of the turbine (2) ,

22. The method according to claim 20 or 21, characterized in that with the aid of the model (18) an exhaust gas temperature (T _A G) upstream of the turbine (2) as a function of the fuel mass (m _kr. ) Injected into the combustion chamber of the internal combustion engine (1) ) and the speed (s) of the internal combustion engine (1) is determined.

23. The method according to claim 22, characterized in that for determining the exhaust gas temperature (T _A G) upstream of the turbine (2) to the temperature (T _sr ) of the gas mixture between the mixing point (10) and the

Internal combustion engine (1) -related temperature change depending on the injected fuel mass (m _kr ) and the speed (n) of the internal combustion engine is determined.

24. The method according to claim 23, characterized in that the temperature change is corrected as a function of a start of delivery (FB) of the fuel to be injected in the combustion chamber of the internal combustion engine (1).

25. The method according to any one of claims 20-24, characterized in that an exhaust gas mass flow (dm _A s _A ) ejected by the internal combustion engine (1) with the aid of the model (18) from a total mass flow (dm.) Supplied to the combustion chamber of the internal combustion engine (1) _ges ) of the gas mixture and a fuel mass (m _kr ) injected into the combustion chamber of the internal combustion engine (1).

26. The method according to claim 25, characterized in that with the aid of the model (18) from the exhaust gas mass flow (dm _A s _A ) expelled by the internal combustion engine (1) and one via the exhaust gas recirculation flowing exhaust gas recirculation mass flow (dm _EGR ) an exhaust gas mass flow (dm _τ ) flowing through the turbine (2) is determined.

27. The method according to any one of claims 20-26, characterized in that with the help of the model (18) a speed (Π _ATL ) of an exhaust gas turbocharger shaft coupled to the turbine (2) and a compressor (7) associated with the internal combustion engine (1) ( 14) as a function of a fresh air mass flow (dm _v ) flowing through the compressor (7) and a pressure ratio (p _vV , p _n v) above the compressor (7).

28. The method according to claim 27, characterized in that to determine the pressure ratio above the compressor (7) a pressure (p _vV ) _upstream of the compressor (7) from the atmospheric pressure (p _A ), a measured

Fresh air mass flow (dm _HF ivι) and the atmospheric temperature (T _A ) as well as a pressure (p _nV ) after the compressor (7) from a boost pressure (pι _ad ) with which the fresh air from the compressor (7) is fed to the mixing point (10) is, the measured fresh air mass flow (dm _HFM ) and a charging temperature (Tι _ad ) with which the fresh air is supplied from the compressor (7) to the mixing point (10).

29. The method according to claim 28, characterized in that the fresh air _mass flow (dm) flowing through the compressor (7) from the measured fresh air _{mass flow} (dm _HFM ) and the pressure (p _vV ) _upstream of the compressor

(7) is determined with the aid of a normalization based on the atmospheric temperature (T _A ) and a reference temperature of the compressor (7).

30. The method according to any one of claims 20-29, characterized in that with the aid of the model (18) an exhaust gas temperature (T _nT ) after the turbine (2) as a function of an exhaust gas temperature (T _vT ) before the turbine (2), a temperature change over the turbine (2) and a turbine efficiency is determined.

31. The method according to claim 30, characterized in that the temperature change over the turbine (2) is determined as a function of a pressure ratio over the turbine (2).

32. The method according to claim 30 or 31, characterized in that the efficiency of the turbine is determined as a function of a blade travel (s) of the turbine (2).

33. The method according to any one of claims 30-32, characterized in that the exhaust gas temperature T _n τ after the turbine (2) in dependence on the exhaust gas temperature T _v τ before the turbine (2), the temperature change ΔT _T over the turbine (2 ) and the efficiency η _{τ of} the turbine (2) is determined as follows: T _n τ = T _vT ^■ (1 - ΔT _T ^• η _τ ).

34. The method according to any one of claims 20-33, characterized in that an exhaust gas back pressure (p _nT ) behind the turbine (2) using the model (18) from a pressure difference value, which is the difference between the exhaust gas back pressure behind the turbine and the atmospheric pressure (p _A ) denotes, is derived.

35. The method according to claim 34, characterized in that the pressure difference value is determined as a function of an exhaust gas mass flow (dm _τ ) through the turbine (2).

36. The method according to claim 35, characterized in that the pressure difference value as a function of the exhaust gas mass flow (dm _τ ) through the turbine (2) is determined after the exhaust gas mass flow (dm _τ ) through the

Turbine (2) has been corrected multiplicatively with the aid of a factor which corresponds to the root of an exhaust gas temperature (T _n τ) after the turbine (2).

37. The method according to any one of claims 34-36, characterized in that the exhaust gas back pressure (p _n τ) behind the turbine (2) is determined by adding the pressure difference value with the atmospheric pressure (p _A ).

38. The method according to claim 37, characterized in that with the aid of an exhaust gas back pressure sensor, a pressure difference in the exhaust tract behind the turbine (2) is measured, the exhaust gas back pressure (p _nT ) behind the turbine (2) by adding the pressure difference measured by the exhaust gas back pressure sensor the first-mentioned pressure difference value and the atmospheric pressure (p _A ) is determined.

39. The method according to any one of claims 20-38, characterized in that an exhaust gas back pressure (p _vT ) before the turbine (2) from an exhaust gas back pressure (p _nT ) after the turbine (2), an exhaust gas mass flow flowing through the turbine (2) (dm _τ ), a blade travel (s) of the turbine (2) and a speed (n _ATL ) of an exhaust gas turbocharger shaft (14) coupled to the turbine (2).

40. The method according to claim 39, characterized in that the exhaust gas back pressure p _vT before the turbine (2) from the exhaust gas back pressure (p _n τ) after the turbine, the exhaust gas mass flow dm _τ through the turbine (2), the blade travel s of the turbine ( 2) and the speed n _{ATL of} the exhaust gas turbocharger shaft (14) is determined as follows:

p _vT = Z ^■ p _nT with

Z = b ₀ + bi • dm-r + b ₂ • (s - 0.5) + b ₃ • s ² + b ₄ ^• (n _ATL - 0.5) ²

+ b ₅ ^■ (dm _τ + 0.5) • (s + 0.5)

+ b ₆ - (dm-r -0.5) -s ²

+ b ₇ (s-1) - (n _ATL -0.5) ²

+ b ₈ - (s-1) - (s-0.5) ² + b ₉ - (dm _τ -1) ²

+ b ₁₀ • (dm _τ - 1) ^• (dm _τ + 0.5) ^{2 •} dm _τ

+ b ₁₁ - [(dm _τ -1) - (s-0.5) ³ -0.5] -b ₁₂

+ b ₁₃ ,

where b ₀ -b denote ₁₃ coefficients.

41. The method according to any one of claims 20-40, characterized in that with the help of the model (18) an exhaust gas back pressure (P _EGR ) of the exhaust gas returned via the exhaust gas recirculation as a function of an exhaust gas recirculation _mass flow (dm _EGR ) flowing through the exhaust gas recirculation, an exhaust gas temperature (T _vT ) _upstream of the turbine (2) and an exhaust gas back pressure (p _vT ) of the exhaust gas emitted by the internal combustion engine (1) in front of the turbine (2) is determined.

42. The method according to claim 41, characterized in that the exhaust gas back pressure P _EGR in the exhaust gas recirculation line as a function of the exhaust gas back pressure p _{vT in} front of the turbine (2), the exhaust gas recirculation _{mass flow} dm _EGR in the exhaust gas recirculation line and the exhaust gas temperature T _{vT in} front of the turbine (2) is determined as follows:

where PF denotes an exhaust gas back pressure constant.

43. The method according to claim 42, characterized in that the exhaust gas back pressure constant (PF) is determined as a function of an effective cross-sectional area (A _ef ) of the exhaust gas recirculation line.

44. The method according to claim 43, characterized in that the exhaust gas back pressure constant PF is derived as follows from the gas constant R and the effective cross-sectional area A _{eff of} the exhaust gas recirculation line:

PF = ^R

A εff

45. The method according to any one of the preceding claims, characterized in that that the storage behavior of an intake tract of the internal combustion engine (1) between a compressor (7) assigned to the internal combustion engine (1) and the mixing point (10) is simulated using a physically based model (19).

46. The method according to claim 45, characterized in that the storage of the intake tract is modeled by modeling a storage volume (V _L ) for the fresh air drawn in by the intake tract with a subsequent throttle point with a specific effective cross-sectional area (A _dr ).

47. The method according to claim 46, characterized in that a fresh air mass flow (dm _L ) flowing via the intake tract to the mixing point (10) as a function of a temperature (Tι _ad ) and a pressure (pι _admod ) of the fresh air and the effective cross-sectional area ( A _r ) the throttle point is determined.

48. The method according to claim 47, characterized in that the fresh air mass flow dm as a function of the temperature Tι _ad the fresh air, the pressure Pι _{ad od} the fresh air, the effective cross-sectional area A _dr

Throttle point and a flow characteristic DF is determined as follows:

2 dm _L = A _dr ^■ p _{lad mod} • • DF.

^ ^R - ¹ lad

49. The method according to claim 47 or 48, characterized in that the pressure (pia _dm o _d ) of the fresh air as a function of a fresh air mass (m ^ _d ) between the compressor (7) and the mixing point (10) and the temperature ( T | ₃ ) of the fresh air is determined.

50. The method according to claim 49, characterized in that the pressure the storage volume V and the temperature Tι _{a of} the fresh air is determined as follows:

Piadmod - ~ ^m iad ^" _7.?7_ ^"'■ T * lad * L where R denotes a gas constant.

51. The method according to claim 49 or 50, characterized in that the fresh air mass (m ^) by temporally integrating a fresh air mass flow difference between one flowing into the compressor (7)

Fresh air mass flow (dm _H FM) and the fresh air mass flow (dm) flowing from the compressor (7) to the mixing point (10) is determined.

52. The method according to any one of the preceding claims, characterized in that with the help of a model (20) the behavior of a connection leading from the mixing point (10) to the internal combustion engine (1), which the combustion chamber of the internal combustion engine (1) the gas mixture of the mixing point (10) feeds, is reproduced.

53. The method according to claim 52, characterized in that a fresh air mass (m _L ) and an exhaust gas mass (m _EGR ) in the connection by integrating over time a difference between a fresh air mass flow supplied to the connection (dm _L ) and one of the connection to the internal combustion engine

(1) supplied fresh air mass flow (dm _m o _t ) or by integrating in time a difference between an exhaust gas mass flow (dm _A GR) supplied to the connection via the exhaust gas recirculation and an exhaust gas mass flow (dm _A G _Rm o _t ) is determined.

54. The method according to claim 53, characterized in that the fresh air mass m in the connection and the exhaust gas mass ΓTI _A GR in the connection as a function of the fresh air mass flow dm supplied to the connection, the exhaust gas mass flow dnri _EGR supplied to the connection, and the connection The fresh air mass flow dm mot supplied to the internal combustion engine (1) and the exhaust gas mass flow dm _AGRm o _t supplied to the internal combustion engine (1) by the connection can be determined as follows: ι + τ _n m _L = j (dm _L - dm _Lmot ) dτ for dm _EGR ≥ 0

z = - dm _Lmo , + dm _EGR ) for dm _EGR <0

and

t + T, m EGR = / (dm AGR - ^dm AGRπ, o) ^{dτ for dm} _A GR ^≥ ° tt tt ++ TT ₀ , m AGR = J ~ ^d AGRmot ^{dτ for d} AGR < ⁰

where t denotes an integration time and T ₀ denotes an integration interval.

55. The method according to claim 53 or 54, characterized in that the total gas mass in the connection (m _sr ) by adding the in the

Connection fresh air mass (m _L ) and the exhaust gas mass in the connection (ΓΠA _G R) is determined.

56. The method according to claim 55, characterized in that with the aid of the model (20) the pressure prevailing in the connection from the total gas mass (m _sr ) in the connection, a temperature (T _sr ) in the connection and a volume (V _sr ) the connection is determined.

57. Method according to claim 55 or 56, characterized in that an exhaust gas recirculation rate (Γ _AG R) is determined by relating the exhaust gas mass (IΓ) _A GR) in connection to the total gas mass (m _sr ) in the connection.

58. Method according to one of claims 52-57, characterized in that a temperature (T _sr ) of the gas mixture supplied via the connection in the combustion chamber of the internal combustion engine (1) is determined with the aid of the model (20, 21).

59. The method according to claim 58 and one of claims 53-57, characterized in that with the aid of the model (20, 21) the temperature (T _sr ) of the gas mixture supplied via the connection to the combustion chamber of the internal combustion engine (1) as a function of the Exhaust mass (IΎI _A GR) in the connection, a temperature O _AG R) of the over

Exhaust gas recirculation exhaust gas recirculation mass flow (dm _EGR ), the fresh air mass (m _L ) in the connection and a temperature (Tι _ad ) of the fresh air mass flow supplied to the connection (dm _L ) is determined.

60. The method according to claim 59, characterized in that the temperature T _{sr of} the gas mixture in the connection as a function of the exhaust gas recirculation mass I AGI _AG R in the connection, the temperature T _EGR of the exhaust gas recirculation mass flow, the fresh air mass m in the connection and the temperature T | _{ad of} the supplied fresh air mass flow is determined as follows:

_{τ =} T _EGR - m _{EGR +} T _lad . m _L ^ _Q m _EGR + m _L and

T _sr = T _lad for m _EGR + m _L = 0.

61. The method according to any one of claims 58-60, characterized in that the temperature (T _sr ) of the gas mixture in the connection is corrected by an amount which is based on a difference between a wall temperature (T _w ) of the connection and the temperature (T _sr ) of the gas mixture in the connection and a factor multiplied by it, the factor in turn dependent on the speed (n) of the internal combustion engine (1) and on the connection the combustion chamber of the

Fresh air mass flow (dm _m o _t ) supplied to the internal combustion engine.

62. The method according to claim 61, characterized in that the wall temperature (T _w ) of the connection is derived from a cooling water temperature (T _κ ) of the internal combustion engine (1) and a wall heat factor (WF) of the connection.

63. The method according to any one of the preceding claims, characterized in that the method is carried out automatically by a control device (4), which is part of an engine management system of the internal combustion engine (1).

64.Control system for an internal combustion engine with exhaust gas recirculation, wherein fresh air is mixed with an exhaust gas recirculated exhaust gas from the internal combustion engine (1) in a mixing point (10) and the resulting gas mixture is fed to a combustion chamber of the internal combustion engine (1), characterized in that Control system is configured such that it automatically controls the composition of the gas mixture in the combustion chamber of the internal combustion engine (1) for calculating the internal combustion engine (1) by calculating corresponding state variables of the internal combustion engine (1) by using corresponding physically based models (16-21), which are described in Reproduce the behavior of the internal combustion engine (1) based on the state variable to be calculated.

65. Control system according to claim 64, characterized in that the control system is designed to carry out the method according to one of claims 1- 63.