DE10158249B4 - Method for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation and appropriately designed control system for an internal combustion engine - Google Patents

Abstract

Method for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation,
wherein fresh air is mixed with a recirculated via the exhaust gas recirculation exhaust gas of the internal combustion engine (1) at a mixing point (10) and the resulting gas mixture is supplied to the combustion chamber of the internal combustion engine (1),
wherein an exhaust gas recirculation valve (13) arranged in the exhaust gas recirculation is simulated with the aid of a physically based model (17) for a throttle point and the exhaust gas recirculation mass flow (dm _AGR ) flowing via the exhaust gas recirculation to the mixing point (10) is determined with the aid of this model (17), and
wherein the exhaust gas recirculation mass flow (dm _AGR ) is dependent on an exhaust back pressure (p _AGR ), a recirculated exhaust gas temperature (T _AGR ), a flow characteristic (DF), an effective sectional area (A _AGR ) of the exhaust gas recirculation valve (13), and a gas constant (R ) is determined
characterized,
in that the effective cross-sectional area (A _AGR ) of the exhaust gas recirculation valve (13) depends on a comparison between a measured charge pressure (p _lad ), with which the fresh air is supplied to the mixing point (10) and with the aid of another model ...

Description

The The present invention relates to a method for determining the Exhaust gas recirculation mass flow an internal combustion engine with exhaust gas recirculation and a corresponding designed control system for an internal combustion engine, for example a diesel engine.

For the emission-optimal Control, for example, a supercharged diesel engine with exhaust gas recirculation is the exact knowledge of a possible huge Variety of state variables or Operating parameters of importance.

From the pamphlets DE 199 63 358 A1 and DE 198 44 637 C1 In each case, methods for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation are known, wherein an exhaust gas recirculation valve arranged in the respective exhaust gas recirculation is simulated with the aid of a physically based model for a throttle point in order to determine the exhaust gas recirculation mass flow flowing via the exhaust gas recirculation on the basis of this model. In particular, the exhaust gas recirculation mass flow is determined as a function of the exhaust back pressure, the exhaust gas temperature, the effective cross-sectional area of the exhaust gas recirculation valve, a flow characteristic and a gas constant.

In usual However, engine management systems is the number of recorded or known state variables relative low, or for the detection of the state variables are each separate sensors required. This applies, for example also the over the exhaust gas recirculation line an exhaust gas recirculation mass flow flowing in an internal combustion engine with exhaust gas recirculation. About the Exhaust gas recirculation line an internal combustion engine, exhaust gas from the exhaust system in the intake system returned.

Of the The present invention is therefore based on the object, a method for determining the exhaust gas recirculation mass flow an internal combustion engine with exhaust gas recirculation and a corresponding designed control system for to propose an internal combustion engine, which with as possible simple means an exact determination of the exhaust gas recirculation mass flow without using a separate sensor is possible.

These Task is achieved by a method with the features of claim 1 and a control system solved with the features of claim 6. Define the subclaims respectively preferred and advantageous embodiments of the present invention Invention.

According to the invention is for Determination of exhaust gas recirculation mass flow a physically based model applied, which in the Exhaust gas recirculation of the Combustion engine arranged exhaust gas recirculation valve as a throttle point replicates. As a result, the exhaust gas recirculation mass flow in dependence from the exhaust back pressure and the temperature of the recirculated exhaust gas be determined, where also a flow characteristic, a effective cross-sectional area the exhaust gas recirculation valve and the gas constant in the physically based model for determination the exhaust gas recirculation mass flow considered become.

Of the Course of the temperature of the over the exhaust gas recirculation recirculated exhaust gas is preferably modeled using the model to make it the respective current temperature of the recirculated exhaust gas before the exhaust gas recirculation valve derive.

The effective cross-sectional area the exhaust gas recirculation valve becomes dependent from a comparison between the respectively measured charge pressure, with where the fresh air is supplied to the mixing point, and one with the help Another model modeled boost pressure by using a adapted to the corresponding correction factor.

The previously mentioned Flow characteristic can for example, from the pressure ratio over the Exhaust gas recirculation valve be derived.

Preferably come several physically based (or empirically determined) Models for use, which are partially closely related stand, so preferably for the calculation of a certain State size in one Model is accessed on the results of another model. It should be noted that the dependencies described herein by formula of the individual sizes in the Usually only the proportional relationships should clarify, so that dependent of the particular application or implementation, if applicable (not specified) normalization or correction factors for further Conversion of the corresponding quantities are.

All in all can with the help of the present invention, the exhaust gas recirculation mass flow an internal combustion engine, such as a diesel engine, exactly and determined by simple means by evaluating already known state variables become. The use of additional Sensors is for this not mandatory. By thus easily possible determination of the exhaust gas recirculation mass flow are new regulatory and diagnostic procedures within each Engine management system possible, which, for example, an emission-optimal control of the internal combustion engine allowed.

The The present invention will be explained in more detail below with reference to FIGS attached Drawing explained with reference to a preferred embodiment.

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,

2 shows a representation for explaining a motor filling model,

3 shows a representation for explaining an exhaust gas recirculation mass flow model,

4 shows a representation for explaining a turbine model,

5 shows a schematic representation of the intake tract of in 1 illustrated internal combustion engine,

6 shows a representation for explaining a fresh air mass flow model,

7 shows a schematic representation of the intake tract of in 1 illustrated internal combustion engine,

8th shows a representation for explaining a Saugrohrmodells,

9 shows a representation for explaining a Saugrohrtemperaturmodells, and

10 shows the course of a flow characteristic as a function of a pressure ratio.

In 1 is an internal combustion engine 1 shown with four combustion chambers or cylinders. The internal combustion engine 1 is coupled to an exhaust gas turbocharger (ATL), which is a turbine 2 and a compressor 7 includes, wherein the turbine and the compressor 7 on a common shaft, the so-called turbocharger shaft 14 , are appropriate. The turbine 2 uses the exhaust gas of the internal combustion engine 1 contained energy to drive the compressor 7 , which has an air filter 6 Fresh air sucks and pre-compressed air into the individual combustion chambers of the internal combustion engine 1 suppressed. The one by the turbine 2 , the compressor 7 and the turbocharger shaft 14 formed exhaust gas turbocharger is fluid only with the internal combustion engine through the air and exhaust gas mass flow 1 coupled.

The of the compressor 7 over the air filter 6 sucked in and pre-compressed air is supplied via a charge air cooler (LLK) 8th , which reduces the exhaust gas temperature and thus the NO _x emission and the fuel consumption, a so-called replacement volume (ERS) 9 fed. The individual combustion chambers of the internal combustion engine 1 is an intake collector (ELS) 10 upstream. That in the combustion chambers of the internal combustion engine 1 generated exhaust gas is from an exhaust collector (ASA) 11 collected and the turbine 2 fed. The turbine 2 is in the exhaust gas flow direction the exhaust system (APU) 12 downstream of the motor vehicle, which the pollutant components during operation of the internal combustion engine 1 degrades resulting exhaust gases and derived the remaining exhaust gases as quietly as possible. Part of the in the combustion chambers of the internal combustion engine 1 generated exhaust gas is from the exhaust manifold 11 via an exhaust gas recirculation (EGR) to the intake manifold 10 returned and mixed there with the sucked fresh air. With the reference number 13 are respectively arranged in corresponding air or gas paths arranged valves. With the reference number 15 is an actuator for the guide vane adjustment of the turbine 2 designated.

Furthermore, in 1 a control unit 4 which is a component of a corresponding engine management system of the motor vehicle. From the controller 4 For example, various sizes or parameters of the illustrated engine system are monitored and converted to various intermediate and output quantities using appropriate stored physically-based models, that of the controller 4 Monitored variables or parameters via an interface 3 the control unit 4 supplied become. The individual of the control unit 4 evaluated quantities are explained in more detail below with reference to the individual physically based models. In particular, by the controller 4 in this way the mass and composition of the combustion chambers of the internal combustion engine 1 located gas mixture, ie, the fresh air and exhaust gas contained therein, determined and implemented to achieve an emission-optimal control in corresponding control signals for the engine system, which - as in 1 is indicated - via the interface 3 can be applied to various components of the engine system.

For a stable calculation of the overall model formed by the individual physically based models by the controller 4 For some parts of the overall model, a certain minimum effective computation time, for example of the order of 2 ms, is required. Since this is not feasible with conventional control unit concepts, an existing time-synchronous grid is preferably used as a basis and the overall model is calculated several times in this grid (over-sampling). For example, in order to achieve an effective computing time of 2 ms for an existing 20 ms raster, the overall model must be calculated ten times within the specified raster. Since the overall model, which is composed of the individual previously mentioned physically based submodels, serves for charge detection of internal combustion engines, ie for the exact determination of 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 charge model.

One of these from the controller 4 executed physically based submodels used to replicate 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 in 1 illustrated mixing point 10 from which the over the compressor 7 sucked fresh air is mixed with the exhaust gas recirculated via the exhaust gas recirculation line, and the intake valves of the internal combustion engine 1 designated. This model can therefore also be referred to as engine fill model.

With the aid of this engine filling model, the aspirated gas mass in the combustion chamber in dependence on the pressure p _sr and the temperature T _{sr of} the intake gas, which define the density of the intake gas, taking into account the gas constant R, before the engine intake valves, ie in the intake manifold, determined a linear approach is chosen depending on the density of the intake gas:

In this case, m _ges denotes the aspirated gas mass in the combustion chamber, ie the mass of the intake fresh air / exhaust gas mixture, 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 ₁ and d ₂ are therefore a function of the engine speed n ₀ . This dependence on engine speed can be represented by quadratic polynomials: d 1 = a 1 + a 2 · n 0 + a 3 · n 0 2 d 2 = a 4 + a 5 · n 0 + a 6 · n 0 2 , (2)

Here, a ₁ -a denote ₆ coefficients of these quadratic polynomials. Optionally, the above-described dependence on the engine speed can also be realized by speed-dependent characteristics, wherein in the control unit 4 between these alternatives, for example, depending on the instantaneous value of a corresponding variable can be switched.

The filling of the combustion chamber of the internal combustion engine 1 is composed of proportions of fresh air and recirculated exhaust gas. From the previously determined gas mass m _ges in the combustion chamber and the current engine speed n of the internal combustion engine 1 the intake gas mass flow dm _ges can be calculated. The fresh air _{mass flow} dm _Lmot in the internal combustion engine 1 results as a function of the intake gas mass flow dm _ges and the current exhaust gas recirculation rate r _AGR as follows: dm Lmot = (1 - r AGR )·dm ges (3)

The aspirated gas mass flow dm _ges and the intake air mass flow dm _Lmot are preferably calculated in the unit kg / s. Of course, a conversion in kg / h is also possible.

As a further intermediate variable, which serves as the basis for the calculation of the air ratio in can serve the combustion chamber, the fresh air _mass m _Lmot in the combustion chamber of the internal combustion engine 1 be determined as follows: m Lmot = (1 - r AGR ) · M ges (4)

Analogously, in the internal combustion engine 1 sucked exhaust gas recirculation _{mass flow} dm _{AGRmot be calculated} from the intake gas mass _flow dm _ges and the current exhaust gas recirculation rate r _EGR as follows: dm AGRmot = r AGR ·dm ges (5)

Analogous to the fresh air _mass in the combustion chamber of the internal combustion engine, the exhaust gas mass m _AGRmot in the combustion chamber of the internal combustion engine can also be determined from the already known aspirated gas mass m _ges : m AGRmot = r AGR · m ges (6)

As a further output variable from the engine filling model, the air ratio R _L in the combustion chamber of the internal combustion engine 1 determined from the now known fresh air _mass m _Lmot and the injected fuel _mass m _kr :

With the aid of the aforementioned correction factor KORR, the engine fill model can be adapted to the actual behavior of the internal combustion engine, for which purpose a comparison is made between a modeled boost pressure p _ladmod and a measured actual boost pressure p _lad . This comparison can be carried out in a further submodel, which can be referred to as a correction model. By the difference of these two variables, an integrator can be fed, whose output value the proportionate correction factor KORR for the total filling of the internal combustion engine 1 results. Defined conditions, such as stationary engine operation without exhaust gas recirculation, must preferably be present for this adaptation process. The control unit 4 For this purpose, it may contain a separate function block which controls the adaptation release, ie the integrator, and for this purpose evaluates certain input variables which define, for example, the permitted adaptation range with regard to injection quantity and rotational speed or monitor the temporal change of these variables. In addition, this function block can be fed additional parameters, with the help of the maximum dynamic range of the fresh air mass flow and the boost pressure can be adjusted, preferably an on and Auschaltverhalten can be implemented with hysteresis. The output KORR of this function block of the controller 4 corrects according to formula (1) the slope of the filling line and thus fits the engine filling model to the actual behavior of the internal combustion engine 1 at.

The previously described engine fill model 16 which is in the control unit 4 is implemented schematically in terms of its input and output variables in 2 shown.

About the in 1 indicated exhaust gas recirculation line is - as already mentioned - returned exhaust gas from the exhaust system in the intake. There is therefore provided a further physically based model which determines the exhaust gas recirculation mass flow through the exhaust gas recirculation line as well as the temperature of the recirculated exhaust gases upstream of the exhaust gas recirculation mixing point 10 calculated, so that this model is also referred to below as exhaust gas recirculation mass flow model.

The determination of the exhaust gas recirculation mass flow dm _EGR takes place with the aid of a model approach for a throttle point of the exhaust gas recirculation valve present in the exhaust gas recirculation line 13 (see 1 ) in response to 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 back pressure p _AGR and the temperature T _AGR before the exhaust gas recirculation valve 13 :

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

The root functions contained in formulas (8) and (9) may preferably be approximated by a quadratic polynomial valid, for example, in the temperature range of 200-1200 K of interest here. To account for 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 the controller becomes 4 preferably delayed by a PT1 member.

As already mentioned, with the aid of this model not only the exhaust gas recirculation mass _flow dm _AGR but also the temperature T _EGR of the recirculated exhaust gases before the mixing point with the fresh air are calculated. The temperature T _AGR 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 in front of the exhaust gas recirculation valve, a distinction must also be made between forward flow and reverse flow. Where: T AGR = T AG - RF · (T AG - T K ) for dm AGR ≥ 0 (10) T AGR = T sr for dm AGR <0 (11)

In the case of the forward flow (dm _EGR ≥ 0), hot exhaust gases are passed through the exhaust gas recirculation line, while in the case of the reverse 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) characterized in that of the exhaust gas temperature TAG before the turbine 2 RF · (T _AG - T _K ) is subtracted, wherein RF denotes a pipe factor of the exhaust gas recirculation line, with the aid of which the cooling can be adapted to the type of exhaust gas recirculation line (eg distinction between cooled and uncooled exhaust gas recirculation), while T _K der Cooling water temperature of the internal combustion engine 1 corresponds and thus is a measure of the cooling of the exhaust gas temperature TAG. The exhaust temperature TAG in front of the turbine 2 is generated by a further physical-based model explained in more detail below.

The according to the formula (8) and (9) needed Flow characteristic DF is a function of the pressure ratio over the from this exhaust gas recirculation mass flow model simulated throttle point, d. H. above the exhaust gas recirculation valve. As the flow parameter DF also is used in other models of the overall system, it is preferable also realized as a separate method by the others Models can be called. The corresponding method evaluates the pressure in front of the corresponding throttle point and the pressure behind the corresponding throttle point and delivers depending on one specific value for the flow characteristic DF back. there must be between a so-called supercritical Flow case where the pressure ratio over the Throttling point is less than a predetermined critical pressure ratio, and a subcritical case where the pressure ratio is greater than the critical pressure ratio is to be distinguished.

The course of the flow _parameter DF as a function of the pressure ratio between the pressure p _{vdr upstream} of the throttle point and the pressure p _{ndr downstream of} the throttle point is in 10 shown. Out 10 It can be seen that in the supercritical flow case, which according to 10 is separated from the subcritical flow case by a dashed line, the flow characteristic DF can be set to a certain maximum value. In the subcritical case, on the other hand, the flow parameter DF is calculated according to a substitute function which corresponds to the in 10 shown in dependence on the pressure ratio continuously decreasing curve for the subcritical case corresponds. In particular, a distinction is made between the case of the forward flow and the case of the reverse flow. The forward flow may be distinguished from the reverse flow by, for example, setting a corresponding bit in a corresponding variable.

The determination of the effective cross-sectional area _{AGR of} the exhaust gas recirculation valve is done by means of a corrected by a correction factor AKORR characteristic field, wherein as an input of this characteristic field in response to the instantaneous value of a corresponding bit optionally the measured valve or the drive duty cycle of this valve by the control unit 4 is used. wel If these input variables are used to determine the effective cross-sectional area of the exhaust gas recirculation valve, this depends on the type of actuator used. An electrical exhaust gas recirculation controller becomes the drive duty cycle of the controller 4 is used as the input variable for the corresponding characteristic map, while in the case of a controller 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 in an adjustment, the thus calculated effective cross-sectional area of the exhaust gas recirculation valve can be delayed by a PT1 member.

By the aforementioned correction factor AKORR, similar to the previously described engine fill model, the calculated valve cross-sectional area of the exhaust gas recirculation valve may be dependent on a comparison between the measured and the modeled boost pressure in the stationary operating phases of the internal combustion engine 1 Getting corrected. In this regard, an integrator can be used for this purpose, which evaluates the difference between the measured and modeled boost pressure and supplies the correction value AKORR for the calculated cross-sectional area of the exhaust gas recirculation valve as output value.

In 3 is the exhaust gas recirculation mass flow model described above 17 shown schematically with its input and output variables.

With the help of another model, which is also referred to below as a turbine model, the behavior of the exhaust tract before and after the in 1 shown turbine 2 be reproduced. The most important output variable of the turbine model is the exhaust backpressure in front of the turbine 2 determined. In addition, further initial and intermediate sizes are calculated, which will be discussed in more detail below.

Within the turbine model, the blade path s of the turbine 2 an important parameter for determining the exhaust backpressure in front of the turbine 2 , The blade travel s can be measured either directly in combination with a corresponding analog / digital conversion or via the drive duty cycle of the in 1 shown actuator 15 be determined. The determination of the instantaneous vane travel s over this drive duty cycle may be accomplished by accessing a corresponding characteristic curve representing each value of the drive duty cycle corresponding to a value of the vane travel s of the turbine 2 assigns. The dynamics of the blade movement of the turbine 2 is preferably taken into account by a PT1 element in order to simulate the time behavior of the blade travel s as well as possible.

The exhaust temperature _{TAG in} front of the turbine 2 becomes n ₀ (normalized engine speed) or n (non-normalized engine speed) depending on the injection amount to me and the engine speed n over a difference temperature approach between the exhaust gas temperature upstream of the turbine 2 and the intake manifold temperature, ie the temperature in the intake tract. In this case, the differential temperature, ie 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 difference temperature value .DELTA.T1 _ASA obtained in this manner can be determined as a function of the start of delivery, ie the beginning of the 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 by the combustion before the turbine 2 , to obtain: .DELTA.T ASA = ΔT1 ASA · .DELTA.T2 ASA (12)

Alternatively, an additive correction can also take place: .DELTA.T ASA = ΔT1 ASA + ΔT2 ASA (13)

The differential temperature correction value ΔT2 _ASA is thereby determined with the aid of a further characteristic as a function of the start of delivery FB. The switching between the two aforementioned alternatives (compare formulas (12) and (13)) can be done in dependence on the position of a corresponding switch or a corresponding bit.

The expelled exhaust gas mass flow dm _{ASA of} the internal combustion engine 1 gets out of the engine 1 or the corresponding combustion chamber sucked gas mass flow dm _ges and the injected fuel mass flow dm _kr or one of the injected fuel mass m _kr and the engine speed n dependent proportion calculated: dm ASA = dm ges + dm kr = dm ges + f (n, m kr ) (14)

The gas mass flow dm _T through the turbine 2 can be from that of the internal combustion engine 1 emitted exhaust mass flow dm _ASA and the exhaust gas recirculation mass flow dm _{EGR are} determined: dm T = dm ASA - dm AGR (15)

Furthermore, one on the compressor 7 related exhaust gas turbocharger or compressor speed n _V using a map as a function of the fresh air mass flow dm _V through the compressor 7 and the pressure ratio across the compressor 7 be determined. For calculating the pressure ratio over the compressor 7 will be the pressure behind the compressor 7 and the pressure in front of the compressor 7 determined, and then from the pressure ratio over the compressor 7 to calculate. The pressure p _vV before the compressor 7 or _pnV after the compressor 7 can be calculated as follows from the atmospheric pressure p _A , that measured by the hot-film air mass sensor and in the in 1 shown model incoming fresh air mass flow dm _HFM , the atmospheric temperature T _A , the boost pressure p _lad and the charging temperature T _{lad are} determined:

In this case, a loss factor VFAK1 or VFAK2 is used according to equation (16), in each case the pressure loss before or after the compressor 7 to be considered, which are determined in each case by quotient formation from the gas constant R and the square of a corresponding equivalent area A ² _vV or A ² _nV .

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

Here K denotes a constant and T _0V 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 _{ATL is} calculated from the on the compressor 7 related exhaust gas turbocharger speed n _V as a function of the ambient or atmospheric temperature T _A and the reference temperature T _{0V of} the compressor 7 as follows:

The root function contained in the formulas (17) and (18) can be calculated by a quadratic polynomial as a function of T _A / T _0V for reasons of computing time.

As a further output variable, the temperature T _nT in the exhaust gas tract behind the turbine is determined with the aid of the turbine model 2 calculated. This is done depending on the temperature T _{vT of} the turbine 2 by replicating the temperature drop over the turbine 2 In addition, wherein the turbine efficiency η _T is taken into account as follows: T nT = T vT · (1 - ΔT T · η T ) (19)

The temperature change ΔT _T over the turbine 2 is determined by means of a corresponding characteristic 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 _{T of} the turbine 2 with the help of a corresponding characteristic as a function of the blade path s of the turbine 2 is applied. The temperature T _vT before 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 before the turbine 2 the modeled exhaust back pressure p _{AG in} front of the turbine 2 ,

As a further variable, the exhaust back pressure p _{nT is} calculated behind the turbine, for which purpose a pressure difference between the exhaust tract behind the turbine 2 and the atmospheric pressure p _{A is} determined. This can likewise take place via a corresponding characteristic curve, the gas mass flow dm _{T being used} as an input variable by the turbine for this characteristic curve 2 is used, which multiplicatively as follows by the root of the exhaust gas temperature T _nT after the turbine 2 is corrected:

Depending on the corrected gas mass flow dm * _T through the turbine 2 can the pressure difference Δp _nT between the exhaust tract behind the turbine 2 and the atmospheric pressure p _A can be determined by means of a quadratic equation as a function of Δp _nT , the coefficients of this quadratic equation being applicable. The exhaust back pressure p _nT after the turbine 2 (in bar) results in the event that no exhaust back pressure sensor in the exhaust tract after the turbine 2 is present as follows from the addition of the atmospheric pressure p _A and the calculated pressure difference Δp _nT : p nT = (Δp nT + p A ) / 10 5 (21)

On the other hand, an exhaust backpressure sensor in the exhaust tract or a differential pressure sensor behind the turbine 2 provided, the measured pressure difference Δp _AG measured by this exhaust backpressure sensor in addition to the modulated exhaust back pressure behind the turbine 2 added: p nT = (Δp nT + p A + Δp AG ) / 10 5 (22)

The exhaust back pressure p _{vT in} front of the turbine 2 can from the exhaust back pressure p _nT after the turbine 2 is calculated with the help of a polynomial having 13 coefficients as a function of the input variables turbine mass flow dm, blade travel s and exhaust gas turbocharger speed n _ATL , wherein the three last-mentioned variables are preferably used standardized with the aid of corresponding applicable parameters. An exemplary and preferred calculation _rule for determining the exhaust gas back pressure p _vT before the turbine 2 is given below, but in principle any combinations of the input variables are possible: p vT = Z · p nT Z = b 0 + b 1 ·dm T + b 2 · (S - 0.5) + b 3 · s 2 + b 4 · (N ATL - 0.5) 2 + b 5 ·(dm T + 0.5) · (s + 0.5) + b 6 ·(dm T - 0.5) · s 2 + b 7 · (S - 1) · (n ATL - 0.5) 2 + b 8th · (S - 1) · (s - 0.5) 2 + b 9 ·(dm T - 1) 2 + b 10 ·(dm T - 1) · (dm T + 0.5) 2 ·dm T + b 11 ·[(dm T - 1) · (s - 0.5) 3 - 0.5] · b 12 + b 13 (23)

The coefficients b ₀ -b ₁₃ are preferably variable.

As a further output variable, the exhaust backpressure p _AGR before the in 1 shown exhaust gas recirculation valve 13 calculated. It results as follows as a function of the exhaust _{backpressure upstream of} the turbine p _vT , the exhaust gas recirculation _{mass flow} dm _AGR , the exhaust gas temperature _{upstream of} the turbine T _vT and a constant PF:

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

In this approach, a pressure drop in the exhaust gas recirculation line before and after 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 controller 4 From this the parameter PF can be calculated as follows, where R denotes the gas constant:

The turbine model explained in detail above 18 is schematically in terms of its input and output variables 4 shown.

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

As output variables of this fresh air mass flow model, in particular the fresh air mass flow dm _L through the aforementioned throttle point into 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 as well as the modeled boost pressure p _ladmod determined. In addition, the difference Δdm _L between the measured fresh air mass flow dm _{HFM of} the hot-film air mass sensor and that in the internal combustion engine 1 flowing fresh air mass flow dm _L determined.

The modeled boost pressure p _ladmod can be calculated as follows from the fresh air _mass m _lad in the volume between the compressor 7 and the intake manifold or engine inlet and a measured charging temperature T _{lad of} the fresh air are calculated:

The charge air temperature T _lad is preferably used PT1 filtered.

The fresh air _{mass flow} dm _L into the intake manifold may vary as a function of the PT1 filtered, measured charge air _temperature T _lad , the modeled boost pressure p _ladmod , the gas constant R, the modeled intake manifold pressure p _sr , ie the pressure of the intake gas upstream of the intake valves of the internal combustion engine 1 , and the effective cross-sectional area A _{dr of} the throttle valve before the exhaust gas recirculation fresh air mixing point are determined:

The fresh air mass flow dm _L determined in this way can also be filtered by means of a corresponding PT1 element in order to simulate the inertia of the fresh air mass flow. The time constants used in the PT1 filtering, which simulate the inertia of the fresh air mass flow for a positive or negative change, should be as small as possible (eg <20 ms). The root in formula (27) can again be approximated by a third-order polynomial. As has already been described with reference to the engine fill model explained above, 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 drive pulse duty factor of the control device, which is likewise delayed by a PT1 element 4 In this case, the time constants of the PT1 element should be chosen so that they largely correspond to the time constants for the opening and closing of the throttle valve.

From the mass flow balance of the volume between the compressor 7 and the intake manifold or engine intake of the internal combustion engine 1 the fresh air mass m _lad results from the integration of the difference mass flow Δdm _L between the inflowing, measured fresh air mass flow dm _HFM and the outflowing, modeled fresh air mass flow dm _L into the intake manifold:

In this case, T ₀ denotes the respectively selected temporal integration interval. The fresh air mass m _lad 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 is preferably variably adjustable by means of a corresponding parameter.

The fresh air mass m _lad determined in this way forms, as has been described above, the basis for determining the modeled _charge pressure p _ladmod according to formula (26) by using the ideal gas law.

The fresh air mass flow model explained in detail above 19 is schematically in terms of its input and output variables 6 shown.

With the help of another model, the behavior of the intake manifold, ie the connection between the exhaust gas recirculation / fresh air mixing point and the engine intake valves, is modeled, wherein the intake manifold is also modeled by a container with a volume V _sr . This container is hereinafter referred to as suction tube, so that the corresponding model can be referred to as Saugrohrmodell. A schematic representation of the intake manifold suction pipe is based on the in 5 shown schematic representation of the intake in 7 shown.

In the intake manifold, the incoming exhaust gas recirculation mass flow dm _AGR and the fresh air mass flow dm _L mix to a fresh air / exhaust gas mixture from which the internal combustion engine 1 his filling relates. 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 _L results from the integration of the difference between the incoming and outgoing fresh air mass flow into the intake manifold:

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 intake manifold (dm _EGR ≥ 0) or fresh air from the intake manifold into the exhaust tract (dm _EGR < 0) flows. Within the intake manifold model, the integrator outputs and thus the calculated fresh air mass m _{L are} preferably limited to a minimum value and to a maximum value.

The calculation of the recirculated exhaust mass m _EGR in the intake manifold is analogous to the calculation of the fresh air mass. By integrating the difference between the incoming exhaust gas recirculation mass flow dm _AGR and the exhaust gas recirculation mass flow dm _AGRmot flowing into the engine, the mass m _{EGR of} the recirculated exhaust gases in the intake manifold is obtained:

In the case of a negative exhaust gas recirculation mass flow, for the sake of simplification, it is assumed that only the fresh air flows into the exhaust tract via the exhaust gas recirculation line, ie it is assumed that: _AGR = 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 validity ranges are preferably over entspre changeable parameters.

The total gas mass m _sr then results from the addition of the fresh air mass m _L and the exhaust gas mass m _EGR in the intake manifold. In an initialization phase of the controller 4 can be calculated in each case for the fresh air mass m _L and the exhaust gas mass m _AGR an initial value in dependence on a predeterminable temperature and a predetermined pressure.

The pressure p _sr in the intake manifold results from the ideal gas law as a function of the volume V _sr :

Finally, as a further output variable of the intake manifold model, the exhaust gas recirculation rate r _AGR is calculated from the mass fractions in the intake manifold as follows:

The intake manifold model explained in detail above 20 is schematically in terms of its input and output variables 8th shown.

In another physically based model, which in the control unit 4 is realized, the intake pipe temperature T _{sr of} the fresh air / exhaust gas mixture is determined in the intake manifold.

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

In the event that the denominator of the formula (33), ie the total mass in the intake manifold, the value 0, the intake pipe temperature T _{sr is set} to the value of the temperature T _{lad of} the incoming fresh air mass flow: T sr = T lad for m AGR + m L = 0 (34)

The wall heat transitions in the intake manifold lead to a 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 . T sr = T sr + ΔT sr (35)

The temperature change ΔT _sr is dependent 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 through the wall heat transfers is also dependent on the current engine operating point. This can be taken into account with the aid of a characteristic map in the calculation of the temperature change .DELTA.T _sr , wherein in this map, a factor as a function of the rotational speed n and into the combustion chamber of the internal combustion engine 1 sucked fresh air _{mass flow} dm _Lmot can be adjusted: .DELTA.T sr = (T w - T sr ) · F (n, dm Lmot ) (36)

It can be seen from the formula (36) that the temperature change ΔT _{sr is} calculated from a difference value of the wall temperature T _{W of} the suction pipe and the temperature T _{sr of} the fresh air / exhaust gas mixture in the intake pipe, this difference value being multiplied by a factor which is multiplied by the _Engine speed n and the intake fresh air _{mass flow} dm _{Lmot of} the internal combustion engine 1 is dependent. The wall temperature T _{W of} the suction pipe can be calculated from the PT1-filtered cooling water temperature T _K and an applicable wall heat factor WF: T W = T K · WF (37)

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

The previously discussed intake manifold temperature model 21 is schematically in terms of its input and output variables 9 shown.

As previously explained has been able to other physically based models or functions or methods be provided, which from the individual previously described in detail Models can be called, depending on specific sizes each handed over Be able to determine parameters. Such a function or method can - as already explained is - to Determination of flow characteristic DF provided be.

In addition, a function can be provided to convert certain quantities and to provide certain constants and parameters that the individual models can access. For the sake of clarity, the overall model is mainly calculated in terms of SI units. For this reason, preprocessing or conversions of the variables used in each case are sometimes required for the individual submodels or submodels, which can be performed by this function block. In this case, separate sections or methods for the processing of speed-synchronous variables and time-synchronous variables can be provided. An example of such a conversion is the conversion of the input signals for position measuring systems. If the exhaust gas recirculation valve and blade position position measuring systems are used, the signal processing in the control unit 4 make necessary, additional characteristics must be provided in each case, which allow the conversion of the voltage signals in the corresponding path. Another example of such a conversion is the consideration of the hysteresis and lots of the blade pitch systems of the turbine 2 , By design, the vane adjustment system has a hysteresis behavior which results from the loosening between the guide pins of the respective control linkage and the vanes on the turbine adjusting ring 2 declining. As a result, dead spots may result in reversing the direction of the control linkage, in which no blade adjustment takes place. In order to take this behavior into account, the blade travel determined by a path system is preferably displaced unidirectionally depending on the direction, wherein the dead travel can be adjusted by application via a corresponding parameter.

A another centrally provided method or another centrally provided function block can for the realization of already previously explained PT1 filtering of various sizes provided be. For this purpose, this method is realized such that They have two time constants, which depend on the input signal direction (ascending or descending), is called. Furthermore this method is preferred in addition to the indication of the respective input signal Also called with a parameter which the respective time grid describes. The return value this method or function block is then the PT1 filtered Input signal.

Claims (7)

Method for determining the exhaust gas recirculation mass flow of an internal combustion engine with exhaust gas recirculation, fresh air having an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine ( 1 ) at a mixing point ( 10 ) and the resulting gas mixture to the combustion chamber of the internal combustion engine ( 1 ), wherein an exhaust gas recirculation valve arranged in the exhaust gas recirculation ( 13 ) using a physically based model ( 17 ) for a throttle point and the exhaust gas recirculation to the mixing point ( 10 ) flowing exhaust gas recirculation mass flow (dm _EGR ) using this model ( 17 ) and wherein the exhaust gas recirculation mass flow (dm _AGR ) in dependence on an exhaust back pressure (p _AGR ), a temperature of the recirculated exhaust gas (T _AGR ), a flow characteristic (DF), an effective cross-sectional area (A _EGR ) of the exhaust gas recirculation valve ( 13 ) and a gas constant (R), characterized in that the effective cross-sectional area (A _AGR ) of the exhaust gas recirculation valve ( 13 ) as a function of a comparison between a measured charge pressure (p _lad ), with which the fresh air of the mixing point ( 10 ), and one with the aid of another model ( 19 ) modeled boost pressure (p _ladmod ) is _adjusted by using a corresponding correction factor (AKORR). Method according to Claim 1, characterized in that the profile of the temperature of the exhaust gas recirculated via the exhaust gas recirculation is determined by means of the model ( 17 ) and from this the temperature (T _EGR ) of the recirculated exhaust gas upstream of the exhaust gas recirculation valve ( 13 ) is derived. A method according to claim 1 or 2, characterized in that the exhaust gas recirculation mass _flow dm _AGR from the effective cross-sectional area A _{AGR of} the exhaust gas recirculation valve ( 13 ), the exhaust back pressure p _AGR , the temperature T _{EGR of} the recirculated exhaust gas, the gas constant R and the flow characteristic DF are determined as follows:

wherein, in the event that the exhaust gas recirculation mass flow from an intake tract of the internal combustion engine ( 1 ) in an exhaust tract of the internal combustion engine ( 1 ) flows as value for the exhaust back pressure of the boost pressure of the fresh air in the intake tract and is used as a value for the temperature, the charging temperature of the fresh air in the intake tract. Method according to one of the preceding claims, characterized in that the flow characteristic (DF) from a pressure ratio over the exhaust gas recirculation valve ( 13 ) is derived. Method according to one of the preceding claims, characterized in that the method is automatically controlled by a control unit ( 4 ), which is part of an engine management system of the internal combustion engine ( 1 ) is executed. Control system for an internal combustion engine with exhaust gas recirculation, wherein fresh air with a recirculated via the exhaust gas recirculation exhaust gas of the internal combustion engine ( 1 ) in a mixing point ( 10 ) mixed and the resulting gas mixture a combustion chamber of the internal combustion engine ( 1 ) is supplied, wherein the control system is designed such that it for controlling the internal combustion engine ( 1 ) automatically via the exhaust gas recirculation to the mixing point ( 10 ) flowing exhaust gas recirculation mass flow (dm _EGR ) using a physically based model ( 17 ), which is arranged in the exhaust gas recirculation exhaust gas recirculation valve ( 13 The control system is configured to calculate the exhaust gas recirculation mass flow (dm _EGR ) as a function of an exhaust back pressure (p _AGR ), a temperature of the recirculated exhaust gas (p _AGR ), a flow characteristic (DS), an effective cross-sectional area (A _EGR ) of the exhaust gas recirculation valve ( 13 ) and a gas constant (R), characterized in that the control system is designed such that it the effective cross-sectional area (A _AGR ) of the exhaust gas recirculation valve ( 13 ) as a function of a comparison between a measured charge pressure (p _lad ), with which the fresh air of the mixing point ( 10 ), and one with the aid of another model ( 19 ) modeled boost pressure (p _ladmod ) by using a corresponding correction factor (AKORR) adapts. Control system according to Claim 6, characterized that the tax system to carry out the Method according to one of the claims 1-5 designed is.