DE102004019315B4 - Method for determining state variables of a gas mixture in an air gap associated with an internal combustion engine and correspondingly configured engine override - Google Patents

Abstract

Method for detecting state variables of a gas mixture in an air gap assigned to an internal combustion engine with exhaust gas recirculation, fresh air mixed with an exhaust gas recirculated via the exhaust gas recirculation of the internal combustion engine (1) at a mixing point (11) of the air gap and the resulting gas mixture fed to the internal combustion engine (1) is characterized in that the mixing point (11) in front of a combustion engine (1) associated compressor (3) is arranged and that state variables of the gas mixture, which comprise an exhaust gas recirculation rate and an oxygen rate, at different points of the air gap with a model (21) for a low-pressure exhaust gas recirculation line as part of the air gap, which in relation to the state variable to be determined the behavior of the air gap or devices contained in the air gap, a compressor (3), a charge air cooler (4), a particulate filter (8) and an Ab Gas return cooler (9) include, simulates, is determined.

Description

The present invention relates to a method for detecting state variables of a gas mixture in the air gap of an internal combustion engine with exhaust gas recirculation as well as a correspondingly designed engine control system for an internal combustion engine, for example a diesel engine.

The DE 199 63 358 A1 describes the control of an internal combustion engine with an air system. Here are determined by models sizes that characterize the air system. The models include a partial model for a compressor, a partial model for the internal combustion engine, a sub-model for a high-pressure fresh-air line and an exhaust-gas recirculation model.

The DE 94 21 145 U1 shows an internal combustion engine in which the recirculated exhaust gas is mixed with fresh air before the compressor.

The DE 101 58 262 A1 discloses the composition of the gas mixture in a combustion chamber of an internal combustion engine. Physically based models are used to determine the composition and the mass of the fresh air / exhaust gas mixture sucked in by the internal combustion engine, each of which simulates the behavior of the internal combustion engine or of the corresponding engine system with respect to certain state variables. The recirculated exhaust gas is mixed after the compressor with fresh air.

Exhaust gas recirculation is used to reduce nitrogen oxides in conventional combustion methods of a direct-injection internal combustion engine, thereby lowering local peak temperatures in a cylinder charge of the internal combustion engine and an oxygen concentration in the cylinder charge. With increasing exhaust gas recirculation rate, however, soot particle formation is favored at the same time and post-oxidation is weakened, as a result of which certain exhaust gas standards can no longer be met. In order to solve a trade-off between the reduction of nitrogen oxides and favoring particle formation, a so-called homogeneous combustion in internal combustion engines is known. In this case, a large amount of well cooled exhaust gas is returned to the engine and initiated an injection very early before the top dead center. As a result, a time for air-fuel mixture formation within the cylinder to ignition is significantly prolonged (ignition delay), since on the one hand, a temperature level during compression is very low and on the other hand inert gases of the exhaust gases cause isolation between fuel and charge. The extended ignition delay, in turn, results in the formation of soot particles being avoided by a more homogeneous mixture formation. In combustion, due to a high proportion of the exhaust gas in the mixture, almost all of the mixture uniformly heats, thereby avoiding local temperature spikes in the mixture. For this reason and because a residual oxygen content in the mixture is low, almost no nitrogen oxides are produced during the combustion. Therefore, an internal combustion engine which is operated with the homogeneous combustion, over an internal combustion engine, which is operated with the conventional combustion, has advantages in terms of the emissions contained in the exhaust gas.

For an emission-optimal control of an internal combustion engine with exhaust gas recirculation, the exact knowledge of different state variables in the air gap of the internal combustion engine is of crucial importance. Methods for detecting these state variables exist in the prior art only for internal combustion engines, which are operated with the conventional combustion.

The present invention is therefore based on the object to provide a method for detecting various non-measurable state variables from the air gap of an internal combustion engine with exhaust gas recirculation and a correspondingly designed engine control for an internal combustion engine, with the simplest possible means a nearly exact determination of these non-measurable state variables, d , H. z. B. a nearly exact determination of the exhaust gas recirculation rate is possible.

This object is achieved by a method with the features of claim 1 and a motor controller with the features of claim 23. The dependent claims each define preferred and advantageous embodiments of the present invention.

In the context of the present invention, a method is provided for detecting state variables of a gas mixture in the air gap assigned to an internal combustion engine with exhaust gas recirculation. In this case, fresh air with a recirculated via the exhaust gas recirculation exhaust gas of the internal combustion engine at a mixing point of the air gap, which is assigned to a the internal combustion engine Compressor is arranged, mixed and fed the resulting gas mixture to the engine. The state variables of the gas mixture are determined at different points in the air gap with a model for a low-pressure exhaust gas recirculation path as part of the air gap, which simulates the behavior of the air gap or devices associated with the air gap with respect to the respective state variable to be determined.

As a result, it is possible according to the invention for various non-measurable state variables of the air gap to be detected, as a result of which an engine controller can better control the internal combustion engine.

Under the air gap is understood both a the internal combustion engine fresh air or a mixture of fresh air and exhaust gas supplying part as well as an exhaust gas from the engine laxative part, as well as an exhaust recirculating part.

According to the state of the art, methods for detecting state variables in air gaps in which the mixing point is arranged after the compressor are known, the pressure in the exhaust gas recirculation of such an air gap being high. In contrast, the mixing point is arranged in the present invention in front of the compressor, which is why the pressure in the exhaust gas recirculation of the air gap is lower. Due to the different pressure conditions, a method which requires the mixing point after the compressor can not be used for an air gap in which the mixing point is arranged in front of the compressor.

In the method according to the invention, a first replacement volume model, which replicates a first replacement volume in front of the compressor, can be used. In this case, a mass of a certain gas (eg oxygen or exhaust gas) stored in the first substitute volume can be determined by integrating all the mass flows flowing into and flowing out of the first substitute volume, which contain the particular gas, each of the inflowing and outflowing mass flows in one Ratio is considered, which corresponds to a rate of the particular gas in the respective mass flow.

By determining the stored mass of a specific gas in the first substitute volume, further state variables (for example an exhaust gas recirculation rate) can advantageously be determined, as will be explained below.

In the same way, in the case of a second replacement volume model, which simulates a second replacement volume between the compressor and the engine, a stored mass of a specific gas in the second substitute volume can be determined, which in turn enables further state variables to be determined.

As other models, a compressor speed model, a compressor model, a model for an exhaust gas recirculation valve and a radiator, and an exhaust pipe model for determining or determining other state variables can be used. For each model, output quantities of the model, which correspond to state variables, can be determined as a function of input variables of the model.

It should be noted that the output variables of the model determined by the models described above can be determined independently of one another depending on respective input variables of the model, whereby the determination of one of the output variables of the model can usually only depend on a subset of the input variables of the model.

In the context of the present invention, an engine control system for an internal combustion engine with exhaust gas recirculation is also provided. In this case, the engine control requires that the fresh air is mixed with the recirculated exhaust gas upstream of the compressor. In this case, the engine control is designed such that it controls the internal combustion engine and the air gap associated devices (compressor, charge air cooler, particulate filter, exhaust gas recirculation valve) depending on state variables of the gas mixture, which determines the engine control at different points of the air gap with the model for the low pressure exhaust gas recirculation line.

As a result, the engine control according to the invention is able to control a homogeneous combustion of the engine almost optimally.

The present invention is preferably suitable for a diesel engine with direct injection. Of course, the invention is not limited to this application, but can also be used in other internal combustion engines, such as. As a gasoline engine, are used.

The present invention will be explained in more detail below with reference to the accompanying drawings with reference to a preferred embodiment.

1 represents the air gap of an internal combustion engine.

2 shows a representation for explaining a motor filling model, which is composed of a known in the prior art and a part of the invention.

3 shows a coarse structure of the inventive part of 2 ,

In 1 is an embodiment of a diesel engine 1 , which is also often called abbreviated engine below, with a motor control 5 and associated air gap shown. In this case, fresh air passes through a first throttle 2 in the diesel engine 1 one. In a mixing point, which before a subsequent compressor 3 is arranged, the fresh air mixed with a recirculated exhaust gas. About a charge air cooler 4 and a second throttle 6 the gas mixture of fresh air and recirculated exhaust gas enters the combustion chamber of the cylinders of the diesel engine 1 , The exhaust gas from the combustion chambers is then passed through a turbine 7 and a particle filter 8th to a point at which a portion of the exhaust gas is recycled and the remainder leaves the air gap. The recirculated exhaust gas is via an exhaust gas recirculation cooler 9 and an exhaust gas recirculation valve 10 to the mixing point 11 returned. There is also a motor control 5 communicates with all air gap devices to record and control readings.

The first throttle valve serves this purpose 2 to increase a pressure difference across the exhaust gas recirculation valve 10 , An existing from fresh air and recirculated exhaust gas total mass flow is through the compressor 3 compressed and in the intercooler 4 cooled. The second throttle 6 has the task in a regeneration of the particulate filter 8th adjust the air ratio over the fresh air mass without exhaust gas recirculation. The compressed total mass flow arrives at a charge change together with an injected fuel mass into the combustion chamber.

2 schematically represents a motor filling model 21 which is from a known from the prior art part 23 and a new part 22 composed. Here, the amount of input variables of the engine filling model is set 21 from input sizes 24 for the known part 23 and input sizes 25 for the new part 22 together. The output sizes of the new part 22 are simultaneously input quantities for the known part 23 , where the output sizes of the known part 23 in output sizes 26 of the engine fill model 21 and input sizes for the new part.

3 shows the new part 22 out 2 in detail, which is a compressor speed model 101 , a first replacement volume model 102 , a compressor model 103 , a second replacement volume model 104 , a model 105 for the exhaust gas recirculation valve 10 and intercooler 4 , an exhaust pipe model 106 and an interface 107 between new and known part of the engine fill model 21 includes.

The compressor speed model 101 determined from a temperature T _{vV upstream} of the compressor, a boost pressure p _lad , with which the gas mixture from the compressor 3 in the direction of the engine 1 is passed, a pressure p _vV before the compressor 3 , a charge temperature T _lad , which the gas mixture after the compressor 3 and a reference _{total mass flow} dm _motref into the engine 1 a compressor speed n _V.

The first replacement volume model 102 , which is a first replacement volume before the compressor 3 replicates, determined depending on one in the compressor 3 flowing compressor total mass flow dm, one in front of the compressor 3 supplied fresh air mass flow dm _HFM , an ambient temperature T _A , an exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, an exhaust gas temperature T _{AGR of} the recirculated exhaust gas and an oxygen rate r _O2AGR in the _recirculated exhaust gas a modeled temperature T _ERS1 in the first replacement _volume , a modeled pressure p _ERS1 in the first replacement _volume , an exhaust gas recirculation rate r _AGRV in the compressor 3 and an oxygen rate r _O2V in the compressor 3 ,

The compressor model 103 determined depending on a modeled boost pressure p _ladmod , the oxygen rate r _O2V in the compressor 3 , the exhaust gas recirculation rate r _AGRV in the compressor 3 , the modeled pressure p _ERS1 im first replacement _volume , the modeled temperature T _ERS1 in the first replacement _volume and a compressor _speed n _V the compressor _{total mass flow} dm, the oxygen rate r _O2V in the compressor 3 and the exhaust gas recirculation rate r _O2V in the compressor 3 ,

The second replacement volume model 104 , which provides a second volume of replacement between the compressor 3 and the engine 1 simulates, determined depending on the _total compressor _{mass flow} dm, the oxygen rate r _O2V in the compressor, the exhaust gas recirculation rate r _AGRV in the compressor 3 , the charge _temperature T _lad , the boost pressure p _lad and an air _demand LA the modeled boost pressure p _ladmod , the reference _{total mass flow} dm _motref into the engine 1 , a total mass flow dm _mot into the engine 1 and an engine exhaust _{gas recirculation rate} r _AGRmot .

The model 105 for the exhaust gas recirculation valve 10 and the exhaust gas recirculation cooler 9 determined depending on an opening degree ÖG _{AGR of} the exhaust gas recirculation valve 10 , a temperature T _nPF after the particle filter 8th , a pressure p _nPF after the particulate filter 8th , the modeled pressure p _ERS1 in the first substitute _volume and the modeled temperature T _ERS1 in the first substitute _volume, the exhaust gas temperature T _AGR in the recirculated exhaust gas and the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas.

The exhaust pipe model 106 , which is an exhaust pipe after the particulate filter 8th modeled, determined depending on the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, an ambient pressure or atmospheric pressure p _A , the temperature T _nPF after the particulate filter 8th and a total mass flow dm _PF through the particulate filter 8th the pressure p _nPF after the particle filter 8th ,

The engine fill model 21 is based on a physical description of the processes in the air system or the air gap of the diesel engine 1 , The intake, which is in the flow direction in front of the diesel engine 1 located, and the exhaust tract, which is in the flow direction to the diesel engine 1 are mapped with a filling and emptying method as a zero-dimensional model. For a zero-dimensional model, changes in state variables are not modeled as a function of spatial dimensions but only as a function of time. It is assumed that transient processes for small time intervals within a replacement system (first and second substitution volume model) can be considered as constant, assuming immediate mixing as well as pressure and temperature compensation through infinitely high velocities of sound.

In the first replacement volume model 102 a mixing of fresh air and exhaust gas is simulated and the modeled pressure p _{ERS1 calculated} in the first replacement _volume model from a mass balance of incoming and outgoing gases. It is assumed in the modeling that the specific heat capacities of fresh air and exhaust gas are the same and no heat transfer takes place.

A total mass m _ERS1 stored in the first substitute _volume is determined from a temporal integration of a mass flow balance depending on the fresh air mass flow dm _HFM , the exhaust mass flow dm _{AGR of} the recirculated exhaust gas and the total _compressor mass flow dm _V , as the following formula shows. m _ERS1 = ∫ (dm _HFM + dm _AGR - dm _V ) dt (1).

Among the previously mentioned assumptions, the modeled temperature T _ERS1 in the first spare volume using a mixture balance exercise dependent on the fresh air mass flow dm _HFM, the ambient temperature T _A, the exhaust gas mass flow dm _EGR of recirculated exhaust gas, the exhaust gas temperature T _AGR of recirculated exhaust gas and the compressor total mass flow dm _V calculate as follows.

The modeled pressure p _ERS1 in the first replacement _volume is determined with the ideal gas equation as a function of the modeled temperature T _ERS1 in the first replacement volume, the gas constant R, a volume V _{ERS1 of} the first replacement volume, and the total mass m _{ERS1 of} the first replacement volume.

The exhaust gas recirculation rate r _AGRV in the compressor 3 is defined as the quotient of an exhaust gas mass m _{AGRERS1 of} the recirculated exhaust gas stored in the first substitute _volume to the total mass m _{ERS1 of} the first substitute _volume .

The exhaust gas mass m _{AGRERS1 of} the recirculated exhaust gas stored in the first substitute _{volume is thereby} integrated by the exhaust gas mass flow flowing in and out depending on the exhaust gas mass flow dm _AGR of the recirculated exhaust gas, the _total compressor mass flow dm _AGRV and the exhaust gas recirculation rate r _AGRV in the compressor 3 calculated. m _AGRERS1 = ∫ (dm _AGR - dm _V × r _AGRV ) dt (5).

Analogously, the oxygen rate r _O2V in the compressor 3 is calculated as the quotient of the oxygen mass m _O2ERS1 stored in the first substitute _volume to the total mass m _ERS1 stored in the first substitute _volume .

The oxygen _mass m _O2ERS1 in the first replacement _volume is determined by an integral depending on the fresh air mass flow dm _HFM , an oxygen percentage r _O2A (r _O2A is normally 23.15%) in the ambient air, the exhaust gas mass flow dm _{EGR of} the recirculated exhaust gas, the oxygen rate r _O2AGR im _recirculated exhaust gas, the compressor _{mass flow} dm _V and the oxygen rate r _O2V in the compressor 3 calculated. m _O2FRS1 = ∫ (dm _HFM × r _O2A + dm _AGR × r _O2AGR - dm _V × r _O2V ) dt (7).

In the model 105 for the exhaust gas recirculation valve 10 and the exhaust gas recirculation cooler 9 the exhaust gas mass flow dm _AGR and the exhaust gas temperature T _{AGR of} the recirculated exhaust gas are determined. In this case, the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas from the pressure p _nPF after the particulate filter 8th or after the turbine 7 and the modeled pressure p _ERS1 in the first substitute volume determined oversampling, that is, that calculations required for the determination are carried out more frequently or with a narrower time grid than with determinations of other state variables. It is lift out that, as previously indicated, the modeled pressure p is _ERS1 drawn from the first spare volume and not a measured pressure for the calculation. As a result, the model stabilizes itself and provides an undisturbed signal, that is, a fluctuation range of the outputs (exhaust gas mass flow dm _AGR and exhaust gas temperature T _{EGR of} the recirculated exhaust gases) is lower than a concept in which the measured pressure is used in the concept where the measured pressure is used.

The compressor _speed n _V is dependent on a corrected compressor _speed n _Vk , the temperature T _vV before the compressor 3 and a reference _temperature T _0V in the compressor 3 when mapping maps (the recording or recording of compressor maps is carried out at the reference _temperature T _0V ) calculated as follows.

The corrected compressor _speed n _Vk is determined from a first map KF ₁ as a function of a compressor pressure ratio Π _V and a corrected _total compressor mass _flow dm _Vk . n _Vk = KF1 (Π _V , dm _Vk ) (9)

In order to avoid a feedback in the model, the input variables (the compressor pressure ratio Π _V and the corrected compressor mass _flow dm _Vk ) are determined from measured values. The Compressor pressure ratio Π _V is calculated from the measured boost pressure p _lad and the pressure p _vV before the compressor as follows.

The corrected _{total compressor mass flow} dm _Vk can be _calculated as follows from the reference _{total mass flow} dm _motref into the engine 1 , A reference pressure p _0V in the compressor when recording maps (the recording or detection of compressor _maps takes place at the reference pressure p _0V ), the pressure p _vV before the compressor 3 , the temperature T _vV before the compressor 3 and the reference _temperature T _0V in the compressor 3 determine when recording maps.

It is assumed that an error can be neglected by the mass storage behavior of the intake.

Here, a reference mass flow (eg dm _motref ) is understood as _meaning a mass flow which is determined from measured variables. In contrast, for example, the modeled pressure p _ERS1 and the modeled temperature T _ERS1 in the first substitute _{volume are} not determined from measured quantities, but are modeled or determined using a model.

In the compressor model 103 is the corrected compressor _{mass flow} dm _Vk , as follows, calculated from a second map KF ₂ in response to a modeled compressor _pressure ratio Π _Vmod and the corrected compressor _speed n _Vk . dm _Vk = KF ₂ (Π _Vmod , n _Vk ) (12)

The modeled compressor pressure ratio Π _Vmod is determined from the modeled boost pressure p _ladmod or the modeled pressure after the compressor and the modeled pressure p _ERS1 in the first replacement _volume as follows.

The compressor _{mass flow} dm is calculated from the corrected compressor _{mass flow} dm _Vk , the pressure p _ERS1 in the first substitute _volume , the reference pressure p _0V in the compressor 3 when recording maps, the reference _temperature T _0V in the compressor 3 when _{assuming maps} and the temperature T _ERS1 in the first replacement _volume determined as follows.

The second replacement volume model 104 Holds part of the air gap from the compressor 3 up to a suction pipe of the engine 1 together. In the second replacement volume model 104 the modeled boost pressure p _{ladmod is calculated} from the incoming and outgoing mass flows and a mixture of exhaust gas and fresh air is simulated.

The engine 1 sucked modeled total mass flow dm _mot into the engine 1 is calculated from the modeled charge pressure p _ladmod , the charge _temperature T _lad , a displacement V _H , an engine speed n _mot , the air demand LA, the gas constant R and a number a of cylinders of the engine 1 calculated as follows.

The reference _{total mass flow} dm _motref into the engine 1 is also calculated from the measured boost pressure p _lad , the charge temperature T _lad , the gas constant R, the stroke volume V _H , the engine speed n _mot and the number of cylinders a as follows, using the air effort LA.

The total mass m _ERS2 stored in the second substitute _volume results from a temporal integration of a mass flow _{balance as a} function of the total compressor mass flow dm _V and the total mass flow dm _mot into the engine 1 as follows. m _ERS2 = ∫ (dm _V - dm _mot ) dt (17)

In the following, it is assumed that the temperature in the second substitute volume of the charge temperature T _lad or the measured temperature after the charge air cooler 4 equivalent. Thus, the pressure p _ladmod in the second substitute _volume with the ideal gas equation depends on the mass m _ERS2 in the second substitute volume, the gas constant R, the volume V _{ERS2 of} the second substitute volume and the charge _temperature T _lad as follows.

The engine exhaust _{gas recirculation rate} r _AGRmot is determined from a quotient of an exhaust gas mass m _AGRERS2 stored in the second substitute _volume and the total mass m _ERS2 in the second substitute _volume as follows.

The exhaust gas mass m _AGRERS2 stored in the second substitute _volume is dependent on the total _mass compressor flow dm _V , the exhaust gas rate r _AGRV in the compressor, by an integration of the incoming and outgoing exhaust gas mass flows 3 , the total mass flow dm _mot into the engine 1 and the engine exhaust _{gas recirculation rate} r _AGRmot 1 calculated as follows. m _AGRERS2 = ∫ (dm _V × r _AGRV -dm _mot × r _AGRmot ) dt (20)

Analogously, the oxygen rate r _{O2mot results} in the total mass _flow into the engine 1 depending on an oxygen mass m _O2ERS2 in the second replacement volume and the total mass m _ERS2 in the second volume replacement as follows.

The oxygen _mass m _O2ERS2 in the second substitute _volume can be determined by an integration of the incoming and outgoing mass flows depending on the compressor _total mass flow dm, the oxygen rate r _O2V in the compressor 3 , the total mass flow dm _mot into the engine 1 and the oxygen rate r _O2mot in the total mass _flow into the engine 1 determine as follows. m _O2ERS2 = ∫ (dm _V × r _O2V - dm _mot × r _O2mot ) dt (22)

In the exhaust pipe model 106 become the total mass flow and the pressure in an exhaust pipe 12 calculated the air gap or the air system. After continuity law, the total mass flow dm _AGL results in the exhaust pipe 12 from a difference of a total mass flow dm _PF through the particulate filter 8th and the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas as follows. dm _AGL = dm _PF - dm _EGR (23)

The pressure p _nPF after the particle filter 8th or in the exhaust pipe 12 can from the ambient pressure p _A and a pressure drop Δp _AGL in the exhaust pipe 12 be approximated as follows, wherein the pressure loss Δp _AGL from a characteristic as a function of the total mass flow dm _AGL in the exhaust pipe 12 is approximated. p _NPF = p + Ap _{AGL A} (24)

LIST OF REFERENCE NUMBERS

1

diesel engine

2

first throttle

3

compressor

4

Intercooler

5

motor control

6

second throttle

7

turbine

8th

particulate Filter

9

Exhaust gas recirculation cooler

10

Exhaust gas recirculation valve

11

Mixing point of fresh air and recirculated exhaust gas

12

exhaust pipe

21

Motor filling model

22

new part

23

known part

24

input variables

25

input variables

26

outputs

101

Compressor speed model

102

first replacement volume model

103

compressor model

104

second replacement volume model

105

Model for exhaust gas recirculation valve and radiator

106

Exhaust pipe model

107

Interface between new and known part of the engine fill model

T _vV

Temperature before the compressor

p _lad

boost pressure

p _vV

Pressure in front of the compressor

T _lad

charge temperature

dm _V

Compressor total mass flow

dm _HFM

Fresh air mass flow

T _A

ambient temperature

ÖG _AGR

Opening degree of the exhaust gas recirculation valve

T _nPF

Temperature after the particle filter

dm _motref

Reference total mass flow into the engine

n. _V

Compressor speed

r _O2V

Oxygen rate in the compressor

r _AGRV

Exhaust gas recirculation rate in the compressor

p _ERS1

modeled pressure in the first replacement volume

T _ERS1

modeled temperature in the first replacement volume

T _AGR

Exhaust gas temperature in the recirculated exhaust gas

dm _AGR

recirculated exhaust gas mass flow

_pnPF

Pressure after the particle filter

p _A

Ambient pressure or atmospheric pressure

r _O2AGR

Oxygen rate in recirculated exhaust gas

dm _PF

Total mass flow through particle filter

p _ladmod

modeled boost pressure after compressor

dm _mot

Total mass flow into the engine

r _AGRmot

Engine exhaust gas recirculation rate

LA

volumetric efficiency

dm _Vk

corrected total compressor mass flow

r _O2A

Oxygen rate in the ambient air

T _0V

Reference temperature in the compressor when recording maps

p _0V

Reference pressure in the compressor when recording maps

dm _AGL

Total mass flow in the exhaust pipe

KF ₁

first map

KF ₂

second map

Δp _AGL

Pressure loss in the exhaust pipe

Π _Vmod

modeled compressor pressure ratio

m _ERS1

Total mass in the first replacement volume

m _AGRERS1

Exhaust mass of the recirculated exhaust gas in the first replacement volume

n _Vk

corrected compressor speed

dm _Vk

corrected compressor mass flow

Π _V

Compressor pressure ratio

n _mot

Engine speed

a

number of cylinders

dmag

Exhaust gas mass flow

Claims (24)

Method for detecting state variables of a gas mixture in an air gap assigned to an internal combustion engine with exhaust gas recirculation, fresh air being supplied with an exhaust gas of the internal combustion engine which is recirculated via the exhaust gas recirculation ( 1 ) at a mixing point ( 11 ) mixed the air gap and the resulting gas mixture the internal combustion engine ( 1 ), characterized in that the mixing point ( 11 ) in front of a combustion engine ( 1 ) associated compressor ( 3 ) and that state variables of the gas mixture, which comprise an exhaust gas recirculation rate and an oxygen rate, at different points of the air gap with a model ( 21 ) for a low-pressure exhaust gas recirculation path as part of the air path, which with respect to the state variable to be determined in each case the behavior of the air gap or devices contained in the air gap, which a compressor ( 3 ), a charge air cooler ( 4 ), a particle filter ( 8th ) and an exhaust gas recirculation cooler ( 9 ), imitates, is determined. Method according to claim 1, characterized in that by a first substitute volume model ( 102 ), which has a first replacement volume in front of the compressor ( 3 ), depending on one in the compressor ( 3 ) compressor mass flow dm _V , one in front of the compressor ( 3 supplied fresh air mass flow dm _HFM , an ambient temperature T _A , an exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, an exhaust gas temperature T _{AGR of} the recirculated exhaust gas and an oxygen rate r _O2AGR in the _recirculated exhaust gas a modeled temperature T _ERS1 in the first substitute _volume , a modeled pressure p _ERS1 , im first replacement _volume , an exhaust gas recirculation rate r _AGRV in the compressor ( 3 ) and an oxygen rate r _O2V in the compressor ( 3 ) is determined. A method according to claim 2, characterized in that a stored in the first spare volume total mass m _ERS1 using a mass flow balance of the fresh air mass flow dm _HFM, the exhaust gas mass flow dm _EGR of recirculated exhaust gas and the compressor total mass flow dm _V is determined. Method according to Claim 3, characterized in that the total mass m _ERS1 stored in the first substitute _{volume is determined} from the fresh air mass flow dm _HFM , the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas and the total _compressor mass flow dm as follows: m _ERS1 = ∫ (dm _HFM + dm _AGR - dm _V ) dt. Method according to claim 3 or 4, characterized in that the modeled pressure p _ERS1 in the first substitute _{volume is determined} using the ideal gas equation from the total mass m _ERS1 stored in the first substitute _volume and the modeled temperature T _ERS1 in the first substitute _volume . Method according to one of claims 3-5, characterized in that the modeled pressure p _ERS1 in the first substitute volume from the total mass m _ERS1 stored in the first substitute _volume , the gas constant R, a volume V _{ERS1 of} the first substitute volume and the modeled temperature T _ERS1 in the first substitute volume Replacement volume is determined as follows:

Method according to one of claims 2-6, characterized in that the modeled temperature T _ERS1 in the first replacement _volume using a mixture _balance of the fresh air mass flow dm _HFM , the ambient temperature T _A , the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, the exhaust gas temperature T _{AGR of} the recirculated exhaust gas and the compressor total mass flow dm is determined. Method according to one of claims 2-7, characterized in that the modeled temperature T _ERS1 in the first substitute _volume from the fresh air mass flow dm _HFM , the ambient temperature T _A , the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, the exhaust gas temperature T _{AGR of} the recirculated exhaust gas and the compressor total mass flow dm is determined as follows:

A method according to claims 2-8, characterized in that a mass of a given gas of the gas mixture stored in the first substitute volume is determined by integrating a difference of all mass flows flowing into and out of the first substitute volume containing the particular gas, each of the inflowing and outflowing masses outgoing mass flows are taken into account in a ratio which corresponds to a rate of the particular gas in the respective mass flow. Method according to one of Claims 2-9, characterized in that an exhaust gas mass M _{AGRERS1 of} the recirculated exhaust gas stored in the first substitute _volume is determined from the exhaust gas mass flow dm _AGR of the recirculated exhaust gas, the _total compressor mass flow dm and the exhaust gas recirculation rate r _AGRV in the compressor ( 3 ) is determined as follows: m _AGRERS1 = ∫ (dm _AGR - dm _V × r _AGRV ) dt. Method according to one of claims 2-10, characterized in that an oxygen _mass m _O2ERS1 stored in the first substitute _volume from the fresh air mass flow dm _HFM , an oxygen rate r _O2A in the ambient air, the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, the oxygen rate r _O2AGR in the _recirculated exhaust gas , the _total compressor _{mass flow} dm _V and the oxygen rate r _O2V in the compressor are determined as follows: m _O2ERS1 = ∫ (dm _HFM × r _O2A + dm _AGR × r _O2AGR - dm _V × r _O2V ) dt. Method according to one of claims 2-11, characterized in that by a model ( 105 ) for an exhaust gas recirculation valve ( 10 ) and an exhaust gas recirculation cooler ( 9 ), which are both contained in the air gap, depending on an opening degree ÖG _{AGR of} the exhaust gas recirculation valve ( 10 ), a temperature T _nPF after a particle filter contained in the air _gap ( 8th ), a pressure p _nPF after the particulate filter ( 8th ), the modeled pressure p _ERS1 in the first substitute _volume and the modeled temperature T _ERS1 in the first substitute _volume, the exhaust gas temperature T _AGR in the recirculated exhaust gas and the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas is determined. A method according to claim 12, characterized in that by an exhaust pipe model ( 106 ) associated with the air gap exhaust pipe ( 12 ) after the particle filter ( 8th ) depending on the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas, an ambient pressure p _A , the temperature T _nPF after the particulate filter ( 8th ) and a total mass flow dm _PF through the particulate filter ( 8th ) the pressure p _nPF after the particle filter ( 8th ) is determined. A method according to claim 13, characterized in that a total mass flow dm _AGL in the exhaust pipe ( 12 ) from the difference of the mass flow dm _PF through the particulate filter ( 8th ) and the exhaust gas mass flow dm _{AGR of} the recirculated exhaust gas is calculated. A method according to claim 13 or 14, characterized in that a pressure in a sampling point at a the internal combustion engine ( 1 ) subsequent section of the air gap, at which the exhaust gas is recirculated, from a sum of the ambient pressure p _A and a pressure loss Δp _AGL in the Exhaust pipe ( 12 ), wherein the pressure loss from a characteristic curve depends on a total mass flow dm _AGL in the exhaust gas line ( 12 ) is determined. A method according to any of claims 12-15, characterized in that the exhaust gas mass flow dm _EGR of recirculated exhaust gas from the pressure p _NPF after the particulate filter ( 8th ) and the modeled pressure p _ERS1 in the first spare _{volume is} oversampled. Method according to one of claims 1-16, characterized in that by a compressor speed model ( 101 ) depending on a temperature T _vV before the compressor ( 3 ), a boost pressure p _lad , with which the gas mixture from the compressor ( 3 ) in the direction of the internal combustion engine ( 1 ), a pressure p _vV before the compressor ( 3 ), a charge temperature T _lad , which the gas mixture after the compressor ( 3 ), and a reference _{mass flow} dm _motref into the internal combustion engine ( 1 ) a compressor speed n _{V is} determined. A method according to claim 17, characterized in that a corrected _{total mass compressor flow} dm _Vk from the reference _{total mass flow} dm _motref into the internal combustion engine ( 1 ), a reference _pressure p _0V in the compressor ( 3 ) when determining _maps , the pressure p _vV before the compressor ( 3 ), the temperature T _vV before the compressor ( 3 ) and a reference _temperature T _0V in the compressor ( 3 ) is determined when determining the maps as follows:

Method according to one of claims 17-18, characterized in that by a second replacement volume model ( 104 ), which provides a second volume of replacement between the compressor ( 3 ) and the internal combustion engine ( 1 ), depending on the compressor _{total mass flow} dm _V , the oxygen rate r _O2V in the compressor ( 3 ), the exhaust gas recirculation rate r _AGRV in the compressor ( 3 ), the charge _temperature T _lad , the boost pressure p _lad and an air _effort LA the modeled boost pressure p _ladmod , the reference _{total mass flow} dm _motref into the internal combustion engine ( 1 ), a total mass flow dm _mot in the internal combustion engine ( 1 ) and the engine exhaust _{gas recirculation rate} r _{AGRmot is} determined. A method according to claim 19, characterized in that a mass of a given gas stored in the second substitute volume is determined by integrating a difference of all the second volume of the incoming and outgoing mass flows containing the particular gas, each of the incoming and outgoing mass flows in one Ratio is considered, which corresponds to a rate of the particular gas in the respective mass flow. Method according to claim 19 or 20, characterized in that an exhaust mass m _AGRERS2 in the second replacement _volume depends on the _total compressor mass flow dm _V , the exhaust gas recirculation rate r _AGRV in the compressor ( 3 ), the total mass flow dm _mot in the internal combustion engine ( 1 ) and the engine exhaust _{gas recirculation rate} r _AGRmot is determined as follows: m _AGRERS2 = ∫ (dm _V × r _AGRV - dm _mot × r _AGRmot ) dt. Method according to one of the preceding claims, characterized in that the internal combustion engine is a diesel engine ( 1 ). Engine control 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 ) at a mixing point ( 11 ) of an internal combustion engine ( 1 ) associated air gap and a resulting gas mixture the internal combustion engine ( 1 ), characterized in that the engine control ( 5 ) is configured such that it the internal combustion engine ( 1 ) depending on state variables of the gas mixture, which comprise an exhaust gas recirculation rate and an oxygen rate, which controls them at different points of the air gap by using a model ( 21 ) for a low-pressure exhaust gas recirculation path as part of the air path, which with respect to the state variable to be determined in each case the behavior of the air gap or devices contained in the air gap, which a compressor ( 3 ), a charge air cooler ( 4 ), a particle filter ( 8th ) and an exhaust gas recirculation cooler ( 9 ), replicates, determined, with the mixing point ( 11 ) in front of the internal combustion engine ( 1 ) associated compressor ( 3 ) is arranged. Motor control according to claim 23, characterized in that the engine control ( 5 ) is configured for carrying out the method according to one of claims 1-22.

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