DE102021207547A1 - Method and device for controlling a gas temperature at the outlet of a charge air cooler of an internal combustion engine with low-pressure exhaust gas recirculation - Google Patents

Abstract

Method for controlling a gas temperature (TLLKDs,Soll) at the outlet of a charge air cooler (14) of an internal combustion engine (25) with a low-pressure exhaust gas recirculation (41), wherein exhaust gas is fed back via the low-pressure exhaust gas recirculation (41), wherein in A consumption-optimized gas temperature (TLLKDs,Eco) at the outlet of the intercooler (14) for the internal combustion engine (25) is determined as a function of a current speed (neng) and a current target cylinder charge (rlsoll), the gas temperature (TLLKDs, Target) at the outlet of the intercooler (14) via a control of the intercooler (14) is set in such a way that the consumption-optimized gas temperature (TLLKDs,Eco) is continuously increased until a current ignition angle efficiency (ηZW) falls below a predeterminable ignition angle efficiency (ηZW,Lim). or a predetermined maximum allowable gas temperature (TLLKDS, Lim) is exceeded, where if the current Ignition angle efficiency (ηZW,Lim) falls below the predefinable ignition angle efficiency (ηZW,Lim), the recirculation of exhaust gas via the low-pressure exhaust gas recirculation (41) is reduced and/or the gas temperature (TLLKDs,Soll) at the outlet of the intercooler (14) is reduced , in order to avoid a further entry of moisture by recirculated exhaust gas via the low-pressure exhaust gas recirculation system (41).

Description

The invention relates to a method and a device for controlling a gas temperature at the outlet of a charge air cooler of an internal combustion engine with low-pressure exhaust gas recirculation.

State of the art

The task of the intercooler in supercharged internal combustion engines is to cool the air after compression and before it is fed to the combustion chamber in order to achieve an increase in air density. This means that more fuel can be converted and the performance of the engine can be increased. If the humidity of the fresh air sucked in is correspondingly high, the cooling can lead to condensation. This effect is further increased in systems with built-in low-pressure exhaust gas recirculation, since part of the exhaust gas, which contains the additional proportion of water from combustion, is mixed with the fresh air and then flows through the intercooler. In order to avoid damaging the engine due to too much condensed water entering the combustion chamber or damaging the charge air cooler due to corrosion, the amount of condensation in the charge air cooler must be limited. Upstream of the compressor, e.g. B. in the EGR cooler or at the mixing point with the fresh air, no condensation should take place, otherwise the compressor wheel can be damaged by falling drops.

Disclosure of Invention

The invention relates to a method and a device for controlling a gas temperature at the outlet of a charge air cooler of an internal combustion engine with low-pressure exhaust gas recirculation and a computer program on a storage medium for executing the method.

In a first aspect, the invention relates to a method for controlling a gas temperature at the outlet of a charge air cooler of a supercharged internal combustion engine with low-pressure exhaust gas recirculation, exhaust gas being returned via the low-pressure exhaust gas recirculation, depending on a current speed and a current setpoint -Cylinder filling, a consumption-optimized gas temperature at the outlet of the charge air cooler for the internal combustion engine is determined, with the gas temperature at the outlet of the charge air cooler being adjusted via control of the charge air cooler as a function of a condensation avoidance strategy in such a way that the consumption-optimized gas temperature is continuously increased until a current ignition angle efficiency reaches a specifiable ignition angle efficiency falls below or a specifiable maximum permissible gas temperature is exceeded, wherein when the current ignition angle efficiency falls below the specifiable ignition angle efficiency , reducing the recirculation of exhaust gas via the low-pressure exhaust gas recirculation and/or reducing the gas temperature at the outlet of the intercooler in order to avoid further moisture input from exhaust gas that is recirculated via the low-pressure exhaust gas recirculation.
By continuously increasing the gas temperature, the condensation in the intercooler can be adjusted in a controlled manner.
The method has the particular advantage that, depending on the ignition angle efficiency, an optimized strategy for avoiding condensation
can be carried out in the intercooler. By limiting the gas temperature at the outlet of the intercooler, the condensation in the intercooler can be adjusted in a controlled manner. By intervening in the recirculated exhaust gas rate of the low-pressure exhaust gas recirculation, condensation on the compressor that forms droplets can be prevented. Furthermore, damage to components such as the internal combustion engine due to condensation in the form of corrosion can be avoided or reduced.

Furthermore, the predefinable maximum permissible gas temperature can correspond to a gas temperature for which a full-load request from the internal combustion engine is necessary at the current speed.
It is advantageous to limit the maximum permissible gas temperature by the full-load requirement at the current speed in order to ensure optimal operation of the internal combustion engine. As a result, the necessary cooling capacity can be made available by the intercooler.
The load point-dependent gas temperature for optimal fuel consumption can be increased through condensation management up to the gas temperature for maximum engine performance.

Furthermore, the definable ignition angle efficiency can be determined as a function of the current speed and the current setpoint cylinder charge, and if the current ignition angle efficiency falls below the definable ignition angle efficiency, the condensation avoidance strategy can be deactivated. This is particularly advantageous since the condensation for the internal combustion engine can be controlled particularly easily by means of the ignition angle efficiency. The ignition angle efficiency is a particularly advantageous control variable for the internal combustion engine.

Furthermore, the definable ignition angle efficiency can correspond to an ignition angle efficiency Chen at which a knock limit occurs, preferably an ignition angle efficiency before a knock limit is reached.
It is particularly advantageous if the knock limit of the combustion of the internal combustion engine can be used to control the method, or an ignition angle efficiency before the knock limit is reached.

Furthermore, with an active condensation avoidance strategy, the consumption-optimized gas temperature for the intercooler is gradually increased by a control difference, with the control difference being determined as a function of a difference between a dew point temperature and a corrected consumption-optimized gas temperature and a water rate for the intercooler.
This regulation has the particular advantage that the condensation in the intercooler can be controlled in a controlled manner. The control difference is continuously increased as long as the condensation avoidance strategy is active.

Furthermore, the dew point temperature can be determined as a function of a second saturation vapor pressure and a relative humidity downstream of the intercooler and upstream of the throttle valve.

Furthermore, the one water rate for the intercooler can be determined as a function of a first water mass.

Furthermore, the first water mass can be determined as a function of a gas mass flow through the intercooler, a second pressure downstream of the intercooler and upstream of the throttle valve and a second specific humidity downstream of the compressor and upstream of the intercooler.

Furthermore, the second specific humidity can be determined as a function of a first specific humidity at the location of a mixing point, a first pressure and a first saturation vapor pressure downstream of the compressor and upstream of the intercooler.

Furthermore, the relative humidity can be determined as a function of the second pressure and a second saturation vapor pressure downstream of the compressor and upstream of the intercooler.

Furthermore, the second saturation vapor pressure can be determined as a function of a second temperature downstream of the intercooler and upstream of the throttle valve.

Furthermore, the first saturation vapor pressure can be determined as a function of a first temperature downstream of the compressor and upstream of the intercooler.

In further aspects, the invention relates to a device, in particular a control device and a computer program, which are set up, in particular programmed, to carry out one of the methods. In yet another aspect, the invention relates to a machine-readable storage medium on which the computer program is stored.

character list

• 1 eine schematische Darstellung einer Brennkraftmaschine mit einer Niederdruck-Abgasrückführung,
• 2 den beispielhaften Ablauf eines Verfahrens zur Steuerung einer Gastemperatur am Austritt eines Ladeluftkühlers einer Brennkraftmaschine mit einer Niederdruck-Abgas-Rückführung.

The invention is described in more detail below with reference to the accompanying drawings and using exemplary embodiments. show it

• 1 a schematic representation of an internal combustion engine with a low-pressure exhaust gas recirculation,
• 2 the exemplary sequence of a method for controlling a gas temperature at the outlet of a charge air cooler of an internal combustion engine with low-pressure exhaust gas recirculation.

Description of the exemplary embodiments

1 shows a schematic representation of an internal combustion engine 25 with a fresh air system 48, via which combustion air is supplied to the internal combustion engine 25, and an exhaust system 49, via which exhaust gases 51 are discharged from the internal combustion engine 25 in the direction of flow.

The following is arranged in the fresh air system 48, seen in the flow direction of the fresh air 50: a pressure sensor 1, which determines a pressure p ₀ , a temperature sensor 2, which determines a temperature T ₀ , an air filter 3, an air mass hot film sensor (HFM) 5, a temperature sensor 6 , which determines a temperature T ₁₀ , a pressure sensor 7, which determines a pressure p ₁₀ , a fresh air throttle valve 8, a pressure sensor 9, which determines a pressure p ₁₁ , a temperature sensor 10, which determines a temperature T ₁₁ , a compressor 12 of an exhaust gas turbocharger 47, a pressure and temperature sensor 13, which determines a temperature T ₂₀ and a pressure p ₂₀ , an intercooler 14 with a volume V ₂₁ , a temperature sensor 16, which determines a temperature T ₂₁ , a pressure sensor 17, which determines a pressure p ₂₁ determined, a throttle valve 19, a pressure sensor 20, which determines a pressure p ₂₂ , and a temperature sensor 21, which determines a temperature T ₂₂ . The values described can e.g. B. as sensor values or as model values. In a preferred embodiment, the pressure sensor 1 determines the ambient pressure and the temperature sensor 2 determines the ambient temperature. A fresh air throttle flap 8 and a LP EGR valve 40 for throttling the LP EGR mass flow are installed.

The intercooler 14 is also connected to a coolant device. The coolant device consists of a controllable coolant pump and a coolant circuit, with coolant being pumped through a cooler 60 connected to the intercooler 14 . The temperature of charge air cooler 14 is controlled by a model calculated on control unit 100, control unit 100 regulating a flow rate or a coolant mass flow, preferably by controlling the coolant pump. Furthermore, a fan for cooling the radiator 60 can be attached to the radiator 60, with the controller 100 adapting the speed of the fan to a desired coolant temperature. In an alternate embodiment, the radiator 60 may include an adjustable fin structure covering the radiator 60, where the fin structure may be opened or closed by the controller 100 to control coolant temperature. Furthermore, a 3-way valve 61 can be installed in the coolant circuit, a bypass around the intercooler 14 being produced by means of the 3-way valve 61 , so that the coolant circuit no longer runs via the cooler 60 . The 3-way valve 61 is controlled via the control unit 100 in order to regulate the coolant temperature.

The following is arranged in the exhaust system 49, starting from the internal combustion engine 25 in the flow direction of the exhaust gas 51: a pressure sensor 26, which determines a pressure p ₃ , a temperature sensor 27, which determines a temperature T ₃ , an exhaust gas turbine 30, a lambda probe 56, which determines the air-fuel ratio in the exhaust system 49, an oxidation catalytic converter (DOC) 31, a nitrogen oxide storage catalytic converter 32, a particulate filter (PF) 33, a temperature sensor 34 that determines a temperature T ₅₀ , a pressure sensor 35 that determines a pressure p ₅₀ determined. The values of the engine speed n _eng and the fuel mass m _fuel supplied are also available as sensor values or model values, provided by a control unit 100, for example.

A low-pressure exhaust gas recirculation line (LP-EGR line) 41 branches off from the exhaust system 49 downstream of the DPF 33, i.e. on a low-pressure side of the exhaust system 49 Fresh air system 48 opens.

The following is arranged along the LP-EGR line 41, starting from the branch of the exhaust system 49 in the flow direction of a mass flow: an LP-EGR cooler 37 with an LP-EGR bypass 38, a temperature sensor 39, which determines a temperature T _LPEGR . The pressure drop across the LP EGR valve 40 can be determined via a differential pressure sensor 42 . The values described can e.g. B. as sensor values or as model values. Alternatively or additionally, an exhaust gas flap 8 can also be installed in the exhaust gas system 49 instead of the fresh air throttle 8 .
This point is also referred to as a mixing point 43, at which the returned exhaust gas 51 mixes with the fresh air 50 from the air supply system.
The pressure p ₁₁ and the temperature T ₁₁ are preferably determined in the area downstream of the mixing point 43 and upstream of the compressor 12 . Furthermore, the pressure p ₂₀ and the temperature T ₂₀ are preferably determined downstream of the compressor 12 and upstream of the charge air cooler 14 . The pressure p ₂₁ and the temperature T ₂₁ are preferably determined downstream of the intercooler 14 and upstream of the throttle valve 19 . Furthermore, the pressure p ₂₂ and the temperature T ₂₂ downstream of the throttle valve 19 and upstream of the internal combustion engine 25 are determined.

In the 2 shows the example of the procedure. In a first step 200, a first temperature T ₂₀ is read in by control unit 100 and a first saturation vapor pressure p _Sαt,20 is determined as a function of a first model M ₁ stored on control unit 100 . Alternatively, the saturation vapor pressure p _Sat,20 can also be determined as a function of a first map K ₁ as a function of the first temperature T ₂₀ , the saturation pressures being stored as a function of the temperature in the map K ₁ . The first temperature T ₂₀ is determined downstream of the compressor 12 and upstream of the charge air cooler 14 .

Subsequently, in a step 205, depending on a second model M ₂ stored on the control unit 100, a second specific _{humidity φ 20 , mdl} determined. The second specific humidity φ _20,mdl corresponds to a modeled specific humidity at the location downstream of the compressor 12 and upstream of the charge air cooler 14.
The first pressure p ₂₀ is preferably determined between the compressor 12 and the charge air cooler 14 . The first specific humidity φ ₁₂ is preferably at the site of the mixing point 43, downstream of the mixing point 43 and upstream of the compressor 12, determined.
Alternatively, the second specific humidity φ _20,mdl can be determined as a function of a second characteristic map K ₂ as a function of the determined first saturation vapor pressure p _Sat,20 , the first pressure p ₂₀ and the first specific humidity φ ₁₂ , with the specific humidity φ _{20, mdl} is stored in the second map K ₂ as a function of these variables.

In a step 210, a second saturation vapor pressure p _sat,21 is determined from a second temperature T ₂₁ as a function of a third model M ₃ stored on control unit 100 .
The second temperature T ₂₁ is determined downstream of the intercooler 14 and upstream of the throttle valve 19 .
Alternatively, the second saturation vapor pressure p _sat,21 can be determined from the second temperature T ₂₁ as a function of a third map K ₃ , the second saturation vapor pressure p _sat,21 being stored in the third map K ₃ as a function of the second temperature T ₂₁ .

In a step 215, the second saturation vapor pressure p _sat,21 determined in step 210, a second pressure p ₂₁ and the second specific humidity φ _20,mdl determined in step 205 are converted into a modeled relative humidity φ _21,mdl from a value stored on the control unit 100 stored fourth model M ₄ determined. The second pressure p ₂₁ corresponds to a pressure at the location downstream of the charge air cooler 14 and upstream of the throttle valve 19. The second relative humidity φ _21,mdl corresponds to a relative humidity at the location downstream of the charge air cooler 14 and upstream of the throttle valve 19.
Alternatively, the modeled relative humidity φ _21,mdl can be determined as a function of a fourth characteristic map K ₄ from the determined second saturation vapor pressure p _sat,21 , the second pressure p ₂₁ and the second specific humidity φ _20,mdl , with the modeled relative humidity φ _{21, mdl} is stored in the fourth map K ₄ as a function of these variables.

_Then , in _{step 220 ,} a first water mass _{m H20,LLK} determined. The first water mass m _H2O,LLK corresponds to a water mass stored in charge air cooler 14 .
Alternatively, the first mass of water m _H2O,LLK can be determined as a function of a fifth map K ₅ from the second temperature T ₂₁ , the second pressure p ₂₁ and the second specific humidity φ _20,mdl and the gas mass flow dm _LLK , with the first mass of water m _H2O,LLK is stored in the fifth map K ₅ as a function of these variables.

In step 225, a check is made as to whether the first water mass m _H2O,LLK determined in step 220 exceeds a predefinable maximum permissible water mass m _max . If the first mass of water m _H2O,LLK falls below the maximum permissible mass of water m _max , a water rate rat _H20,LLK is determined as a function of the first mass of water m _H20,LLK and the method is continued in step 230 . If the first water mass m _{H2O,LLK exceeds} the maximum permissible water mass m _max , the first water rate rat _H20,LLK is determined as a function of the maximum permissible water mass m _max and the method is continued in step 230 .

In step 230, a check is made as to whether an ignition angle efficiency η _ZW of the combustion of the internal combustion engine 25 is above an operating point-dependent, specifiable ignition angle efficiency η _ZW,Lim . Furthermore, a consumption-optimized gas temperature T _LLKDs,Eco is determined from a current engine operating point. The engine operating point _is determined by a sixth model M ₆ stored on control unit 100 as a function of engine speed n _eng and setpoint cylinder charge rl set . Alternatively, the consumption-optimized gas temperature T _LLKDs,Eco can be determined as a function of a sixth map K ₆ from the speed n _eng and the target cylinder charge rl _soll , with the consumption-optimized gas temperature T _{LLKDs,Eco being} stored in the sixth map K ₆ as a function of these variables .
If the ignition angle efficiency η _{ZW exceeds} the operating point-dependent predeterminable ignition angle efficiency η _ZW,Lim , the gas temperature supplied to the internal combustion engine 25 is increased as a strategy for avoiding condensation effects in the fresh air system 48 and the method is continued in step 240.
If the ignition angle efficiency η _ZW falls below the operating point-dependent predeterminable ignition angle efficiency η _ZW,Lim , the method is continued in step 235 and the low-pressure exhaust gas recirculation rate rat _LPEGR,Lim is reduced in order to avoid condensation effects in the fresh air system 48, preferably on the compressor 14 and to reduce condensation effects in the charge air cooler 15.

In step 235, the condensation avoidance strategy is deactivated in control unit 100 and the setpoint low-pressure exhaust gas recirculation rate rat _LPEGR,Soll is reduced. This is preferably carried out by limiting a setpoint low-pressure exhaust gas recirculation rate rat _LPEGR,Soll to a predefinable low-pressure exhaust gas recirculation rate rat _LPEGR,lim . Preferably, the target low-pressure exhaust gas Feedback rate rat _{LPEGR, Shall} be adjusted by fully closing the LP EGR valve 40. The limitation of the low-pressure exhaust gas recirculation rate is preferably carried out until the ignition angle efficiency η _ZW again exceeds the operating point-dependent predeterminable ignition angle efficiency η _ZW,Lim . The method is continued in step 255, with the consumption-optimized gas temperature T _LLKDs,Eco corresponding to a corrected gas temperature T _LLKDs,EcoCtl .

In step 240, a dew point temperature T _Dew,21 is determined from the second saturation vapor pressure p _Sat,21 determined in step 210 and the relative humidity φ _21,mdl modeled in step 215 as a function of a seventh model M ₇ stored on control unit 100.
Alternatively, the dew point temperature T _Dew,21 can be determined as a function of a seventh map K ₇ from the second saturation vapor pressure p _Sat,21 and the modeled relative humidity φ _21,mdl , with the dew point temperature T _Dew,21 in the seventh map K ₇ depending on this sizes is stored.

Then, in step 245, a temperature change ΔT _Kond is determined as a function of the water rate rat _H2O,LLK determined in step 225 and a difference Diff between the determined dew point temperature T _Dew,21 and a corrected consumption-optimized gas temperature T _LLKDs,EcoCtl . The corrected, consumption-optimized gas temperature T _LLKDs,EcoCtl can be given a predefinable value or zero in a first calculation grid. The water rate _H20,LLK and the difference Diff are preferably fed to an integrator I and the temperature change ΔT _Kond is determined via this.
If the determined water rate rat _{H2O,LLK falls below} a first predefinable threshold value S ₁ , then the temperature change ΔT _Kond is reduced by a predefinable temperature step, and more condensation is thus permitted. If the determined water rate rat _H2O,LLK exceeds the first threshold value S ₁ , then the temperature change ΔT _Kond is increased as a function of the definable temperature step in order to evaporate condensate. The gas temperature at the outlet of the charge air cooler 14 is adjusted primarily by specifying a coolant mass flow and/or a coolant temperature of the charge air cooler 14 .

The temperature change ΔT _cond determined in step 250 is then added to the consumption-optimized gas temperature T _LLKDs,Eco determined in step 230 and a corrected gas temperature T _{LLKDs,EcoCtl is obtained} .

In step 255 it is checked whether the corrected gas temperature T _LLKDs,EcoCtl ascertained in step 250 exceeds a predefinable maximum permissible gas temperature T _LLKDs,Lim . The predefinable maximum permissible gas temperature T _LLKDs,Lim corresponds to a gas temperature that would be necessary to carry out a full-load request at the current speed n _eng of internal combustion engine 25 . If the determined corrected gas temperature T _{LLKDs,EcoCtl exceeds} the maximum, the predeterminable maximum permissible gas temperature T _LLKDs,Lim , the corrected gas temperature T _{LLKDs,EcoCtl is} limited to the maximum permissible gas temperature T _LLKDs,Lim and output as the desired target gas temperature T _LLKDs,Soll . If the determined corrected gas temperature T _{LLKDs,EcoCtl falls below} the predefinable maximum permissible gas temperature T _LLKDs,Lim , the determined corrected gas temperature T _LLKDs,EcoCtl is output as the desired target gas temperature T _LLKDs,Soll .

In a step 260, the desired setpoint gas temperature T _{LLKDs,Soll is set} via a model on control unit 100 for the coolant device, preferably by controlling the coolant mass flow and/or by adjusting the coolant temperature.
The coolant temperature can B. be adjusted by controlling a fan, which flows through the radiator grille with air.
In an alternate embodiment, the adjustable louvers covering the grille can be opened or closed to further control coolant temperature.
The method then continues from the beginning in step 200 .

Claims (15)

Method for controlling a gas temperature (T _LLKDs,Soll ) at the outlet of a charge air cooler (14) of an internal combustion engine (25) with a low-pressure exhaust gas recirculation (41), wherein exhaust gas is returned via the low-pressure exhaust gas recirculation (41), wherein A consumption-optimized gas temperature (T _LLKDs,Eco ) at the outlet of the intercooler (14) for the internal combustion engine (25) is determined as a function of a current speed (n _eng ) and a current target cylinder charge (rl set ) _, with the Gas temperature (T _LLKDs,Soll ) at the outlet of the charge air cooler (14) is set via a control of the charge air cooler (14) in such a way that the consumption-optimized gas temperature (T _LLKDs,Eco ) is continuously increased until a current ignition angle efficiency (η _ZW ) reaches a specifiable Ignition angle efficiency (η _ZW,Lim ) falls below or a specifiable maximum allowable gas temperature (T _LLKDS,Lim ) is exceeded, where if the current ignition angle efficiency (η _ZW,Lim ) falls below the predefinable ignition angle efficiency (η _ZW,Lim ), exhaust gas is returned via the nozzle the pressure exhaust gas recirculation (41) is reduced and/or the gas temperature (T _LLKDs,Soll ) at the outlet of the intercooler (14) is reduced in order to avoid further moisture input through recirculated exhaust gas via the low-pressure exhaust gas recirculation (41). procedure after claim 1 , characterized in that the predeterminable maximum permissible gas temperature (T _LLKDs,Lim ) corresponds to a gas temperature for which a full load requirement of the internal combustion engine (25) at the current speed (n _eng ) is necessary. procedure after claim 1 , characterized in that the predeterminable ignition angle efficiency (η ZW _{, lim} ) is determined as a function of the current speed (n _eng ) and the current setpoint cylinder charge (rl set ) and if the current ignition angle efficiency (η _ZW ) exceeds the predeterminable ignition angle efficiency (η _{ZW ,lim} ) falls below the condensation avoidance strategy is deactivated. procedure after claim 3 , characterized in that the predeterminable ignition angle efficiency (η _ZW,Lim ) corresponds to an ignition angle efficiency at which a knock limit occurs, preferably an ignition angle efficiency before a knock limit is reached. procedure after claim 1 , characterized in that when the condensation avoidance strategy is active, the consumption-optimized gas temperature (T _LLKDs,Eco ) for the intercooler (14) is gradually increased by a control difference (ΔT _Kond ), the control difference (ΔT _Kond ) depending on a difference between a dew point temperature (T _Dew,21 ) and a corrected consumption-optimized gas temperature (T _LLKDs,EcoCtl ) and a water rate (rat _H2O,LLK ) for the intercooler (14). procedure after claim 5 , characterized in that the dew point temperature (T _Dew,21 ) is determined as a function of a second saturation vapor pressure (p _Sat,21 ) and a relative humidity (φ _21,mdl ) downstream of the intercooler (14) and upstream of the throttle valve (19). procedure after claim 5 , characterized in that the one water rate (rat _{H2O, LLK} ) for the intercooler (14) is determined as a function of a first water mass (m _{H2O, LLK} ). procedure after claim 7 , characterized in that the first water mass (m _H2O,LLK ) as a function of a gas mass flow (dm _LLK ) via the intercooler (14), a second pressure (p ₂₁ ) downstream of the intercooler (14) and upstream of the throttle valve (19) and a second specific humidity (φ _20,mdl ) is determined downstream of the compressor (12) and upstream of the intercooler (14). procedure after claim 7 , characterized in that the second specific humidity (φ _20,mdl ) as a function of a first specific humidity (φ ₁₂ ) at the location of a mixing point (43), a first pressure (p ₂₀ ) and a first saturation vapor pressure (p _Sat,20 ) is determined downstream of the compressor (12) and upstream of the charge air cooler (14). procedure after claim 5 , characterized in that the relative humidity (φ _21,mdl ) as a function of the second pressure (p ₂₁ ) and a second saturation vapor pressure (p _Sat,21 ) downstream of the compressor (12) and upstream of the charge air cooler (14) is determined . procedure after claim 9 that the first saturation vapor pressure (p _Sat,20 ) is determined as a function of a first temperature (T ₂₀ ) downstream of the compressor (12) and upstream of the intercooler (14). procedure after claim 10 that the second saturation vapor pressure (p _Sat,21 ) is determined as a function of a second temperature (T ₂₁ ) downstream of the intercooler (14) and upstream of the throttle valve (19). Computer program which is set up to carry out the method according to one of Claims 1 until 12 to perform. Electronic storage medium with a computer program Claim 13 . Device, in particular control unit, which is set up for the method claim 1 until 12 to execute.

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