DE102011013481A1 - Method for determining temperature of gas in combustion chamber of e.g. diesel engine, for passenger car, involves determining temperature of gas based on total mass and pressure in chamber, rotation speed of engine and volume of chamber - Google Patents

Abstract

The method involves determining total mass comprising a total gas mass (m) and mass of fuel in a combustion chamber (101) of an internal combustion engine at a predetermined crank angle. Pressure (p) in the combustion chamber is determined at the predetermined crank angle. Rotation speed of the combustion engine is determined. Volume (V) of the combustion chamber is determined at the predetermined crank angle. Temperature (T) of gas in the combustion chamber is determined at the predetermined crank angle based on the total mass, the pressure, the rotation speed and the volume. Independent claims are also included for the following: (1) a method for determining a gas temperature gradient in a combustion chamber of an internal combustion engine (2) a method for determining a residual gas amount in a combustion chamber of an internal combustion engine (3) a method for determining internal exhaust gas reconducting mass in a combustion chamber of an internal combustion engine (4) a method for determining total gas mass in a combustion chamber of an internal combustion engine (5) a method for determining oxygen contents during a compression stroke in a combustion chamber of an internal combustion engine (6) a method for controlling an internal combustion engine (7) a control device for an internal combustion engine.

Description

The present invention relates to a method for controlling an internal combustion engine and a corresponding control device for an internal combustion engine. More particularly, the present invention relates to a control of an internal combustion engine in which a real time gas state model is used for the internal combustion engine and the internal combustion engine is controlled on the basis of this gas state model. The method is particularly suitable for engines with an exhaust gas recirculation and a variable or fully variable valve timing of intake and exhaust valves of the internal combustion engine and an exhaust gas turbocharger.

In internal combustion engines of vehicles, such as passenger cars or trucks, there is a great need to reduce emissions and to reduce fuel consumption. Especially in diesel engines while methods for exhaust aftertreatment are used. The main focus of emission control is the emission of nitrogen oxide emissions and soot particles. In order to achieve this, for example, a combined reduction of nitrogen oxides and soot particles using a combustion process optimized in the engine map or an exhaust aftertreatment by diesel particulate filter and nitrogen oxide catalyst is used. Although exhaust aftertreatment systems offer the advantage of a high degree of conversion, but have the disadvantage of higher fuel consumption and high component and integration costs due to the systematic requirements. Therefore, the primary focus is on optimizing the combustion process. Manipulated variables for optimizing the combustion process are, on the one hand, fixed parameters such as combustion chamber, piston bowl and channel geometries as well as the geometric compression ratio. Furthermore, the nozzle geometry of the injector offers optimization potential. Variable parameters during engine operation are the gas state in the combustion chamber and the injection characteristic. The gas state in the combustion chamber is largely defined by the charge density, charge composition and charge movement and is controlled by means of gas system controllers. The injection characteristic is characterized by the amount of fuel to be injected, the number of injections per cycle, the injection timing and injection pressure and by the time course of the individual injections.

Controlling and control algorithms, which calculate, adjust and adjust the correct manipulated variable values for the current operating strategy on the basis of sensor and model variables, play a significant role in combustion optimization. In doing so, stationary and dynamic interventions are equally important. A significant proportion of pollutants is emitted during dynamic operating phases and in higher-load steady-state operation. The dynamic disadvantages arise on the one hand from quickly changing gas and combustion conditions in the engine and on the other hand from the inertness of the gas system path. This manifests itself for example in the response of an exhaust gas turbocharger. Another reason may be inadequate control and regulation strategies, which consist predominantly of influence regulations with map-based feedforward control. These are often developed for stationary operation and applied to emission-relevant cycle areas. Therefore, these control algorithms can only conditionally respond to rapid changes in the operating conditions of the engine.

In the prior art, a variety of proposals are known which determine and simulate an operation of an internal combustion engine as accurately as possible in order to carry out a suitable control, and moreover known methods for exhaust gas recirculation in order to solve the abovementioned emission, consumption and control problems ,

From the DE 10 2004 030 258 A1 In this context, an exhaust gas recirculation arrangement and a method for operating an exhaust gas recirculation arrangement are known. In the method, emission levels are improved when a regulator device is provided for adjusting an oxygen concentration in the intake manifold and a total gas mass in at least one cylinder of the internal combustion engine. For example, a broadband lambda probe can be used to determine the oxygen concentration. In the method, the oxygen concentration in the draft tube is determined and set and it is assumed that the oxygen concentration is equally present in the cylinder. However, this is not the case in particular in engines with a variable or fully variable valve control, since these engines may have internal exhaust gas recirculation which changes the oxygen content in the cylinder relative to the oxygen content in the intake manifold. An internal exhaust gas recirculation can be achieved by a corresponding control of the valves in a variable or fully variable valve timing and allow, for example, exhaust gas recirculation or exhaust gas pre-storage. During exhaust gas recirculation, exhaust gas is sucked back from the exhaust manifold in the intake stroke, and exhaust gas is stored in front of exhaust gas during an exhaust stroke into the intake manifold. In addition, the method does not take into account a residual amount of gas remaining in the cylinder.

The DE 4428976 C1 relates to an internal combustion engine with an exhaust gas recirculation opening into an intake tract. For a lean operation of the internal combustion engine it is provided that the intake tract is divided into two at least in the mouth region of the exhaust gas recirculation in the flow direction to form an exhaust passage and a fresh air passage, and that an oxygen sensor which is to allow control of the recirculated exhaust gas flow, between the exhaust passage and the fresh air passage is arranged so that it sets the oxygen levels in the exhaust duct and the fresh air duct to each other in a ratio. In addition to the additional costs for the oxygen sensor internal internal exhaust gas recirculation and a residual amount of exhaust gas in the cylinder is not considered in this internal combustion engine.

For an optimal combustion control, in particular a control of the gas state, in particular a cylinder-specific control of the gas state is required. However, the gas condition may depend on a variety of gas system actuators, such as turbocharger, throttle, swirl damper, high pressure exhaust gas recirculation adjustment, low pressure exhaust gas recirculation adjustment, and variable valve timing. For optimal gas adjustment therefore a real-time simulation and modeling of the gas system of the engine is required.

In this context, for example, from the DE 10 2006 061 936 A1 a method and system for simulating the operation of an internal combustion engine, from the DE 10 2008 032935 A1 a method for calculating a combustion chamber pressure in real time, from the WO 03/046356 A2 a method for determining the composition of the gas mixture in a combustion chamber of an internal combustion engine with exhaust gas recirculation and from DE 10 2004 044 814 A1 a method for simulating an internal combustion engine known. However, the methods do not consider an internal exhaust gas recirculation, which can be achieved with the aid of a fully variable valve operation and whose variabilities can be used as fast positioning possibilities for the dynamic and individual cylinder conditioning of the gas state in the combustion chamber.

Object of the present invention is therefore to provide a holistic concept for a cylinder-specific control of the gas state, in particular for diesel engines with the involvement of all gas system controller, which can be used real-time capable of combustion control, for example, requirements for the control of various diesel combustion processes to reduce emissions in stationary and transient To meet engine operation.

This object is according to the present invention by a method for determining a gas temperature in a combustion chamber of an internal combustion engine according to claim 1, a method for determining a gas temperature profile in a combustion chamber of an internal combustion engine according to claim 9, a method for determining a residual gas amount in a combustion chamber of an internal combustion engine after Claim 10, a method for determining an internal exhaust gas recirculation mass in a combustion chamber of an internal combustion engine according to claim 11, a method for determining a total amount of gas in a combustion chamber of an internal combustion engine according to claim 12, a method for determining an oxygen content during a compression stroke in a combustion chamber of an internal combustion engine according to claim 14, a method of controlling an internal combustion exhaust gas recirculation internal combustion engine according to claim 17, a control apparatus for an internal combustion engine according to claim 22, and a vehicle according to claim 23 öst. The dependent claims define preferred and advantageous embodiments of the invention.

In the context of the present invention, the terms "internal exhaust gas recirculation" and "external exhaust gas recirculation" as well as "residual gas increase" are used. The external exhaust gas recirculation comprises, for example, a high-pressure exhaust gas recirculation via which exhaust gases are returned from a region upstream of the turbine of the exhaust gas turbocharger into the intake region of the internal combustion engine, and a low-pressure exhaust gas recirculation, via which exhaust gases are returned from a region downstream of the exhaust gas turbocharger to the intake region of the internal combustion engine. The internal exhaust gas recirculation includes, for example, an exhaust pre-storage in which exhaust gas is exhausted into the intake portion (at least one of the intake ports) of the engine by opening an intake valve, and an exhaust gas recirculation in which exhaust gas is exhausted from the exhaust portion by opening an exhaust valve during the intake stroke is sucked into a combustion chamber. The residual gas increase includes, for example, the retention of exhaust gas during the Ausschiebetaktes by early closing one of the two exhaust valves of a combustion chamber of the internal combustion engine.

In the context of the present invention, crankshaft angles are defined during different strokes of the internal combustion engine, which is typically a four-stroke engine. In this context, a crankshaft angle between 0 ° and 180 ° refers to an intake stroke, a crankshaft angle between 180 ° and 360 ° to a compression stroke or compression stroke, a crankshaft angle between 360 ° and 540 ° Expansion stroke or power stroke and a crankshaft angle between 540 ° and 720 ° a Ausschiebetakt or exhaust stroke. After 720 ° the work cycle of the engine is repeated.

According to the present invention, a method for determining a gas temperature in a combustion chamber of an internal combustion engine is provided. In the method, a total mass in the combustion chamber is determined at a predetermined crankshaft angle. The total mass comprises a total mass of gas and a mass of an amount of fuel in the combustion chamber at the predetermined crankshaft angle. Furthermore, a combustion chamber pressure in the combustion chamber is determined at the predetermined crankshaft angle and the gas temperature in the combustion chamber is determined at the predetermined crankshaft angle as a function of the total mass and the combustion chamber pressure. The combustion chamber pressure can be determined individually for each cylinder with the aid of a pressure gauge. On the basis of the total mass, the combustion chamber pressure and the combustion chamber volume, which can be easily determined at the predetermined crankshaft angle from geometric considerations, the gas temperature in the combustion chamber can be determined in a simple manner in real time from the general equation of state for gases.

According to one embodiment, in addition to the combustion chamber volume of the combustion chamber at the predetermined crankshaft angle, the rotational speed of the internal combustion engine is determined and the gas temperature is determined as a function of the total gas mass, the combustion chamber pressure, the rotational speed and the combustion chamber volume. By determining the speed of the internal combustion engine, the gas constant of the general equation of state for gases speed and combustion chamber pressure can be determined depending on empirical models for the operating state of the internal combustion engine. As a result, the gas temperature can be determined in a simple real-time manner.

For example, a first gas temperature may be determined at a predetermined first crankshaft angle during a power stroke of the internal combustion engine. The first crankshaft angle is for example 460 °. According to one embodiment, a second gas temperature at a second crankshaft angle at the end of the power stroke is determined as a function of the rotational speed, the combustion chamber pressure at the first crankshaft angle, the first gas temperature and empirically determined parameters. The second crankshaft angle is for example 540 °. The empirical parameters can be determined, for example during a development of the internal combustion engine, such that only simple equations with the empirical parameters and the determined variables for determining the gas temperature have to be calculated in real-time operation. As a result, a real-time capable determination of the gas temperature is possible.

According to another embodiment, a third gas temperature at a third crankshaft angle of, for example, 640 ° or 700 ° during an exhaust stroke of the internal combustion engine depending on the rotational speed, a second combustion chamber pressure at the aforementioned second crankshaft angle, a third combustion chamber pressure at the third crankshaft angle and the previously determined certain second gas temperature. Thus, a gas temperature during the exhaust stroke of the internal combustion engine in real time can be determined.

According to another embodiment, a fourth gas temperature of a residual gas in the combustion chamber is determined at a fourth crankshaft angle of, for example, 720 ° at the end of the exhaust stroke in dependence on the rotational speed, the first combustion chamber pressure, the third combustion chamber pressure and the second gas temperature. Also the fourth gas temperature, d. H. the temperature of the residual gas, can be determined depending on the aforementioned parameters, so that a simple and real-time capable gas temperature determination of the residual gas is possible.

According to a further embodiment, an inlet gas temperature is determined at inlet valves of the internal combustion engine and an indexed moment of the internal combustion engine is determined. Depending on the rotational speed, the fourth gas temperature, the indicated torque and the intake gas temperature, a fifth gas temperature is determined at a fifth crankshaft angle of, for example, 100 ° during an intake stroke of the internal combustion engine. The fifth gas temperature can be determined from the aforementioned quantities with simple algebraic means and therefore this determination is real-time capable.

According to a further embodiment, a fifth combustion chamber pressure at the fifth crankshaft angle is determined with, for example, a pressure sensor and a sixth gas temperature at a sixth crankshaft angle at the end of the Ansautaktts, for example at 180 ° determined depending on the speed, the fifth gas temperature and the fifth combustion chamber pressure. The sixth gas temperature is also determined from the mentioned input variables and can therefore be determined in real time during operation of the internal combustion engine.

According to a further embodiment, a sixth combustion chamber pressure is determined at the sixth crankshaft angle, a combustion chamber volume at the sixth crankshaft angle is determined from, for example, geometric data of the internal combustion engine, a seventh combustion chamber pressure is determined at the seventh crankshaft angle and a combustion chamber volume is determined at the seventh crankshaft angle. The seventh crankshaft angle is a crankshaft angle during a compression stroke or working stroke of the internal combustion engine, in particular a crankshaft angle in the range of 180 ° to 460 °. Depending on the sixth gas temperature, the fifth combustion chamber pressure, the sixth combustion chamber pressure, the combustion chamber volume at the sixth crankshaft angle, the seventh combustion chamber pressure, the combustion chamber volume at the seventh crankshaft angle, the rotational speed and the indicated torque, a seventh gas temperature at the seventh crankshaft angle, for example, by means of the ideal gas equation. Thus, the seventh gas temperature can be determined in a simple manner in real time from the mentioned input variables.

With the methods described above, it is thus possible to determine the actual gas temperature in the combustion chamber of the internal combustion engine in real time for a large number of different crankshaft angles during the different cycles of the internal combustion engine.

Therefore, in the context of the present invention, a method is provided for determining a gas temperature profile in a combustion chamber of an internal combustion engine, wherein a plurality of gas temperatures in the combustion chamber of the internal combustion engine at different crankshaft angles of the internal combustion engine are determined and the gas temperature profile is determined over the crankshaft angle by interpolating the plurality of gas temperatures. The multiple gas temperatures are determined as previously described. With the aid of the method described above, it is therefore possible to determine a current gas temperature in the combustion chamber of the internal combustion engine without a temperature sensor in the combustion chamber.

According to the present invention, there is further provided a method of determining a total amount of gas in a combustion chamber of an internal combustion engine. In the method, a cylinder pressure characteristic is determined at the end of an intake stroke of the internal combustion engine. The cylinder pressure characteristic value at the end of an intake stroke can be determined, for example, by a cylinder pressure characteristic which is determined by means of a pressure sensor in the cylinder at a specific crankshaft angle during a compression stroke and calculated back to a cylinder pressure value for the end of the intake stroke with the aid of an isentropic state change. Furthermore, in the method at the end of the intake stroke, a combustion chamber volume of, for example, geometric data of the internal combustion engine and a cylinder temperature is determined. The total gas mass in the combustion chamber is determined as a function of the cylinder pressure characteristic at the end of the intake stroke, the combustion chamber volume at the end of the intake stroke and the cylinder temperature at the end of the intake stroke. The determination of the total gas mass can be determined, for example, with the aid of the equation of state for ideal gases in a simple manner from the parameters described above. This makes a determination in real time possible.

According to one embodiment, the gas temperature is determined by means of the above-described method for determining the sixth gas temperature at the end of the intake stroke at, for example, a crankshaft angle of 180 °. The cylinder temperature at the end of the intake stroke may then be determined as a function of the sixth gas temperature, the engine speed, the cylinder pressure characteristic at the end of the intake stroke, the fifth combustion chamber pressure and the indicated torque of the cylinder. The cylinder temperature may be determined, for example, by means of a correction term from the sixth gas temperature, i. H. the gas temperature at the end of the intake stroke. For example, the correction term may be determined from the engine speed, the cylinder pressure rating, the combustion chamber pressure, and the indicated torque using the ideal gas equation and a charge cycle analysis.

According to the present invention, there is further provided a method of determining an oxygen content during a compression stroke in a combustion chamber of an internal combustion engine with exhaust gas recirculation. For this purpose, an exhaust gas recirculation rate is determined via the exhaust gas recirculation and a fresh air oxygen content of a fresh air supplied to the internal combustion engine is determined. Furthermore, an exhaust gas oxygen content of an exhaust gas of the internal combustion engine, which is supplied to the internal combustion engine via the exhaust gas recirculation, determined and determines the oxygen content in the combustion chamber in dependence on the exhaust gas recirculation rate, the fresh air oxygen content and the exhaust gas oxygen content. The exhaust gas recirculation may include an internal and an external exhaust gas recirculation as well as a residual gas increase. The internal exhaust gas recirculation and the residual gas increase can be controlled by a variable valve control of valves be adjustable to the internal combustion engine. The exhaust gas recirculation rate in this case is determined on the basis of an external exhaust gas recirculation EGR rate, an internal exhaust gas recirculation EGR rate, and a residual gas rate of remaining gas remaining in the combustion chamber after combustion, if necessary increased. By taking into account the external exhaust gas recirculation, the internal exhaust gas recirculation and the amount of residual gas, the oxygen content in the combustion chamber of the internal combustion engine can be accurately determined in a simple manner, especially in engines with variable valve timing, especially in a fully variable valve timing. In addition, no complex sensors for determining the oxygen content, for example, in the intake tract required.

According to one embodiment, a real pressure ratio of two measured pressure characteristics during the compression stroke is determined at two different crankshaft angles. Furthermore, a modeled pressure ratio at the two different crankshaft angles is determined as a function of an engine speed, an indicated torque, a gas mass in the combustion chamber and the oxygen content in the combustion chamber. The modeled pressure ratio represents an expected pressure ratio, which is modeled empirically using stationary measurements using the aforementioned parameters. Depending on the engine speed, the indicated torque, the real pressure ratio and the modeled pressure ratio, a correction value for the oxygen content is determined. For example, the previously determined oxygen content in the corresponding engine operation is additively corrected with the aid of the difference between the modeled and the real pressure ratio. This correction is particularly necessary in transient operating phases of the internal combustion engine to determine an accurate oxygen content during the compression stroke.

According to a further aspect of the present invention, a method for controlling an internal combustion engine with an external and internal exhaust gas recirculation and with a residual gas increase is provided. The internal exhaust gas recirculation and residual gas increase are adjustable via a variable valve control, in particular a fully variable valve control. In the method, a total gas mass and an oxygen content in a combustion chamber of the internal combustion engine is determined. Furthermore, a target total gas mass and a desired oxygen content in the combustion chamber are determined for a given operating point of the internal combustion engine. Depending on the determined total gas mass, the determined oxygen content, the target total gas mass and the desired oxygen content, an actuator of the internal combustion engine is set.

The actuator of the internal combustion engine may comprise, for example, the variable or fully variable valve control, a throttle valve of the internal combustion engine, a turbocharger adjustment device for adjusting a turbocharging pressure of at least one exhaust gas turbocharger of the internal combustion engine. A setting of the valves, the throttle valve and / or the turbocharger adjustment device can be carried out in particular with the total gas mass as a reference variable. The setting of the valves may include, for example, adjusting an intake open timing, an intake close timing, and a valve lift amount.

According to a further embodiment, the actuator comprises the internal exhaust gas recirculation, the residual gas increase, an external exhaust gas recirculation of the internal combustion engine and / or a swirl flap of the internal combustion engine. The internal exhaust gas recirculation, the residual gas increase, the external exhaust gas recirculation and / or the swirl flap are adjusted as a function of the determined oxygen content and the desired oxygen content. The oxygen content serves as a reference variable in this case. By setting some of the internal combustion engine (throttle, variable valve control, turbocharger) with the total gas mass as a reference and adjusting other actuators (exhaust gas recirculation and swirl damper) with the oxygen content as a reference, easy and reliable control of the engine can be ensured.

According to a further embodiment, the internal combustion engine comprises a plurality of cylinders and the variable valve control is a fully variable cylinder-individual valve control. The total gas mass, the oxygen content, target total gas mass and the desired oxygen content are determined in this case cylinder-specific and the valves of the fully variable cylinder-individual valve control are set individually for each cylinder in dependence on the above variables. As a result, differences in the cylinder, for example different temperatures in the combustion chambers or a different charge cycle behavior (gas mass and oxygen content) of the individual cylinders, can be taken into account individually and thus emissions and fuel consumption can be optimized for each cylinder.

Furthermore, the present invention provides a control device for an internal combustion engine, which is designed to carry out one or more of the previously described methods and the embodiments described above. Furthermore, according to the present invention, there is provided a vehicle having an internal combustion engine and the control device. The control device and the vehicle therefore also include the advantages of the previously described methods and embodiments.

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

1 shows description variables for the gas state in the combustion chamber of an internal combustion engine.

2 FIG. 12 shows an overall structure of cylinder pressure based gas state modeling according to one embodiment of the present invention. FIG.

3 schematically shows an internal combustion engine system and a sensor configuration of the internal combustion engine system.

4 shows a structure of a combustion model according to an embodiment of the present invention.

5 shows a representation of a physical temperature profile and an approximated according to the present invention, temperature profile in the cylinder.

6 shows method steps for determining a gas temperature profile according to an embodiment of the present invention.

7 shows method steps for determining a gas temperature according to an embodiment of the present invention.

8th shows method steps for determining a total amount of gas in the combustion chamber according to an embodiment of the present invention.

9 shows a method for determining an oxygen content in the combustion chamber according to an embodiment of the present invention.

10 shows a control structure with a model-based Führungsgrößenvorsteuerung according to an embodiment of the present invention.

11 shows a schematic structure of a control structure for a gas system according to an embodiment of the present invention.

12 schematically shows a cylinder-specific gas mass control according to an embodiment of the present invention.

13 schematically shows a cylinder-specific regulation of an oxygen content according to an embodiment of the present invention.

14 shows an overall construction of a cylinder-individual control and regulation concept according to an embodiment of the present invention.

15 shows a method for adjusting an actuator of an internal combustion engine according to an embodiment of the present invention.

16 shows a vehicle with a control device according to an embodiment of the present invention.

According to the present invention, a gas state modeling is provided, which includes real-time capable and series-compatible model approaches for the control-relevant process variables cylinder gas mass and oxygen content in the combustion chamber. Here, a physically motivated modeling is in the foreground, ie the process variables are purely based on the changing process state calculated physical relationships and not in response to varying manipulated variable values, such as the setpoint for a valve position. The result is thus an integrated modeling approach for the gas state, which is independent of the engine design and of the operating strategy or the operating state. The main input variable for the overall model is the cylinder-specific measured cylinder pressure. In order to determine physical dependencies of the model sizes of measurable process variables, methods of correlation analysis can be applied. On the basis of experimental measurement results, the development of so-called gray-box models takes place, which include physical model structures in the base. These are supplemented by empirical parts in the form of non-linear equations with approximated parameters. In the following, the empirical model components are described by the general form z = a ₀ + a ₁ · x · y or z = b ₀ + b ₁ · x · y. The variables x and y describe the input variables. The size z is the starting size. The parameters a ₀ and a ₁ or b ₀ and b ₁ are applicational variables with different values and result from the model data.

The gas state in the combustion chamber is defined by its thermodynamic properties. 1 schematically shows a cylinder 100 an internal combustion engine with a combustion chamber 101 as a control volume in which the process variables pressure p, volume V, temperature T, gas mass m, oxygen content x _O2 prevail, taking into account an exhaust gas recirculation rate r _AGR and charge movement (as a spin level DN) at the beginning of the compression. As soon as the injection has taken place, these state variables are decisive for the subsequent mixture formation process (ignition delay time) and thus decisively influence the ignition conditions. Furthermore, the gas state is critical to the combustion process (heat release) and the resulting combustion efficiency.

The gas state in the combustion chamber is determined for all cylinders equal to the external gas state in the intake and exhaust system and by means of cylinder-specific parameters. Possible control elements in the gas system are, for example, the charge pressure of an exhaust gas turbocharger, a throttle valve, valves in a high-pressure exhaust gas recirculation (HP-EGR) or low-pressure exhaust gas recirculation (LP EGR), a swirl flap, a valve train of the intake and exhaust valves, an injection pressure or an injection time.

• – Es erfolgt nur die Modellierung konzeptrelevanter Prozessgrößen mit ausreichender Genauigkeit auf Basis von Mittelwertmodellen. Das systematische Verhalten des Gesamtmodells steht im Vordergrund und nicht die Genauigkeit der Teilmodelle.
• – Drosselstellen im System (Drosselklappe, Hochdruck- und Niederdruckabgasrückführventile, Abgasklappe) werden als ideale Drosseln betrachtet.
• – Das den Motor durchströmende Gas wird in seinen Eigenschaften als ideal betrachtet, d. h. es ist ideal durchmischt und besitzt in den Kontrollvolumina keine örtlichen Ungleichverteilungen.
• – Es erfolgt keine direkte Modellierung des Wandwärmeübergangsverhalten in den Kontrollvolumina. Es wird nur die Erwärmung der Gase infolge von Mischungsvorgängen betrachtet.

To make the overall model suitable for series production and real-time, some basic simplifications and assumptions can be made.

• - Only the modeling of process-relevant process variables with sufficient accuracy on the basis of mean value models takes place. The systematic behavior of the overall model is in the foreground and not the accuracy of the submodels.
• - Throttles in the system (throttle, high and low pressure EGR valves, exhaust flap) are considered ideal throttles.
• - The gas flowing through the engine is considered to be ideal in its properties, ie it is ideally mixed and has no local unequal distributions in the control volumes.
• - There is no direct modeling of the wall heat transfer behavior in the control volumes. Only the heating of the gases due to mixing operations is considered.

As references for the individual submodels, predominantly measured values of a sensor system of a test stand are used. These include, for example, a mass flow sensor of the test bench as well as pressures and temperatures at different positions in the gas system. In addition, the exhaust gas recirculation rate, the exhaust gas oxygen content and the air ratio in the exhaust gas are calculated from the exhaust gas analysis.

As references for the state variables in the cylinder, the result values of an offline charge change analysis and indexing evaluation are used. For this purpose, all performed stationary tests are recalculated and evaluated in their course with the help of a process analysis cycle- and cylinder-selective. The calculation results for the total gas mass, residual gas mass and temperature characteristics continue to serve as a target for the model data.

After completion of the assessment, the quality of the individual models is assessed on the basis of a correlation coefficient (KK). This is a measure of the degree of linearity of two sets of magnitude N. For a coefficient of 1 or -1, there is a complete linear relationship between the target and model values. For the value 0 there is no correlation. With the help of the correlation coefficient, the influence of superimposed nonlinear cross-influences on the relationship between model and target size can be quantified. The following equation 1 describes the calculation rule.

The detection of the gas state in the combustion chamber is based on the description of the physical processes in the combustion chamber and in the inlet and outlet tract. Therefore, the overall model is divided into two subsystems. The one system, the external gas system, is open and flows continuously. The second system, the combustion chamber, is connected to the external gas system during the charge exchange. In the high pressure phase, both systems are decoupled.

2 shows a forest 200 cylinder pressure based gas state modeling with an external gas system model 201 , a combustion model 202 , an engine control unit 203 and a cylinder pressure evaluation 204 , For the model-based detection of the total gas mass in the combustion chamber, the temperature is required in addition to the pressure and the volume. The total gas mass is thus not determinable exclusively from the combustion chamber pressure signal. Furthermore, the pressure signal contains no direct information about the amount of recirculated exhaust gas. It is therefore necessary to combine information of the external gas system with those of the combustion chamber.

p₀ = p₅, T₀

Umgebungsdruck, -temperatur

p₁, T₁

Ladedruck, -lufttemperatur

p₂, T₂

Saugrohrdruck, -temperatur

p₃, T₃

Abgasgegendruck, -temperatur vor Turbine

T'₃

Temperatur vor HD-AGR-Ventil

p₄, T₄

Abgasdruck, -temperatur nach Partikelfilter

T'₄

Temperatur vor ND-AGR-Ventil

P_Zyl1,2,3,4

Zylinderdrucksignale

T_Mot

Kühlwassertemperatur des Motors

n_Mot

Motordrehzahl

sowie die Referenzwerte:

ṁ_Ref

Referenzmassenstrom vom Prüfstand

λ_LSU

Serienbreitband-Lambdasonde

The real-time indexing system supplies cylinder-specific parameters of the processes to the combustion chamber model on the basis of the measured pressure signals. As additional information of the external gas system, in addition to the modeled high-pressure and low-pressure exhaust gas recirculation mass flows, sensor signals are also used at certain installation locations and read in via the engine control. 3 shows the sensor configuration used for the overall model of an engine system 300 , The engine system includes the actual internal combustion engine 301 with for example four cylinders, a fresh air supply 302 , an exhaust gas turbocharger 303 , an injection system 304 , a high-pressure exhaust gas recirculation (HD-EGR) 305 , a low pressure exhaust gas recirculation (ND-EGR) 306 , an air flow meter 307 , an exhaust flap 308 , an exhaust system 309 , a ND-EGR control valve 310 , a HD EGR control valve 311 and a throttle 312 , In the engine system 300 the following sizes are used:

p ₀ = p ₅ , T ₀

Ambient pressure, temperature

p ₁ , T ₁

Boost pressure, air temperature

p ₂ , T ₂

Intake manifold pressure, temperature

p ₃ , T ₃

Exhaust back pressure, temperature before turbine

T ' ₃

Temperature before HD EGR valve

p ₄ , T ₄

Exhaust gas pressure, temperature after particle filter

T ' ₄

Temperature before LP EGR valve

P _Cyl1,2,3,4

Cylinder pressure signals

T _Mot

Cooling water temperature of the engine

_Motive

Engine speed

as well as the reference values:

_Ref

Reference mass flow from the test bench

λ _LSU

Series broadband lambda probe

The equations for the external gas system are calculated synchronously. The combustion chamber model calculates the respective process variables for a cylinder in an angular synchronous manner and in firing order on the basis of the input variables.

4 shows a combustion chamber pressure model 400 in which various variables such as cylinder gas mass, residual gas mass, aspirated gas mass, internal exhaust gas recirculation mass and cylinder oxygen content from the cylinder pressure signal and from external process variables are determined synchronously in the working cycle. The model sizes are calculated individually for each cylinder and summarized to mean values for linking with the external gas system model (eg, aspirated gas mass flow).

All submodels are based on physical equations, some of which are augmented with empirical correction terms. On the one hand, this procedure makes it possible to map the physical relationships of the process variables very well. On the other hand, the addition of correction terms makes it possible to take account of systematic errors in cylinder pressure measurement and difficult to model operable influences (eg wall heat transfer behavior in the combustion chamber).

The calculation of the partial models requires special measurement data, which are determined directly on an engine test bench or with the help of an offline process analysis afterwards. The test programs mainly include stationary measurements with inlet and outlet valve variations as well as exhaust gas recirculation and boost pressure variations throughout the engine map. Furthermore, methods of statistical experimental design can be used to independently investigate the process influence of the respective variation.

Gas temperature in the combustion chamber

For the determination of the gas mass, residual gas mass and for the state of the internal exhaust gas recirculation mass, the course of the gas temperature in the cylinder is required in addition to the measurable pressure curve. Therefore, an approach for an empirical model for approximating the physical temperature history in the combustion chamber is provided. 6 clarifies the procedure. Based on various temperature characteristics in the charge cycle, which are calculated from cylinder pressure characteristics to specific crank angle positions (step 601 ), is approximated by a straight line connection and subsequent combination with a relative temperature curve in the high pressure phase of the overall physical temperature course over a working cycle (step 602 ). 5 shows the approximate temperature appr. T (φ) and a physically determined in a test facility temperature phys. T (φ) compared with a cycle of the internal combustion engine. Further shows 5 an opening course of an intake valve (EV) and an exhaust valve (AV) of the internal combustion engine over the working cycle. The individual calculation steps are shown below.

Temperatures in the expansion and Ausschiebetakt

m₄₆₀ = m_Zyl + m_Kr (3) R₄₆₀ = R_Gas·(a₁ + a₀·n_Mot·p₄₆₀) (4) The calculation of the temperature characteristics starts in the expansion phase. Corresponding process steps are in 7 shown. For the crank angle φ = 460 ° CA, the first temperature characteristic T _{460 is} determined using Equation 2 on the basis of the ideal gas equation (step 703 ). For this purpose, the total mass m _{460 is} determined (equation 3), which is the sum of the total gas mass m _Cyl (result of the last model calculation) and the injected fuel quantity m _Kr (step 701 ). The pressure value p ₄₆₀ is measured (step 702 ). Volume V ₄₆₀ is known from the geometry of the engine. Depending on the operating condition of the engine, the ideal gas constant R _Gas must be corrected (Equation 4).

m ₄₆₀ = m _Cyl + m _Kr (3) R ₄₆₀ = R _gas · (a ₁ + a ₀ · n _Mot · p ₄₆₀ ) (4)

On the basis of T _{460 it is} possible with the help of the cylinder pressure value at φ = 460 ° KW and equation 5 to empirically determine a following temperature characteristic value T ₅₄₀ at φ = 540 ° CA at the end of the expansion stroke. T ₅₄₀ = T ₄₆₀ * (a ₁ + a ₀ * n _Mot * p ₄₆₀ ) (5)

Based on T ₅₄₀ , using equation 6, a temperature characteristic T _{640 is} calculated at φ 640 ° CA during the exhaust stroke. The pressure ratio p ₆₄₀ / p ₅₄₀ is a measure of the temperature drop in the cylinder during Ausschiebens.

On the basis of T ₅₄₀ , a further temperature characteristic T ₇₀₀ at φ = 700 ° CA during the Ausschiebetaktes is calculated using Equation 7. The pressure ratio p ₇₀₀ / p ₅₄₀ is also a measure of the temperature drop in the cylinder during Ausschiebens.

By comparing the empirically modeled temperature characteristics with the physically calculated values of the offline process analysis, which serve as measured variables (reference variables), a model quality can be determined. After completion of the model, for example, a correlation coefficient (KK) of more than 95% was achieved. On the basis of the procedure shown, further temperature characteristics can be modeled and thus the approximated temperature profile further refined. Wall heat effects are not explicitly taken into account.

Residual gas temperature

The temperature of the residual gas mass T _RG at top dead center of the charge cycle is the result of the previous expansion and Ausschiebeprocessess. It can be calculated by changing the equation 14 with known residual gas mass (from the charge exchange analysis) and with the help of a residual gas characteristic value determined in the indexing system. The desired residual gas temperature is then estimated from Equation 8 on the basis of T ₅₄₀ and measured pressure characteristics in the exhaust stroke.

By comparing the modeled residual gas temperature with the desired temperature value, a corresponding model quality can be determined. For example, a model quality with a correlation coefficient of 98% can be achieved. It should be noted that the modeled temperature does not necessarily have to correspond to the real value at top dead center of the charge cycle.

Intake valve temperature

As a starting point for the modeling of the gas temperature in the intake phase in the cylinder, the temperature of the incoming gas must be known. For this purpose, the temperatures at the intake valves are determined on the basis of the intake pipe temperature T ₂ on the basis of a mean value model. From a physical point of view, the incoming gas in the inlet ducts experiences a temperature increase due to the convective heat transfer from the engine head to the gas. Based on the Newtonian heat transfer, the temperature of the gas at the intake valves T _{EV is} calculated according to equation 9. In this case, for T _Mot the cooling water temperature of the engine is assumed.

Upon activation of an internal exhaust gas recirculation functionality (exhaust pre-storage, ie the opening of one or more intake valves during the Ausschiebetaktes) increases due to the hot exhaust gas constituents the temperature in the inlet channels. Equation 10 takes this relationship into account and corrects the base valve temperature as a function of the modeled exhaust gas recirculation _temperature T _iAGR and the modeled internal exhaust gas recirculation _mass m _iAGRVL by means of an additional temperature _component . The heating and cooling process is approximated by a proportional element with delay I. order (PT ₁ element). The time constant T ₁ is stored speed-dependent and mass-dependent. The gain K _p is equated with one.

Temperatures in the intake stroke

During the intake phase, the incoming gas mixes with the residual gas remaining in the cylinder. Thereby a temperature compensation takes place. The incoming gas cools the combustion chamber. Based on the modeled intake _{gas temperature} T _EViAGR and the residual gas _temperature T _RG , the mixed intake temperature T ₁₀₀ in the cylinder can then be calculated at φ = 100 ° CA (Equation 11). In this case, an ideal cooling process of the cylinder is assumed by the incoming gas. In addition, upon activation of a second internal exhaust gas recirculation functionality (exhaust gas recirculation, ie, opening of one or more exhaust valves during the intake stroke), there is an increase in the in-cylinder temperature due to the mixing of the intake gas with the hot exhaust gas constituents of the exhaust mass being drawn. With the known temperature in the exhaust manifold T ₃ , this effect can be corrected. Measurements and recalculations provided the result that at this point the mixing of the individual gas mass fractions takes place very quickly. Thus, a time mapping of the temperature mixing process can be dispensed with.

The temperature characteristic in the bottom dead center of the intake stroke T _{180 is} calculated on the basis of the state at φ = 100 ° CA according to equation 12 and describes predominantly the heating of the gas mass at the cylinder walls.

Temperature profile in the compression, combustion and expansion phase

During the high-pressure phase, a simplified temperature curve can be calculated from the knowledge of the temperature at a reference time using the ideal gas equation and assuming m _Cyl = const. And R _Gas = const. As the reference time, the temperature at the bottom dead center of the intake stroke T _{180 is} selected here. The calculation rule is given in Equation 13.

On the basis of this procedure, further temperature characteristics can be determined and used for process monitoring. B. the maximum combustion chamber temperature.

Residual gas mass in the combustion chamber

The residual gas mass is defined as the minimum gas mass in the combustion chamber during a work cycle. It is the residual gas remaining in the cylinder from previous work cycles after the exhaust valves have been closed. The value of the residual gas mass is dependent on the one hand on the ratio of exhaust back pressure and intake manifold pressure and the second of the choice of the timing for the intake and exhaust valves.

A simple and production-ready method is the calculation of the residual gas mass m _RG based on the ideal gas equation at top dead center of the charge cycle. Here, the assumption is made that at this time the cylinder is completely closed. For the standard diesel engine, this assumption is permissible due to the geometry of the combustion chamber with a very small valve overlap. The measurable at this time in the compression volume V _{RG RG} pressure p is therefore directly proportional to the residual gas mass. For full application of the gas equation, the residual gas temperature T _{RG is} required. This is replaced in a first approximation with the value of the exhaust gas temperature or calculated as in this case cylinder-specific during the previous Ausschiebetaktes. The gas constant R _RG is the value of the exhaust gas constant R ₄₆₀ . The calculation is made according to Equation 14.

Internal exhaust gas recirculation mass

The realization of an internal exhaust gas recirculation is made possible by exhaust gas recirculation and exhaust gas pre-storage. The shifted gas mass fractions are difficult to detect by measurement and therefore have to be modeled indirectly from the available sensor signals.

One approach is to calculate the average internal EGR mass m _IAGR based on the average total cylinder mass m _Zyl , the average residual gas mass m _{RG ,} the mean injected fuel mass m _Kr and the measured oxygen content in the exhaust gas x _O2Abg (Equation 15). For this purpose, a stationary test program can be put together in which in the first step, the lift curves of both inlet valves during the Ausschiebetaktes be varied in different operating points equal. In the second step, the same operating points are driven with an equal variation of the lift curves of both exhaust valves during the intake stroke. For the calculation approach to be valid, the same target stroke curve must be set for a valve type on all cylinders. This results in a mean value of the internal mass for preloads and back suction for all cylinders. These values can serve as target variables for a cylinder-specific modeling.

In order to use the internal exhaust gas recirculation cylinder-individually and in a pilot control, a further approach is required. This approach relies on the computation equation for mass flow at orifices, as will be described later in connection with Equation 42. The difference is that the processes exhaust gas recirculation and pre-aging have no continuous behavior and the effective cross-sectional area is variable depending on the lift curve. The mass flow to be calculated thus changes per crank angle position constantly. Therefore, based on the currently prevailing gas state and the integral of the valve lift curve, a direct exhaust gas recirculation mass value is calculated here. There is thus a linking of process variables with geometric parameters. The equations 16 and 17 show the calculation _rules for the cylinder- _specific internal exhaust gas recirculation _{mass by pre-storage} m _iAGRVL and suck back m _iAGRRS . The target size for the model condition is the average return mass calculated for each type, using equation 15.

beschreiben den aktuellen Zustand der Gasmasse. In Abhängigkeit von den zylinderinneren und -äußeren Zuständen,

wird das Strömungsverhalten der Gasmassen berücksichtigt und korrigiert. In beiden Gleichungen werden Mess- und Modellgrößen miteinander verknüpft. In both equations, the area integral becomes below the lift curve A _iAGR = Δφ · h approximated by the product of opening duration and maximum lifting height. The terms

describe the current state of the gas mass. Depending on the inside and outside of the cylinder,

the flow behavior of the gas masses is considered and corrected. Both equations link measurement and model quantities.

Total gas mass in the combustion chamber

Calculation of m _cyl

During compression, the combustion chamber is considered to be an ideally closed container. The thermodynamic state of the gas mixture can thus be described using the equation of state 18 for ideal gases. p · V = m · R · T (18)

After the ideal gas equation has been changed, the total gas mass m _Cyl can be calculated after the gas _exchange according to Equation 19. 8th shows process steps for calculating the total gas mass m _cyl . In it, p _{Zyl is} a pressure characteristic determined in the compression phase (step 801 ). The cylinder _volume V _Zyl is calculated from the geometric data of the engine (step 802 ). The specific gas constant of the cylinder _gas R _Gas = 287 J / (kgK) is assumed to be constant. The fourth unknown in the gas equation is the cylinder temperature T _Zyl taken from the gas temperature model (step 803 ). Finally, the total gas mass m _{Cyl is} calculated according to equation 19 (step 804 ).

As a reference time for the calculation of the gas mass, the inlet closing time can be selected. In the present case, the reference time is set to the bottom dead center before the start of compression at φ = 180 ° CA. Thus, the calculation approach is independent of varying intake-close timing as would be expected with a fully variable valve train.

Calculation of p _cyl

During the compression phase in internal combustion engines, the thermodynamic behavior of the gas mixture can be approximated by a polytropic change of state.

With equation 20, the pressure component p _Zyl can be determined for the calculation of the total gas mass with the aid of a substitute pressure value. The substitute pressure value is determined in the compression phase at a point in time at which the correlation between the substitute pressure value and the cylinder gas mass assumes as high an amount as possible. For this purpose, the correlation coefficient between cylinder pressure value and charge mass is plotted above the crank angle and the largest correlation is determined. The evaluations show that the cylinder pressure value at φ = 340 ° KW is optimally suitable as a substitute pressure value. The polytropic state change for p _{Zyl is} calculated according to Equation 21 assuming a constant gas mass and using a constant polytropic exponent of n = 1.37.

It should be noted that the calculated pressure value p _Zyl for φ = 180 ° CA does not have to correspond to a measured pressure value at this point.

Calculation of T _cyl

To solve the ideal gas equation further, the temperature _component T _{Zyl is} required. The basis for this is formed by the equation 12 calculated temperature characteristic at φ = 180 ° KW. In order to adapt this temperature to the cylinder gas mass sought, it is necessary to correct the base value. The determination of the correction terms is carried out by changing the equation 19 according to the sought temperature. The reference gas mass value used here is the gas mass determined from the offline gas exchange analysis in the compression phase. Equation 22 describes the correction calculation.

With the help of the adaptation of T ₁₈₀ to the sought-after temperature component T _Zyl further non-linear heating effects during the suction phase, the heat transfer behavior during the compression phase as well as the simplifications are compensated by the ideal polytropic state change with n = 1.37 for the pressure calculation.

From the total cylinder mass m _cyl , the intake gas mass m _A can be calculated by subtracting the residual gas mass m _RG . m _A = m _Zyl - m _RG (23)

With knowledge of the engine speed n _Mot , the total engine mass flow ṁ _Cyl and the intake mass flow ṁ _{Ein are} calculated as follows for a four-cylinder four-stroke engine. ṁ _Cyl = 2 · 60 · n _Mot · m _Cyl · 10 ^-6 (24) ṁ _On = 2 × 60 × n _Mot × m _On × 10 ^-6 (25)

When the internal and external exhaust gas recirculation in stationary operation is switched off, the mean value of the cylinder-individually drawn-in mass flow is comparable to the reference mass flow sensor in the intake path of the engine. (Blow-by is neglected.)

Oxygen content in the combustion chamber

Oxygen content x _O2Zyl during compression

The combustion-relevant oxygen content in the combustion chamber after completion of the charge cycle is the result of the mixing of fresh air with recirculated exhaust gas and existing residual gas content. To determine the total oxygen content in each cylinder, according to equation 26, all mass fractions are balanced with their oxygen content and related to the total cylinder mass. 9 describes corresponding method steps.

In this equation, a value of 20.94% by volume is assumed for the fresh air _{oxygen content} x _O2FL (step 902 ). The fresh air mass m _FL is calculated from the difference between total cylinder mass and total exhaust gas recirculation mass. The sum of the externally recirculated exhaust _masses m _eAGRHD + m _eAGRND are calculated using the mean value for the exhaust gas oxygen content x _O2ZylAbg calculated from all cylinders. The residual oxygen content of the internal feedback m _iAGRVL + M _iAGRRS and of the residual gas m _RG is determined by the cylinder-specific exhaust gas oxygen content x _O2ZylAbg (step 903 ).

The individual mass constituents can be linked to other parameters describing the gas system, ignoring external gas run times (equations 27-31).

With Equation 30, an exhaust gas recirculation rate may be determined (step 901 ). By substituting these equations in equation 26, the total oxygen content in the combustion chamber can be calculated individually for each cylinder according to equation 32 (step 904 ).

By equating the exhaust gas oxygen contents, the following simple description (equation 33) results for the oxygen content in the combustion chamber, which can be used for a precontrol of the exhaust gas recirculation rate. x _O2Zyl = x _O2FL - r _AGRZyl · (x _O2FL - x _O2ZylAbg ) (33)

Cylinder _individual correction of x _O2Zyl

Due to expected model inaccuracies for the internal exhaust gas recirculation masses as well as the generally existing cylinder inequality distribution of recirculated exhaust gas (especially in dynamic operating phases), it is necessary to re-correct the basic oxygen content, calculated according to equation 32 individually for each cylinder. This correction is made using a cylinder pressure based approach. For this purpose, two cylinder pressure values are determined in the compression phase and, in analogy to equation 21, converted into two pressure characteristic values for bottom dead center (UT) (equation 34). The ratio Π of these calculated characteristics correlates with the total oxygen content x _O2Zyl in the cylinder.

The calculated pressure ratio Π from two measured pressure values in the compression is a measure of the deviation of the compression curve from the ideal curve, assuming a constant polytropic component n.

Using stationary measurements, the expected pressure ratio Π _{mod is} modeled according to the empirical equation 35. Π _mod = f (n _Mot , M _ind , m _Zyl , x _O2Zyl ) (35)

With the help of the difference between modeled and measured pressure ratio, the basic oxygen content is corrected additively according to equation 36 in engine operation. The correction is required especially in transient operating phases. x = x _{O2ZylKorr O2Zyl} + DELTA x = _{O2Zyl O2Zyl} + (a ₀ + a ₁ · n · M _{Mot ind)} · _(mod Π - Π) (36)

Oxygen content x _O2ZylAbg after combustion

The cylinder- _specific residual oxygen content of the combustion gas x _O2ZylAbg can be calculated from the previously calculated oxygen content in the compression assuming complete combustion (λ = 1) according to equation 37. In this case, the stoichiometric ratio of oxygen O is _st = 3.045 used.

Exhaust air ratio λ _Abg

λ_Abg = f(x _O2ZylAbg)|_PST/LSU (39) On the basis of the modeled mean exhaust gas oxygen content simplified (Equation 38) or, more precisely by means of a characteristic curve (Equation 39), the exhaust air ratio λ _Abg be determined. The characteristic curve can be optionally calibrated to the test bench measured value λ _PST from the exhaust gas analysis or to the lambda probe λ _LSU value.

λ _exhaust = f ( x _O2ZylAbg ) | _{PST / LSU} (39)

From the previously calculated model sizes, further gas system characteristics can be determined. Equation 40 detects the oxygen ratio λ _O2 in the combustion chamber and Equation 41 the global air ratio λ _L.

Modeling sizes of the external gas system

As in 3 As shown, the state of the gas in the cylinder-peripheral engine system is detected via a series of pressure and temperature sensors. The last two unknown quantities are the exhaust gas recirculation mass flows through high and low pressure valves. These are particularly important for the modeling of the external exhaust gas recirculation rate and furthermore crucial for the accuracy of the cylinder oxygen calculation.

The basis for the model-based calculation of the mass flows through the valves or valves is the continuity equation for friction-free flows at throttle points (equation 42).

This states that the mass flow rate remains constant due to a changing flow cross-section due to the changing density. By opening and closing the valves or flaps, the cross-sectional area of the exhaust gas recirculation lines is changed. It turns due to the surrounding state, a new mass flow.

With the help of the Bernoulli equation for compressible fluids and assuming a stationary, frictionless and isentropic flow for ideal gases, the flow rate can be calculated. By substituting this quantity in Equation 42 and using the ideal gas equation, the mass flow through a restriction is calculated as follows (Equation 43).

ist lediglich vom Druckverhältnis

über dem Drosselelement und vom Isentropenexponent

abhängig. Der maximale Durchflusswert wird im sogenannten kritischen Druckverhältnis

erreicht. In diesem Zustand strömt das ideale Gas mit Schallgeschwindigkeit durch die Drosselstelle. Unterhalb des kritischen Druckverhältnisses ist keine weitere Geschwindigkeitserhöhung möglich. Folglich muss in einen unter- und überkritischen Fall unterschieden werden (Gleichung 45).The flow function Ψ according to de Saint-Venant (equation 44) contained in this throttling equation

is only the pressure ratio

above the throttle element and the isentropic exponent

dependent. The maximum flow value is in the so-called critical pressure ratio

reached. In this condition, the ideal gas flows at the speed of sound through the throttle. Below the critical pressure ratio no further speed increase is possible. Consequently, a distinction must be made between a subcritical and supercritical case (Equation 45).

To reduce the computational effort in the control unit, the following approximation equation 46 can be used for the calculation of the flow function in the subcritical pressure range.

The remaining unknown quantity in the throttling equation is the effective cross-sectional area A _eff . This can be determined as a function of the triggering duty cycle of the actuator under consideration. The relationship between the Ansteuertastverhältnis the respective valve or flap actuator and associated effective cross-sectional area can be determined directly on the test carrier. For this purpose, the duty cycle is varied in an operating point of the engine over the entire adjustment range. Subsequently, the effective cross-sectional area is calculated using the throttle equation using the reference mass flow and the measured gas state before and after the actuator. The reference mass flow can be determined from the test mass mass flow value or else with a method for determining the exhaust gas recirculation mass flow from oxygen concentrations, depending on the condition of the control. The calculated relationship between duty ratio and effective cross-sectional area is then stored in a characteristic curve as a valve or flap characteristic.

The throttle model just presented can subsequently be used as an inverted model for the precontrol of the effective valve and flap surfaces of the actuators of the external gas system. It thus forms a central element of the external gas system control.

High-pressure exhaust gas recirculation mass flow

Equations 43 and 46 and the corresponding sensor values can be _{used to} calculate the high pressure _{exhaust gas recirculation mass flow} ṁ _eAGRHD (equations 47-50). The effective valve cross-sectional area A _effHD is determined from a characteristic as a function of the TV _HD valve position. Due to the short gas lines and the thermal state of the high-pressure exhaust gas recirculation, it is assumed that the modeled mass flow is supplied to the cylinders without time delays.

Low-pressure exhaust gas recirculation mass flow

Δp_HD = p₄ – p₀ (54) The low pressure _{exhaust gas recirculation mass flow} ṁ _eAGRND is calculated analogously according to equation 47 (equations 51-54).

Δp _HD = p ₄ - p ₀ (54)

External exhaust gas recirculation rate

welche sich aus dem sich ändernden Zustand vor der Drosselklappe und dem zugehörigen Massenstrom ergibt. Dieser Massenstrom entspricht dem Massenstrom durch die Drosselklappe als Differenz aus zylinderangesaugtem und Hochdruckabgasrückführmassenstrom ṁ_DK = ṁ_Ein – ṁ_eAGRHD. Die in das Saugrohr einströmende Niederdruckrate r_NDSgr berechnet sich dann nach Gleichung 55.The external exhaust gas recirculation rate is the sum of the exhaust gas components in the intake manifold which are recirculated in the high and low pressure branches. Due to the system, the current low-pressure exhaust gas recirculation mass flow at the valve does not correspond to the current proportion in the intake manifold. Especially in dynamic operating phases of the engine, this disadvantage comes into play. The reason for this is the storage behavior of the intake section and the running time the gas due to the longer distance. For the modeling of these effects, the gas path of the low-pressure exhaust gas recirculation from the inlet point upstream of the compressor to the throttle valve is understood as a container of constant volume with inflowing and outflowing mass flows and with connecting pipe sections of constant cross section. The aperiodic transport process of the low-pressure exhaust gas recirculation can then be approximately modeled by a dead-time proportional element with delay I. order (PT ₁ T _t element). Both elements have the same variable time constant at constant gain K.

which results from the changing state before the throttle and the associated mass flow. This mass flow corresponds to the mass flow through the throttle valve as the difference between cylinder- _suctioned and high-pressure _{exhaust gas recirculation} mass flow ṁ _DK = ṁ _Ein - ṁ _eAGRHD . The low pressure rate r _NDSgr flowing into the suction pipe is then calculated according to equation 55.

Neglecting the low storage effect of the intake manifold, the external exhaust gas recirculation rate in the intake manifold can be calculated. The sum of low-pressure ṁ _DKrNDSgr and high-pressure _return ṁ _eAGRHD is based on the total cylinder _{gas mass flow} ṁ _cyl .

This dynamically determined rate r _eAGRdyn (equation 56) replaces the steady-state calculated external rate r _eAGR in equation 32 for calculating the cylinder _{oxygen content} .

Strategy for controlling the gas state in the cylinder

A cylinder-specific regulation of the gas state is described below, which is an essential component of a cylinder-pressure-based engine management concept for diesel combustion processes. The main parameters for controlling the gas state in the combustion chamber are the total gas mass and the oxygen content. The gas system control integrates a large number of gas system actuators and functionalities of a fully variable valve train. The nonlinear system behavior of the engine gas system in stationary operating points is considered as linear and controlled by methods of linear control and control theory.

In terms of control engineering, the gas system of modern diesel engines is a multi-size system. In the following, this control task will be solved by the use of inference controls with model-based reference variable feedforward control and coordinated setpoint specification for the individual partial control loops. In part, a summary of cascaded structures is given. 10 illustrates the basic principle, where w is a command / setpoint, y is a control / actual size, e is a control deviation, u _{V is} a pilot control output, u _{R is} a controller output and u is a manipulated variable.

The precontrol takes the place of an inverted system model to improve the dynamic control behavior of the control loop. The purpose of the parallel-connected linear controller element (PI or PID element) is to correct pilot uncertainties and disturbances on the basis of sensor or model values to ensure stable control behavior. In this servo control, the command variable is continuously given from a higher level and is not known in advance. In dynamic engine operation, this results in a constant change of operating point with changed setpoint values. In addition, setting parameters are provided in the functions with which the behavior of the closed loop can be adjusted if necessary (application).

Against the background of the limited computing capacity in the control unit and the requirement for production-ready approaches, this approach makes sense. The fundamental requirement is fulfilled to build stable and robust controls with good dynamic and interference-free properties.

Structure of the regulatory structure

The gas state is inherently dependent on the out-of-cylinder condition in the intake manifold and in the supply lines. Compared to the injection path, the gas path has a rather sluggish adjustment behavior. Due to this, additional novel actuators, in particular a fully variable valve controller, are used in the gas path. Requirements for a precise adjustment of the gas state in the combustion chamber at the beginning of compression can be met with the standard structure for the gas system control, as implemented in conventional engine control units, or only with great effort. This means that the serial structure is not defined by the effect of the actuator. Thus, the physical relationships of the overall system can be mapped only to a small extent. The mass air flow mass variable can be adjusted for example via the high pressure, Niederdruckbabgasrückführung or even indirectly via a boost pressure change with the exhaust gas turbocharger. All three variants have different effects on engine dynamics and emissions. Another problem of the conventional structure is the incorporation of new actuators in increasingly separate modes. This means that the actuators are only used under certain conditions in the respective operating mode. This is caused by a structure change, in which both the controlled variables and their assignment to the individual actuators change. Inevitably, such transitions lead to difficult-to-describe states in the overall system. This results in an increased complexity of the motor control, which does not have the necessary flexibility for the integration of new positioning options.

Therefore, a holistic approach to the structure of the gas system regulation is proposed below. The physically motivated assignment of setpoints to the individual actuators forms an important principle according to their function. Due to the fixed allocation, structural changes of the control structure should be prevented. Another important feature is the modular design of the concept, which allows increased variability in the integration of new control elements regardless of their design and type.

The gas system control described below predominantly uses the gas-chamber-external gas system controllers as the basis for the integration of the cylinder-specific adjustment possibilities of a fully variable valve train. Furthermore, there is an extension with respect to the main parameters of the structure. Instead of the exhaust gas recirculation rate, the oxygen content in the combustion chamber is used.

In connection with 11 the gas system regulation will be explained in more detail. Essential features are the hierarchical structure with grouped function groups as well as the clear and fixed assignment of the setpoints. Depending on the design, passive or active, each structural element consists of a pre-control model and a control function.

At the lowest level all control elements in the engine gas system are shown. The intake and exhaust valves of a cylinder (i) are represented as representative of the 16 gas exchange valves. Each actuator is assigned only a fixed target size. For the two external exhaust gas recirculation valves, these are, for example, the associated mass flows. With the help of the exhaust flap, a required pressure drop over the low pressure valve should be set. Therefore, the required differential pressure is used as the setpoint for this controller. Throttle and turbocharger should realize according to their position in the gas path predetermined pressure levels. Valve opening and valve closing angles as well as the maximum lifting height are important parameters for the stroke curve generation of the gas exchange valves.

The individual control elements are combined to the next higher level to function groups. Low-pressure valve and exhaust flap form a common group for setting a required low-pressure exhaust gas recirculation rate. In the associated coordinator, the required setpoint values for mass flow and pressure ratio are calculated according to the rate setpoint. Similarly, the other actuators also generalize towards higher levels. Because of its central importance for the integration of a fully variable valve train the coordinator for the gas exchange valves will be described in more detail below.

Due to the manifold system freedoms for influencing the gas state, corresponding coordinators are required at higher hierarchical levels. An example is the breakdown of the gas mass adjustment to turbocharger, throttle and valve train. In parallel, the division of the Total exhaust gas recirculation rate from a required oxygen content in the combustion chamber to be coordinated to internal, external and residual gas rate. Further, the division takes place to external high pressure and low pressure rate or to an internal Vorlagern- and Rücksauganteil. The coordinators offer another advantage. In dynamic engine operation, intervention in the distribution functions can be carried out in a simple manner. For example, the running time of the low-pressure exhaust gas recirculation can be compensated by a short-term increased external high-pressure exhaust gas recirculation or cylinder-specific internal gas recirculation. In this way, the desired oxygen content would continue to be set in the combustion chamber.

The highest hierarchical level is the Cylinder Coordinator. Its central task is the setpoint generation of the two main parameters for the gas system: gas mass and oxygen content in the combustion chamber. Both quantities are divided into two parallel gas state paths corresponding to those in 11 introduced control structure set. The Cylinder Coordinator is not dedicated to the gas system, but similarly coordinates the injection system.

Below, the two gas state paths are described separately in their operation.

Cylinder-specific adjustment of the gas mass in the combustion chamber

Using the existing gas system controller, the gas mass in the combustion chamber is adjusted starting from the boost pressure. Accordingly, the setpoint values for the intake manifold and boost pressure are derived from the target value for the gas mass. Further, the calculation of the pilot value for the intake-closing angle of the valve train is made from the specification for the effective compression ratio. When lowering the compression ratio, for example by early or late closing of the intake valves, it is necessary to keep the predetermined gas mass in the combustion chamber to keep the thermodynamic efficiency of the combustion constant. This can only be realized by a correspondingly higher suction pipe or boost pressure, whose increase is inversely proportional to the effective compression ratio. In other words, this method corresponds to a kind of bias of the system, since a pressure reserve is built up over the valve train. At the same time, a further pressure reserve above the throttle valve can be set by specifying a higher charge pressure setpoint. Both pressure reserves can be released in transient operating phases in the short term and thus be used for a rapid increase in the gas charge mass, which in turn has an improved response of the turbocharger result.

With appropriate distribution of the total pressure reserve on the valve train and throttle valve, this approach is also consumable neutral. It makes sense here a combined use of both control elements. For stationary operation, the greater proportion of the pressure reserve should be assigned to the valve train, since the variation of the effective compression in contrast to throttling generates no increased losses and thus does not cause unwanted fuel consumption. In the dynamics, however, it can be quite helpful to use the throttle as a supporting element reinforced. In addition to the valve train can then be compensated with the help of this strategy, a possible overshoot of the boost pressure setpoint using the throttle to continue to set the required intake manifold pressure. Due to the objective of using the generated pressure reserve in dynamic engine operation specifically for a better response, the total inventory and its distribution to the gas system controller are only controlled and not specified as direct control variables. Instead, the control of the absolute state variables takes place after the actuators.

12 shows the functional diagram of the cylinder-specific gas mass control. The structure consists of a model-based reference variable feedforward control and a control component. The purpose of the feedforward control is to calculate the setpoint values for intake manifold and boost pressure as well as the pre-control value for the inlet-closing angle on the basis of the reference variables gas mass and effective compression or pressure reserve. Intake manifold and charge pressure setpoints are forwarded to the appropriate control circuits. Parallel to the cylinder-identical pilot control branch, a cylinder-individual control loop is connected, which adjusts additively the pilot-controlled inlet closing angle via a control component as a function of the control deviation. In the following, the functional blocks are described in more detail. Target quantities are generally identified with a star symbol (*) here.

Function block "feedforward control"

The pilot control for setting a required gas mass in the combustion chamber unites the coordinators "cylinder gas mass" and "intake manifold state" 11 , The following equations are used to calculate the specified setpoints. In this case, the desired suction pipe density ρ * _{Sgr is} determined as an intermediate variable. In this case, the actual thermal state between cylinder and channel T ₁₀₀ / T _EV As well as operating point-dependent _Channel losses C _channel taken into account. Thus, even at zero pres- sure, the target intake manifold density is always slightly different than the in-cylinder density at the inlet-close timing. The quantities π * _VT and π * _DK describe the already defined pressure reserves via valve train and throttle valve. They are calculated from a given distribution factor ξ * and the ratio of geometric to required effective compression ratio

In accordance with the required pressure reserve via valve train, an absolute delta inlet closing angle | Δφ * _ESV | certainly. By the calculation of the delta angle with the reference angle φ _UT and a factor K _ES , according to the specification "early" or "late-intake close" (FES / SES), the pilot-controlled target intake-closed angle results φ * _ESV for the intake valves of all cylinders from Equations 62-64.

In this case, the pilot control characteristic is the inverted mapping of the effect of an inlet-closed-angle variation from early to late at various operating points, averaged out of a set of curves.

Function block "Cylinder-individual controller"

For each cylinder, an individual controller determines a correction value for the previously calculated pilot control value based on the control deviation for the gas mass in the combustion chamber. As a result, the control correction acts directly on the valve train via the manipulated variable inlet closing angle. There is no correction of the setpoint values for the external gas state.

Setpoint m * _Cyl and actual value m _{Cyl (i)} are filtered with an equal parameterized IIR filter to improve the signal-to-noise ratio. In order for the two values to be comparable to one another, the setpoint value is additionally delayed by an idling time element by an evaluation step. This is necessary because the actual value is determined on the basis of the cylinder pressure. A current value is thus available only after the charge has been changed. The determined gas mass deviation e _{m (i)} is converted by means of a characteristic-based model into a control deviation e _{φ (i)} for the inlet-closing angle. This is then amplified by means of a linear proportional-integral controller (PI controller) with constant factors. The controller output φ * _{ESR (i)} is then held within the applied maximum and minimum values by a limiting function in the controller. Furthermore, the I component of the controller is limited to these values when the manipulated variable limits are exceeded (anti-windup function). The I component of the controller acts as an adaptive algorithm during the stationary transient process and adjusts the setpoint and actual masses to each other.

According to Equation 65, based on the pilot control value φ * _ESV and the controller output variable φ * _{ESR (i),} the cylinder-specific target value for the inlet-closing angle φ * _{ES (i) is} calculated. At this point, the factor K _ES according to the inlet-closing strategy must also be taken into account.

The feedforward and control algorithms for the gas mass adjustment are calculated in order to guarantee a working cycle synchronous specification in the angle-synchronous grid. That is, the functions of the feedforward control and regulation are called every 180 ° KW, wherein the control of the individual cylinders in firing order from the top dead center of the combustion takes place.

Function block "intake manifold pressure control"

The intake manifold pressure control includes a model-based pilot control and a parallel-connected loop with a linear PI element to improve the dynamic control behavior. Based on the target intake manifold pressure p * ₂ , the actual mass flow m * _DK and the current actual state before the throttle valve, p ₁ and T ₁ , the effective flow area A * _effDK at the valve is _{pre-controlled} using the throttle model according to Equation 43 (Equation 66). The precontrol value is corrected by an additive control component as a function of the calculated and amplified control deviation. With the aid of a characteristic curve, the conversion of the area value into a nominal sampling ratio TV * _DK for the flap position takes place. The calculation of these control functions takes place in a time-synchronized grid.

Function block "boost pressure control"

The subordinate loop for boost pressure adjustment includes a conventional boost pressure control. This includes a map-based, operating point-dependent precontrol of the duty cycle TV * _VTG for the VTG controller (controller for variable turbine geometry) on the turbocharger and a parallel control loop, whose input variable is the boost pressure desired value p * ₁ derived from the gas mass.

Illustration of the functional principle of the gas mass control

The control structure described above leads accordingly 11 to a decoupling of the cylinder-specific gas mass control of the external intake manifold and boost pressure control. A connection exists only via the setpoint specification.

Based on a rather poor control behavior of the boost pressure control, at least the overshoot can be reduced by the temporary buildup of pressure reserves generated by the intake manifold pressure control. It is inevitable that deviations from the specified setpoint for stationary operation. For the entire measurement section, no pressure reserve is requested via the throttle valve. Consequently, a boost pressure deficiency can not be compensated. Nevertheless, the intake manifold pressure control achieves a higher control quality than the boost pressure control. This is due to the type of control elements and the structure of the control circuits. Between intake manifold and cylinder also a pressure or density is required. This can not be achieved by the intake manifold pressure control in all operating points. At this point, the interaction between requested gas mass, pressure reserve and system dynamics becomes apparent. With the help of the valve train, however, these control errors can be compensated and the required gas mass in the combustion chamber can be set with sufficient control accuracy. In this case, the lowered effective compression in the form of the pressure reserve above the valve train is opened or released and the inlet closing angle is shifted from the pre-controlled late position in the direction of early. If necessary, the pressure reserve is additionally increased with further retardation. Due to this cascaded procedure results for the adjustment of the gas mass in the combustion chamber, a better follow-up behavior than by a sole adjustment using the boost pressure control.

Cylinder-specific adjustment of the oxygen content in the combustion chamber

The oxygen content in the combustion chamber forms a further important gas state variable, which can be adjusted starting from the oxygen content of pure fresh air (about 21% by volume) with the aid of the exhaust gas recirculation. External and internal options are available to provide the recirculated exhaust gas. With the external feedback, it is possible to mix large amounts of recirculated exhaust gas with the fresh air. Thus, in steady state operation, the required oxygen content in the combustion chamber should always be precisely set with the aid of the external actuator. In this case, a controlled division of the high and low pressure branch makes sense. Depending on the system, the long gas transit times and the pulsation behavior of the external components can have a negative effect on the current oxygen content in the combustion chamber, especially during transient engine operating phases. By means of the selective internal exhaust gas recirculation, made possible by a variable valve train, occurring oxygen overshoots can be corrected. Any undershoots in the follow-up behavior of the controlled variable, caused by an excessive proportion of recirculated exhaust gas, can be corrected in the same way for all cylinders with already reduced internal exhaust gas recirculation by briefly lowering the external high-pressure exhaust gas recirculation.

13 illustrates the functional scheme of the cylinder-specific control of the oxygen content. Analogous to the gas mass control, the structure consists of a model-based reference variable pilot control and a cylinder-specific control component.

The oxygen control uses the exhaust gas recirculation rate as a key variable. The division of the entire pilot-controlled rate and the cylinder-specific rule portion on the subsequent subordinate controller or controllers for the external and internal exhaust gas recirculation is carried out by an additional coordinator.

Function block "feedforward control"

The _{pilot-controlled} total _{exhaust gas recirculation} rate r * _{AGRV is} calculated from the ratio of the oxygen concentrations before and after the cylinder and represents the inversion of the model presented above for the cylinder oxygen content. The following equation 67 describes the calculation specification with the input variables setpoint oxygen content x * _O2Zyl and the actual values for Fresh air x _O2FL and exhaust oxygen content x * _O2ZylAbg . In deceleration phases of the engine, ie with the injection switched off, a fixed control value for the total exhaust gas recirculation rate is specified, since in such situations it would exceed the value 1.

In the subsequent coordinator, the total rate is divided into the internal and external rates as a function of the distribution factor v * _int (equations 68 and 69).

With the distribution factor v * _ND the pilot control values for the external low and high pressure exhaust gas recirculation can be determined (equations 70 and 71).

The internal exhaust gas recirculation rate is not further divided at this point. For the internal residual gas mass only a Basisrestgasrate r _{RGV (i) is formed} from the Istmassenverhältnis. This is necessary in order to specify a meaningful input variable for the gas exchange valve coordinator. Furthermore, due to the system, there will always be a proportion of residual gas mass in the combustion chamber.

Function block "Cylinder-individual controller"

For each cylinder, an individual controller determines a correction value for the instantaneous exhaust gas recirculation rate on the basis of the control deviation for the oxygen content in the combustion chamber. The rule component is divided according to the application strategy to the individual internal and external positioning options and offset against the pilot control values. Setpoint x * _O2Zyl and actual value x _{O2Zyl (i)} are filtered with an equal parameterized IR filter to improve the signal-to-noise ratio. In order for the two values to be comparable to one another, the setpoint value is additionally delayed by an idling time element by an evaluation step. This is necessary because the actual value is determined on the basis of the cylinder pressure. A current value as well as the actual gas mass is only available after the charge has been changed. The determined oxygen deviation e _{O2 (i)} is converted into a control deviation e _{r (i)} for the exhaust gas recirculation rate using the rate model (equation 74) (equation 73).

This is then amplified by a linear proportional-integral (PI) controller with constant factors. The controller output Δr * _{AGR (i)} is subsequently held within the applied maximum and minimum values by a limiting function in the controller. The function of the P component is based on the short-term change in the control variable to compensate for occurring exhaust gas recirculation fluctuations. The I component in turn acts as an adaptive algorithm for equalizing the setpoint and the actual value.

Function block "EGR coordinator"

The "AGR coordinator" function block divides and offsets the control components with the pilot exhaust gas recirculation rates based on the following equations. The set value for the external low pressure rate r * _ND is equal to the pilot value r * _NDV (Equation 74). She learns by the regulation part no correction.

In contrast to this, a correction of the pilot control value r * _{HDV of} the external high-pressure rate r * _HD takes place as a function of the control component averaged over all cylinders Δ r * _AGR of the oxygen regulator. This is necessary as soon as an increased control deviation is detected in all cylinders. Then the global oxygen content in the intake manifold is tracked. The activation of the external high-pressure rate correction takes place when exceeding or falling below specified limits ( Δ r _max , Δ r _min ) for the mean rule component (Equation 75).

In addition, the correction of the pilot rate for the internal rate r * _iAGRV and the residual gas rate r _{RGV (i) is carried out} with a corresponding distribution. From the individual correction rate Δr * _{AGR (i)} the mean correction rate Δ r * _AGR be deducted, so that the sole individual cylinder fraction can be calculated within and outside the aforementioned limit. Thus, even with external rate correction, a change in the cylinder rate is maintained. The division factor v * _Δ is calculated from an applicable map as a function of the engine speed and the rate correction value. Consequently, the cylinder- _specific setpoint values for the internal exhaust gas recirculation rate r * _{iAGR (i)} and the residual gas rate r * _{RG (i)} can be determined (equations 76-78).

The calculated rate setpoints are passed on to the subordinate modules. It should be noted that the rate setpoints range from 0 to 1.

Function block "External EGR control"

According to the presented regulatory structure ( 11 ), the actuators of the high and low pressure circuits are used to set the external exhaust gas recirculation. Each actuator is assigned a model-based feedforward control and a linear controller. As physical setpoint values, a mass flow value is specified for each of the LP and HP EGR valve controllers. These are calculated from the desired rates over the valves. The setpoint for the exhaust flap control is defined by the required pressure drop Δp _ND = p ₄ P ₀ above the low pressure valve. For this purpose, the throttle equation (equation 51) after the pressure before the exhaust flap (also corresponds to the pressure before the low pressure valve) dissolved, and determines the required for the adjustment of the mass flow at maximum valve cross-sectional area pressure upstream of the valve. The following equations 79-81 describe the setpoint calculations.

The model-based pilot control of each actuator is based on the already used throttling model according to Equation 43. Configured accordingly, the respectively required effective valve area A * _{eff is precontrolled} for each actuator (Equation 82). The pre-control value is corrected by an additive control component as a function of the calculated and amplified control deviation. With the aid of the associated characteristic curve, the conversion of the area value into a desired sampling ratio TV * for the respective valve or flap position takes place. The calculation of these control functions takes place in a time-synchronized grid.

Function block "Internal EGR control"

Since the internal EGR in transient engine operation and in cylinder-specific use is very difficult to capture with the existing measurement technology, this is only precontrolled with the aid of a model-based approach. The compensation occurring pilot error is shifted to the parent oxygen regulator. The pilot model is the inversion of the actual model for the internal exhaust gas recirculation mass.

Based on the following equations 83-86, the setpoint rate is converted into a desired mass m * _{iAGR (i)} and divided by means of an applicable factor v * _VL into "Vorlagern" (VL) and "Rücksaugen" (RS).

The functions for "Vorlagern" and "Rücksaugen" calculate from the respective mass setpoint depending on the current gas state per an effective area. These manipulated variables correspond to the Area integrals under the valve lift curves of the intake and exhaust valves. The two target areas are calculated as follows (equations 87 and 88).

For the release of the small valve lift curves, these must be adjustable with the valve actuators, i. H. the pre-controlled area under the curve must always be larger than a minimum area. Furthermore, when a maximum value is exceeded, the respective pilot control area is divided equally between both inlet and outlet valves. This maximum value can be specified dependent on the speed. Based on these criteria, the release for "pre-storage" or "back-suction" is coordinated via one or two inlet and outlet valves.

With the final respective setpoint integral for a valve, the required maximum lift height h * _{VL / RS (i) 1/2 is} determined by means of a characteristic curve and the maximum permissible valve lift for "pre-storage" or "back-suction". Subsequently, the calculation of the required opening duration Δφ * _{VL / RS (i) 1/2 takes place} on the basis of the desired surface integral and the lifting height, assuming a trapezoidal valve lift curve. Opening time and lifting height are therefore always in a certain ratio to each other. The setpoint for the starting angle φ * _{VL / RS (i) 1/2 of} the small valve lift curves is defined as a function of the engine speed and the desired mass for "Vorlagern" or "Rücksaugen". Due to this procedure, the three parameters corresponding to the activation apply to two valves of the same type. At this point, an extension of the approach with respect to a different Hubkurvenvorgabe per valve type is conceivable. For example, "pre-storing" with intake valve 1 earlier could be done with a smaller valve lift curve and via intake valve 2 offset thereto with a larger valve lift curve.

Function block "residual gas control"

The desired residual gas mass per cylinder is calculated in an analogous manner from the desired residual gas rate and the current cylinder gas mass (equation 89). The setpoint is forwarded directly to the gas exchange valve coordinator. m * / RG (i) = r * / RG (i) * m _{Zyl (i)} (89)

The control of the residual gas mass uses the closing positions and the lift heights of the outlet valves as manipulated variables. The compensation occurring pilot error is also relocated to the parent oxygen regulator. Due to the availability of a sensor-based actual model, it would also be possible to regulate the residual gas mass at this point. However, this only makes sense if a very accurate dosage of the residual gas mass is required. Since the possibilities of internal exhaust gas recirculation ("pre-storage" and "back-suction") in normal operation offer a higher potential for the oxygen control with constant efficiency, these methods are preferred over the residual gas mass increase.

Control of the charge movement

Another important gas state property is the charge motion. For the diesel combustion process, this is a rotational movement of the gas around the cylinder's longitudinal axis. This swirling motion makes a significant contribution to improved mixture formation at the time of injection. By means of the fully variable valve train, for example, a twisting movement can be generated by staggered opening of the intake valves or by lowering the valve lift of an intake valve. These variations must be performed on the engine test bench at various operating points with exhaust gas recirculation. The resulting swirl level can be estimated from a so-called Tippelmann test rig and a model calculation based on stationary measurements of the intake manifold-valve group become. Furthermore, an assessment of the lift curve strategy with regard to the influenced emissions of soot particles and nitrogen oxides can be made.

On the basis of these preliminary investigations, it is possible to determine a correlation between swirl level and delta opening angle or lifting height of the inlet valves. This correlation is transferred in a further step in a feedforward control. The required setpoint value for the swirl level DN * can be applied in a characteristic field as a function of the engine speed n _Mot and the total nominal exhaust gas recirculation rate r * _AGRZyl .

Coordination of gas exchange valves

The main tasks of the coordinator for the gas exchange valves is the processing of the requirements of the input variables to the valve lift curve parameters and their calculation. According to the definition of the regulatory structure ( 11 ) is every physical state quantity. at least one Hubkurvenparameter assigned as a manipulated variable. The following equations summarize the description of the main parameters of the valve lift curves.

Parameter calculation for the inlet valves

The position of the inlet opening φ * _{EOE (i) 1/2} for both inlet valves depends on the required residual gas mass m * _{RG (i)} and the swirl level DN * and can be applied. The parameter inlet-closing angle is a manipulated variable for setting a required cylinder _{gas mass} m * _cyl and a predetermined pressure reserve π * _VT via valve train. Furthermore, the maximum lift height per intake valve is not defined directly from a control function, but as a function of the opening duration of the respective intake valve (Δφ = φ * _{ES (i) 1/2} - φ * _{EOE (i) 1/2} ) (Equations 90- 92).

Parameter calculation for the exhaust valves

The opening angle of the exhaust valves φ * _{AOE (i) 1/2} are applied depending on operating point and loading pressure. The parameters exhaust-close angle φ * _{AS (i) 1/2} and maximum lift height h * _{AV (i) i / 2} are determined by engine speed and residual mass demand (equations 93-95).

Parameter calculation for setting the internal exhaust gas recirculation

The curve parameters for setting the internal exhaust gas recirculation are taken directly from the "Internal EGR control" function block. These are for "preload" or "back suction" the angular positions for valve opening φ * _{VL / RS (i) OE1 / 2} and valve closing φ * _{VL / RS (i) S1 / 2} and the lifting height h * _{VL / RS (i) 1/2} (equations 96-98).

In a first approach, the parameters φ * _{EOE (i) 1/2} , φ * _{AOE (i) 1/2} and φ * _{AS (i) 1/2} are calculated by applicable, map-based relationships. These can be extended with additional model functions.

For a complete definition of the 16 valve lift curves h * _(ij) (φ) further parameters are required, which are calculated depending on the main parameters (open, close, lift height). In addition, limits for the respective manipulated variables for safe operation can be applied in this function block. If the stroke curve parameters are suitably configured, almost all status requirements are met.

14 schematically shows an overall structure of a cylinder-individual control and control concept. The control and concept can, for example, in an engine electronics 1601 for an internal combustion engine of a vehicle 1600 as it is in 16 is shown, realized and an in 15 perform the procedure shown. There is a setpoint generator 1401 Setpoint values for a total gas mass m _Cyl and an oxygen content x02 _Cyl for each cylinder of the internal combustion engine before (step 1503 ). A device 1402 For signal evaluation and model calculation, the corresponding actual values of the total gas mass are determined (step 1501 ) and the oxygen content in the combustion chamber of the cylinder (step 1502 ) from sensor signals of the internal combustion engine, as described above. By comparing the actual values with the setpoints for the total gas mass and the oxygen content in step 1504 Adjusted actuators of the internal combustion engine.

In particular, depending on the _{reference variable total gas mass} in conjunction with a fully variable valve train, an inlet closing time phi _ES and a valve lift height h _Ven and an intake manifold pressure p _Sgr above a throttle position TV _ThrVlv and a charging pressure p _Lade above a turbocharger position TV _VTG can be set. Manipulated variables for the reference variable oxygen content can include, for example, the exhaust gas recirculation and a swirl function T _VSwVlv . The exhaust gas recirculation rate includes, for example, an internal exhaust gas recirculation iAGR by preliminary storage (VL) and / or backwashing (RS), an external exhaust gas recirculation eAGR (high pressure and low pressure circuit, eHD, eND) by a high pressure EGR valve (TV _HD ) and / or Low pressure EGR valve (TV _ND ) and residual gas mass (RG). The nominal values for the gas mass and oxygen content in the cylinder are derived from the higher-level setpoint generator 1401 specified. The cylinder-specific gas state control of 14 can be combined with a combustion control. Further reference variables in this case are, for example, an indicated mean pressure p _mi , a combustion center of gravity AQ50 and an injection strategy InjCrv with the manipulated variables injection quantity m _Kr over control period ASD and injection start phi _ASB over control start ASB.

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 102004030258 A1 [0005]
• DE 4428976 C1 [0006]
• DE 102006061936 A1 [0008]
• DE 102008032935 A1 [0008]
• WO 03/046356 A2 [0008]
• DE 102004044814 A1 [0008]

Claims (23)

Method for determining a gas temperature in a combustion chamber of an internal combustion engine, comprising: Determining a total mass, wherein the total mass comprises a total mass of gas and a mass of an amount of fuel in the combustion chamber at a predetermined crankshaft angle, Determining a combustion chamber pressure in the combustion chamber at the predetermined crankshaft angle, and Determining the gas temperature in the combustion chamber at the predetermined crankshaft angle as a function of the total mass and the combustion chamber pressure. The method of claim 1, further comprising: - Determining a speed of the internal combustion engine, and Determining a combustion chamber volume of the combustion chamber at the predetermined crankshaft angle, wherein determining the gas temperature comprises determining the gas temperature as a function of the total mass, the combustion chamber pressure, the rotational speed and the combustion chamber volume. The method of claim 2, wherein the predetermined crankshaft angle comprises a first crankshaft angle during an engine operating stroke and the determined gas temperature comprises a first gas temperature T ₄₆₀ , wherein a second gas temperature T ₅₄₀ at a second crankshaft angle at the end of the power stroke in response to the engine speed n _Mot , the combustion chamber pressure p ₄₆₀ at the first crankshaft angle, the first gas temperature T ₄₆₀ and empirically determined parameters a ₀ , a _{1 is} determined. The method of claim 3, further comprising determining a second combustion chamber pressure p ₅₄₀ at the second crankshaft angle, determining a third combustion chamber pressure p ₆₄₀ at a third crankshaft angle during an exhaust stroke of the internal combustion engine, and determining a third gas temperature T ₆₄₀ at the third crankshaft angle in FIG Depending on the speed, the second combustion chamber pressure p ₅₄₀ , the third combustion chamber pressure p ₆₄₀ and the second gas temperature T _540th The method of claim 4, further comprising determining a fourth gas temperature T _{RG of} a residual gas in the combustion chamber at a fourth crankshaft angle at the end of the exhaust stroke in dependence on the speed, the first combustion chamber pressure p ₄₆₀ , the third combustion chamber pressure T ₆₄₀ and the second gas temperature T ₅₄₀ , The method of claim 5, further comprising: determining an intake gas temperature T _EV at intake valves of the internal combustion engine, determining an indicated torque M _{ind of} the internal combustion engine, and determining a fifth gas temperature T ₁₀₀ at a fifth crankshaft angle during an intake stroke of the internal combustion engine as a function of Speed, the fourth gas temperature T _RG , the indicated torque M _ind and the inlet gas temperature T _EV . The method of claim 6, further comprising: determining a fifth combustion chamber pressure p ₁₀₀ at the fifth crankshaft angle, and determining a sixth gas temperature T ₁₈₀ at a sixth crankshaft angle at the end of the intake stroke in response to the speed, the fifth gas temperature T _100, and the fifth Combustion chamber pressure p ₁₀₀ . The method of claim 7, further comprising: determining a sixth combustion chamber pressure p ₁₈₀ at the sixth crankshaft angle, determining a combustion chamber volume V ₁₈₀ at the sixth crankshaft angle, determining a seventh combustion chamber pressure p at a seventh crankshaft angle during a compression stroke or a combustion engine's working stroke; and determining a combustion chamber volume V at the seventh crankshaft angle, determining a seventh gas temperature T at the seventh crankshaft angle as a function of the sixth gas temperature T ₁₈₀ , the fifth combustion chamber pressure p ₁₀₀ , the sixth combustion chamber pressure p ₁₈₀ , the combustion chamber volume V ₁₈₀ at the sixth crankshaft angle , the seventh combustion chamber pressure p, the combustion chamber volume V at the seventh crankshaft angle, the rotational speed and the indicated torque M _ind . A method for determining a gas temperature profile in a combustion chamber of an internal combustion engine, comprising: determining a plurality of gas temperatures in the combustion chamber of the internal combustion engine at different crankshaft angles of the internal combustion engine, and determining the gas temperature profile over the crankshaft angle by interpolating the plurality of gas temperatures, wherein the plurality of gas temperatures each according to one of previous claims are determined. Method for determining a residual gas quantity in a combustion chamber of an internal combustion engine, comprising: Determining a combustion chamber volume of the combustion chamber at the end of an exhaust stroke of the internal combustion engine, Determining a combustion chamber pressure in the combustion chamber at the end of an exhaust stroke, Determining a residual gas temperature of the residual gas quantity in the combustion chamber, and Determining the residual gas mass at the end of an exhaust stroke as a function of the combustion chamber volume, the combustion chamber pressure and the residual gas temperature at the end of the exhaust stroke. Method for determining an internal exhaust gas recirculation mass in a combustion chamber of an internal combustion engine, comprising: Determining a mean total cylinder mass in the combustion chamber of the internal combustion engine, Determining a mean residual gas mass in the combustion chamber, Determining a mean injected fuel mass in the combustion chamber, - Determining an oxygen content in the exhaust gas, and - Determining the average internal exhaust gas recirculation mass as a function of the average total cylinder mass, the average residual gas mass, the average injected fuel mass and the oxygen content in the exhaust gas. Method for determining a total gas mass in a combustion chamber of an internal combustion engine, comprising: determining a cylinder pressure characteristic p _Zyl at the beginning of a compression stroke of the internal combustion engine, determining a combustion chamber volume V _Zyl at the beginning of the compression stroke, determining a cylinder temperature T _Zyl at the beginning of the compression stroke, and Determining the total gas mass m _Cyl in the combustion chamber as a function of the cylinder pressure characteristic p _Zyl , the combustion chamber volume V _Zyl and the cylinder _temperature T _cyl . The method of claim 12, further comprising: determining the sixth gas temperature T ₁₈₀ according to the method of claim 7, and determining the cylinder temperature T _Zyl at the beginning of the compression stroke as a function of the seventh temperature according to the method of claim 8. Method for determining an oxygen content during a compression _stroke in a combustion chamber of an internal combustion engine with exhaust gas recirculation, comprising: determining an exhaust gas recirculation rate r _AGRZyl via the exhaust gas recirculation, determining a fresh air _oxygen content x _O2FL of fresh air supplied to the internal combustion engine, determining an exhaust gas oxygen content x _{O2ZylAbg of} an exhaust gas of the exhaust gas Internal combustion engine, which is supplied to the internal combustion engine via the exhaust gas recirculation, and - Determining the oxygen content x _O2Zyl in the combustion chamber as a function of the exhaust gas recirculation rate, the fresh air oxygen content and the exhaust gas oxygen content. The method of claim 14, wherein the exhaust gas recirculation comprises an internal exhaust gas recirculation, an external exhaust gas recirculation and a residual gas increase, wherein the internal exhaust gas recirculation and the residual gas increase is adjustable via a variable valve timing of valves of the internal combustion engine, wherein determining the exhaust gas recirculation rate r _{AGRCyl on} determining an exhaust gas recirculation rate the basis of an external exhaust gas recirculation rate r _eAGR the external exhaust gas recirculation, an internal exhaust gas recirculation rate r _iAGR the internal exhaust gas recirculation and a residual gas rate r _RG comprises a remaining in the combustion chamber after combustion residual gas. The method of claim 15, further comprising: Determining a real pressure ratio π of two measured pressure values during the compression stroke at two different crankshaft angles, determining a modeled pressure ratio π _mod at the two different crankshaft angles in dependence on an engine speed n _Mot , an indicated torque M _ind , a gas mass m _Zyl in the Combustion chamber and the determined oxygen content x _O2Zyl in the combustion chamber, and - determining a correction _value x _O2ZylKorr for the oxygen content x _O2Zyl as a function of the engine speed, the indicated torque, the real pressure ratio and the modeled pressure ratio. A method for controlling an internal combustion engine with an internal exhaust gas recirculation, wherein the internal exhaust gas recirculation is adjustable via a variable valve control of valves of the internal combustion engine, comprising: - determining a total gas mass in a combustion chamber of the internal combustion engine, - determining an oxygen content in the combustion chamber of the internal combustion engine, - determining a Target total gas _mass m _Cyl and a target oxygen content x _O2Zyl in the combustion chamber for a given operating point of the internal combustion engine, and - adjusting an actuator of the internal combustion engine as a function of the determined total gas mass, the determined oxygen content, the target total gas mass and the desired oxygen content. The method of claim 17, wherein the total gas mass is determined according to one of claims 12 or 13, and wherein the oxygen content is determined according to any one of claims 14-16. The method of claim 17 or 18, wherein the actuator comprises the variable valve timing, a throttle of the internal combustion engine and / or a turbocharger adjustment device for adjusting a turbocharging pressure of the internal combustion engine, wherein adjusting the actuator adjusting the valves, the throttle and / or the turbocharger according to from the determined total gas mass and the target total gas mass. Method according to one of claims 17-19, wherein the actuator the internal exhaust gas recirculation, via the variable valve control adjustable residual gas increase, an external exhaust gas recirculation ( 305 . 306 ) of the internal combustion engine ( 300 ) and / or a swirl flap of the internal combustion engine, wherein adjusting the actuator setting the internal exhaust gas recirculation, the residual gas increase, the external exhaust gas recirculation ( 305 . 306 ) and / or the swirl flap as a function of the determined oxygen content and the desired oxygen content. The method of claim 17, wherein the internal combustion engine comprises a plurality of cylinders, wherein the variable valve control comprises a fully variable cylinder-specific valve control, wherein the total gas mass, the oxygen content, the target total gas mass and the desired oxygen content are determined individually for each cylinder, and wherein the adjusting of the actuator comprises a cylinder-individual adjustment of the valves. Control device for an internal combustion engine, characterized in that the control device ( 1601 ) is configured for carrying out the method according to one of the preceding claims. Vehicle with an internal combustion engine and a control device ( 1601 ) according to claim 22.

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