DE102016212945A1 - Method and device for controlling an internal combustion engine with an exhaust gas turbocharger - Google Patents

Abstract

The subject of the invention relates to a method and a device for controlling an internal combustion engine having at least one exhaust gas turbocharger, wherein an increase in the power requirement to the internal combustion engine, a speed increase of the turbocharger is supported by an increase in the exhaust gas temperature. In this case, depending on the power requirement and the respective current operating point of the exhaust gas turbocharger, means of corresponding calculation models, a current target exhaust gas temperature and on the basis of required to achieve the target exhaust gas temperature target setpoint control value of at least one exhaust gas temperature influencing control parameter determined and to control the internal combustion engine applied to increase the exhaust gas temperature. The combustion center position is used as a control parameter, and the combustion center position is shifted only for a part number of the total existing cylinders of the internal combustion engine with respect to the combustion center position at the time before the increase in the power requirement at a later time. As a result, while avoiding thermal overloading of the exhaust gas turbocharger, the response time of the internal combustion engine is significantly shortened, in particular in the event of a sudden increase in the power requirement.

Description

The invention relates to a method and a device for controlling an internal combustion engine, which is equipped with an exhaust gas turbocharger for increasing the power.

Internal combustion engines, in particular for motor vehicles, are in the majority designed as reciprocating internal combustion engines, to which reference is made in the present invention in particular. However, rotary internal combustion engines (Wankel principle) are also encompassed by this generic term and the scope of the invention.

Reciprocating internal combustion engines have at least one reciprocating cylinder, hereinafter referred to as cylinder, but often at least three or more cylinders. Per cylinder, a reciprocating piston is provided which limits a combustion chamber with the cylinder and can move up and down in the cylinder over a distance called a piston stroke. By filling the combustion chamber with a fuel-air mixture and the subsequent combustion of the fuel of the reciprocating piston is driven and the movement of the reciprocating piston is transmitted via a connecting rod to the crankshaft of the internal combustion engine, which is thereby rotated and the torque required to drive a motor vehicle emits. In this case, so-called gasoline engines and diesel engines are distinguished, with gasoline engines with normal or premium gasoline in foreign ignition by means of at least one ignition device, for. As spark plug, per cylinder and diesel engines are operated with diesel fuel in auto-ignition.

Due to the ever stricter legislation with regard to the exhaust gas performance of internal combustion engines and the increased performance expectation of users, ever greater demands are placed on the design, control and regulation of the combustion process of the fuel in the combustion chamber and the exhaust aftertreatment. To meet these requirements, modern internal combustion engines are equipped with electronic, usually by means of a central control unit, the so-called ECU (Electronic Control Unit), controlled systems for fuel and Luftzumessung and ignition, valve opening and exhaust gas recirculation and exhaust aftertreatment.

For fuel metering direct fuel injection systems are used in which the fuel under high pressure, for example in a so-called common rail and electronically controlled injectors depending on the current power requirement, high-precision doses, high pressure, often in several parts -Injection quantities split and injected into the combustion chamber timely to the desired time.

Also, the oxygen required for combustion is in the form of ambient air, possibly under, by a supercharging system, for example, an exhaust gas turbocharger, increased pressure, as needed metered via throttle valves and intake valves into the combustion chamber of the respective cylinder.

The ignition of the fuel-air mixture in the combustion chamber is carried out as a function of the instantaneous operating point of the engine timely with respect to the correlating with the current crankshaft rotation angle Hubkolbenposition, said instantaneous crankshaft rotation angle is also referred to as a so-called ignition angle. The ignition is specified in gasoline engines via an actuated by the central control device ignition device and in diesel engines by the injection timing of the fuel.

Another possibility of influencing the combustion process is in internal combustion engines, which are equipped with a so-called variable valve train, so a system for variable adjustment of the valve opening intervals. With such a system, the opening cross sections and the opening times of the intake valves, via which the intake air from the intake system into the combustion chamber, and the exhaust valves, via which the exhaust gas from the combustion chamber passes into the exhaust system after combustion, can be variably controlled within certain limits , As a result, the so-called charge change in the combustion chamber, ie the exchange of exhaust gas for fuel-air mixture and thus the power output can be significantly influenced.

As a further, the combustion process and thus the performance of the internal combustion engine significantly influencing aggregate come especially in diesel engines, but increasingly often in gasoline engines charging systems, in particular exhaust gas turbocharger used. This is increasingly done with the aim of reducing the internal combustion engine with the same or even increased performance in size and weight while reducing consumption and thus the CO ₂ emissions, in view of increasingly stringent legal requirements in this regard. The operating principle is to use the energy contained in the exhaust stream to increase the pressure in the intake system before the cylinders of the engine and thus to better fill the combustion chamber with air-oxygen and thus implement more fuel, gasoline or diesel, per combustion process can, so to increase the performance of the internal combustion engine.

For this purpose, the exhaust gas turbocharger has an exhaust gas turbine arranged in the exhaust system of the internal combustion engine, a fresh air compressor arranged in the intake system, and a rotor bearing arranged therebetween. The exhaust gas turbine has a turbine housing and a turbine runner, which is arranged therein and driven by the exhaust gas mass flow. The fresh air compressor has a compressor housing and a compressor impeller arranged therein, which builds up a boost pressure. The turbine runner and the compressor runner are arranged on the opposite ends of a common shaft, the so-called rotor shaft, rotatably and thus form the so-called turbocharger rotor. The rotor shaft extends axially between the turbine runner and the compressor runner through the rotor bearing arranged between the exhaust gas turbine and the fresh air compressor and is radially and axially rotatably mounted therein, with respect to the rotor shaft axis. According to this construction, the turbine wheel driven by the exhaust gas mass flow drives the compressor wheel via the rotor shaft, whereby the pressure in the intake system of the internal combustion engine, based on the fresh air mass flow behind the fresh air compressor, is increased and thereby a better filling of the combustion chamber with air oxygen is effected.

In a sudden increase in load, for example, in connection with an acceleration process of a motor vehicle, the turbocharger rotor is delayed accelerated by the increasing exhaust pressure, which is based on the fact that at low exhaust gas mass flow of the engine and low starting speed of the turbocharger rotor both the turbine, and the compressor of the turbocharger very have poor efficiencies. For this reason and reinforced by the inertial mass of the turbocharger rotor, the pressure build-up in the intake system is delayed, which in turn has a delayed response of the internal combustion engine, which is generally referred to as a so-called turbo lag, the result. This turbo lag is not only perceived as unpleasant by the driver of the vehicle, since the vehicle only reacts with delay to a deliberately dynamic power requirement, for example during an overtaking maneuver, but also represents a security risk which should not be underestimated, for example if an over-hauling operation occurs the "turbo lag delay" can not be completed on time.

This undesirable effect is countered with different approaches, such as the parallel or sequential arrangement of multiple turbochargers of different size, power and response. Another approach consists in the additional arrangement of a compressor unit operable independently of the exhaust gas flow in the intake system, which is used as a supplement to an exhaust gas turbocharger, specifically in transient operating phases of the internal combustion engine, for rapid pressure build-up in the intake system, ie for elimination of the turbo lag. However, these solutions are constructive and conceptually very expensive and correspondingly expensive.

Another approach to at least downsizing said turbo lag is, for example, in internal combustion engines equipped with a variable valve train, to vary the opening times of the intake and exhaust valves of the respective cylinder with respect to the crankshaft angle so as to overlap the opening times of the intake and exhaust valves, so comes to a persisting over a certain time simultaneous opening of intake and exhaust valves. This leads to the so-called over-flushing, which is also known by the term "scavening", wherein the pressure gradient between the intake system and the exhaust system promotes the charge cycle in the cylinder and better filling of the combustion chamber with air can be achieved. Thus, a larger amount of fuel can be metered, resulting in an immediate increased power output. The disadvantage of this method is a worsened exhaust gas behavior, since in this mode of operation an excess of oxygen (lambda> 1) must therefore be run with a so-called lean fuel-air mixture, for example to limit the catalyst temperature, but this leads to increased NOx emissions. as a result of which the limit values established in existing, but at least in future emissions legislation can no longer be met.

Another way to reduce the turbo lag is the targeted increase in the exhaust gas temperature in front of the turbine of the exhaust gas turbocharger. This results in greater exhaust enthalpy which results in faster turbocharger rotor acceleration and thus faster pressure build-up by the compressor in the intake system.

Such a method is for example in the document DE 10 2011 081 844 A1 disclosed. This relates to a method for operating a supercharged internal combustion engine having at least one exhaust gas turbocharger, wherein the response of the exhaust gas turbocharger or the charged by turbocharging internal combustion engine is improved at a load jump. This is achieved by a method which is characterized in that at a load jump of the ignition timing is shifted from an optimized in terms of efficiency ignition timing late and indeed one to avoid one knocking combustion required ignition timing addition.

Also the document DE 101 40 120 A1 discloses a method for operating an internal combustion engine with an exhaust gas turbocharger, in which the response of the internal combustion engine is improved by increasing the exhaust gas enthalpy. For this purpose, when overlapping the opening angle of the intake and exhaust valves of a cylinder post-injection of fuel is made, which leads to an air-fuel mixture in the exhaust system, which is combustible and ignited by driving an additional ignition in the exhaust system in front of the turbine of the exhaust gas turbocharger and burned.

Similarly, in the document DE 199 44 190 A1 disclosed method, an increase in the efficiency of the exhaust gas turbocharger and thus the torque in an internal direct fuel injection engine achieved in that in addition to a primary injection post injection into one or more cylinders of the internal combustion engine is caused, resulting in an increase in the exhaust gas temperature in front of the turbine of the exhaust gas turbocharger leads.

However, the abovementioned processes for increasing the exhaust gas enthalpy have the disadvantage that the exhaust gas temperature can increase uncontrollably and, especially in gasoline engines, there is a risk of thermal damage to the exhaust gas turbine or the downstream components in the exhaust system, such as soot particle filters and catalysts. Furthermore, there is the danger that the control interventions, which are oriented only on the exhaust gas enthalpy optimization, have a negative effect on the operating behavior of the internal combustion engine as a whole. This can result in increased emission levels or even additional torque reduction, for example in the case of an ignition point shifted too far to the rear or exhaust valves that are opened too early.

The present invention is therefore based on the object of specifying a method and an apparatus for controlling an internal combustion engine having at least one exhaust-gas turbocharger by which the torque weakness is safely and significantly reduced in the event of a sudden power demand from low engine speeds. while at the same time reliably avoiding thermal overloading of the system components as well as excessively increased pollutant emissions.

This object is achieved by a method and a device for controlling an internal combustion engine with the features of the independent claims. Advantageous training and further education, which individually or, if it is not mutually exclusive alternatives, can be used in combination, are the subject of the dependent claims.

In the method according to the invention for controlling an internal combustion engine having a plurality of cylinders and having an exhaust-gas turbocharger, an increase, in particular in the event of a sudden increase, of the power requirement on the internal combustion engine will increase the speed of the turbocharger and thus increase the retrievable torque by increasing it the exhaust gas temperature supported. In this case, the method according to the invention is characterized in that initially as a function of the power requirement and the respective current operating point of the exhaust gas turbocharger, an instantaneous desired exhaust gas temperature is determined by means of an exhaust gas turbocharger calculation model. Subsequently, based on the determined target exhaust gas temperature, the desired control value required to achieve the target exhaust gas temperature is determined by means of an exhaust gas temperature calculation model of at least one control parameter influencing the exhaust gas temperature. From the target control value, a default control value is then determined, including at least one limit control value, and the determined default control value of the at least one control parameter influencing the exhaust gas temperature is used to control the internal combustion engine to increase the exhaust gas temperature, the combustion center position being the control parameter and the combustion center position is shifted only for a part number of the total existing cylinders of the internal combustion engine with respect to the combustion center position at the time before the increase of the power requirement at a later time.

The calculation models used in the method according to the invention and the further embodiments of the method, the exhaust-gas turbocharger calculation model, the exhaust-gas temperature calculation model, the combustion boundary calculation model and the torque calculation model are mathematical calculation models based on the physical laws of the internal combustion engine or of the exhaust-gas turbocharger Form of appropriate program algorithms, for example, stored in the central control unit. These calculation models can calculate the desired variable manipulated variables or control values for controlling the subsystems on the basis of or as a function of sensor-determined, constructively determined system parameters available in operating maps or prescribed for operational reasons. Such models can be designed and programmed in such a way that, depending on which variables are specified, the other variable can be determined as a result variable. For example, the calculation path can also be reversed.

Thus, the exhaust gas turbocharger calculation model is provided in this case, starting from a torque or power demand on the internal combustion engine, which is specified for example by accelerator pedal or accelerator pedal by the driver, and including the current, for example, sensory determined operating variables such as Example engine speed, boost pressure in the intake system, rotor speed of the turbocharger, exhaust enthalpy (pressure and temperature) before and after the exhaust turbine, etc., to calculate the required for a rapid increase in speed of the turbocharger rotor increase in exhaust enthalpy and thus the target exhaust gas temperature in front of the turbine.

The exhaust-gas temperature calculation model, on the other hand, is intended to determine at least one desired control value of at least one control parameter influencing the exhaust-gas temperature on the basis of the required desired exhaust-gas temperature determined by means of the exhaust-gas turbocharger calculation model. Parameters that can be included in this calculation are, for example, the mass and the temperature of the intake air and the metered fuel mass. The desired control value to be calculated can then be, for example, the ignition time or the injection time of the fuel or else the opening times or opening angles of the intake and exhaust valves. A combined adaptation of the control values of a plurality of said control parameters can also be provided.

An essential point of the method according to the invention is that a default control value is determined from the desired control value including at least one limit control value. For the desired control values to be determined, the control parameters influencing the exhaust gas temperature generally have limits beyond which the operating behavior of the internal combustion engine as a whole or in relation to individual features, such as the emission behavior or the engine torque, undesirably changes negatively. In order to avoid this, corresponding limit control values can now be specified fixedly or also as a function of the operating point and used in the determination of the default control values ultimately used for the temperature increase. This happens in the sense that, for example, when the limit control value is exceeded by the determined desired control value, the default control values are set to the value of the limit control value, whereas when the limit control value is undershot, the determined desired control value, the default control values are set to the value of the desired control value.

This method has the significant advantage over the previously known methods that the intervention in the control of the internal combustion engine to increase the exhaust gas temperature is limited so that an optimal increase in torque can be achieved without thereby negatively affecting the negative pressure on the performance of the internal combustion engine Need to become.

If the combustion focus position is used as a control parameter and the combustion center of gravity is shifted for a part number of the total existing cylinders of the internal combustion engine with respect to the combustion focus position at a time before increasing the power requirement at a later time, not only results in an advantageous increase in the exhaust gas temperature and thus a Enthalpy increase, but it is also achieved a Abgasimpulserhöhung. It is the Ausschiebeimpuls increased significantly at the same Enthalpieerhöhung, which in turn leads to an acceleration of the turbine of the exhaust gas turbocharger.

If the internal combustion engine has an even number of cylinders, then it is expedient to shift the center of gravity of the combustion in half of the total cylinders present. The cylinders in which a shift of the combustion center position is performed can also be selected alternately. It is only to be noted that a shift of not all existing cylinders at the same time takes place.

The combustion center position in this context is referred to as the timing during the combustion process at which a certain proportion (e.g., 50%) of fuel introduced into the combustion chamber has been completely burned.

A refinement of the method according to the invention is characterized in that the at least one limit control value is determined by means of a combustion limit calculation model for the current operating point of the internal combustion engine. The so-called burn limit represents a maximum value for the shift of the ignition angle or the ignition point to late. The retarded ignition angle must not exceed a maximum latest value, the combustion limit, otherwise an allowable exhaust gas temperature or a limit for emissions is exceeded. The burning limit is often determined by the prior art on the basis of the limit value for the emission of hydrocarbons and the limit value for the Manifold temperature determined. The firing limit also depends to a certain extent on the instantaneous operating point and can be determined by means of the mentioned firing boundary calculation model for the respective instantaneous operating point. This in turn results in a maximum ignition angle which can be specified as limit control value or also a latest injection time. This has the advantage that the exhaust gas temperature can always be increased to an optimum at the respective operating point and does not have to be prematurely reversed for safety reasons or kept too low overall.

A further embodiment of the method according to the invention is characterized in that the at least one limit control value is determined by means of a torque calculation model for the current operating point of the internal combustion engine. The torque calculation model determines based on current operating and control variables, the expected torque at the crankshaft of the internal combustion engine. As previously mentioned, as the change in certain control parameters continues after an initial increase in torque due to the increase in the exhaust gas temperature when a respective limit control value is exceeded, undesirable torque drops may occur. The advantage of this embodiment of the method lies in the fact that the torque maximum that is possible in the respective operating point can be determined by means of the torque calculation model and a respective limit control value correlating therewith can be defined for the respective control parameter. Thus, the maximum possible torque can be generated at each operating point.

In another embodiment of the method according to the invention, the determined target exhaust gas temperature is limited to a predetermined limit exhaust gas temperature. The limit exhaust gas temperature is chosen so that it is ensured that no thermal damage to the components of the exhaust gas turbocharger and the other components in the exhaust system is to be feared. As a result, relatively easy overheating of the system can be prevented.

In a further advantageous embodiment of the method according to the invention, the at least one control parameter influencing the exhaust gas temperature is an ignition point of at least one cylinder of the internal combustion engine which is displaced to increase the exhaust gas temperature to a later point in time than a normal ignition point the ignition point is late, the so-called center of gravity is shifted late, so temporally closer to the opening of the exhaust gas outlet valve out. As a result, the exhaust gas at the outlet from the combustion chamber of the cylinder has a higher temperature. The spark timing can be done in gasoline engines by the time-controlled spark ignition and diesel engines through the timed injection of diesel fuel into the combustion chamber. The advantage of this design is seen in the relatively simple implementation without additional required system components.

A further embodiment of the method according to the invention provides that the internal combustion engine is equipped with a direct injection system for direct fuel injection into combustion chambers of the plurality of cylinders and the combustion center position is carried out by changing the injection timing and / or the injection mass. This allows, for example, the distribution of the injected fuel quantity into several subsets and a late injection of a partial fuel, whose late combustion has a temperature increase in the exhaust gas result. In this way, even with diesel engines, the focus of combustion can be shifted late.

In a particularly advantageous embodiment, the method is repeatedly carried out until a desired boost pressure in the intake system is reached or the internal combustion engine has reached a requested performance level or the increased power requirement is withdrawn. This ensures that the method, optimized for the current and dynamically changing operating conditions in transient operation of the internal combustion engine, is optimized and executed as long as necessary. The higher the repetition rate of the execution of the method, the more accurate the execution is in direct dependence on the current operating point.

The features and combinations of features of the above mentioned in the description or below in the description of the embodiments of the subject invention are, as far as they are not exclusively alternatively applicable or mutually exclusive, individually, in part or in total, even in mutual combination or mutual complement, in To use training of the subject invention, without departing from the scope of the invention.

With reference to the figures, particularly advantageous embodiments, details or developments of the invention will be explained in more detail below, although the subject matter of the invention should not be limited to these examples.

Show it:

1 a simplified flow diagram of a method according to the invention for controlling an internal combustion engine and

2 a schematically simplified illustration of an internal combustion engine according to the invention with an exhaust gas turbocharger.

The flowchart in 1 shows in rough steps the sequence of the method for controlling an internal combustion engine with an exhaust gas turbocharger, in which an increase in the power requirement to the internal combustion engine, a speed increase of the turbocharger is supported by an increase in the exhaust gas temperature. The process sequence for increasing the exhaust gas temperature is embedded in a higher-level method for controlling the internal combustion engine, which is executed for example in the form of an executable sequence program in the central control unit ECU, wherein on the one hand sensed by sensors current state variables of the internal combustion engine read via inputs of the control unit ECU and according to the program On the other hand, control values of various control parameters are output to corresponding actuators determining the operation.

From an upstream method step SX of the higher-level method for controlling the internal combustion engine, it is monitored in a branch VX whether there is an erratic power requirement, ie a torque difference between the instantaneous increased torque and the suddenly requested increased torque, also referred to as load step TQ_D. This is the case, for example, when starting from an operating point of low load, at low engine speed of the internal combustion engine and the turbocharger rotor, suddenly a greatly increased power is requested, and thus a so-called load step TQ_D is present. This is given, for example, in the rapid startup of a vehicle from idle speed or at the start of an overtaking when the driver, for example, by a so-called "kickdown" of the accelerator, suddenly requests a greatly increased performance or even the maximum power of the engine. As long as such a sudden increase in the power requirement is not determined, the higher-level method is continued according to the program in a method step SY.

However, as soon as a load jump TQ_D is detected in the branch VX, in a subsequent method step SX1 a momentary desired exhaust gas temperature is determined in dependence on the power requirement and the respective current operating point of the exhaust gas turbocharger, ie depending on the height of the load step TQ_D by means of an exhaust gas turbocharger calculation model AT_S determined. Depending on the embodiment of the method, as already shown in the flowchart, the determined target exhaust gas temperature AT_S can already be limited by means of a predetermined limit exhaust gas temperature AT_GR, whereby a first protection against overheating of subsequent system components lying in the exhaust gas flow is realized.

Then, in the subsequent method step SX2, based on the determined target exhaust gas temperature AT_S, by means of an exhaust gas temperature calculation model, the desired control value StW_S required to achieve the target exhaust gas temperature AT_S is determined at least one control parameter influencing the exhaust gas temperature. Corresponding control parameters are, for example, the ignition times or the injection times of the fuel in the individual cylinders, the injected fuel mass or the opening times of the exhaust valves.

In the subsequent method step SX3, a default control value StW_VG is determined, taking into account at least one limit control value StW_GR, from the desired control value StW_S. The corresponding respective limit control value StW_GR can be predefined in maps, for example, fixed or as a function of the operating point of the internal combustion engine. In addition or instead, however, at least one limit control value StW_GR_BG can be determined by means of a combustion limit calculation model for the current operating point of the internal combustion engine, or a limit control value StW_GR_TQ can additionally be determined or alternatively determined by means of a torque calculation model for the current operating point of the internal combustion engine Determination of the default control value StW_VG be used. The thus determined default control value of the at least one exhaust gas temperature influencing control parameter for controlling the internal combustion engine is then output and applied to the control of the internal combustion engine to increase the exhaust gas temperature.

After the determination of the default control value StW_VG in method step SX3 and the application of the default control value StW_VG takes place in the branch VX1 monitoring of the already achieved torque or the currently present load step TQ_D between the requested and the output torque or power. Likewise, also in the fresh air manifold 71 Pending pressure P_L be monitored.

If there is still a relevant load step TQ_D or a torque difference or the pending boost pressure P_L has the desired value, a target boost pressure, not yet reached, there is a return to step SX1 and it is again the now-current target exhaust gas temperature determined. In this way, the process is repeatedly carried out until the load jump is overcome or a desired boost pressure in the intake system is reached or the engine has reached the requested power level or the increased power requirement is withdrawn.

2 shows a schematic representation of an internal combustion engine 1 , the four cylinder 11 , a fuel direct injection system 3 , an intake system 7 , an exhaust system 8th and at least one exhaust gas turbocharger 2 with one in the exhaust system 8th arranged exhaust gas turbine 22 and one in the intake system 7 arranged compressor 21 having. Every cylinder 11 has an ignition device 4 , an injection valve 32 , an inlet valve 51 and an exhaust outlet valve 52 on. For better clarity of 1 are the aforementioned components per cylinder 11 only on the first cylinder 11 marked with reference number. Furthermore, the internal combustion engine in this example has a variable valve train 5 on, here symbolic with a camshaft actuator 53 and a camshaft 54 is shown.

The intake system 7 has a fresh air manifold 71 each with a supply line for each cylinder 11 and a throttle disposed upstream with respect to the fresh air flow, see arrows 72 for controlling the fresh air mass supplied. Upstream of the throttle 72 is the compressor 21 the exhaust gas turbocharger 2 in the intake system 7 arranged. Upstream of the compressor 21 is an air filter system 73 for cleaning the compressor 21 sucked outside air in the intake system 7 arranged. Furthermore, in the fresh air manifold 71 a charge air pressure sensor 74 , for measuring the boost pressure upstream of the cylinders, and upstream of the compressor 21 an air mass sensor 75 , for detecting the air mass supplied, in the intake system 7 arranged. The charge air pressure sensor 74 and the air mass sensor 75 are connected via electrical connections with the electronic control unit ECU and thus feed the recorded measured values into the ECU.

The exhaust system 8th has an exhaust manifold 81 with one connection to each of the cylinders 11 on. In the flow direction of the exhaust gas, see arrows, downstream of the exhaust manifold 81 is the exhaust gas turbine 22 arranged the exhaust gas turbocharger in the exhaust system. Downstream of the exhaust gas turbine 22 then closes the exhaust aftertreatment system 82 which is simplified here as a single component, but may have a plurality of components, such as a catalyst, and a soot filter. In the exhaust system 8th are before the inlet of the exhaust gas turbine, an exhaust gas pressure sensor 83 and an exhaust gas temperature sensor 84 and downstream of the exhaust gas turbine 22 another exhaust gas pressure sensor 85 and, for example, a catalyst temperature sensor 86 are also connected via electrical connections to the ECU and so feed the recorded readings into the ECU.

The fuel direct injection system 3 also has a common rail 31 designated fuel pressure accumulator, which via a fuel supply line 33 from a high pressure fuel pump (not shown). Furthermore, those with the common rail 31 in hydraulically connected injection valves 32 the single cylinder 11 to the fuel direct injection system 3 , The injectors 32 continue to be in electrical communication with the ECU, which controls the injectors 32 in accordance with the determined values for the injection time and injection mass and thus drives the fuel into the individual combustion chambers of the cylinders 11 attaches. At the common rail 31 is a fuel pressure sensor 34 which is also in electrical communication with the ECU and detects the fuel pressure in the common rail and transmits corresponding values to the ECU. Usually have fuel direct injection system 3 additionally still different valve units, such as a pressure control valve or a flow control valve for controlling the fuel pressure in the common rail, on. For the sake of clarity, the representation of these components has been described 2 waived. It is understood, however, that in the context of the present invention, all conventional or known to those skilled configurations of direct fuel injection systems can be used. In gasoline engines beyond other injection systems, such as intake manifold injection systems can be used, since the ignition timing set by a separate ignition of the ECU and thus an increase in the exhaust gas temperature can be achieved.

The variable valve train 5 has a camshaft actuator 53 on, which is electrically connected to the electronic control unit ECU and is controlled by this for the variable adjustment of the valve opening intervals. Of course, all known in the art systems for variable adjustment of the valve opening intervals can be used in the invention.

Every cylinder 11 of the internal combustion engine 1 is in the example shown with an igniter 4 , for example, a spark plug, which is respectively electrically connected to the ECU and the ECU for igniting the fuel mixture in the combustion chamber of the respective cylinder 11 , in accordance with the determined ignition timing, is controlled directly or indirectly. In diesel engines, such a device can be omitted and the ignition point or the center of gravity is set by means of the high-pressure injection of the diesel fuel.

For controlling the operation of the internal combustion engine 1 an electronic control device ECU is provided, which is electrically connected via electrical connections to sensors and control systems of the internal combustion engine, as already described above. Furthermore, the ECU is provided with a turbocharger rotor speed sensor 23 and a crankshaft speed sensor 12 as well as a driver interface 15 , For example, an accelerator pedal or accelerator pedal to specify the power or torque requirement, in electrical connection.

About the various sensors as well as the driver interface 15 become the respective current operating parameters of the internal combustion engine 1 and the driver's request is detected and fed to the ECU. In the ECU, those are for controlling the internal combustion engine 1 or the internal combustion engine 1 assigned sub-systems required procedures and algorithms stored in flow programs, and deposited corresponding default values and maps that are used to calculate or determine the respective control values of the operating control parameters, depending on the driver's request. The corresponding control values are then output by the ECU directly or via corresponding so-called power amplifiers to the actuators in order to set the corresponding control parameters.

In particular, the electronic control device ECU is program-technically and system-technically designed to operate the internal combustion engine 1 to control in a proper operation according to a respective embodiment of the aforementioned control method, wherein an increase in the power requirement to the internal combustion engine 1 a speed increase of the turbocharger 2 is supported by increasing the exhaust gas temperature.

Finally, it should be pointed out once again that the in 2 illustrated internal combustion engine according to the invention 1 is merely an example so to speak in maximum equipment with respect to the different embodiments of the method according to the invention. For example, a diesel engine does not require separate ignition devices. A variable valve train is only required if the exhaust gas temperature is to be influenced by means of the valve opening intervals.

Furthermore, the invention has been explained with reference to an internal combustion engine having 4 cylinders. However, the described method or the device is also applicable to internal combustion engines other cylinder number, provided that they have at least 2 cylinders.

Furthermore, the presented, compared to a modern internal combustion engine modern design, greatly simplified configuration of the internal combustion engine not restrictive effect with respect to the configuration and function of the subsystems such as the direct injection fuel system, the variable valve train, the ignition device, the exhaust gas turbocharger and the intake system and the exhaust system.

LIST OF REFERENCE NUMBERS

1

 internal combustion engine

2

 turbocharger

3

 Direct fuel injection system

4

 detonator

5

 Variable valve train

6

 Electronic control device

7

 intake system

8th

 exhaust system

11

 cylinder

12

 Crankshaft rotational speed sensor

15

 Driver interface

21

 compressor

22

 exhaust turbine

23

 Turbocharger rotor speed sensor

31

 Common rail

32

 Injector

33

 Fuel supply

34

 Fuel pressure sensor

51

 intake valve

52

 Exhaust outlet

53

 camshaft actuator

54

 camshaft

71

 Fresh air distribution

72

 throttle

73

 Air filtration system

74

 Charge-air pressure sensor

75

 Air mass sensor

81

 exhaust manifold

82

 aftertreatment system

83

 Exhaust pressure sensor

84

 Exhaust gas temperature sensor

85

 Exhaust pressure sensor

86

 Catalyst temperature sensor

ECU

 electronic control device

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 102011081844 A1 [0015]
• DE 10140120 A1 [0016]
• DE 19944190 A1 [0017]

Claims (9)

Method for controlling one, a plurality of cylinders ( 11 ) having an internal combustion engine ( 1 ), which has an exhaust gas turbocharger ( 2 ), wherein with an increase in the power requirement to the internal combustion engine ( 1 ) a speed increase of the turbocharger ( 2 ) is supported by an increase in the exhaust gas temperature, characterized in that - in dependence on the power requirement and the respective current operating point of the exhaust gas turbocharger ( 2 ), is determined by means of an exhaust gas turbocharger calculation model, a current target exhaust gas temperature (AT_S), - starting from the determined target exhaust gas temperature (AT_S) by means of an exhaust gas temperature computing model, the required to achieve the target exhaust gas temperature (AT_S) required control value ( StW_S) at least one control parameter influencing the exhaust gas temperature is determined and a default control value (StW_VG) is determined from the desired control value (StW_S) including at least one limit control value (StW_GR, StW_GR_BG, StW_GR_TQ) and the determined default control value (StW_VG ) of the at least one exhaust gas temperature influencing control parameter for controlling the internal combustion engine ( 1 ) is used to increase the exhaust gas temperature, wherein the combustion center position is used as a control parameter and the combustion center position is used only for a part number of the total existing cylinders ( 11 ) of the internal combustion engine ( 1 ) is shifted with respect to the combustion center position at the time before the increase in the power demand at a later time. Method for controlling an internal combustion engine ( 1 ) according to claim 1, characterized in that in an internal combustion engine ( 11 ) with an even number of cylinders, the combustion center position is shifted at half the number of cylinders. Method for controlling an internal combustion engine ( 1 ) according to claim 1 or 2, characterized in that the at least one limit control value (StW_GR_BG) by means of a combustion boundary computing model for the current operating point of the internal combustion engine ( 1 ) is determined. Method for controlling an internal combustion engine ( 1 ) according to one of claims 1 to 3, characterized in that the at least one limit control value (StW_GR_TQ) by means of a torque calculation model for the current operating point of the internal combustion engine ( 1 ) is determined. Method for controlling an internal combustion engine ( 1 ) according to one of claims 1 to 4, characterized in that the determined target exhaust gas temperature (AT_S) is limited to a predetermined limit exhaust gas temperature (AT_GR). Method for controlling an internal combustion engine ( 1 ) according to one of claims 1 to 5, characterized in that the focal point of combustion is carried out by shifting the ignition timing to a later than a normal ignition timing later. Method for controlling an internal combustion engine ( 1 ) according to one of claims 1 to 6, characterized in that the internal combustion engine ( 1 ) with a direct injection system ( 3 ) for direct fuel injection into e combustion chambers of the plurality of cylinders ( 11 ) and the combustion center position is performed by changing the injection timing and / or the injection mass. Method for controlling an internal combustion engine according to one of Claims 1 to 6, characterized in that the method is carried out repeatedly until a desired boost pressure in the intake system ( 7 ) is reached or the internal combustion engine ( 1 ) has reached a requested performance level or the increased performance requirement is being withdrawn. Device for controlling one, a plurality of cylinders ( 11 ) having an internal combustion engine ( 1 ), which has an exhaust gas turbocharger ( 2 ), wherein with an increase in the power requirement to the internal combustion engine ( 1 ) a speed increase of the turbocharger ( 2 ) is supported by an increase in the exhaust gas temperature, wherein the apparatus is adapted to perform a method according to any one of claims 1-8.

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