HK1221275B - Method and device for operating a gas internal combustion engine - Google Patents

Description

Method and device for operating a gas internal combustion engine

Technical Field

The present invention relates to a method for operating a gas internal combustion engine having a gas mixer, an intake section, and an engine with a plurality of cylinders. The present invention also relates to a control unit for an internal combustion engine according to the invention and an internal combustion engine according to the invention. The gas internal combustion engine further comprises a supercharging section in the intake section and a bypass section for bypassing the supercharging section. The gas internal combustion engine is, in particular, a spark-ignition gas internal combustion engine.

Background Art

Fields of application for gas internal combustion engines of this type are mobile applications (for example in the marine sector or in the commercial and heavy vehicle sector) and stationary applications such as in block power plants, which are particularly advantageously designed for fluctuating gas supplies.

Document 6,131,552 generally discloses a fuel control system that can regulate the gas supply to the mixing chamber based on the measured engine operating state. The control method of document US Pat. No. 6,131,552 A or other gas distributions that are solely load-dependent have also proven to be insufficient in complex control systems.

Air consumption is generally a measure of the gaseous fresh charge of the charge air in the charge mixture that is supplied to the internal combustion engine. Air consumption also allows for the quality of the intake system and intake process to be accounted for. Actual air consumption generally expresses the proportion of the fresh air mass actually supplied to the engine or its cylinders during a working cycle in the charge mixture. This actual mixture mass is determined relative to the theoretical fresh charge mass from the geometric displacement and the theoretical charge density in ambient conditions (for freely aspirated engines) or in supercharged engines, taking into account the state of the fresh charge downstream of the compressor or downstream of the compressed air cooler.

A number of factors influence the fresh charge delivered to the cylinders, such as the valve opening cross-section or valve control time. In principle, this can be determined by the module for determining the engine boost pressure, in which the intake section model is stored. In practice, however, the fresh charge delivered to the engine in the charge mixture only corresponds to the target value in exceptional cases. Air consumption is not a constant value for the engine, but rather depends significantly on the engine speed and the current geometry of the intake system and combustion chamber. To understand this relationship, suitable characteristic maps can be used, for example.

For example, EP 1 398 490 A2 discloses intake section models in engine controllers for conventional internal combustion engines. These models share the broad basic concept of designing the intake section model (in the simplest case, as a uniform pressure vessel to capture the dynamic processes in the air path) by designing the storage characteristics of the intake section (also known as the intake pipe) through the filling and emptying process. The intake pipe is treated as a pressure vessel that is continuously filled with air via a throttle valve, and the engine draws air from it via the intake valve according to the power stroke, using its intake characteristics.

However, it has been shown that the fuel supply of a gas internal combustion engine is considerably more complex, particularly in the transient operating range of the internal combustion engine and with variable fuel qualities. In particular, it has been shown that in gas internal combustion engines, particularly those designed for spark ignition, operation in the low load range and/or in the transient load range can be problematic.

During operation of a gas engine, mixture formation typically occurs upstream of the exhaust gas turbocharger's compressor. The intake section between the compressor outlet and the combustion chamber inlet also includes a partially large volume, which stores or emits a significant mixture mass. This is particularly true when changes in engine load and/or speed result in pressure and/or temperature changes in the various sub-volumes. Due to mixture mass formation in gas internal combustion engines, which is only conditionally adapted to the operating point (because it is imprecise, particularly due to the partially large volumes), high hydrocarbon (HC) emissions or other increased emissions (NOx, CO, particulates, etc.) can be expected, as well as poor efficiency due to unburned combustion gases.

It is desirable to design the gas operation of a gas internal combustion engine more advantageously, particularly in transient operation, in accordance with load requirements and emission conditions. This is particularly desirable at least in the low-load range. It is particularly desirable to achieve this in the full load range, that is, preferably up to 100%.

Summary of the Invention

In this regard, the present invention has the object of specifying a method and a device by means of which an improved operation of an internal combustion engine as a gas internal combustion engine is achieved. The present invention is particularly aimed at achieving improvements in transient operation and/or improvements in the area of emissions. Preferably, the fuel composition problems existing in the low-load range should be addressed in an improved manner, in particular while avoiding sudden changes in the engine's torque in otherwise transient operation, also when the load requirement changes and, if possible, while avoiding excessive emissions. In particular, the mixture mass formation should be designed as advantageously as possible. In particular, the engine's air consumption should be designed in an improved manner. The present invention is particularly aimed at solving at least one of the above-mentioned problems. At least an alternative solution should be proposed.

The object relating to the method is achieved by the invention by a method according to the invention.

In particular, a method for operating a gas internal combustion engine of the type mentioned at the outset is used here, which has a gas mixer, an intake section and an engine with a plurality of cylinders, wherein in the method

A fuel mixture, including a charge mixture, is supplied to the engine and the engine is operated in gas operation using the gas in the charge mixture as fuel.

Furthermore, it is provided that an output mixture portion of the gas-air mixture associated with a subsequent mixture state is determined with the aid of an input mixture portion of the gas-air mixture associated with at least one previous mixture state, and that this determination is carried out with the aid of an intake section model serving as a basis for a calculation model for the intake section.

According to the present invention, an output mixture fraction of the gas-air mixture at the engine supply is determined, and an input mixture fraction is determined from the output mixture fraction in a plurality of associated volumes of the intake section via a plurality of intermediate states of the mixture fraction. To this end, the input mixture fraction of the gas-air mixture at the gas mixer is determined, and the air flow and/or gas flow at the gas mixer is adjusted as a function of the input mixture fraction. Advantageously, the air flow and/or gas flow at the gas mixer can be adjusted as a function of the input mixture fraction, in particular the mixture mass flow, within the scope of a simultaneous real-time calculation.

The present invention is based on the idea that a favorable transient operation of a gas internal combustion engine should be achieved while taking into account the state pressure of the intake section, in particular while taking into account the throttling of the charge air or fuel mixture, and/or while taking into account the air consumption as much as possible, taking into account the characteristics of gas engines. Although intake section models in engine controllers are generally known, the present invention is based on the idea that the aforementioned modeling is generally insufficient for applications in gas engines, in particular with mixture formation upstream of the compressor.

The present invention primarily considers the fact that in gas engines, the path between the location where the mixture is formed, for example, upstream of the compressor of the exhaust gas turbocharger and the location of the combustion chamber inlet, can be relatively long and thus encompass a relatively large volume, which can hardly be realistically described in a static state. The present invention also primarily considers that in gas engines, the intake path between the compressor outlet and the combustion chamber inlet consists of partially large volumes (which store or emit a significant mixture mass when changes in the engine load and/or speed result in changes in the pressure and/or temperature in the individual subvolumes).

The present invention recognizes that, in gas engines in particular, the mixture mass flow at the outlet of the gas mixer is temporarily decoupled from that at the combustion chamber inlet. Previously, this situation was not adequately accounted for in intake section models, particularly for gas engines. Consequently, as the present invention demonstrates, conventional static considerations of gas engines (in which only the mixture pressure upstream of the cylinder is considered) are insufficient to ensure that a defined combustion air ratio is maintained. The concept of the present invention therefore provides a mixture mass sensor with improved reliability and better adapted to transient operation, in particular a virtual mixture mass sensor for mixture formation upstream of the compressor.

The object relating to the device is achieved with the invention by a control unit according to the invention for a gas internal combustion engine.

The solution according to the invention also achieves this object with respect to an arrangement according to the invention for a gas internal combustion engine.

The internal combustion engine can in particular have an engine with a plurality of cylinders and an intake system with a gas mixer and an intake section. In addition, it has proven advantageous to place a receiver volume (which can be implemented, for example, in the form of a bend or mixing section) upstream of the cylinders before the plurality of cylinders.

It has proven particularly advantageous to provide a gas internal combustion engine with a supercharging section in the intake system, in particular a supercharging section including a charge heat exchanger. Depending on the size of the gas internal combustion engine, in particular for large engines, the supercharging section can be single-stage or two-stage, preferably with an exhaust gas recirculation section. In particular, a bypass section to the intake section of the intake system can also be provided to bypass the supercharging section.

As demonstrated in cycle simulations and tests, particularly on gas engines, the present invention improves the stability of the combustion air ratio, particularly during transient engine operation (on-off). This allows for larger load increases to be displayed, while also making load switching operations easier to coordinate on the test bench. Furthermore, the use of virtual gas sensors according to the present invention contributes to compliance with current and future exhaust gas standards.

These and other advantageous refinements of the invention are apparent from the present invention and advantageous possibilities for realizing the concept of the invention are explained in detail within the scope of the refinements and with a description of further advantages.

In a particularly preferred embodiment, it is essentially ensured that the air and/or gas flow at the gas mixer is adjusted as a function of the input mixture proportions within the scope of simultaneous real-time determination. In particular, a mixture mass sensor designed as a virtual mixture mass sensor is shown, which advantageously provides the mixture mass flow at the gas mixer outlet; this is advantageous compared to previous approaches.

In particular, an input mixture portion of the gas-air mixture at the outlet of the gas mixer and/or an output mixture portion of the gas-air mixture at a cylinder or cylinder inlet or receiver of the engine can be determined.

It is particularly advantageous to determine the input mixture proportion of the gas-air mixture at the outlet of the gas mixer and/or the output mixture proportion of the gas-air mixture at the cylinder or cylinder inlet of the engine, in particular at the receiver. It is particularly advantageous to determine the state pressure upstream of the cylinder of the engine, preferably as the receiver pressure in the receiver volume, where the receiver volume can be understood to mean any type of volume upstream of the cylinder and downstream of the supercharging unit and/or bypass section. For example, the receiver volume can be the volume of a bend or other structural space expansion of the intake section. The receiver volume can be understood in particular to mean a volume that exceeds the normal volume of the intake section; within the scope of the improved solution, it is shown that the receiver pressure preset in the receiver volume is particularly important for reliable regulation of gas internal combustion engines, because an increase in the uncertainty of the state of the combustion gas mixture is associated with an increase in the size of the receiver volume. The regulation of the receiver pressure in the receiver volume thus eliminates the uncertainty inherent in the static assumption of the intake section.

Preferably, the determination of the mixture fraction includes determining the mixture mass flow, in particular using flow equations for components and/or determined volumes of the throttle mechanism and/or the intake section. In particular, the inlet mixture fraction at the gas mixer is correlated with the mixture mass flow and/or the outlet mixture fraction is correlated with the mixture mass flow. In particular, the mixture mass flow at the throttle can be determined, wherein the flow rate of the return flow and/or the boost flow is determined. This can preferably be achieved using flow equations for compressible media at an ideal or actual nozzle, assuming an ideal or actual frictionless or frictional flow of the gas.

Preferably, the determination of the mixture fraction additionally or alternatively includes the determination of the mixture state, in particular the temperature and/or state pressure of the mixture fraction at least for the volume of the intake section. Preferably, this is achieved with the aid of a thermodynamic state equation of the actual or ideal gas.

It is advantageous to consider the state pressure in the suction section, for example in a large volume, or as a pressure loss due to throttling. Its actuator can be provided to influence the state pressure, for example at a throttle flap, a throttle valve or other throttling mechanism. Here, the throttling mechanism of the suction section is considered to be any device for reducing or generally regulating the pressure, and in addition to the engine throttle, the compressor bypass throttle may also belong to this category if necessary. The engine throttle can be, in particular, a valve, a flap or a throttle or also the variable turbine geometry of the compressor. Here, the adjustment angle α between the fully open and fully closed positions of the throttle is usually used to describe the throttle position of such a throttle mechanism; multiple throttle mechanisms of the above type can also be used independently of each other or in a combination of different throttle mechanisms or in a coordinated manner.

In particular, an engine throttle can be provided upstream of the receiver volume and/or a compressor bypass throttle can be provided in the bypass section. In particular, the suction section can be throttled, in particular for the engine and/or the bypass section, depending on the setpoint and/or actual state pressure of the suction section.

In a particularly preferred embodiment, a development provides that the intake section between the gas mixer and the engine is divided into a plurality of volumes, in particular at least two volumes, preferably exactly two large volumes.

This approach has proven suitable, in which, in particular, one or more of the following steps are used:

- filling and emptying methods (especially when using information such as pressure and temperature at existing measuring points),

- real-time calculation of the mixture mass flow at different points in the intake section in the engine controller,

- The same mixer is mixed via a gas metering cell, which results in a virtually defined mixture quality with a desired combustion air ratio.

Advantageously, the input mixture proportions are linked to the combustion air ratio, in particular its setpoint value (LAMBDA_SOLL), wherein the gas metering unit of the gas mixer is guided by the stoichiometric air requirement (L_st) and/or the combustion air ratio. This measure has proven suitable for controlling the gas mixer.

In the volume of the suction section, in particular, intermediate mixture portions, in particular mixture mass flows and/or mixture states can be determined for a plurality of theoretically calculated suction section volumes and/or for a plurality of actual suction section housing volumes, wherein the number of intermediate mixture portions in the suction section is correlated with the number of at least one large suction section volume. This allows the suction section model to be designed in a particularly realistic manner.

Preferably, the state pressure is determined virtually, in particular simulated and/or calculated, based on a calculation model of the intake section (which includes at least the calculated volume of the charge heat exchanger and/or the receiver volume). In a particularly preferred development, the state pressure is determined as the receiver pressure in the receiver volume upstream of the cylinder of the engine (which is upstream of the cylinder and downstream of the supercharging section and/or the bypass section). In principle, the charge heat exchanger volume can be understood as any type of volume with which temperature is exchanged; this can in particular include the volume of the engine supply section and/or the intake section of the intake system. The above-mentioned volumes have proven to be particularly relevant for the description of the intake section.

Preferably, the plurality of large volumes of the suction section include one or more component volumes of the suction section, which are selected from: at least one, preferably two, receiver volumes, in particular at least one cylinder volume in an engine block, at least one charge heat exchanger volume, at least one compressor volume.

Further preferably, the plurality of at least one large volumes of the suction section include one or more component volumes of the suction section, which are further selected from: at least one compressor bypass volume, in particular at a bypass line section and/or a compressor bypass valve; at least one suction section volume, in particular at a suction line section and/or an engine and/or an inlet throttle valve.

Preferably, the gas operation is spark-ignited gas operation; this has proven to be particularly efficient and suitable for many applications. However, other ignition principles, such as ignition systems for diesel or other liquid fuels, are also generally applicable. Alternatively, the gas operation can also be ignition jet operation with external mixture formation of the gas-air mixture, and in this case, a diesel ignition jet or another liquid fuel ignition jet is used.

Although not specifically mentioned above, due to the relatively constant fuel quality, gas-fired internal combustion engines can be operated with either natural gas or liquid fuels, such as diesel or liquefied natural gas. During operation, a fuel mixture (including the charge mixture and/or the liquid fuel) is supplied to the engine. In this case, the engine can operate with diesel fuel (using diesel or another liquid fuel) in a first operating state and with natural gas fuel (using natural gas as the fuel in the supercharged mixture) in a second operating state. This type of gas-fired internal combustion engine is also known as a multi-fuel internal combustion engine (dual-fuel internal combustion engine) and, in addition to the preferred fuel options of diesel and natural gas, can also operate with a wide variety of other fuels. Alternatively, particularly in gas-fired operation, the gas-fired internal combustion engine can operate in jet-ignition operation with external mixture formation of the gas-air mixture and a diesel jet-ignition operation. Thus, gas-fired internal combustion engines are also known as jet-ignition engines and are typically based on diesel engine design and are state-of-the-art, particularly within the scope of environmentally friendly applications for large engines. Spark-ignition jet engines can also be operated with liquid fuels such as diesel or other liquefied fuels such as liquefied natural gas (LNG) or liquefied petroleum gas (LPG); gas internal combustion engines can usually have a gas-diesel engine to form a gas-diesel internal combustion engine.

Gas internal combustion engines, in particular, have an injection system that is preferably electronically adjustable. For this purpose, the internal combustion engine can also include an injection system that is advantageously designed as a common rail injection system. The injection system can be adjusted, in particular, for different gas qualities, such as biogas or natural gas in liquid form, or even for oils such as vegetable oils used as liquid fuels. Common rail injection systems, but also pump-nozzle-injection systems with electronic regulation, if necessary, have proven suitable. In gas operation, the ignition medium can be added to the actual gaseous fuel charge mixture in the cylinder under high compression, or also to the intake duct. Gas engines operating with external mixture formation in gas operation, in particular spark-ignited gas operation or ignition jet operation, are generally more flexible in fuel usage and have lower emissions.

BRIEF DESCRIPTION OF THE DRAWINGS

Now, embodiments of the present invention will be described below in comparison with the prior art, which is also partially shown, based on the accompanying drawings. These embodiments are not necessarily to be shown to scale, but rather the drawings are shown in schematic and/or slightly distorted form for the purpose of illustration. In addition to the teachings that can be directly identified from the drawings, reference is made to the relevant prior art. It should be taken into account that various modifications and changes can be made in the form and details of the embodiments without departing from the general idea of the invention. The features of the invention disclosed in the description and in the drawings are important for improving the invention not only individually but also in any combination. In addition, all combinations consisting of at least two of the features disclosed in the description and/or in the drawings fall within the scope of the present invention. The general idea of the invention is not limited to the precise form or details of the preferred embodiments shown and described below, nor to objects that are limited compared to the claimed object. In the described size ranges, values within the limits should be disclosed as limit values and can be applied arbitrarily and claimed. Further advantages, features and details of the invention can be obtained from the following description of the preferred embodiments and from the drawings; among them:

FIG1 shows a schematic diagram of a gas internal combustion engine having a gas mixer and an intake section with a supercharging section consisting of a turbocharger and a charge heat exchanger, and an engine with a plurality of cylinders downstream of a receiver volume, wherein the supercharging section can be bypassed via a bypass section, the gas internal combustion engine being designed for spark-ignited gas operation;

In an alternative shown by dashed lines, the gas internal combustion engine can also be designed as a gas-diesel internal combustion engine and can be operated with pure diesel as well as in mixed operation or pure gas operation (for example, with an ignition mixture in the form of diesel injected as an ignition jet), wherein the injection system is formed in the form of a common rail system shown by dashed lines;

2 shows a flow chart of a preferred embodiment of a method for determining an input mixture portion of a gas-air mixture associated with a previous mixture state by means of an output mixture portion of a gas-air mixture associated with a subsequent mixture state within the scope of simultaneous real-time determination, wherein the input mixture portion of the gas-air mixture is determined at a gas mixer and the air flow and/or gas flow at the gas mixer is adjusted as a function of the input mixture portion.

3 shows a schematic diagram of a preferred embodiment of a regulator structure for dual-fuel operation, wherein the input mixture portion of the gas-air mixture of the gas mixer is determined from the output mixture portion of the gas-air mixture supplied to the engine in the intake section by means of an intake section model via a plurality of intermediate states of the mixture portions.

DETAILED DESCRIPTION

1 shows a gas internal combustion engine 100 having an engine 10 and an intake system with a branched intake section 30. In particular, a gas mixer 40 and (to form a supercharging section) a turbocharger 50 and a charge heat exchanger 60, here in the form of a charge air cooler, as well as a bypass 70 are arranged in the intake section.

Here, the engine is designed as a V-engine with 16 cylinders, with 8 cylinders Ai,i=1...8 on the A side and 8 cylinders Bi,i=1...8 on the B side. This type of cylinder arrangement and number is shown here only as an example. Engine designs with 10, 12, 20, 24, or 28 cylinders, or other numbers of cylinders, are also particularly suitable for large engine applications.

In an alternative or additional design as a dual-fuel internal combustion engine, the internal combustion engine further comprises an injection system 20 (shown in dashed lines). This is designed here as a common rail system with a common rail 21, from which multiple injection lines 22 (each having an injector 23 and a separate reservoir 24 upstream of the injector) branch off to the individual cylinders Ai, Bi, i=1...8 of the engine 10. The injection system 20 is designed to fractionate a liquid fuel, such as diesel or another liquefied or liquid fuel, so that the fuel is injected as a liquid fuel in diesel operation or as an ignition jet in gas or ignition jet operation at the beginning of the working cycle of each cylinder Ai, Bi. This occurs at a very high injection pressure. Accordingly, in this variant, the engine 10 further comprises a common rail injection system 20 for liquid fuel, in this case diesel fuel, as well as a supercharger 50 with a charge heat exchanger 60 and a bypass 70 for bypassing the supercharger 50 and the charge heat exchanger 60.

Referring to the important parts of this embodiment shown in solid lines, a gas mixer 40 connected to the intake section 30 at the inlet-side end of the intake system draws charge air LL from the environment and mixes combustion gas BG therein. In gas operation, the charge mixture (hereinafter also referred to as "mixture G"), which can also be referred to as a combustion gas mixture, is supplied at a mass flow rate m(')_G ("(')" is shown in the figure as a point on mass m or other parameters for clarity) via the compressor section 32 at an intake pressure p1 and an intake temperature T1 approximately corresponding to the ambient temperature to the compressor 51 of the turbocharger 50, where it is compressed at a compression temperature T2 to a compression pressure p2. Compressor 51 is driven by a turbine 52 and lies together with it on a common turbocharger axis 53; turbine 52 of the exhaust system 90 is in turn driven by exhaust gas AG leaving the engine 10 in the exhaust system 90. The mass flow m(′)_G of the mixture G, which has been heated to the compression temperature T2 due to the compression, is supplied to the cooling section 31 of the intake section 30 and there conducted in the charge heat exchanger 60 through the cooler structure 61; in the heat exchanger volume 62, shown here as a symbol, heat exchange with the coolant in the cooler structure 61 takes place, thereby cooling the mixture G. The combustion gas mixture leaves the heat exchanger volume of size V3 in cooled form at the charge temperature T3 and the charge pressure p3 in the direction of the charging section 33 for supplying the mixture G to the engine 10.

In the intake section model, the state of the mixture G before compressor 51 is typically described using pressure and temperature state variables, here intake air temperature T1 and intake air pressure p1 before compressor 51, or using a suitable compressor model to describe the state after compressor 51 at an increased compressor pressure p2 and increased compressor temperature T2 with state variables p2, T2 after compressor 51. This is derived, for example, from a gas state equation for an ideal or real gas. The large volume of the intake section 30 is particularly important in conjunction with components of the heat exchanger 60 and receiver 80, such as the elbow and/or the collecting section. Therefore, according to the present invention, this and other spaces in the intake section are associated with heat exchanger volume V3 or receiver volume V5 in the intake section model for modeling additional gas states. Accordingly, the combustion gas mixture G in heat exchanger volume V3 exhibits state variables p3, T3, which, due to cooling and volume increase, reduces the charge pressure and charge temperature p3, T3.

The state of the mixture G in the bypass 70 is essentially also determined by the state parameters p1, T1 at the inlet of the bypass 70 or the state parameters p3, T3 at the outlet, or vice versa in the case of return flow through the bypass 70; that is, the bypass mixture G_BP in the bypass section 71 of the bypass 70 depends on the existing pressure conditions and the position of the compressor bypass throttle 72 (here according to the adjustment angle αVBP of the compressor bypass valve), and the bypass section 71 can be used in particular to guide the excess mixture G back in front of the compressor 51 so that it can be recompressed and delivered to the cylinders Ai, Bi of the engine 10 for combustion.

Before the gas mixture G in state p3, T3 is delivered to engine 10, it is introduced into receiver 80 at a pressure and temperature that varies according to mass flow m(')_DK introduced into receiver volume 81 via engine throttle 82 (and according to receiver volume V5 at receiver pressure p5 and receiver temperature T5). First and second receiver volumes 81.B and 81.A are associated with the B-side and A-side of engine 10, respectively. They are situated upstream before cylinders Ai, Bi and after first and second supercharging sections 33.B and 33.A on the B-side and A-side, respectively, and after heat exchanger volume 62. Engine throttle 82 is formed by first and second engine throttle flaps 82.B and 82.A, respectively, associated with first and second receiver volumes 81.B and 81.A. These first and second engine throttle flaps 82.B and 82.A are adjustable independently of one another; they are collectively referred to below as engine throttle 82 for simplicity. Receptacle volume 81 can be understood as the sum of first and second reservoir volumes 81.A and 81.B. In reservoir volume 81, mixture G, due to the volume increase and depending on the position αDK of engine throttle flaps 82.A, 82.B, assumes a gaseous state characterized by p5 and T5 in volume V5 of reservoir volume 81. This is correlated with the B-side or A-side mass flow rate m(')_DK,B or m(')_DK,A, depending on the position of engine throttle flaps 82.B and 82.A.

In this embodiment, the state of the gas mixture G, characterized by pi, Ti, i = 1, 2 or Vj, pj, Tj, j = 3, 5, is determined primarily in the region provided by the compressor 51, the heat exchanger volume 62 and the receiver volume 81, specifically relative to the limits provided by the engine throttle 82 and the compressor bypass throttle 72 or the compressor 51.

Below, based on the intake section model of the gas internal combustion engine 100 shown here, for the receiver pressure p5 in the receiver volume V5 or for the control parameters established based on the receiver pressure p5 (for example, the actual receiver pressure p5_IST or the theoretical receiver pressure p5_SOLL or the simulated receiver pressure p5), the input-mixture proportion of the gas-air mixture at the gas mixer 40 is determined by multiple intermediate states of the mixture proportion in multiple associated volumes of the intake section 30.

It is shown that the adjustment of the mass flow m(')_G of the combustion gas BG and the mass flow m(')_LL of the charge air LL at the gas mixer 40 to the combustion air ratio LAMBDA_SOLL or the stoichiometric air ratio Lst may not necessarily be achieved along the intake section under static conditions. Therefore, the scheme of this embodiment considers at least two large volumes, namely the receiver volume 80 and the charge heat exchanger volume 62, in the intake section model (as described with reference to FIG. 1 ) for the comprehensive volume of the intake section. Within the scope of the intake section model, the intake section 30 is modeled according to the principle of the filling and emptying method (known in principle). The state changes in these volumes are considered to be approximately isothermal. This simplifies the system by limiting it to mass conservation relative to an adiabatic approach and in particular simplifies the simultaneous calculation of the internal combustion engine or its intake section in real time. However, in principle, an adiabatic or variable approach or targeted heat transfer can also be used to simulate the state changes in the intake section when computing power is sufficient.

As can also be seen from FIG. 3 , special assumptions for setting up the intake section can be implemented within the scope of additional models, particularly if correspondingly set-up measured values for the intake section 30 are not available. This involves, for example, an additional model for the compressor (the p2_T2 module, which describes the activity of the compressor 51 and describes the state of the mixture G upstream of the compressor via temperature and pressure (G(p1, T1)) and the state downstream of the compressor (G(p2, T2)). This also involves, for example, an additional model for the compressor bypass valve (mass flow_VBP), which describes the mass flow through the compressor bypass valve within the scope of the flow equation.

During operation, a fuel mixture, including a charge mixture, is supplied to the engine 10. The engine is operated using natural gas as fuel in the supercharged mixture. Based on an intake section model (as described in FIG. 3 using the structure of controller 200), at least one intake section (mass flow m(')_G,SOLL) of the gas-air mixture associated with a subsequent mixture state (with state parameters p5, T5) is determined using an output section (mass flow m(')_G,ZYL) of the gas-air mixture associated with a subsequent mixture state (with state parameters p5, T5). This determination is performed using the intake section model as the basis for a calculation model for the intake section 30, the gas mixer 40, and the receiver 80 or the inlet to the engine 10.

According to this embodiment, the output mixture portion (mass flow m(')_G,ZYL) of the gas-air mixture at the engine supply (here, at the receiver 80) is determined. In this case, the input mixture portion (mass flow m(')_G,SOLL) is determined from the output mixture portion (mass flow m(')_G,ZYL) by a simultaneous real-time calculation using a plurality of intermediate states of the mixture portion (with state parameters pi,Ti,i=5,3,2,1). The number of intermediate states (with state parameters pi,Ti,i=5,3,2,1) is correlated with the volume (Vi,i=5,3,2,1) or the number of components E1, E2, E3, E4, and C3 and C5 in the intake section. The input mixture portion of the gas-air mixture is determined at the gas mixer 40, and the air flow and/or gas flow BG at the gas mixer 40 is adjusted as a function of the input mixture portion.

Here, in the suction section model, multiple large volumes are associated with the suction section; they include: two receiver volumes 81.B, 81.A (with an associated volume V5 and with state parameters p5, T5 of the charge mixture therein), at least one cylinder volume in the engine group, at least one charge heat exchanger volume 62 (with an associated volume V3 and with state parameters p3, T3 of the charge mixture therein), at least one compressor volume at the compressor 51 (with value V2 and with state parameters p2, T2 of the charge mixture at its outlet) or the state before the compressor 51 (with state parameters p1, T1 of the charge mixture and with state parameters p0, T0 of the environment (atmosphere) of the suction air), as shown in Figure 1.

In the present intake section model, the components E1, E2, E3, E4 of the intake system, as well as C3 and C5, which are also shown in FIG1 , are taken into account, and the state or mass flow of the charge mixture is simulated or calculated in this respect. This is merely a preferred example selected from a range of other possible examples; for example, in the case of two-stage supercharging, a modified intake section model can also take into account further components. In principle, further volumes can also be taken into account in the intake section model. In the extreme case, the intake section can be divided into finitely small or infinitely small sub-volumes and the determination of the state and mass flow of the charge mixture is then obtained as the solution of a system of difference or differential equations; a balance can be taken in refining this model, taking into account the computational costs and the complexity of the intake section to be simulated; in particular, taking into account the real-time capabilities.

FIG2 , as a flow chart, schematically illustrates the flow of a preferred embodiment of an operating method for a gas internal combustion engine 100 having a gas mixer 40 , an intake section 30 , and an engine 10 with a plurality of cylinders Ai, Bi, i=1..8, also referred to herein as a gas engine, as described with reference to FIG1 . In principle, the operating method can be implemented by implementing the method shown in FIG2 for determining control parameters at the gas mixer 40 , by means of which the air flow LL and/or the gas flow BG can be adjusted. Referring to FIG2 , the method, based on the intake section model shown here, includes further steps S1 through S9 in step S0 , which are located at the aforementioned components E1, E2, E3, E4, as well as C3 and C5 of the intake system (which are correspondingly shown in FIG1 ) determined by the intake section model.

Specifically, the input mixture fraction of element E0 (here associated with gas mixer 40) is determined "backwards," that is, by back-calculating from the known output mixture fraction at element E4 in the intake section. Element E4 corresponds approximately to engine 10 or cylinders Ai, Bi, i=1..8 of engine 10. In a first step S1 of this determination method, a calculation module in control unit R4 uses corresponding characteristic parameters for engine speed nMOT, receiver pressure p5, and receiver temperature T5 to determine the cylinder filling, for example, a measure of the required air consumption LAMBDA_a at engine 10. Then, in a second step S2 of the determination method, a mixture mass flow m(')_G,ZYL can be predefined at element E4 of the intake system (here, for example, at the cylinder inlets of cylinders Ai, Bi, i=1..8).

3 shows the controller structure of a controller 200, which has a first control unit R4 for carrying out steps S1 and S2. The input variables are the engine speed nMOT and the cylinder inlet pressure (in particular, the receiver pressure p5 in the receiver 80 or the receiver volumes 81.B, 81.A with a total volume V5) and, if necessary, other associated gas state variables of the mixture fraction in the receiver 80 (with a receiver volume V5).

Then, in the third step S3 of FIG. 2 , as a further component E1 of the intake system, the receiver volumes 81.B, 81.A with the total volume V5 of the receiver 80 are taken into account (for example, associated with a bend or mixing section in the intake section between the cylinder inlet and the engine throttle valve DK).

Given the aforementioned known pressure and temperature conditions (receiver pressure p5, receiver temperature T5) in receiver volume 81.A, 81.B V5, the mixture mass flow m(')_RECEIVER through the receiver can be determined in a fourth step S4 using control unit R5 shown in FIG3. Control unit R5 includes a calculation unit that determines the mixture mass flow m(')_RECEIVER from the mixture state of the gas-air mixture in receiver 80 (often also referred to as a receiver tube) (i.e., the state parameters p5, T5 of the charge mixture in receiver volume 81.B, 81.A having volume V5) using a charge mixture state equation and a mass flow equation according to the filling and emptying method for receiver tubes 81.B, 81.A.

Furthermore, a preferred embodiment of the determination method in FIG2 provides a charge heat exchanger 60 for a further component E2 based on the suction section model. In the associated volume V3 of the heat exchanger volume 62, the charge mixture has a gaseous state, which can be calculated in a fifth step S5 by means of the charge pressure p3 and the charge temperature T3 of the volume V3. Similarly, according to the determination method shown here, the mixture mass flow m(')_LLK through the charge heat exchanger 62 is determined. Correspondingly, FIG3 shows a regulating unit R3 for determining the mass flow m(')_LLK through the charge heat exchanger 60 in the form of a charge air cooler having a cooler structure 61. This determination is carried out in a sixth step S6 by means of a filling and emptying method for the volume V3 with a known mixture state or its thermodynamic characteristic parameters and the corresponding mixture mass flow equation.

In a seventh step S7, the target value LAMBDA_SOLL for the combustion air ratio and the stoichiometric air requirement (Lst) can then be derived from the mixture mass flows m(')_G,ZYL, m(')_RECEIVER, and m(')_LLK, that is, the important target characteristic parameters of the fuel at another component E0 of the gas mixer 40 in the intake section 30. To this end, a first summing element R45 and a second summing element R43 are provided in the controller structure shown in FIG. 3 to determine the mixture mass flow m(')_G,SOLL from the mixture mass flow equation; this is achieved by taking into account a further controller unit R0 for implementing the seventh step S7, i.e., for calculating the combustion air ratio LAMBDA_SOLL and the stoichiometric air requirement Lst.

In this way, suitable control parameters can then be predefined at the gas mixer 40 in an eighth step S8 in order to adjust the input mixture proportions at the gas mixer 40 as a function of the current engine speed nMOT and the pressure and temperature conditions at the cylinder inlet (approximately p5, T5).

Here, the intake section model is designed with two large volumes (i.e., receiver volumes 81.B, 81.A (V5) and charge heat exchanger volume 62 (V3)). Taking into account the filling/emptying methods for both volumes, the intake section model is incorporated into the simulation or calculation of the intake section 30, along with corresponding flow equations for at least the throttle valves 82, 72 and a compressor model for the compressor 51 (with state parameters p1, T1 → p2, T2). This is sufficient to overcome the aforementioned problem; large volumes (which were not adequately accounted for in previously known calculation methods and thus resulted in a decoupling of the gas mass flow at the gas mixer from the gas mass flow actually present at the engine) are fully accounted for in the intake section model of the determination method shown in FIG. It also turns out that the determination method in the form shown here can be implemented relatively simply and, therefore, computationally efficiently, so that it can be used within the scope of simultaneous real-time determination. This allows the gas mixer 40 to be adjusted to the current requirements of the engine 10 in practically real time.

In a refinement of the determination method shown in FIG. 2 , FIG. 3 shows the control unit R3 and the feedback of the mass flows m(′)_DK, m(′)_RECEIVER and m(′)_LLK in detail. This makes sense for the case where a bypass 70 is provided. According to FIG. 3 , the state parameters of the charge mixture G of the mixture state after the compressor 51 are taken into account as further input parameters for the control unit R3, i.e., in addition to the mixture state in or after the charge heat exchanger 60 as compression pressure p2 and compression temperature T2, as charge pressure p3 and charge temperature T3. The mixture state after the compressor 51 (state parameters p2, T2) is obtained within the scope of a further control unit R2 in the compressor model for describing the compressor 51.

In the intake section model, the compressor 51 is considered as another intake section device E3 with an associated control unit R2. The mixture mass flow m(')_LLK of the charge air cooler 60 serves as the inlet of the control unit. This in turn is correlated with the mixture mass flow m(')_VBP at the compressor bypass valve 72 and the mixture mass flow at the engine throttle valves 82.A and 82.B. These mixture mass flows are summed in the summing unit R42 to determine the mixture mass flow m(')_LLK supplied at the inlet of the heat exchanger 60. The mixture mass flow m(')_DK at the engine throttle valves 82.B and 82.A is in turn obtained as the output of the summing element R45, that is, it is back-calculated from the mixture mass flows m(')_G,ZYL at the cylinder inlet and the receiver volume.

Another input of the mixture mass flow m(')_DK returned by the throttle valve DK and fed back to the regulating unit R2 (that is, influencing the suction section device C5 for adjusting the return flow of the charge mixture through the compressor bypass 70) is obtained by feeding back the output of the adding unit R45 shown by the dotted line in Figure 3 to the adding unit R42.

The mixture mass flow through the compressor bypass 70 is in turn determined within the flow equation at the other intake section device C3, namely, the compressor bypass flap 72. While the mixture mass flow m(')_DK at the engine throttle flap DK is primarily determined by the position αDK of the throttle flap as the intake section device C5, the mixture mass flow m(')_VBP through the compressor bypass is primarily determined by the position αVBP of the compressor bypass throttle 72 and by taking into account the pressure conditions at the beginning and end of the compressor bypass 70, namely, p1, T1 (i.e., the charge mixture state upstream of the compressor 51) and p3, T3 (i.e., the charge mixture state upstream of the throttle flap DK).

Thus, the further control unit R1 receives as input variables not only the position αVBP of the compressor bypass throttle 72 but also the gas state parameters p3, T3 in the charge heat exchanger 60 and at least the intake pressure p1, T1 upstream of the compressor, wherein the temperature T1 can approximately correspond to the ambient temperature T0.

As a result, an improved method for operating the gas internal combustion engine 100 is achieved using the controller structure 200 shown in detail in FIG. 3 .

Claims (30)

1. A method for operating a gas-fired internal combustion engine (100), The gas internal combustion engine has - Gas mixer (40), suction section (30) and - An engine (10) with multiple cylinders, wherein, in the method - A fuel mixture is supplied to the engine (10), the fuel mixture comprising a charge mixture, and the engine (10) is operated using the gas in the charge mixture as fuel during gas combustion, wherein - The output-mixture share of the gas-air mixture associated with a subsequent mixture state is determined by means of the input-mixture share of the gas-air mixture associated with at least one previous mixture state, and wherein this determination is achieved by means of a suction section model, which serves as the basis for the calculation model for the suction section (30). Its features are, - Determine the output of the fuel-air mixture at the engine supply section - mixture ratio. - The input-mixture share is determined by the output-mixture share through multiple intermediate states of the mixture share in multiple associated volumes of the suction section (30), wherein - Determine the input-mixture ratio of the gas-air mixture at the gas mixer (40) and adjust the air flow and/or gas flow at the gas mixer (40) according to the input-mixture ratio. 2. The method according to claim 1, characterized in that, within a range determined in real time simultaneously with the operation of the engine, the input-mixture ratio is determined by the output-mixture ratio through multiple intermediate states of the mixture ratio in multiple associated volumes of the suction section (30). 3. The method according to claim 1 or 2, characterized in that, - Determine the input-mixture ratio of the gas-air mixture at the outlet of the gas mixer (40) and/or - Determine the output-mixture share of the gas-air mixture at the cylinder (Ai, Bi) or the cylinder inlet or receiver (80) of the engine (10). 4. The method according to claim 1 or 2, characterized in that the determination of the mixture proportion includes: - The mass flow of the mixture is determined by means of a flow equation for the volume of the suction section (30), wherein the mass flow of the mixture is associated with the input-mixture share at the gas mixer (40) and/or the mass flow of the mixture at the engine supply section is associated with the output-mixture share; and/or - The determination of the state of the mixture includes at least the determination of the state pressure and/or temperature of the mixture fraction for the associated volume of the suction section (30) by means of the thermodynamic state equation of the actual or ideal gas in the associated volume. 5. The method according to claim 1 or 2, characterized in that the combustion air ratio is associated with the input-mixture share, wherein the gas measurement unit of the gas mixer (40) is guided by means of stoichiometric air demand (L_st) and/or combustion air ratio. 6. The method according to claim 5, characterized in that the theoretical value of the combustion air ratio (Lambda_SOLL) is associated with the input-mixture ratio. 7. The method according to claim 1 or 2, characterized in that the determination of the intermediate mixture proportion in the associated volume of the suction section (30) is realized for a plurality of theoretically calculated volumes of the suction section (30) and/or for a plurality of actual shell volumes of the suction section (30), wherein the number of intermediate states of the mixture proportion in the suction section (30) is associated with at least one large volume of the suction section (30). 8. The method according to claim 7, characterized in that the determination of the intermediate state of the mixture mass flow and/or the state of the mixture is realized for a plurality of theoretically calculated volumes of the suction section (30) and/or for a plurality of actual shell volumes of the suction section (30). 9. The method according to claim 1 or 2, characterized in that the plurality of large volumes of the suction section (30) include one or more component volumes of the suction section (30), selected from: - At least one receiver volume (V5, p5, T5), - At least one cylinder volume in the engine assembly - At least one charge heat exchanger volume (V3, p3, T3), - At least one compressor volume (V2, p2, T2). 10. The method according to claim 9, wherein the volume of one or more components of the suction section (30) is selected from: - Two receiver volumes (V5, p5, T5). 11. The method according to claim 1 or 2, characterized in that the state pressure is defined as the receiver pressure (p5) in the receiver volume (81) before the cylinder (Ai, Bi) of the engine (10), the receiver volume being positioned upstream of the cylinder (Ai, Bi) and downstream of the booster section and/or bypass section. 12. The method according to claim 1 or 2, wherein the state pressure is virtually determined based on a calculation model of the suction section (30), which includes at least one calculated volume of the charge heat exchanger (60) and/or receiver volume (81). 13. The method according to claim 12, wherein the state pressure is determined simulcast and/or computationally based on a computational model of the suction section (30). 14. The method according to claim 1 or 2, characterized in that the number of at least one large volume of the suction section (30) includes one or more component volumes of the suction section (30), which are further selected from: - At least one compressor bypass volume (Vl,pl,Tl to V3,p3,T3); - At least one suction section volume. 15. The method according to claim 14, wherein at least one of the compressor bypass volumes (V1, p1, T1 to V3, p3, T3) is the compressor bypass volume at the bypass pipe section and/or the compressor bypass valve (72). 16. The method according to claim 14, wherein at least one of the suction section volumes is the suction section volume at the suction pipe section and/or the engine and/or the input throttle valve (82). 17. The method according to claim 1 or 2, characterized in that the mass flow of the mixture at the throttling section (72, 82) is determined, wherein the flow rates of the backflow (p3, T3, p1, Tl) and/or the booster flow are determined by means of the flow equation of the compressible medium at the ideal or practical nozzle, assuming the frictionless or frictional flow of the ideal or practical gas. 18. The method according to claim 1 or 2, wherein the gas operation is a spark-ignition gas operation. 19. An adjustment unit for a gas internal combustion engine (100), comprising: - Gas mixer (40), suction section (30) and - An engine with multiple cylinders (10), - The pressurization section of the suction section (30), and - A bypass section with the suction section (30) is used to bypass the pressurization section. The engine (10) is capable of operating in gas combustion mode with a supplied fuel mixture comprising a charge mixture containing gas as fuel. The adjustment unit is configured such that: - Determine the output mixture share of the gas-air mixture associated with subsequent mixture states by means of the input mixture share associated with at least one previous mixture state. - The suction section model, which serves as the basis for the computational model used for the suction section (30), determines the input-mixture share in the previous mixture state by means of the output-mixture share in the subsequent mixture state. The characteristic is that the adjusting part is configured as follows: - Determine the output of the gas-air mixture at the engine supply section - mixture ratio, and - The input-mixture share is determined by the mixture state of the output-mixture share through multiple intermediate states of the mixture share in multiple associated volumes of the suction section (30), wherein - Determine the input-mixture ratio of the gas-air mixture at the gas mixer (40) and adjust the air flow and/or gas flow at the gas mixer (40) according to the input-mixture ratio. 20. The adjusting unit according to claim 19, wherein the adjusting unit is used in a spark-ignition gas internal combustion engine. 21. The regulating unit according to claim 19 or 20, characterized in that the engine (10) has a receiver volume (81) arranged upstream of the cylinder. 22. The regulating section according to claim 19 or 20, characterized in that the pressurizing section of the suction section (30) is equipped with a charge heat exchanger. 23. The adjusting section according to claim 19 or 20, characterized in that the pressurizing section of the suction section (30) has a single-stage or two-stage pressurizing section. 24. The regulating section according to claim 19 or 20, characterized in that the pressurizing section of the suction section (30) is equipped with exhaust gas recirculation. 25. An internal combustion engine, configured as a gas internal combustion engine (100), having - Gas mixer (40), suction section (30) and - An engine with multiple cylinders (10), - The pressurization section of the suction section (30), and - A bypass section with the suction section (30) is used to bypass the pressurization section. The engine (10) is capable of operating in gas combustion mode with a supplied fuel mixture comprising a charge mixture containing gas as fuel, and It has an adjustment section, which is constructed as follows: - Determine the output mixture share of the gas-air mixture associated with subsequent mixture states by means of the input mixture share associated with at least one previous mixture state. - The suction section model, which serves as the basis for the computational model used for the suction section (30), determines the input-mixture share in the previous mixture state by means of the output-mixture share in the subsequent mixture state. The characteristic is that the adjusting part is configured as follows: - Determine the output of the gas-air mixture at the engine supply section - mixture ratio, and - The input-mixture share is determined by the mixture state of the output-mixture share through multiple intermediate states of the mixture share in multiple associated volumes of the suction section (30), wherein - Determine the input-mixture ratio of the gas-air mixture at the gas mixer (40) and adjust the air flow and/or gas flow at the gas mixer (40) according to the input-mixture ratio. 26. The internal combustion engine according to claim 25, wherein the internal combustion engine is configured as a spark-ignition gas internal combustion engine. 27. The internal combustion engine according to claim 25 or 26, characterized in that the engine (10) has a receiver volume (81) arranged upstream of the cylinder. 28. The internal combustion engine according to claim 25 or 26, characterized in that the pressurization section of the suction section (30) is equipped with a charge heat exchanger. 29. The internal combustion engine according to claim 25 or 26, characterized in that the supercharging section of the suction section (30) has a single-stage or two-stage supercharging section. 30. The internal combustion engine according to claim 25 or 26, characterized in that the pressurization section of the suction section (30) is equipped with exhaust gas recirculation.