METHOD AND DEVICE FOR OPERATING A GAS ENGINE

Description

[0001] The invention relates to a method for operating a gas-powered internal combustion engine with a gas mixer, an intake manifold, and an engine with a number of cylinders. The invention further relates to a control system for the internal combustion engine according to the preamble of claim 13 and an internal combustion engine according to the preamble of claim 14. The gas-powered internal combustion engine further comprises a supercharger and a bypass section in the intake manifold for circumventing the supercharger. In particular, the gas-powered internal combustion engine is a spark-ignition gas-powered internal combustion engine.

[0002] Application areas for this type of gas combustion engine include mobile applications --such as in the shipping sector or in the commercial and heavy vehicle sector -- as well as stationary applications such as combined heat and power plants, which are particularly advantageous to design for fluctuating gas supplies.

[0003] US 6,131,552 discloses a fuel control system in general terms that can regulate the supply of gas to a mixing chamber depending on a measured operating condition of the engine. The control method of US 6,131,552 A or other gas metering methods that are solely load-dependent prove to be insufficient even in complex control systems.

[0004] Such a method is described in WO 2011 / 000474 A1; this describes a method for controlling a gas engine with a generator, in which a control torque is calculated from a speed control deviation via a speed controller, and in which a target volume flow is calculated at least as a function of the control torque. A deviation of the control torque from a generator torque is calculated, and a target receiver pipe pressure is corrected based on this deviation.

[0005] Other similar methods, in which a speed control deviation is calculated from a target speed and an actual speed, are described in DE 10 2007 045 195 B3 and DE 10 2007 056 623 B3.

[0006] Air consumption is typically a measure of the amount of fresh gaseous charge supplied to an internal combustion engine within a charge air mixture. Air consumption also provides information about the quality of the intake system and the intake process. The actual air consumption regularly represents the ratio of the actual mass of fresh air supplied to the engine, or to a single cylinder, during a combustion cycle within a charge air mixture. This actual mixture mass is related to the theoretical fresh charge mass, which is determined from the geometric displacement and the theoretical charge density under ambient conditions (for naturally aspirated engines). For turbocharged engines, the state of the fresh charge downstream of the compressor or intercooler is considered.

[0007] Several factors influence the fresh charge supplied to a cylinder, such as valve timing and valve opening area. In principle, this can be determined using a module for calculating engine boost pressure, which incorporates an intake manifold model. However, in reality, the fresh charge supplied to the engine in a charge mixture only corresponds to the theoretical value in exceptional cases. The air requirement is not a constant value for an engine, but rather depends heavily on the engine speed and the specific geometric conditions of the intake tract and combustion chamber; a suitable map can be used, for example, to account for this dependency.

[0008] Intake manifold models, on the other hand, are fundamentally known in engine controllers for general internal combustion engines, such as those described in EP 1 398 490 A2. These models share the common feature that, when modeling the intake manifold—in the simplest case as a homogeneous pressure vessel to capture the dynamic processes in the air path—the storage behavior of the intake manifold (also referred to as the intake pipe) is modeled using the filling and emptying method. In this approach, the intake pipe is treated as a pressure vessel that is continuously filled with air by a throttle valve, and from which the engine, through its suction behavior corresponding to the power stroke, draws air via the intake valve.

[0009] However, it turns out that supplying fuel to a gas-fired internal combustion engine, especially in the engine's transient operating range and with variable fuel qualities, is far more complex. In particular, it has become apparent that gas-fired internal combustion engines, especially spark-ignition gas engines, can be problematic to operate in the low-load and / or transient load range.

[0010] In gas engines, mixture formation typically takes place upstream of the exhaust gas turbocharger compressor. Simultaneously, the intake tract between the compressor outlet and the combustion chamber inlet consists of partially large volumes, which consequently store or release significant mixture masses. This is particularly true when changes in engine load and / or speed lead to pressure and / or temperature fluctuations in the individual sub-volumes. As a consequence of mixture mass formation in a gas combustion engine that is only partially adapted to the operating point—because it is inherently imprecise due to the partially large volumes—high hydrocarbon emissions (HC emissions) or other elevated emissions (NOx, CO, particulate matter, etc.) as well as poor efficiency due to unburned fuel gas are to be expected.

[0011] It is desirable to make the operation of a gas-powered internal combustion engine more advantageous, particularly during transient periods, in accordance with load requirements and emission conditions. It is especially desirable to implement this at least in the low-load range. Ideally, this should be implemented across the full load range, preferably up to 100%.

[0012] The invention addresses this point, aiming to provide a method and a device for achieving improved operation of an internal combustion engine as a gas-powered engine. In particular, the invention aims to achieve improvements in transient operation and / or emissions. Preferably, it seeks to solve a problem of fuel composition existing in the low-load range in an improved manner, especially by avoiding torque jumps in the engine during transient operation, even under changing load requirements, and, if possible, by avoiding excessive emissions. In particular, mixture formation should be optimized as much as possible. Specifically, the engine's air consumption should be optimized. The invention aims, in particular, to address at least one of the problems described above.At least an alternative solution should be proposed.

[0013] The problem concerning the method is solved by the invention with a method according to claim 1.

[0014] In particular, it is assumed that a method of the type mentioned at the outset is used for operating a gas combustion engine, with a gas mixer, an intake manifold and an engine with a number of cylinders, wherein in the method The engine is supplied with a fuel mixture comprising a charging mixture, and the engine is operated in gas mode with gas as fuel in the charging mixture.

[0015] Furthermore, it is provided that, by means of an input mixture fraction of a gas-air mixture assigned to at least one earlier mixture state, an output mixture fraction of the gas-air mixture assigned to a later mixture state is determined, and wherein the determination is carried out by means of an intake path model serving as the basis of a computational model for the intake path.

[0016] According to the invention, the output mixture fraction of the gas-air mixture is determined at an engine inlet, and the input mixture fraction is determined from the output mixture fraction via a number of intermediate mixture fraction states in a number of associated volumes of the intake tract. For this purpose, the input mixture fraction of a gas-air mixture is determined at the gas mixer, and an airflow and / or a gas flow at the gas mixer is adjusted according to the input mixture fraction.

[0017] The invention is based on the consideration that advantageous transient operation of a gas combustion engine should ideally take into account a state pressure of the intake manifold, in particular by considering a throttling of the charge air or the fuel mixture, and / or by taking the air consumption into account as much as possible, while considering the specific characteristics of a gas engine. Although intake manifold models are generally known in engine control units, the invention is based on the premise that the aforementioned modeling is inherently inadequate, especially for the application of a gas engine, particularly with mixture formation upstream of a compressor.

[0018] The invention takes into account for the first time the consideration that, in a gas engine, the distance between the point of mixture formation, e.g., upstream of the compressor of the exhaust gas turbocharger, and the point of combustion chamber inlet can be comparatively long and thus inherently encompasses a comparatively large volume that is hardly realistically describable in a steady state. The invention also takes into account for the first time that, in a gas engine, the intake path between the compressor outlet and the combustion chamber inlet consists of partially large volumes that store or release significant mixture masses when pressure and / or temperature changes occur in the individual partial volumes due to changes in engine load and / or speed.

[0019] The invention recognizes that, therefore, particularly in a gas engine, the mixture mass flow at the outlet of the gas mixer is temporarily decoupled from that at the combustion chamber inlet. This fact has not yet been adequately considered in intake manifold models, especially for a gas engine. In particular, conventional steady-state analyses of a gas engine, which only consider the mixture pressure upstream of the cylinder, are therefore—as recognized by the invention—unsuitable for ensuring compliance with defined combustion air conditions. Thus, the concept of the invention serves to provide a mixture mass sensor with improved reliability and better adapted for transient operation, in particular a virtual mixture mass sensor for mixture formation upstream of the compressor.

[0020] The problem concerning the device is solved by the invention through a control device for a gas combustion engine according to claim 13.

[0021] The internal combustion engine comprises a motor with a number of cylinders and an intake system including the gas mixer and intake manifold. Furthermore, upstream of the cylinders, a receiver volume is located, for example, in the form of a manifold, a mixing section, or the like.

[0022] In particular, it has proven advantageous to equip the gas-powered internal combustion engine with a turbocharger in the intake system, especially one comprising a charge heat exchanger. Depending on the dimensions of the gas-powered internal combustion engine, particularly if based on a large engine, the turbocharger can be single- or two-stage, preferably with exhaust gas recirculation. Furthermore, a bypass section to the intake manifold of the intake system can be provided to bypass the turbocharger.

[0023] As demonstrated particularly in cycle simulations and tests on gas engines, the invention improves the stability of the air-fuel ratio, especially during transient engine operation (load ramp-up / off), thus enabling the simulation of larger load changes and facilitating simpler adjustment of load-shifting operation, even on the test bench. Furthermore, the use of a virtual gas sensor according to the invention's concept contributes to achieving current and future emissions standards.

[0024] These and other advantageous embodiments of the invention can be found in the dependent claims and specify in detail advantageous ways of realizing the concept of the invention within the framework of the embodiments and with further advantages.

[0025] Advantageously, an airflow and / or a gasflow can be adjusted at the gas mixer according to the input mixture fraction, in particular a mixture mass flow rate, within the framework of a simultaneous real-time calculation. In a particularly preferred embodiment, it is ensured that an airflow and / or a gas flow is adjusted at the gas mixer according to the input mixture fraction within the framework of a simultaneous real-time determination. In particular, a mixture mass sensor is represented according to the concept as a virtual mixture mass sensor, which advantageously provides the mixture mass flow rate at the gas mixer outlet; this is an advantage over previous approaches.

[0026] According to the invention, the input mixture fraction of the gas-air mixture is determined at an output of the gas mixer, and the output mixture fraction of the gas-air mixture is determined at a cylinder or a cylinder input of the engine, in particular at a receiver.

[0027] Particularly advantageous is the determination of the state pressure upstream of a cylinder of the engine, preferably as a receiver pressure in a receiver volume. A receiver volume is understood to be any type of volume located upstream of the cylinder in the upstream flow direction and downstream of a turbocharger and / or a bypass section. For example, a receiver volume can be the volume of a manifold or other extension of the intake manifold. In particular, a receiver volume is understood to be a volume that exceeds the usual volumes of an intake manifold. Further development shows that specifying the receiver pressure in the receiver volume is particularly significant for the reliable control of the gas-powered internal combustion engine because increasing receiver volume is associated with increasing uncertainty in the state of the fuel-air mixture.Controlling the receiver pressure within the receiver volume thus eliminates uncertainties that exist with stationary assumptions about an intake path.

[0028] Preferably, the determination of a mixture fraction comprises the determination of a mixture mass flow rate, in particular by means of a flow equation for a specific volume and / or a component of the intake tract and / or a throttle device. In particular, a mixture mass flow rate is assigned to the input mixture fraction at the gas mixer and / or a mixture mass flow rate to the output mixture fraction. In particular, a mixture mass flow rate can be determined at a throttle, whereby a flow rate of a backflow and / or a charging flow is determined. This can preferably be implemented by means of a flow equation for compressible media at an ideal or real nozzle, assuming a frictionless or frictional flow of ideal or real gases.

[0029] Preferably, the determination of a mixture fraction additionally or alternatively includes the determination of a mixture state, in particular at least the determination of a temperature and / or a state pressure of the mixture fraction for the volume of the intake tract. Preferably, this is done using a thermodynamic equation of state for real or ideal gases.

[0030] It has proven advantageous to consider the state pressure in the intake manifold, e.g., in large volumes or as a pressure drop across throttles. Actuators for these throttles can be provided to influence the state pressure, such as a throttle flap, a throttle valve, or other throttling devices. In this context, a throttling device in the intake manifold is defined as any means of pressure reduction or, more generally, pressure regulation; this may include, in addition to an engine throttle, a compressor bypass throttle. In particular, an engine throttle can also be a valve, a flap, a throttle, or even the variable turbine geometry of a compressor.An angle of attack α between the fully open and fully closed positions of the throttle is generally used here to describe a throttle position of such throttle organs; several throttle organs of the aforementioned type can also be used independently or in combination with one another, either in plurality or in combination.

[0031] In particular, a motor throttle can be provided upstream of the receiver volume and / or a compressor bypass throttle in the bypass section. Depending on a target and / or actual operating pressure of the intake section, a throttle on the intake section can be implemented, especially to throttle the motor and / or the bypass section.

[0032] In a particularly preferred embodiment, a further development provides for a division of the intake path --i.e. between gas mixer and engine-- into several, in particular at least two volumes, preferably exactly two large volumes.

[0033] This approach has proven effective, particularly when one or more of the following steps are used: the filling and emptying method (especially using pressure and temperature information, among other things, from existing measuring points), a real-time calculation of the mixture mass flows at various points in the intake path in the engine controller, a mixing of those gas mixers by a gas metering unit that leads to the virtually determined mixture mass in the desired combustion air ratio.

[0034] Advantageously, a combustion air ratio, in particular a setpoint value (Lambda_SOLL), is assigned to the input mixture fraction, whereby a gas metering unit of the gas mixer is controlled by means of the stoichiometric air requirement (L_st) and / or the combustion air ratio. This measure has proven effective for controlling a gas mixer.

[0035] In particular, the determination of the mixture fraction, especially a mixture mass flow rate and / or a mixture state of an intermediate state, in the volume of the intake manifold can be carried out for a number of theoretical calculation volumes of the intake manifold and / or, according to the invention, for a number of actual housing volumes of the intake manifold, wherein the number of intermediate states of the mixture fraction in the intake manifold is assigned to a number of at least one large volume of the intake manifold. This allows the intake manifold model to be designed to be particularly realistic.

[0036] According to the invention, the state pressure is determined—particularly virtually, especially by simulation and / or calculation—based on a computational model of the intake tract comprising at least the computational volumes of the receiver volume and the charge heat exchanger. According to the invention, the state pressure is determined as the receiver pressure in a receiver volume upstream of a cylinder of the engine, which is located upstream of the cylinder in the upstream flow direction and downstream of a charging system and / or a bypass section. A charge heat exchanger volume is to be understood as any volume involved in temperature exchange; this can, in particular, include volumes of the intake tract and / or the engine feed section of the intake manifold. The aforementioned volumes have proven to be particularly relevant for describing the intake tract.

[0037] The number of large volumes of the intake tract comprises one or more component volumes of the intake tract selected from the group comprising: at least one, preferably two, receiver volumes, at least one cylinder volume, in particular in the engine block, at least one charge heat exchanger volume, at least one compressor volume.

[0038] More preferably, the number of at least one large volume of the intake path comprises one or more component volumes of the intake path, which are selected from the group further comprising: at least one compressor bypass volume, in particular at a bypass pipe section and / or a compressor bypass flap; at least one intake path volume, in particular at an intake pipe section and / or an engine and / or inlet throttle valve.

[0039] The preferred gas-powered ignition system is spark-ignited; this proves to be particularly efficient and suitable for many applications. However, other ignition principles are also suitable, such as a diesel or other liquid fuel ignition system. In particular, the gas-powered ignition can alternatively be a pilot-ignition system with external mixture formation of a gas-air mixture, using a diesel or other liquid fuel pilot jet.

[0040] Generally, although not the primary focus here, it is possible, not least because of the comparatively consistent fuel quality, to operate the gas-powered engine either with gas or with liquid fuel such as diesel or liquefied petroleum gas (LPG). During operation, the engine is supplied with a fuel mixture comprising a charge mixture and / or a liquid fuel. The engine can operate in two states: first, on diesel – with diesel or another liquid fuel – and second, on gas – with gas as the fuel in the charge mixture. These types of gas-powered engines are also known as dual-fuel engines and, in addition to the preferred fuels of diesel and gas, can also be operated with a wide variety of other fuels.In particular, the gas-powered internal combustion engine can alternatively be operated in pilot-injection mode, with external mixture formation of a gas-air mixture and a diesel pilot jet. Gas-powered internal combustion engines are thus also referred to as pilot-injection engines and are regularly based on a diesel engine design. They represent one of the latest technologies, especially in the area of ​​environmentally friendly applications for large engines. A pilot-injection engine can also be operated with liquid fuel such as diesel or other liquefied fuels such as liquefied natural gas (LNG) or liquefied petroleum gas (LPG); in most cases, a gas-powered internal combustion engine can be combined with a gas-diesel engine to form a gas-diesel engine.

[0041] In particular, a gas-powered internal combustion engine has an injection system that is preferably electronically controlled. Specifically, an internal combustion engine can further feature an injection system that is advantageously designed as a common-rail injection system. In particular, an injection system can be controlled for different gas qualities, such as biogas or natural gas, in liquid form, or for the use of oils such as vegetable oils or the like as liquid fuel. Common-rail injection systems, and possibly also pump-nozzle injection systems with electronic control, have proven particularly effective in this regard. The ignition medium, in gas operation, can be added to the actual gaseous fuel of the charge mixture in the cylinder at high compression or added to the intake manifold.Gas engines operating on gas, especially spark-ignited gas engines or pilot-jet engines, with external mixture formation, are generally more flexible in their fuel use and produce even fewer emissions.

[0042] Exemplary embodiments of the invention are now described below with reference to the drawings and comparison with the prior art, some of which is also shown. These exemplary embodiments are not necessarily to scale; rather, where explanatory, the drawings are presented in a schematic and / or slightly distorted form. For further details regarding the teachings directly apparent from the drawings, reference is made to the relevant prior art. It should be noted that numerous modifications and changes concerning the form and details of an embodiment can be made without deviating from the general idea of ​​the invention. The features of the invention disclosed in the description, the drawings, and the claims can be essential for the further development of the invention, both individually and in any combination.Furthermore, the invention encompasses all combinations of at least two of the features disclosed in the description, the drawing, and / or the claims. The general idea of ​​the invention is not limited to the exact form or detail of the preferred embodiment shown and described below, nor is it limited to an object that would be restricted compared to the object claimed in the claims. For specified design ranges, values ​​lying within the stated limits are also disclosed as limit values ​​and can be used and claimed as desired. Further advantages, features, and details of the invention will become apparent from the following description of the preferred embodiments and from the drawing; this is shown in: . Fig. 1 shows a schematic of a gas-powered internal combustion engine with a gas mixer and an intake manifold with turbocharging and a charge heat exchanger, and an engine with a number of cylinders after a receiver volume, wherein the turbocharging can be bypassed by a bypass section. The gas-powered internal combustion engine is designed for spark-ignition gas operation. In an alternative shown in dashed lines, the gas-powered internal combustion engine can also be designed as a gas-diesel engine and can be operated in pure diesel mode as well as in mixed mode or in pure gas mode (e.g., as pilot-ignition operation with injection of an ignition mixture in the form of diesel), wherein the injection system is formed in the form of a common-rail system shown in dashed lines.2. A flowchart of a preferred embodiment of a method for determining, within the framework of a simultaneous real-time determination, an input mixture fraction of a gas-air mixture associated with an earlier mixture state by means of an output mixture fraction of a gas-air mixture associated with a later mixture state, wherein the input mixture fraction of the gas-air mixture is determined at the gas mixer and an airflow and / or a gas flow is set at the gas mixer according to the input mixture fraction; Fig. 3. A schematic representation of a preferred embodiment of a controller structure for dual-fuel operation, wherein the input mixture fraction of a gas-air mixture of the gas mixer is determined from the output mixture fraction of a gas-air mixture supplied to the engine via a number of intermediate states of the mixture fraction in the intake manifold by means of the intake manifold model.

[0043] Fig. 1 Figure 1 shows a gas combustion engine 100 with an engine 10 and an intake system with a branched intake path 30. The intake path includes, among other things, a gas mixer 40 and, for the purpose of creating a charge, a turbocharger 50 and a charge heat exchanger 60, here in the form of an intercooler, as well as a bypass 70.

[0044] The engine is presented here as a V-engine with sixteen cylinders, comprising eight cylinders Ai, i=1..8 on one A-side and eight cylinders Bi, i=1..8 on the other B-side; this cylinder arrangement and number are shown here only as an example. For large-engine applications, in particular, the engine can also be configured with ten, twelve, twenty, twenty-four, or twenty-eight cylinders, or any other number of cylinders.

[0045] In the case of an alternative or additional configuration as a dual-fuel internal combustion engine, the engine also has an injection system 20, shown with dashed lines, which in this case is a common-rail system with a common rail 21 from which a number of injection lines 22—each with an injector 23 and an individual accumulator 24 upstream of the injector—branch off to each of the cylinders Ai, Bi, i=1..8 of the engine 10. The injection system 20 is designed to ration liquid fuel such as diesel or another liquefied or liquid fuel in order to inject it as liquid fuel in diesel operation or as a pilot jet in gas or pilot-injection operation at the beginning of each working cycle of cylinder Ai, Bi; this is done at very high injection pressures.Accordingly, in this variant, the engine 10 further features a common-rail injection system 20 for a liquid fuel, in particular diesel fuel, as well as a charging system 50 with a charging heat exchanger 60 and with a bypass 70 to bypass the charging system 50 and the charging heat exchanger 60.

[0046] Referring further to the essential part of the embodiment shown with solid lines, the gas mixer 40, connected to the intake section 30 at the inlet end of the intake system, draws in charge air LL from the environment and mixes it with fuel gas BG. The charge mixture, also referred to as the fuel gas mixture in gas operation, --hereinafter also referred to as "mixture G"-- with mass flow rate m(')_G ("( ')" is shown in the drawing for clarity as a dot above the mass m or other quantity) is fed at the intake pressure p1 and at an intake temperature T1, which essentially corresponds to the ambient temperature, via a compressor section 32 to a compressor 51 of the turbocharger 50 and is compressed there to a compression pressure p2 at a compression temperature T2.The compressor 51 is driven by a turbine 52 and is mounted on a common compressor shaft 53 with it; the turbine 52 of the exhaust manifold 90, in turn, is driven by the exhaust gas AG leaving the engine 10 in the exhaust manifold 90. The mass flow m(')_G of the mixture G, heated to the compression temperature T2 as a result of compression, is fed to a cooling section 31 of the intake manifold 30 and there passed through a cooler structure 61 in a charge heat exchanger 60; in the heat exchanger volume 62, shown here symbolically, heat exchange takes place with a cooling medium in the cooler structure 61, so that the mixture G is cooled. The fuel-air mixture leaves the heat exchanger volume of size V3 in a cooled form at a charging temperature T3 and a boost pressure p3 towards a charge section 33 for supplying the mixture G to the engine 10.

[0047] In an intake manifold model, the state of the mixture G upstream of the compressor 51 is described relatively generally using the state variables for pressure and temperature, here intake temperature T1 and intake pressure p1 upstream of the compressor 51, and downstream of the compressor 51 at increased compression pressure p2 and increased compression temperature T2, using the state variables p2 and T2 downstream of the compressor 51, by means of a suitable compressor model; this is done, for example, according to a gas law equation, such as for an ideal or real gas. The following components of the heat exchanger 60 and the receiver 80, such as a manifold and / or a collection section, are of particular importance as large volumes of the intake manifold 30 according to the concept of the invention, so that a heat exchanger volume V3 and a receiver volume V5, respectively, are assigned to these and to the further space of the intake manifold for modeling the further gas states in the intake manifold model.Accordingly, the fuel gas mixture G in the heat exchanger volume V3 assumes the state variables p3, T3 as a result of cooling and an increase in volume with decreasing boost pressure and charging temperature p3, T3.

[0048] The state of the mixture G in bypass 70 is also fundamentally determined according to the state variables p1, T1 at the inlet and p3, T3 at the outlet of bypass 70, or vice versa in the case of backflow through bypass 70; i.e., a bypass gas mixture G_BP in the bypass section 71 of bypass 70 is determined depending on the prevailing pressure conditions and the position of the compressor bypass throttle 72 --here according to the actuation angle αVBP of the compressor bypass flap -- a bypass section 71 can in particular serve to return excess mixture G to the compressor 51 in order to recompress it and supply it again for combustion in the cylinders Ai, Bi of the engine 10.

[0049] Before the gas mixture G in state p3, T3 is supplied to the engine 10, it is fed into the receiver 80 under varying pressure and temperature according to a mass flow rate m(')_DK, which is directed into the receiver volume 81 via the engine throttle 82, and according to the receiver volume V5 at a receiver pressure p5 and a receiver temperature T5. In this case, a first and a second receiver volume 81.B, 81.A are each assigned to a B-side and an A-side of the engine 10, respectively; that is, these are arranged upstream of the cylinders Ai, Bi and after the first and second charging sections 33.B, 33.A of the B-side and A-side, and after the heat exchanger volume 62. The motor throttle 82 is formed in this case by a first and a second motor throttle valve 82.B, 82.A, which are each assigned to the first and second receiver volume 81.B, 81.A, wherein the first and the second motor throttle valve 82.B, 82.The motor throttle valves 82A and 82B are independently adjustable; where simpler, they are referred to collectively as the motor throttle valve 82. The receiver volume 81 is to be understood as the sum of the first and second receiver volumes 81.A and 81.B. Within the receiver volume 81, the mixture G, as a result of the volume increase and depending on the position αDK of the motor throttle valves 82.A and 82.B, assumes the gas states characterized by p5 and T5 in volume V5 of the receiver volume 81; this depends on the B-side and A-side mass flow rates m(')_DK,B and m(')_DK,A, respectively, depending on the position of the motor throttle valves 82.B and 82.A.

[0050] The states of the gas mixture G characterized by pi, Ti, i=1,2 or Vj, pj, Tj, j=3,5 are essentially determined in this embodiment in the areas as given by the compressor 51, the heat exchanger volume 62 and the receiver volume 81, or against the limits given by the motor throttle 82 and the compressor bypass throttle 72 or the compressor 51.

[0051] In the following, the receiver pressure p5 in the receiver volume V5, or the control variables based on the receiver pressure p5 --such as an actual receiver pressure p5_actual or a target receiver pressure p5_target or a simulated receiver pressure p5 - plays a role in determining an input mixture fraction of a gas-air mixture at the gas mixer 40 via a number of intermediate states of the mixture fraction in a number of assigned volumes of the intake tract 30, based on the intake path model of a gas combustion engine 100 shown here.

[0052] It turns out that the adjustment of the mass flows m(')_G for fuel gas BG and m(')_LL for charge air LL at the gas mixer 40, corresponding to a combustion air ratio LAMBDA_SOLL or a stoichiometric air ratio Lst, cannot necessarily be carried out under the assumption of steady-state conditions along the intake path. The concept of the embodiment takes into account in an intake path model --as shown by Fig. 1 As described, at least two large volumes are used to summarize the volume of the intake manifold: the receiver volume 80 and the charge heat exchanger volume 62. The intake manifold 30 is modeled within the intake manifold model based on the—generally known—principle of a filling and emptying method. The changes of state in the volumes are considered quasi-isothermal here. This simplifies the system by limiting it to mass conservation compared to an adiabatic approach and, in particular, simplifies the simultaneous calculation of the internal combustion engine and its intake manifold in real time. However, an adiabatic or polytropic approach, or a targeted heat transfer, can also be used to simulate the changes of state in the intake manifold if sufficient computing capacity is available.

[0053] Out of Fig. 3 Furthermore, it becomes apparent that special assumptions for components of the intake manifold can be implemented within the framework of supplementary models, particularly when measured values ​​for the corresponding component of the intake manifold 30 are not available. This applies, for example, to the supplementary model of a compressor (p2_T2 module), which describes the operation of the compressor 51 and the states of the mixture G before the compressor via temperature and pressure (G(p1, T1)) and after the compressor (G(p2, T2)). This also applies, for example, to the supplementary model for a compressor bypass valve (Massnstrom_VBP), which describes a mass flow through the compressor bypass valve within the framework of a flow equation.

[0054] During operation, a fuel mixture comprising a charging mixture is supplied to engine 10. The engine is operated using gas as fuel in this charging mixture. Using an output mixture fraction (mass flow rate m(')_G,ZYL) of the gas-air mixture, assigned to a later mixture state (with state variables p5, T5), an intake path model—as described in Fig. 3 The structure of the controller 200 describes how an input mixture fraction (mass flow rate m(')_G,SOLL) of a gas-air mixture is determined, which is assigned to at least one previous mixture state (with state variables p0, T0). The determination is carried out using the intake manifold model as the basis of a computational model for the intake manifold 30, the gas mixer 40 and the receiver 80, or the input to the engine 10.

[0055] According to the concept of this embodiment, the output mixture fraction (mass flow rate m(')_G,ZYL) of the gas-air mixture is determined at an engine inlet—in this case, a receiver 80. The input mixture fraction (mass flow rate m(')_G,SOLL) is determined from the output mixture fraction (mass flow rate m(')_G,ZYL) via a number of intermediate states (with state variables pi, Ti, i=5,3,2,1) of the mixture fraction—in this case, within the framework of a simultaneous real-time calculation. The number of intermediate states (with state variables pi, Ti, i=5,3,2,1) are assigned to a number of volumes (Vi, i=5,3,2,1) or to the components E1, E2, E3, E4, as well as C3 and C5 in the intake manifold. The input mixture ratio of a gas-air mixture is determined at the gas mixer 40 and, according to the input mixture ratio, an air flow and / or a gas flow BG is set at the gas mixer 40.

[0056] In the present intake tract model, a number of large volumes are assigned to the intake tract; these include: two receiver volumes 81.B, 81.A (with assigned volume V5 and with state variables p5, T5 of a charging mixture therein), at least one cylinder volume in the engine block, at least one charge heat exchanger volume 62 (with assigned volume V3 and with state variables p3, T3 of a charging mixture therein), at least one compressor volume at the compressor 51 (with value V2 and with state variables p2, T2 of a charging mixture at its outlet) or the states upstream of the compressor 51 (with state variables p1, T1 of a charging mixture) and with state variables p0, T0 of the surroundings (atmosphere) of an intake air, as described in Fig. 1 is displayed.

[0057] The also in Fig. 1 The components E1, E2, E3, E4, C3, and C5 of the intake system shown are included in the present intake manifold model, and the states and mass flow rates of the charge mixture are simulated and calculated accordingly. This is merely a preferred example from a number of other possible examples; for instance, a modified intake manifold model could also include further components in the case of two-stage turbocharging. In principle, additional volumes can also be considered in an intake manifold model. In the limiting case, the intake manifold can be divided into finite or infinitesimal subvolumes, and the determination of the states and mass flow rates of the charge mixture then results from, among other things, the solution of a difference or...Differential equation system; this consideration of model refinement can be made taking into account the computational effort and the complexity of the intake path to be simulated; in particular taking into account real-time capability.

[0058] Fig. 2 The flowchart shows, in principle, the sequence of a preferred embodiment of an operating method for a gas combustion engine 100 with a gas mixer 40, an intake manifold 30 and an engine 10, also referred to herein as a gas engine, with a number of cylinders Ai, Bi, i=1..8, as can be seen from Fig. 1 was described. Basically, the operating procedure can be implemented by implementing a Fig. 2 The illustrated determination method for control variables on a gas mixer 40 is to be implemented, by means of which an airflow LL and / or a gas flow BG can be adjusted. Referring to Fig. 2 Based on the intake tract model shown here, the procedure in this case provides for further steps S1 to S9 in step S0, which are based on the previously described components E1, E2, E3, E4, C3 and C5 of the intake system defined by the intake tract model, which are in Fig. 1 are displayed accordingly.

[0059] Specifically, the determination of an input mixture fraction of component E0 – here assigned to the gas mixer 40 – according to the invention is carried out "backwards," i.e., by calculating it from the knowledge of the output mixture fraction at component E4 of the intake manifold. In this case, component E4 essentially corresponds to engine 10 or a cylinder Ai, Bi, i=1..8 of engine 10. In a first step S1 of the determination method, a cylinder charge is determined using corresponding parameters for engine speed nMOT and receiver pressure p5 as well as receiver temperature T5 via a calculation module in the control unit R4, for example, a measure for the required air input LAMBDA_a at engine 10 is specified. In a second step S2 of the determination method, a mixture mass flow rate m(')_ZYL can then be specified at component E4 of the intake system – here a cylinder inlet, for example, to a cylinder Ai, Bi, i=1..8.

[0060] Accordingly, in Fig.3 A controller structure of a controller 200 is shown, which—for the execution of steps S1 and S2—has a first control unit R4. The input variables are an engine speed nMOT, as well as a cylinder inlet pressure—specifically, here a receiver pressure p5 in a receiver 80, or receiver volume 81.B, 81.A with a total volume V5—and, if applicable, further associated gas state variables of a mixture component in the receiver 80 (with receiver volume V5).

[0061] In a third step S3 of the Fig. 2 The receiver volumes 81.A, 81.B with a total volume V5 of receiver 80 are then taken into account as a further component E1 of the intake system --for example, assigned to a manifold or a mixing section that is integrated into the intake path between cylinder inlets and the engine throttle valve DK.

[0062] Given the aforementioned knowledge of the pressure and temperature conditions in the receiver volume 81.A, 81.B V5 (receiver pressure p5, receiver temperature T5), a mixture mass flow through the receiver m(') _RECEIVER can be determined in a fourth step S4 by means of a Fig. 3 The control unit R5 shown determines the mixture mass flow rate. The control unit R5 has a processing unit which, from the mixture state of a gas / air mixture in the receiver 80 (also often referred to as the receiver tube) --i.e., in the receiver volumes 81.A, 81.B with volume V5, determines the state parameters p5, T5 of the charging mixture -- using a state equation of the charging mixture and a mass flow equation according to the filling and emptying method for the receiver tube 81.A, 81.B.

[0063] Furthermore, the preferred embodiment of the determination method provides for Fig. 2 For the further component E2, based on the intake manifold model, the charging heat exchanger 60 is presented. In the associated volume V3 of the heat exchanger volume 62, the charge mixture assumes a gaseous state, which can be calculated in a fifth step S5 for volume V3 using the charge pressure p3 and the charge temperature T3. Analogously, a mixture mass flow rate m(')_LLK through the charging heat exchanger 62 is determined according to the determination method shown here. Accordingly, Fig. 3 This includes a control unit R3 for determining the mass flow rate m(')_LLK through the charge air cooler 60 in the form of the charge air cooler with the cooler structure 61. The determination is carried out in a sixth step S6 using the filling and emptying method for the volume V3, knowing the mixture state or its thermodynamic properties and a corresponding mixture mass flow equation.

[0064] Using the mixture mass flows m(')_CYL, m(')_RECEIVER and m(')_LLK, a target value for a combustion air ratio LAMBDA-target and a stoichiometric air requirement (Lst) can then be determined in a seventh step S7, i.e., the essential target parameters of the fuel at the further component E0 of the gas mixer 40 in the intake manifold 30. For this purpose, in the Fig. 3 The controller structure shown includes a first adding element R45 and a second adding element R43 to determine the mixture mass flow rate m(')_G,SOLL from a mixture mass flow equation; this takes into account a further controller unit R0 for implementing the seventh step S7, namely for calculating the combustion air ratio LAMBDA_SOLL and the stoichiometric air requirement Lst.

[0065] This allows a suitable control variable to be specified at the gas mixer 40 in the eighth step S8, in order to adjust the input mixture proportion at the gas mixer 40 according to a current engine speed nMOT and the pressure and temperature conditions (essentially p5, T5) at the cylinder inlet.

[0066] The intake manifold model is designed with two large volumes—namely the receiver volume 81.B, 81.A (V5) and the charge heat exchanger volume 62 (V3)—and, taking into account the filling / emptying method for both volumes in combination with corresponding flow equations for at least the throttles 82, 72, as well as a compressor model at the compressor 51 (with state variables p1, T1 → p2, T2), is incorporated into a simulation or calculation of the intake manifold 30. This is sufficient to eliminate the aforementioned problem; the large volumes, which are not adequately considered in previously known calculation methods and therefore cause a decoupling of the gas mass flow at the gas mixer and the actual gas mass flow at the engine, are included in the intake manifold model of the [reference to be added]. Fig. 2 The determination method presented here adequately takes this into account. At the same time, it is evident that the determination method, in the form presented here, is comparatively simple and therefore computationally efficient, making it available for simultaneous real-time determination. In this way, the gas mixer 40 can be adjusted to the immediate demand of the motor 10 practically in real time.

[0067] In further training of the in Fig. 2 The determination method shown demonstrates Fig. 3 Specifically, the control unit R3 as well as feedback of the mass flows m(')_DK, m(')_RECEIVER and m(')_LLK. This makes sense if a bypass 70 is provided. According to Fig. 3 The state variables of the charging mixture G for a mixture state after the compressor 51 are considered as further input variables to the control unit R3, namely as compression pressure p2 and compression temperature T2 in addition to the mixture state in or after the charging heat exchanger 60, as boost pressure p3 and charging temperature T3. The mixture state after the compressor 51 (state parameters p2, T2) is determined within the framework of a further control unit R2 for describing the compressor 51 in a compressor model.

[0068] In the intake manifold model, the compressor 51 is considered as a further intake manifold component E3 with an associated control unit R2. The input to the control unit is a mixture mass flow m(')_LLK from the charge air cooler 60. This, in turn, depends on a mixture mass flow m(')_VBP at the compressor bypass valve 72 as well as the mixture mass flow at an engine throttle valve 82.A, 82.B. These mixture mass flows are summed in the addition 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 valve 82.B, 82.A is then derived as the output of the addition element R45, i.e., calculated back from the mixture mass flow m(')_ZYL at the cylinder inlet and receiver volume.

[0069] The further input of a mixture mass flow m(')_DK returned by the throttle valve DK in feedback to the control unit R2, --i.e. the influence of the intake manifold device C5 for adjusting the return flow of a charging mixture through the compressor bypass 70-- results from the in Fig. 3 The dashed line shows the feedback from the output of the adding unit R45 to the input of the adding unit R42.

[0070] The mixture mass flow through the compressor bypass 70 is determined by a flow equation at the further intake manifold device C3, namely the compressor bypass valve 72. While the mixture mass flow m(')_DK at the engine throttle valve DK is essentially determined by the position αDK of the throttle valve as intake manifold device C5, the mixture mass flow of the compressor bypass m(')_VBP is essentially determined by the position αVBP of the compressor bypass valve 72 and by taking into account the pressure conditions at the beginning and end of the compressor bypass 70, namely by taking into account p1,T1 (i.e. the charge mixture state before the compressor 51) and p3,T3 (i.e. the charge mixture state before the throttle valve DK).

[0071] 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 variables p3, T3 in the charging heat exchanger 60 and at least the intake pressure p1, T1 before the compressor, whereby the temperature T1 can essentially correspond to the ambient temperature T0.

[0072] As a result, using the in Fig. 3 The regulator structure 200 shown in more detail implements an improved method for operating the gas combustion engine 100.

Claims

1. Method for operating a gas-fired internal combustion engine (100), with - a gas mixer (40), an intake path (30), an intercooler (60), and - an engine (10) with a number of cylinders (Ai, Bi), and a receiver (80) with a receiver volume (81) upstream of a cylinder (Ai, Bi) of the engine (10), wherein in the method - the engine (10) is supplied with a fuel mixture comprising a charge mixture and the engine (10) is operated in gas operation with gas as fuel in the charge mixture, wherein - an output mixture fraction (m(')_G, ZYL), assigned to a later mixture state (p5, T5), of a gas-air mixture (G) is determined by means of an input mixture fraction (m(')_G, SOLL), assigned to at least one earlier mixture state (p0, T0), of the gas-air mixture (G), with a determination which takes place by means of an intake path model serving as the basis for a computational model for the intake path (30), wherein with the determination - (S1, S2, R4) the output mixture fraction (m(')_G, ZYL) of the gas-air mixture (G) is determined at an engine supply, and (S7, R0) the input mixture fraction (m(')_G, SOLL) of the gas-air mixture (G) is determined at the gas mixer (40) and (S8) an air flow and / or a gas flow at the gas mixer (40) is set in accordance with the input mixture fraction (m(')_G, SOLL), - specifically, the input mixture fraction of the gas-air mixture is determined at an outlet of the gas mixer (40) and the output mixture fraction of the gas-air mixture is determined at a cylinder (Ai, Bi) or a cylinder inlet of the engine (10), wherein - the input mixture fraction is determined from the output mixture fraction over a number of intermediate states of the mixture fraction in a number of assigned volumes of the intake path (30), wherein - the determination of the input mixture fraction takes place by reverse calculation from the output mixture fraction, and with the determination - the mixture fraction is determined in the assigned volume of the intake path (30) for a number of real housing volumes of the intake path (30), wherein the number of intermediate states of the mixture fraction in the intake path (30) is assigned to at least one large volume of the intake path (30), and a number of large volumes of the intake path (30) comprises one or more component volumes of the intake path (30) which comprises at least one receiver volume (V5, p5, T5, 81, 80) and an intercooler volume (V3, p3, T3, 62, 60), and - a state pressure is determined on the basis of the computational model of the intake path (30) comprising (R45, R43) at least one computational volume of the receiver volume (V5, p5, T5, 81, 80) and of the intercooler volume (V3, p3, T3, 62, 60), wherein the state pressure is defined as the receiver pressure (p5) in the receiver volume (81) upstream of a cylinder (Ai, Bi) of the engine (10) which is arranged before the cylinder (Ai, Bi) in the upstream direction and is arranged after a forced-induction unit and / or a bypass path in the downstream direction.

2. Method according to Claim 1, characterized in that the input mixture fraction is determined from the output mixture fraction over the number of intermediate states of the mixture fraction in the number of assigned volumes of the intake path (30) in the course of a real-time determination simultaneous with the operation of the engine.

3. Method according to Claim 1 or 2, characterized in that - the output mixture fraction of the gas-air mixture is determined at the receiver (80).

4. Method according to one of the preceding claims, characterized in that the determination of a mixture fraction comprises: - the determination of a mixture mass flow by means of a throughflow equation for the volume of the intake path (30), wherein a mixture mass flow is assigned to the input mixture fraction at the gas mixer (40) and / or a mixture mass flow at the engine supply is assigned to the output mixture fraction; and / or - the determination of a mixture state comprises at least the determination of a temperature and / or a state pressure of the mixture fraction for the assigned volume of the intake path (30) by means of a thermodynamic state equation for real or ideal gases in the assigned volume.

5. Method according to one of the preceding claims, characterized in that a combustion air ratio is assigned to the input mixture fraction, specifically a SETPOINT value of the combustion air ratio (Lambda_SOLL), wherein a gas metering unit of the gas mixer (40) is controlled by means of a stoichiometric air requirement (L_st) and / or by means of the combustion air ratio.

6. Method according to one of the preceding claims, characterized in that the determination of the mixture fraction, specifically a mixture mass flow and / or a mixture state of an intermediate state, takes place in the assigned volume of the intake path (30) for a number of theoretical computational volumes of the intake path (30), wherein the number of intermediate states of the mixture fraction in the intake path (30) is assigned to at least one large volume of the intake path (30).

7. Method according to one of the preceding claims, characterized in that the number of large volumes of the intake path (30) comprises a component volume of the intake path (30) which - comprises at least one compressor volume (V2, p2, T2).

8. Method according to one of the preceding claims, characterized in that the state pressure is determined virtually, on the basis of the computational model of the intake path (30) comprising at least the computational volume of the receiver volume (V5, p5, T5, 81, 80) and of the intercooler volume (V3, p3, T3, 62, 60).

9. Method according to one of the preceding claims, characterized in that the state pressure is determined virtually, specifically is simulated and / or calculated.

10. Method according to one of the preceding claims, characterized in that the number of at least one large volume(s) of the intake path (30) comprises one or more component volumes of the intake path (30) which are chosen from the group further comprising: - at least one compressor bypass volume (V1, p1, T1 to V3, p3, T3) at a bypass pipe section and / or a compressor bypass flap (72); - at least one intake path volume at an intake pipe section and / or at an engine throttle flap and / or inlet throttle flap (82).

11. Method according to one of the preceding claims, characterized in that a mixture mass flow is determined at a throttle (72, 82), wherein a throughflow of a return flow (p3, T3, p1, T1) and / or of a forced-induction flow is determined, by means of a throughflow equation for compressible media at an ideal or real nozzle, assuming a flow of ideal or real gases which is frictionless or is subject to friction.

12. Method according to one of the preceding claims, characterized in that the gas operation is spark-ignition gas operation.

13. Control device with a controller (200) for a gas-fired internal combustion engine (100) with - a gas mixer (40), an intake path (30) and - an engine (10) with a number of cylinders (Ai, Bi) and a receiver (80) arranged upstream of the cylinders and with a receiver volume (81) arranged upstream of the cylinders (Ai, Bi) of the engine (10), - a forced-induction unit for the intake path (30) with an intercooler (60), and - with a bypass path for the intake path (30) for bypassing the forced-induction unit, wherein the engine (10) can be operated in gas operation with gas as fuel with a supplied fuel mixture comprising a charge mixture, and the control device with the controller (200) is designed: - to determine an output mixture fraction (m(')_G, ZYL), assigned to a later mixture state (p5, T5), of a gas-air mixture (G) by means of an input mixture fraction (m(')_G, SOLL), assigned to at least one earlier mixture state (p0, T0), of the gas-air mixture (G) with a determination which takes place by means of an intake path model serving as the basis for a computational model for the intake path (30), wherein the control device with the controller (200) is designed with a determination: - (S1, S2, R4) to determine the output mixture fraction (m(')_G, ZYL) of the gas-air mixture (G) at an engine supply and to determine the input mixture fraction (m(')_G, SOLL) of the gas-air mixture (G) at the gas mixer (40), and to set an air flow and / or a gas flow at the gas mixer (40) in accordance with the input mixture fraction, - specifically to determine the input mixture fraction of the gas-air mixture at an outlet of the gas mixer (40) and to determine the output mixture fraction of the gas-air mixture at a cylinder (Ai, Bi) or a cylinder inlet of the engine (10), wherein the input mixture fraction is determined from the output mixture fraction over a number of intermediate states of the mixture fraction in a number of assigned volumes of the intake path (30), wherein - the determination of the input mixture fraction takes place by reverse calculation from the output mixture fraction, and with the determination - the mixture fraction is determined in the assigned volume of the intake path (30) for a number of real housing volumes of the intake path (30), wherein the number of intermediate states of the mixture fraction in the intake path (30) is assigned to at least one large volume of the intake path (30), and a number of large volumes of the intake path (30) comprises one or more component volumes of the intake path (30) which comprises at least the receiver volume (V5, p5, T5, 81, 80) and an intercooler volume (V3, p3, T3, 62, 60), and a state pressure is determined on the basis of the computational model of the intake path (30) comprising (R45, R43) at least one computational volume of the receiver volume (V5, p5, T5, 81, 80) and of the intercooler volume (V3, p3, T3, 62, 60), wherein the state pressure is defined as the receiver pressure (p5) in the receiver volume (81) upstream of a cylinder (Ai, Bi) of the engine (10) which is arranged before the cylinder (Ai, Bi) in the upstream direction and is arranged after a forced-induction unit and / or a bypass path in the downstream direction.

14. Internal combustion engine designed as a gas-fired internal combustion engine (100), with - a gas mixer (40), an intake path (30) and - an engine (10) with a number of cylinders (Ai, Bi) and a receiver (80) arranged upstream of the cylinders and with a receiver volume (81) arranged upstream of the cylinders (Ai, Bi) of the engine (10), - a forced-induction unit for the intake path (30) with an intercooler (60), and - with a bypass path for the intake path (30) for bypassing the forced-induction unit, and - with a control device according to Claim 13 with the controller (200), wherein the engine (10) can be operated in gas operation with gas as fuel with a supplied fuel mixture comprising a charge mixture.

15. Internal combustion engine designed as a gas-fired internal combustion engine (100), according to Claim 14, characterized in that - the gas-fired internal combustion engine (100) is designed as a spark-ignition gas-fired internal combustion engine, and / or - the forced-induction unit for the intake path (30) with the intercooler (60) is designed as a one- or two-stage forced-induction unit and optionally with exhaust-gas recirculation.