KR102581651B1 - A large two-stroke turbocharged uniflow scavenged internal combustion engine and method of operating the engine - Google Patents

Abstract

A large two-stroke turbocharged internal combustion engine includes a controller (60) that: monitors operating conditions, when the engine is operating under steady-state operating conditions, and when the engine is operating under steady-state conditions. Consists of determining whether the air-fuel ratio is controlled individually for each combustion chamber as a function of the operating conditions with a value for the air-fuel ratio smaller than a known operating condition dependent critical level by a margin initially set to a first value (p1). step, reducing for each combustion chamber individually the margin value decreasing over time from the actual value towards the second value (p2), each combustion chamber separately for partial misfire event, misfire event and pre-ignition event. The step of monitoring, when such an event is detected, increases the margin value from the actual value to the first value (p1) and uses that method until such an event is no longer detected.

Description

A LARGE TWO-STROKE TURBOCHARGED UNIFLOW SCAVENGED INTERNAL COMBUSTION ENGINE AND METHOD OF OPERATING THE ENGINE}

The present disclosure relates to large two-stroke gaseous fuel internal combustion engines, particularly large two-stroke uniflow scavenge internal combustion engines with a crosshead running on gaseous fuel flowing from a fuel valve during the piston stroke from BDC to TDC.

Large two-stroke turbocharged Uniflo scavenge internal combustion engines with crossheads are used, for example, for propulsion of large oceangoing vessels or as main power plants in power plants. This two-stroke diesel engine is constructed differently from other internal combustion engines not just because of its size. The exhaust valve weighs up to 400 kg, the piston diameter is up to 100 cm, and the maximum operating pressure in the combustion chamber is typically several hundred bar. The forces associated with these high pressure levels and piston sizes are enormous.

A large two-stroke turbocharged internal combustion engine powered by a gaseous fuel introduced by a fuel valve located centrally along the length of the cylinder liner or located within the cylinder cover, i.e. the gas is pumped during the upward stroke of the piston, which begins well before the exhaust valve closes. An engine that receives fuel, compresses the mixture of gaseous fuel and scavenging air in the combustion chamber, and ignites the compressed mixture, uses timed ignition means such as pilot oil injection to remove the gas from the combustion chamber. The mixture of fuel and scavenging air is compressed and the compressed mixture is ignited.

This type of gas intake, using a fuel valve (gas intake valve) placed in the cylinder liner or cylinder cover, has the advantage of allowing the use of much lower fuel injection pressures because the gaseous fuel is injected when the compression pressure is relatively low. . This is low compared to large two-stroke turbocharged internal combustion engines that inject gaseous fuel when the piston is close to top dead center (TDC), i.e. when the compression pressure in the combustion chamber is at or near maximum. The latter engines require much higher fuel injection pressures than the already high maximum combustion pressures. Fuel systems that can handle gaseous pressures at these extremely high pressures are expensive and complex due to the volatile nature of gaseous fuels and their behavior at these high pressures.

Therefore, a fuel supply system for an engine that injects gaseous fuel during the compression stroke is significantly cheaper compared to an engine that injects gaseous fuel at high pressure when the piston is near TDC.

However, when injecting gaseous fuel during the compression stroke, the piston compresses the mixture of gaseous fuel and scavenge air, resulting in the risk of pre-ignition. The risk of pre-ignition can be reduced by operating with very lean mixtures, but lean mixtures increase the risk of misfire or partial misfire and consequently fuel slip.

Therefore, there is a need for improved control of the conditions of the combustion chamber during compression in these large two-stroke turbocharged internal combustion engines to overcome or at least reduce problems associated with misfire and pre-ignition/diesel-knock. Conditions in the combustion chamber must be controlled very precisely to prevent pre-ignition and misfire.

During normal operation of the engine, the performance layout of the engine generally ensures that pre-ignition is prevented. This is achieved through careful selection of combustion chamber design, fuel injection timing and exhaust valve timing. However, it must be operated at a safe distance from combustion conditions where misfire or partial misfire or pre-ignition are likely to occur. This large safety distance results in suboptimal combustion conditions, especially with regard to fuel efficiency.

DK201970370 discloses a large two-stroke turbocharged uniflow scavenged gas operated internal combustion engine with multiple combustion chambers, at least one controller associated with the engine, a controller that determines the average compressed air-fuel ratio and bulk compression temperature of the combustion chamber at the start of combustion, The controller is configured as follows:

- perform at least one compressed air-fuel ratio increase measurement when the air-fuel ratio is below the lower compressed air-fuel ratio threshold,

- perform at least one compressed air-fuel ratio reduction measurement when the determined or measured average compressed air-fuel ratio exceeds the upper compressed air-fuel ratio threshold,

- perform at least one bulk compression temperature increase measurement when the determined or measured bulk compression temperature is below a low bulk compression temperature threshold,

- Perform at least one bulk compression temperature reduction measurement when the determined or measured bulk compression temperature is above the upper bulk compression temperature threshold.

The goal is to provide an engine and method that overcomes or at least reduces the above-described problems.

The above and other objects are achieved by the features of the independent claims. Additional embodiment forms become apparent from the dependent claims, description and drawings.

According to a first aspect, a large two-stroke turbocharged uniflow scavenge internal combustion engine configured to operate using gaseous fuel as the main fuel in a gaseous operation mode, said engine comprising: a cylinder liner, a reciprocating piston and a cylinder cover. A plurality of combustion chambers, each distinct, a scavenging port arranged in the cylinder liner for introducing scavenging air into the combustion chamber, an exhaust gas outlet disposed in the cylinder cover and controlled by an exhaust valve, and the exhaust valve timing for each combustion chamber can be controlled. a variable timing exhaust valve actuation system, one or more gaseous fuel inlet openings arranged in the cylinder liner or cylinder cover configured to introduce gaseous fuel during the stroke of the piston toward the cylinder cover, and at least one controller associated with the engine; The at least one controller is configured to determine and control the opening and closing timing of the exhaust valve for each combustion chamber and to control the amount of gaseous fuel flowing into the combustion chamber through the gaseous fuel inlet opening for each combustion chamber, At least one controller is configured to monitor operating conditions of the engine and determine when the engine is operating at steady-state operating conditions, wherein the combustion chamber is configured to initiate a partial misfire event, a misfire event, when the air-fuel ratio exceeds a known operating condition dependent on a threshold level. having known undesirable combustion conditions at least for steady-state operation at which events and/or pre-ignitions are likely to occur, wherein the variable timing exhaust valve actuation system allows individual control of exhaust valve timing for each combustion chamber, wherein at least One controller is configured to determine and control the opening and closing timing of the exhaust valve individually for each combustion chamber and to control the amount of gaseous fuel flowing into the combustion chamber through the gaseous fuel inlet opening individually for each combustion chamber, The at least one controller is configured to operate in a steady state mode when the at least one controller determines that the engine is operating in steady state conditions, wherein in the steady state operating mode the at least one controller is configured to: Controlling the air-fuel ratio for the combustion chamber to a value of the air-fuel ratio that is less than the known operating condition dependent threshold level by a margin set to a first value (p1) earlier than the known operation, and over time for each combustion chamber the actual value individually decreasing the margin value decreasing toward a second value (p2), wherein the second value (p2) is less than the first value (p1) and greater than 0, and includes a partial misfire event, a misfire event, and individually monitoring each combustion chamber for pre-ignition events, and upon detection of a partial misfire event, a misfire event and/or a pre-ignition event, the actual combustion chamber until the partial misfire event, misfire event and/or pre-ignition event are no longer detected. It may consist of increasing the margin value incrementally from a value toward the first value (p1).

By allowing the margin to decrease relative to the margin for steady-state operating conditions, engine operating parameters such as (maximum) fuel efficiency (energy efficiency), (minimized) NOx emissions or (minimized) hydrocarbon slip (HC slip) ( It allows safe use of unburned or partially unburned fuel at values close to critical levels, thus improving the possibility of optimizing engine operation and combustion processes. Engines can therefore be designed to optimize these operating parameters.

Optimization specifies how the setpoint for the actuator should deviate from normal steady-state operating conditions to optimize the combustion process. Because uncontrollable factors (e.g. ambient conditions and component maintenance) determine how much optimization is possible, optimization approaches only specify deviation rules and are at least partially reversed when undesirable combustion conditions are detected. Therefore, uncontrollable factors determine the extent to which the rule (the magnitude of the deviation from the actuator setpoint) can be applied.

An undesirable combustion state is one in which a misfire event, partial misfire event, or pre-ignition event occurs.

Depending on the critical level of air-to-fuel ratio for steady-state operation, operating conditions are determined during engine design and/or based on test runs and/or computer simulations.

The first value of the margin is determined during engine design and/or based on test runs and/or computer simulations. Margin values as a function of operating conditions are stored as lookup tables or implemented as algorithms.

The process of reducing the margin value from an initially set first value begins when the controller determines that the engine is operating under steady-state conditions either immediately or with a predetermined delay (a predetermined length of time).

In a possible implementation of the first aspect, at least one controller is notified of undesirable combustion conditions and known operating conditions according to threshold levels.

In a possible implementation of the first aspect, the controller may adjust the margin value individually for each combustion chamber to decrease slightly from the actual value when a predetermined length of time has elapsed since the last increase in value and the value is not equal to the second value. It is configured to resume decline over time.

In a possible implementation of the first aspect, the controller is configured to reduce the margin value according to a predetermined reduction by advancing the exhaust valve closing timing, preferably by advancing the exhaust valve closing timing in steps.

In a possible implementation of the first aspect, the controller is configured to increase the margin value according to a predetermined increment by retarding the exhaust valve closing timing, preferably by retarding the exhaust valve closing timing in steps.

In a possible implementation of the first aspect, the engine includes an exhaust gas bypass having an exhaust gas bypass control valve, and the controller is configured to reduce the reduction margin value by closing or increasing the throttling of the exhaust gas bypass control valve. do.

In a possible implementation of the first aspect, the engine includes an exhaust gas bypass having an exhaust gas bypass control valve, and the controller adjusts the margin value in predetermined increments by opening or reducing the throttling of the exhaust gas bypass control valve. It is configured to increase.

In a possible implementation of the first aspect, the engine includes an exhaust gas recirculation conduit having an exhaust gas recirculation blower in the exhaust gas recirculation conduit, wherein the controller adjusts the margin value in predetermined increments by activating or increasing the speed of the exhaust gas recirculation blower. It is configured to increase.

In a possible implementation of the first aspect, the engine includes an exhaust gas recirculation conduit having an exhaust gas recirculation blower in the exhaust gas recirculation conduit, wherein the controller is configured to reduce the reduction margin value by deactivating or reducing the speed of the exhaust gas recirculation blower. It is composed.

In a possible implementation of the first aspect, the engine includes a cylinder bypass upstream of the main scavenge air cooler, and the controller adjusts the air-fuel ratio by opening the hot cylinder bypass conduit or reducing throttling of a control valve in the hot cylinder bypass conduit. It is configured to increase and vice versa.

In a possible implementation of the first aspect, the controller is configured to activate liquid fuel injection when operating conditions require it, and to reset the margin value to the first value when injection of liquid fuel, for example diesel oil, is activated. do.

In a possible implementation of the first aspect, an increment is a small increment, a decrease is a small decrease, and a step is a small step.

In a possible implementation of the first aspect, the engine includes a sensor for individually sensing the cylinder pressure for each cylinder, wherein the controller is configured to individually monitor the sensed cylinder pressure for each cylinder, and the controller is configured to individually monitor the sensed cylinder pressure for each cylinder. It is configured to individually determine for each cylinder whether a misfire event, partial misfire event and/or pre-ignition event has occurred in that cylinder.

In a possible implementation of the first aspect, the controller determines the deviation of the cylinder pressure evolution from the expected evolution of the cylinder pressure in the absence of a misfire event, when a partial misfire event and/or a pre-ignition event occurs, thereby determining a misfire event, a partial misfire event and/or a pre-ignition event. /or configured to determine a pre-ignition event.

In a possible implementation of the first aspect, the controller is configured to determine that the engine operates in a steady state if the difference between the desired engine speed and the actual engine speed is below the deviation threshold and at the same time the engine load is above the engine load threshold.

In a possible implementation of the first aspect, one or more of the gaseous fuel inlet openings are configured to introduce gaseous fuel received from a supply of pressurized gaseous fuel into the combustion chamber through a fuel inlet valve.

According to a second aspect, a method of operating a large two-stroke turbocharged uniflow scavenge internal combustion engine with multiple combustion chambers in gaseous operation mode, wherein an air-fuel ratio mixture having an air-fuel ratio is present in the combustion chamber prior to ignition; said combustion chamber having, at least for steady-state operation, known undesirable combustion conditions in which partial misfire events, misfire events and/or pre-ignition are likely to occur when the air-fuel ratio exceeds known operating conditions dependent on a critical level; and is: monitoring the operating conditions of the engine and determining when the engine is operating at steady-state operating conditions, and when the steady-state operating conditions are determined: above by a margin initially set to the first value (p1); individually controlling the air-fuel ratio for each combustion chamber as a function of the operating conditions with a value for the air-fuel ratio less than a known operating condition-dependent critical level, determining a second value (p2) from the actual value over time for each combustion chamber. individually decreasing the margin value decreasing toward, wherein the second value (p2) is less than the first (p1) value and is greater than 0, and separately monitoring, and upon detection of a partial misfire event, a misfire event, and/or a pre-ignition event, the first value from the actual value until the partial misfire event, misfire event, and pre-ignition event are no longer detected. and increasing the margin value in increments toward (p1).

These and other aspects will become apparent from the examples described below.

In the following detailed portions of the present disclosure, aspects, embodiments and implementations will be described in more detail with reference to example embodiments shown in the drawings, wherein:
1 is a front view of a large two-stroke diesel engine according to an embodiment of the present invention;
Figure 2 is a side view of the large two-stroke engine of Figure 1;
Figure 3 is a first schematic diagram of the large two-stroke engine according to Figure 1;
FIG. 4 is a cross-sectional view of the cylinder frame, cylinder liner, and piston of the engine of FIG. 1 equipped with a cylinder cover and an exhaust valve, with TDC and BDC shown;
Figure 5 is a second schematic diagram of the engine of Figure 1;
Figure 6 is a schematic diagram of a compression temperature observer and a compression air-fuel ratio observer;
7 is a diagram showing the compressed air-fuel ratio on the vertical axis and the bulk cylinder temperature on the horizontal axis, showing a safety zone surrounded by zones requiring measures to return to the safe zone;
8 is a process illustrating one embodiment of a method for controlling a large two-stroke engine;
Figure 9 is a diagram showing the individual optimization process for each cylinder.

In the following detailed description, internal combustion engines will be described with reference to a large two-stroke, low-speed turbocharged internal combustion crosshead engine in an exemplary embodiment. 1, 2 and 3 show an embodiment of a large, low-speed turbocharged two-stroke diesel engine with a crankshaft (8) and a crosshead (9). 1 and 2 are front and side views, respectively. Figure 3 is a schematic diagram of the large, low-speed turbocharged two-stroke diesel engine of Figures 1 and 2 with intake and exhaust systems. In this exemplary embodiment, the engine has four cylinders in a row. Large, low-speed turbocharged two-stroke internal combustion engines typically have between four and fourteen cylinders in a row supported by an engine frame (11). The engine can be used, for example, as the main engine of a ship or as a generator in a power plant. For example, the total output of the engine may range from 1,000 to 110,000 kW.

In this exemplary embodiment the engine is a two-stroke uniflow scavenging type engine with a scavenging port 18 in the lower region of the cylinder liner 1 and a central exhaust valve 4 at the top of the cylinder liner 1. Scavenging air passes from the scavenging air receiver (2) through the scavenging port (18) of the individual cylinder liner (1) when the piston (10) is below the scavenging port (18). When the piston moves upward and before the piston passes the fuel valve 30 (gas inlet valve), gaseous fuel is introduced from the gaseous fuel injection valve 30 under the control of the electronic controller 60.

The gas is introduced at a relatively low pressure, less than 30 bar, preferably less than 25 bar and more preferably less than 20 bar. The fuel valves 30 are preferably evenly distributed around the circumference of the cylinder liner and are placed somewhere in the central region of the length of the cylinder liner 1. Therefore, the influx of gaseous fuel occurs when the compression pressure is relatively low. When the piston reaches TDC, it is well below the compression pressure, so relatively low pressures are also acceptable.

The piston 10 of the cylinder liner 1 compresses the gaseous fuel and scavenge air, the compression taking place at or near TDC ignition, for example via a pilot oil fuel valve 50, preferably arranged in the cylinder cover 22. ) is injected with pilot oil (or any other suitable ignition liquid), combustion follows and exhaust gases are produced. Instead of or in addition to the pilot oil fuel valve 50, an alternative ignition system may be used, for example a pre-chamber (not shown), a laser ignition (not shown) or a glow plug (not shown). The form can also be used to initiate ignition.

When the exhaust valve (4) opens, the exhaust gas flows through the exhaust duct connected to the cylinder (1) to the exhaust gas receiver (3) and then through the first exhaust conduit (19) to the turbine (6) of the turbocharger (5). Flowing, the exhaust gases flow from the turbine through the second exhaust conduit through the economizer 20 to the outlet 21 and into the atmosphere.

Via the shaft the turbine (6) drives the compressor (7), from which fresh air is supplied through the air inlet (12). The compressor (7) delivers compressed scavenge air to the scavenge air conduit (13) leading to the scavenge air receiver (2). The scavenge air in the conduit 13 passes through an intercooler 14 to cool the scavenge air.

When the compressor 7 of the turbocharger 5 does not deliver sufficient pressure to the scavenge receiver 2, i.e. under low or partial load conditions, the cooled scavenge air is supplied to the electric motor for pressurization. It passes through the auxiliary blower (16) driven by 17).

At higher engine loads the turbocharger compressor (7) delivers sufficient compressed scavenge air and the auxiliary blower (16) is bypassed via the check valve (15).

Figure 4 shows a cylinder liner 1 generally designated for large two-stroke crosshead engines. Depending on the engine size, cylinder liners 1 can be manufactured in various sizes, with cylinder bores typically ranging from 250 mm to 1000 mm and typical lengths ranging from 1000 mm to 4500 mm.

In Figure 4 the cylinder liner 1 is shown mounted on the cylinder frame 23 and the cylinder cover 22 is disposed on top of the cylinder liner 1 with an airtight interface therebetween. 4 the piston 10 is positioned at both bottom dead center (BDC) and top dead center (TDC), although of course it is clear that these two positions do not occur simultaneously and are separated by a 180 degree rotation of the crankshaft 8. It is schematically depicted with a line. The cylinder liner (1) is provided with a cylinder lubrication hole (25) and a cylinder lubrication line (24) which provide a supply of cylinder lubricant when the piston (10) passes through the lubrication line (24), which is then connected to the piston ring (not shown). ) distributes the cylinder lubricant onto the working surface of the cylinder liner (1). Engines typically have a geometric compression ratio between 8 and 15, but in engines where high-pressure gas injection is provided from a fuel injection valve in the cylinder cover, the geometric compression ratio may be more than 20 for gaseous fuel injected at high pressure at or near TDC.

A pilot fuel valve 50 (typically more than one per cylinder), or a pre-chamber with a pilot oil valve 50, is mounted on the cylinder cover 22 and connected to a pilot oil source (not shown). The timing of pilot oil injection is controlled by the electronic control unit 60.

The fuel valve 30 is installed in the cylinder liner 1 (or cylinder cover 22), its nozzle is substantially at the same level as the inner surface of the cylinder liner 1, and the rear end of the fuel valve 30 is located at the cylinder liner 1 (or cylinder cover 22). It protrudes from the outer wall of the liner 1. Typically, one or two, but possibly three or four fuel valves 30, distributed circumferentially (preferably evenly distributed circumferentially) around the cylinder liner 1, are provided in each cylinder. Provided in liner (1). The fuel valve 30 is arranged substantially centrally along the length of the cylinder liner 1 in one embodiment.

Figure 4 also schematically shows a gaseous fuel supply system comprising a pressurized gaseous fuel source 40 connected to each inlet of the gaseous fuel valve 30 via a gaseous fuel supply conduit 41.

Figure 5 shows a schematic diagram of an engine similar to Figure 2, but with a more detailed description of the engine's gas exchange infrastructure. Ambient air is sucked in at ambient air pressure and temperature and delivered to the compressor (7) of the turbocharger (5) through the air inlet (12). The scavenging air compressed in the compressor (7) is conveyed through the air conduit (32) to the distribution point (28).

The distribution point 28 allows scavenge air to branch off via the hot cylinder bypass conduit 29 to the turbine connection 32 of the first exhaust conduit 19. Flow through the hot cylinder bypass conduit (29) is regulated by the hot cylinder bypass control valve (31). The cylinder bypass control valve 31 is electronically controlled by the controller 60. The effect of opening the hot cylinder bypass conduit 29 or reducing the throttling of the control valve 31 in the hot cylinder bypass is an increase in the air-fuel ratio and an increase in the bulk compression temperature and vice versa.

The air conduit 13 further includes a first scavenge air control valve 33 upstream of the intercooler 14. The second scavenging control valve 34 is arranged downstream of the intercooler 14. The air conduit (13) continues into the scavenge air receiver (2). The auxiliary blower 16 branches off from the intercooler 14.

A cold cylinder bypass conduit (35) connects the scavenge air receiver (2) to the turbine connection (32) of the first exhaust conduit (19). Flow through the coat of bypass (35) is regulated by the cold cylinder bypass control valve (36). Cold cylinder bypass control valve 36 is electronically controlled by controller 60. The effect of opening the cold cylinder bypass 35 or reducing the throttling of the cold cylinder bypass valve 36 is an increase in bulk compression temperature.

The cold scavenge bypass conduit 37 allows scavenge air to escape from the scavenge air receiver 26 from the environment. Flow through the cold scavenge bypass conduit (37) is controlled by a cold scavenge bypass control valve (38). The cold scavenge bypass control valve 38 is electronically controlled by a controller 60. The cold scavenge bypass control valve 38 reduces the scavenge air pressure and reduces the air-fuel ratio, and vice versa. The cooled scavenge bypass conduit 37 need not branch from the scavenge air receiver 2, but may branch from the air conduit 13 at any location downstream of the intercooler 14.

The exhaust gas recirculation conduit 42 connects the exhaust gas receiver 3 to the scavenge receiver 2 and connects the exhaust gas recirculation control valve 45, the exhaust gas recirculation cooler 44 and the exhaust gas recirculation blower 43. Includes. The exhaust gas recirculation blower 43 and the exhaust gas recirculation control valve 45 are both used to regulate the flow through the exhaust gas recirculation conduit 42 under electronic control of the controller 60. Under normal operating conditions, the exhaust gas recirculation blower 43 is activated because the pressure in the exhaust gas receiver 42 is generally lower than the pressure in the scavenge receiver 2 (therefore the exhaust gas recirculation blower 43 will not operate). (the exhaust gas recirculation control valve (45) must be closed). The exhaust gas recirculation conduit (42) does not need to be connected from the exhaust gas receiver (3) but can be connected at any point to the first exhaust conduit (19) and does not need to be connected to the scavenge air receiver (2) and the intercooler (14). It is well connected to any location in the air conduit 13 downstream of.

Activating or increasing the speed of the exhaust gas recirculation blower 43 in the exhaust gas recirculation conduit 42 reduces the compressed air-fuel ratio and slightly reduces the bulk compression temperature, and vice versa.

The exhaust gas bypass 39 branches off from the exhaust gas receiver 3 or the first exhaust conduit 19 and connects to the atmosphere 27 at a given back pressure 27 . Exhaust gas bypass control valve 49 regulates flow through exhaust gas bypass conduit 39 under electronic control of controller 60.

Opening the exhaust gas bypass control valve 49 or reducing the throttling of the exhaust gas bypass control valve 49 reduces the compressed air-fuel ratio of the cylinder and vice versa.

In engines provided with a selective catalyst receiver (SVR) reactor and a reactor bypass valve (RVB), the turbine (6) of the turbocharger (5) is fed from the scavenge air receiver (3) passing through the SCR reactor under electronic control of the controller (60). ) to adjust the fraction of flow to.

All the above-mentioned components controlled by the controller 60 are connected to these components by signal lines indicated as interrupted lines in FIG. 5 .

Figure 6 shows the air-fuel ratio observer 46 and the bulk compression temperature observer 47.

The air-fuel ratio observer 46 is a computer-implemented algorithm that holds information about scavenge air pressure, exhaust valve closing timing, cylinder geometry, stoichiometric air-fuel ratio, and injected gas volume. Compressed air-fuel ratio observer 46 may be part of controller 60 or may be a separate computer or controller. The compressed air-fuel ratio observer 46 provides an output that is an estimate of the compressed air-fuel ratio of the (complete) compressed air-fuel ratio mixture (when the piston 10 is at TDC) and sends this to the controller 60. The estimate is based on the ratio of the fresh air mass captured in the combustion chamber when the exhaust valve 4 lands on its seat divided by the fresh air mass required for stoichiometric combustion of the total injected gas mass.

Bulk compression temperature observer 47 is a computer implemented algorithm that holds information about scavenge pressure, scavenge air temperature, exhaust valve closing timing and crankshaft speed. Bulk compression temperature observer 47 may be part of controller 60 or may be a separate computer or controller. Bulk compression temperature observer 47 provides an output that is an estimate of the maximum bulk compression temperature of the combustion chamber, Tcomp(Tc), in the time window from the start of gas injection to pilot injection. Bulk compression temperature observer 47 sends an estimate to controller 60. In one embodiment, the Tcomp estimate refers to piston 10 at TDC.

Figure 7 is a graph setting the bulk compression temperature (Tcomp) to the air-fuel ratio (λ). The steady state basic zone 51 lies within boundaries defined by the lower air-fuel ratio threshold, the upper air-fuel ratio threshold, the lower bulk compression temperature threshold, and the upper bulk compression temperature threshold. In this steady-state default region 51, the controller 60 provides each cylinder individually with the amount of fuel required for the current engine load and the controller 60 takes no action to change the bulk compression temperature. Control it with Air-fuel ratio up to a level that is a function of engine operating conditions with a safety distance in the form of a margin from known undesirable combustion conditions at which partial misfire events, misfire events and/or pre-ignition are likely to occur. exceeds The level of the first margin has the first value greater than 0.

When combustion conditions in the cylinder liner 1 threaten to leave the normal running zone 51 and enter the operating zone 52, the controller 60 will take action to prevent this from happening.

Here the controller 60 is configured individually for each cylinder as follows:

- perform at least one compressed air-fuel ratio increase measurement (CAFRIM) if the determined or measured average compressed air-fuel ratio is below a lower compressed air-fuel ratio threshold, and

- perform at least one compressed air-fuel ratio reduction measurement (CAFRDM) when the determined or measured average compressed air-fuel ratio exceeds the upper compressed air-fuel ratio threshold,

- perform at least one bulk compression temperature increase measurement (BCTIM) when the determined or measured bulk compression temperature is below a lower bulk compression temperature threshold,

- Perform at least one bulk compression temperature reduction measurement (BCTDM) when the determined or measured bulk compression temperature is higher than the upper bulk compression temperature threshold.

By performing these actions, the controller 60 maintains the state of each cylinder liner 1 inside the normal running zone 51 and, at least temporarily, moves the condition outside the normal running zone 51 to the action zone ( 52) is allowed to enter. Zone 52 is surrounded by a critical zone 53 where a pre-ignition and/or misfire event is highly likely to occur.

The boundaries for zones 51, 52, and 53 may be defined by upper and lower thresholds for the bulk compression temperature and upper and lower limits for the compression air-fuel ratio. These thresholds can be determined for a particular engine empirically through trial and error or through computer simulation of the engine cycle.

When the observers indicate that both the compressed air-fuel ratio and bulk compression temperature are outside the normal operating zone 51, the controller 60 monitors the compressed air-fuel ratio and bulk compression temperature to move the condition of the cylinder liner 1 back to the normal operating zone 51. Two measures are taken to individually adjust the compression temperature.

Adjust the exhaust gas bypass control valve (49) (move the exhaust gas bypass control valve (49) to a more open position) to open the exhaust gas bypass (EGB) conduit (39) (from the TC turbine inlet to the turbine outlet or surrounding Opening the flow causes a decrease in the scavenging air pressure, thereby reducing the air mass captured in the combustion chamber. As a result, this measure is suitable for reducing the compressed air-fuel ratio with only a minor effect on the compressed bulk temperature. If the engine has more turbochargers, a single EGB can still be used in the exhaust gas receiver as long as the location is chosen according to the different potential mixing points in the different flows to the exhaust gas receiver.

Opening the hot cylinder bypass control valve 31 (flow from TC compressor outlet to TC turbine inlet) increases the compressed air-fuel ratio and bulk compression temperature in the combustion chamber.

Opening the scavenge bypass control valve (38) creates flow from the scavenge air receiver (2) to or around the compressor inlet and its opening has a qualitative effect similar to that of an exhaust gas bypass on the air compression-fuel ratio, but on the scavenging process ( Therefore, it has a different effect on the bulk compression temperature of the combustion chamber. The effect of opening the scavenge bypass control valve 38 on combustion chamber conditions is faster compared to exhaust gas bypass.

Opening the cold cylinder bypass valve 36 increases the flow from the scavenge air receiver to the TC turbine inlet and increases the bulk compression temperature, but has very little effect on the compressed air-fuel ratio.

Exhaust valve closing timing determines the ratio between compression and scavenge pressures in the combustion chamber. Varying timing has a significant effect on both the compressed air-fuel ratio and the bulk compression temperature in the combustion chamber.

Exhaust valve opening timing affects the first step in the combustion chamber cleaning process. Changing the timing affects engine efficiency and the cleaning process. As the scavenging process changes, the resulting bulk temperature also changes. If the exhaust valve (4) is opened very early, there is no flow to the scavenge air receiver (2) when the piston (10) subsequently opens the scavenge port (18). When the exhaust valve (4) opens very slowly there is a large flow into the scavenging receiver (2). The piston 10 then opens the scavenge port 18. This action changes the scavenging process and therefore the fraction of 'dirty hot' gases from the previous combustion joining the next compression stroke.

Therefore, if the exhaust valve (4) is opened later, more “dirty hot” gases are generated from the previous combustion, which reduces the compressed air-fuel ratio and increases the bulk compression temperature. Closing the exhaust valve (4) very late will reduce the “dirty hot” gases from previous combustion and thus increase the compressed air-fuel ratio at the bulk compression temperature. Increasing compression by closing the exhaust valve (4) earlier means less gas escapes through the exhaust valve (4) and thus more gas is trapped in the combustion chamber. This increases the air-fuel ratio. Additionally, increasing compression leads to more compression work performed by the piston 10 on the gases in the combustion chamber. This leads to higher gas temperatures in the combustion chamber.

By increasing the exhaust gas recirculation flow rate by activating the exhaust gas recirculation blower (43) or increasing the speed of the exhaust gas recirculation blower (43), more air flows from the exhaust gas receiver (3) to the turbocharger compressor outlet or scavenge air receiver (2). Exhaust gases flow, which reduces the compressed air-fuel ratio.

Increasing the speed of the auxiliary blower 16 will slightly increase the compressed air-fuel ratio.

In the case of water injection engines, injecting water into the combustion chamber during compression lowers the bulk compression temperature.

Scavenge Air Cooler Bypass (not shown): Bypassing the intercooler (14) significantly increases the bulk compression temperature in the combustion chamber with a negligible effect on the compressed air-fuel ratio.

In the case of engines provided with variable geometry turbines 6, the effect of reducing the turbine flow area is an increase in scavenging air pressure and therefore an increase in the mass of air trapped in the combustion chamber. As a result, this measure is suitable for reducing the compressed air-fuel ratio with only a minor effect on the compressed bulk temperature.

For engines provided with turbocharger assist, increasing the assist to increase the speed of the turbocharger (5) increases the compression air-fuel ratio, but has little effect on the compression temperature.

Another measure is to change the ratio between gaseous and liquid fuels (e.g. diesel oil or marine diesel). Reducing the gas-fuel ratio of the total injection fuel energy increases the compressed air-fuel ratio during compression. This increases the liquid fuel ratio and maintains crankshaft torque.

In the case of engines where the heat exchanger is installed in the exhaust gas receiver (or if there is a heat exchanger that receives part of the exhaust gases), increasing the proportion of exhaust gases passing through the heat exchanger, i.e. extracting more heat from the exhaust gases, reduces the scavenge air pressure. This decreases, and thus the trapped air mass in the combustion chamber decreases. As a result, this measure is suitable for reducing the compressed air-fuel ratio with only a minor effect on the compressed bulk temperature. Heat exchangers can be used for steam production.

For engines with hot scavenge bypass, opening the hot scavenge bypass control valve establishes or increases flow from the compressor outlet to the ambient or compressor inlet, significantly reducing scavenge air pressure and thus trapping air in the combustion chamber. Air mass decreases. As a result, this measure is suitable for reducing the compressed air-fuel ratio.

In one embodiment, the lower compressed air-fuel ratio threshold, upper compressed air-fuel ratio threshold, lower bulk compression temperature threshold, and upper bulk compression temperature threshold are engine operating condition dependent parameters. Engine operating conditions are determined by parameters such as engine load, ambient temperature, ambient humidity, engine speed, etc. Values for these operating condition dependent parameters may be available to the controller 60, for example via lookup tables or algorithms or a combination thereof.

In an embodiment, controller 60 is configured as follows:

If the determined or measured average compressed air-fuel ratio is below a lower minimum compressed air-fuel ratio threshold than a lower minimum compressed air-fuel ratio threshold, taking an additional compressed air-fuel ratio increase action (e.g., selected from the actions mentioned above);

If the determined or measured average compressed air-fuel ratio is higher than the maximum compressed air-fuel ratio threshold, that is, higher than the maximum compressed air-fuel ratio threshold, performing additional compressed air-fuel ratio reduction measurements, and

and performing at least one additional bulk compression temperature reduction measurement when the determined or measured bulk compression temperature is above a maximum bulk compression temperature threshold that is higher than the upper bulk compression temperature threshold.

These additional measures are taken when the state of the combustion chamber moves from the operating area 52 to the critical area 53 surrounding the operating area 52 . Therefore, the controller 60 processes back to the work area 52 and then back to the normal execution area 51.

The controller 60 is configured to minimize the constraints, i.e. the measures mentioned above in order to move the engine back to operating conditions within the normal zone 51. Accordingly, the controller is configured to terminate all the above-mentioned actions when the state of the combustion chamber returns to the normal operating region.

Figure 8 is a flowchart showing the process of operating the engine according to the configuration of the controller 60 described above.

After the process controller starts, it verifies that the compressed air-fuel ratio is below the lower threshold. If the answer is no, the controller moves on to verifying that the engine is running in a normal state. If the answer is yes, the controller moves on to executing the air-fuel ratio optimization process. This process will be explained in detail with reference to FIG. 9. If the response is 'no', the controller 60 checks whether the compressed air-fuel ratio exceeds the upper limit, and if the response is 'yes', the controller 60 takes the compressed air-fuel ratio. Take steps to increase the compressed air-fuel ratio in one of the measures mentioned above. Next, controller 60 determines whether the compressed air-fuel ratio is below the minimum threshold. If the answer is no, the controller moves to the step of checking whether the upper compression ratio threshold is exceeded, and if the answer is yes, the controller 60 moves to the step of taking the action of increasing the additional compression fuel ratio in the above-mentioned actions. This is to check whether the compression-fuel ratio is above the upper threshold.

Controller 60 determines whether the compressed air-fuel ratio exceeds the upper threshold. If the answer is no then controller 60 moves on to checking if the lower bulk compression temperature threshold has been exceeded, if yes then controller 60 takes action to reduce the compressed air-fuel ratio from one of the actions mentioned above. Next, the controller 60 checks whether the compressed air-fuel ratio is greater than the maximum threshold. If the answer is no, the controller moves on to checking whether the lower bulk compression temperature threshold has been exceeded, and if the answer is yes, the controller 60 takes further compression-to-fuel ratio reduction measures in the above-mentioned actions and then bulk compression. Go to the step of checking if the temperature is below the lower threshold.

Controller 60 verifies that the bulk compression temperature is below a lower threshold. If the answer is no, the controller 60 moves to the next step to check if the bulk compression temperature is above the upper threshold, and if the response is yes, the controller 60 takes action to increase the bulk compression temperature. Controller 60 then checks if the bulk compression temperature is below a minimum threshold, and if the answer is no, controller 60 moves to checking if the bulk compression temperature is above an upper threshold, and if the answer is yes, Controller 60 takes further bulk temperature increasing actions from the actions mentioned above and then moves on to checking if the bulk compression temperature threshold has been exceeded.

Controller 60 checks whether the bulk compression temperature threshold is exceeded, and if the response is no, controller 60 returns to checking if the compressed air-fuel ratio is below the lower threshold, and if the response is yes, controller 60 returns to the step of checking if the compressed air-fuel ratio is below the lower limit threshold. Take measures to reduce bulk temperature. Next, the controller 60 checks if the bulk compression temperature is above the maximum threshold, if the answer is "no" then moves back to checking if the compressed air-fuel ratio is below the lower threshold and if the answer is yes, the controller 60 After taking additional bulk temperature reduction measures from the measures mentioned above, we move on to checking whether the compressed air-fuel ratio is below the lower threshold.

In one embodiment, controller 60 is provided with an algorithm, lookup table, or other information to determine which of the available actions to increase or decrease the air-fuel ratio is the most appropriate action under the current operating conditions of the engine.

Figure 9 shows air-fuel ratio optimization for a single cylinder, where each cylinder of the engine is provided with an individual air-fuel ratio optimization unit. In one embodiment, the air-fuel ratio optimization unit and related units are integrated into controller 60. In other embodiments, these units are part of a controller (not shown) associated with controller 60.

The output of the air-fuel ratio optimizer for a single cylinder is individually controlled, i.e. each cylinder has its own individual cylinder control unit which receives signals from the cylinder-specific air-fuel ratio optimizer.

The steady-state basic mode value for the air-fuel ratio is transmitted to the summation point, and the result of the summation point is transmitted to the individual cylinder control unit for that cylinder. The output of the air-fuel ratio optimization unit (λ optimization unit) is also added to the summation point.

The air fuel optimizer determines at least the engine load, speed error (difference between desired engine speed (RPM) and actual engine speed (RPM)) and FRC (high pressure injection of liquid fuel (e.g. diesel oil or marine diesel) at or near TDC. used to stabilize the combustion process).

Additionally, the cylinder pressure variation estimation module receives the measured cylinder pressure for the individual cylinder involved. The cylinder pressure fluctuation estimation module determines the desired combustion event and whether a desired combustion event, such as a misfire event, partial misfire event, and/or pre-ignition (pre-ignition is determined by the deviation and not the fluctuation) occurs. The cylinder pressure change estimation module determines the deviation between the expected cylinder pressure compared to the actual (measured) evolution of the cylinder pressure and determines the occurrence of an undesired combustion event based on the deviation and/or deviation from the estimation.

The optimizer integrates (decreases over time) the margin values from an initial first level p1, gradually decreasing towards the minimum value p2 (where p2 is greater than 0). Once the value p2 is reached, it remains until the event is forced to change. The process of integrating margin values from the first level p1 to the minimum level p2 is relatively slow, typically taking at least a few minutes and at most 10-15 minutes.

The process of reducing the margin value from an initially set first value begins when the controller 60 determines that the engine is operating steady state conditions either immediately or with a predetermined delay (a predetermined length of time).

As previously discussed, the margin value is the margin from the air-fuel ratio level for actual operating conditions known to be in the region where undesirable combustion events are likely to occur. The margin value can be considered a safety margin.

As long as the engine is in steady-state operation and the controller 60 can confirm that this is the case, the optimization unit will integrate the margin values in small increments over time towards the minimum level p2.

However, if the load signal indicates that the load is below the load threshold or the speed error is above the speed error threshold, processor 60 will conclude that the engine is no longer running in steady state and will cancel the optimization process and The value is set to level p1.

Additionally, if the cylinder pressure estimation unit detects an undesirable combustion event such as a misfire event, partial misfire event and/or pre-ignition event, the air-fuel ratio optimization process is reversed and the air-fuel ratio optimizer returns to the first level p1 when no misfire event is detected. Integrate the margin values in small increments towards the first level p1 until .

In one embodiment, controller 60 controls each combustion chamber individually to increment the margin value slightly and decrease the margin value slightly from the actual value as the predetermined time elapses because the margin is not equal to the second value. is configured to resume decline over time. The predetermined length of time is a predefined period. The predefined period can be in the range of seconds or minutes.

In one embodiment, the controller 60 is configured to decrease the margin value by a predetermined reduction by advancing the exhaust valve closing timing, preferably by advancing the exhaust valve closing timing in small steps.

In one embodiment, the controller 60 is configured to increase the margin value in predetermined increments by retarding the exhaust valve closing timing, preferably by retarding the exhaust valve closing timing in small steps.

In one embodiment, the controller 60 is configured to reduce the margin value by a predetermined reduction by closing or increasing the throttling of the exhaust gas bypass control valve 49.

In one embodiment, the controller 60 is configured to increase the margin value in predetermined increments by opening or reducing the throttling of the exhaust gas bypass control valve 49.

In one embodiment, the controller 60 is configured to increase the margin value in predetermined increments by activating or increasing the speed of the exhaust gas recirculation blower 43.

In one embodiment, the controller 60 is configured to reduce the margin value by a predetermined reduction by deactivating or reducing the speed of the exhaust gas recirculation blower 43.

In one embodiment, the controller 60 is configured to activate liquid fuel injection (FRC) when operating conditions require (e.g., to prevent unreliable ignition or series of misfires), and the liquid It is configured to reset the margin value to the first value (p1) when fuel injection is activated. Liquid fuels are for example diesel oil, marine diesel or other liquid fuels known to have good and reliable ignition properties for compression ignition.

Various aspects and implementations have been described in connection with various embodiments herein. However, other modifications to the disclosed embodiments may be understood and effected by those skilled in the art in practicing the claimed subject matter from a study of the drawings, disclosure, and appended claims. In a claim, the word "comprising" does not exclude another element or step, and the indefinite article "a" or "an" does not exclude plurality. A single processor, controller, or other device may perform the functions of multiple items recited in the claims. The fact that certain measures are recited in different dependent claims does not indicate that a combination of them cannot be used to advantage. Reference signs used in the claims should not be construed as limiting the scope.

Claims (17)

A large two-stroke turbocharged Uniflow scavenge internal combustion engine configured to operate using gaseous fuel as the primary fuel in a gaseous operating mode, said engine comprising:
A plurality of combustion chambers each divided into a cylinder liner (1), a reciprocating piston (10), and a cylinder cover (22);
a scavenging port (18) arranged in the cylinder liner (1) for introducing scavenging air into the combustion chamber;
an exhaust gas outlet disposed in the cylinder cover (22) and controlled by an exhaust valve (4);
A variable timing exhaust valve actuation system capable of controlling the exhaust valve timing for each combustion chamber;
one or more gaseous fuel inlet openings arranged in the cylinder liner (1) or cylinder cover (22) configured to admit gaseous fuel during the stroke of the piston (10) towards the cylinder cover (22), and
Comprising at least one controller (60) associated with the engine,
The at least one controller 60 determines and controls the opening and closing timing of the exhaust valve 4 for each combustion chamber and determines the amount of gaseous fuel flowing into the combustion chamber through the gaseous fuel inlet opening for each combustion chamber. configured to control.
wherein the at least one controller (60) is configured to monitor operating conditions of the engine and determine when the engine is operating at steady-state operating conditions;
said combustion chamber having, at least for steady-state operation, a known undesirable combustion condition such that any one or more of a partial misfire event, a misfire event, or pre-ignition is likely to occur when the air-fuel ratio exceeds a known operating condition dependent threshold level;
said variable timing exhaust valve actuation system allows individual control of exhaust valve timing for each combustion chamber;
At least one of the controllers 60 determines and controls the opening and closing timing of the exhaust valve 4 individually for each combustion chamber and determines the timing of opening and closing the exhaust valve 4 individually for each combustion chamber and controls the gaseous fuel flowing into the combustion chamber through the gaseous fuel inlet opening individually for each combustion chamber. configured to control the amount,
at least one controller (60) is configured to operate in a steady-state operating mode when at least one controller (60) determines that the engine is operating in steady-state conditions;
In the steady-state mode of operation, at least one controller 60:
controlling the air-fuel ratio for each combustion chamber as a function of operating conditions to an air-fuel ratio value that is less than the known operating condition dependent threshold level by a margin value set to a first value (p1) earlier than the known operating condition;
Reducing the margin value, individually for each combustion chamber over time, by a predetermined decrease from an actual value toward a second value (p2), wherein the second value (p2) is the first value. Less than (p1) and greater than 0,
individually monitoring each combustion chamber for partial misfire events, misfire events and pre-ignition events, and
When any one or more of a partial misfire event, a misfire event or a pre-ignition event is detected, a predetermined amount is moved from the actual value toward the first value p1 until the partial misfire event, misfire event or pre-ignition event is no longer detected. Operates by increasing the margin value in increments,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the at least one controller (60) is configured to notify the undesirable combustion condition and the known operating condition according to a threshold level.
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the controller (60) is configured to resume decreasing the margin over time individually for each combustion chamber with a small decrease from the actual value when a predetermined length of time has elapsed since the last increase in the margin. engine that works. value and the value is not equal to the second value (p2),
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the controller (60) is configured to reduce the margin value according to a predetermined reduction by advancing the exhaust valve closing timing.
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the controller (60) is configured to increase the margin value according to a predetermined increment by delaying the exhaust valve closing timing.
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, comprising an exhaust gas bypass (39) with an exhaust gas bypass control valve (49), wherein the controller (60) closes or increases throttling of the gas bypass control valve (49). Configured to reduce the margin value by
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, comprising an exhaust gas bypass (39) with an exhaust gas bypass control valve (49), wherein the controller (60) adjusts the margin value by opening or reducing throttling of the exhaust gas. configured to increase in increments,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, comprising an exhaust gas recirculation conduit (42) with an exhaust gas recirculation blower (43) in the exhaust gas recirculation conduit (42), wherein the controller (60) controls the speed of the exhaust gas recirculation blower (43). Increased by activating or increasing ,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, comprising an exhaust gas recirculation conduit (42) with an exhaust gas recirculation blower (43) in the exhaust gas recirculation conduit (42), wherein the controller (60) controls the speed of the exhaust gas recirculation blower (43). configured to reduce the margin value of reduction by deactivating or reducing ,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The system of claim 1, comprising a cylinder bypass upstream of the main scavenge air cooler, wherein the controller (60) opens the hot cylinder bypass conduit (29) or closes the control valve (31) in the hot cylinder bypass conduit (29). configured to increase the air-fuel ratio by reducing the throttling of or vice versa,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the controller (60) is configured to activate liquid fuel injection (FRC) when operating conditions require, and to reset the margin value to the first value (p1) when liquid fuel injection is activated. felled,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. The method of claim 1, wherein the increment is a small increment, the decrease is a small decrease, and the step is a small step.
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, comprising a sensor for individually sensing the cylinder pressure for each cylinder, wherein the controller is configured to individually monitor the sensed cylinder pressure for each cylinder, wherein the controller is configured to individually monitor the sensed cylinder pressure for each cylinder. configured to individually determine whether a misfire event, partial misfire and/or pre-ignition event has occurred in the corresponding cylinder,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 14. The method of claim 13, wherein the controller (60) is further configured to determine a deviation of the cylinder pressure evolution from the expected evolution of the cylinder pressure when the misfire event, partial misfire event and/or pre-ignition event occurs by determining the deviation of the cylinder pressure evolution from the expected evolution of the cylinder pressure. /or configured to determine a pre-ignition event,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the controller (60) determines that the engine is operating under steady-state conditions when the difference between the desired engine speed and the actual engine speed is less than the deviation threshold, while at the same time the engine load is above the engine load threshold. Consisting to decide that,
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 2. The method of claim 1, wherein the at least one controller (60) includes or is connected to a compressed air-fuel ratio observer (46) for determining an instantaneous average compressed air-fuel ratio in the combustion chamber.
Large two-stroke turbocharged Uniflow scavenge internal combustion engine. 1. A method of operating a large two-stroke turbocharged uniflow scavenge internal combustion engine with multiple combustion chambers in gaseous operation mode, wherein an air-fuel ratio mixture having an air-fuel ratio is present in the combustion chamber prior to ignition;
said combustion chamber having, at least for steady-state operation, a known undesirable combustion condition such that any one or more of a partial misfire event, a misfire event, or pre-ignition is likely to occur when the air-fuel ratio exceeds a known operating condition dependent threshold level;
The above method is:
monitor the operating conditions of the engine and determine when the engine is operating at steady-state operating conditions;
When steady-state operating conditions are determined:
individually controlling the air-fuel ratio for each combustion chamber as a function of operating conditions with a value for the air-fuel ratio being less than said known operating condition dependent threshold level by a margin value initially set to a first value (p1);
Reducing the margin value, individually for each combustion chamber over time, by a predetermined decrease from an actual value toward a second value (p2), wherein the second value (p2) is equal to the first value (p2). Less than value (p1) and greater than 0,
individually monitoring each combustion chamber for partial misfire events, misfire events and pre-ignition events, and
Upon detection of any one or more of a partial misfire event, a misfire event, or a pre-ignition event, a predetermined decrease is made from the actual value toward the first value p1 until the partial misfire event, misfire event, or pre-ignition event is no longer detected. Including increasing the margin value in increments,
A method of operating a large two-stroke turbocharged Uniflow scavenge internal combustion engine with multiple combustion chambers in gaseous operation mode.