JP7230265B2 - Large two-stroke turbocharged uniflow scavenging internal combustion engine and method of operating same - Google Patents

Description

This specification relates to a large two-stroke internal combustion engine using gas fuel, in particular a large two-stroke crosshead uniflow scavenging internal combustion engine operated by gas fuel introduced through a fuel valve on the way of the piston from BDC to TDC. Regarding institutions.

background

A large two-stroke uniflow scavenging internal combustion engine of a crosshead type is used, for example, as a propulsion system for a large ship or as a prime mover of a power plant. The size of this large two-stroke engine is huge. Not only because of its enormous size, this large two-stroke engine has a different construction than other internal combustion engines. For example, the weight of the exhaust valve can reach 400 kg and the diameter of the piston can reach 100 cm. The maximum pressure in the combustion chamber during operation is typically several hundred bar. The forces generated by such high pressure levels and piston sizes are enormous.

Some large two-stroke turbocharged internal combustion engines are operated with gas fuel introduced from a fuel valve located near the longitudinal center of the cylinder liner or on the cylinder cover. In this type of engine, gas fuel is introduced into the cylinder during the upstroke of the piston, well before the exhaust valve closes. The engine compresses a mixture of gaseous fuel and scavenging air in the combustion chamber and ignites the compressed mixture at or near top dead center (TDC) by synchronous ignition means (eg, pilot oil ignition means).

In a large two-stroke turbocharged internal combustion engine, when the piston injects gaseous fuel at or near top dead center (TDC), the compression pressure in the combustion chamber is near maximum. In comparison, the above type of gas introduction using a fuel valve (gas inlet valve) located in the cylinder liner or cylinder cover injects gaseous fuel when the pressure in the combustion chamber is relatively low, resulting in a significantly lower fuel injection rate. It has the advantage that pressure can be used. For engines of the type that inject gaseous fuel at or near TDC, the fuel injection pressure must be achieved well above the pressure in the combustion chamber, which is already near maximum pressure. Fuel systems capable of handling gaseous fuels at such extremely high pressures are expensive and complex. Reasons for this include the volatility of gaseous fuels and their behavior under high pressure, which allows them to diffuse into (or through) the steel components of the fuel system.

Thus, a fuel delivery system for an engine that injects gas fuel midway through the compression stroke costs much less than that for an engine that injects gas fuel at high pressure when the piston is near TDC.

However, when injecting gas fuel in the middle of the compression stroke, the piston will compress a mixture of gas fuel and scavenging air, with the risk of pre-ignition. By operating with a very lean mixture, the risk of pre-ignition can be reduced. However, using a lean mixture increases the risk of misfire or partial misfire, resulting in fuel slip.

It is therefore desirable to improve the control of combustion chamber conditions during compression in order to eliminate or at least reduce the problems of misfire, pre-ignition and diesel knock in large two-stroke turbocharged internal combustion engines such as those described above. There is a need for In order to prevent pre-ignition and misfires from occurring, it is necessary to control the state of the combustion chamber very precisely.

Engine performance design is typically such that pre-ignition does not occur when the engine is operating at steady state. This is achieved through careful selection of combustion chamber design, fuel injection timing, and exhaust valve timing. However, this requires operating at a safe distance from combustion conditions where misfires, partial misfires and pre-ignition are likely to occur. This large safety distance results in less than optimal combustion conditions. This results in less than optimal combustion conditions, especially in terms of fuel efficiency.

Dk201970370 discloses a large two-stroke turbocharged uniflow scavenging internal combustion engine. The engine comprises a plurality of combustion chambers and at least one controller associated with the engine for determining an average compressed air-fuel ratio and bulk compression temperature at the start of combustion;
- if the determined average compressed air-fuel ratio is below the lower compressed air-fuel ratio threshold, then taking a strategy to increase the compressed air-fuel ratio;
- if the determined average compressed air excess ratio exceeds the compressed air-fuel ratio upper threshold, then implementing a strategy to reduce the compressed air-fuel ratio;
- if the determined or measured bulk compaction temperature is below the bulk compaction temperature below threshold, performing at least one strategy to increase the bulk compaction temperature;
- if the determined or measured bulk compaction temperature is above the bulk compaction temperature threshold, performing at least one strategy to reduce the bulk compaction temperature;

configured to perform

Summary

It is an object to provide an engine and method that solves or at least mitigates the above-mentioned problems.

The above-mentioned and other problems are solved by the features of the independent claims. More specific implementations will become apparent from the dependent claims, the detailed description of the invention and the drawings.

According to a first aspect, there is provided a large two-stroke turbocharged uniflow scavenging internal combustion engine operating on gas fuel as the primary fuel in a gas mode of operation. This institution
a plurality of combustion chambers each defined by a cylinder liner, a reciprocating piston and a cylinder cover;
a scavenging port for introducing scavenging air into the combustion chamber, the scavenging port disposed in the cylinder liner;
an exhaust gas outlet disposed in the cylinder cover and controlled by an exhaust valve;
a variable timing exhaust valve actuation system that allows control of exhaust valve timing for each combustion chamber individually;
one or more gas inlets disposed in the cylinder liner or the cylinder cover and configured to introduce gaseous fuel during the stroke of the reciprocating piston into the cylinder cover;
at least one controller associated with the engine;
The at least one controller is configured to individually determine and control the opening and closing timing of the exhaust valve for each combustion chamber, and for each combustion chamber individually, the gas is introduced into the combustion chamber from the inlet. configured to control the amount of gas fuel,
the at least one controller configured to monitor operating conditions of the engine and determine when the engine is operating at steady state operating conditions;

the at least one controller is configured to operate in a steady state mode when determining that the engine is operating in a steady state operating condition;

said combustion chamber, at least at steady state operation, having known undesired combustion conditions, wherein said known undesired combustion conditions, when an air-fuel ratio exceeds a known operating condition dependent critical value, a partial misfire event; Misfire events and/or preignition are likely to occur,

The at least one controller in the steady state mode comprises:
- for each combustion chamber individually, controlling the air-fuel ratio as a function of operating conditions, with a margin initially set to a first value such that said air-fuel ratio value is below said known operating-condition dependent critical value; control to
decrementing the margin from the actual value to a second value less than the first value and greater than zero for each combustion chamber individually;
monitor for misfire events, partial misfire events, and preignition events for each combustion chamber separately;
upon detection of a partial misfire event, misfire event and/or preignition event, reducing said margin from the actual value until no misfire event, partial misfire event or preignition event is detected; increments towards the value,
configured as

Potential for optimizing engine operation and combustion processes, allowing engine operating parameters to be used at values close to critical values by allowing a smaller margin compared to the values under normal steady-state operating conditions. increase These engine operating parameters are, for example, (maximized) fuel efficiency (energy efficiency), (minimized) NOx emissions, (minimized) hydrocarbon slip (HC slip) (hydrocarbons not combusted fuel or partially combusted fuel). Engines can therefore be designed to optimize any of these operating parameters.

This optimization specifies how far the actuator setpoints deviate from normal steady-state operating conditions in order to push the combustion process in the optimal direction. To determine how much uncontrollable factors (such as outside air conditions and component maintenance) can be optimized, the only way to optimize is to specify deviation rules. The optimization strategy is then at least partially reversed when undesirable combustion conditions are detected. Therefore, uncontrollable factors will determine to what extent the rule (size of actuator setpoint deviation) is applicable.

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

An operating condition dependent critical value of air-fuel ratio for steady state operation is determined during engine design and/or based on test runs and/or based on computer simulations.

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

As soon as the controller determines that the engine is operating in steady state conditions, or with a predetermined delay (i.e., after a predetermined time), the controller begins reducing the margin from the initial value it initially set.

In one example of implementation of the first aspect, the at least one controller is informed of the undesirable combustion state and the known operating condition dependent critical value.

In one example of an implementation of the first aspect, when a predetermined amount of time has elapsed since the value was last increased and when the value is not equal to the second value: Individually for each combustion chamber, the margin is arranged to be decremented smaller from the actual value and to resume the gradual decrease.

In one example implementation of the first aspect, the controller is configured to decrement the margin smaller by advancing exhaust valve closing timing, preferably in steps.

In one example implementation of the first aspect, the controller is configured to incrementally increase the margin by retarding exhaust valve closing timing, preferably in stages.

In one example of implementation of the first aspect, the engine comprises an exhaust bypass having an exhaust bypass control valve, and the controller decrements the margin by closing or tightening the exhaust bypass control valve. configured to be as small as possible.

In one example of an implementation of the first aspect, the engine comprises an exhaust bypass having an exhaust bypass control valve, and the controller increments the margin by opening or throttling the exhaust bypass control valve. configured to grow

In one example of an implementation of the first aspect, the engine has an exhaust recirculation pipe with an exhaust recirculation blower therein, and the controller activates or speeds up the exhaust recirculation blower. is configured to increment and enlarge the margin.

In one example of implementation of the first aspect, the engine has an exhaust recirculation pipe with an exhaust recirculation blower therein, and the controller deactivates or slows down the exhaust recirculation blower. Thereby, the margin is decremented to be smaller.

In one example of an implementation of the first aspect, the engine has a cylinder bypass upstream of a main scavenge air cooler, and the controller opens a hot cylinder bypass line or controls a control valve in the hot cylinder bypass line. The air-fuel ratio is raised by loosening the throttle, and the air-fuel ratio is lowered by closing the hot cylinder bypass pipe or tightening the throttle of the control valve in the hot cylinder bypass pipe.

In one example of an implementation of the first aspect, the controller is configured to activate injection of liquid fuel (e.g., diesel oil) when operating conditions dictate; Sometimes configured to reset the margin to the first value.

In one example implementation of the first aspect, the increment is a small increment, the decrement is a small decrement, and the steps are small steps.

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

In one example implementation of the first aspect, the controller controls the deviation of cylinder pressure change from an expected change in cylinder pressure when no misfire event, partial misfire event, and/or pre-ignition event occurs. is configured to determine a misfire event, a partial misfire event, and/or a preignition event.

In one example of an implementation of the first aspect, the controller controls the engine to enter steady state if the difference between the desired engine speed and the actual engine speed is below a deviation threshold and at the same time the engine load is above the engine load threshold. configured to determine that it is operating in

In one example of the implementation of the first aspect, one or more gas fuel introduction holes are provided so as to introduce gas fuel supplied from a pressurized gas fuel supply unit through a fuel introduction valve into the combustion chamber. Configured.

According to a second aspect, there is provided a method of operating a large two-stroke turbocharged uniflow scavenge internal combustion engine having multiple combustion chambers in a gas mode of operation. wherein a mixture having an air-fuel ratio prior to combustion exists in the combustion chamber,
said combustion chamber, at least at steady state operation, having known undesired combustion conditions, wherein said known undesired combustion conditions, when an air-fuel ratio exceeds a known operating condition dependent critical value, a partial misfire event; Misfire events and/or preignition are likely to occur,
and the method
monitoring operating conditions of the engine to determine when the engine is operating at steady state operating conditions;
Having determined the steady state operating conditions,
Controlling the air/fuel ratio as a function of operating conditions for each combustion chamber individually, with a margin initially set to a first value such that said air/fuel ratio value is below said known operating condition dependent critical value. control and
decrementing the margin from the actual value to a second value less than the first value and greater than zero for each combustion chamber individually;
monitor misfire, partial misfire, and preignition events for each combustion chamber separately;
upon detection of a partial misfire event, misfire event, and/or preignition event, reducing said margin from the actual value to said first value until no misfire event, partial misfire event, or preignition event is detected; Increment toward to make it bigger.

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

Hereinafter, various concepts, embodiments, and implementation examples will be described in detail with reference to exemplary embodiments shown in the drawings.
1 is a front view of a large two-stroke diesel engine in accordance with certain exemplary embodiments; FIG. 2 is a side view of the large two-stroke engine of FIG. 1; FIG. 2 is a first schematic representation of the large two-stroke engine of FIG. 1; Figure 2 is a cross-sectional view of the cylinder frame and cylinder liner of the engine of Figure 1; Cylinder covers and exhaust valves are attached, and the pistons at TDC and BDC are also depicted. Figure 2 is a second schematic representation of the engine of Figure 1; 4 is a schematic representation of a compression temperature observer and a compression air-fuel ratio observer; A graph is drawn with the horizontal axis representing the bulk cylinder temperature and the vertical axis representing the compressed air-fuel ratio. A safe zone and surrounding zones are shown where some action must be taken to return to the safe zone. 1 depicts an embodiment of a method for controlling a large two-stroke engine; FIG. FIG. 4 depicts the optimization process for individual cylinders;

Detailed explanation

In the following detailed description, the internal combustion engine will be described with reference to the crosshead type large low speed two-stroke turbocharged internal combustion engine of the embodiment. 1-3 depict an example of a turbocharged large low speed two-stroke diesel engine. This engine has a crankshaft 8 and a crosshead 9 . 1 is a front view, and FIG. 2 is a side view. FIG. 3 is a schematic representation of the turbocharged large low-speed two-stroke diesel engine of FIGS. 1 and 2 together with its intake and exhaust systems. In this example, the engine has four cylinders in series. Turbocharged large low speed two-stroke internal combustion engines typically have from 4 to 14 cylinders arranged in series. These cylinders are carried on the engine frame 11 . Such an engine can also be used, for example, as a main engine on ships or as a stationary engine for powering generators in power plants. The total engine power can be, for example, 1000 kW to 110000 kW.

The engine in this embodiment is a two-stroke uniflow scavenging engine, provided with a scavenging port 18 in the lower region of the cylinder liner 1 and an exhaust valve 4 centrally located at the top of the cylinder liner 1 . Scavenging air is directed from the scavenging receiver 2 to the scavenging port 18 of each cylinder liner 1 when the piston is below the scavenging port 18 . Gas fuel is introduced from the gas fuel introduction valve 30 under the control of the electronic control unit 60 . This is done during the upstroke of the piston and before the piston passes the fuel valve (gas fuel inlet valve) 30 . Gas fuel is introduced at a relatively low pressure, less than 30 bar, preferably less than 25 bar, more preferably less than 20 bar. The fuel valves are preferably evenly distributed around the circumference of the cylinder liner. Also preferably, it is arranged near the center in the longitudinal direction of the cylinder liner. The introduction of gaseous fuel therefore takes place when the compression pressure is relatively low. That is, it can be introduced at a relatively low pressure, since it is done at a much lower compression pressure than the piston reaches TDC.

A piston 10 compresses a mixture of gaseous fuel and scavenging air in the cylinder liner 1 . Compression occurs and ignition is initiated at or near TDC. Ignition may be initiated, for example, by injecting pilot oil (or other suitable ignition fluid) from pilot oil fuel valve 50 . A pilot oil fuel valve 50 is preferably located in the cylinder cover 22 . Combustion then occurs and exhaust gases are produced. Other forms of ignition systems may use prechambers, laser ignition, glow plugs (none shown), etc. to facilitate ignition instead of or in addition to pilot oil.

When the exhaust valve 4 is open, the exhaust gas flows through the exhaust duct provided in the cylinder 1 to the exhaust receiver 3 and further through the first exhaust pipe 19 to the turbine 6 of the turbocharger 5 . From there the exhaust gases flow through a second exhaust pipe 25 to the economizer 20 and out of an outlet 21 to the atmosphere. Turbine 6 drives compressor 7 via a shaft. Outside air is supplied to the compressor 9 through an air intake 12 . The compressor 7 sends the compressed scavenging air to the scavenging pipe 13 connected to the scavenging receiver 2 . The scavenging air in tube 13 passes through an intercooler 14 for cooling the scavenging air.

The cooled scavenging air passes through an auxiliary blower 16 driven by an electric motor 17 . The auxiliary blower 16 compresses the scavenging air flow when the compressor 7 of the turbocharger 5 cannot supply the scavenging air receiver 2 with the required pressure, ie when the engine is under low or part load. At high engine loads, the turbocharger compressor 7 can supply sufficiently compressed scavenging air so that the auxiliary blower 16 is bypassed by the non-return valve 15 .

FIG. 4 shows a cylinder liner 1 designed for a large crosshead two-stroke engine. Depending on the size of the engine, the cylinder liner 1 is made in various sizes. Typical sizes are diameters of 250 mm to 1000 mm and corresponding total lengths of 1000 mm to 4500 mm.

FIG. 4 depicts a state in which the cylinder liner 1 is mounted on the cylinder frame 23 and the cylinder cover 22 is mounted on the cylinder liner 1 . The cylinder liner 1 and the cylinder cover 22 are connected so that gas does not leak therebetween. In FIG. 4, the states of the piston 10 at its bottom dead center (BDC) and top dead center (TDC) are indicated by dashed lines. Of course, these two states do not occur simultaneously, but are separated by 180 degrees of rotation of the crankshaft 8 . The cylinder liner 1 is provided with a cylinder lubrication hole 25 and a cylinder lubrication line 24 . These supply cylinder lubrication oil as the piston 10 passes through the lubrication line 24 . Piston rings (not shown) then distribute the cylinder lubricating oil over the entire running surface of the cylinder liner. The above engines typically have a geometric compression ratio between 8 and 15. However, the geometric compression ratio can exceed 20 for engines with high pressure gas injection mechanisms that inject gaseous fuel from fuel injectors located in the cylinder covers at high pressures near TDC.

A pilot oil valve 50 or a pre-chamber with a pilot oil valve 50 is normally mounted on the cylinder cover 22 . Pilot oil valve 50 Normally one is mounted on each cylinder. Pilot oil valve 50 is connected to a source of pilot oil, not shown. The injection timing of the pilot oil is controlled by the electronic control unit 60 .

A fuel valve 30 is mounted on the cylinder liner 1 (or the cylinder cover 22). The fuel valve 30 has a nozzle that lies substantially flush with the inner surface of the cylinder liner 1 . Also, the rear end of the fuel valve 30 protrudes from the outer wall of the cylinder liner 1 . Typically one or two, at most three or four fuel valves 30 are provided in each cylinder liner 1 . They are arranged (preferably evenly spaced) around the circumference of the cylinder liner 1 . In this embodiment, the fuel valve 30 is arranged exactly in the center of the cylinder liner 1 in the longitudinal direction.

FIG. 4 shows a schematic representation of a gas fuel supply system comprising a pressurized gas fuel source 40 connected to respective inlets of a plurality of gas fuel valves 30 through a gas fuel supply tube 41 .

FIG. 5 is a schematic representation of an engine similar to that of FIG. 2, but with the gas exchange apparatus of the engine drawn in detail. Outside air is taken in at a pressure and temperature similar to the surroundings and fed through the air inlet 12 to the compressor 7 of the turbocharger 5 . Compressed scavenging air is sent from the compressor 7 through the scavenging pipe 13 to the branch point 28 .

The branch point 28 allows the scavenging air to branch through the hot cylinder bypass pipe 29 to the turbine connection 32 of the first exhaust pipe 19 . The flow rate of hot cylinder bypass pipe 29 is controlled by hot cylinder bypass control valve 31 . Hot cylinder bypass control valve 31 is electronically controlled by controller 60 . Opening the hot cylinder bypass pipe 29 or loosening the throttle of the hot cylinder bypass control valve 31 has the effect of increasing the air-fuel ratio and also has the effect of increasing the bulk compression temperature. On the other hand, closing the hot cylinder bypass 29 or tightening the throttle of the hot cylinder bypass control valve 31 has the effect of lowering the air-fuel ratio and also has the effect of lowering the bulk compression temperature.

The scavenging pipe 13 is provided with a first scavenging control valve 33 upstream of the intercooler 14 . A second scavenging control valve 34 is provided downstream of the intercooler 14 . The scavenging pipe 13 is connected to the scavenging receiver 2 . A pipe with an auxiliary blower 16 branches off from the intercooler 14 .

A cold cylinder bypass pipe 35 connects the scavenging receiver 2 to the turbine connection 32 of the first exhaust pipe 19 . A cold cylinder bypass control valve 36 controls the flow rate of the cold cylinder bypass pipe 35 . Cold cylinder bypass control valve 36 is electronically controlled by controller 60 . Opening the cold cylinder bypass 35 or throttling the cold cylinder bypass control valve 36 has the effect of increasing the bulk compression temperature.

A cold scavenging air bypass pipe 37 allows scavenging air to escape from the scavenging pipe 2 to the ambient environment 26 . The cold scavenging bypass pipe 37 flow rate is controlled by a cold scavenging bypass control valve 38 . Cold scavenging bypass control valve 38 is electronically controlled by controller 60 . Opening the cold scavenging bypass control valve 39 or loosening the throttle of the cold scavenging bypass control valve 38 has the effect of lowering the scavenging pressure and has the effect of lowering the air-fuel ratio. On the other hand, closing the cold scavenging bypass control valve 39 or tightening the throttle of the cold scavenging bypass control valve 38 has the effect of increasing the scavenging pressure and has the effect of increasing the air-fuel ratio. The cold scavenging bypass pipe 37 does not need to branch from the scavenging receiver 2 and may branch from any position in the scavenging pipe 13 downstream of the intercooler 14 .

An exhaust recirculation pipe 42 connects between the exhaust receiver 3 and the scavenging receiver 2 . The exhaust recirculation pipe 42 has an exhaust recirculation control valve 45 , a recirculation exhaust gas cooler 44 and a recirculation exhaust gas blower 43 . Both the recirculation exhaust gas blower 43 and the exhaust recirculation control valve 45 are used to regulate the flow rate of the exhaust recirculation line 42 under the electronic control of the controller 60 . Under normal operating conditions, no exhaust flows through the exhaust recirculation line 42 unless the recirculation exhaust gas blower 43 is activated. However, this is because the pressure in the exhaust receiver 3 is usually lower than the pressure in the scavenging receiver 2 . Therefore, when the recirculation exhaust gas blower 43 is not active, the exhaust gas recirculation control valve 45 must be closed. The exhaust gas recirculation pipe 42 may not be directly connected to the exhaust receiver 3 and may be connected to some position on the first exhaust pipe 19 . Further, the exhaust gas recirculation pipe 42 may not be directly connected to the scavenging receiver 2 and may be connected to any position of the scavenging pipe 13 as long as it is downstream of the intercooler 14 .

Activating the recirculation exhaust gas blower 43 or increasing the speed of the recirculation exhaust gas blower 43 in the exhaust recirculation pipe 42 lowers the air/fuel ratio and also slightly lowers the bulk compression temperature. On the other hand, in the exhaust recirculation pipe 42, deactivating the recirculation exhaust gas blower 43 or lowering the recirculation exhaust gas blower 43 speed increases the air/fuel ratio and also slightly increases the bulk compression temperature.

An exhaust bypass pipe 39 branches off from the exhaust receiver 3 or the first exhaust pipe 19 and is connected to the ambient environment 27 at a given back pressure. Exhaust bypass control valve 49 is used to regulate the flow rate of exhaust bypass pipe 39 under the electronic control of controller 60 .

Opening the exhaust bypass control valve 49 or loosening the throttle of the exhaust bypass control valve 49 lowers the air-fuel ratio in the cylinder. On the other hand, closing the exhaust bypass control valve 49 or tightening the throttle of the exhaust bypass control valve 49 increases the air-fuel ratio in the cylinder.

In an engine with a selective catalytic reduction (SCR) reactor and a reactor bypass valve, the flow rate passing through the SCR reactor out of the flow rate from the exhaust receiver 3 to the turbine 6 of the turbocharger 5 is controlled under the electronic control of the controller 60. adjusted.

In FIG. 5, connections to controller 60 of all the above-described elements controlled by controller 60 are represented using dashed lines.

FIG. 6 is a diagram for explaining the air-fuel ratio observer 46 and the bulk compression temperature observer.

The air/fuel ratio observer 46 is a computer implemented algorithm that uses information about scavenge pressure, exhaust valve closing timing, cylinder geometry, stoichiometric air/fuel ratio, and amount of gas injected. Compressed air-fuel ratio observer 46 may be implemented as part of controller 60 or may be implemented as a computer or controller separate from controller 60 . Compressed air-fuel ratio observer 46 provides as an output an estimate of the compressed air-fuel ratio for the (fully) compressed mixture (i.e., the mixture when piston 10 is at TDC), which is sent to controller 60 . This estimate is based on the ratio of the mass of ambient air trapped in the combustion chamber when the exhaust valve 4 is seated divided by the mass of air required to completely burn the total mass of injected gas. is.

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

FIG. 7 is a graph of bulk compression temperature (Tcomp) versus air-fuel ratio (λ). The zone within the boundaries defined by Air-Fuel Ratio Lower Threshold, Air-Fuel Ratio Upper Threshold, Bulk Compression Temperature Lower Threshold, and Bulk Compression Temperature Upper Threshold is the steady state default zone 51 . In this steady-state default zone 51, the controller 60 provides each cylinder individually with the amount of fuel required for the current engine load and takes no measures to vary the bulk compression temperature. The controller 60 also individually controls the air/fuel ratio of each cylinder 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. These known undesirable combustion conditions are conditions in which pre-ignition or events such as partial or misfires are likely to occur when the air/fuel ratio exceeds a known critical level depending on operating conditions. The first margin level has a first value greater than zero.

When the combustion conditions in the cylinder liner 1 change and leave the steady state default zone 51 and enter the action zone 52, the controller 60 takes measures to prevent such an occurrence.

To this end, controller 60 is configured to perform the following for each cylinder individually.
If the determined or measured average compressed air-fuel ratio is below the lower compressed air-fuel ratio threshold, then implement at least one measure to increase the compressed air-fuel ratio (Compression Air-fuel Ratio Increasing Measure; CAFRIM).
If the determined or measured average compressed air-fuel ratio is above a compressed air-fuel ratio upper threshold, implement at least one Compression Air-fuel Ratio Decreasing Measure (CAFRDM).
• If the determined or measured bulk compression temperature is below the bulk compression temperature lower threshold, implement at least one measure to increase the bulk compression temperature (Bulk Compression Temperature Increasing Measure; BCTIM).
• If the determined or measured bulk compression temperature is above the bulk compression temperature upper threshold, implement at least one measure for decreasing the bulk compression temperature (Bulk Compression Temperature Decreasing Measure; BCTDM).

By taking these measures, the controller 60 keeps the conditions within each cylinder liner 1 within the normal zone 51 . The state inside the cylinder liner 1 is made to go out of the normal zone 51 and into the action zone 52 only temporarily. Action zone 52 is surrounded by danger zone 53 . The danger zone 53 is a zone in which an event such as preignition or misfire is highly likely to occur.

Zones 51, 52, 53 are bounded by upper and lower bulk compression temperature and compressed air-fuel ratio thresholds. These thresholds can be empirically and experimentally determined for each engine, for example, by trial and error, or by computer simulation of engine cycles.

If observers 46 , 47 indicate that both the bulk compression temperature and the compressed air-fuel ratio are outside the steady state default zone 51 , the controller 60 adjusts the conditions within the cylinder liner 1 to within the steady state default zone 51 . To return, both the strategy of returning the bulk compression temperature to the steady-state default zone 51 and the strategy of returning the compressed air-fuel ratio to the steady-state default zone 51 are performed for each cylinder individually.

Adjusting the exhaust bypass control valve 49 (moving the exhaust bypass control valve 49 to the open position) to open the Exhaust Gas Bypass (EGB) pipe 39 (from the supercharger turbine inlet to the turbine outlet or increasing the flow to the outside environment) results in a lower scavenging pressure and thus less air mass trapped in the combustion chamber. Therefore, this measure is suitable as a measure for lowering the compressed air-fuel ratio. This measure has only a minor effect on the bulk compaction temperature. Even if the engine has multiple turbochargers, only one EGB may be sufficient by connecting the EGB to the exhaust receiver. However, the connection must be chosen so that the multiple streams to the exhaust receiver are properly mixed.

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

Opening the scavenging bypass control valve 38 creates flow from the scavenging receiver 2 to the compressor inlet or the outside environment. This has a qualitatively similar effect on the compressed air-fuel ratio as the exhaust bypass. However, the scavenging process (and thus the bulk compression temperature of the combustion chamber) is affected differently. The effect of opening the scavenge bypass control valve 38 on combustion chamber conditions is more rapid than with exhaust bypass.

Opening the cold cylinder bypass 36 increases the flow from the exhaust receiver to the supercharger turbine inlet. This results in an increase in bulk compaction temperature. However, it has little effect on the compressed air-fuel ratio.

Exhaust valve closing timing determines the ratio of compression to scavenging pressure in the combustion chamber. Varying this timing has a large effect on both the compressed air-fuel ratio in the combustion chamber and the bulk compression temperature.

Exhaust valve opening timing affects the early stages of the scavenging process within the combustion chamber. Varying this timing affects engine efficiency and the scavenging process. Changes in the scavenging process result in changes in bulk temperature. If the exhaust valve 4 opens too early, there will be no flow to the scavenging receiver 2 when the piston 10 opens the scavenging port 18 . If the exhaust valve 4 opens too late, there will be a large flow to the scavenging receiver 2 when the piston 10 opens the scavenging port 18 . These measures change the scavenging process and change the rate at which the 'dirty/hot' gases produced by the combustion are entrained in the next combustion stroke.

That is, by opening the exhaust valve 4 later, more 'dirty/hot' gases are entrained from the previous combustion, lowering the compressed air/fuel ratio and increasing the bulk compression temperature. Closing the exhaust valve 4 very late reduces the entrained 'dirty/hot' gases from the previous combustion, increases the compressed air/fuel ratio and lowers the bulk compression temperature. Closing the exhaust valve 4 earlier to increase the compression reduces the amount of gas leaving the exhaust valve 4 and leaves more gas in the combustion chamber. This increases the air-fuel ratio. Increasing compression increases the compression work done on the gases by the piston 10 in the combustion chamber. This raises the temperature of the gases in the combustion chamber.

Increasing the flow of exhaust recirculation increases the flow of exhaust from the exhaust receiver 3 to the turbocharger compressor outlet (or scavenging receiver). The flow rate of exhaust gas recirculation can be increased by activating the exhaust gas recirculation blower 43 or increasing the rotational speed of the exhaust gas recirculation blower 43 . As the flow of exhaust gas recirculation increases, the compressed air-fuel ratio decreases.

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

For engines with water injection, injecting water into the combustion chamber during the compression stroke reduces the bulk compression temperature.

Scavenging air cooler bypass (not shown): Bypassing the intercooler 14 greatly increases the bulk compression temperature of the combustion chamber. On the other hand, it has only a small effect on the compressed air-fuel ratio.

In engines with variable geometry turbines 6, reducing the turbine flow area results in an increase in scavenging air pressure, thus reducing the trapped air mass in the combustion chamber. Therefore, this measure is suitable as a measure for lowering the compressed air-fuel ratio. This measure has only a minor effect on the bulk compaction temperature.

For engines with turbocharger assist, increasing the speed of the turbocharger 5 with more assist results in an increase in the compressed air-fuel ratio. This measure has only a minor effect on compression temperature.

Altering the ratio of gas fuel to liquid fuel (eg diesel or marine diesel) is also a strategy. Reducing the proportion of gas fuel in the total fuel energy injected increases the compressed air-fuel ratio during compression. A corresponding increase in the proportion of liquid fuel ensures that the crankshaft torque is maintained.

For engines with a heat exchanger mounted in the exhaust receiver (or for engines with a heat exchanger that receives a portion of the exhaust gas), increasing the amount of exhaust gas passing through the heat exchanger, i.e. Extracting more heat results in a lower scavenging pressure and thus less air trapped in the combustion chamber. Therefore, this measure is suitable as a measure for lowering the compressed air-fuel ratio. This measure has only a minor effect on the bulk compaction temperature. A heat exchanger may be used to generate steam.

For engines with hot scavenging bypass, opening the hot scavenging bypass control valve creates or increases flow from the compressor outlet to the environment or to the compressor inlet. This greatly reduces the scavenging pressure. Therefore, the amount of air trapped in the combustion chamber is reduced. Therefore, this measure is suitable as a measure for lowering the compressed air-fuel ratio.

In one embodiment, the Compressed Air-Fuel Ratio Lower Threshold, Compressed Air-Fuel Ratio Upper Threshold, Bulk Compression Temperature Lower Threshold, and Bulk Compression Temperature Upper Threshold are all engine operating condition dependent parameters. Engine operating conditions are determined by parameters such as engine load, ambient temperature, ambient humidity, and engine speed. These operating condition parameters are available to controller 60 through, for example, lookup tables, algorithms, combinations thereof, and the like.

In some embodiments, controller 60 is configured as follows.
• If the determined or measured average compressed air-fuel ratio is below a minimum compressed air-fuel ratio threshold that is less than the compressed air-fuel ratio lower threshold, then take further measures to increase the compressed air-fuel ratio. This further measure may for example be selected from the measures described above.
If the determined or measured average compressed air-fuel ratio exceeds a maximum compressed air-fuel ratio threshold greater than a compressed air-fuel ratio upper maximum threshold greater than a compressed air-fuel ratio upper threshold, then take further measures to reduce the compressed air-fuel ratio.
• If the determined or measured bulk compaction temperature is below a minimum bulk compaction temperature threshold that is less than the bulk compaction temperature below threshold, then take further measures to increase the bulk compaction temperature.
• If the determined or measured bulk compaction temperature is above the maximum bulk compaction temperature threshold above the bulk compaction temperature upper threshold, then take further measures to lower the bulk compaction temperature.

These further measures are taken when combustion chamber conditions move from the action zone 52 to the danger zone surrounding the action zone. Controller 60 is configured to take as many necessary actions as possible to return combustion chamber conditions to action zone 52 and to steady state default zone 51 .

The controller 60 is configured to minimize taking action to return the operating conditions of the engine to the steady state default zone 51 (ie, taking the above-described measures). The controller 60 is therefore configured to cease all of the above measures when the combustion chamber conditions return to the normal operating zone.

FIG. 8 is a flow chart for explaining the process of operating the engine according to the configuration of the controller 60 described above.

When processing begins, the controller checks to see if the compressed air-fuel ratio is below the lower threshold. If the result of the inquiry is No, the controller proceeds to investigate whether the engine is operating at steady state. If the result of the inquiry is Yes, the controller proceeds to perform the air/fuel ratio optimization process. This process is described in detail with reference to FIG. If the result of the inquiry is No, the controller 60 investigates whether the compressed air-fuel ratio exceeds the upper threshold. If the result of the inquiry is Yes, controller 60 implements one of the strategies described above to increase the compressed air-fuel ratio. Controller 60 then checks to see if the compressed air-fuel ratio is below a minimum threshold. If the result of the inquiry is No, the controller proceeds to investigate whether the compressed air-fuel ratio exceeds the upper threshold. If the result of the inquiry is yes, the controller 60 performs one more of the above-described measures of increasing the compressed air-fuel ratio and proceeds to the step of examining whether the compressed air-fuel ratio exceeds the upper threshold.

The controller 60 checks to see if the compressed air-fuel ratio is above the upper threshold. If the result of the inquiry is No, the controller moves to investigate whether the Bulk Compression Under Temperature Threshold has been violated. If the result of the inquiry is Yes, the controller 60 performs one of the above to reduce the compressed air-fuel ratio. The controller 60 checks to see if the compressed air-fuel ratio is above the maximum threshold. If the result of the inquiry is No, the controller moves to investigate whether the Bulk Compression Under Temperature Threshold has been violated. If the result of the inquiry is yes, the controller 60 performs one more of the above-described measures of lowering the compressed air-fuel ratio and moves to the inquiry step of whether the bulk compression temperature is below the lower threshold.

The controller 60 checks to see if the bulk compaction temperature is below the lower threshold. If the result of the inquiry is No, the controller 60 moves to the next step, which is to investigate whether the bulk compaction temperature is above the upper threshold. And if the result of the inquiry is Yes, the controller 60 implements measures to increase the bulk compaction temperature. Controller 60 then checks to see if the bulk compaction temperature is below a minimum threshold. If the result of the inquiry is No, the controller 60 proceeds to inquire whether the bulk compaction temperature is above the upper threshold. If the result of the inquiry is yes, the controller 60 performs one more of the above measures of increasing the bulk compaction temperature and then moves to the next step of examining whether the bulk compaction temperature threshold has been exceeded. .

The controller 60 checks to see if the bulk compaction temperature threshold has been exceeded. If the result of the inquiry is No, the controller 60 returns to the step of inquiring whether the compressed air-fuel ratio is below the lower threshold. If the result of the inquiry is Yes, the controller 60 implements one of the above-described measures to reduce the bulk compaction temperature. The controller 60 then checks to see if the bulk compaction temperature is above the maximum threshold. If the result of the inquiry is No, the controller 60 returns to the step of inquiring whether the compressed air-fuel ratio is below the lower threshold. If the result of the inquiry is yes, the controller 60 performs one of the above-described measures to lower the bulk compression temperature, and then proceeds to inquire whether the compressed air-fuel ratio is below the lower threshold. .

In some embodiments, controller 60 includes an algorithm or lookup for determining which of the available strategies for increasing or decreasing the compressed air-fuel ratio is the most optimal strategy under current engine operating conditions. Information such as an up table is provided.

FIG. 9 illustrates air/fuel ratio optimization for a single cylinder. A separate air/fuel ratio optimization unit is provided for each of the plurality of cylinders of the engine. In some embodiments, the air/fuel ratio optimization unit and related units are integrated into controller 60 . In some embodiments, these units are part of a controller (not shown) associated with controller 60 .

The output of the air/fuel ratio optimization unit for a single cylinder (i.e., for each cylinder) is fed separately to a cylinder control unit (CCU) that receives signals from the cylinder-specific air/fuel ratio optimization unit. have.

The steady state default mode value for air/fuel ratio is sent to a summing point and the result of that summing point is sent to the cylinder control unit for the associated cylinder. The output from the air/fuel ratio optimization unit (λ optimization unit) is also sent to the summing point.

The air/fuel ratio optimization unit provides at least a signal representing engine load, a signal representing speed error (difference between desired engine speed (RPM) and actual engine speed (RPM)), and FRC (combustion process stabilization). A signal indicating that the high pressure injection pressure of liquid fuel (eg, diesel oil or marine diesel fuel) at or near TDC, used for the injection of liquid fuel, is active.

In addition, a cylinder pressure variation estimation module receives cylinder pressure measurements for the cylinder being estimated. A cylinder pressure variation estimation module determines if an expected combustion event has occurred, such as a misfire event, partial misfire event, or pre-ignition. (Pre-ignition is determined from deviations rather than from variations.) The cylinder pressure variation estimation module combines actual (i.e., measured) changes in cylinder pressure with expected raw events in cylinder pressure. and determining the occurrence of an unexpected combustion event based on the change and/or deviation from the estimate.

The optimization unit integrates from the initial level p1 towards a minimum value p2 (p2 has a value greater than or equal to 0) with small decrementing margin values. (decrease gradually). Once the value p2 is reached, it remains in that state until an event forces a change. The process of integrating the values of the margins from the initial level p1 to the minimum level p2 is relatively slow, typically taking at least several minutes, sometimes 10 to 15 minutes.

Once the controller 60 determines that the engine is operating in steady state conditions, the controller 60 either immediately or with a predetermined delay (i.e. after a predetermined time) begins reducing the margin from the initially set initial value. do.

As previously stated, this margin is a margin from the air/fuel ratio level for actual operating conditions that are known to be in a zone where undesirable combustion events are likely to occur. This margin can be considered a safety margin.

As long as the engine is in steady state operation, and the controller 60 is sure of it, the optimization unit integrates the value of the margin over time in small increments 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, the processor 60 will determine that the engine is no longer operating at steady state. Conclude and cancel the optimization process. The margin value is set to level p1.

Also, if the cylinder pressure estimation unit detects an unwanted combustion event such as a misfire event, partial misfire event, or pre-ignition, the air/fuel ratio optimization process is reversed and the air/fuel ratio optimization unit Integrate the value in small increments towards the initial level p1. This continues until no misfire events are detected or the initial value p1 is reached.

In some embodiments, when a predetermined amount of time has elapsed since the value of the margin was last increased, and when the margin is not equal to the second value, the controller 60, for each combustion chamber individually, actually Resume gradually decreasing the margin by small decrements from the value of . This default length of time is a default duration. This predetermined period may be seconds long and may be minutes long.

In some embodiments, the controller 60 is configured to decrement the margin smaller by advancing the exhaust valve closing timing (preferably in small steps).

In some embodiments, the controller 60 is configured to incrementally increase the margin by retarding the exhaust valve closing timing (preferably in small steps).

In some embodiments, the controller 60 is configured to decrement the margin smaller by closing or tightening the exhaust bypass control valve 49 .

In some embodiments, the controller 60 is configured to incrementally increase the margin by opening or throttling the exhaust bypass control valve 49 .

In some embodiments, the controller 60 is configured to increment the margin by activating or speeding up the exhaust recirculation blower 43 .

In some embodiments, the controller 60 is configured to decrement the margin smaller by deactivating or slowing down the exhaust recirculation blower 43 .

In some embodiments, controller 60 activates liquid fuel injection (FRC) when operating conditions dictate (e.g., to prevent unstable combustion or misfire continuation). Configured. It is also configured to reset the margin to the initial value p1 when the liquid fuel injection is activated. The liquid fuel may be any liquid fuel that has good and reliable combustion characteristics for compression ignition, such as diesel oil and marine diesel fuel.

A number of aspects and implementations have been described with some examples. However, many variations in addition to those described exist in the practice of the claimed invention by one of ordinary skill in the art upon review of the specification, drawings, and claims of this application. understand and be able to implement it. The verbs "comprising," "having," and "including" in the claims do not exclude the presence of elements or steps not recited. The absence of an explicit plural number of an element in a claim does not exclude the presence of a plurality of such elements. The functions of several elements recited in the claims may be performed by a single processor, controller or other unit. The mere fact that certain items are recited in separate dependent claims does not preclude their joint practice, which may be practiced to advantage.

Any reference signs used in the claims shall not be construed as limiting the scope of the invention.

Claims (17)

A large two-stroke uniflow scavenging internal combustion engine operating with gas fuel as the primary fuel in a gas mode of operation, comprising:
a plurality of combustion chambers each defined by a cylinder liner, a reciprocating piston and a cylinder cover;
a scavenging port for introducing scavenging air into the combustion chamber, the scavenging port disposed in the cylinder liner;
an exhaust gas outlet disposed in the cylinder cover and controlled by an exhaust valve;
a variable timing exhaust valve actuation system that allows control of exhaust valve timing for each combustion chamber;
one or more gas inlets disposed in the cylinder liner or the cylinder cover and configured to introduce gaseous fuel during the stroke of the reciprocating piston into the cylinder cover;
at least one controller associated with the engine;
The at least one controller is configured to determine and control opening and closing timing of the exhaust valves for each combustion chamber, and for each combustion chamber, an amount of gaseous fuel introduced into the combustion chamber from the gas inlet. is configured to control the
the at least one controller configured to monitor operating conditions of the engine and determine when the engine is operating at steady state operating conditions;
The combustion chamber has a known undesired combustion state at least at operating conditions deviating from the steady state operating condition, wherein the air-fuel ratio exceeds a known operating condition dependent critical value at the known undesired combustion conditions. and partial misfire events, misfire events, and/or preignition are likely to occur,
the variable timing exhaust valve actuation system permits control of exhaust valve timing independently for each combustion chamber;
The at least one controller is configured to individually determine and control the opening and closing timing of the exhaust valve for each combustion chamber, and for each combustion chamber individually, the gas is introduced into the combustion chamber from the inlet. configured to control the amount of gas fuel,
the at least one controller is configured to operate in a steady state mode when determining that the engine is operating in a steady state operating condition;
The at least one controller in the steady state mode comprises:
Controlling the air/fuel ratio as a function of operating conditions for each combustion chamber individually, with a margin initially set to a first value such that said air/fuel ratio value is below said known operating condition dependent critical value. control and
decrementing the margin from the actual value to a second value less than the first value and greater than zero for each combustion chamber individually;
monitor misfire, partial misfire, and preignition events for each combustion chamber separately;
upon detection of a partial misfire event, misfire event, and/or preignition event, reducing said margin from the actual value to said first value until no misfire event, partial misfire event, or preignition event is detected; Increment towards and increase,
An institution that is configured to 2. The engine of claim 1, wherein the at least one controller is configured to be informed of the undesirable combustion state and the known operating condition dependent critical value. The controller reduces the margin to the actual value, individually for each combustion chamber, when a predetermined amount of time has elapsed since the value was last increased and when the value is not equal to the second value. 2. The engine of claim 1, configured to decrement small from the value and resume the gradual decrease. 2. The engine of claim 1, wherein the controller is configured to decrement the margin smaller by advancing exhaust valve closing timing, preferably in steps. 2. The engine of claim 1, wherein the controller is configured to incrementally increase the margin by retarding exhaust valve closing timing, preferably in stages. an exhaust bypass having an exhaust bypass control valve;
The controller is configured to decrement the margin smaller by closing or tightening the exhaust bypass control valve.
The engine of Claim 1. an exhaust bypass having an exhaust bypass control valve;
wherein the controller is configured to increment and increase the margin by opening or loosening the throttle of the exhaust bypass control valve;
The engine of Claim 1. having an exhaust recirculation pipe with an exhaust recirculation blower therein;
wherein the controller is configured to increment and increase the margin by activating or speeding up the exhaust recirculation blower;
The engine of Claim 1. having an exhaust recirculation pipe with an exhaust recirculation blower therein;
the controller is configured to decrement the margin smaller by deactivating or slowing the exhaust recirculation blower;
The engine of Claim 1. having a cylinder bypass upstream of the main scavenging air cooler,
The controller raises the air-fuel ratio by opening the hot cylinder bypass pipe or throttling the control valve in the hot cylinder bypass pipe, and closing the hot cylinder bypass pipe or throttling the control valve in the hot cylinder bypass pipe. configured to reduce the air-fuel ratio by tightening
The engine of Claim 1. The controller is configured to activate liquid fuel injection when operating conditions require it, and configured to reset the margin to the first value when liquid fuel injection is activated. , an engine according to claim 1. 6. An engine according to claim 4 or 5, wherein said increments are small increments, said decrements are small decrements, and said stepwise steps are small steps. Equipped with a sensor that detects the cylinder pressure individually for each cylinder,
The controller monitors sensed cylinder pressure for each cylinder individually and determines whether a misfire, partial misfire, and/or preignition event has occurred for each cylinder individually. configured to
The engine of Claim 1. The controller determines the deviation of cylinder pressure change from an expected change in cylinder pressure when no misfire event, partial misfire event, and/or pre-ignition event occurs. 14. An engine according to claim 13, adapted to determine fire events and/or pre-ignition events. The controller determines that the engine is operating under steady state conditions when the difference between the desired engine speed and the actual engine speed is below a deviation threshold while the engine load is above the engine load threshold. 2. The engine of claim 1, wherein the engine comprises: 2. An engine according to claim 1, wherein said at least one controller comprises or is connected to compressed air-fuel ratio monitoring means for determining an instantaneous average compressed air-fuel ratio in said combustion chamber. . A method of operating a large two-stroke turbocharged uniflow scavenging internal combustion engine having multiple combustion chambers in a gas mode of operation, comprising:
there is an air-fuel mixture in the combustion chamber with a pre-combustion air-fuel ratio;
The combustion chamber has known undesired combustion conditions at least at operating conditions deviating from steady state operating conditions , and at the known undesired combustion conditions when the air-fuel ratio exceeds a known operating condition dependent critical value. , partial misfire events, misfire events, and/or preignition are likely to occur,
Said method is:
monitoring operating conditions of the engine to determine when the engine is operating at the steady state operating conditions ;
Upon determining that the steady state operating condition is
Controlling the air/fuel ratio as a function of operating conditions for each combustion chamber individually, with a margin initially set to a first value such that said air/fuel ratio value is below said known operating condition dependent critical value. control and
decrementing the margin from the actual value to a second value less than the first value and greater than zero for each combustion chamber individually;
monitor misfire, partial misfire, and preignition events for each combustion chamber separately;
upon detection of a partial misfire event, misfire event, and/or preignition event, reducing said margin from the actual value to said first value until no misfire event, partial misfire event, or preignition event is detected; Increment towards and increase,
Method.

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