WO2018207307A1

WO2018207307A1 - Engine control method and engine system

Info

Publication number: WO2018207307A1
Application number: PCT/JP2017/017853
Authority: WO
Inventors: 隆典黒岩; 橋本　徹; 渡辺　孝一; 和広結城; 美和子山澤
Original assignee: 新潟原動機株式会社
Priority date: 2017-05-11
Filing date: 2017-05-11
Publication date: 2018-11-15
Also published as: CN109247022B; JP6793824B2; JPWO2018207307A1; CN109247022A

Abstract

In the present invention, a mixture of fuel gas and air is combusted in a combustion chamber in a gas mode of a dual-fuel engine (1) for ships. The engine system is provided with: a control unit (22) for advancing the closing timing of an intake valve (8) of the engine and, so as to accord with the changes from said advancement, also advancing the opening timing of a fuel gas supply valve (15), when the output of an output shaft (2) of the engine (1) increases; a variable intake valve timing mechanism (30) for advancing the closing timing of the intake valve (8) in accordance with the intake valve (8) closing timing set by the control unit; and a fuel gas supply valve timing mechanism (45) for advancing the opening timing of the fuel gas supply valve (15) so as to accord with the changes from the advancement of the intake valve (8) set by the control unit. Control is performed such that with the increase in the output of the output shaft of the engine (1), the compression ratio of the gas-air mixture in the engine is further reduced by the variable intake valve timing mechanism (30).

Description

Engine control method and engine system

The present invention relates to an engine control method and an engine system using a gas fuel such as natural gas, and more particularly, to an engine control method and an engine using a gas fuel having a variable intake valve timing (VIVT) mechanism. It is about the system.

Examples of specific configurations of the variable valve timing mechanism are described in Patent Documents 1 to 3 below.
As shown in FIGS. 20 and 21, the drive mechanism of the variable valve timing mechanism 100 described in Patent Document 1 includes a link mechanism 101 and an actuator 102. In the link mechanism 101, the exhaust valve swing arm 103 connected to the push rod of the engine exhaust valve is supported by the link shaft 104, and the intake valve swing arm 105 connected to the push rod of the intake valve is eccentric from the link shaft 104. It is supported by the tappet shaft 106 of the eccentric shaft portion.
The exhaust valve swing arm 103 and the intake valve swing arm 105 can be advanced and retracted by an eccentric cam 108a of the cam shaft 108, respectively. The link shaft 104 is connected to a piston rod 109 provided on the actuator 102.
When the position shown in FIG. 21 is before the pop-out operation of the piston rod 109, all the

swing arms

105 and 103 connected to each other are rotated in one direction by the pop-out operation of the piston rod 109 by the actuator 102. Therefore, the rotation angle of all the

swing arms

105 and 103 can be controlled by the actuator 102 via the link mechanism 101.

As another example, variable valve timing mechanisms described in

Patent Documents

2 and 3 are described in FIGS. The same parts as those of the variable valve timing mechanism 100 shown in FIGS. 20 and 21 will be described using the same reference numerals.
In the variable valve timing mechanism shown in FIG. 22, the rotational range of the link shaft 104 is restricted by the range of the teeth of the sector gear 120 connected to the actuator 102, and an eccentric disk 123 (which is eccentrically fixed to the link shaft 104). Is equivalent to the exhaust valve swing arm 103 and the intake valve swing arm 105.
Therefore, the position where the eccentric cam 108a of the cam shaft 108 contacts and pushes up against the exhaust valve swing arm 103 or the intake valve swing arm 105 changes with respect to the displacement of the rotational angle position of each eccentric disk 123 with respect to the rotational position of the link shaft 104. To do.

In the example shown in FIG. 23, the exhaust valve swing arm 103 and the intake valve swing arm 105 connected to the rocker arm 127 via a push rod 128 are connected to the tappet shaft 106 of the crank-shaped link shaft 104 (the fulcrum position of the swing arm). It is connected. By changing (turning) the phase of the crank-shaped link shaft 104 by the actuator 102, the fulcrum positions of the intake valve swing arm 105 and the exhaust valve swing arm 103 are changed, and as a result, the contact position to the cam shaft 108 is changed. .
As a result, the timing at which the eccentric cam 108a of the cam shaft 108 pushes the exhaust valve swing arm 103 or the intake valve swing arm 105 with the cam shaft 108 to advance and retract is variable.

Patent Document 4 discloses a fuel injection device that is provided in a first intake valve and a second intake valve provided in each combustion chamber, and injects fuel toward each of the first intake valve and the second intake valve provided in an intake port. And a variable valve mechanism capable of changing the valve timing of the first intake valve and the valve timing of the second intake valve to different valve timings are disclosed. . The internal combustion engine controls the variable valve mechanism to set the opening timing of the first intake valve before top dead center, set the opening timing of the second intake valve after top dead center, In the valve timing setting means for setting the closing timing of the second intake valve and the closing timing of the second intake valve after bottom dead center, and the fuel directed toward the first intake valve by the fuel injection device in the valve timing setting state by the valve timing setting means Injection timing setting means for setting the injection start timing to a time delayed from the top dead center as the amount of gas blown back to the intake port side through the first intake valve increases.

Patent Document 4 is for realizing stratified combustion in the internal EGR of a vehicle engine, and the fuel injected toward the first intake valve by the fuel injection device relates to a liquid fuel that can end the injection in an instant. It is.

In Patent Document 5, the negative change rate of the target EGR rate is compared with a predetermined threshold value, and when the negative change rate is equal to or higher than the threshold value, the EGR valve closing operation is performed during the EGR valve opening period. When the EGR gas response delay is detected, and when the EGR gas response delay is detected, the closing timing of the intake valve is corrected so that the actual compression ratio becomes high, and the compression top dead center It is disclosed that the fuel injection timing is corrected so as to approach.

Patent document 5 is for coping with the response delay of EGR gas in external EGR of a diesel engine. The fuel injected into each cylinder via the injector is a liquid fuel that can be filled in the combustion chamber in an instant.

Patent document 6 is a self-ignition combustion type premixed compression ignition engine, which detects a set output of the engine, and sets an equivalence ratio of the premixed gas (set by a fuel supply amount) as the set output increases. The actual compression ratio is decreased while increasing. The engine setting output, in other words, the required engine output, is set manually or by detecting a load applied to the crankshaft of the engine. This “load applied to the crankshaft” is “set output, in other words, required engine output”, so it is set and requested for the engine, not the output of the output shaft that the engine actually outputs. Output.

Patent Document 7 describes that in a sub-chamber gas engine, the output of the gas fuel engine is defined by the rotational speed and torque.

International Publication No. 2015/060117 European Patent Publication No. 2136054 JP-A-62-99606 Japanese Patent No. 5502033 Japanese Patent No. 5338977 JP 2002-21608 JP 2013-185515 A

In the conventional gas fuel engine, even if the supply amount of gas fuel is increased at a fast rate in order to rapidly increase the output, the supercharger cannot follow and cannot supply the necessary air amount. This is because the supercharger is driven by exhaust gas, so that it does not work effectively unless the output of the gas fuel engine is increased and exhaust gas is sufficiently supplied to the supercharger. If the amount of air is insufficient, the air-fuel ratio becomes gas rich and knocking occurs, leading to engine failure. In order to suppress the knocking, if the output is increased at a speed that can be followed by the supercharger, it takes about 10 minutes to increase the output (load increase). In this specification, as a general rule, the term “output” is used for the power that the engine actually outputs, and the term “load” is used for the power that is set and required for the engine. In many cases, “output” = “load”, and the term “load” is used instead of “output” (or vice versa).

On the other hand, emission regulations for harmful exhaust gases are becoming stricter every year in marine engines, and there is a demand for the introduction of a dual fuel engine that can meet the emission regulations with little emission of harmful exhaust gases derived from fuel. ing. However, in order to introduce such a dual fuel engine, it has been necessary to reduce the load increase time to about 20 seconds in order to satisfy the operation mode of the marine engine.

The present invention has been made in view of the above-described problems, and can suppress knocking that occurs when the output of a gas fuel engine is increased to shorten the load increase time, and also provides unburned gas fuel during valve overlap. An object of the present invention is to provide an engine control method and an engine system in which fuel consumption does not deteriorate due to air blow-through.

That is, the present inventors change the valve opening (supply start) timing of the fuel gas supply valve at each VIVT angle, determine the optimum value from the THC concentration and the combustion state, and the fuel gas supply valve according to the VIVT angle. By setting the valve opening timing, it is possible to suppress the knocking that occurs when the output of the gas fuel engine is increased, and to shorten the load increase time. Further, the fuel gas supply valve in the torque rich region and the torque poor region The present inventors have found that combustion fluctuations and rotational speed hunting caused by the valve opening timing can be improved.

The engine control method according to the present invention is an engine control method using gas as fuel, and combustion is performed by adjusting the advance angle from the suction bottom dead center at the timing when the intake valve closes as the engine output increases. Control is performed to lower the compression ratio of the air-fuel mixture in the room, and the opening timing of the fuel gas supply valve is advanced in response to the change in the advance angle.

As an engine knocking suppression technique, an effective compression ratio can be lowered using a variable intake valve timing (VIVT) mechanism. A knocking suppression technique in this regard will be described with reference to FIGS. 19A and 19B. FIG. 19A shows a normal 4-stroke cycle process, and FIG. 19B shows a mirror cycle process.
For example, in a gas fuel engine, the intake valve is normally closed at the bottom dead center of the piston (see FIG. 19A). On the other hand, if the closing timing is made earlier than the bottom dead center as shown in FIG. 19B, the in-cylinder temperature Ts decreases from the case of FIG. 19A because the air-fuel mixture continues to expand after the intake valve closes (Ts ^* <Ts). . Accordingly, since the maximum compression temperature at the top dead center is also lowered (Tc ^* <Tc), self-ignition can be prevented and knocking is suppressed.
As a disadvantage of the Miller cycle, since the compression temperature is lowered and the ignitability in the low load region is deteriorated, the normal intake valve opening timing shown in FIG. The valve opening timing needs to be advanced.

In the engine control method according to the present invention, in the gas fuel engine in which the supply amount of the gas fuel is increased to increase the output, the closing timing of the intake valve is advanced (advanced) from the suction bottom dead center. Control to lower the compression ratio of the air-fuel mixture. By reducing the compression ratio, the temperature in the combustion chamber at the time of compression is lowered, so that knocking can be suppressed.
If the compression ratio of the air-fuel mixture is lowered in the combustion chamber, the ignitability deteriorates at the time of start-up and at a low output, and it departs from an advantageous condition in terms of combustion efficiency, resulting in a demerit that fuel consumption deteriorates. Therefore, in the present invention, control is performed to lower the compression ratio at a larger rate by changing the timing at which the intake valve closes more greatly in an operation region where the output that is more likely to cause knocking is higher. Thereby, knocking can be suppressed according to the change in output, and the load increase time can be shortened while preventing deterioration of fuel consumption. In addition, by advancing the opening timing of the fuel gas supply valve in response to a change in the advance angle of the intake valve, it is possible to reduce blowout of unburned fuel gas when the intake valve and the exhaust valve overlap each other. .

Further, it is preferable that the degree of advancement of the opening timing of the fuel gas supply valve becomes larger as the advancement of the closing timing of the intake valve proceeds.
Thereby, it is possible to further reduce blowout of unburned fuel gas in the combustion chamber when the intake valve and the exhaust valve overlap each other.

The closing timing of the fuel gas supply valve is calculated based on the deviation between the target rotational speed and the actual rotational speed, and the opening period is set based on the opening timing of the fuel gas supply valve. Is preferably determined by:
Based on the deviation between the target engine speed and the actual engine speed, the valve opening period of the fuel gas supply valve is calculated by feedback control such as PID control. By setting this valve opening period starting from the opening timing of the fuel gas supply valve, the closing timing of the fuel gas supply valve can be determined.

Further, it is preferable to increase the supply pressure of the fuel gas as the intake valve closes at the advanced timing.
By increasing the supply pressure of the fuel gas in accordance with the advance timing of the intake valve closing timing, it is possible to supply an appropriate amount of gas fuel for maintaining the output within the valve opening period.

Further, it is preferable to advance the ignition timing of the engine in accordance with the advance timing of the closing timing of the intake valve.
The thermal efficiency can be increased by advancing the ignition timing in a range where NOx falls within a predetermined value in accordance with the advance angle of the intake valve closing timing.

The ignition timing advance angle peaks at the ship's cube characteristic line that is the boundary between the torque rich region and the torque poor region, which are set using the engine output and the rotational speed measured in advance as parameters. The advance angle may be reduced from the cube characteristic line.
Even in this case, since the ignition timing is advanced in the front region of the torque rich region and the torque poor region with the ship's cube characteristic line at the peak rather than before the timing at which the intake valve is closed, NOx is generally within the regulation range. There is an advantage that the thermal efficiency is increased while suppressing the temperature. The “marine cube characteristic” is a characteristic of a marine main engine whose output is proportional to the cube of the rotational speed. However, an actual ship is not necessarily proportional to the cube, and some deviation occurs.

The output of the engine is the output shaft obtained from the torque measurement value obtained by measuring the torque of the engine output shaft with a torque sensor and the rotation speed measurement value obtained by measuring the rotation speed of the engine output shaft with a rotation speed sensor. An output value is preferred.
In the engine according to the present invention, since the fuel gas is an elastic body, it is relatively difficult to obtain an accurate fuel supply amount compared to the liquid fuel. Therefore, it is preferable to calculate the output in relation to the rotational speed by actually measuring the torque with a torque sensor. In addition, the output (load) of the output shaft of the engine can be obtained in real time by taking the product of the measured value of the rotational speed of the output shaft obtained by providing the rotational speed sensor and the measured torque value of the torque sensor. Therefore, since the advance angle of the closing timing of the intake valve of the engine and the advance angle of the opening timing of the fuel gas supply valve can be set with high accuracy, even if the output increases, the combustion efficiency is improved and the gas fuel engine is appropriately You can drive.

Further, it is preferable that the advance angle at the timing when the intake valve is closed is set based on advance values set as parameters using the output values and rotation speed data of a plurality of output shafts measured in advance.
A suitable value for the advance angle of the closing timing of the intake valve becomes larger when the output is large, but in addition to this, it also depends on the rotational speed. Therefore, knocking can be further suppressed by creating a map including at least these two parameters in advance and controlling the advance timing of the intake valve closing timing in accordance with changes in engine output and rotational speed.

The engine system according to the present invention is an engine system having a four-stroke engine using gas as fuel, and advances the timing at which the engine intake valve closes when the output of the engine output shaft increases. A control unit that advances the opening timing of the fuel gas supply valve in response to a change in the advance angle, and variable intake that changes the closing timing of the intake valve according to the closing timing of the intake valve set by the control unit A valve timing mechanism, and a fuel gas supply valve timing mechanism for advancing the valve opening timing of the fuel gas supply valve in response to a change in the advance angle of the intake valve set by the control unit. With this increase, the variable intake valve timing mechanism controls to lower the compression ratio of the gas / air mixture in the engine.
In the engine system control method according to the present invention, when the supply amount of gas fuel is increased and the output of the output shaft of the engine is increased, the timing at which the intake valve closes is advanced from the suction bottom dead center. Control to lower the compression ratio of the air-fuel mixture in the room. By reducing the compression ratio, the temperature in the combustion chamber at the time of compression is lowered, so that knocking can be suppressed.
In addition, in the present invention, control is performed to lower the compression ratio at a larger rate by changing the timing at which the intake valve closes more greatly in an operation region where the output that is more likely to cause knocking is higher. Thereby, knocking can be suppressed according to the change in output, and the load increase time can be shortened while preventing deterioration of fuel consumption. In addition, the fuel gas supply valve opening timing is advanced in response to the change in the advance angle of the intake valve, so that the unburned fuel is generated when the intake valve and the exhaust valve overlap with respect to the crank angle of the intake valve closing. Gas blow-by can be reduced.

In addition, a torque sensor for measuring the torque of the output shaft of the engine and a rotation speed sensor for measuring the rotation speed of the output shaft of the engine are provided, and the output shaft is determined from the torque measurement value by the torque sensor and the rotation speed measurement value by the rotation speed sensor. It is preferable to set a change in the closing timing of the intake valve in the control unit.
In the gas fuel engine according to the present invention, since the gas serving as the fuel is an elastic body, it is relatively difficult to obtain an accurate fuel supply amount as compared with the liquid fuel. Therefore, it is preferable to calculate the output in relation to the rotational speed by actually measuring the torque with a torque sensor. Moreover, the output of the output shaft of the engine can be obtained in real time by taking the product of the rotational speed measurement value of the output shaft obtained by providing the rotational speed sensor and the torque measurement value of the torque sensor.

According to the engine control method and the engine system of the present invention, when the output of the output shaft of the engine increases, the compression ratio of the air-fuel mixture in the engine can be lowered. The raising time can be shortened.
Moreover, in order to advance the opening timing of the fuel gas supply valve in response to the adjustment of the advance angle of the intake valve, the combustion fluctuation and the rotational speed hunting caused by the opening timing of the fuel gas supply valve are improved. The gas fuel engine can be appropriately operated.

It is a block diagram which shows the principal part structure of the marine dual fuel engine by embodiment of this invention. It is a figure which shows the diesel mode and gas mode in the dual fuel engine shown in FIG. It is a three-dimensional map which shows the relationship between an output, a rotational speed, and a VIVT command value. It is a three-dimensional map which shows the relationship between an output, a rotational speed, and the valve closing timing of an intake valve. It is a graph which shows the relationship between an output and the optimal VIVT command value in the case where the rotational speed of an output shaft changes with the case where it is constant. It is a graph which shows the relationship between the valve opening timing of a fuel gas supply valve, and THC density | concentration when a VIVT command value differs variously. It is a graph which shows the relationship between an output and the valve-opening timing of a fuel gas supply valve in the case where the rotational speed of an output shaft changes with a case. It is a graph which shows the fuel gas supply start time corresponding to a VIVT command value. It is a figure which shows the structure of the control apparatus which performs PID control of the marine dual fuel engine. It is a figure which shows the process of carrying out PID control so that an actual rotational speed may become target rotational speed. 6 is a timing chart of opening and closing operations of an intake valve and a fuel gas supply valve at normal time and advance angle. It is a figure which shows the relationship between the output, the valve opening period of a fuel gas supply valve, and the pressure (DELTA) P of fuel gas. It is a three-dimensional map which shows the relationship between an output, rotation speed, and pressure (DELTA) P of fuel gas. It is a graph which shows the change of NOx corresponding to an air fuel ratio, and the usable range. It is a three-dimensional map which shows the relationship between a rotational speed, an output, and an air supply pressure. It is a graph which shows the relationship between the output in the case where intake valve closing timing is constant, and the case of advance, and supply air pressure. It is a three-dimensional map which shows the relationship between a rotational speed, an output, and ignition timing. It is a graph which shows the relationship between NOx at the time of ignition timing change, and thermal efficiency. It is process drawing of the normal cycle of an engine combustion cycle. It is process drawing of the mirror cycle of a combustion cycle of an engine. It is a perspective view which shows an example of the conventional variable intake valve timing mechanism. It is a front view which shows an example of the conventional variable intake valve timing mechanism. It is a perspective view which shows the other example of the conventional variable intake valve timing mechanism. It is a figure which shows the further another example of the conventional variable intake valve timing mechanism. It is a figure which shows the relationship between the actuator and link shaft in the conventional variable intake valve timing mechanism shown in FIG.

Hereinafter, for example, a four-stroke dual fuel engine 1 used as a marine engine as an engine according to an embodiment of the present invention will be described with reference to the accompanying drawings.
A marine dual fuel engine 1 (hereinafter, simply referred to as an engine 1) shown in FIGS. 1 and 2 includes diesel mode D and gas mode G engines. During operation, the diesel mode D and gas mode G are provided. It is an engine that can be switched between. A dual fuel engine 1 shown in FIG. 1 has a mechanism of a crankshaft 2 as an output shaft connected to a propeller or the like, and the crankshaft 2 is connected to a piston 4 installed in a cylinder block 3. A combustion chamber 6 is formed by a piston 4 and an engine head 5 provided in the cylinder block 3.

The combustion chamber 6 is sealed by an intake valve 8 and an exhaust valve 9 attached to the engine head 5 and a fuel injection valve 10 used in the diesel mode D. The engine head 5 is provided with a micro pilot oil injection valve 11 used in the gas mode. An intake pipe 13 is connected to the intake port where the intake valve 8 of the engine head 5 is installed, and an exhaust pipe 14 is installed to the exhaust port where the exhaust valve 9 is installed. A fuel gas supply valve 15 that is an electromagnetic valve that controls gas injection is installed in the intake pipe 13, and an air cooler 16 and a supercharger 17 that communicates with the exhaust pipe 14 are installed upstream of the intake pipe 13.

Here, the dual fuel engine 1 according to the present embodiment can be operated by switching between the diesel mode D and the gas mode G as shown in FIG. In the diesel mode D, for example, heavy fuel oil A or the like can be mechanically injected from the fuel injection valve 10 into the compressed air in the combustion chamber 6 as fuel oil, ignited and burned. In the gas mode G, a fuel gas such as natural gas is supplied to the intake pipe 13 by the fuel gas supply valve 15 and premixed with the air flow to supply the air-fuel mixture into the combustion chamber 6. Pilot fuel is injected from the pilot oil injection valve 11 to ignite and burn. The micro pilot oil injection valve 11 is electronically controlled, for example, and injects a small amount of pilot fuel as a powerful ignition source. The fuel gas supply valve 15 is an electromagnetic valve that forms a large opening with a small stroke and allows a large amount of gas to flow in a short time.

The engine 1 is started in a diesel mode D in which liquid fuel is injected into the combustion chamber 6 from the fuel injection valve 10. After confirming that the gas pressure exceeding the reference value is supplied to the engine 1, the fuel gas supply valve 15 supplies gas fuel to the intake pipe 13 and mixes it with air, and then flows into the combustion chamber 6. The operation is performed in the gas mode G in which the gas fuel is burned.
When stopping, change to diesel mode D again and stop. The diesel mode D and the gas mode G can be changed except when starting and stopping.

The dual fuel engine 1 according to the present embodiment includes a gas engine system that performs output control when the load increases in the gas mode G. The structure of this gas engine system will be further described.
In FIG. 1, a rotational speed sensor 20 and a torque sensor 21 are attached to the crankshaft 2. The rotational speed sensor 20 measures the rotational speed (number of rotations) of the crankshaft 2, and the torque sensor 21 measures engine torque. measure. As the torque sensor 21, for example, a sensor that detects the torque applied to the shaft by strain can be used. Measurement data measured by the rotational speed sensor 20 and the torque sensor 21 are output as signals to the control unit 22 that controls the engine 1.
The control unit 22 detects the operating state of the engine 1 based on signals from the rotational speed sensor 20 and the torque sensor 21. That is, assuming that the rotational speed (number of revolutions) of the crankshaft 2 measured by the rotational speed sensor 20 is n and the torque measured by the torque sensor 21 is T, the output of the engine 1 is expressed by the following formulas (1) and (2). (Load) A is calculated. However, Lt is the rated output of the engine 1.
Output Lo = 2πTn / 60 (1)
Output (load) A = Lo / Lt × 100 (2)

In addition, as a method for obtaining the output (load) of the engine 1, a method of estimating from the fuel supply amount and other information related to the operating state of the engine 1, and a power transmission system of the output shaft of the engine 1 includes a torque sensor 21. There is a method of actually measuring torque and obtaining an output. In a gas fuel engine, since the gas used as the fuel is an elastic body, it is relatively difficult to obtain an accurate fuel supply amount compared to the liquid fuel. Therefore, it is preferable to calculate the output by actually measuring the torque with the torque sensor 21.
When the rotational speed n is constant, the output A and the torque measurement value T are in a directly proportional relationship. Under a condition where the rotational speed n is constant, it is desirable to set the advance timing of the closing timing of the intake valve 8 at a larger rate as the output A is larger, that is, as the torque data T is larger.

The control unit 22 stores a first map 24 for determining a first electric signal of intake valve opening / closing timing prepared in advance and a second map 25 for determining opening / closing timing corresponding to the first electric signal. In the control unit 22, based on the rotational speed data n and the torque data T corresponding to the output A of the engine 1 measured by the rotational speed sensor 20 and the torque sensor 21, The output A is calculated. Then, the first electric signal corresponding to the opening / closing timing of the intake valve 8 is selected on the first map 24 based on the rotational speed n and the output A. Based on this first electric signal, the opening / closing timing of the intake valve 8 corresponding to the first electric signal is determined in the second map 25. A method for creating the first map 24 and the second map 25 will be described later.
The second electrical signal at the opening / closing timing set by the control unit 22 is transmitted to the electropneumatic converter 27, and the electropneumatic converter 27 converts the opening / closing timing signal into air pressure. This air pressure is sent to the actuator 28 to control the drive of the variable intake valve timing mechanism 30. The actuator 28 is supplied with air pressures P1 and P2 for driving and control from the first pressure reducing regulator 34 and the electropneumatic converter 27.

The air pressure supplied to the actuator 28 is compressed by the air compressor 32 and stored in the air tank 33. The air pressure in the air tank 33 is reduced to a required pressure by the first pressure reducing regulator 34. The pressure at this time is adjusted by changing the valve opening degree of the first pressure-reducing regulator 34, and is supplied to the actuator 28 as the driving air pressure P1. When the pressure P1 measured by the pressure gauge 36 is equal to or less than a specified value, the engine 1 cannot be started.
The air pressure for driving the electropneumatic converter 27 is supplied after the pressure is further reduced from the first pressure reduction regulator 34 by the second pressure reduction regulator 37. The electropneumatic converter 27 supplies the air pressure corresponding to the input second electrical signal at the opening / closing timing to the actuator 28 as the air pressure P2 for adjusting the operation of the actuator 28. Based on these air pressures P1 and P2, the rod 28a of the actuator 28 is operated to operate the variable intake valve timing mechanism 30.

The actuator 28 is, for example, a known P-cylinder (cylinder with a positioner), and controls the advance / retreat of the rod 28 a based on the pressures P 1 and P 2 input from the first pressure-reducing regulator 34 and the electropneumatic converter 27. By changing the moving length of the rod 28a of the actuator 28, the drive of the variable intake valve timing mechanism 30 is controlled, and the closing timing of the intake valve 8 is advanced (advance) or delayed (delay). By controlling the angle, the compression ratio is lowered and control is performed.
Since the time between the opening timing and the closing timing of the intake valve 8 does not change, when the opening timing advances from the suction bottom dead center, the closing timing also advances from the suction top dead center by the same time. Moreover, in the present invention, the timing for opening and closing the valve is changed in accordance with the output of the engine 1 to suppress knocking and shorten the load increase time. The opening / closing timing of the intake valve 8 is set by the first map 24 and the second map 25 in the control unit 22 based on the output A and the rotational speed n of the engine 1, and the intake valve 8 is controlled by the actuator 28 and the variable intake valve timing mechanism 30. The valve opening and closing timing is adjusted so that knocking can be suppressed.

The configuration of the variable intake valve timing mechanism 30 is conventionally known, and has the same structure as that shown in FIGS. That is, in the variable intake valve timing mechanism 30, for example, a link shaft whose rotation angle range is set via a sector gear according to the moving length of the rod 28a of the actuator 28 and a cam shaft having an eccentric cam are arranged in parallel. Yes. An exhaust swing arm is connected to the link shaft, and an intake swing arm is connected to a tappet shaft provided at an eccentric position of the link shaft. An intake valve 8 is connected to the intake swing arm via a push rod and a rocker arm, and an exhaust valve 9 is connected to the exhaust swing arm via a push rod and a rocker arm.

The distance between the cam shaft and the intake swing arm changes depending on the rotation angle of the tappet shaft according to the rotation of the link shaft, and the timing at which the eccentric cam of the cam shaft starts to change changes. As a result, the valve closing timing can be changed to an advance angle (or a delay angle). As the distance from the tappet shaft to the cam shaft center increases, the closing timing of the intake valve 8 becomes earlier. The rotation angle of the tappet shaft is changed by the moving length of the rod 28a of the actuator 28. The moving length of the rod 28 a is arbitrarily changed by the control air pressures P 1 and P 2 supplied to the actuator 28.
The magnitude of the advance angle, which is the closing timing of the intake valve 8, is determined by the timing at which the eccentric cam of the cam shaft starts to hit the intake swing arm connected to the tappet shaft of the link shaft.

The tappet shaft rotating device in the variable intake valve timing mechanism 30 may use a servo motor (not shown) instead of the actuator 28. In this case, a signal of the opening / closing timing transmitted from the second map 25 of the control unit 22 is input to the servo motor. The servo motor can change the opening / closing timing of the intake valve 8 by rotating the link shaft by an amount corresponding to the received signal and rotating the tappet shaft so as to approach and separate from the cam shaft. When a servo motor is used, the configuration from the actuator 28 and the air compressor 32 to the pressure gauge 38 is not necessary. Further, the servo motor is driven by a controller instead of the electropneumatic converter 27.

A mechanism for supplying gas fuel to the fuel gas supply valve 15 that controls gas injection into the intake pipe 13 will be described. In FIG. 1, gas fuel is supplied to a gas vaporizer 41 from an LNG gas tank 40 in which gas fuel such as natural gas is stored, and the gas pressure is reduced to a necessary gas pressure by a gas regulator 42.
The gas pressure at this time is displayed on the fuel gas pressure gauge 43, adjusted by changing the valve opening of the gas regulator 42, and supplied as gas fuel for combustion from the fuel gas supply valve 15 into the intake pipe 13. . In the intake pipe 13, the gas fuel and supercharged air cooled by the air cooler 16 are mixed and supplied to the combustion chamber 6. When increasing the load, the amount of gas fuel supplied is increased by the operation of the fuel gas supply valve 15.

In addition, the second electrical signal of the opening / closing timing set by the control unit 22 is transmitted to the fuel gas supply valve 15 via the fuel gas supply timing means 44 separately from the electropneumatic converter 27. The fuel gas supply timing means 44 controls to open the fuel gas supply valve 15 and advance the valve opening timing for supplying gaseous fuel into the intake pipe 13 in accordance with the advance timing of the closing timing of the intake valve 8. To do. The gas regulator 42 for adjusting the gas pressure and the fuel gas supply timing means 44 for advancing the opening timing of the fuel gas supply valve 15 are included in the fuel gas supply valve timing mechanism 45.
The fuel gas supply timing means 44 may be installed outside the control unit 22. If the fuel gas supply valve timing mechanism 45 can receive the second electric signal from the second map 25 and advance the opening timing of the fuel gas supply valve 15 according to the advance timing of the closing timing of the intake valve 8. Good.

Next, a method for creating the first map 24 and the second map 25 stored in the control unit 22 will be described. FIG. 3 shows the crank angle when the intake valve 8 is closed, which is a VIVT command value (Intake Valve Closed timing, IVC), based on the rotational speed of the crankshaft 2 and the output (load factor) of the engine 1. It is a three-dimensional map which shows the detail of the 1st map 24 to determine.
In FIG. 3, a region B that is regularly (practically) operated is indicated by a broken line. On the other hand, the change (advance angle) of the VIVT command value with respect to the change in output when the rotational speed performed by power generation is constant is indicated by an arrow line C, and the rotational speed and output (load factor) performed for marine use are A change (advance angle) in the VIVT command value when changing simultaneously is indicated by an arrow line D. The arrow line D indicates the marine cube characteristic. The marine cube characteristic indicates a typical characteristic of a marine main engine whose output is proportional to the cube of the rotational speed, and is a characteristic curve of the rotational speed and output determined by the rated rotational speed of the engine and the rated output. In the normal operation region B, the region where the output (load factor) is higher than the marine cube characteristic line D indicates the torque rich region, and the region where the output (load factor) is low indicates the torque poor region.

The first map 24 was created based on the steps of the following experimental procedures (1) to (18).
In the experiment, a dual fuel engine 1 of the same model that was actually used was used.
(1) The engine 1 is started, the rotational speed (the number of revolutions) n is set to 400 min ⁻¹ , the output (load) A is set to 10%, and the closing timing of the intake valve 8 is set to 545 degrees (the slowest closing timing in terms of structure). Set.
(2) Then, abnormal combustion called knocking that occurs when the engine 1 is driven and the exhaust temperature at that time are measured.
Knocking is detected by a knock sensor (not shown) attached to each engine head 5. When knocking occurs, the normal combustion waveform is superimposed on the high-frequency pressure fluctuation.

The exhaust temperature at the time of knocking measurement is measured by a temperature sensor attached to the exhaust pipe 14.
(3) After the exhaust temperature measurement at the time of the knocking measurement is completed, the closing timing of the intake valve 8 is decreased by 5 degrees, and the measurement of (2) is performed again. Measurement is performed by changing the valve closing timing to 500 deg (the earliest valve closing timing in terms of structure).
(4) When the measurement of the above (3) is completed, the output A is increased stepwise until it becomes 110% by 10%, and the measurements of (2) and (3) are repeated again.

(5) According to the measurements (1) to (4) above, when the knocking strength is less than the reference value and the exhaust temperature is 500 ° C. or less, knocking is suppressed and the engine 1 can be operated safely. Judge.
(6) From the measurement result of (5) above, in the three-dimensional graph of FIG. 4 where the X axis is output A, the Y axis is the rotation speed n, and the Z axis is the opening / closing timing, ● (black circle), x is plotted at unsafe measurement points. Thereby, the knocking suppression range in the relationship among the output A, the rotation speed n, and the valve closing timing can be selected.
(7) a measuring step of the above (1) to (6), the rotational speed n rises to 900 min ^-1 by 100 min ^-1, measured safely driving can range for each rotation speed n.

(8) And FIG. 4 is a graph showing the measurement result of (7) above with three axes of the rotational speed n, the output A, and the valve closing timing. In FIG. 4, a range surrounded by a straight line is a range in which knocking is suppressed and the engine 1 can be operated safely.
(9) Next, nitrogen oxides (hereinafter referred to as NOx) are within a three-dimensional region surrounded by a straight line shown in FIG. 4 that can be operated safely, as measured by the experiments (1) to (8). Further experiments are conducted to find a setting that is below the reference value and has the highest thermal efficiency.
The engine speed n is set to 400 min ⁻¹ , the output A is set to 10%, and the valve closing timing of the intake valve 8 is set to 545 deg.

(10) Next, NOx and thermal efficiency are measured. NOx is measured by an exhaust gas analyzer attached to the exhaust pipe 14. The thermal efficiency is calculated by the following equation (3) based on the fuel flow rate L measured from the fuel flow meter attached to the fuel pipe and the output A calculated from the measurement result of the torque sensor 21.
Thermal efficiency η = 360Lo / H / L (3)
However, H: Lower heating value of fuel gas (J / Nm ³ )
Lo: Current output L: Fuel flow rate

(11) After the measurement of (10) is completed, the closing timing of the intake valve 8 is decreased by 5 degrees and the measurement of (10) is performed again. The valve closing timing is changed to 505 deg and measured (see FIG. 9).
(12) When the measurements of (10) and (11) are completed, the output is increased stepwise by 10% to 110%, and the measurements of (10) and (11) are repeated again. The valve closing timing is changed within a range where the operation can be safely performed as shown in FIG.

(13) the measurement of the (9) to (12), the rotational speed n rises stepwise to 900 min ^-1 by 100 min ^-1, to determine a good measurement point most performance for each rotational speed.
(14) Then, the closing timing of the intake valve 8 at which NOx is equal to or less than the predetermined value and has the highest thermal efficiency is set for each rotational speed n and output A. As a result, a draft of the first map shown in FIG. 3 is created.

(15) Further, knocking is detected by increasing the rotational speed n and the output A with an arbitrary load increasing pattern. The load increase pattern is a change state per time of the output A (load factor) and the rotation speed n, and changes depending on the propeller specifications (shape, rotation speed) of the marine propulsion device.
(16) The valve closing timing of the measurement point where the knocking strength detected in (15) is equal to or greater than the reference value is decreased by 3 degrees.
(17) Next, the steps (15) and (16) are repeated until the knocking strength is equal to or less than the reference value, and the valve closing timing at which knocking is suppressed is determined. If the valve closing timing is decreased, the thermal efficiency deteriorates. The setting of the valve closing timing at which NOx and knocking strength are equal to or less than the reference values and the highest thermal efficiency is obtained is set as the setting value of the rotational speed n and output A.
(18) The valve closing timing at which knocking was suppressed from the above (17) was measured at each rotational speed n and output A, and the final first map 24 shown in FIG.

In FIG. 3, the VIVT command value corresponding to the rotation speed and the output is shown in a three-dimensional plane graph, and the upper side in the figure is the direction in which the valve closing timing is advanced. A region indicated by a broken line on the three-dimensional plane is a practical operation region used in the actual operation of the ship propulsion device, and an example of a good load increase pattern is indicated by a ship cube characteristic line D. In increasing the load in a practical operating range, control is performed to increase the advance angle of the valve closing timing as the engine output increases.
In an example of a good load increasing pattern indicated by the marine cube characteristic line D, the advance angle is minimized at the lower right position in the figure where the rotation speed and output are small, and the advance angle increases as the rotation speed and output increase. To do. The ratio of increasing the advance angle is not constant, but the advance angle is increased as the output increases as a whole. Since the output (load factor) is obtained by the product of the torque and the rotational speed, it can be expressed that the advance angle increases as the torque of the output shaft increases.

Next, the second map 25 was created by the following experiment.
When the variable intake valve timing mechanism 30 is rotationally controlled by the actuator 28, the second map 25 is created by the following procedure.
(1) The valve closing timing is changed by the actuator 28, and the pressure when changing to each valve closing timing is measured.
(2) Based on the specifications of the electropneumatic converter 27, the second electrical signal necessary for supplying the pressure (1) is investigated.
(3) From the results of (1) and (2) above, the first electric signal selected on the first map 24 on the horizontal axis and the second map 25 indicating the valve closing timing (second electric signal) on the vertical axis. create.

The above description is for the case where the actuator 28 is used. When the variable intake valve timing mechanism 30 is controlled to rotate by a servo motor instead of the actuator 28, the following description is given.
(1) The valve closing timing is changed based on the servo motor, and the second electric signal when changing to each valve closing timing is measured.
(2) Based on the result of (1) above, a second map 25 is created that indicates the first electric signal on the horizontal axis and the valve closing timing (second electric signal) on the vertical axis.
The second map 25 is a map representing the relationship between the valve closing timing (second electric signal) and the first electric signal.

In the three-dimensional map shown in FIG. 3, the optimum VIVT command value differs depending on the output between the power generation characteristic line indicated by the solid line C and the marine cube characteristic line D. That is, as shown as an example in FIG. 5, even when the output is the same, the intake valve closing crank angle of the optimum VIVT command value is different when the rotational speed is different.
In the present embodiment, in response to changes in the VIVT command value, that is, for various intake valve closing crank angles, to the exhaust pipe 14 of unburned fuel gas due to valve overlap between the intake valve 8 and the exhaust valve 9. The valve opening timing of the fuel gas supply valve 15 for supplying the fuel gas to the intake pipe 13 is set so that the blow-through of the fuel is reduced. For this purpose, first, a VIVT command value corresponding to the rotational speed and output is set. Strictly speaking, it is preferable that the air-fuel ratio and the ignition timing are set to optimum values with reference to thermal efficiency and NOx, but here, these conditions are not set because the engine 1 can be stably operated. .

As an example of setting the opening timing of the fuel gas supply valve 15, the opening timing of the fuel gas supply valve 15 by the fuel gas supply timing means 44 under the engine operating condition in accordance with the optimum VIVT command value in the ship cube characteristic line D How to decide is described below.
First, in each operating condition where the engine output (load factor) is 25%, 50%, 75%, and 100%, the fuel gas is supplied from the fuel gas supply valve 15 with the opening timing of the intake valve 8 as a guideline. Since the fuel gas is supplied into the intake pipe 13, the fuel gas does not reach the intake valve 8 instantaneously. Therefore, the crank angle position of the valve opening timing of the fuel gas supply valve 15 in consideration of the distance from the fuel gas supply valve 15 to the intake valve 8 is assumed. Then, the crank angle position at the opening timing of the fuel gas supply valve 15 is changed to 5 deg increments before and after that, and the total hydrocarbon concentration (unburned gas) in the exhaust gas at the gas turbine outlet of the supercharger 17 at that time ( (THC concentration) is measured. Measurement of THC concentration is repeatedly performed under each operating condition. The THC concentration is preferably measured by a flame ionization method (JIS B 7956).

The opening timing of the fuel gas supply valve 15 is changed according to each condition, and each VIVT command value (intake valve closing crank angle) is set to 40%, 65%, 85%, 100%, for example. FIG. 6 shows the relationship between the fuel gas valve opening timing at the command value and the measured THC concentration.
As shown in FIG. 6, the opening timing of the fuel gas supply valve 15 is based on the crank angle at which the unburned fuel gas hardly blows through and the THC concentration becomes the lowest when the valve overlaps. On the other hand, if the valve opening timing of the fuel gas supply valve 15 changes suddenly due to the output change, it leads to the above-described combustion fluctuation and rotation speed fluctuation. For this reason, ± 5 deg. From the selected standard so that the amount of change in the valve opening timing of the fuel gas supply valve 15 according to the output becomes as small as possible. C. A crank angle that is the optimum fuel gas opening timing of the fuel gas supply valve 15 is selected within the range of A, and the crank angle corresponding to the optimum opening timing of the fuel gas supply valve 15 is determined under each condition.

The relationship between the crank angle at the optimum valve opening timing of the fuel gas supply valve 15 and the output (load factor) in the optimum VIVT command value of the marine cube characteristic line D determined as described above is shown in FIG. It is shown by the line of “change”.
Similarly, the relationship between the crank angle of the optimum valve opening timing of the fuel gas supply valve 15 and the output (load factor) at the optimum VIVT command value at the output with constant rotation speed performed on the power generation characteristic line C in FIG. Is indicated by a broken line of “constant rotational speed” in FIG.
As shown in FIG. 7, the optimum opening timing of the fuel gas supply valve 15 is different depending on the condition that the rotational speed changes and the constant rotational speed even if the output is the same.

However, the optimal fuel gas supply valve 15 at the optimum opening timing of the fuel gas supply valve 15 at the optimal VIV command value of the marine cube characteristic line D and the optimal fuel gas at the optimal VIVT command value of the power generation characteristic line C (constant rotation speed). As shown in FIG. 8, the crank angle at the valve opening timing of the supply valve 15 exhibits one consistent diagram characteristic when the VIVT command value is arranged on the horizontal axis instead of the output. That is, it can be seen that the valve opening timing of the optimum fuel gas supply valve 15 does not depend on the output, but depends on the VIVT designated value (the intake valve closing crank angle).

As is apparent from FIG. 8, when the closing timing of the intake valve 8 is changed by the variable intake valve timing mechanism 30, the advance of the supply start timing of the fuel gas supply valve 15 increases as the advance timing of the closing timing of the intake valve 8 advances. The degree of corners becomes larger.
Therefore, the fuel gas supply timing means 44 sets the crank angle of the optimal valve opening timing of the fuel gas supply valve 15 determined under each condition on the basis of the VIVT command value, so that the fuel gas supply valve 15 based on the VIVT command value is set. The valve opening timing can be optimized. In FIG. 8, the unmeasured VIVT command value, fuel gas supply start timing, etc. may be determined by an approximate line connecting data before and after the measurement point.

Next, timing control of the end of fuel gas supply by the fuel gas supply valve 15 will be described with reference to FIGS.
FIG. 9 shows a main configuration of the engine 1 shown in FIG. In FIG. 9, a target rotation speed command unit 50 is installed outside the control unit 22, and a preset target rotation speed is input to the control unit 22. The gas supply time calculation unit 51 of the control unit 22 directly PID-controls the valve opening period of the fuel gas supply valve 15 based on the deviation between the actual rotation speed calculated from the measured value of the rotation speed sensor 20 and the target rotation speed. To do.
The gas supply valve control unit 52 connected to the gas supply time calculation unit 51 calculates the time to be opened from the opening timing of the fuel gas supply valve 15 as a starting point and outputs the calculated time to the fuel gas supply valve 15. Feedback control is performed so that the fuel gas supply valve 15 is opened for the time to be used.

The closing timing control of the fuel gas supply valve 15 is performed as follows. That is, as shown in FIG. 10, the control unit 22 directly sets the valve opening period of the fuel gas supply valve 15 to PID based on the deviation between the target rotational speed set by the target rotational speed command unit 50 and the actual rotational speed. Control. Specifically, based on the deviation between the target rotational speed and the actual rotational speed, the time during which each fuel gas supply valve 15 is open is controlled by feedback control so that the actual rotational speed follows the target rotational speed.
The gas supply valve control unit 52 controls the closing timing of each fuel gas supply valve 15 based on the valve opening period calculated from the opening timing of the fuel gas supply valve 15 as a starting point. The controller 22 directly performs PID control of the opening period of the fuel gas supply valve 15 so that the actual rotational speed matches the target rotational speed without calculating the amount of fuel gas to be supplied in advance.

In the fuel gas supply pressure control, which will be described later, the supply pressure detected by the supply pressure gauge 54 provided in the intake pipe 13 is added to the pressure ΔP value set with the output and rotation speed data of the engine 1 as parameters. The fuel gas pressure regulator 55 is feedback-controlled so that there is no deviation between the value and the value of the fuel gas pressure gauge 43.

FIG. 11 shows the relationship between the advance angle of the variable intake valve timing mechanism 30 displaying the above results and the start and end timings of the supply of the fuel gas supply valve 15 by the fuel gas supply timing means 44.
FIG. 11 shows the relationship between the crank angle of the engine 1 and the valve lifts of the intake valve 8 and the exhaust valve 9. In the curve indicating the opening / closing operation of the intake valve 8, the solid line indicates the case where the VIVT (variable intake valve timing) command value is 0%, and the alternate long and short dash line indicates the advance angle (the VIVT command value is 100%). The opening and closing operation image in the case of is shown. Then, the valve opening period of the fuel gas supply valve 15 at the time of advance (the VIVT command value is 100%) is longer than the valve opening period of the fuel gas supply valve 15 when the VIVT command value is 0%.

Further, it is preferable to increase the fuel supply amount by increasing the supply pressure of the fuel gas as the valve closing timing of the intake valve 8 advances. Therefore, in FIG. 9, the supply pressure of the fuel gas into the intake pipe 13 is set to a magnitude obtained by adding the pressure ΔP value to the supply pressure detected by the supply pressure gauge 54 provided in the intake pipe 13. As will be described later, the pressure ΔP is set with parameters of output and rotational speed data of a plurality of engines 1 measured in advance. As a result, the supply pressure of the fuel gas supplied from the fuel gas supply valve 15 increases as the intake valve 8 closes.

A method of setting the fuel gas pressure ΔP will be described.
FIG. 12 is a graph showing the relationship between the output, the valve opening period of the fuel gas supply valve 15, and the pressure ΔP. The operating condition of the engine 1 is changed with the output (load factor) and the rotation speed as parameters, and the pressure ΔP is changed under each condition to obtain the pressure ΔP at each output and the valve opening period of the fuel gas supply valve 15. If the pressure ΔP is set low based on the output and the rotation speed, the valve opening period of the fuel gas supply valve 15 becomes long. If the valve opening period is too long, it becomes impossible to supply appropriate gas fuel while the intake valve 8 is open. . Conversely, when the pressure ΔP is set high, the valve opening period of the fuel gas supply valve 15 is shortened, and the controllability of the supply amount is deteriorated. Therefore, the pressure ΔP is set to a pressure ΔP that does not adversely affect the operating state of the engine 1.

When the rotational speed is parameterized, the same procedure is repeated to determine the pressure ΔP. In FIG. 12, regarding the fuel gas supply valve 15, the lower limit value of the valve opening time is set by setting the upper limit of the pressure ΔP, and the upper limit value of the valve opening period is set by setting the lower limit of the pressure ΔP. The set value of the pressure ΔP is set as appropriate within the range of the upper limit value and the lower limit value of the valve opening period, and preferably the median value of the upper limit value and the lower limit value is set as the set value.
In addition, since the distribution of the air-fuel mixture changes depending on the supply pressure of the fuel gas, it is necessary to pay attention to the combustion state. In FIG. 12, the condition that is not measured may be determined by an approximate line that connects the data before and after the measurement point.

In this manner, when the set ΔP is plotted using the output (load factor) and the rotation speed as parameters, the result is as shown in FIG. In the three-dimensional map using the output, the rotation speed, and the pressure ΔP shown in FIG. 13 as parameters, the range indicated by the broken line is a normal (practical) operating region, and the solid line indicates the marine cube characteristic.
Since the fuel gas supply pressure is determined by the pressure ΔP, when the supply air pressure changes, the fuel gas supply pressure also changes. That is, it indicates that the differential pressure upstream and downstream of the fuel gas supply valve 15 is set, and the amount of fuel gas is determined by the differential pressure before and after the fuel gas supply valve 15 and the valve opening period. Even if it changes, the relationship between the fuel gas supply pressure determined here and the opening period of the fuel gas supply valve 15 is not significantly affected.
As can be seen from FIG. 13, when the output (load factor) of the engine 1 increases, that is, as the intake valve 8 closes, the fuel gas supply pressure including the pressure ΔP also increases.

Next, how to determine the supply air pressure (air-fuel ratio) in the engine 1 will be described.
Corresponding to the change in the VIVT command value (intake valve closing crank angle) according to the present embodiment described above, the supply air pressure supplied from the intake pipe 13 to the intake valve 8 depends on the output and rotational speed of the engine 1 measured in advance. According to the target supply air pressure set as a parameter, for example, a flow rate adjusting valve provided with a bypass line on each of the compressor side and the turbine side of the supercharger 17 shown in FIG. 9 is controlled to control the supply air pressure ( For example, see what is shown in Japanese Patent Application No. 2016-027359.) The supply pressure control is not limited to this control method. A conventionally known supply air pressure control method may be used.

A method for determining the above-described supply air pressure (air-fuel ratio) will be described below.
The amount of fuel serving as the base of the air-fuel ratio is determined by the relationship between the supply pressure of the fuel gas and the valve opening period of the fuel gas supply valve 15 shown in FIGS. A suitable amount of fuel is determined by the output (load factor) and the rotation speed. The air-fuel ratio is determined by the ratio of the air amount and the fuel gas amount. Therefore, the amount of air supplied to the combustion chamber 6 is changed by changing the supply air pressure. That is, the air-fuel ratio is adjusted by the supply air pressure.
The air-fuel ratio is set by changing the operating conditions of the engine 1 using the output (load factor) and the rotational speed as parameters, for example, setting the output (load factor) to 25%, 50%, 75%, 100%, etc. Then, the thermal efficiency and NOx data at various air-fuel ratios (supply pressure) are measured. FIG. 14 shows an example of NOx data when the air-fuel ratio is changed at an arbitrary output and rotational speed.

In FIG. 14, the measured value of the NOx data corresponding to the change in the air-fuel ratio is shown as a “measurement data” with a curve. In the measurement data, it can be recognized that NOx increases by adjusting the air-fuel ratio small (lowering the supply air pressure).
Here, the reference value of the value of NOx varies depending on the application. For the marine use, for example, the upper limit value and the lower limit value are regulated by the NOx emission amount indicated by the Air Pollution Control Law and the like based on the IMO NOx regulation based on the revised MARPOL Convention Annex VI rule 13 for marine use.
The lower limit value of the air-fuel ratio is limited by the upper limit value based on the above-mentioned NOx emission amount regulation. However, when abnormal combustion such as knocking occurs before the NOx data reaches the upper limit regulation value, the air-fuel ratio immediately before the abnormal combustion occurs is set to the lower limit value.
On the other hand, when the air-fuel ratio is increased, NOx decreases, but the engine 1 cannot be stably operated due to misfire or the like. Therefore, the upper limit of the air / fuel ratio at which stable operation can be continued is set to the upper limit value. These determine the range of air-fuel ratios that can be set.

Here, a settable air-fuel ratio range is determined, and an air-fuel ratio in the middle of the settable range is set as a suitable value, and the supply pressure at that time is set. The same measurement is repeated at any output and rotational speed within the range that is regularly used. In other words, a suitable supply air pressure (air-fuel ratio) that satisfies the target performance is obtained by changing the supply air pressure (air-fuel ratio) under the operating conditions of each output and rotation speed.
FIG. 15 is a three-dimensional map in which the intake pressure set corresponding to the output and the rotational speed is plotted. In the figure, a region indicated by a broken line is an image of a regular (practical) operation region, and a marine cube characteristic indicated by a solid line is set within the range. In addition, what is necessary is just to determine the conditions which are not measured by the approximate line which ties the data before and behind a measurement point.
As is apparent from FIG. 15, when the output of the engine 1 increases, that is, as the intake valve 8 closes, the required supply air pressure increases.
As described above, the air-fuel ratio control is performed by the supply start and end timing control of the fuel gas supply valve 15, the supply pressure control of the fuel gas, and the supply air pressure control.

When the timing for closing the intake valve 8 is advanced, the air-fuel ratio becomes small at the same supply pressure. For this reason, in the relationship between the output and the intake pressure shown in FIG. 16, as compared with the case where the closing timing of the intake valve 8 is constant, the optimum as the output increases with the advance timing of the closing timing of the intake valve 8. The rate of increase (increase in gradient) of the supply air pressure increases.

Next, it is necessary to change the fuel ignition timing in response to the change in the VIVT specified value (the crank angle of the intake valve closing timing). The ignition timing of the engine 1 is, for example, a fuel oil injection timing for ignition by the micropilot oil injection valve 11, and this ignition timing is set by using data of the output value and the rotational speed of the engine 1 measured in advance as parameters. Is set from the ignition timing value shown in FIG.

A method for determining the ignition timing will be described.
The operating conditions of the engine 1 are changed using the output (load factor) and the rotation speed as parameters, and thermal efficiency and NOx data of various ignition timings such as a load factor of 25%, 50%, 75%, and 100% are obtained. FIG. 18 shows an example of data on thermal efficiency and NOx when the ignition timing is changed at an arbitrary output and rotational speed. As shown by the solid line in FIG. 18, as the ignition timing is advanced, NOx increases and the thermal efficiency is improved. Therefore, thermal efficiency and NOx are in a trade-off relationship.
As described above, since NOx has a predetermined reference value, the ignition timing advanced to the highest thermal efficiency is set as a suitable timing within a setting range that satisfies NOx, which is the same reference as the supply air pressure. .

However, if abnormal combustion such as knocking starts before the NOx data reaches a predetermined reference value, the ignition timing before the start of abnormal combustion is set as a suitable ignition timing. Regarding the ignition timing of the micropilot oil injection valve 11, measurement is repeated at an arbitrary output and rotational speed within a range that is routinely used. In other words, under the operating conditions of “output (load factor) and rotational speed”, the ignition timing is changed to obtain an ignition timing that satisfies the target performance. The ignition timing was determined by repeating the same adjustment at an arbitrary rotation speed and output within the range of regular operation. In addition, what is necessary is just to determine the conditions which are not measured by the approximate line which ties the data before and behind a measurement point.

As apparent from FIG. 17, the region where the output is lower than the solid line indicating the image of the ship cubic power characteristic within the range of the regular operation region indicated by the broken line indicates the torque poor region. In the torque poor region, if the output of the engine 1 increases, that is, as the intake valve 8 closes, the ignition timing of the micro pilot oil injection valve 11 of the engine 1 also advances.

On the other hand, the region where the output is higher than the marine cube characteristic line indicated by the solid line indicates the torque rich region. In the torque rich region, the ignition timing peaks on the broken line in the normal operation region, and the advancement degree can be reduced when the ignition timing is in the range of 30 to 50% at the maximum with respect to the peak value. However, within the torque poor region and the torque rich region, the ignition timing can be kept advanced at the same rotational speed as compared with the case where the VIVT command value is not advanced. That is, with the advancement of the intake valve closing crank angle of the VIVT command value, the ignition timing is also advanced with a peak on the solid line indicating the ship cube characteristic. Although the degree of advance is reduced in the torque rich region, the fuel injection timing (ignition timing) of the micropilot oil injection valve 11 of the engine 1 as a whole is greater than when the VIVT command value is not advanced. Also advance.

As described above, according to the control method of the engine 1 and the engine 1 according to the embodiment of the present invention, it is possible to reduce the amount of unburned gas fuel that is discharged in a regular operation region. Advantages can also be gained for the environment, such as reduction
Further, the higher the output of the engine 1, the narrower the operable range with respect to the air-fuel ratio. Therefore, maintaining a stable operating state is particularly effective for expanding the operating range in the torque rich region where the output is higher than the marine cube characteristic. Can be obtained.
Further, when the output of the output shaft of the engine 1 increases, the compression ratio of the air-fuel mixture in the engine 1 can be lowered, so that knocking at the time of increasing the load can be suppressed and the load increasing time can be shortened.

The engine according to the present invention is not limited to the dual fuel engine 1 according to the above-described embodiment, and can be appropriately changed or replaced without departing from the gist of the present invention. In the following, modifications and the like of the present invention will be described, but the same or similar parts and components as those described in the above-described embodiments will be described using the same reference numerals.

The engine according to the present invention is not limited to the dual fuel engine 1 capable of switching between the diesel mode D in which liquid fuel is the main fuel and the gas mode G in which gas is the main fuel, and uses gas as the fuel. It can also be applied to gas fuel engines.
In addition, the present invention is not limited to the load increase pattern of the marine engine, but can be applied to a load increase pattern that can be used in a vehicle or an emergency generator.

In the above-described embodiment, the variable intake valve timing mechanism 30 changes both the valve opening timing and the valve closing timing and does not change the time during which the intake valve 8 is open. One or both of the valve opening timing and the valve opening timing may be selected and changed.

The present invention provides an engine control method and an engine system that use a premixed gas fuel and air to suppress knocking at the time of load increase and to shorten the load increase time.

DESCRIPTION OF SYMBOLS 1 Dual fuel engine 2 Crankshaft 8 Intake valve 9 Exhaust valve 13 Intake pipe 14 Exhaust pipe 15 Fuel gas supply valve 17 Supercharger 20 Rotational speed sensor 21 Torque sensor 22 Control part 24 First map 25 Second map 27 Electropneumatic conversion 28 Actuator 30 Variable intake valve timing mechanism 42 Gas regulator 44 Fuel gas supply timing means 45 Fuel gas supply valve timing mechanism

Claims

An engine control method using gas as fuel,
As the output of the engine increases, control is performed to lower the compression ratio of the air-fuel mixture in the combustion chamber by adjusting the advance angle from the suction bottom dead center when the intake valve closes.
An engine control method characterized by advancing the opening timing of a fuel gas supply valve in response to a change in the advance angle.
2. The engine control method according to claim 1, wherein the degree of advancement of the opening timing of the fuel gas supply valve increases as the advancement of the closing timing of the intake valve proceeds.
The closing timing of the fuel gas supply valve is calculated by calculating a valve opening period based on a deviation between the target rotation speed and the actual rotation speed, and setting the valve opening period from the valve opening timing of the fuel gas supply valve as a starting point. The engine control method according to claim 1, wherein the engine control method is determined by:
The engine control method according to any one of claims 1 to 3, wherein the supply pressure of the fuel gas is increased as the intake valve is closed.
The engine control method according to any one of claims 1 to 4, wherein the ignition timing of the engine is advanced in accordance with an advance angle of the closing timing of the intake valve.
The advance of the ignition timing has a peak at the ship's cube characteristic line that is a boundary between the torque rich region and the torque poor region set as parameters of the engine output and the rotational speed measured in advance.
6. The engine control method according to claim 5, wherein in the torque rich region, an advance angle is reduced from the marine cube characteristic line.
The output of the engine is an output shaft obtained from a torque measurement value obtained by measuring the torque of the output shaft of the engine with a torque sensor and a rotation speed measurement value obtained by measuring the rotation speed of the output shaft of the engine with a rotation speed sensor. The engine control method according to any one of claims 1 to 6, wherein the output value is an output value of the engine.
2. The advance angle of the closing timing of the intake valve is set from an advance value set by using the output and rotation speed data of a plurality of output shafts measured in advance as parameters. 8. The engine control method according to any one of 1 to 7.
An engine system comprising a four-stroke engine powered by gas,
A control unit that advances the closing timing of the intake valve of the engine when the output of the output shaft of the engine increases, and advances the opening timing of the fuel gas supply valve in response to a change in the advance angle When,
A variable intake valve timing mechanism that changes the timing at which the intake valve closes in accordance with the timing at which the intake valve closes set by the control unit;
A fuel gas supply valve timing mechanism for advancing the valve opening timing of the fuel gas supply valve in response to a change in the advance angle of the intake valve set by the control unit,
The engine system is characterized in that, as the output of the output shaft of the engine increases, the variable intake valve timing mechanism controls to lower the compression ratio of the gas / air mixture in the engine.
An engine system according to claim 9, wherein
A torque sensor for measuring the torque of the output shaft of the engine;
A rotational speed sensor for measuring the rotational speed of the output shaft of the engine,
An engine system in which an output of the output shaft is obtained from a torque measurement value obtained by the torque sensor and a rotation speed measurement value obtained by the rotation speed sensor, and a change in the closing timing of the intake valve in the control unit is set.
The engine system according to claim 9 or 10, wherein
The closing timing of the fuel gas supply valve is
A gas supply time calculation unit for calculating a valve opening period based on a deviation between the target rotation speed and the actual rotation speed;
An engine system set by a gas supply valve control unit for instructing a closing timing of the fuel gas by the fuel gas supply valve based on the valve opening period;
The engine system according to any one of claims 9 to 11,
An engine system set to increase the supply pressure of the fuel gas as the intake valve closes at an advanced timing.
The engine system according to any one of claims 9 to 12,
An engine system that advances the ignition timing of the engine as the intake valve closes.
The engine system according to any one of claims 9 to 13,
The degree of advance of the ignition timing of the engine accompanying the advance of the closing timing of the intake valve is a marine vessel which is a boundary between a torque rich region and a torque poor region set with the output value and the rotational speed of the engine measured in advance as parameters. An engine system in which the cube characteristic line is a peak, and the advance angle is reduced in the torque rich region.