CN109247022B - Control method of engine and engine system - Google Patents

Control method of engine and engine system Download PDF

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Publication number
CN109247022B
CN109247022B CN201780025208.0A CN201780025208A CN109247022B CN 109247022 B CN109247022 B CN 109247022B CN 201780025208 A CN201780025208 A CN 201780025208A CN 109247022 B CN109247022 B CN 109247022B
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China
Prior art keywords
engine
valve
output
intake valve
fuel gas
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CN109247022A (en
Inventor
黑岩隆典
桥本彻
渡边孝一
结城和广
山泽美和子
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Ihi Prime Mover
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Ihi Prime Mover
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D13/00—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D43/00—Conjoint electrical control of two or more functions, e.g. ignition, fuel-air mixture, recirculation, supercharging or exhaust-gas treatment
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M51/00—Fuel-injection apparatus characterised by being operated electrically
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02P—IGNITION, OTHER THAN COMPRESSION IGNITION, FOR INTERNAL-COMBUSTION ENGINES; TESTING OF IGNITION TIMING IN COMPRESSION-IGNITION ENGINES
    • F02P5/00—Advancing or retarding ignition; Control therefor
    • F02P5/04—Advancing or retarding ignition; Control therefor automatically, as a function of the working conditions of the engine or vehicle or of the atmospheric conditions
    • F02P5/145—Advancing or retarding ignition; Control therefor automatically, as a function of the working conditions of the engine or vehicle or of the atmospheric conditions using electrical means
    • F02P5/15—Digital data processing
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/12—Improving ICE efficiencies
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/30—Use of alternative fuels, e.g. biofuels
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/40—Engine management systems

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Signal Processing (AREA)
  • Output Control And Ontrol Of Special Type Engine (AREA)
  • Electrical Control Of Ignition Timing (AREA)
  • Electrical Control Of Air Or Fuel Supplied To Internal-Combustion Engine (AREA)
  • Combined Controls Of Internal Combustion Engines (AREA)

Abstract

In a gas mode of a dual fuel engine (1) for a marine vessel, a mixed gas of fuel gas and air is combusted in a combustion chamber. The method comprises the following steps: a control unit (22) that advances the timing at which an intake valve (8) of the engine is closed when the output of an output shaft (2) of the engine (1) increases, and that advances the valve opening timing of a fuel gas supply valve (15) in accordance with a change in the advance angle; a variable intake valve timing mechanism (30) that advances the timing at which the intake valve (8) closes, in accordance with the closing timing of the intake valve (8) provided by the control unit; and a fuel gas supply valve timing mechanism (45) that advances the valve opening time of the fuel gas supply valve (15) in accordance with a change in the advance angle of the intake valve (8) provided in the control unit. As the output of the output shaft of the engine (1) increases, a control is performed to further reduce the compression ratio of the mixture of gas and air in the engine by means of the variable intake valve timing mechanism (30).

Description

Control method of engine and engine system
Technical Field
The present invention relates to a method for controlling an engine using a gas fuel such as natural gas, and an engine system, and more particularly, to a method for controlling an engine using a gas fuel, which has a Variable Intake Valve Timing (VIVT) mechanism, and an engine system.
Background
Specific examples of the variable valve timing mechanism are described in the following patent documents 1 to 3.
As shown in fig. 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, an exhaust valve rocker arm 103 connected to a push rod (プッシュロッド) of an engine exhaust valve is supported by a link shaft 104, and an intake valve rocker arm 105 connected to a push rod of an intake valve is supported by a tappet shaft (タペット axle) 106 having an eccentric shaft portion deviating from the link shaft 104.
The exhaust valve rocker arm 103 and the intake valve rocker arm 105 can be advanced and retracted by an eccentric cam 108a of the camshaft 108, respectively. The link shaft 104 is coupled to a piston rod 109 provided in the actuator 102.
If the position shown in fig. 21 is set before the ejection operation of the piston rod 109, the piston rod 109 is subjected to the ejection operation by the actuator 102 ( is retracted し), and all the linked rocker arms 105, 103 are rotated in one direction. Therefore, the rotation angles of all the rocker arms 105, 103 can be controlled by the actuator 102 through the link mechanism 101.
As another example, fig. 22 and 23 show a variable valve timing mechanism described in patent documents 2 and 3. These mechanisms will be described with the same reference numerals in the same portions as those of the variable valve timing mechanism 100 shown in fig. 20 and 21.
In the variable valve timing mechanism shown in fig. 22, the rotation range of the link shaft 104 is restricted within the range of the tooth portion of the sector gear 120 coupled to the actuator 102, and an eccentric disc 123 (corresponding to a tappet shaft) fixed eccentrically to the link shaft 104 is held by the base portions of the exhaust valve rocker arm 103 and the intake valve rocker arm 105.
Therefore, the position at which the eccentric cam 108a of the camshaft 108 abuts against the exhaust valve rocker arm 103 and the intake valve rocker arm 105 and is pushed upward changes with respect to the deviation of the rotational angle position of each eccentric disc 123 from the rotational position of the link shaft 104.
In the example shown in fig. 23, the exhaust valve rocker arm 103 and the intake valve rocker arm 105 coupled to the rocker 127 via the pushrod 128 are connected to the tappet shaft 106 (the fulcrum position of the rocker arm) of the crank-like link shaft 104. By changing (rotating) the phase of the crank-like link shaft 104 by the actuator 102, the fulcrum positions of the intake rocker arm 105 and the exhaust rocker arm 103 are changed, and as a result, the position of the tangent point to the camshaft 108 is changed.
Thereby, the time for which the eccentric cam 108a of the camshaft 108 pushes the exhaust valve rocker arm 103 or the intake valve rocker arm 105 to advance and retract by the camshaft 108 is made variable.
Patent document 4 discloses a control device applied to an internal combustion engine, including: a first intake valve and a second intake valve provided in each combustion chamber; a fuel injection device provided in the intake port and injecting fuel to the first intake valve and the second intake valve, respectively; 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. The internal combustion engine includes: a valve timing setting unit that controls the variable valve mechanism, sets an opening timing of the first intake valve before top dead center, an opening timing of the second intake valve after top dead center, and a closing timing of the first intake valve and a closing timing of the second intake valve after bottom dead center; and an injection timing setting unit that sets a start time of fuel injection from the fuel injection device to the first intake valve to a timing that is retarded from the top dead center as an amount of gas blown back to the intake port side via the first intake valve increases in a set state of the valve timing setting unit.
Patent document 4 relates to an apparatus for realizing stratified combustion in internal EGR of a vehicle engine, and the fuel injected from a fuel injection device into a first intake valve is liquid fuel that can be injected at one instant.
Patent document 5 discloses: comparing a negative rate of change of the target EGR rate with a predetermined threshold value, and detecting that there is a response delay of the EGR gas accompanying a closing operation of the EGR valve during a valve opening period of the EGR valve when the negative rate of change is equal to or greater than the threshold value; when a delay in response of the EGR gas is detected, the closing period of the intake valve is corrected to raise the actual compression ratio, and the fuel injection period is corrected to approach compression top dead center.
Patent document 5 is for coping with a response delay of EGR gas in external EGR of a diesel engine. The fuel injected into each cylinder via the injector is liquid fuel that can be instantaneously filled into the combustion chamber.
Patent document 6 is a self-ignition combustion type premixed compression ignition engine that detects a set output of the engine, increases the equivalence ratio of premixed gas (set according to the amount of fuel supplied) as the set output increases, and decreases the actual compression ratio. The set output of the engine, in other words, the required output of the engine is set manually or by detecting a load applied to a crankshaft of the engine. The "load applied to the crankshaft" is the "set output, in other words, the output of the required engine", and therefore is not the output of the output shaft which is actually output by the engine, but the output which is set and required for the engine.
Patent document 7 describes that the output of a gas-fueled engine (ガス fuel エンジン) is regulated by the number of revolutions and torque in a subchamber gas engine.
Documents of the prior art
Patent document
Patent document 1: international publication No. 2015/060117
Patent document 2: european patent publication No. 2136054
Patent document 3: japanese patent laid-open No. 62-99606
Patent document 4: japanese patent No. 5502033
Patent document 5: japanese patent No. 5338977
Patent document 6: japanese laid-open patent publication No. 2002-21608
Patent document 7: japanese laid-open patent publication No. 2013-185515
Disclosure of Invention
Problems to be solved by the invention
In the conventional gas fuel engine, even if the supply amount of gas fuel is increased early in order to increase the output quickly, the supercharger cannot follow up, and the required amount of air cannot be supplied. This is because the supercharger is driven by the exhaust gas, and therefore, when the output of the gas fuel engine is increased and the exhaust gas is not sufficiently supplied to the supercharger, the supercharger cannot be operated efficiently. When the amount of air is insufficient, the air-fuel ratio becomes excessive, so that knocking occurs, resulting in a failure of the apparatus. When the output is raised at a speed that the supercharger can follow to suppress such knocking, the raising of the output (load raising) requires a time of about 10 minutes. In the present specification, the term "output" is used in principle for the power actually output by the engine, and the term "load" is used in principle for the power set and requested for the engine, but in many cases, the term "output" is "load", and the term "load" is used instead of "output" (or vice versa) in some cases.
On the other hand, in marine devices, since the emission limit of harmful exhaust gas is becoming strict year by year, there is a demand for a dual-fuel engine that can reduce the emission amount of harmful exhaust gas derived from fuel and satisfy the emission limit. However, in order to introduce such a dual-fuel engine, it is necessary to shorten the load-raising time to about 20 seconds to satisfy the operation mode of the marine device.
The present invention has been made in view of the above-described problems, and an object of the present invention is to provide an engine control method and an engine system that can suppress knocking that occurs when the output of a gas-fueled engine is increased, shorten the load increase time, and prevent deterioration in fuel consumption due to blow-by of unburned gas fuel in valve overlap.
Means for solving the problems
Namely, the present inventors have found that: the present invention has been completed by the finding that the open valve (supply start) time of a fuel gas supply valve is changed at each VIVT angle, an optimal value is determined from the THC concentration and the combustion state, and the open valve time of the fuel gas supply valve is set according to the VIVT angle, whereby knocking occurring when the output of a gas-fueled engine is boosted can be suppressed and the load boost time can be shortened, and further, combustion fluctuation and rotational speed fluctuation caused by the open valve time of the fuel gas supply valve in an excessive torque region (トルクリッチ region) and a torque shortage region (トルクプア region) can be improved.
The method for controlling an engine according to the present invention is a method for controlling an engine using gas as fuel, wherein control is performed to reduce the compression ratio of the air-fuel mixture in the combustion chamber by adjusting an advance angle by which the closing time of the intake valve is advanced from the intake bottom dead center as the engine output increases, and the opening time of the fuel gas supply valve is advanced in accordance with a change in the advance angle.
As a knock suppression technique for an engine, a Variable Intake Valve Timing (VIVT) mechanism may be used to reduce the effective compression ratio. In this regard, the knock suppression technique will be explained by fig. 19A, 19B. Fig. 19A shows the steps of a normal four-stroke cycle, and fig. 19B shows the steps of a miller cycle.
For example, in a gas fuel engine, the intake valve is normally closed at the bottom dead center of the piston (refer to fig. 19A). On the other hand, as shown in fig. 19B, when the closing time is earlier than the bottom dead center, the mixture gas continues to expand even after the intake valve closes, and therefore the in-cylinder temperature Ts is decreased from that in fig. 19A (Ts × Ts). The highest compression temperature at the top dead center is also lowered only so much (Tc × Tc), whereby self-ignition can be prevented and knocking can be suppressed.
The miller cycle has a disadvantage in that the compression temperature is lowered, and the ignitability in the low load region is deteriorated, so that it is necessary to return to the normal valve opening time of the intake valve shown in fig. 19A at the time of start-up and at the time of low load, and the valve opening time of the intake valve is advanced only at the time of high load.
In a method for controlling an engine, in a gas fuel engine in which the output is increased by increasing the amount of gas fuel supplied, control is performed to reduce the compression ratio of the air-fuel mixture in the combustion chamber of the engine by advancing (advancing) the timing at which the intake valve is closed from the intake bottom dead center. By lowering the compression ratio, the temperature in the combustion chamber at the time of compression is lowered, and therefore knocking can be suppressed.
Further, when the compression ratio of the air-fuel mixture is lowered in the combustion chamber, ignitability is deteriorated at the time of start-up and at the time of low output, and combustion efficiency is deviated from favorable conditions, resulting in a disadvantage of deterioration of fuel consumption. Therefore, the present invention performs control of changing the timing of closing the intake valve more largely in an operation region of higher output where knocking is more likely to occur, and lowering the compression ratio by a larger ratio. Thus, knocking can be suppressed in accordance with a change in output, deterioration in fuel consumption can be prevented, and the load raising time can be shortened. Further, by advancing the valve opening time of the fuel gas supply valve in accordance with the change in the advance angle of the intake valve, blow-by of unburned fuel gas can be reduced when the valves of the intake valve and the exhaust valve overlap.
Preferably, the degree of advance of the valve opening time of the fuel gas supply valve is greater as the advance of the closing time of the intake valve progresses.
This can further reduce the blow-by of unburned fuel gas in the combustion chamber when the intake valve and the exhaust valve overlap.
Further, it is preferable that the valve closing time of the fuel gas supply valve is determined by: the valve opening period is calculated from the deviation between the target rotation speed and the actual rotation speed, and is set starting from the valve opening time of the fuel gas supply valve.
The valve opening period of the fuel gas supply valve is calculated by feedback control such as PID control based on the deviation between the target rotational speed and the actual rotational speed of the engine. The valve opening period is set with the valve opening time of the fuel gas supply valve as a starting point, whereby the valve closing time of the fuel gas supply valve can be determined.
Further, it is preferable that the supply pressure of the fuel gas is increased as the advance angle of the closing time of the intake valve is advanced.
The supply pressure of the fuel gas is increased in accordance with the advance of the valve closing timing of the intake valve, whereby an appropriate amount of fuel gas for maintaining the output can be supplied during the valve opening period.
Further, it is preferable that the ignition time of the engine is advanced as the closing time of the intake valve is advanced.
The heat efficiency can be improved by advancing the ignition timing within a range in which NOx is limited to a predetermined value, in accordance with the advance angle of the closing timing of the intake valve.
The degree of advance of the ignition timing may be set such that a ship-use cubic characteristic line, which is a boundary between an excessive torque region and a deficient torque region and is set using an output and a rotational speed of an engine measured in advance as parameters, is set as a peak, and the degree of advance may be reduced in the excessive torque region as compared with the ship-use cubic characteristic line.
In this case, the ignition timing is advanced in the front region of the excessive torque region and the insufficient torque region with the cubic characteristic line for the ship as a peak, compared with before the intake valve closing timing is advanced, and therefore, there is an advantage that NOx is suppressed in a limited range as a whole and the thermal efficiency is improved. The "ship cubic characteristic" is a characteristic of the ship main engine that outputs an output proportional to the third power of the rotation speed, but in an actual ship, the output is not necessarily exactly proportional to the third power, and there is a certain degree of variation.
Further, it is preferable that the output of the engine is an output value of the output shaft obtained from a torque measurement value (obtained by measuring a torque of the output shaft of the engine by a torque sensor) and a rotation speed measurement value (obtained by measuring a rotation speed of the output shaft of the engine by a rotation speed sensor).
In the engine of the present invention, since the fuel gas as the fuel is an elastomer, it is relatively difficult to obtain an accurate fuel supply amount as compared with a liquid fuel. Therefore, it is preferable to actually measure the torque by the torque sensor and calculate the output from the relationship with the rotational speed. The product of the rotation speed measurement value of the output shaft obtained by the rotation speed sensor and the torque measurement value measured by the torque sensor can be used to obtain the output (load) of the engine output shaft in real time. Therefore, the advance angle of the closing time of the intake valve of the engine and the advance angle of the opening time of the fuel gas supply valve can be accurately set, and therefore, even if the output is increased, the combustion efficiency can be improved, and the operation of the gas fuel engine can be appropriately performed.
Further, it is preferable that the advance angle of the intake valve closing time is set based on a value of an advance angle set using data of output values of a plurality of output shafts and rotational speeds measured in advance as parameters.
The preferred value of the advance angle of the closing timing of the intake valve becomes larger when the output is large, but depends on the rotation speed. Therefore, by preparing a map including at least these two parameters and controlling the advance of the closing timing of the intake valve in accordance with the change in the output and rotational speed of the engine, knocking can be further suppressed.
An engine system according to the present invention is an engine system including a four-stroke engine using gas as fuel, comprising: a control unit that advances a closing time of an intake valve of the engine and advances a valve opening time of the fuel gas supply valve in accordance with a change in the advance angle when an output of an output shaft of the engine increases; a variable intake valve timing mechanism that changes the timing of closing the intake valve in accordance with the timing of closing the intake valve provided by the control unit; and a fuel gas supply valve timing mechanism that advances a valve opening time of the fuel gas supply valve in accordance with a change in an advance angle of an intake valve provided in the control unit, wherein the variable intake valve timing mechanism performs control to further reduce a compression ratio of a gas-air mixture in the engine as an output of an output shaft of the engine increases.
In the control method of the engine system according to the present invention, when the supply amount of the fuel gas increases and the output of the output shaft of the engine increases, the timing at which the intake valve is closed is advanced from the intake bottom dead center, thereby performing control to reduce the compression ratio of the air-fuel mixture in the combustion chamber of the engine. By lowering the compression ratio, the temperature in the combustion chamber at the time of compression is lowered, and therefore knocking can be suppressed.
Further, the present invention performs control of reducing the compression ratio at a larger ratio by changing the timing of closing the intake valve more largely in an operation region with higher output in which knocking is more likely to occur. Thus, knocking can be suppressed in accordance with a change in output, deterioration in fuel consumption can be prevented, and the load raising time can be shortened. Further, by advancing the valve opening time of the fuel gas supply valve in accordance with the change in the advance angle of the intake valve, the crank angle at which the intake valve is closed can reduce blow-by of unburned fuel gas when the valves of the intake valve and the exhaust valve overlap.
In addition, preferably, it comprises: a torque sensor that measures torque of an output shaft of the engine; and a rotation speed sensor for measuring the rotation speed of the output shaft of the engine, wherein the output of the output shaft is obtained according to the torque measurement value measured by the torque sensor and the rotation speed measurement value measured by the rotation speed sensor, and the change of the closing time of the air intake valve in the control part is arranged.
In the gas fuel engine of the present invention, since the gas as the fuel is an elastomer, it is relatively difficult to obtain an accurate fuel supply amount as compared with the liquid fuel. Therefore, it is preferable to actually measure the torque by the torque sensor and calculate the output from the relationship with the rotational speed. The output of the engine output shaft can be obtained in real time by using the product of the rotation speed measurement value of the output shaft obtained by the rotation speed sensor and the torque measurement value measured by the torque sensor.
Effects of the invention
With 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 reduced, so knocking at the time of load increase can be suppressed and the load increase time can be shortened.
Further, since the valve opening time of the fuel gas supply valve is advanced in accordance with the adjustment of the advance angle of the intake valve, it is possible to improve combustion variation and rotation speed fluctuation caused by the valve opening time of the fuel gas supply valve, and to appropriately perform the operation of the gas fuel engine.
Drawings
Fig. 1 is a block diagram showing the configuration of the main parts of a dual fuel engine for a ship according to an embodiment of the present invention.
Fig. 2 is a diagram showing a diesel mode and a gas mode in the dual fuel engine shown in fig. 1.
Fig. 3 is a perspective view (3 rd order element マップ) showing a relationship between an output, a rotational speed, and a VIVT command value.
Fig. 4 is a perspective view showing a relationship between an output, a rotation speed, and a valve closing time of an intake valve.
Fig. 5 is a graph showing a relationship between an output and an optimum VIVT command value when the rotational speed of the output shaft is constant or varies.
Fig. 6 is a graph showing the relationship between the open time of the fuel gas supply valve and the THC concentration when various differences occur in the VIVT command value.
Fig. 7 is a graph showing a relationship between an output and a valve opening time of the fuel gas supply valve when the rotation speed of the output shaft is constant or when the rotation speed is changed.
Fig. 8 is a graph showing a fuel gas supply start timing corresponding to the VIVT command value.
Fig. 9 is a diagram showing the configuration of a control device that performs PID control of a marine dual-fuel engine.
Fig. 10 is a diagram showing a procedure of performing PID control so as to change the actual rotational speed to the target rotational speed.
Fig. 11 is a timing chart showing the opening and closing operations of the intake valve and the fuel gas supply valve at the normal time and at the advanced time.
Fig. 12 is a diagram showing a relationship between the output, the valve opening period of the fuel gas supply valve, and the pressure Δ P of the fuel gas.
Fig. 13 is a perspective view showing a relationship between the output, the rotation speed, and the pressure Δ P of the fuel gas.
Fig. 14 is a graph showing the change in NOx and the usable range corresponding to the air-fuel ratio.
Fig. 15 is a perspective view showing a relationship among the rotation speed, the output, and the air supply pressure.
Fig. 16 is a graph showing the relationship between the output and the intake air pressure when the intake valve closing time is constant and when the advance angle is advanced.
Fig. 17 is a perspective view showing a relationship among the rotation speed, the output, and the ignition time.
Fig. 18 is a graph showing the relationship between NOx and thermal efficiency when the ignition timing is changed.
Fig. 19A is a step chart of a normal cycle in the combustion cycle of the engine.
Fig. 19B is a step chart of the miller cycle in the combustion cycle of the engine.
Fig. 20 is a perspective view showing an example of a conventional variable intake valve timing mechanism.
Fig. 21 is a front view showing an example of a conventional variable intake valve timing mechanism.
Fig. 22 is a perspective view showing another example of a conventional variable intake valve timing mechanism.
Fig. 23 is a diagram showing another example of a conventional variable intake valve timing mechanism.
Fig. 24 is a diagram showing a relationship between an actuator and a link shaft in the conventional variable intake valve timing mechanism shown in fig. 22.
EMBODIMENTS FOR CARRYING OUT THE INVENTION
Next, as an engine according to an embodiment of the present invention, for example, a four-stroke dual fuel engine 1 used as a marine engine will be described based on the drawings.
The marine dual fuel engine 1 shown in fig. 1 and 2 (hereinafter, may be simply referred to as an engine 1) has respective devices in a diesel mode D and a gas mode G, and can be switched to the diesel mode D and the gas mode G during operation. The dual fuel engine 1 shown in fig. 1 has a crankshaft 2 as an output shaft coupled to a propeller (プロペラ) or the like, and the crankshaft 2 is coupled to a piston 4 provided in a cylinder block 3. A piston 4 and an engine cover 5 provided in the cylinder block 3 form a combustion chamber 6
The combustion chamber 6 is sealed by an intake valve 8 and an exhaust valve 9 mounted in the engine cover 5, and a fuel injection valve 10 used in the diesel mode D. The engine cover 5 is provided with a pilot fuel injection valve (マイクロパイロット oil injection barrier) 11 for use in the gas mode. An intake pipe 13 is connected to an intake port of the intake valve 8 provided with the engine cover 5, and an exhaust pipe 14 is provided to an exhaust port provided with the exhaust valve 9. The intake pipe 13 is provided with a fuel gas supply valve 15 formed of an electromagnetic valve for controlling the injection of fuel gas, and an air cooler 16 and a supercharger 17 communicating with the exhaust pipe 14 are provided on the upstream side thereof.
Here, as shown in fig. 2, the dual-fuel engine 1 of the present embodiment can be switched to the diesel mode D and the gas mode G. In the diesel mode D, for example, heavy oil a or the like is mechanically injected as fuel oil from the fuel injection valve 10 into the compressed air in the combustion chamber 6 to ignite and burn. In the gas mode G, a fuel gas such as natural gas is supplied to the intake pipe 13 through the fuel gas supply valve 15, premixed with the air flow, and the mixed gas is supplied into the combustion chamber 6, and the pilot fuel is injected from the pilot fuel injection valve 11 in a compressed state of the mixed gas to ignite and burn. The micro pilot fuel injection valve 11 injects a small amount of pilot fuel as a powerful ignition source, for example, by electronic control. The fuel gas supply valve 15 is an electromagnetic valve that has a large opening formed with a very small stroke and can flow a large amount of gas in a short time.
The engine 1 is started in a diesel mode D in which liquid fuel is injected from the fuel injection valve 10 into the combustion chamber 6. After the supply of the gas pressure equal to or higher than the standard value to the engine 1 is confirmed, the gas mode G is operated in which the gas fuel is supplied to the intake pipe 13 by the fuel gas supply valve 15, mixed with air, and then flows into the combustion chamber 6, and the gas fuel is combusted.
When the vehicle stops, the vehicle stops after changing to the diesel mode D again. The diesel mode D and the gas mode G may be changed at a time other than the start time and the stop time.
The dual-fuel engine 1 of the present embodiment includes a gas engine system that performs output control during a load increase in the gas mode G. The structure of the gas engine system will be further described below.
In fig. 1, a rotation speed sensor 20 and a torque sensor 21 are attached to a crankshaft 2, the rotation speed sensor 20 measures the rotation speed (rotation speed) of the crankshaft 2, and the torque sensor 21 measures the engine torque. As the torque sensor 21, for example, a sensor that detects torque applied to a shaft from deformation can be used. The measurement data measured by the rotation speed sensor 20 and the torque sensor 21 are each output in the form of a signal to a 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, the torque sensor 21, and the like. That is, the output (load) a of the engine 1 is calculated by the following equations (1) and (2) with n being the rotational speed (rotational speed) of the crankshaft 2 measured by the rotational speed sensor 20 and T being the torque measured by the torque sensor 21. Where Lt is the rated output of the engine 1.
Output Lo 2 pi Tn/60 (1)
Output (load) A Lo/Lt × 100 (2)
Further, as a method of determining the output (load) of the engine 1, there are a method of estimating the output (load) based on information related to the operating state of the engine 1 such as the fuel supply amount, and a method of obtaining the output by providing a torque sensor 21 in a power transmission system of an output shaft of the engine 1 and actually measuring the torque. In a gas fuel engine, since a gas as a fuel is an elastomer, it is relatively difficult to obtain an accurate fuel supply amount as compared with a liquid fuel. Therefore, it is preferable to calculate the output by actually measuring the torque by the torque sensor 21.
When the rotation speed n is constant, the output a and the torque measurement value T are in a proportional relationship. In the case where the rotation speed n is constant, it is desirable to set the advance angle of the closing timing of the intake valve 8 at a larger ratio as the output a, that is, as the torque data T is larger.
The control unit 22 stores a first map 24 of a first electric signal for specifying the opening/closing time of the intake valve and a second map 25 for specifying the opening/closing time corresponding to the first electric signal, which are created in advance. The control unit 22 calculates the output a of the engine 1 by the above equations (1) and (2) 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. Then, the first electric signal corresponding to the opening/closing time of the suction valve 8 is selected in the first map 24 based on the rotation speed n and the output a. On the basis of this first electrical signal, the switching time of the suction valve 8 is determined in the second diagram 25 in accordance with the first electrical signal. In addition, a method of creating the first chart 24 and the second chart 25 will be described later.
The second electric signal of the switching time set by the control section 22 is sent to the electric idling converter 27, and the electric idling converter 27 converts the signal of the switching time into air pressure. This air pressure is sent to the actuator 28, which controls the driving of the variable intake valve timing mechanism 30. The air pressures P1, P2 for driving and control are supplied from the first pressure reducing regulator 34 and the electrical idle converter 27 to the actuator 28.
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 depressurized to a desired 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 air pressure P1 for driving. When the pressure P1 measured by the pressure gauge 36 is equal to or lower than a predetermined value, the engine 1 cannot be started.
The air pressure for driving the air-to-air converter 27 is further reduced in pressure by the second pressure reducing regulator 37 from the first pressure reducing regulator 34 and then supplied thereto. The electrical idling converter 27 supplies the air pressure corresponding to the input second electrical signal of the switching time to the actuator 28 as the air pressure P2 for adjusting the operation of the actuator 28. The air pressures P1 and P2 move the rod 28a of the actuator 28, thereby operating the variable intake valve timing mechanism 30.
The actuator 28 is, for example, a known P cylinder (a incidental point cylinder (ポジショナリ to きシリンダ)), and controls the forward and backward movement of the rod 28a based on pressures P1 and P2 input from the first pressure reducing regulator 34 and the electric idle converter 27. The driving of the variable intake valve timing mechanism 30 is controlled by changing the moving length of the rod 28a of the actuator 28, and the closing timing of the intake valve 8 is advanced (advanced) or retarded (retarded) from the intake bottom dead center, thereby performing control for lowering the compression ratio.
Since the time between the valve opening time and the valve closing time of the intake valve 8 is constant, if the valve opening time is advanced from the intake bottom dead center, the valve closing time is also advanced from the intake top dead center by the same time. Further, the present invention suppresses knocking and shortens the load-raising time by changing the time for opening and closing the valve in accordance with the output of the engine 1. 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 rotation speed n of the engine 1, and the opening/closing timing of the intake valve 8 is adjusted by the actuator 28 and the variable intake valve timing mechanism 30 so that knocking can be suppressed.
The structure of the variable intake valve timing mechanism 30 is conventionally known, and has the same structure as that shown in fig. 20 to 24. That is, in the variable intake valve timing mechanism 30, for example, a link shaft that sets a rotation angle range via a sector gear by the movement length of the rod 28a of the actuator 28 is disposed in parallel with a camshaft having an eccentric cam. An exhaust rocker arm is connected to the link shaft, and an intake rocker arm is connected to a tappet shaft provided at a position offset from the link shaft. An intake valve 8 is connected to the intake rocker arm via a push rod and a rocker arm; an exhaust valve 9 is connected to the exhaust rocker arm via a pushrod and a rocker arm.
The distance between the camshaft and the intake rocker arm changes according to the rotation angle of the tappet shaft corresponding to the rotation of the link shaft, and the time at which the eccentric cam of the camshaft starts to contact changes. This makes it possible to change the valve closing timing to an advanced angle (or a retarded angle). The farther the distance from the tappet shaft to the camshaft center, the earlier the closing time of the intake valve 8. The rotation angle of the tappet shaft varies depending on the length of movement of the rod 28a of the actuator 28. The moving length of the rod 28a is arbitrarily changed according to the control air pressures P1 and P2 supplied to the actuator 28.
The magnitude of the advance angle as the opening and closing timing of the intake valve 8 is determined by the timing at which the intake rocker arm connected to the tappet shaft of the link shaft starts to contact the eccentric cam of the camshaft.
In addition, a servomotor, not shown, may be used in place of the actuator 28 as a means for rotating the tappet shaft in the variable intake valve timing mechanism 30. At this time, the signal of the switching time from the second map 25 of the control unit 22 is input to the servo motor. The servo motor rotates the link shaft by an amount corresponding to the received signal and rotates the tappet shaft, thereby making it possible to change the opening and closing timing of the intake valve 8 with respect to the approach and the separation of the camshaft. In addition, when the servo motor is used, the configuration of the actuator 28, the air compressor 32, and the pressure gauge 38 is not necessary. In addition, instead of the electrical idle converter 27, a controller is used to drive the servo motor.
A supply mechanism for supplying the fuel gas to the fuel gas supply valve 15 for controlling the injection of the fuel gas into the intake pipe 13 will be described. In fig. 1, a gas fuel is supplied from an LNG gas tank 40 storing a gas fuel such as natural gas to a gas vaporizer 41, and the gas pressure is reduced to a desired gas pressure by a gas regulator 42.
The gas pressure at this time is displayed on the fuel gas pressure gauge 43, and is adjusted by changing the valve opening of the gas regulator 42, and is supplied from the fuel gas supply valve 15 into the intake pipe 13 as fuel gas for combustion. In the intake pipe 13, the gas fuel and the charge air cooled by the air cooler 16 are mixed and supplied to the combustion chamber 6. When the load is increased, the supply amount of the fuel gas is increased by the operation of the fuel gas supply valve 15.
Also, the second electric signal of the switching time set by the control portion 22 is sent to the fuel gas supply valve 15 via a fuel gas supply timing unit 44 different from the electrical idle converter 27. The fuel gas supply timing unit 44 controls the fuel gas supply valve 15 to be opened to advance the valve opening time for supplying the fuel gas into the intake pipe 13 in accordance with the advance angle of the closing time of the intake valve 8. Both 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 unit 44 may be provided outside the control unit 22. The fuel gas supply valve timing mechanism 45 may receive the second electric signal from the second map 25 and advance the open time of the fuel gas supply valve 15 in accordance with the advance of the close time of the intake valve 8.
Next, a method of creating the first chart 24 and the second chart 25 stored in the control unit 22 will be described. Fig. 3 is a perspective view showing the first map 24 in detail, and the first map 24 determines the VIVT command value (Intake Valve Closed crank angle, IVC), that is, the crank angle at which the Intake Valve 8 is Closed, from the rotational speed of the crankshaft 2 and the output (load factor) of the engine 1.
In fig. 3, a region B in which a usual (utility) operation is performed is indicated by a broken line. On the other hand, the change (advance angle) in the VIVT command value with respect to the change in the output at a constant rotational speed during power generation is indicated by an arrow line C; the change (advance angle) of the VIVT command value when the rotation speed and the output (load factor) change simultaneously in the marine vessel is indicated by an arrow line D. The arrow line D indicates the cubic characteristic for the ship. The ship cubic characteristic represents a representative characteristic of a ship main engine whose output is proportional to the third power of the rotational speed, and is a characteristic curve of the rotational speed and the output determined from the rated rotational speed and the rated output of the engine. In the region of the usual operation region B, a region with a high output (load factor) indicates a torque excess region and a region with a low output (load factor) indicates a torque shortage region, compared to the ship cubic characteristic line D.
The first chart 24 is created based on the flow of the following experimental steps (1) to (18).
The same type of bi-fuel engine 1 used in practice was used in the experiments.
(1) The engine 1 is started with a rotation speed (rotation speed) n of 400min-1The output (load) a is 10%, and the valve closing time of the intake valve 8 is 545deg (structurally latest valve closing time).
(2) Then, abnormal combustion called knocking generated when the engine 1 is driven and the exhaust gas temperature at that time are measured.
The occurrence of knocking is detected by a knock sensor, not shown, attached to each engine cover 5. When the knocking phenomenon occurs, a waveform is formed by superimposing a high-frequency pressure fluctuation on a normal combustion waveform.
Further, the exhaust temperature at the time of knock measurement is measured by a temperature sensor attached to the exhaust pipe 14.
(3) After the exhaust temperature at the time of the knock measurement is measured, the valve closing time of the intake valve 8 is reduced by 5deg, and the measurement of (2) is performed again. The valve closing time was changed to 500deg (structurally the earliest valve closing time) and measured.
(4) After the measurement of the above (3) is completed, the output a is increased to 110% every 10%, and the measurements of (2) and (3) are repeated again.
(5) By the measurement of the above (1) to (4), it is determined that knocking is suppressed and the engine 1 can be operated safely when the knock intensity is equal to or less than the standard value and the exhaust gas temperature is equal to or less than 500 ℃.
(6) From the measurement results of the above (5), in the perspective graph of fig. 4 in which the X axis is the output a, the Y axis is the rotation speed n, and the Z axis is the switching time, ● (black circles) is drawn at the measurement points that can be operated safely, and X is drawn at the measurement points that are not safe. Thereby, the knock suppression range in the relationship between the output a, the rotation speed n, and the valve closing time can be selected.
(7) The rotation speed n is set every 100min-1Lifting to 900min-1The measurement steps (1) to (6) are performed to measure the range of safe operation for each rotation speed n.
(8) Fig. 4 is a graph showing the measurement result of (7) above on three axes of the rotation speed n, the output a, and the valve closing time. In fig. 4, the range enclosed by the straight line is a range in which knocking is suppressed and the engine 1 can be operated safely.
(9) Next, in the range of the three-dimensional region surrounded by the straight lines capable of safely operating the engine shown in fig. 4 measured by the above-described experiments (1) to (8), the experiment was further performed for the purpose of setting nitrogen oxide (hereinafter, referred to as NOx) to a standard value or less and thereby maximizing the thermal efficiency.
The engine rotation speed n was set to 400min-1The output a is set to 10%, and the valve closing time 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 according to the following equation (3) based on the fuel flow rate L (measured by a fuel flow meter attached to the fuel pipe) and the output a (calculated from the measurement result of the torque sensor 21).
Thermal efficiency eta is 360Lo/H/L (3)
Wherein, H: lower heating value (J/Nm) of fuel gas3)
Lo: current output
L: flow rate of fuel
(11) After the measurement of the above (10) is completed, the valve closing time of the suction valve 8 is reduced every 5deg, and the measurement of the above (10) is performed again. The valve closing time was changed to 505deg and measured (see fig. 9).
(12) After the measurement of (10) and (11) is completed, the output is increased to 110% every 10%, and the measurement of (10) and (11) is repeated again. The valve closing time is varied within the range of safe operation shown in fig. 4.
(13) The rotation speed n is set every 100min-1Gradually increasing to 900min-1The above measurements (9) to (12) are performed, and the measurement point with the best performance for each rotation speed is determined.
(14) Then, the closing time of the intake valve 8 in which NOx is equal to or less than a predetermined value and the thermal efficiency is highest is set for each rotation speed n and output a. From the result, a prototype of the first chart shown in fig. 3 was created.
(15) Then, the rotation speed n and the output a are raised in accordance with an arbitrary load raising pattern, and knocking is detected. The load boost mode is a state of change per unit time of the output a (load factor) and the rotation speed n, and changes according to the propeller specification (shape, rotation speed) of the marine propulsion device.
(16) And (3) reducing the valve closing time of the measuring point with the detected knock intensity of (15) above a standard value by 3 deg.
(17) Thereafter, the steps of (15) and (16) are repeated until the knock intensity becomes a standard value or less, and a valve closing time at which knocking is suppressed is determined. When the valve closing time is reduced, the thermal efficiency is deteriorated. The setting of the valve closing time at which NOx and the result that the knock intensity is equal to or less than the standard value and the thermal efficiency is the highest are obtained as the set values of the rotation speed n and the output a.
(18) The valve closing time for suppressing knocking in (17) is measured at each of the rotation speed n and the output a, and a final first graph 24 shown in fig. 3 is created based on the result.
In fig. 3, the VIVT command value corresponding to the rotation speed and the output is shown by a graph (i.e., the 3 rd-order layer height グラフ) of a three-dimensional plane, and the upper side in the figure is a direction in which the closing time is further advanced. On the solid plane, the region indicated by the broken line is a practical operation region used for the actual operation of the ship propulsion device, and an example of a favorable load lifting pattern is indicated by a ship cubic characteristic line D. In the load increase in the practical operation region, the advance angle of the valve closing time is controlled to be increased as the output of the engine is increased.
In an example of a favorable load raising mode indicated by the ship cubic characteristic line D, the advance angle is set to be minimum at a position on the lower right in the figure where the rotation speed and the output are small, and the advance angle is increased as the rotation speed and the output increase. The rate of increasing the advance angle is not constant, but the advance angle is set to be larger as the output increases as a whole. Since the output (load factor) is obtained from the product of the torque and the rotational speed, the advance angle may be increased as the torque of the output shaft increases.
Thereafter, a second chart 25 was created by the following experiment.
When the variable intake valve timing mechanism 30 is rotation-controlled by the actuator 28, the second map 25 is created as follows.
(1) The valve closing time is changed by the actuator 28, and the pressure at the time of the change to each valve closing time is measured.
(2) The second electric signal required to supply the pressure of (1) above was investigated according to the specification of the electric idling transformer 27.
(3) From the results of the above (1) and (2), a second graph 25 is created in which the horizontal axis represents the first electric signal selected by the above first graph 24 and the vertical axis represents the valve-closing time (second electric signal).
In the case where the actuator 28 is used as described above, the rotation control of the variable intake valve timing mechanism 30 using a servo motor instead of the actuator 28 is performed as follows.
(1) The valve closing time is changed based on the servo motor, and a second electric signal is measured when the valve closing time is changed.
(2) From the result of the above (1), a second graph 25 is created in which the horizontal axis represents the first electric signal and the vertical axis represents the valve-closing time (second electric signal).
The second graph 25 is a graph showing a relationship between the valve closing time (second electric signal) and the first electric signal.
In the perspective view shown in fig. 3, the optimum VIVT command value differs depending on the output between the characteristic line for power generation indicated by the solid line C and the cubic characteristic line D for the ship. That is, as shown in an example in fig. 5, even when the output is the same, if the rotation speed is different, the intake valve closing crank angle of the optimum VIVT command value is different.
In the present embodiment, the opening time of the fuel gas supply valve 15 for supplying fuel gas to the intake pipe 13 is set in accordance with the change in the VIVT command value, that is, the crank angle for closing the intake valves, so as to reduce blow-by of unburned fuel gas to the exhaust pipe 14 when the intake valve 8 and the exhaust valve 9 overlap each other. For this purpose, a VIVT command value corresponding to the rotation speed and output is first set. Strictly speaking, it is preferable to set the air-fuel ratio and the ignition timing to optimum values also in advance with the thermal efficiency and NOx as targets, but here it is assumed that the engine 1 can be stably operated without setting these conditions.
As an example of setting the open time of the fuel gas supply valve 15, a method of determining the open time of the fuel gas supply valve 15 by the fuel gas supply timing unit 44 under the engine operating condition corresponding to the optimum VIVT command value in the ship cubic characteristic line D will be described below.
First, under each of the operating conditions in which the output (load factor) of the engine is 25%, 50%, 75%, and 100%, the fuel gas is supplied from the fuel gas supply valve 15 with the open time of the intake valve 8 as a standard, but the fuel gas is supplied into the intake pipe 13, so the fuel gas does not instantaneously reach the intake valve 8. Therefore, the crank angle position of the open time of the fuel gas supply valve 15 is assumed in consideration of the distance from the fuel gas supply valve 15 to the intake valve 8. Then, the crank angle position of the fuel gas supply valve 15 during the valve opening time is changed to 5deg marks before and after the change, and the total hydrocarbon concentration (THC concentration) as unburned gas in the exhaust gas at the gas turbine outlet of the supercharger 17 at that time is measured. The measurement of the THC concentration was repeatedly performed under each operating condition. The THC concentration is preferably measured by a hydrogen flame ionization method (JIS B7956).
The open time of the fuel gas supply valve 15 was changed according to each condition, and each VIVT command value (intake valve closing crank angle) was set to 40%, 65%, 85%, 100%, for example, and the relationship between the open time of the fuel gas and the measured THC concentration in each VIVT command value was shown as shown in fig. 6.
As shown in fig. 6, the opening time of the fuel gas supply valve 15 is set based on the crank angle at which the blow-by of unburned fuel gas is small and the THC concentration is the lowest at the time of valve overlap. On the other hand, if the output change causes a rapid change in the valve opening time of the fuel gas supply valve 15, the change is related to the combustion variation and the rotation speed variation. Therefore, in order to make the amount of change in the valve opening time of the fuel gas supply valve 15 according to the output tend to be extremely small, the crank angle of the fuel gas valve opening time of the fuel gas supply valve 15 that is optimal is selected within the range of ± 5deg.c.a from the selected reference, and the crank angle corresponding to the optimal valve opening time of the fuel gas supply valve 15 is determined under each condition.
The relationship between the crank angle and the output (load factor) of the optimal opening time of the fuel gas supply valve 15 in the optimal VIVT command value of the ship cubic characteristic line D thus determined is represented by a broken line of "change in rotational speed" in fig. 7.
Similarly, the relationship between the crank angle of the optimal opening time of the fuel gas supply valve 15 and the output (load factor) of the optimal VIVT command value in the output with a constant rotational speed indicated by the power generation characteristic line C in fig. 3 is indicated by a broken line of "constant rotational speed" in fig. 7.
As shown in fig. 7, even if the output is the same, the optimum open time of the fuel gas supply valve 15 differs between the condition where the rotation speed is changed and the condition where the rotation speed is constant.
However, as shown in fig. 8, the crank angle of the optimum opening time of the fuel gas supply valve 15 among the optimum VIVT command values of the ship cubic characteristic line D and the crank angle of the optimum opening time of the fuel gas supply valve 15 among the optimum VIVT command values of the power generation characteristic line C (constant rotational speed) are adjusted on the horizontal axis instead of the output, and both show a single line characteristic. That is, it is understood that the open time of the optimum fuel gas supply valve 15 is not dependent on the output, but is dependent on the VIVT specified value (intake valve closing crank angle).
As can be seen from fig. 8, when the closing timing of the intake valve 8 is changed by the variable intake valve timing mechanism 30, the degree of advance of the supply start timing of the fuel gas supply valve 15 becomes greater as the advance of the closing timing of the intake valve 8 progresses.
Therefore, the fuel gas supply timing unit 44 sets the crank angle of the optimum open time of the fuel gas supply valve 15 determined under each condition based on the VIVT command value, thereby optimizing the open time of the fuel gas supply valve 15 by the VIVT command value. In fig. 8, the unmeasured VIVT command value, the fuel gas supply start timing, and the like can be identified by an approximate line connecting data before and after the measurement point.
Next, the timing control for ending the supply of the fuel gas by the fuel gas supply valve 15 will be described with reference to fig. 9 and 10.
Fig. 9 shows a main part configuration of the engine 1 shown in fig. 1. In fig. 9, a target rotation speed command unit 50 is provided 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 performs PID control of the valve opening period of the fuel gas supply valve 15 based on the deviation between the actual rotation speed calculated from the measurement value of the rotation speed sensor 20 and the target rotation speed.
The gas supply valve control unit 52 connected to the gas supply time calculation unit 51 performs feedback control so that the time to open the valve is calculated from the time to open the fuel gas supply valve 15, and the calculated time is output to the fuel gas supply valve 15 so that the fuel gas supply valve 15 opens only during the time to open the valve.
The valve 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 performs PID control of the valve opening period of the fuel gas supply valve 15 based on the deviation between the target rotation speed set by the target rotation speed command unit 50 and the actual rotation speed. Specifically, the time for which each fuel gas supply valve 15 is opened is controlled by feedback control based on the deviation between the target rotation speed and the actual rotation speed so that the actual rotation speed follows the target rotation speed.
The gas supply valve control unit 52 controls the valve opening time of each fuel gas supply valve 15 based on the valve opening period calculated from the valve opening time of the fuel gas supply valve 15 as a starting point. The control unit 22 directly performs PID control on the valve opening period of the fuel gas supply valve 15 so that the actual rotation speed and the target rotation speed match without calculating the amount of fuel gas to be supplied in advance.
The fuel gas supply pressure control described later performs feedback control of the fuel gas pressure regulator 55 so that there is no difference between a value obtained by adding the supply pressure detected by the supply pressure gauge 54 provided in the intake pipe 13 to a pressure Δ P value set using data of the output and the rotation speed of the engine 1 as parameters and the value of the fuel gas pressure gauge 43.
The relationship between the advance angle of the variable intake valve timing mechanism 30 and the timing at which the supply of the fuel gas from the fuel gas supply valve 15 by the fuel gas supply timing means 44 is started and ended is shown in fig. 11.
Fig. 11 shows a relationship between a crank angle of the engine 1 and valve lifts (バルブリフト) of the intake valve 8 and the exhaust valve 9. In the graph showing the opening and closing operation of the intake valve 8, the solid line shows the opening and closing operation image when the VIVT (variable intake valve timing) command value is 0%, and the alternate long and short dash line shows the opening and closing operation image when the angle is advanced (the VIVT command value is 100%). Further, the valve-opening period of the fuel gas supply valve 15 at the time of the advance (the VIVT command value is 100%) is longer than the valve-opening period of the fuel gas supply valve 15 at the time of the VIVT command value being 0%.
Further, it is preferable to increase the supply pressure of the fuel gas and increase the fuel supply amount as the valve closing timing of the intake valve 8 advances. Therefore, in fig. 9, the supply pressure for supplying the fuel gas into the intake pipe 13 is set to a value obtained by adding the pressure Δ P to the supply air pressure detected by the supply air pressure gauge 54 provided in the intake pipe 13. The pressure Δ P is set using data of the output and the rotational speed of the plurality of engines 1 measured in advance as parameters, as will be described later. As a result, the supply pressure of the fuel gas supplied from the fuel gas supply valve 15 increases with the advance of the closing timing of the intake valve 8.
Next, a method of setting the pressure Δ P of the fuel gas will be described.
Fig. 12 is a diagram showing a relationship between the output, the valve opening period of the fuel gas supply valve 15, and the pressure Δ P. The operating conditions of the engine 1 are changed using the output (load factor) and the rotation speed as parameters, and the pressure Δ P is changed under each condition to obtain the pressure Δ P and the valve opening period of the fuel gas supply valve 15 at each output. If the pressure Δ P is set low in accordance with the output and the rotation speed, the valve opening period of the fuel gas supply valve 15 becomes long, and if the valve opening period is too long, appropriate fuel gas cannot be supplied while the intake valve 8 is open. Conversely, when the pressure Δ P 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 Δ P should be set to a pressure Δ P that does not adversely affect the operating state of the engine 1.
The same steps are also repeated while parameterizing the rotation speed, determining the pressure Δ P. In fig. 12, the fuel gas supply valve 15 is provided with an upper limit of the pressure Δ P to set a lower limit of the valve opening time, and is provided with a lower limit of the pressure Δ P to set an upper limit of the valve opening period. The range between the upper limit value and the lower limit value of the valve opening period is set as a set value for appropriately setting the pressure Δ P with a variable width, and preferably, the intermediate value between the upper limit value and the lower limit value is set as the set value.
In addition, since the distribution of the mixed gas changes due to the supply pressure of the fuel gas, attention also needs to be paid to the combustion state. In fig. 12, the unmeasured condition can be identified by an approximate line connecting data before and after the measurement point.
As described above, the set Δ P is plotted using the output (load factor) and the rotation speed as parameters, as shown in fig. 13. In the perspective view of fig. 13 with the output, the rotation speed, and the pressure Δ P as parameters, the range indicated by the broken line is a usual (practical) operating region, and the solid line indicates the cubic characteristic for the ship.
In addition, since the fuel gas supply pressure is determined according to the pressure Δ P, when the supply gas pressure changes, the supply pressure of the fuel gas also changes. That is, since the differential pressure between the upstream and downstream sides of the fuel gas supply valve 15 is set, and the amount of the fuel gas is determined by the differential pressure between the front and rear sides of the fuel gas supply valve 15 and the valve opening period, even when the supply pressure changes, the relationship between the supply pressure of the fuel gas and the valve opening period of the fuel gas supply valve 15, which is determined here, is not significantly affected.
As can be seen from fig. 13, if the output (load factor) of the engine 1 becomes large, that is, as the timing of closing the intake valve 8 advances, the supply pressure of the fuel gas including the pressure Δ P also increases.
Next, a method of determining the supply air pressure (air-fuel ratio) in the engine 1 will be described.
In accordance with the change in the VIVT command value (intake valve closing crank angle) of the present embodiment described above, regarding the supply air pressure supplied from the intake pipe 13 to the intake valve 8, the flow rate regulating valves provided with bypass lines on, for example, the compressor side and the turbine side of the supercharger 17 shown in fig. 9 are controlled in accordance with the target supply air pressure set with the output and the rotation speed of the engine 1 measured in advance as parameters to control the supply air pressure (refer to the contents shown in japanese patent application No. 2016 027359, for example). The control of the supply air pressure is not limited to this control method. The known air supply pressure control method can also be adopted.
The following describes a method of determining the intake air pressure (air-fuel ratio).
The fuel amount that is the basis 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 fig. 12 and 13, and an appropriate fuel amount is determined from the predetermined output (load factor) and the rotation speed of the engine 1. The air-fuel ratio is determined according to the ratio of the air amount and the fuel gas amount. Therefore, by changing the intake air pressure, the amount of air supplied to the combustion chamber 6 can be changed. That is, the air-fuel ratio is adjusted by the supply air pressure.
As for the setting method of the air-fuel ratio, the thermal efficiency and NOx data in various air-fuel ratios (supply air pressures) can be measured by changing the operating conditions of the engine 1 using the output (load factor) and the rotation speed as parameters, for example, setting the output (load factor) to 25%, 50%, 75%, 100%, and the like. Fig. 14 shows an example of NOx data when the air-fuel ratio is changed at any output and rotation speed.
In fig. 14, the measured value of the NOx data corresponding to the change in the air-fuel ratio is represented by a curve as "measurement data". From the measured data, it was confirmed that the decrease of the air-fuel ratio (decrease of the supply pressure) increases NOx.
Here, the NOx value varies depending on the application, and the standard value thereof is different. As for the standard values, the upper limit values and the lower limit values for land use are limited by the NOx emission amount shown by the air pollution prevention law or the like according to the IMO NOx limit based on, for example, the revised MARPOL convention VI rule 13 for ships.
The lower limit value of the air-fuel ratio is limited by the upper limit value regulated by the law based on the above-described NOx emission amount. However, if abnormal combustion such as knocking occurs before the NOx data reaches the upper limit 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 may decrease, but misfire or the like may cause the engine 1 to fail to operate stably. Therefore, the upper limit of the air-fuel ratio at which stable operation can be continued should be set to the upper limit value. Thereby determining the range of the settable air-fuel ratio.
Here, the range of the settable air-fuel ratio is determined, and the air-fuel ratio in the middle of the settable range is set as an appropriate value, and the supply air pressure at that time is set. The same measurement is repeated at an arbitrary output and rotation speed within a range that is frequently used. In other words, the supply air pressure (air-fuel ratio) is changed to obtain an appropriate supply air pressure (air-fuel ratio) that satisfies the target performance under the operating conditions of each output and rotation speed.
Fig. 15 is a perspective view depicting the suction pressure set in correspondence with the output and the rotation speed. In the figure, the region indicated by the broken line is an image of a usual (practical) running region, and the cubic characteristic for the ship indicated by the solid line is set in the range. The unmeasured condition can be determined by an approximate line connecting data before and after the measurement point.
As can be seen from fig. 15, if the output of the engine 1 becomes large, i.e., as the timing of closing the intake valve 8 advances, the required supply air pressure also increases.
As described above, the air-fuel ratio control can be performed by the time control of the start and end of the supply of the fuel gas supply valve 15, the supply pressure control of the fuel gas, and the supply pressure control.
In addition, when the timing at which the intake valve 8 is closed is advanced, the air-fuel ratio becomes small at the same supply pressure. Therefore, in the relationship between the output and the intake pressure shown in fig. 16, the proportion of increase in the optimum intake air pressure (the gradient of increase) with an increase in the output increases with the advance in the timing of closing the intake valve 8 as compared with the case where the closing timing of the intake valve 8 is constant.
After that, it is necessary to change the ignition timing of the fuel in correspondence with a change in the VIVT specified value (crank angle of the intake valve closing timing). The ignition timing of the engine 1 is, for example, a timing at which the pilot fuel injection valve 11 injects fuel oil for ignition, and is set based on the value of the ignition timing shown in fig. 17, which is set using data of the output value and the rotational speed of the engine 1 measured in advance as parameters.
The method of determining the ignition timing will be described below.
The operating conditions of the engine 1 are changed using the output (load factor) and the rotational speed as parameters, and thermal efficiency and NOx data at various ignition timings, such as load factors of 25%, 50%, 75%, 100%, are obtained. An example of thermal efficiency and NOx data when the ignition timing is changed is shown in an arbitrary output and rotation speed, as shown in fig. 18. As shown by the solid line in fig. 18, the NOx increases and the thermal efficiency improves when the ignition time is advanced. Therefore, the thermal efficiency and NOx are in a trade-off relationship.
As described above, since NOx has a predetermined standard value, the ignition timing advanced so that the thermal efficiency becomes the highest is set to a preferable timing within the setting range of NOx that satisfies the same standard as the supply air pressure.
When abnormal combustion such as knocking starts before the NOx data reaches a predetermined standard value, the ignition time before the abnormal combustion starts is set to a preferable ignition time. The ignition timing of the pilot fuel injection valve 11 is repeatedly measured by an arbitrary output and rotation speed in a range that is often used. In other words, under each of the operating conditions of "output (load factor) and rotational speed", the ignition timing is changed to obtain the ignition timing that satisfies the target performance. The same adjustment is repeated at an arbitrary rotation speed and output within a frequently used range, thereby determining the ignition timing. The unmeasured condition can be determined by an approximate line connecting data before and after the measurement point.
As can be seen from fig. 17, in the range of the normal operation region indicated by the broken line, the region in which the output is lower than the solid line of the image indicating the ship cubic characteristic indicates the torque shortage region. In the torque shortage region, if the output of the engine 1 becomes large, that is, as the timing of closing the intake valve 8 is advanced, the ignition timing of the micro pilot fuel injection valve 11 of the engine 1 is also advanced.
On the other hand, a region in which the output is higher than the cubic characteristic line for a ship indicated by the solid line indicates a torque excess region. In the excessive torque area, the ignition time takes the broken line of the common operation area as a peak, and the ignition time can reduce the advance angle degree in the range of 30-50% at most relative to the peak value. Wherein, if the torque shortage region and the torque excess region are provided, the ignition timing can be maintained in the advanced state at the same rotation speed as compared with the case where the VIVT command value is not advanced. That is, as the intake valve closing crank angle of the VIVT command value is advanced, the ignition timing is also advanced by the peak on the line of the solid line representing the ship cubic characteristic. The degree of advance is reduced in the excessive torque region, but the fuel injection timing (ignition timing) of the pilot fuel injection valve 11 of the engine 1 is also advanced as a whole compared to the case where the VIVT command value is not advanced.
As described above, according to the control method of the engine 1 and the engine 1 of the embodiment of the present invention, since the unburned fuel gas to be discharged can be reduced in the normal operation region, it is possible to improve the thermal efficiency and obtain the environmental advantages such as reduction of the greenhouse gas.
In addition, since the operable range for the air-fuel ratio becomes narrower as the output of the engine 1 becomes higher, it is possible to obtain an effect of maintaining a stable operation state, particularly, expanding the operation range for the torque excess region having an output higher than the ship-use cubic characteristic.
Further, when the output of the output shaft of the engine 1 is increased, the compression ratio of the mixture gas in the engine 1 can be lowered, and therefore, knocking at the time of load increase can be suppressed and the load increase time can be shortened.
The engine of the present invention is not limited to the dual-fuel engine 1 of the above embodiment, and appropriate modifications, replacements, and the like may be made without departing from the scope of the present invention. Next, modifications of the present invention and the like will be described, and the same or similar parts as those of the components and materials and the like described in the above embodiments will be given the same reference numerals, and the description thereof will be omitted.
The engine of the present invention is not limited to the dual fuel engine 1 that can switch between the diesel mode D using liquid fuel as the main fuel and the gas mode D using gas as the main fuel, and can be applied to a gas fuel engine using gas as fuel.
Further, the present invention is applicable not only to the load raising mode of the marine engine but also to the load raising mode applicable to the vehicle generator, the emergency generator, and the like.
In the above embodiment, the variable intake valve timing mechanism 30 changes both the valve-opening time and the valve-closing time, and the time during which the intake valve 8 is opened does not change, but one or both of the valve-closing time and the valve-opening time of the intake valve 8 may be selected and controlled to be changed.
Industrial applicability
The invention provides a control method of an engine and an engine system capable of suppressing knocking at the time of load increase and shortening the load increase time by using a premixed gas of a gas fuel and air.
Description of the symbols
1 Dual-fuel engine
2 crankshaft
8 air suction valve
9 exhaust valve
13 air suction pipe
14 exhaust pipe
15 fuel gas supply valve
17 pressure booster
20 rotation speed sensor
21 torque sensor
22 control part
24 first diagram
25 second diagram
27 electric idle converter
28 actuator
30 variable suction valve timing mechanism
42 gas regulator
44 fuel gas supply timing unit
45 fuel gas supply valve timing mechanism

Claims (13)

1. A control method of an engine using gas as fuel,
in an operation in which a pilot fuel is injected from a pilot fuel injection valve in a compressed state of a mixture gas and ignited and combusted, control is performed to lower the compression ratio of the mixture gas in a combustion chamber by adjusting an advance angle by which an intake valve closing time is advanced from an intake bottom dead center as an output of the engine increases,
the valve opening time of the fuel gas supply valve is advanced with respect to the opening time of the intake valve in accordance with the change in the advance angle, and the degree of advance of the valve opening time of the fuel gas supply valve is increased as the advance angle of the closing time of the intake valve advances.
2. The control method of an engine according to claim 1, characterized in that the valve closing time of the fuel gas supply valve is determined by: the valve opening period is calculated from a deviation between a target rotation speed and an actual rotation speed, and is set starting from a valve opening time of the fuel gas supply valve.
3. The engine control method according to claim 1 or 2, characterized in that the supply pressure of the fuel gas is increased in accordance with the advance angle of the closing time of the intake valve.
4. The control method of the engine according to claim 1 or 2, characterized in that an ignition time of the engine is advanced along with an advance angle of the intake valve closing time.
5. The control method of an engine according to claim 4,
the degree of advance of the ignition timing is set to a peak on a cubic characteristic line for a ship which is a boundary between an excessive torque region and a deficient torque region and is set using an output and a rotational speed of the engine measured in advance as parameters,
in the excessive torque region, the degree of advance is reduced as compared with the ship cubic characteristic line.
6. The method according to claim 1 or 2, wherein the output of the engine is an output value of an output shaft obtained from a torque measurement value obtained by measuring a torque of the output shaft of the engine by a torque sensor and a rotation speed measurement value obtained by measuring a rotation speed of the output shaft of the engine by a rotation speed sensor.
7. The engine control method according to claim 1 or 2, characterized in that the advance angle of the intake valve closing time is set based on a value of an advance angle set using data of output and rotation speed of a plurality of output shafts measured in advance as parameters.
8. An engine system comprising a gas-fueled four-stroke engine, the engine system comprising:
a control unit that advances a closing timing of an intake valve of the engine by an advance angle and advances a valve opening timing of a fuel gas supply valve with an opening timing of the intake valve as a criterion in accordance with a change in the advance angle when an output of an output shaft of the engine is increased during an operation in which a pilot fuel is injected from a pilot fuel injection valve in a compressed state of a mixed gas and ignited and combusted;
a variable intake valve timing mechanism that changes a time at which the intake valve is closed, in accordance with a closing time of the intake valve provided by the control unit; and
a fuel gas supply valve timing mechanism that advances a valve opening time of the fuel gas supply valve in accordance with a change in an advance angle of the intake valve provided in the control unit,
wherein control is performed to further reduce a compression ratio of a mixture gas of gas and air in the engine by the variable intake valve timing mechanism as an output of the output shaft of the engine increases.
9. The engine system of claim 8, comprising:
a torque sensor that measures a torque of an output shaft of the engine; and
a rotational speed sensor that measures a rotational speed of an output shaft of the engine,
wherein the output of the output shaft is obtained from a torque measurement value measured by the torque sensor and a rotation speed measurement value measured by the rotation speed sensor, and the control unit is provided with a change in closing time of the intake valve.
10. The engine system according to claim 8 or 9,
the valve closing time of the fuel gas supply valve is set by:
a gas supply time calculation unit that calculates a valve opening period from a deviation between a target rotation speed and an actual rotation speed; and
and a gas supply valve control unit that instructs a fuel gas closing time to be executed by the fuel gas supply valve based on the valve opening period.
11. An engine system according to claim 8 or 9, arranged to:
the supply pressure of the fuel gas is increased in accordance with the advance of the closing timing of the intake valve.
12. The engine system according to claim 8 or 9,
the ignition timing of the engine is advanced in accordance with the advance of the closing timing of the intake valve.
13. The engine system according to claim 8 or 9,
the degree of advance of the ignition timing of the engine, which is associated with the advance of the closing timing of the intake valve, is determined by setting a cubic characteristic line for a ship, which is a boundary between an excessive torque region and a deficient torque region and is set using an output value and a rotation speed of the engine measured in advance as parameters, as a peak, and the degree of advance is reduced in the excessive torque region.
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