JPH0658066B2 - Air-fuel ratio controller for internal combustion engine - Google Patents

Description

The present invention relates to a fuel supply amount control device suitable for a two-cycle internal combustion engine. Further, the present invention can be applied to an internal combustion engine having a long valve overlap period and a large amount of fuel blow-through, such as a 4-valve engine, even if it is a 4-cycle internal combustion engine.

[Conventional technology and problems]

In a two-cycle internal combustion engine, the period during which the intake port and the exhaust port communicate with each other for scavenging becomes very long. When fuel is supplied to the cylinder in the form of air-fuel mixture by a vaporizer as usual, it is often discharged as it is to the exhaust system by blow-by. Therefore, there is proposed a system in which a fuel injector is installed and fuel is injected only for a predetermined period in the intake cycle of the engine (Japanese Patent Laid-Open No. 53-27731). However, even if such a fuel injection system is adopted, it is not possible to completely prevent blow-through of fuel. That is, in order to effectively prevent the blow-through of fuel, it is necessary to operate the fuel injection valve at a crank angle after the exhaust valve is closed and before the intake valve is closed. Is not long enough to inject a sufficient amount of fuel. That is, the required amount of fuel can be injected during low load / low speed operation, but the required amount of fuel cannot be injected during high load / high speed operation. Therefore, during high load operation, fuel injection must be started early, that is, while the exhaust valve is open. As a result, fuel blows through the exhaust pipe at high loads. As a result, the amount of fuel becomes insufficient in the transient state, and stable combustion cannot be performed.

An object of the present invention is to enable accurate compensation of fuel blow-through on the high load / high speed side.

[Means for solving problems]

According to the present invention, in FIG. 1, in an internal combustion engine with many blow-throughs such as a two-cycle internal combustion engine, the air-fuel ratio control device includes a fuel supply means 2 for supplying a desired amount of fuel to the internal combustion engine 1, Fuel supply amount calculation means 3 for calculating a fuel supply amount that is determined by operating conditions of the internal combustion engine such as load and rotational speed
And a storage means 4 for storing correction factor data for compensating for the blow-through of fuel according to a plurality of operating conditions of the internal combustion engine, and a storage means for storing the value of the correction factor suitable for the actually measured operating conditions of the internal combustion engine. The fuel supply amount correction means 5 for correcting the fuel supply amount calculated by the fuel supply amount calculation means 3 and the fuel supply means 2 for supplying the corrected amount of fuel to the engine. Means 6 for forming a fuel supply signal to the.

〔Example〕

FIG. 2 shows an overall schematic structure of a 6-cylinder 2-cycle internal combustion engine of the type having an intake valve and an exhaust valve to which the present invention is applied, and FIG. 3 shows one cylinder. As will be described later, this type of two-cycle internal combustion engine causes a stratification action of causing exhaust swirl at the time of reverse flow of exhaust gas after blowdown to concentrate fresh air near the spark plug in the upper part of the combustion chamber, resulting in a light load. This is a device to improve the ignitability during operation. However, the present invention is not limited to this type of two-cycle internal combustion engine, and can be applied to a normal piston valve type two-cycle internal combustion engine. Also,
Even in the case of a 4-cycle internal combustion engine, the concept of the present invention can be applied to the case where the intake air and the fresh air are often blown through because the valve overlap period is long like the 4-valve type. In FIGS. 2 and 3, reference numeral 10 denotes a main body of an internal combustion engine, which includes a cylinder block 12, a cylinder bore 14, a crankshaft 15, a piston 16, a combustion chamber 17, a cylinder head 18, and a spark plug 19. . The cylinder head 18 has two intake ports 20a, 20b and two exhaust ports 22a, 22b, and intake valves 24a, 24b and exhaust valves 26a, 26b for opening and closing the respective intake ports and exhaust ports.
Is a so-called 4-valve type. The intake valve and the exhaust valve are opened and closed by dedicated cams 27 and 28, respectively. 30 and 31 are valve springs. The exhaust ports 22a, 22b are selected in such a shape that swirl of the exhaust gas can be obtained around its vertical axis in the cylinder bore when the exhaust gas flows back into the cylinder bore due to its negative pressure after blowdown. .

In FIG. 2, reference numeral 32 indicates a surge tank, which is connected to a number of intake pipes 33, the number of which corresponds to the number of cylinders. The intake pipe 33 has an internal partition wall 33-1 and has two intake passages 34a and 34b formed therein, which are connected to the intake ports 20a and 20b, respectively. The effective dimension of the second intake passage 34b is the first intake passage 34a.
It is larger and the intake control valve 36 is installed. The intake control valve 36 of each cylinder is connected to an actuator 37 by link means 36 '. The actuator 37 is, for example, a diaphragm mechanism of negative pressure operation, and is switched between negative pressure and atmospheric pressure by a switching valve (not shown), and the intake control valve 36 opens and closes the intake passage 34b. Can be taken selectively. As will be described later, the intake control valve 36 is closed when the load is light and opened when the load is high. The fuel injectors 38a, 38b are connected to the intake passages 34a, 34b.
Is located in. 40a and 40b are reed valves, which are installed as necessary for controlling the reverse flow.

An intercooler 42, a mechanical supercharger 44, a throttle valve 46, an air flow meter 48, and an air cleaner 50 are sequentially arranged in the intake system upstream of the surge tank 32. The mechanical supercharger 44 is constituted by, for example, a roots pump or a vane pump, and a pulley is mounted on its drive shaft 44-1.
44-2 is provided, and the crankshaft 15 is provided by the belt 44-3.
It is connected to the upper pulley 15 '. A bypass control valve 45 is installed in a bypass passage 44 ′ that bypasses the mechanical supercharger 44, and adjusts the pressure between the supercharger 44 and the throttle valve 46. The intercooler 42 is configured as an air-cooled type in this embodiment, and has an inlet container 42-1 and an outlet container 42.
-2, the heat exchange pipe 42-3 communicating between them, and the heat exchange pipe
And a fin 42-4 mounted on 42-3. The exhaust manifold 54 is # 1- # in this embodiment.
Two cylinders are installed separately for each of the cylinder groups up to 3 and the cylinder groups up to # 6. This grouping is such that ignition alternates between these two groups. That is, in this embodiment, the ignition order is #
The order is 1, # 6, # 2, # 4, # 3, # 5. By grouping with alternating ignition, the exhaust pressure of one cylinder in the scavenging stroke can be prevented from being influenced by the exhaust pressure of other cylinders, as will be described later. The exhaust manifolds 54 of the cylinder groups # 1 to # 3 and the cylinder groups # 4 to # 6 are connected to a dedicated catalytic converter (which may also serve as a muffler or a dedicated muffler may be separately installed) 56. .

Reference numeral 58 is a distributor, which is connected to the spark plug 19 of each cylinder as is well known, and is controlled by an igniter and an ignition coil (not shown) so that ignition is performed at a desired crank angle.

The control circuit 60 controls the operation of the injectors 38a and 38b so as to obtain a desired air-fuel ratio according to the present invention, and is configured as a microcomputer system. The control circuit 60 is a microprocessing unit
(MPU) 60-1, memory 60-2, input port 60-3
And output port 60-4 and bus 60-5 connecting them
Composed of and. Each sensor is connected to the input port 60-3 and an operating condition signal is input. The air flow meter 48 may be of a volume flow type and measures the flow rate Q of the intake air passing through the intake pipe. The present invention can also be applied to a fuel injection system in which a pressure sensor that detects the intake pipe pressure is installed instead of the air flow meter. In this case, the anti-conductor type intake pipe pressure sensor 61 is installed downstream of the throttle valve 46 and upstream of the supercharger 44, and generates a signal corresponding to the intake pipe pressure PM. In this embodiment in which the bypass passage 44 'is installed, the pressure sensor 61 is preferably installed upstream of the connection portion of the bypass passage 44' so that the pressure is not affected by the bypass air flow rate. If the supercharging system does not have a bypass, it is possible to install the pressure sensor downstream of the supercharger. The crank angle sensors 62, 64 are installed in the distributor 58. The first crank angle sensor 62 is installed face-to-face with a magnet piece 58-2 fixed on the distributor shaft 58-1, and generates, for example, a pulse signal at every 360 ° (corresponding to one cycle of the engine) at a crank angle, It becomes the reference signal. On the other hand, the second crank angle sensor 64 is installed face-to-face with the magnet piece 58-3 on the distributor shaft 58-1, and for example, the crank angle is 30 °.
A pulse signal is generated every time, the engine speed can be known, and the signal becomes a start signal of the fuel injection routine. A water temperature sensor 68 is installed in the engine body 10 and generates a signal corresponding to the temperature THW of the cooling water in the water jacket 10-1.

The MPU 60-1 executes arithmetic processing according to programs and data stored in the memory 60-2, and executes processing for forming drive signals for the intake control valve actuator 37 and the injectors 38a, 38b. The output port 60-4 is connected to the actuator 37 and the fuel injectors 38a and 38b of each cylinder, and a drive signal is applied.

FIG. 4 shows intake valves 24a, 24b and exhaust valve 26 in one cylinder determined by the profiles and orientations of cams 27 and 28.
The operation timings of a and 26b are shown. First, the exhaust valves 26a and 26b start to open at 80 ° before the bottom dead center (BDC),
Closes at 40 ° after bottom dead center (BDC). On the other hand, intake valves 24a, 2
4b starts to open at 60 ° before bottom dead center (BDC), bottom dead center (BDC)
After that, it closes at 60 °. I indicates a fuel injection period, which is set to end immediately before closing the intake valves 24a and 24b. The start of the fuel injection period I changes depending on the magnitude of the fuel injection amount, and naturally becomes earlier as the combustion injection amount increases. FIG. 5 is a timing chart showing the period during which the exhaust valve is operating in each cylinder with respect to the crank angle. Since it is a two-cycle engine, one cycle is completed at 360 ° CA, and the exhaust valve is opened every 60 ° crank angle for the period EX shown in FIG. 3 according to the ignition sequence. One group (# 1 to ##) where cylinders are placed every other ignition order.
3 or # 4 to # 6), the exhaust valves are opened every 120 °, and in each group, the exhaust valve open periods do not overlap with each other between adjacent cylinders in the ignition order. This prevents the exhaust pressure of one cylinder from affecting the exhaust pressure of the next ignited cylinder in the group. That is, the exhaust pressure pulsates due to the effect of blowdown, but when this pulsation is transmitted to other cylinders, the pressure changes in such a way that it cannot be predicted, and as a result, the predictability of the amount of fresh air and fuel blow-through is lost. This prevents the air-fuel ratio from being accurately compensated according to the amount of blow-through, which is prevented. On the other hand, if the two groups are included, the exhaust valve opening period overlaps between the cylinders adjacent to each other in the ignition sequence, but the exhaust manifold 54 is separate between these cylinders.
The exhaust pressure of one cylinder does not affect the exhaust pressure of the other cylinder.

First, the combustion operation of a two-sided internal combustion engine of the type having an intake valve and an exhaust valve to which the present invention is applied will be described. When the engine is lightly loaded, the intake control valve 36 is closed, and the intake air is introduced into the engine only through the first intake passage 34a. In the process of lowering the piston 16, first the bottom dead center
Exhaust valves 26a and 26b start to open around 80 ° before (BDC).
Therefore, the exhaust gas is discharged from the combustion chamber to the exhaust ports 22a and 22b as shown by the arrow P in FIG. 6 (a), and so-called blowdown occurs, but this blowdown is weak and ends immediately. , 22b are not affected by the exhaust pressure of the cylinders to be ignited next, because they belong to a group other than the exhaust manifold 54. Then, when the piston 16 further descends, the inside of the cylinder bore 14 becomes weak, but a negative pressure is generated.
Exhaust gas flows backward toward the cylinder bore as indicated by arrow Q due to the pressure difference between a and 22b (Fig. 6 (b)). Then, due to the shape of the exhaust ports 26a and 26b, a swirl flow (swirl) of exhaust gas is formed in the cylinder bore as indicated by arrow R. Around this time, the intake valves 24a (24b) start to open, but the lift is still small, the throttle valve 46 is throttled, the intake control valve 36 is closed, and the intake passage 34b having a large effective dimension is closed. Due to the fact that air can flow only through the intake passage 34a having a small effective size, the introduction of fresh air does not substantially occur. When the piston 16 descends further, the swirl of the exhaust gas continues, while the intake valve 24
Since the lifts of a and 24b become large, fresh air is introduced into the cylinder bore as indicated by arrow S, at which time exhaust gas rides on the swirl and moves to the lower part of the cylinder bore 14, while fresh air mixed with the injected fuel swirl. The stratification in which the exhaust gas R gathers in the vicinity of the spark plug electrode above the exhaust gas portion (FIG. 6 (c) is achieved. In such a stratified state of the exhaust gas R and the fresh air S, the piston is at the bottom dead center. It is maintained even after reaching (BDC) (Fig. 6 (d)). In (e), the intake valves 24a and 24b are closed to prevent blowback of fresh air, and then the piston moves upward. The stratified state is maintained until the compression is completed, and the fresh air portion near the spark plug can be easily ignited.

In the high load state of the engine, the intake control valve 36 is opened.
Therefore, the intake passage 34b which has been closed until now is opened. In FIG. 7, when the exhaust valves 26a and 26b are first opened in the process of lowering the piston 16, the exhaust gas in the cylinder bore 14 is discharged to the exhaust ports 22a and 22b by the blowdown P. In comparison, it is strong and has a long duration (Fig. 7 (a)), and a large amount of exhaust gas is discharged to the exhaust port. Intake valve 2 at the time of Fig. 7 (b)
4a and 24b start to open, but this time the intake control valve 36 is open, the opening of the throttle valve 46 is large, and the supercharger 44 is performing sufficient supercharging operation. Is performed as indicated by arrow T. At this time, intake port 20
Fresh air is introduced from both a and 20b, and this fresh air is indicated by the arrow T.
As described above, the exhaust gas flows from the upper side to the lower side along the wall surface of the cylinder bore, and the exhaust gas is caused to flow out to the exhaust ports 22a and 22b as indicated by an arrow U, so that a so-called transverse scavenging is realized. At the time of Fig. 7 (c), a negative pressure component appears in the pressure wave pulse based on the strong blowdown, and the exhaust ports 22a and 22b temporarily become negative pressure, and as a result, the introduction of the fresh air T into the cylinder bore is further promoted. Some of the fresh air is exhaust port 22a,
It is once discharged to 22b and stored. This stored fresh air is
When the pressure in the exhaust ports 22a, 22b returns to positive pressure, the arrow W
As shown in Fig. 7, the air flows back into the cylinder bore to generate fresh air swirl X (Fig. 7 (d)). As a result, turbulence occurs and flame spread after ignition is improved. The intake valves 24a and 24b are completely closed at the time of FIG. 7 (e), and blowback of fresh air is prevented.

Next, the operation of the control circuit 60 for performing the operation of the intake control valve 36 in the above-described combustion operation will be described with reference to the flowchart of FIG. This routine can be executed at regular intervals. In step 100 the flag FTVIS
It is determined whether or not = 1. When FTVIS = 0, step 102
Next, it is determined whether or not the intake air amount-rotation speed ratio Q / NE is larger than a predetermined value (Q / NE) ₀ , and in step 104, the rotation speed NE
Is determined to be greater than a predetermined value (NE) ₀ . When the intake air amount-rotation speed ratio Q / NE> predetermined value (Q / NE) ₀ or the rotation speed NE> predetermined value (NE) ₀ , the routine proceeds to step 106, where the output port 60-
4 outputs a signal to the actuator 37 to open the intake control valve 36. In step 108, the flag FTVIS =
Set to 1. When FTVIS = 1, the routine proceeds to step 110, where it is judged if the intake air amount-rotation speed ratio Q / NE is smaller than a predetermined value (Q / NE) ₁ , and at step 112 the rotation speed NE is a predetermined value (NE) ₁ It is determined whether or not it is smaller. Intake air volume −
When the rotation speed ratio Q / NE <predetermined value (Q / NE) ₁ and the rotation speed NE <predetermined value (NE) ₁ , the process proceeds to step 114, and the output port 60-4
From the actuator 37, a signal for closing the intake control valve 36 is output. In step 116, flag FTVIS = 0
Is set.

Next, the fuel injection control of the present invention will be described. Similar to a normal fuel injection control device in a 4-cycle engine, in the present invention, in principle, the intake air amount is measured, and an amount of fuel is injected according to the measured value to obtain an initial air-fuel ratio. To do. However, although there is a similar problem in a normal two-cycle internal combustion engine of a piston valve, there are many problems of fresh air and fuel blow-through because the exhaust valve and the intake valve are held open at the same time for a long time. And
The proportion of fresh air that blows through changes depending on the load, the number of revolutions, and other operating conditions. Therefore, the blow-through of fresh air can be compensated by algebraic function approximation or the like as in JP-A-53-27731, and more precisely, it can be compensated by a map as proposed in the embodiment of the present invention. However, not only fresh air but also fuel blow-through occurs,
This blow-through amount changes depending on the load and the rotation speed. That is, the operating period I (FIG. 4) of the fuel injection valve is set as long as possible after the exhaust valve is closed and the intake valve is open. Since the fuel cannot be injected, the fuel injection amount inevitably increases during the period when the exhaust valve is open, and the fuel blow-through amount inevitably increases, and the blow-through amount changes depending on the operating conditions of the engine. Therefore, in the present invention, the coefficient f _FTR (that is, the ratio of the blow-through combustion amount to the total injection amount) representing the fuel blow-through amount is introduced, and the basic fuel injection amount is corrected by this, regardless of the fuel blow-through. It is possible, and more precisely, it is designed to obtain the desired air-fuel ratio. FIG. 9 shows the fuel injection routine.
This routine is 30 ° from the second crank angle sensor 64.
This is a crank angle interrupt routine that is executed each time a CA signal arrives. In step 130, it is determined whether or not it is the fuel injection calculation timing. As shown in FIG. 3, the fuel injection is performed within a predetermined angle range after the intake valves 24a and 24b start to open, so that this calculation is executed at a predetermined crank angle slightly preceding this. This timing can be known by the value of the counter which is cleared by the 360 ° CA signal from the first crank angle sensor 62 and incremented by the 30 ° CA signal from the second crank angle sensor 64. If the fuel injection calculation timing is determined, the routine proceeds to step 132, where the basic fuel injection amount T _p is calculated by T _p = k (Q ′ / NE). Here, Q'is the intake air amount Q converted into mass, which is a value after the measured value of the air flow meter 48 is corrected by the intake air temperature or the like. (Intake pipe pressure P
In a system in which the fuel injection amount is known by M, PM can be used instead of Q '/ NE. ) In step 134, the map calculation of the fresh air acquisition coefficient f _TR is executed. Here, the fresh air trapping coefficient f _TR is the amount of fresh air that has flown into the exhaust system due to blow-through minus the amount of fresh air that has been measured by the air flow meter 48. It is a correction factor for the fuel injection amount related to the ratio. FIG. 10 conceptually shows how the fresh air trapping coefficient f _TR changes with respect to the intake air amount-rotational speed ratio and the rotational speed. It can be seen that the intake air amount-rotation speed ratio and the rotation speed change intricately due to the influence of pressure pulsation of the exhaust pipe due to blowdown (the case where there is no influence due to blowdown is shown by the broken line). When the intake control valve 36 is switched between open and closed, the fresh air acquisition coefficient f _TR
It can also be seen that changes discontinuously (see the chain double-dashed line).
According to FIG. 10, the memory 60-2 stores the fresh air acquisition coefficient f _TR for the combination of the intake air amount-rotational speed ratio and the rotational speed.
The data of is stored. Then, the interpolation calculation is executed based on the measured intake air amount-rotation speed ratio and the rotation speed, and the fresh air trapping coefficient f _TR suitable for the current operating condition is calculated. In a system that knows the fuel injection amount from the intake pipe pressure PM, a map of the fresh air trapping coefficient f _TR is created by combining PM and the rotation speed, and interpolation calculation is executed from the intake pipe pressure measured by the pressure sensor 61. It

In step 136, the fuel blow-through coefficient f _FTR is calculated. The blow-through amount of fuel increases as the rotation speed and load increase. This is because the fuel injection valve operates at an early stage when the intake valve and the exhaust valve overlap each other as the engine speed and load increase.
This blow-through coefficient is the amount of intake air as a load on the memory −
There are maps for speed ratio and speed. From this map, the value of the blow-through coefficient f _FTR at the intake air amount-rotational speed ratio and the value of the rotational speed is calculated by interpolation. Instead of constructing a map of the intake air amount-rotational speed ratio and the rotational speed, the blow-through coefficient f _FTR can be calculated from the map with the basic fuel injection amount T _p as in the case of FIG. That is, since the basic fuel injection amount T _p changes depending on the intake air amount-rotation speed ratio and the rotation speed, the basic fuel injection amount is an appropriate index of the blow-through coefficient f _FTR that changes according to operating conditions. In step 138, the final injection amount TAU is calculated by TAU = f _TR × (1 / f _FTR ) × T _p × α × β. That is, by multiplying T _p by 1 / f _FTR , the injection amount is increased by the amount of the fuel that has blown through, and the amount of fuel required by the engine can be constantly supplied.
Note that α and β are representatively shown as correction coefficients and correction amounts, which are not directly related to the present invention and whose description is omitted.

In step 140, it is determined whether or not the flag FTVIS = 1, that is, whether the intake control valve 36 is open or closed. When the intake control valve 36 is open, the routine proceeds to step 142,
Address TAU for storing the fuel injection time of the first fuel injector 38a is stored in address TAUa, and address TAU for storing the fuel injection time of the second fuel injector 38b.
Zero is put in b. That is, only the first injector 38a is activated and the second injector 38b is not activated.
When the intake control valve 36 is closed in step 140, the routine proceeds to step 144, where 1/3 of TAU is put in the address TAUa storing the fuel injection time of the first fuel injector 38a and the second fuel injector 38b is stored. The remaining 2/3 of TAU is put in the address TAUb storing the fuel injection time. Here, 1/3 and 2/3 have no specific meaning, and they are adaptation constants. Since the effective dimension of the second intake passage 34b> the effective dimension of the first intake passage 34a, the air-fuel ratio can be either Since it is constant, it merely indicates that the fuel injection amount from the second injector 38b is larger than the fuel injection amount from the first injector 38a. In step 146, the fuel injection signal forming process is performed so that the injectors 38a, 38b are operated only for a period corresponding to TAUa, TAUb from the desired injection start timing. Since this process itself is well known, detailed description thereof will be omitted. Step 148 generally shows another process that is started every time the 30 ° CA signal arrives.

〔The invention's effect〕

According to the present invention, in an internal combustion engine such as a two-cycle internal combustion engine that has many blow-throughs, the fuel blow-through coefficient is calculated according to the operating conditions of the engine, and the fuel injection amount is increased according to the calculated blow-through coefficient. The fuel can be supplied, and the drivability (acceleration performance) can be improved especially in a transient state. In addition, the fuel consumption rate is improved because the fuel increase is necessary and sufficient.

[Brief description of drawings]

FIG. 1 is a diagram showing the configuration of the present invention. FIG. 2 is a schematic view of the entire system of an embodiment of the present invention. FIG. 3 is a view showing a cross section of one cylinder (III-I in FIG. 2).
Diagram along line II). FIG. 4 is an operation timing angle diagram of an intake valve and an exhaust valve of one cylinder in one cycle of the engine. FIG. 5 is a diagram showing the operation timing of the exhaust valve of each cylinder in one cycle of the engine. FIG. 6 is a diagram for explaining a combustion operation in one cycle of a two-cycle internal combustion engine with an intake valve and an exhaust valve according to an embodiment of the present invention at a light load. FIG. 7 is a diagram for explaining combustion operation in one cycle of a two-cycle internal combustion engine with an intake valve and an exhaust valve according to an embodiment of the present invention at the time of high load. 8 and 9 are flow charts for explaining the operation of the control circuit. FIG. 10 is a conceptual diagram of a change in the fresh air trapping coefficient f _TR with respect to the intake air amount-rotation speed ratio and the rotation speed. FIG. 11 is a graph showing the relationship between the basic injection amount and the fuel blow-through coefficient. 10 ... Engine body, 17 ... Combustion chamber, 24a, 24b ... Intake valve, 26a, 26b ... Exhaust valve, 34a, 34b ... Intake passage, 36 ... Intake control valve, 38a, 38b ... Fuel injector, 42 ... Intercooler, 44 ... Mechanical supercharger, 48 ... Air flow meter, 54 ... Exhaust manifold, 60 ... Control circuit, 62, 64 ... Crank angle sensor.

Claims (1)

[Claims] 1. An air-fuel ratio control device comprising the following components in an internal combustion engine such as a two-cycle internal combustion engine with many blow-throughs, fuel supply means for supplying a desired amount of fuel to the internal combustion engine, load of the internal combustion engine, , A fuel supply amount calculation means for calculating a fuel supply amount determined by a basic operating condition such as a rotational speed, and a storage means for storing correction factor data for compensating the blow-through of fuel according to a plurality of operating conditions of the internal combustion engine , A fuel supply amount correction means for correcting the fuel supply amount calculated by the fuel supply amount calculation means by calculating a correction factor value suitable for the actually measured operating condition of the internal combustion engine from the data stored in the storage means, and corrected Means for activating the fuel supply signal to the fuel supply means such that a later amount of fuel is supplied to the engine.