JP2002221037A - Cylinder injection type gas fuel internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a cylinder injection type gas fuel internal combustion engine capable of maintaining a stable combustion condition even if injection pressure varies. SOLUTION: In an internal combustion engine directly injecting gas fuel in cylinders, a means 78 detecting pressure of gas fuel to be injected, and a means controlling opening of swirl control valve 38 according to a detected pressure are provided. A means controlling opening of EGR valve 68 according to the detected pressure is provided. Moreover, a means controlling ignition timing according to the detected pressure is provided.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to compressed natural gas (C)
The present invention relates to an internal combustion engine that directly injects gas fuel such as NG (Compressed Natural Gas) into a cylinder.

[0002]

2. Description of the Related Art In recent years, there has been a movement to use a gas fuel such as compressed natural gas (CNG) as a fuel for an internal combustion engine for a vehicle. In such a gas-fueled engine, there is an issue of increasing the cruising distance, and therefore, a technique for injecting fuel into a cylinder to realize stratified combustion has been developed. In stratified combustion, incomplete combustion is performed by forming a rich mixture and a lean mixture in a stratified form in the combustion chamber, first igniting the rich mixture, and then burning the lean mixture with the flame. In addition, a lean air-fuel mixture is burned as a whole while avoiding misfires to improve the fuel consumption rate (for example, Japanese Patent Application Laid-Open No. Hei 7-166873).
JP-A-8-42381, etc.).

[0003]

In such a direct injection gas fuel engine, fuel is supplied from a fuel cylinder to an injector via a regulator or directly, but the fuel injection pressure changes, and the Therefore, there is a problem that the combustion state changes. That is, in a system that does not have a regulator and performs fuel injection by cylinder pressure (residual fuel pressure),
The fuel residual pressure, that is, the fuel injection pressure changes with the consumption of fuel. In a system that regulates pressure using a regulator, the fuel injection pressure changes when the regulator characteristics change due to a failure, a change over time, or the like, or when the regulator inlet pressure (residual fuel pressure) becomes equal to or lower than the regulator outlet pressure.

[0004] In a gas fuel engine, the spray speed is very high and the kinetic energy is large, so that the effect of the spray on the formation of flow and turbulence in the cylinder is large. Therefore, when the injection pressure changes and the spray speed changes, the in-cylinder flow and the mixture formation change. As a result, the combustion state changes, and the emission and operability deteriorate.

Another problem is that the optimum ignition timing changes with a change in the fuel injection pressure. The first reason is that the in-cylinder turbulence changes due to the change in the spray speed as described above. The second reason is that, during injection, the spray temperature decreases due to decompression and expansion, but the change in injection pressure changes the temperature reduction rate, and as a result, the charging efficiency and knock margin change. . The third reason is that the fuel injection in the compression stroke (specifically, from after the intake valve is closed to the ignition timing) improves the charging efficiency as compared with the case of the intake stroke injection, but when the injection pressure decreases,
The fuel injection time is prolonged, and it is necessary to inject not only in the compression stroke but also in the intake stroke, and the charging efficiency is reduced.

The present invention has been made in view of the above-mentioned problems, and an object of the present invention is to provide a direct injection gas-fueled internal combustion engine capable of maintaining a stable combustion state even when the injection pressure changes. To provide.

[0007]

According to a first aspect of the present invention, there is provided an internal combustion engine for directly injecting gas fuel into a cylinder, the pressure of the injected gas fuel being detected. And a means for controlling the degree of opening of the swirl control valve in accordance with the detected pressure.

According to a second aspect of the present invention, in an internal combustion engine for injecting gas fuel directly into a cylinder, means for detecting the pressure of the injected gas fuel, and EGR in accordance with the detected pressure Means for controlling the degree of opening of the valve.

According to a third aspect of the present invention, in an internal combustion engine for directly injecting gas fuel into a cylinder, means for detecting the pressure of the injected gas fuel, and ignition in accordance with the detected pressure A gas-fueled internal combustion engine provided with means for controlling timing.

According to the present invention, preferably, there is further provided means for reducing the flow velocity when the injected gas fuel collides with the spark plug.

According to the present invention, preferably, there is further provided means for controlling the injection timing so that the gas fuel injection is started after the intake valve is closed.

According to the present invention, preferably, there is further provided means for controlling the injection timing so that the gas fuel is injected during the compression stroke at the time of starting the engine.

[0013]

Embodiments of the present invention will be described below with reference to the accompanying drawings.

FIG. 1 is an overall schematic diagram of a direct injection type CNG engine as one embodiment of the present invention. This CNG engine 1 includes a cylinder block 2 and a cylinder head 3. A plurality of cylinders 4 extending in the up-down direction are arranged in the cylinder block 2 in parallel in the thickness direction of the drawing, and a piston 5 is accommodated in each cylinder 4 so as to be able to reciprocate. Each piston 5 is connected to a common crankshaft 7 via a connecting rod 6. The reciprocating motion of each piston 5 is converted into a rotational motion of a crankshaft 7 via a connecting rod 6.

A combustion chamber 8 is provided between the cylinder block 2 and the cylinder head 3 above each piston 5. The cylinder head 3 is provided with an intake port 9 and an exhaust port 10 for communicating both outer surfaces thereof with the respective combustion chambers 8. In order to open and close these ports 9 and 10, an intake valve 11 and an exhaust valve 12 are supported on the cylinder head 3 so as to be able to reciprocate substantially vertically. In the cylinder head 3,
Above each valve 11, 12, an intake side camshaft 1 is provided.
3 and the exhaust-side camshaft 14 are provided rotatably. Cams 15 and 16 for driving the valves 11 and 12 are attached to the camshafts 13 and 14, respectively. Timing pulleys 17 and 18 provided at ends of the camshafts 13 and 14 are connected to a timing pulley 19 provided at an end of the crankshaft 7 by a timing belt 20.

The intake port 9 is connected to an intake passage 30 having an air cleaner 31, a throttle valve 32, a surge tank 33, an intake manifold 34 and the like. Air (outside air) outside the engine 1 is supplied to the intake passage 3 toward the combustion chamber 8.
0 sequentially passes through the respective parts 31, 32, 33 and 34. The throttle valve 32 in the present embodiment is a so-called electronic throttle, which is not mechanically directly connected to the driver's seat accelerator pedal.
Driven by

The cylinder head 3 is provided with each combustion chamber 8.
An injector 40 for injecting fuel toward is mounted. Compressed natural gas (CNG), which is a fuel, is stored in a fuel cylinder 41, adjusted to a constant pressure by a regulator 44, and then supplied to an injector 40.
The fuel injected from the injector 40 is supplied to the intake passage 3
0, the combustion chamber 8 via the intake port 9 and the intake valve 11
And the air introduced into the combustion chamber 8 to form a mixture.

To ignite the mixture, an ignition plug 50 is attached to the cylinder head 3. Each cylinder is provided with an ignition coil 52 with a built-in igniter that is independently coupled to the ignition plug 50 for each cylinder. At the time of ignition, the igniter controls the supply and cutoff of the primary current in the igniter built-in ignition coil 52 for each cylinder receiving the ignition signal, and the secondary current is supplied to the ignition plug 50.

The burned air-fuel mixture is guided to an exhaust port 10 via an exhaust valve 12 as exhaust gas. The exhaust port 10 has an exhaust manifold 61 and a catalytic converter 62.
The exhaust passage 60 provided with the above is connected. HC (hydrocarbon), which is an incomplete combustion component, is supplied to the catalytic converter 62.
And a three-way catalyst that simultaneously promotes the oxidation of CO (carbon monoxide) and the reduction of NOx (nitrogen oxide) generated by the reaction between nitrogen in the air and residual oxygen. Alternatively, lean NOx is mounted in consideration of stratified combustion. The exhaust gas thus purified in the catalytic converter 62 is discharged into the atmosphere.

In the first embodiment of the present invention, a swirl control valve (SCV) 38 for generating a swirl (swirl) of air in an intake port around the intake valve 11 and sending the swirl into the combustion chamber 8 is provided. It is assumed that it is provided. Swirl control valve 3
8 is controlled via a swirl control actuator 39.

In the second embodiment of the present invention, it is assumed that the engine 1 performs EGR (exhaust gas recirculation) for the purpose of reducing NOx. That is, a passage 64 for circulating exhaust gas is provided between the exhaust passage 60 and the intake passage 30 on the downstream side of the throttle valve 32, and the EGR gas flow rate is controlled by the EGR gas provided in the middle of the passage 64. Controlled by valve 68.

The engine 1 is provided with various sensors. A water temperature sensor 74 for detecting the temperature of the cooling water of the engine 1 is attached to the cylinder block 2. In the intake passage 30, near the throttle valve 32, a throttle opening sensor 72 for detecting a rotation angle of the shaft and an accelerator opening sensor for detecting an accelerator operation amount (also referred to as an accelerator depression amount, an accelerator opening, etc.). 77 are provided. The surge tank 33 is provided with a vacuum sensor 71 for detecting the internal pressure (intake pipe negative pressure). Further, an injection pressure sensor 78 for detecting a fuel injection pressure is provided in a fuel pipe connecting the regulator 44 and the injector 40.

The camshaft 13 is provided with a crank reference position sensor 80 for generating a reference position detection pulse every 720 ° CA in terms of a crank angle (CA). The crankshaft 7 has a temperature of 30 ° C.
A crank angle sensor 81 for generating a rotation speed detection pulse for each A is provided. In the electronic control unit 90, every time a pulse signal from the crank angle sensor 81 is input, the engine rotation speed is calculated from the pulse interval and stored in an internal memory as engine rotation speed data NE.

An electronic control unit (ECU) 90 is a microcomputer system for executing fuel injection control, ignition timing control, swirl control valve control, EGR valve control, etc., according to programs stored in a memory and various maps. Signals from various sensors are input, arithmetic processing is executed based on the input signals, and control signals for various actuators are output based on the arithmetic results.

As described above, there is a problem that the combustion state changes when the fuel injection pressure changes. Therefore, in the first embodiment of the present invention, by controlling the opening of the swirl control valve (SCV) 38 in accordance with the change in the fuel injection pressure, the intake air necessary for forming the optimum mixture and the in-cylinder flow is formed. I always give the flow.

FIG. 2 is a flowchart for calculating the swirl control valve opening. First, at step 102, the engine speed NE and the accelerator opening ACP as the engine load are detected. Next, in step 104, the swirl control valve opening SCVA according to the NE and the ACP is obtained by referring to a predetermined map (not shown).

At step 106, the current fuel injection pressure P is detected based on the output of the injection pressure sensor 78. Next, at step 108, the detected injection pressure P
ΔP between the reference value and the injection pressure reference value P _REF is calculated. In the next step 110, the absolute value of the calculated deviation | ΔP |
Is determined to be greater than or equal to a predetermined threshold value Pα. If not, the process ends because there is no need for correction.

On the other hand, when | ΔP | exceeds the threshold value Pα, the swirl control valve opening correction amount ΔSCVA corresponding to the deviation ΔP is obtained in step 112 by referring to a map shown in FIG. . Next, at step 114, the correction amount ΔS is added to the previously determined swirl control valve opening SCVA.
CVA is added to determine the final swirl control valve opening SCVA.

According to the swirl control valve opening correction amount ΔSCVA shown in FIG. 3, when the injection pressure P decreases and the formation of air-fuel mixture and the formation of turbulence in the cylinder due to fuel injection are weakened, the swirl control valve is closed. As a result, a large swirl is generated, and the decrease in the injection pressure is compensated. When the injection pressure P increases, the reverse is true.

The above-described swirl control valve control shown in FIG. 2 can be applied to a configuration without the regulator 44. However, in that case, the range of the deviation ΔP covered by the map of FIG. 3 is expanded.

As described above, in the first embodiment of the present invention,
Although the opening degree of the swirl control valve 38 is controlled according to the change of the fuel injection pressure, in the second aspect of the present invention, E
By controlling the degree of opening of the GR valve 68, an EGR amount necessary for forming an optimal air-fuel mixture is provided.
This is because the EGR gas also contributes to the formation and combustion rate of the air-fuel mixture in the cylinder.

FIG. 4 is a flowchart for calculating the EGR valve opening. First, in step 202, the engine speed NE and the accelerator opening ACP as the engine load are detected. Next, in step 204,
The EGR valve opening EGRVA according to the NE and the ACP is obtained by referring to a predetermined map (not shown).

In step 206, the current fuel injection pressure P is detected based on the output of the injection pressure sensor 78. Next, at step 208, the detected injection pressure P
ΔP between the reference value and the injection pressure reference value P _REF is calculated. In the next step 210, the absolute value of the calculated deviation | ΔP |
Is determined to be greater than or equal to a predetermined threshold value Pα. If not, the process ends because there is no need for correction.

On the other hand, when | ΔP | exceeds the threshold value Pα, the EGR corresponding to the deviation ΔP is determined in step 212 by referring to a map as shown in FIG.
The valve opening correction amount ΔEGRVA is obtained. Then
In step 214, the correction amount ΔEGRVA is added to the previously obtained EGR valve opening EGRVA to determine the final EGR valve opening EGRVA.

The EGR valve opening correction amount Δ shown in FIG.
According to EGRVA, when the injection pressure P decreases and the turbulence of the air-fuel mixture due to fuel injection weakens, the EGR valve is closed and the combustion is promoted. As a result,
The reduction in injection pressure is compensated. When the injection pressure P increases, the reverse is true.

The above-described EGR valve control shown in FIG. 4 can be applied to a configuration without the regulator 44. However, in that case, the range of the deviation ΔP covered by the map of FIG. 5 is expanded.

As described above, the CNG engine 1 also has a problem that the optimum ignition timing changes with a change in the fuel injection pressure. Therefore, in a third aspect of the present invention, the ignition timing is controlled according to the fuel injection pressure. That is, from the fuel injection pressure, each change of the mixture turbulence (affected by the spray kinetic energy), the charging efficiency (affected by the spray temperature and the injection time), and the mixture temperature (affected by the spray temperature) are estimated Then, the optimum ignition timing is calculated based on the calculated values. As a result, the output is improved.

FIG. 6 is a flowchart for specifically calculating the ignition timing (ignition advance value) SA. First,
In step 302, the engine speed NE and the accelerator opening ACP as the engine load are detected. Next, at step 304, the basic ignition timing S according to the NE and the ACP is referred by referring to a predetermined map (not shown).
A is required.

In step 306, the current fuel injection pressure P is detected based on the output of the injection pressure sensor 78. Next, at step 308, the detected injection pressure P
ΔP between the reference value and the injection pressure reference value P _REF is calculated. In the next step 310, the absolute value | ΔP |
Is determined to be greater than or equal to a predetermined threshold value Pα. If not, the process ends because there is no need for correction.

On the other hand, when | ΔP | exceeds the threshold value Pα, first, in step 312, the map shown in FIG. 7 is referenced to determine the amount of change in the spray temperature from the reference value based on the deviation ΔP. ΔT is determined. Next, at step 314, an ignition timing correction amount ΔSAT based on ΔT is calculated by referring to a map (not shown). Note that the relationship between ΔT and ΔSAT is experimentally obtained in advance.

In step 316, the amount of change ΔE in the kinetic energy of the spray is determined based on the deviation ΔP by referring to a map as shown in FIG. Next, at step 318, the ignition timing correction amount ΔSAE based on ΔE is calculated by referring to a map (not shown). Note that the relationship between ΔE and ΔSAE is experimentally obtained in advance.

Further, in step 320, by referring to a map (not shown), a change amount ΔηV of the charging efficiency due to a change in the fuel injection time is obtained based on the deviation ΔP. Next, at step 322, the ignition timing correction amount ΔSAηV based on ΔηV is calculated by referring to a map (not shown). Note that ΔηV and ΔSA
The relationship with ηV is obtained experimentally in advance.

Next, at step 324, the correction amounts .DELTA.SAT, .DELTA.SAE and .DELTA.SA.eta.V are added to the previously obtained ignition timing SA to determine the final ignition timing SA.

In addition to the control of the swirl control valve opening, the control of the EGR valve opening and the control of the ignition timing in consideration of the influence of the fuel injection pressure described above, it is preferable to further include the following configuration.

In the cylinder injection type gas fuel engine, since the fuel is gas and there are no liquid droplets, the spray guide combustion which directly collides the fuel spray with the spark plug is realized which is difficult in the cylinder injection type gasoline engine. Could be done. However, if fuel is directly sprayed on the spark plug, the combustion deteriorates. This is because the speed of the injected spray is very high, and the turbulence of the air-fuel mixture exceeds the combustion propagation limit, and the flame cannot be propagated. Therefore, it is preferable to provide a means for reducing the flow velocity when the injected gas fuel collides with the ignition plug.

Therefore, for example, in the embodiment shown in FIG. 9, the speed of the fuel injected toward the spark plug 50 is reduced by an obstruction plate (jamming plate) 54 attached to the spark plug 50, and The mixture flow near the plug 50 is reduced.

Further, in the embodiment shown in FIG. 10, a strong air current (reverse tumble flow) is generated in a direction opposite to the direction of the spray, thereby canceling the speed of the spray and mixing near the spark plug 50. The air flow rate can be reduced. The reverse tumble flow is realized by the shape of the intake port.

In the embodiment shown in FIG. 11, the tip of the injector 40 (the downstream side of the seal portion 40a)
A decompression chamber 40b having an injection hole (pore) 40c is provided. The fuel injected from the seal portion 40a is decompressed and decelerated in the downstream decompression chamber 40b. Further, the decompression chamber 40b
The fuel is injected while being mixed with the air filled with water, so that the kinetic energy of the spray is absorbed by the air and the speed of the air-fuel mixture injected from the injection holes (pores) 40c decreases.

Now, in a gas fuel engine such as a CNG engine,
Since the fuel is gaseous, its charging efficiency is low, a 10% reduction compared to gasoline engines. Therefore, its output is lower than that of a gasoline engine. Therefore, it is preferable to inject the fuel in the compression stroke, so that the fuel is always injected after the cylinder is filled with the intake air, and high filling efficiency is ensured under all engine conditions. In particular, when the intake valve 11 is configured by a variable valve timing (VVT) mechanism, the fuel injection timing is controlled according to the change in the valve timing, and the injection start timing is set after the intake valve is closed.

FIG. 12 is a flowchart for calculating the fuel injection timing. First, in step 402,
It is determined whether or not the load is high. If the load is not high, the process proceeds to step 404, where a normal injection start timing calculation based on the engine operating state is performed. On the other hand, when the load is high, first, at step 406, the current intake valve closing timing (closing angle) INVC is read. Next, at step 408, a margin Aα corresponding to the engine speed NE is set.
Is calculated, and the injection start timing A is calculated from Aα and INVC.
INJ is calculated.

By the way, in a gas fuel engine such as a CNG engine, the charging efficiency is low as described above, and therefore the output is low. Therefore, at the time of a low temperature start, the friction torque cannot be overcome, and as shown in FIG. 13, the rotation speed of the CNG engine increases slowly as compared with the gasoline engine. Therefore, when starting the engine, it is preferable to perform in-cylinder fuel injection during the compression stroke, and by doing so,
Even at a low temperature start, a torque higher than the friction torque can be output, and a smooth start can be ensured.

FIG. 14 is a flowchart for calculating the fuel injection timing at the time of starting the engine. First, at step 502, the rotation speed reference value Nα is determined based on the current engine cooling water temperature THW by referring to the map shown in FIG. In this map, a rotation speed reference value Nα for determining whether or not the engine is started based on the engine rotation speed NE is determined according to the cooling water temperature THW.

Next, at step 504, NE becomes Nα.
If NE is equal to or greater than Nα, that is, if it is not the time of starting, the routine proceeds to step 506,
A normal injection timing calculation based on the engine operating state is performed.
On the other hand, when NE is smaller than Nα, that is, at the time of starting, the routine proceeds to step 508, where the starting injection timing A
The predetermined time AS during the compression stroke is adopted as INJ. The AS is a function of the engine speed NE and the accelerator opening ACP as the engine load, and is obtained experimentally in advance.

Although the embodiment of the present invention has been described above, the present invention is not limited to this, and various embodiments can be adopted.

[0055]

As described above, according to the present invention,
In a cylinder injection type gas-fueled internal combustion engine, a stable combustion state can be maintained even if the injection pressure changes, thereby improving output, improving drivability, reducing emissions, and the like.

[Brief description of the drawings]

FIG. 1 is a cylinder injection type CNG according to an embodiment of the present invention.
FIG. 2 is an overall schematic diagram of the institution.

FIG. 2 is a flowchart for calculating a swirl control valve opening SCVA;

FIG. 3 is a diagram showing a map in which a swirl control valve opening correction amount ΔSCVA is determined according to a fuel injection pressure deviation ΔP.

FIG. 4 is a flowchart for calculating an EGR valve opening EGRVA;

FIG. 5 is a diagram showing a map in which an EGR valve opening correction amount ΔEGRVA is determined according to a fuel injection pressure deviation ΔP.

FIG. 6 is a flowchart for calculating an ignition timing SA.

FIG. 7 is a characteristic diagram showing a relationship between a fuel injection pressure deviation ΔP and a variation ΔT of a spray temperature from a reference value.

FIG. 8 is a characteristic diagram showing a relationship between a fuel injection pressure deviation ΔP and a variation ΔE of a kinetic energy of spray from a reference value.

FIG. 9 is a view for explaining one means for reducing the spray speed.

FIG. 10 is a diagram for explaining another means for reducing the spray speed.

FIG. 11 is a view for explaining still another means for reducing the spray speed.

FIG. 12 is a flowchart for calculating a fuel injection timing.

FIG. 13 is a diagram showing how the engine rotation speed increases at the time of starting.

FIG. 14 is a flowchart for calculating a fuel injection timing at the time of starting the engine.

FIG. 15 is a view showing a map in which a rotation speed reference value for determining whether or not the engine is started based on the engine rotation speed is determined in accordance with the cooling water temperature.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 ... In-cylinder injection type CNG engine 2 ... Cylinder block 3 ... Cylinder head 4 ... Cylinder 5 ... Piston 6 ... Connecting rod 7 ... Crankshaft 8 ... Combustion chamber 9 ... Intake port 10 ... Exhaust port 11 ... Intake valve 12 ... Exhaust valve DESCRIPTION OF SYMBOLS 13 ... Intake side camshaft 14 ... Exhaust side camshaft 15 ... Intake side cam 16 ... Exhaust side cam 17, 18, 19 ... Timing pulley 20 ... Timing belt 30 ... Intake passage 31 ... Air cleaner 32 ... Throttle valve 33 ... Surge tank 34 ... intake manifold 37 ... throttle motor 38 ... swirl control valve 39 ... swirl control actuator 40 ... injector 41 ... fuel cylinder 44 ... regulator 50 ... spark plug 52 ... igniter built-in ignition coil 54 ... obstruction plate 60 ... exhaust passage DESCRIPTION OF SYMBOLS 1 ... Exhaust manifold 62 ... Catalytic converter 64 ... Exhaust gas circulation path 68 ... EGR valve 71 ... Vacuum sensor 74 ... Water temperature sensor 77 ... Accelerator opening sensor 78 ... Injection pressure sensor 80 ... Crank reference position sensor 81 ... Crank angle sensor 90 ... Electronic control unit (ECU)

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI theme coat ゛ (Reference) F02D 41/02 301 F02D 41/02 301E 325 325K 41/06 335 41/06 335S 43/00 301 43/00 301B 301N 301U 301J F02M 21/02 F02M 21/02 L 25/07 550 25/07 550F 570 570A F02P 5/15 F02P 5/15 G F-term (reference) 3G022 AA08 EA07 FA06 GA00 GA05 GA08 3G06 EA10 CA06 GA04 GA06 GA15 3G084 BA15 BA17 BA20 BA21 DA11 EB08 EB12 FA10 FA18 FA33 3G092 AA06 AA10 AA17 AB08 BA09 BB06 DC06 DC09 DE03S DG09 EC10 FA05 GA03 HB02X HB03Z HC09X HD07X HE01Z HF08Z03G03 HA04 PE03 P04

Claims (6)

[Claims] 1. An internal combustion engine for directly injecting gas fuel into a cylinder, means for detecting a pressure of the gas fuel to be injected, means for controlling an opening of a swirl control valve according to the detected pressure, A gas-fueled internal combustion engine comprising: 2. In an internal combustion engine that injects gas fuel directly into a cylinder, means for detecting the pressure of the injected gas fuel, and means for controlling the opening of the EGR valve according to the detected pressure. A gas-fueled internal combustion engine, comprising: 3. An internal combustion engine for directly injecting gas fuel into a cylinder, comprising: means for detecting the pressure of the gas fuel to be injected; and means for controlling the ignition timing in accordance with the detected pressure. A gas-fueled internal combustion engine, characterized by: 4. The gas fuel according to claim 1, further comprising means for reducing a flow velocity when the injected gas fuel collides with the spark plug. Internal combustion engine. 5. The fuel injection system according to claim 1, further comprising means for controlling an injection timing so that gas fuel injection is started after the intake valve is closed. A gas-fueled internal combustion engine according to claim 1. 6. The apparatus according to claim 1, further comprising means for controlling an injection timing so that the gas fuel is injected in a compression stroke when the engine is started. A gas-fueled internal combustion engine as described.