JP2003097357A - Gas engine, and control method thereof - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a gas engine and a control method thereof in which any blow-by of a fuel gas during the scavenging operation is prevented, the intake efficiency is improved, and generation of the knocking is suppressed. SOLUTION: In a gas engine (1), a fuel injection device (7) is provided in an intake system (1S), the fuel injection device (7) ejects the fuel gas at the timing of introducing the pre-mixture of the fuel gas and air into a cylinder (10) after the burnt gas remaining in the cylinder (10) is scavenged, and an exhaust valve (9) is closed. A mechanism for adjusting the opening/closing timing of intake and exhaust valves (8 and 9) is set so that the overlap of opening the intake valve (8) is relatively large (the crank angle of 60-110 deg.) when the exhaust valve (9) is opened.

Description

Detailed Description of the Invention

[0001]

TECHNICAL FIELD The present invention relates to an improvement in a gas engine and a control method thereof.

[0002]

2. Description of the Related Art In a conventional gas engine, for example, as shown in FIG. 28, a premixture Gm obtained by a mixer 3 for mixing and agitating fuel gas Ga and intake air Ai is compressed by a compressor 16 of an exhaust turbo 15. Then, the premixed air-fuel mixture Gm having a high pressure and high temperature is cooled by the intercooler 4, and the flow rate of the cooled air-fuel mixture Gm is adjusted by the throttle valve 6.
The piston 12 performs four strokes of intake, compression, combustion / expansion, and exhaust, and exhaust gas Eg is exhausted through the exhaust turbine 17.

29 to 32 show a known Otto cycle showing this process, in which the intake valve 8A is open and the exhaust valve 9A is closed during intake (FIG. 29).
Both intake and exhaust valves 8A and 9A are closed during compression (FIG. 30) and combustion / expansion (FIG. 31), and intake valve 8A is closed and exhaust valve 9A is open during exhaust. That is, the intake valve 8A and the exhaust valve 9A are not open at the same time.
Reference numeral 12A in the figure indicates the piston at the top dead center,
b shows the pistons at the bottom dead center.

At this typical valve opening / closing timing, since the residual gas remaining in the upper part of the combustion chamber 10 is not discharged,
In an actual engine, a so-called overlap period in which the intake valve 8A and the exhaust valve 9A are simultaneously opened and opened is provided in the process of intake to exhaust of the air-fuel mixture Gm into the combustion chamber 10. Then, the combustion gas remaining in the combustion chamber 10 at the piston top dead center is scavenged and discharged, and the overlap period is preferably the minimum period required for scavenging.

Particularly, in a conventional premixed gas engine for scavenging the residual combustion gas with the mixed gas Gm, since the premixed gas is used for the scavenging, the fuel contained in the premixed gas Gm is proportional to the overlap period. There is a problem of “blowing” (arrow B in FIG. 28) in which the gas Ga is discharged as it is together with the exhaust gas Eg without performing work. FIG. 44 illustrates the valve opening / closing operation state in which the overlap angle is 56 degrees in polar coordinates.

Further, when the scavenging and exhausting of the residual gas is insufficient, the temperature of the new premixed gas Gm to be taken in is raised, so that there is a serious problem that the intake efficiency is lowered and knocking is likely to occur. is there.

33 to 42 show a so-called Miller cycle which is effective in avoiding knocks which lead to a reduction in thermal efficiency and damage to the engine. 33 to 37
Indicates a so-called “late closing mirror cycle”.
During intake, the intake valve 8A is in the open state and the exhaust valve 9A is in the closed state (FIG. 33). After the compression stroke, the intake valve 8A is closed (FIG. 34) and compressed at a low compression ratio ( (Fig. 35)
Then, it is burned and expanded to the bottom dead center (FIG. 36), and is exhausted (FIG. 37). By this "late closing" of the intake valve 8A, the effective compression ratio is reduced and the rise in temperature and pressure at the compression end is suppressed.

38 to 42 show a so-called "early closing mirror cycle". The intake valve 8A is closed at the position of the piston 12Ac before the top dead center during intake (FIG. 38),
The piston 12Ac reciprocates to the bottom dead center for ineffective expansion and ineffective compression (FIG. 39), then effective compression (FIG. 40), full stroke expansion (FIG. 41), and exhaust (FIG. 42). This "early closing" of the intake valve 8A reduces the effective compression ratio and suppresses the rise in temperature and pressure at the compression end.

FIG. 43 shows an example of a valve lift diagram showing the opening / closing characteristics of the intake valve. The vertical axis is the valve lift and the horizontal axis is the crank angle, and "early closing" is the line RC and "late closing" is the line L.
It is represented by C. From this lift and the valve opening period (crank angle period), it can be seen that the lift is low and the valve opening period is short in "early closing", for example.

The selection of "late closing" and "early closing" in the Miller cycle is neglected depending on the structure and output characteristics of the engine, but the "early closing" method closes the intake valve 8A early. Therefore, the seating impact force of the intake valve 8A is large, and the valve lift is not large enough to reduce the flow passage area, but on the other hand, the premixed gas Gm entering the combustion chamber 10 is blown to the intake pipe. There is no return. On the other hand, in the “late closing” method, since the intake valve 8A is slowly closed, a large valve lift can be obtained while suppressing the seating impact force, the flow passage area can be increased, and the intake resistance can be reduced. On the other hand, "blowing back" of the intake premixed gas Gm to the intake pipe due to "late closing" cannot be avoided.

As described above, in the conventional late-closing Miller cycle type gas engine, the fuel gas and air once sucked into the cylinder are blown back into the intake pipe and stay there. These flow into the cylinder again at the beginning of the intake stroke of the next cycle. Therefore, the arrow B in FIG.
The problem of "blowing through" of fuel gas is large as shown in. In order to deal with this problem, as shown in FIG. 44, the valve overlap angle is often set to a relatively small numerical value such as OL = 56 °.

However, if the overlap is made small, the residual portion of the combustion exhaust gas cannot be sufficiently discharged, and it is the ratio of the residual gas (gas remaining in the cylinder at the top dead center) to the new air-fuel mixture. "Ratio of residual gas"
Will increase. The residual gas is at a high temperature, and if the ratio of the residual gas to the new air-fuel mixture is large, the temperature of the whole intake air-fuel mixture rises, causing a problem that the intake efficiency is lowered and knocking occurs.

[0013]

SUMMARY OF THE INVENTION The present invention has been proposed in view of the above-mentioned problems of the prior art, and prevents the fuel gas from passing through during scavenging to improve the intake efficiency and prevent knocking. The purpose is to provide a gas engine and a control method that can be suppressed.

[0014]

As a result of various studies, the inventor has supplied only air from an intake system instead of a mixer for mixing air and fuel gas, and the fuel gas is injected by a fuel injection means. In the case of a gas engine of the type, by appropriately controlling the timing of fuel gas injection, it is possible to prevent blowout of the air-fuel mixture by allowing only air to flow into the cylinder when discharging (scavenging) residual gas from the cylinder. Focused on what can be done.

The gas engine of the present invention was created on the basis of such findings, and a fuel injection device is provided in the intake system, and the fuel injection device scavenges the gas remaining after combustion in the cylinder and exhausts it. It is configured to inject fuel gas at the timing when a premixture of fuel gas and air is introduced into the cylinder after the valve is closed.The mechanism that adjusts the opening and closing timing of the intake and exhaust valves is the exhaust valve. When the intake valve is open, the valve overlap that opens the intake valve is 60 to 110 degrees (preferably 10 degrees).
0 degree: a relatively large angle) is set (claim 1).

In carrying out the present invention, the gas engine is preferably a Miller cycle of an early closing system (claim 2). Further, it is preferable that the fuel injection device is configured to start the injection of the fuel gas within a range from the start time of the intake valve opening to the top dead center (claim 3). More specifically, it is preferable to inject the fuel gas in any of the ranges of 20 to 40 degrees before the top dead center (TDC).

According to the method for controlling a gas engine of the present invention having the above-mentioned structure, in the fuel injection timing of the fuel injection device, only air is introduced into the cylinder when the gas remaining after combustion in the cylinder is scavenged. And the fuel gas is injected at the timing when the fuel gas or a premixture of fuel gas and air is introduced into the cylinder after the gas after combustion is scavenged and the exhaust valve is closed. (Claim 4).

Here, it is preferable that the injection of the fuel gas from the fuel injection device is started within the range from the intake valve opening start time to the top dead center (claim 5). More specifically, fuel gas injection is started in any of the ranges of 20 to 40 degrees from the top dead center (TDC), and the injection period is 30 to 30 degrees.
It is suitable to set it to 100 degrees (preferably about 50 degrees).

According to the present invention having such a configuration, only air is supplied from the intake system instead of the mixer for mixing air and fuel gas, and the fuel gas is injected by the fuel injection means. The fuel gas injection timing from the fuel injection means is the timing at which the pre-mixture of the fuel gas and air is introduced into the cylinder after the combustion gas remaining in the cylinder is scavenged and the exhaust valve is closed. Therefore, when scavenging the residual gas, the fuel gas is not injected from the fuel injection means, and only the air flows into the cylinder. Therefore, when scavenging the residual gas from the inside of the cylinder, only the air is blown from the intake valve to the exhaust valve, and the blowout of the fuel gas is completely prevented.

As a result of preventing the blow-through of the fuel gas, the valve overlap becomes large (crank angle of 60 degrees to
110 degrees: 100 degrees, for example) can be set. As a result, the high-temperature residual gas does not remain in the cylinder, so that the temperature of the new air-fuel mixture can be lowered. And
The intake efficiency is improved, knocking is suppressed, and the pump loss is decreased based on the decrease in the temperature of the new air-fuel mixture, so that the operation efficiency is also improved.

The gas engine of the present invention is constructed so that the combustion cycle is a premature closing Miller cycle. This is because when the late-closing type mirror cycle is adopted, a mixture of air and fuel gas is used for scavenging residual gas by blowback, and the advantage of using only air for scavenging is not utilized.

The fuel injection means is not particularly limited, and any known gas injection device may be used. The timing for starting the fuel gas injection from the fuel injection means may be any one of the top dead center (TDC) from the intake valve opening start time.

[0023]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the accompanying drawings. In FIG. 1, a gas engine 1 of the present invention includes an intake system 1S for supplying fuel Ga and combustion air Ai to an engine combustion chamber 10, and intake, combustion, and exhaust of fuel Ga and air Ai supplied from the intake system 1S. It is composed of a combustion system 1N for causing the exhaust gas Eg to be discharged to the outside through the combustion system 1N.

The intake system 1S includes a first intake pipe 2 for intake air Ai and the intake air Ai for the engine combustion chamber 1.
The compressor 16 of the exhaust turbo 15 that is supercharged to 0, the intercooler 4 that cools the intake air Ai that is compressed and heated by the compressor 16, the second intake pipe 2A that communicates with the intercooler 4, and the second air The fuel injection device 7 is attached to the pipe 2A and injects the fuel gas Ga into the second intake pipe 2A.

The combustion system 1N is provided above the cylinder 14 and supplied with fuel Ga and air Ai from the intake system 1S.
A combustion chamber 10 for combusting the fuel, a cylinder 14 forming the combustion chamber 10, a piston 12, a crankshaft 12A communicating with the piston 12, fuel Ga, air Ai in the combustion chamber 10.
The intake valve 8 is an intake valve for introducing the exhaust gas, and the exhaust valve 9 is an exhaust valve for exhausting the exhaust gas Eg after combustion to the exhaust valve system 1E.

The exhaust system 1E includes a first exhaust pipe 13 that guides the exhaust gas Eg to the exhaust turbo 15, a turbine 17 of the exhaust turbo 15, and a second exhaust pipe 13 that communicates with the turbine 17.
It is composed of A and.

In the fuel injection device 7, after the combustion exhaust gas Eg remaining in the combustion chamber 10 in the cylinder 14 is scavenged to the outside of the combustion chamber 10 and the exhaust valve 9 is closed, the fuel gas Ga and the air Ai are discharged. It is configured to inject and supply the fuel gas Ga at the timing when the premixed gas Gn is introduced into the combustion chamber 10.

In this embodiment, the injection timing is in the range from the time when the intake valve 8 opens and starts opening to the top dead center, more specifically, the top dead center (TDC).
The control is configured so that the fuel injection is started in the range of 20 to 40 degrees before.

2 to 5 show the mounting position of the fuel injection device 7, the timing of injecting the fuel gas Ga, the unburned gas amount ratio, etc. before the intake valve closes early = top dead center (360 degrees) (bTD).
C) 20 degrees, intake valve / exhaust valve opening overlap = 1
The data is shown under the condition of 00 degrees.

FIG. 2 shows the measurement position of the unburned gas Ga, and FIGS. 3 to 5 show the unburned gas amount ratio with respect to the crank angle. In FIG. 2, the fuel injection device 7 is attached to the second intake pipe 2A, the measurement position P7 of the unburned gas Ga is near the injection hole 7a, and the measurement position of the unburned gas Ga is on the intake port 2Ap side of the intake valve 8. P8 is the exhaust port 12p of the exhaust valve 9
The measurement position P9 of the unburned gas Ga is provided on each side.

FIG. 3 shows the unburned gas amount ratio at point P7 when the fuel gas injection timing is 30 degrees before top dead center. The horizontal axis represents the crank angle (degrees) and the vertical axis represents the vertical axis. It is the unburned gas amount ratio. The curve Q1 shows that the unburned gas amount ratio rises at 30 degrees before top dead center, reaches a maximum value at top dead center, and becomes zero at 40 degrees after top dead center.

FIG. 4 shows the unburned gas amount ratio at the measurement position P8. The curve Q2 shows that the fuel gas amount is zero up to 60 degrees after the exhaust valve is closed and after the top dead center. It shows that it is standing up from the middle of the intake stroke.
It becomes zero again before the end of the intake stroke, indicating that all fuel is being sucked.

FIG. 5 shows the unburned gas amount ratio at the measurement position P9, and the curve Q3 shows that there is no unburned gas in the full angle of the crank, that is, there is no blown by unburned gas.

6 and 7 show the relationship between the fuel injection timing, the thermal efficiency, and the amount of fuel blown into the exhaust gas. It is a measured value under conditions such as a net mean effective pressure (BMEP) of 1.50 MPa and a fuel injection period of 50 degrees.

In FIG. 6, the horizontal axis represents the fuel injection start timing (before top dead center) (degrees), and the vertical axis represents the thermal efficiency%. The curve E1 shows that the efficiency decreases significantly when the injection start timing is 20 degrees to 0 degrees before the top dead center with respect to the efficiency when the injection start timing is 40 degrees to 20 degrees before the top dead center.

In FIG. 7, the vertical axis represents the total HC (THC) concentration pp.
The curve P2 shows that when the injection start timing is 20 degrees to 0 degrees before the top dead center, the curve P2 is much worse than before.

From the results shown in FIGS. 6 and 7, the fuel injection timing is preferably 20 to 40 degrees before the top dead center.

The intake valve 8 and the exhaust valve 9 are provided when the exhaust valve 9 is opened and opened by a mechanism (not shown) for adjusting the opening / closing timing of the valve, such as a cam shaft.
Is opened to be in an open state, that is, a so-called valve opening overlap period is a period of 60 degrees to 110 degrees or more in crank angle. The overlap angle period is preferably determined according to the structure of the gas engine and the required performance characteristics.

FIG. 8 shows the opening / closing operation state of the intake valve 8 and the exhaust valve 9 in polar coordinates. A hatched missing outer ring indicated by reference sign V8 is an operation diagram of the exhaust valve 9 and has a valve opening angle period of 272 degrees that starts opening 51 degrees before bottom dead center and closes 41 degrees after top dead center. There is. The hatched inner missing ring indicated by reference sign V9 is an operation diagram of the intake valve 8 and has an opening angle period of 270 degrees that starts opening at 50 degrees before top dead center and closes at 40 degrees after bottom dead center. There is. From the figure, the period during which both valves 8 and 9 overlap and open is 91 degrees both at the top and bottom dead centers.

The overlap angle of 91 degrees in FIG. 8 is as shown in FIG.
Compared with the overlap angle of 56 degrees described in the related art, the increase is slightly less than twice. Simulation data showing the effect of this overlap angle are shown in FIGS. 9 and 10. In FIG. 9, the overlap angle (degree) is taken as a variable on the horizontal axis, and the vertical axis shows the ratio (%) of the residual gas remaining in the combustion chamber 10 and the maximum temperature (K) of the new air-fuel mixture entering the combustion chamber 10. The solid line Gz showing the residual gas ratio shows a high residual gas ratio even though the overlap angle 0 to about 60 degrees tends to be improved, and the solid line Gz is 60 to 160 degrees.
In terms of the degree, the residual gas ratio is low with a gradual improvement trend.

The dotted line T showing the maximum temperature of the new air-fuel mixture
n has a tendency to decrease in temperature almost in proportion to the residual gas ratio as the overlap angle increases. That is, the high temperature residual gas heats and raises the temperature of the new air-fuel mixture.

In FIG. 10, the display of FIG. 9 is changed, and the residual gas ratio (%) is taken as a variable on the abscissa and the ordinate is the maximum temperature (K) of the new air-fuel mixture. As the relationship line Tm shows, the maximum temperature of the new air-fuel mixture becomes high as the residual gas ratio increases. If the maximum temperature of the new air-fuel mixture in FIG. 9 and FIG. 10 is higher, the intake efficiency into the combustion chamber is reduced to the extent that it is more likely to cause knock, resulting in a reduction in output and a reduction in efficiency.

FIG. 11 and FIG. 12 show the function and effect using the overlap angle as a parameter. In FIG. 11, the vertical axis indicates the knocking limit ignition (ignition) angle (front TDC), and the horizontal axis indicates the net mean effective pressure (BMEP) (MPa).
Is taken to show the relationship with the overlap angle. The solid line B1a is the characteristic when the overlap angle is 91 degrees, and the dotted line B1b is the characteristic when the overlap angle is small (56 degrees in FIG. 11). In line B1a and line B1b, as the net mean effective pressure (BEMP) increases, the knocking limit ignition (ignition) timing becomes smaller. Further, the solid line B1a having an overlap angle of 91 degrees is always larger than the knocking limit ignition (ignition) timing (previous TDC) than the dotted line B1b, indicating that the knock resistance is good.

FIG. 13 shows a measured value of the working effect of pumping loss related to engine rotational friction with the overlap angle as a parameter, in the P-V diagram. The vertical axis of FIG. 14 is the in-cylinder pressure (kPa), and the horizontal axis is the actual combustion chamber volume (cc) that changes depending on the stroke. The solid line C1 indicates an overlap angle of 91 degrees, and the dotted line C2 indicates a relatively large overlap angle. The case is small (56 degrees in FIG. 13). In the figure, areas C1S and C2S surrounded by the intersections Gc and below of the lines C1 and C2 respectively indicate pumping loss. As is apparent from the figure, the pumping loss C1S of the solid line C1 representing the overlap angle of 91 degrees is smaller than the overlap angle 51 degrees represented by the dotted line C2.

Here, in the embodiment shown in FIG.
The employed mirror cycle is the "early closing" of the characteristic RC in FIG. Combustion chamber 1 in "late closing" of characteristic LC
This is because the mixture of the air and the fuel gas blown back into the scavenging of the residual gas in 0 is used, and the effect of preventing blow-through becomes small.

14 to 21 are schematic diagrams for explaining the mixture blow-through of "early closing" and "late closing" when the valve opening overlap angle is large, and FIGS. 14 to 17 show "early closing". ",
18 to 21 show "late closing". Figure 1 of "early closing"
4, the fuel gas Ga is in the combustion chamber 10 in the first half of the intake stroke.
Has been inhaled.

In the latter half of the intake stroke of FIG. 15, the intake valve 8 closes early to completely contain the fuel gas Ga once entering the combustion chamber 10. In the compression stroke of FIG. 16, the intake port 2A
Only air is stagnant at p. In the scavenging stroke of FIG. 17, only the air in the intake port 2Ap blows through and scavenges during valve opening overlap.

In FIG. 18 of "late closing", the intake of the fuel gas Ga into the combustion chamber 10 is shown. This process is still similar to "early closing". In FIG. 19, since the intake valve 8 closes late, a part of the fuel gas mixture in the compression stroke is blown back to the intake port 2Ap. In FIG. 20, the intake valve 8 is closed in the compression stroke, so that the air-fuel mixture becomes the intake port 2A.
stay in p. In FIG. 21, when the valve opening overlaps in the scavenging stroke, the air-fuel mixture staying in the intake port 2Ap blows out.

As described above, the blow-through in the scavenging process by "early closing" can be performed only by air.
In “late closing”, the air-fuel mixture blows out during the scavenging process and fuel is wasted. This is apparent by comparing FIGS. 2 to 5 showing the flow characteristic of “early closing” and FIGS. 22 to 25 showing the measured flow characteristic of “late closing”.

22 to 25 showing the "late closing", the installation position of the fuel injection device 7, the timing of injecting the fuel gas Gm, the unburned gas amount ratio, etc. are shown by the intake valve late closing = top dead center (36).
(0 degree) (aTDC) 80 degrees, and the data is shown under the condition that the intake valve / exhaust valve opening overlap = 100 degrees.

FIG. 22 shows the measurement position of the unburned gas amount, and FIGS. 23 to 25 show the unburned gas amount ratio with respect to the crank angle. 22, the mounting position of the fuel injection device 7, the measurement position P7 of the unburned gas, the measurement position P8 of the unburned gas, and the measurement position P9 of the unburned gas are substantially the same as those in FIG.

FIG. 23 is a diagram showing the unburned gas amount ratio at the point P7 when the fuel gas injection timing is 0 degrees after the top dead center. The horizontal axis shows the crank angle (degrees) and the vertical axis shows the unburned gas. It is a ratio. The curve Q1d shows that the unburned gas amount ratio rises at 0 degrees after top dead center, takes a maximum value at 20 degrees after top dead center, and becomes zero at 60 degrees after top dead center.

FIG. 24 shows the unburned gas amount ratio at the measurement position P8. The curve Q2d rises in the intake stroke of 60 degrees after the top dead center, but the fuel is delayed by closing after the intake stroke bottom dead center. It shows that the gas is being blown back.

FIG. 25 shows the unburned gas amount ratio at the measurement position P9, and the curve Q3d shows that the fuel gas is blown to the exhaust side at the overlap timing of the exhaust stroke and the intake stroke.

Comparing the remaining amount of unburned gas in the above "late closing" FIGS. 23 to 25 with the above "early closing" FIGS. 3 to 5, retention of unburned gas that is blown by "late closing" ( It can be seen that there are many blow-throughs (Q3d) due to the large number of Q2d hatching portions) and the large loss of fuel, which is disadvantageous.

FIGS. 26 and 27 compare the “late closing” and the “early closing” with the ignition timing taken as a variable on the horizontal axis. FIG. 26 is a diagram in which the thermal efficiency% is plotted on the vertical axis, and “early closing” at 0 degrees after top dead center is represented by a symbol Rh, and “late closing” at 80 degrees after top dead center is represented by a symbol Lh. There is. "Early closing"
The line Rh is always higher in thermal efficiency than the “late closing” Lh.

In FIG. 27, the vertical axis represents total hydrocarbon (THC) p.
pm and sign Rt for "early closing" at 0 degrees after top dead center
In addition, "late closing" of 80 degrees after the top dead center is represented by a line diagram with the symbol Lt. The line Rt of "early closing" is always "late closing"
The total hydrocarbon concentration is lower than Lt. That is, it indicates that the amount of fuel gas blown through is decreasing.

14 to 27 and FIGS. 2 to 5 described above.
Therefore, in the Miller cycle, “early closing” is advantageous on the assumption that the fuel injection timing can be freely selected.

[0059]

The effects of the present invention are listed below. (1) According to the present invention, a fuel injection device that can change the generation of a premixed gas by a conventional mixer is provided, and fuel is supplied to the combustion chamber only when necessary. Therefore, the valve opening overlap angle can be increased and residual gas The air can be scavenged with only air to prevent loss of fuel due to blow-by, and it also improves knock resistance. (2) Since the residual gas is eliminated, the temperature of the new air-fuel mixture is not increased, and the suction efficiency and the combustion efficiency are increased. (3) By combining with the early closing mirror cycle,
Pumping loss can be reduced and knock resistance can be further improved, improving efficiency. (4) By adopting the early closing method, the thermal efficiency can be improved and the THC concentration of the exhaust gas can be reduced.

[Brief description of drawings]

FIG. 1 is a configuration diagram showing an embodiment of the present invention.

FIG. 2 is a side sectional view showing a measurement position of unburned gas near the intake and exhaust valves.

FIG. 3 is an actual measurement diagram showing unburned gas existing in an intake pipe.

FIG. 4 is an actual measurement diagram showing unburned gas existing in an intake port on an intake valve.

FIG. 5 is an actual measurement diagram showing unburned gas existing in an exhaust port on an exhaust valve.

FIG. 6 is an actual measurement diagram showing the relationship between fuel injection start timing and thermal efficiency.

FIG. 7 is an actual measurement diagram showing the relationship between fuel injection start timing and THC concentration in exhaust gas.

FIG. 8 is a diagram showing the operation of intake / exhaust valves with a large overlap angle (91 degrees).

FIG. 9 is an actual measurement diagram showing the relationship between the overlap angle, the residual gas ratio, and the maximum temperature of a new air-fuel mixture.

FIG. 10 is an actual measurement diagram showing the relationship between the residual gas ratio and the maximum temperature of a new air-fuel mixture.

FIG. 11: Large overlap angle (91 degrees), small overlap (56
Measurement diagram showing the relationship between knock limit injection timing and BMEP in degrees).

FIG. 12 Large overlap angle (91 degrees), small overlap angle (56)
The measured figure which shows the relationship between thermal efficiency and BMEP in degrees.

FIG. 13 is an actual measurement diagram showing a relationship between a large overlap angle (91 degrees) and a small overlap angle (56 degrees) and a pumping loss in a mirror cycle.

FIG. 14 is a diagram of an intake stroke in the explanation of the action of “large overlap angle + early closing”.

FIG. 15 is a diagram of an early closing state in the intake stroke of “large overlap angle + early closing”.

FIG. 16 is a diagram of a compression process of “large overlap angle + early closing”.

FIG. 17 is a compression stroke diagram showing scavenging of “large overlap angle + early closing”.

FIG. 18 is a diagram illustrating an intake stroke of “large overlap angle + late closing”.

FIG. 19 is a diagram in which the air-fuel mixture is blown back in the compression stroke of “large overlap angle + late closing”.

FIG. 20 is a diagram in which the air-fuel mixture blown back in the compression stroke of “large overlap angle + late closing” is retained in the intake port.

FIG. 21 is a diagram in which the air-fuel mixture staying in the blow-back intake port blows through during scavenging in the compression stroke of “large overlap angle + late closing”.

FIG. 22 is a side sectional view showing a measurement position of unburned gas near the intake and exhaust valves.

FIG. 23 is an actual measurement diagram showing unburned gas existing in the intake pipe.

FIG. 24 is an actual measurement diagram showing unburned gas existing in an intake port on an intake valve.

FIG. 25 is an actual measurement diagram showing unburned gas existing in an exhaust port on an exhaust valve.

FIG. 26 is an actual measurement diagram showing a relationship between ignition timing and thermal efficiency in “early closing” and “late closing”.

[Fig. 27] Ignition timing and TH in "early closing" and "late closing"
The measurement figure which shows the relationship of C.

FIG. 28 is a configuration diagram of a conventional gas engine.

FIG. 29 is a view showing an intake stroke of a conventional Otto cycle.

FIG. 30 is a view showing a compression stroke of a conventional Otto cycle.

FIG. 31 is a view showing an expansion stroke of a conventional Otto cycle.

FIG. 32 is a view showing an exhaust stroke of a conventional Otto cycle.

FIG. 33 is a diagram showing an intake stroke of a “late closing” Miller cycle.

FIG. 34 is a diagram showing an intake valve late closing invalid compression stroke of a “late closing” Miller cycle.

FIG. 35 shows the compression stroke of a “late closing” Miller cycle.

FIG. 36 shows the expansion stroke of a “late closing” Miller cycle.

FIG. 37 is a diagram showing a scavenging stroke of a “late closing” Miller cycle.

FIG. 38 is a diagram showing an intake stroke of an “early closing” Miller cycle.

FIG. 39 is a view showing invalid expansion and invalid compression strokes by early closing of the “early closing” mirror cycle.

FIG. 40 shows the compression stroke for an “early close” Miller cycle.

FIG. 41 shows the expansion stroke of an “early close” Miller cycle.

FIG. 42 is a diagram showing a scavenging stroke of an “early closing” Miller cycle.

FIG. 43 is a diagram showing “early closing” and “late closing” intake valve lift.

FIG. 44 is a polar coordinate diagram showing a valve opening / closing operation with a small valve opening overlap between the intake valve and the exhaust valve (56 degrees).

[Explanation of symbols]

1 ... Gas engine 1S ··· Inhalation system 1E ... Exhaust system 1N ... Combustion system 2 ... the first intake pipe 2A ... Second intake pipe 4 ... Intercooler 7 ... Fuel injection device 8 ... Intake valve 9 ... Exhaust valve 10 ... Combustion chamber 12 ... Piston 12A ... Crankshaft 13 ... First exhaust pipe 13A ... Second exhaust pipe 14 ... Cylinder liner 15 ... Turbocharger 16 ... Compressor

   ─────────────────────────────────────────────────── ─── Continued front page    (72) Inventor Satoshi Morimoto             1-5-20 Kaigan, Minato-ku, Tokyo Tokyo Gas             Within the corporation (72) Inventor Makoto Akaike             1-5-20 Kaigan, Minato-ku, Tokyo Tokyo Gas             Within the corporation (72) Inventor Sakurai Teruhiro             1-5-20 Kaigan, Minato-ku, Tokyo Tokyo Gas             Within the corporation (72) Inventor Toru Nakazono             1-32 Yanmar, Chayamachi, Kita-ku, Osaka-shi, Osaka             Diesel Co., Ltd. (72) Inventor Toru Takemoto             1-32 Yanmar, Chayamachi, Kita-ku, Osaka-shi, Osaka             Diesel Co., Ltd. F term (reference) 3G092 AA18 AB06 BB06 DA12 DE01S                       FA02 FA16

Claims (5)

[Claims] 1. A fuel injection device is provided in an intake system, and the fuel injection device generates a premixed mixture of fuel gas and air after the burned gas remaining in the cylinder is scavenged and the exhaust valve is closed. It is configured to inject fuel gas at the timing when it is introduced into the cylinder, and the mechanism that adjusts the opening and closing timing of the intake and exhaust valves has a valve overlap that opens the intake valve when the exhaust valve is open. A gas engine, which is set to have a crank angle of 60 degrees to 110 degrees. 2. The gas engine according to claim 1, wherein the gas engine is an early closing Miller cycle. 3. The gas engine according to claim 1, wherein the fuel injection device is configured to start the injection of the fuel gas within a range from the intake valve opening start time to the top dead center. 4. The method of controlling a gas engine according to claim 1, wherein the fuel injection timing of the fuel injection device is such that only air is scavenged when the gas remaining after combustion in the cylinder is scavenged. The fuel gas is injected into the cylinder, and after the combustion gas is scavenged and the exhaust valve is closed, the fuel gas or the mixture of the fuel gas and air is injected into the cylinder at the timing of injecting the fuel gas. A method for controlling a gas engine, characterized by being set. 5. The method of controlling a gas engine according to claim 4, wherein the fuel gas is injected from the fuel injection device within the range from the intake valve opening start time to the top dead center.