JPS6363726B2 - - Google Patents

Description

[Detailed description of the invention] [Industrial application field]

The present invention relates to an intake system for an internal combustion engine having two intake holes per cylinder, and more particularly to an intake system for an internal combustion engine that forms a swirling flow within a combustion chamber to improve combustion.

[Prior art] As a method of reducing harmful components in exhaust gas,
It is known to use lean mixtures or to recirculate large amounts of exhaust gas into the engine intake system. However, when these methods are used, it is difficult to ensure stable combustion because the combustion speed of the air-fuel mixture becomes slow. Particularly during low-load operation, there is a large amount of burned gas remaining in the combustion chamber, which tends to slow down the combustion rate. Therefore, during low load operation, it is necessary to increase the combustion rate. In addition, there is a recent trend to highly compress the air-fuel mixture in order to improve fuel consumption, but increasing the compression ratio makes it more likely that knocking will occur, especially during high-load operation at low and medium speeds. In order to prevent this, it is necessary to increase the combustion rate. One method of increasing the combustion rate is to generate strong swirling flow or turbulence within the combustion chamber. As a method of generating a swirling flow in the combustion chamber, there is a method of using a helical intake port (for example, see Japanese Patent Application Laid-open No. 104612/1983),
Methods that utilize directional jets from narrow passages (for example, Japanese Patent Laid-Open No. 53-4109;
97605 and Japanese Patent Application Laid-Open No. 127914/1983) are known. In addition, the intake system that communicates with each combustion chamber from the confluence space formed in the intake passage downstream of the carburetor is composed of a low-speed intake port and a high-speed intake port, and the low-speed intake port has a smaller diameter than the high-speed intake port. The high-speed intake port is a large-diameter helical-type intake port in the opposite direction to the low-speed intake port, and the high-speed intake port is equipped with a control valve that opens when the engine is under high load or at high speeds. As a result, the swirling flow from the high-speed intake port collides with the swirling flow from the low-speed intake port from the side, promoting turbulence in the combustible mixture and improving combustion during high engine loads. It is known (for example, Utility Model Application Publication No. 54-91513).

[Problems to be solved by the invention]

By the way, during engine operation, there is a cooling loss due to heat transfer from the high-temperature gas in the combustion chamber during combustion or immediately after combustion to the relatively low-temperature inner wall of the cylinder, and this cooling loss is caused by a strong swirling flow. It's reasonably big. Since cooling loss causes an increase in fuel consumption rate, if the operating conditions are such that swirling flow is not very necessary, it is advantageous to have a weak swirling flow. An example of an operating condition in which this swirling flow is not so necessary is during medium load operation. That is, during medium-load operation, the air-fuel mixture in the cylinder has a smaller amount of residual burnt gas than during low-load operation, and the condition is such that it is relatively easy to burn.
Furthermore, the conditions are such that knocking is unlikely to occur, and a strong swirling flow is unnecessary, and cooling loss becomes a problem. However, in the conventional swirling flow generation mechanism described above, the swirling flow is formed not only during low load operation and high load operation when swirling flow is necessary, but also during medium load operation, so cooling loss increases. The problem was that the thermal efficiency decreased, resulting in an increase in the fuel consumption rate of the engine. Therefore, the present invention was proposed in view of the above problems, and its purpose is to form a swirling flow during low-load operation and high-load operation when swirling flow is necessary, and to form a swirling flow when swirling flow is not so necessary. On the contrary, during medium load operation where cooling loss becomes a problem, the swirling flow is reduced and fine air-fuel mixture turbulence is generated to improve combustion and reduce fuel consumption.

[Means to solve the problem]

Therefore, the present invention adopts the following configuration as a means for solving the above-mentioned problems. That is, the present invention provides the above-mentioned 2 cylinders per cylinder.
In an intake system for an internal combustion engine that has a large-diameter intake passage and a small-diameter intake passage that are bifurcated, the large-diameter intake passage is a helical intake port, and the small-diameter intake passage is 1/5 of the cross-sectional area of the large-diameter intake passage. It has a cross-sectional area of ~1/7 and forms a swirling flow in the combustion chamber in the opposite direction to the swirling flow caused by the large-diameter intake passage. A control valve is provided that is controlled to gradually increase the opening degree from the fully closed state as the pressure increases, and the swirling flow formed by the small diameter intake passage and the large diameter intake passage collides during medium load operation. and is characterized by being reduced. Specifically, to explain this using FIGS. 1 and 2 as examples, an intake system for an internal combustion engine has a merging space 14 in the intake passage downstream of the carburetor 15, and from the merging space 14 per cylinder. Large diameter intake passages 7a, 7b and small diameter intake passage 8a branched into two,
The combustion chamber 8b is connected to the combustion chamber 25 via intake valves 4 and 5 provided independently in each passage. The large diameter intake passages 7a, 7b are configured in a helical shape near the combustion chamber 25, and the intake passages 7a, 7b are configured in a helical shape near the combustion chamber 25.
As a result, a swirling flow is formed within the combustion chamber 25. The small diameter intake passages 8a and 8b are the same as the large diameter intake passage 7.
It has a cross-sectional area of 1/5 to 1/7 of the cross-sectional area of the intake passages a and 7b, and has an almost straight shape that opens tangentially near the outer periphery of the combustion chamber, and the swirling flow due to the large diameter intake passages 7a and 7b is It is connected to form a swirling flow in the combustion chamber 25 in the opposite direction. Further, a control valve 10 is provided in the large-diameter intake passages 7a, 7b, and the control valve 10 is controlled to gradually increase its opening degree from a fully closed state as the vent valve negative pressure of the carburetor 15 increases.

[Effect]

According to the above-mentioned means, during low-load operation, the control valve 10 is closed because no vent negative pressure is generated in the carburetor 15. Therefore, the entire air-fuel mixture is supplied into the combustion chamber 25 through the small-diameter intake passages 8a and 8b, and a strong swirling flow is formed within the combustion chamber 25, increasing the combustion speed. [Figure 3]. During medium load operation, as the amount of intake air increases, the vent lily negative pressure of the carburetor 15 also increases, and the control valve 1 is closed by an amount corresponding to the vent lily negative pressure.
0 is released. Therefore, the air-fuel mixture is supplied into the combustion chamber 25 from both the small-diameter intake passages 8a, 8b and the large-diameter intake passages 7a, 7b. Since the swirling flows formed by the small-diameter intake passages 8a, 8b and the large-diameter intake passages 7a, 7b are in opposite directions, they collide at the position where the flows meet. As a result, the swirling flow is reduced and fine turbulence is formed.
Cooling loss is reduced (Figure 4). Furthermore, during high-load operation, a large negative pressure is generated in the vent lily, and the control valve 10 is almost fully opened. Therefore, most of the air-fuel mixture is supplied into the combustion chamber 25 through the large-diameter intake passages 7a and 7b having a large cross-sectional area and low flow resistance. As a result, a strong swirling flow is formed in the combustion chamber 25 by the helical-shaped large-diameter intake passages 7a and 7b, increasing the combustion speed and suppressing the occurrence of knocking (FIG. 5).

【Example】

Embodiments of the present invention will be described in detail below with reference to the drawings. 1 and 2 are a partially sectional plan view and a vertical sectional view showing an intake system of an internal combustion engine according to an embodiment of the present invention, which is a four-cylinder spark ignition engine. The cylinder block 2 has four cylinders 23, and a piston 3 is inserted into each cylinder so as to be movable up and down. A cylinder head 1 is fixed above the cylinder block 2 via a gasket 24. The cylinder head 1 is connected to each cylinder 2.
Each valve 3 has a large-diameter intake valve 4, a small-diameter intake valve 5, an exhaust valve 6, and a spark plug 22. On the other hand, the carburetor 15 is of a two-barrel type, and has a primary vent lily 16 and a secondary vent lily 17, and is equipped with throttle valves 116 and 117, respectively.
A merging space 14 is provided downstream of the vaporizer 15. From the confluence space 14, there are two intake pipes per cylinder, namely a large diameter intake pipe 7a and a small diameter intake pipe 8.
a is branched and connected to a large diameter intake port 7b and a small diameter intake port 8b, respectively, and is further connected to a large diameter intake valve 4.
and a combustion chamber 25 via a small-diameter intake valve 5. Further, the exhaust port 9 is connected to the combustion chamber 25 via the exhaust valve 6, and the other end is ultimately opened to the atmosphere as in a general exhaust system. Each large-diameter intake pipe 7a is provided with a control valve 10 at a position close to the cylinder head 1. The movement of the diaphragm 12 that divides the space inside the diaphragm device 11 into an A chamber 20 and a B chamber 21 is transmitted to the control valve 10 via a link mechanism 13 . The B chamber 21 is open to the atmosphere, and the A chamber 20 is connected to the two in the vaporizer 15 via the pipe 19.
Negative pressure port 1 provided in the constriction part of the next bench lily 17
8 is connected. Further, the diaphragm 12 closes the control valve 10 when the diaphragm 12 is located on the B chamber 21 side by the spring 26, and opens the control valve 10 when negative pressure is introduced into the A chamber 20 and the diaphragm 12 moves toward the A chamber 20. work like that. The large-diameter intake port 7b is configured as a so-called helical-type intake port consisting of a straight pipe part and a spiral part, and guides the air-fuel mixture into the combustion chamber 25 so as to wrap around the upper part of the cap of the large-diameter intake valve 4. Forms a swirling flow.
Note that the large diameter intake pipe 7a and the large diameter intake port 7b are
In order to prevent deterioration of filling efficiency during high-load operation, a shape with a large cross-sectional area and a large radius of curvature of the spiral portion is preferable. The small-diameter intake pipe 8a and the small-diameter intake port 8b are arranged to form a substantially straight passage, and are configured as a so-called biased flow port connected (tangentially) to the combustion chamber 25 near the outer periphery of the combustion chamber. The direction of the small-diameter intake port 8b is arranged to form a swirling flow in the opposite direction to the swirling flow caused by the helical-type large-diameter intake port 7b. Furthermore, the cross-sectional area of the small-diameter intake pipe 8a and the small-diameter intake port 8b is the same as that of the large-diameter intake pipe 7a and the large-diameter intake port 7b.
is set to 1/5 to 1/7 of the cross-sectional area of This is because during medium load operation, both swirling flows formed by the small-diameter intake port 8b, which is a biased flow port, and the helical-type large-diameter intake port 7b collide with each other, and the swirling flow is reduced and the inside of the combustion chamber 25 is reduced. This value was experimentally determined to generate fine turbulence in the combustion chamber 25 and to form a strong swirling flow in the combustion chamber 25 mainly due to the helical large-diameter intake port 7b during high-load operation. Next, the operation of the intake system for an internal combustion engine having the above configuration will be explained. During low load operation, the amount of air taken in is small. Therefore, there is no or only a small amount of air passing through the secondary bench lily 17, and no vent lily negative pressure is generated. The diaphragm 12 of the diaphragm device 11 remains biased toward the B chamber 21 by the spring 26 and does not move, and the control valve 10 is closed.
Therefore, during low-load operation, the entire air-fuel mixture is supplied into the combustion chamber 25 through the small-diameter intake pipe 8a and the small-diameter intake port 8b, and the air-fuel mixture is supplied into the combustion chamber 25 as shown by arrow A in FIG.
A strong swirling flow is formed within 5. Therefore, even with a flame-retardant mixture containing a large amount of burnt residual gas, the combustion rate can be increased and stable combustion can be obtained. During medium load operation, as the intake air amount increases, the bench lily negative pressure also increases. This negative pressure is transferred to A of the diaphragm device 11 via the pipe 19.
The diaphragm 12 is introduced into the chamber 20 and the diaphragm 12 is connected to the spring 26.
The displacement of the diaphragm 12 is transmitted to the control valve 10 via the link mechanism 13, and the displacement of the diaphragm 12 is transmitted to the control valve 10 via the link mechanism 13. Open the amount according to the bench lily negative pressure. Therefore, as shown in FIG. 4, a part of the air-fuel mixture flows into the combustion chamber 25 through the small-diameter intake pipe 8a and the small-diameter intake port 8b, forming a swirling flow as shown by arrow A. Further, the remaining air-fuel mixture passes through the large-diameter intake pipe 7a and forms a swirling flow as shown by arrow B at the helical-type large-diameter intake port 7b. At this time, since the rotation directions of both swirling flows are opposite, they collide at the position where the flows meet.
As a result, swirling flow is reduced and fine turbulence occurs. By reducing the swirl flow and forming fine turbulence in this way, it is possible to reduce cooling loss, reduce fuel consumption rate, and further stabilize combustion. Furthermore, during high-load operation, a large amount of air passes through the secondary vent lily 17, generating a large vent lily negative pressure, and the control valve 10 becomes almost fully open. In this case, most of the air-fuel mixture is supplied into the combustion chamber 25 through the large-diameter intake pipe 7a with a large cross-sectional area and low flow resistance, and the helical-type large-diameter intake port 7b. A strong swirling flow is formed by port 7b. Note that although some of the air-fuel mixture passes through the small-diameter intake pipe 8a and the small-diameter intake port 8b, the amount thereof is so small that it hardly affects the swirling flow. Arrow B in FIG. 5 shows this situation. By forming a strong swirling flow, it is possible to increase the combustion rate and suppress the occurrence of knocking. Although the present invention has been described above with reference to specific embodiments, the present invention is not limited to the above embodiments, and includes various embodiments within the scope of the claims. For example, in the above embodiment, the carburetor is a two-barrel type, and the negative pressure port is provided in the secondary bench lily, but the negative pressure port is not limited to this embodiment, and the negative pressure port may be provided in both the secondary bench lily and the primary bench lily. It goes without saying that a single barrel type carburetor is fine. It can also be adopted in the case of a gasoline-injected internal combustion engine.

【Effect of the invention】

As described above, according to the present invention, stable combustion is achieved by accelerating the combustion rate by strong swirling flow during low-load operation and high-load operation, and by fine turbulence during medium-load operation. At the same time, during medium load operation, the swirling flow is reduced to reduce the cooling loss to the cylinder inner wall surface, and the fuel consumption rate can be lowered, which is an excellent practical effect.

[Brief explanation of drawings]

1 and 2 show an intake system for an internal combustion engine according to an embodiment of the present invention, and FIG. 1 is a plan view showing a cylinder head and a part of the intake system of the internal combustion engine in cross section. , FIG. 2 is a longitudinal cross-sectional view taken along the intake passage of the internal combustion engine (a cross-sectional view taken along the line - in FIG. 1), and FIGS. 3, 4, and 5 each show one embodiment of the present invention. This figure shows the intake air-fuel mixture flow in the combustion chamber of the internal combustion engine according to the example. Fig. 3 is a schematic explanatory diagram during low load operation, Fig. 4 is a schematic explanatory diagram during medium load operation, and Fig. 5 is a schematic explanatory diagram during high load operation. It is a schematic explanatory diagram at the time of load operation. Explanation of symbols: 1...Cylinder head, 4...Large diameter intake valve, 5...Small diameter intake valve, 7a...Large diameter intake pipe (large diameter intake passage) 7b...Large diameter intake port (large diameter intake passage) ), 8a...Small diameter intake pipe (small diameter intake passage), 8b...Small diameter intake port (small diameter intake passage),
10...Control valve, 11...Diaphragm device, 1
4...merging space, 15...carburizer, 17...secondary bench lily, 25...combustion chamber.

Claims (1)

[Scope of Claims] 1. A merging space is provided in the intake passage downstream of the carburetor, and a large-diameter intake passage and a small-diameter intake passage that branch out from the merging space into two per cylinder are provided independently. In an intake system for an internal combustion engine that is connected to a combustion chamber via an intake valve, the large-diameter intake passage is configured in a helical shape near the combustion chamber, and the intake passage forms a swirling flow within the combustion chamber. The small-diameter intake passage has a cross-sectional area of 1/5 to 1/7 of the large-diameter intake passage.
It has a cross-sectional area of , and has a substantially straight shape that opens tangentially near the outer periphery of the combustion chamber, and is connected to form a swirling flow in the combustion chamber in a direction opposite to the swirling flow caused by the large-diameter intake passage. At the same time, a control valve is provided in the large-diameter intake passage, and the control valve is controlled to gradually increase the opening degree from the fully closed state as the vent valve negative pressure of the carburetor increases. An intake system for an internal combustion engine, wherein swirling flows formed by the small-diameter intake passage and the large-diameter intake passage collide and are reduced.