JP2010090849A - Internal combustion engine and intake device therefor - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an intake device for an internal combustion engine, promoting the vaporization and mixing of fuel in the internal combustion engine, forming a vertical swirl of sufficient intensity in a cylinder irrespective of a load and the rotation speed of the internal combustion engine without reducing intake air volume even in high rotation and a high load and less in manufacturing cost. <P>SOLUTION: In the intake passage of the internal combustion engine, which is divided into two symmetrical intake ports from an intake manifold, at any cross-sectional position from the bifurcation start position of the intake ports to intake port outlets, intake port cross sections are formed in almost isosceles triangle shapes having a symmetry plane at bases. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to an intake device for an internal combustion engine, and more particularly, to an internal combustion engine that can reinforce a longitudinal vortex flow generated in a cylinder without reducing an intake air amount at a high load, and an intake device thereof.

  Conventionally, as shown in, for example, Japanese Patent Application Laid-Open No. 2004-169570, a tumble control capable of opening and closing one of a partition that divides a cross section of an intake port vertically and two air passages configured by the partition and an intake port wall surface Closing by a valve is performed. In this method, one side of the passage composed of the partition wall and the intake port wall surface is closed by a tumble control valve, so that a high-speed gas flow is generated from the intake port in the tangential direction of the cylinder, and a strong vertical vortex is generated in the cylinder. It is formed.

  Further, as shown in Japanese Utility Model Publication No. 5-44526, it is introduced from a direction substantially perpendicular to the cylinder axis, branches into two while being bent in a direction substantially parallel to the cylinder axis, and opens to the combustion chamber. In the suction passage of the cylinder head, the entire side of the suction passage is formed in a cross-sectional shape in which the centrifugal side curved from the axis is expanded in the lateral width direction and the centripetal side is reduced in the lateral width direction. In this method, a large amount of fuel gas flows to the centrifugal side of the curve due to the centrifugal force in the curved path, so by making the cross-sectional area of the centrifugal side larger than that of the centripetal side, the intake resistance of the gas is reduced and the cylinder The amount of inhaled gas can be increased.

JP 2004-169570 A Japanese Utility Model Publication No. 5-44526

  In the technique described in Japanese Patent Application Laid-Open No. 2004-169570, when the internal combustion engine is operated at a high load and a high speed, the tumble control valve is opened to reduce the intake resistance as much as possible, and the two divided by the partition walls are used. Gas is supplied into the cylinder from both sides of the passage. For this reason, at the time of high load and high speed operation, the gas speeding up action by the tumble valve is lost, and a sufficiently strong vertical vortex cannot be formed in the cylinder. Even when the tumble valve is opened, an increase in intake resistance due to the tumble valve and the partition wall occurs, so the amount of air taken into the cylinder decreases. Furthermore, since a tumble valve and a partition wall are installed, there is a problem that the manufacturing cost becomes high.

  Further, in the technique disclosed in Japanese Utility Model Publication No. 5-44526, the flow rate of gas colliding with the cylinder wall surface is increased by increasing the air density on the centrifugal side of the curved passage, so that the gas velocity after entering the cylinder is increased. There is a problem that the speed is reduced and a sufficiently strong vertical vortex cannot be formed in the cylinder.

  This problem will be described in more detail with reference to FIGS. FIG. 1 is a plan view showing a schematic shape of an internal combustion engine using an intake device according to the prior art. FIG. 2 is a longitudinal sectional view showing a schematic shape of an internal combustion engine using an intake device according to the prior art. This internal combustion engine has a so-called 4-valve internal combustion engine configuration having two intake valves and two exhaust valves on the cylinder head. 1 and 2, a piston 6 is fitted into the cylinder 3, and a combustion chamber 3C is formed surrounded by the cylinder 3, the cylinder head 3h, and the crown surface of the piston 6.

  The cylinder head 3h is provided with two intake ports 1 and 1 and two exhaust ports 2 and 2 that open to the combustion chamber 3C. The two intake ports 1 and 1 are grouped together on the upstream side of the intake partition wall 1C to form an intake manifold INT. Similarly, the two exhaust ports 2 and 2 are combined into one on the upstream side of the exhaust partition 2C to form an exhaust manifold EXH.

  An intake valve 9 is provided at the opening between the intake port 1 and the combustion chamber 3C, and an exhaust valve 10 is provided at the opening between the exhaust port 2 and the combustion chamber 3C.

  A spark plug 4 is provided substantially at the center of the cylinder head 3h, and a fuel injector 5 is provided on the intake side surface of the cylinder head 3h. That is, this internal combustion engine has a configuration of a direct injection internal combustion engine that directly injects fuel into the combustion chamber 3C.

  FIG. 3 is a simplified view of the internal combustion engine shape shown in FIGS. 3 shows only the intake manifold INT, the intake port 1, the intake valve 9, the cylinder head 3h, and the cylinder 3 in FIGS.

  As shown in FIG. 3, intake passage cross sections A1-A1, A2-A2, and A3-A3 are respectively defined. That is, each cross section A1-A1... A3-A3 is a cross section perpendicular to the central axis AX of the intake passage. A1-A1 is a cross section of the intake manifold INT, A2-A2 is a cross section at a substantially central position of the intake port 1, and A3-A3 is downstream from A2-A2 and slightly upstream from the opening of the intake port 1 to the combustion chamber. It is a cross section of the side. The target surface of the two intake ports 1 and 1 is CL.

  An example of a cross-sectional shape of an intake device according to the prior art is shown in FIGS. FIG. 4 shows an example of a substantially oval cross-sectional shape in the cross section A1-A1 of the intake pipe portion, and a substantially circular shape in the intake port cross sections A2-A2 and A3-A3 after branching. FIG. 5 shows a substantially oval cross-sectional shape in the cross section A1-A1 of the intake pipe portion, a substantially triangular shape in the cross section A2-A2 in the substantially central portion of the intake port, and A3-A3 in the downstream of the intake port. It is an example of a substantially circular shape.

  FIG. 6 and FIG. 7 show the gas flow in the intake stroke in the intake device according to the prior art. In the intake stroke, the intake valve 9 opens toward the combustion chamber 3C, so that an opening S9 is formed between the intake port 1 and the combustion chamber 3C. Further, as the piston 6 descends, the pressure in the combustion chamber 3C is lower than that in the intake port 1, so that gas is drawn from the intake port 1 through the opening S9 into the combustion chamber 3C.

  As shown in FIG. 7, since the intake port 1 is inclined toward the intake side with respect to the center axis CC of the combustion chamber, the flow velocity VF on the left side (combustion chamber center side) of the intake valve is caused by the inertia force of the gas. As a result, the counterclockwise vertical vortex TU is generated in the combustion chamber 3C.

  This vertical vortex TU has an effect of promoting the mixing of air and fuel in the combustion chamber 3C, and increasing the combustion speed by being collapsed in the later stage of the compression stroke and changing to a turbulent vortex. These are known to play an important role in improving the basic performance of an internal combustion engine, such as promoting fuel combustion, reducing fuel consumption, improving output, suppressing knocking, and purifying exhaust.

  8 and 9 are examples showing the velocity distribution of the gas entering the combustion chamber 3C from the intake valve 9 in the intake stroke. FIG. 8 shows a velocity distribution in the intake valve when the port cross-sectional shape shown in FIG. 4 is used. FIG. 9 shows the velocity distribution at the intake valve when the port cross-sectional shape shown in FIG. 5 is used.

  In FIG. 8, since the intake port cross section (A2-A2) is axisymmetric as shown in FIG. 4, the flow velocity distribution in the intake port is substantially uniform. For this reason, the gas velocity component VF that mainly generates the vertical vortex in the combustion chamber 3C is not sufficiently strong compared with the gas velocity component VR that enters from the intake side of the intake valve 9 and suppresses the generation of the vertical vortex. The longitudinal vortex generated inside becomes relatively weak.

  On the other hand, in FIG. 9, since the intake port cross section (A2-A2) has a substantially triangular shape as shown in FIG. 5, the intake port cross section is widened (in the outward direction of the intake port curvature). A lot of gas flows. FIG. 10 shows an example of the flow rate distribution in each cross section of the port. The center of flow rate is biased to the outside of the curvature of the intake port in the substantially triangular A2-A2 cross section that spreads upward, and this tendency remains in the A3-A3 cross section immediately before the intake valve opening.

  As a result, as shown in FIG. 9, the flow velocity distribution of the gas entering from the valve is parallel to the plane of symmetry CL from the intake valve and is directed toward the exhaust side, and the gas velocity component VC toward the center of the combustion chamber 3C, The gas velocity component VS toward the outside of the cylinder is larger than the gas velocity component VR toward the intake side cylinder wall.

  In the velocity distribution shown in FIG. 9, the increase in the gas velocity components VF and VC has an effect of strengthening the vertical vortex in the combustion chamber 3C. However, the gas velocity component VS toward the outside of the cylinder collides with the wall of the cylinder 3 immediately after entering the combustion chamber from the intake valve opening, and its momentum is attenuated. For this reason, the increase in the gas velocity component VS does not work effectively to strengthen the vertical vortex in the combustion chamber 3C, and there is a possibility that the intake resistance increases due to the wall surface collision.

  The present invention has been made in view of the above points, and does not reduce the amount of intake air even at high rotation and high load, and has sufficient strength in the cylinder regardless of the load and rotation speed of the internal combustion engine. It is an object of the present invention to provide an intake device for an internal combustion engine that forms a vertical vortex and has a low manufacturing cost.

  In order to strengthen the vertical vortex generated in the combustion chamber without reducing the intake air amount, it is desirable that the gas velocity entering the combustion chamber from the intake valve has a distribution as shown in FIG.

  That is, the gas velocity component VF parallel to the plane of symmetry CL from the intake valve 9 and toward the exhaust side and the gas velocity component VC toward the center of the combustion chamber 3C are larger than the gas velocity component VR toward the intake side cylinder wall. To do. Further, the gas velocity components VS1 and VS2 toward the outside of the cylinder are made smaller than the VF and VC. When the axis passing through the center of the intake valve 9 and parallel to the symmetry plane CL is CVY, the gas flow velocity inside (in the symmetry plane CL) of the intake valve opening CVY is outside (symmetric) of the intake valve opening CVY. The gas flow velocity on the side facing the surface CL is made faster than the gas flow velocity. When the axis passing through the center of the intake valve 9 and perpendicular to the plane of symmetry CL is CVX, the gas flow rate entering the combustion chamber from the exhaust side with respect to CVX is greater than the gas flow rate entering the combustion chamber from the intake side with respect to CVX. To be faster.

  As described above, the gas velocity component VF parallel to the plane of symmetry CL from the intake valve and toward the exhaust side coincides with the flow direction of the vertical vortex generated in the combustion chamber 3C. There is a higher effect.

  Further, the gas velocity component VC from the intake valve 9 toward the center of the combustion chamber 3C collides with each other on the symmetry plane CL after entering the combustion chamber 3C, and is a combined velocity component VC ′ parallel to the symmetry plane CL and directed toward the exhaust side. Form. Since the velocity component VC ′ coincides with the flow direction of the longitudinal vortex generated in the combustion chamber 3C, similarly to the velocity component VF, there is an effect of further increasing the strength of the longitudinal vortex. When a plane-symmetric velocity component collides on the symmetry plane CL, both gases move at the same speed along the collision plane on the boundary plane (collision plane). The loss of momentum of VC ′ generated in this way is small.

  On the other hand, since the gas velocity components VS1 and VS2 toward the outside of the cylinder are small, the loss of momentum due to collision with the wall of the cylinder 3 after flowing into the combustion chamber 3C is small. That is, the inflow loss due to the wall collision is kept low.

  As described above, the velocity of the gas entering the combustion chamber from the intake valve in the intake stroke is distributed as shown in FIG. 11, so that the vertical vortex generated in the combustion chamber can be further reduced without reducing the intake air amount. It becomes possible to strengthen.

FIG. 45 shows a comparison result of the longitudinal vortex strength (tumble ratio) in the combustion chamber and the amount of air taken into the combustion chamber by the intake device (FIG. 14) of the internal combustion engine according to the present invention and the conventional intake device (FIG. 5). Show. This result is a result calculated using three-dimensional fluid simulation analysis software (CFD). The engine displacement was 500 cm ³ , the engine speed was 6500 rpm, and the throttle valve was fully open (WOT). As shown in FIG. 45, in the present invention, the longitudinal vortex strength in the combustion chamber can be improved by about 20% without reducing the amount of air sucked into the combustion chamber.

  By improving the longitudinal vortex strength in the combustion chamber, fuel vaporization, fuel and air mixing, combustion speed can be improved, exhaust emissions are reduced, fuel efficiency is improved, and output is improved. Basic performance can be improved.

  FIG. 12 shows an embodiment of the present invention. FIG. 12 is a perspective view of an intake manifold INT, two intake ports 1, 1, 2 intake valves 9, 9, a cylinder head 3 h, and a cylinder 3 in an internal combustion engine having two intake valves and two exhaust valves.

  The intake manifold INT is divided into two intake ports 1 and 1 by a partition wall 1C, and the intake ports 1 and 1 open to the cylinder head 3h. An intake valve 9 is provided in each of the two openings, and the port opening can be opened and closed.

  The exhaust port, exhaust valve, spark plug, fuel injector, and the like are the same as those shown in FIGS. 1 and 2, and are omitted in FIG.

  12, S1, S2,..., S5 indicate cross-sectional shapes of the intake manifold INT and the intake port 1.

  FIG. 13 shows a longitudinal sectional view of one embodiment of the present invention shown in FIG. In FIG. 13, an axis connecting the cross-sectional centers of the intake manifold INT and the intake port 1 is defined as AX, and cross sections perpendicular to AX are defined as A1-A1, A2-A2, and A3-A3. Here, A1-A1 is a downstream section of the intake manifold INT, A2-A2 is an intake port section at a substantially central position between the partition wall 1C and the combustion chamber opening of the intake port 1, and A3-A3 is an opening of the combustion chamber of the intake port 1. It is an intake port cross section just before a part.

  The shape of the piston 6 viewed from above is shown in FIG. 40, and the AA cross section of FIG. 40 is shown in FIG. At the center of the crown surface of the piston 6, there is provided a cavity 6C having a deepest central portion from the intake valve side to the exhaust valve side, a rectangular outer periphery, and an arcuate curved bottom. Both side walls 6 </ b> W of 6 </ b> C are formed as walls higher than the horizontal surface 6 </ b> H of the piston 6.

  Further, the bottom surface of the cavity 6C is a flat surface in the direction along the section AA, and gradually becomes shallower in the direction perpendicular to the section AA as it goes to the intake / exhaust valve side as shown in FIG. It is arcuate. In this cavity 6C, an elongated small step (convex portion) 6S as shown in FIGS. 40 and 41 is formed so as to cross the position of the piston 6 facing the ignition plug 4, and the step 6S is the bottom surface of the cavity 6C. Is connected with a smooth curved surface. The height of the step 6S is, for example, about 2 mm from the bottom surface of the cavity 6C. As shown in FIG. 41, the side wall 6W of the cavity 6S is curved toward the outside, and is connected to the crown surface of the piston 6 and the bottom surface of the cavity 6C. No step other than the step 6S is provided in the cavity 6C. The side wall 6W of the cavity 6C is formed, for example, about 5 mm from the bottom surface of the cavity.

  The spray injected from the fuel injector 5 into the combustion chamber is selected in the injection direction and spray pattern so as to avoid adhesion to the cylinder and the piston wall surface as much as possible and to mix well with air. As the fuel injector 5 that generates spray, a porous nozzle valve, a high-pressure swirl valve, a slit nozzle valve, an outer opening valve, or the like is used.

  An example of the spray shape of the fuel injector 5 in the present embodiment is shown in FIG. FIG. 42 shows a cross-sectional shape of the spray 1 ms after fuel injection in a cross-section BB 30 mm below the tip of the fuel injector nozzle under the conditions of a fuel pressure of 11 MPa and an atmospheric pressure of atmospheric pressure. This injector is a porous nozzle valve, and is composed of six spray beams consisting of SP1 to SP4.

  The spray beam SP1 is directed on the symmetry plane CL and slightly below the electrode position of the spark plug. The spray beam SP1 is directed on the plane of symmetry CL and below the spray beam SP1. The spray beam SP2 is between SP1 and SP2, and is directed outward with respect to the symmetry plane CL. The spray beam SP4 is below SP3 and is directed outward with respect to the symmetry plane CL. Further, SP4 spreads outward than SP2, and as a whole, a substantially triangular spray beam arrangement with SP1, SP4 and SP4 as vertices is provided.

  FIG. 14 shows the shape of the intake port of each cross section of A1-A1, A2-A2, and A3-A3. In the intake manifold (cross-section A1-A1), from the substantially oval cross section to the substantially intermediate position (cross-section A2-A2) of the intake port, it becomes a substantially isosceles triangle shape with the wall surface on the symmetric plane CL side as the bottom, and near the intake port opening ( The cross section A3-A3) has a substantially circular cross section.

  FIG. 15 shows changes in the passage cross-sectional area at the three cross-sectional positions shown in FIG. In addition, about the cross sections A2 and A3 in FIG. 15, the sum total of two intake port cross-sectional areas on both sides of the symmetry plane CL is shown. As shown in FIG. 15, the passage sectional area is kept substantially constant between the sectional positions A2 and A3. In addition, the cross-sectional area of the intake manifold indicated by the A1 position and the cross-sectional area of the intake port indicated by the A2 position are substantially the same (broken line), or the cross-sectional area of the intake port is increased (solid line). The size of is determined. Thus, by keeping the cross-sectional areas of the intake manifold and the intake port constant in the axial direction, an increase in flow velocity in the intake manifold and the intake port can be suppressed, and the flow resistance can be reduced as much as possible. Alternatively, by increasing the cross-sectional area of the intake port from the cross-sectional area of the intake manifold, it is possible to prevent an increase in flow resistance due to drift or vortex that occurs when the intake manifold branches into two intake ports.

  In addition, as shown in FIG. 16, the cross-sectional area of the intake port at the cross-sectional position A2 may be wider than the passage area at other cross-sectional positions. By making the intake port into a substantially isosceles triangular cross section with the base of the symmetry plane at the cross section A2-A2, the flow velocity distribution in the cross section may change greatly with respect to other cross sections, and the flow loss may increase. is there. Therefore, it is possible to reduce this flow loss by expanding the cross-sectional area of the intake port at the cross-sectional position A2 beyond the passage area at the other cross-sectional position.

  FIG. 17 shows the flow rate distribution of the gas in the intake manifold INT and the intake port 1 in the intake stroke. The intake manifold cross section A1-A1 has a parabolic flow distribution, but the intake port A2-A2 cross section has a substantially triangular cross section with the base of the symmetry plane CL as shown in FIG. The inner gas is biased toward the symmetry plane CL, resulting in a flow rate distribution having a maximum value on the symmetry plane CL side with respect to the center of the intake valve 9. Due to the inertial force of the fluid, this distribution is also maintained in the downstream A3-A3 cross section. Therefore, at the opening of the intake valve 9, the gas flow rate entering the combustion chamber from the side of the symmetry plane CL is greater than the valve center, and the gas flow rate entering the combustion chamber from the opposite side of the symmetry plane CL is greater than the valve center. The gas flow rate at the intake valve opening is the flow rate divided by the opening cross-sectional area. Therefore, the gas flow velocity at the intake valve portion in this intake port shape has a distribution as shown in FIG. 11, and for the reason described above, the vertical vortex generated in the combustion chamber can be made stronger without increasing the flow resistance. .

  In the present invention, in order to generate the velocity distribution around the intake valve shown in FIG. 11, a part of the cross section of the intake port has a substantially isosceles triangle shape with the symmetry plane side as the base. As such a port cross-sectional shape, the following shapes are conceivable. Since the two intake port shapes are in a plane-symmetrical relationship with respect to the symmetry plane CL, the following description only describes the features of the intake port shape on the right side with respect to the symmetry plane CL.

  The shape shown in FIG. 18 can be considered as the shape of the A2-A2 cross section in the present invention. In FIG. 18, the tangent of the surface of the port inner wall facing the symmetry plane CL is P1, and the tangent of the port inner wall outside the port curvature is P2. The intake port cross-sectional shape shown in FIG. 18 is characterized in that the intersection angle α between P1 and P2 is an acute angle, that is, narrower than 90 degrees.

  Moreover, the shape shown in FIG. 19 can be considered as the shape of the A2-A2 cross section in the present invention. The width of the intake port cross section is W, and the height of the intake port cross section is H. Further, a rectangular coordinate system XY is defined with the coordinate in the W direction as X, the coordinate in the H direction as Y, and the midpoint of W and the midpoint of H as the origin. The intake port cross-sectional area included in the first quadrant of the XY coordinates is S1, the intake port cross-sectional area included in the second quadrant is S2, the intake port cross-sectional area included in the third quadrant is S3, The intake port cross-sectional area included in the four quadrants is S4. The intake port cross-sectional shape shown in FIG. 19 is characterized in that the total area of S2 and S3 is larger than the total area of S1 and S4, and S1 is equal to S4 or S1 is smaller than S4.

  Moreover, the shape shown in FIG. 20 can be considered as the shape of the A2-A2 cross section in the present invention. The width of the intake port cross section is W, and the height of the intake port cross section is H. Further, a rectangular coordinate system XY is defined with the coordinate in the W direction as X, the coordinate in the H direction as Y, and the midpoint of W and the midpoint of H as the origin. The intake port cross-sectional area included in the first quadrant of the XY coordinates is S1, the intake port cross-sectional area included in the second quadrant is S2, the intake port cross-sectional area included in the third quadrant is S3, The intake port cross-sectional area included in the four quadrants is S4. The intersection of the intake port wall surface on the side facing the symmetry plane CL and the X axis is defined as Cx. The intake port cross-sectional shape shown in FIG. 20 is that the maximum X-coordinate value Cxmax of the intake port wall facing the symmetry plane CL is equal to Cx, and the total cross-sectional area of S2 and S3 is the total cross-sectional area of S1 and S4. It is characterized by being larger than.

  Moreover, the shape shown in FIG. 21 can be considered as the shape of the A2-A2 cross section in the present invention. The width of the intake port cross section is W, and the height of the intake port cross section is H. Further, a rectangular coordinate system XY is defined with the coordinate in the W direction as X, the coordinate in the H direction as Y, and the midpoint of W and the midpoint of H as the origin. In addition, the maximum value of the Y coordinate of the intake port wall surface in the outward direction of the port curvature is defined as Cymax. The intake port cross-sectional shape shown in FIG. 21 is characterized in that the X coordinate of the intake port wall surface is always smaller than Cymax when X> 0.

  In the embodiment of the present invention shown in FIGS. 12 to 14, the cross section of the intake manifold INT has a substantially oval shape as shown in FIG. 14 and the cross section A1-A1. However, as shown in FIGS. 22 and 23, the A1-A1 cross section of the intake manifold INT may be substantially rhombus-shaped. By making the A1-A1 cross section into a substantially rhombus shape in this way, the change in the cross-sectional shape up to the A2-A2 cross section is small, and the gas flow in the intake manifold is more smoothly transferred to the intake port of the A2-A2 cross section. Flowing. As a result, the flow resistance can be further reduced.

  The same effect can be obtained by combining the rhombus cross-sectional shape of the intake manifold with the intake port-shaped cross section of the A2-A2 cross section shown in FIGS.

  FIG. 24 shows a longitudinal sectional shape of the intake port in the embodiment of the present invention. As described above, in the embodiment of the present invention, the A2-A2 cross section has a substantially isosceles triangular shape with the symmetry plane as the base, and the area of the A2-A2 cross section is different as shown in FIGS. It should be equal to or larger than the area of the cross section. For this reason, as shown in FIG. 24, the vertical cross-sectional shape of the intake port has an A2-A2 cross-section that extends in the vertical direction of the intake port, resulting in a substantially barrel-shaped cross-sectional shape. Here, 1U and 1B in FIG. 24 indicate the uppermost and lowermost ridgelines of the intake port having a substantially triangular cross section as shown in FIG. Reference numerals 1U 'and 1B' denote the uppermost and lowermost ridgelines of the intake port having a substantially circular cross section as shown in FIG. By changing the cross-sectional shape of the intake port from the substantially triangular shape to the substantially circular shape from the A2-A2 cross section toward the A3-A3 cross section, the inflection points PB and PU are between 1B and 1B ′ and between 1U and 1U ′. Each can. The tangent vector at the inflection point PB of the curved surface 1B is Vtb, and the tangent vector at the inflection point PU of the curved surface 1U is Vtu.

  In the intake port shape in the present invention, it is desirable that the tangent vector Vtb on the lower surface of the intake port is directed to the opening VO, where VO is an opening on the cylinder center side that is generated when the intake valve lifts.

  In the shape of the intake port in the present invention, if the center coordinate of the bottom surface of the umbrella portion of the intake valve 9 is PVC, the tangent vector Vtu on the top surface of the intake port is directed to the exhaust side (cylinder center side) than the center coordinate PVC of the intake valve. It is desirable that

  In addition, with respect to the above two points, it is possible to generate a stronger vertical vortex by satisfying both, but either one has an effect of strengthening the vertical vortex. Hereinafter, the reason why the vertical vortex is strengthened by this shape will be described.

  In the intake port having the longitudinal cross-sectional shape as shown in FIG. 24, the gas flow along the lower wall surface 1B of the intake port is directed to the opening VO as shown in FIG. Therefore, the combustion chamber 3C is entered through the exhaust side opening VO of the intake valve 9. In addition, the gas flow along the upper wall surface 1U of the intake port has the exhaust valve 9 exhaust gas because the tangent vector Vtu on the upper surface of the intake port is directed to the exhaust side (cylinder center side) from the center coordinate PVC of the intake valve. It enters the combustion chamber 3C through the side opening VO. These flows become a gas flow VT from the intake side toward the exhaust side in the upper part of the combustion chamber 3C. In the upper part of the combustion chamber, since the gas flow VT in the combustion chamber 3C acts in the direction of promoting the vertical vortex TU in the combustion chamber 3C, the intake port having the vertical cross-sectional shape as shown in FIG. 24 should be used. By this, the vertical vortex TU can be made stronger.

  As shown in FIG. 27, the vertical cross-sectional shape of the intake port is a cross-sectional shape such that Rb is smaller than Ru when the radius of curvature of the lower surface of the intake port is Rb and the radius of curvature of the lower surface of the intake port is Ru. You may decide. By making Rb smaller, the flow along the intake port lower surface 1B can be guided to the exhaust-side opening VO of the intake valve. Further, by making Ru larger, the flow along the intake port upper surface 1U can be guided to the exhaust valve side opening VO of the intake valve. These act so that the longitudinal vortex TU in the combustion chamber 3C becomes stronger as described in FIG.

  Further, in order to generate a stronger vertical vortex when the internal combustion engine is at a low rotation speed and a low load, a rectifying plate TP and a vertical vortex control valve TVL may be provided in the intake port as shown in FIGS. The rectifying plate TP is formed of a thin plate that divides the cross section of the intake port 1 vertically. 28 and 29 show an example in which the rectifying plate TP is provided in the substantially central cross section of the intake port, the required longitudinal vortex strength is appropriately adjusted depending on the gas flow rate with respect to the cross-sectional position. That is, when a strong vertical vortex is required at low rotation and low load, the rectifying plate TP may be attached at a position above the central cross-sectional position of the intake port. The vertical vortex control valve TVL is provided upstream of the rectifying plate TP, and closes or opens the lower intake port passage partitioned by the rectifying plate TP by a driving mechanism such as a motor (not shown). When a strong vertical vortex is required at low rotation and low load, the lower intake port passage partitioned by the rectifying plate TP is closed by the vertical vortex control valve TVL and the upper intake vortex partitioned by the rectifying plate TP is closed. A strong vertical vortex TU is generated in the combustion chamber 3C by injecting gas from the intake port. At the time of high rotation and high load, as shown in FIG. 30, the vertical vortex control valve TVL is opened to reduce the flow loss, and gas is introduced into the combustion chamber from the entire intake port cross section. Since the rectifying plate TP has the effect of suppressing the disturbance of the gas flow in the intake port and rectifying in the axial direction of the intake port, even when the vertical vortex control valve TVL is opened, the rectifying plate TP is not provided. This has the effect of strengthening the vertical vortex TU in the combustion chamber. Further, the installation of the rectifying plate TP does not hinder the improvement of the inflow velocity from the exhaust side and the symmetry plane side of the intake port, which is the vertical vortex enhancement mechanism according to the present invention. For this reason, a stronger vertical vortex TU can be generated by combining the intake port shape of the present invention and the current plate TP.

  FIG. 31 shows the spray and gas flow in the combustion chamber 3C in the intake stroke. When performing homogeneous combustion in a direct injection internal combustion engine, normal fuel is injected in the intake stroke. This is to promote the vaporization of the spray and the mixing of the vaporized fuel and air using the gas flow generated during the intake stroke. An appropriate fuel injection timing is set according to the rotational speed and load (fuel injection amount) of the internal combustion engine. For example, in a full load operation of 2000 rpm, fuel is started to be injected from the vicinity of 60 to 90 degrees crank angle after top dead center.

  FIG. 32 shows the spray and gas flow in the intake stroke as seen from above the cylinder. The lift of the intake valve 9 is the maximum (around 100 degrees crank angle after top dead center).

  In the shape of the intake port according to the present invention, in order to improve the inflow speed from the exhaust side and the symmetrical plane side of the intake port and to strengthen the vertical vortex in the combustion chamber, the speed distribution of the inflow part is as shown in FIG. The distribution has a velocity peak on the symmetry plane CL side with respect to the central axis CVY. In order to promote the vaporization of the fuel injected in the intake stroke and to promote the mixing of the air and the fuel sucked into the combustion chamber, it is effective to arrange the spray in a portion where the gas flow is strong. This is due to the following reason.

  Normally, the vaporization rate from the droplets increases as the velocity difference between the droplets and air increases, so that vaporization is promoted if there is a spray at a portion where the gas flow rate is high. As the velocity difference between the droplet and air increases, atomization of the droplet due to the shearing force of air is promoted, and vaporization is promoted when there is a spray in a portion where the gas flow rate is fast. Further, when the gas flow rate is high, mixing by convection of gas and mixing by turbulent vortex are performed, and mixing of air and fuel is promoted.

  As shown in FIG. 32, when the intake valve is at the maximum lift, it is defined by Pv-Pi-Pv where Pv, Pv are the two intersections of the valve center axis CVY and the outer periphery of the intake valve umbrella, and Pi is the spray point. Is defined as α1, and the outermost angle of spray is defined as α2. In order to arrange the spray in the portion where the gas flow generated by the intake port according to the present invention is strong, it is desirable to set the angle of the spray SP injected from the fuel injector 5 so that α2 ≦ α1. As a result, the fuel spray SP is injected toward the gas flow having a high speed entering the combustion chamber from the intake port, and fuel atomization, vaporization, and mixing of air and fuel are further promoted.

  Further, when lean combustion is performed at low load, or when catalyst warm-up is performed by ignition timing delay control immediately after cold start, a part or all of the fuel is injected into the compression stroke as shown in FIG. Is done. In the present embodiment, the fuel is injected at a later stage of the compression stroke, for example, in the vicinity of a crank angle of 40 degrees before the compression top dead center. In this case, since the piston 6 is raised to the cylinder head side, part of the fuel spray is injected into the cavity 6C provided on the crown surface of the piston 6. Further, since there is a step 6S on the piston crown surface, the fuel injected into the cavity 6C is suppressed from being dispersed to the exhaust side in the combustion chamber 3C by the step 6S. Further, due to the step 6S, the vertical vortex TU flowing on the surface of the piston 6 becomes an upward flow at the step 6S, and an upward flow is generated at substantially the center of the combustion chamber 3C. As the compression process further proceeds, the spray in the cavity 6C is vaporized. The fuel FV vaporized from the upflow generated in the vicinity of the center of the combustion chamber 3C caused by the vertical vortex TU of the combustion chamber 3C and the upflow generated in the step 6S due to the penetration force of the spray itself is transferred from the step 6S to the electrode portion of the spark plug 4. Ascend towards. As a result, the concentration of the vaporized fuel around the electrode of the spark plug 4 becomes high, and a so-called stratified mixture is formed. By stratifying the air-fuel mixture in this way, it is possible to reliably ignite the lean air-fuel mixture as a whole, or to perform stable combustion when the ignition timing is retarded immediately after starting the cold engine. Become. As described above, in order to stratify the air-fuel mixture around the spark plug, an upward flow in the center of the combustion chamber generated by the vertical vortex in the combustion chamber is required. The strength of this upward flow is almost proportional to the strength of the vertical vortex. For this reason, it is indispensable to form a sufficiently strong vertical vortex in the combustion chamber even in stratified combustion at low load, and the vertical vortex reinforcement according to the present invention is effective.

  Next, application examples of the present invention to a port injection internal combustion engine will be described with reference to FIGS. In FIG. 33, a fuel injector 5 is provided in the intake manifold INT, and fuel is injected into the intake manifold INT and the intake port 1. The spray angle α3 in the height direction of the fuel spray is determined so that the outermost corner of the fuel spray collides with the upper and lower wall surfaces of the intake port on the downstream side of the intake port 1 (slightly upstream from the intake valve opening). Yes.

  Further, as shown in FIG. 34, the fuel spray SP is injected in two directions toward the two intake valves 9 and 9.

  As shown in FIG. 35, the A2-A2 cross section of the intake port 1 has a substantially isosceles triangular cross section with the plane of symmetry CL as the base, and the A3-A3 cross section has a substantially circular shape.

  The form of fuel spray used in the present embodiment will be described with reference to FIG. FIG. 36 shows the flow rate distribution of the spray in a cross section separated from the nozzle of the fuel injector by Li. Here, Li is the distance from the tip of the nozzle of the injector 5 to the intake valve 9 in a state where the injector 5 is attached to the internal combustion engine as shown in FIG. That is, the spray flow rate distribution shown in FIG. 36 shows the flow rate distribution at the intake valve position in the internal combustion engine. When the spray width is W and the spray height is H, the spray sprayed from the injector is characterized by H> W, that is, a vertically long shape. The center of gravity of the flow rate is a distribution that is biased toward the plane of symmetry from the center of the intake valve.

  When the spray having the above-described characteristics is injected into the intake manifold and the intake port, a spray form as shown in FIG. 37 is obtained in the A2-A2 cross section. That is, since the spray center of gravity is biased toward the symmetry plane CL and the spray has a vertically long shape, the spray SP is concentrated near the bottom of the triangle in the intake port 1 of the A2-A2 cross section having a substantially triangular cross section. For this reason, even if it is the vertically long spray SP, it is possible to prevent the spray SP from colliding with the upper wall surface 1U and the lower wall surface 1B of the intake port 1.

  FIG. 38 is a view of the relationship between the spray form and the gas flow velocity distribution in the intake valve section from above the internal combustion engine. Since the spray SP injected from the fuel injector 5 is biased toward the plane of symmetry, it collides inside the two intake valves. As described above, the gas flow rate on the inner side (symmetrical surface side) of the intake valve is increased by the cross-sectional shape of the intake port. FIG. 39 shows the relationship between the liquid film form of the fuel and the gas velocity distribution in the combustion chamber inflow portion from the exhaust side of the internal combustion engine. Since the spray center of gravity has a distribution that is biased toward the symmetry plane CL, most of the fuel spray collides with the inside (symmetry plane) of the intake valve 9 and forms a fuel liquid film LF at the collision position. By making the spray vertically long, the area where the spray adheres to the wall surface increases, and the liquid film generated on the intake valve becomes thin. When a thin fuel liquid film is formed over a wide area, the heat transfer area between the liquid film and the gas and the wall surface, the evaporation area between the liquid film and the gas increase, and heat transfer to the liquid film by heat conduction is promoted. , Vaporization of the liquid film is promoted.

  As shown in FIG. 39, since the gas velocity distribution in the inflow portion is a distribution in which the peak is biased toward the symmetry plane, the portion of the combustion chamber inflow portion where the gas velocity is high flows on the surface of the fuel liquid film LF. The vaporization speed of the fuel liquid film increases as the difference between the speed of the fuel liquid film and the gas velocity flowing on the surface thereof increases. Therefore, a portion where the gas velocity at the inflow portion of the combustion chamber flows at the surface of the fuel liquid film LF The vaporization of the fuel liquid film LF is promoted. In addition, since the difference between the speed of the fuel liquid film and the speed of the gas flowing on the surface is large, atomization of the liquid film due to the shearing force of the gas is promoted. Further, the fuel vaporized on the intake valve 9 quickly flows into the combustion chamber 3C by the high-speed gas flow inside the intake valve, and is well mixed with air.

The effects of the present embodiment in the port injection internal combustion engine are summarized above. By making the A2-A2 cross section of the intake port into a substantially isosceles triangular cross section with the symmetry plane as the base, and making the shape of the fuel spray injected into the intake port into a vertically elongated and symmetrical distribution, The following effects are obtained.
(1) The spray can be wide-angled while preventing the spray interference to the port wall surface, and a thin liquid film can be formed. Thereby, the vaporization of the liquid film can be promoted.
(2) By forming a liquid film on the symmetrical surface side of the intake valve having a high gas flow rate, it is possible to promote the vaporization of the liquid film and to promote the mixing of fuel and air in the combustion chamber.

1 is a schematic configuration diagram (plan view) of a four-valve internal combustion engine. 1 is a schematic configuration diagram (longitudinal sectional view) of a four-valve internal combustion engine. The conventional intake port block diagram. A conventional intake port sectional view. A conventional intake port sectional view. The top view which shows the gas flow by the conventional intake port. The longitudinal cross-sectional view which shows the gas flow by the conventional intake port. The top view which shows the gas flow by the conventional intake port. The top view which shows the gas flow by the conventional intake port. The flow distribution diagram of the conventional intake port cross section. The top view which shows the gas flow by the intake port of this invention. The perspective view which shows the intake port shape of this invention (Example 1). The intake port longitudinal cross-sectional view of this invention. The intake port sectional view of the present invention (example 1). 1 is a cross-sectional area change example 1 of the intake port of the present invention. Fig. 3 is a cross-sectional area change example 2 of the intake port of the present invention. Flow distribution in the intake port according to the invention. The intake port cross-sectional shape in the A2-A2 cross section of the present invention (Example 1). The intake port cross-sectional shape in the A2-A2 cross section of the present invention (Example 2). The intake port cross-sectional shape in the A2-A2 cross section of the present invention (Example 3). The intake port cross-sectional shape in the A2-A2 cross section of the present invention (Example 4). The perspective view which shows the intake port shape of this invention (Example 2). The intake port sectional view of the present invention (example 2). The intake port longitudinal cross-sectional view of this invention. The intake port sectional view of the present invention. The longitudinal cross-sectional view which shows the gas flow by the intake port of this invention. The intake port longitudinal cross-sectional view of this invention. The perspective view which shows the intake port shape of this invention (Example 3). The longitudinal cross-sectional view which shows the gas flow by the intake port of this invention (when a vertical vortex control valve is closed). The longitudinal cross-sectional view which shows the gas flow by the intake port of this invention (when a vertical vortex control valve is opened). The longitudinal cross-sectional view which shows the spray behavior and gas flow at the time of injecting fuel in an intake stroke. The top view which shows the spraying behavior at the time of injecting fuel in an intake stroke, and gas flow. The longitudinal cross-sectional view which shows the structural example of the port injection type engine by this invention. The top view which shows the structural example of the port injection type engine by this invention. The intake port sectional view of the present invention. The figure which shows the spraying shape of the port injection injector by this invention. Sectional drawing which shows the spraying form in the intake port by this invention. The top view which shows the fuel behavior and gas flow velocity distribution of the port injection type engine by this invention. The top view which looked at the inside of a cylinder from the exhaust side which shows the fuel liquid film distribution and gas flow velocity distribution of the port injection type engine by this invention. The top view of the piston shown in the Example of this invention. The longitudinal cross-sectional view of the piston shown in the Example of this invention. The spray shape of the fuel injector shown in the Example of this invention. The longitudinal cross-sectional view which shows the spray behavior and gas flow at the time of injecting fuel in the latter half of a compression stroke. The longitudinal cross-sectional view which shows the vaporization fuel and gas flow in the latter half of a compression stroke. Comparison of longitudinal vortex strength and intake air volume according to the present invention and the prior art.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Intake port 1C Inlet port partition wall 1B Intake port recessed part lower surface 1B 'Intake port straight part lower surface 1U Intake port convex part upper surface 1U' Intake port straight part upper surface 2 Exhaust port 2C Exhaust port partition wall 3 Cylinder 3C Combustion chamber 3h Cylinder head 4 Spark plug 5 Fuel injector 6 Piston 6C Cavity 6S Piston crown surface 6S Step 6W provided on piston crown surface Side wall of cavity 9 provided on piston crown surface 9 Intake valve 10 Exhaust valve AX Intake port center axis CC Cylinder center axis CL Symmetry plane CVY Intake valve central axis INT parallel to symmetry plane Intake manifold EXH Exhaust manifold FV Vaporized fuel H Spray height LF Fuel liquid film SP Fuel spray TP Inlet port cross section TU Vertical vortex TVL Vertical vortex control valve W Spray width

Claims (13)

In the intake device having an intake passage of the internal combustion engine that is divided into two intake ports from the intake manifold, a cross section from the branch start position of the intake port to the outlet of the intake port is
2. An intake device for an internal combustion engine, characterized in that the two intake ports have a substantially isosceles triangle shape with a bottom surface as a facing surface. In any cross-sectional position from the intake port branch start position to the intake port outlet of the intake passage of the internal combustion engine that is divided into two symmetrical intake ports from the intake manifold,
The opposite surfaces of the two intake ports are parallel to the plane of symmetry, and the intersection angle between the tangent to the outer surface of the intake port curvature and the tangent to the opposite surface of the intake port is smaller than 90 degrees. An internal combustion engine intake system. In any cross-sectional position from the intake port branch start position to the intake port outlet of the intake passage of the internal combustion engine that is divided into two symmetrical intake ports from the intake manifold,
The port in the first quadrant in the XY coordinate system with the origin of the midpoint of the width direction of the one-side intake port and the midpoint of the height direction, where the width direction of the one-side intake port is the X-axis and the height direction is the Y-axis S2 + S3> S1 + S4 and S1 ≦ S4, where S1 is the cross-sectional area, S2 is the port cross-sectional area of the second quadrant, S2 is the port cross-sectional area of the third quadrant, and S4 is the port cross-sectional area of the fourth quadrant. An intake device for an internal combustion engine characterized by the above. In any cross-sectional position from the intake port branch start position to the intake port outlet of the intake passage of the internal combustion engine that is divided into two symmetrical intake ports from the intake manifold,
A port in the first quadrant in the XY coordinate system with the origin of the midpoint of the width direction of the one-side intake port and the midpoint of the height direction, where the width direction of the one-side intake port is the X-axis and the height direction is the Y-axis When the cross-sectional area is S1, the port cross-sectional area of the second quadrant is S2, the port cross-sectional area of the third quadrant is S3, the port cross-sectional area of the fourth quadrant is S4, and the intersection of the X axis and the port cross-section outside is Cx An intake device for an internal combustion engine, wherein the outermost X coordinate of the port is equal to Cx, and S2 + S3> S1 + S4. In any cross-sectional position from the intake port branch start position to the intake port outlet of the intake passage of the internal combustion engine that is divided into two symmetrical intake ports from the intake manifold,
In the XY coordinate system, where the width direction of the one-side intake port is the X axis and the height direction is the Y-axis, An intake system for an internal combustion engine, wherein when the height is Cymax, the Y coordinate of the intake port cross section is always Y <Cymax when X ≧ 0.   The intake device for an internal combustion engine according to any one of claims 1 to 5, wherein a ridge line on the lowermost surface of the inner surface of the port has a downwardly convex curvature.   The bottom surface ridgeline of the inner surface of the port has a downwardly convex curvature B, and the bottom surface of the port inner surface has a substantially straight line B ′. The intake device for an internal combustion engine according to claim 6, wherein the vector V1 is directed to a valve opening portion on a cylinder center side during intake valve lift.   The ridgeline of the uppermost surface of the inner surface of the port is composed of a portion U having a convex curvature, and the ridgeline of the uppermost surface of the inner surface of the port is composed of a substantially straight portion U ′, and the tangential direction of U at the joint between U and U ′ The intake device for an internal combustion engine according to claim 6, wherein the vector V2 is directed to the cylinder center side from the intake valve center when the intake valve is fully lifted. The bottom surface ridgeline of the inner surface of the port has a downwardly convex curvature B, and the bottom surface of the port inner surface has a substantially straight line B ′. The vector V1 faces the valve opening on the cylinder center side during intake valve lift,
And,
The ridgeline of the uppermost surface of the inner surface of the port is composed of a portion U having a convex curvature, and the ridgeline of the uppermost surface of the inner surface of the port is composed of a substantially straight portion U ′, and the tangential direction of U at the joint between U and U ′ The intake device for an internal combustion engine according to claim 6, wherein the vector V2 is directed to the cylinder center side of the intake valve center when the intake valve is at the maximum lift.   6. The intake device for an internal combustion engine according to claim 1, wherein a curvature radius Rb of the lower surface of the inner surface of the port and a curvature radius Ru of the upper surface satisfy Rb <Ru. In a direct injection type internal combustion engine in which fuel is directly injected into a combustion chamber by a fuel injector provided between two intake ports,
Using the intake device according to any one of claims 1 to 10,
When the intake valve is at its maximum lift, when the intersection of the valve center axis CVY and the outer periphery of the intake valve umbrella is Pv, Pv, the spray point is Pi, and the angle defined by Pv-Pi-Pv is α1 And an in-cylinder direct injection internal combustion engine characterized in that an outermost angle of spray injected from the fuel injector is α1 or less. In a port injection type internal combustion engine in which fuel is injected into an intake port by a fuel injector provided in an intake pipe,
Using the intake device according to any one of claims 1 to 10,
The port injection type internal combustion engine, wherein the spray injected from the fuel injector has a flow rate center of gravity on a symmetrical plane side with respect to the center of the intake valve.   The port injection type internal combustion engine according to claim 12, wherein the height of the spray injected from the fuel injector is larger than the width of the spray.

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