JP2010249111A - Internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To promote combustion by intensifying the flow of gas in a cylinder without providing a gas flow control valve. <P>SOLUTION: To a crank shaft 4 and an output shaft 11, a pair of elliptical gears 12, 13 are journaled, respectively, which engage each other. A crank angle &theta; is set between a minor axis S1 of the crank shaft side elliptical gear 12 and a major axis L2 of the output side elliptical gear 13 engaging each other so that a crank shaft 4 is rotated at the maximum speed in a valve opening period of an intake valve 8. <P>COPYRIGHT: (C)2011,JPO&amp;INPIT

Description

  The present invention relates to an internal combustion engine in which the descending speed of a piston is set to the highest speed in synchronization with the opening timing of an intake valve.

  2. Description of the Related Art Conventionally, a technique for improving combustion by enhancing gas flow generated in a cylinder is known. As means for enhancing gas flow, for example, in Patent Document 1 (Japanese Patent Laid-Open No. 7-119472), an intake passage upstream of an intake port is divided into two by a partition, and a gas flow control valve is provided in one divided passage. And by closing this gas flow control valve, a technique of generating a strong gas flow (tumble flow) in the cylinder by causing the intake air to flow into the cylinder at high speed from the other divided passage. It is disclosed.

  According to this technique, a gas flow such as a longitudinal vortex flow (tumble flow) or a lateral vortex flow (swirl flow) is generated in the intake air supplied into the cylinder, whereby the combustion speed is advanced and combustion is improved.

  However, as in the technique disclosed in the above-mentioned document, in order to generate gas flow in the cylinder, a control valve (tumble control valve (TGV) or swirl control valve as a gas flow control valve is provided in the intake pipe. (SCV)), there is a problem that the structure of the exhaust pipe becomes complicated.

  Further, in order to generate the gas flow, one passage is closed to increase the intake air flow velocity. Therefore, there is a problem in that the pump loss increases and the fuel consumption deteriorates as the intake air is throttled.

  Furthermore, during low load operation, the flow rate of intake air flowing through the intake passage is small, so that it is difficult to generate gas flow, and it is difficult to improve combustion by this gas flow.

  In view of the above circumstances, the present invention can enhance the gas flow in the cylinder without providing a gas flow control valve, thereby promoting combustion, improving the isovolume, and avoiding knocking and achieving high compression. It is another object of the present invention to provide an internal combustion engine that can achieve a reduction in the ratio and can reduce pump loss and simplify the structure by eliminating the need for a gas flow control valve.

  In order to achieve the above object, the present invention provides an internal combustion engine comprising a crankshaft that is rotated by a vertical movement of a piston and an output shaft that is rotated by a rotational movement from the crankshaft, and is provided between the crankshaft and the output shaft. When the output shaft rotates at a substantially constant speed, the crankshaft is provided with an inconstant speed power transmission mechanism in which the rotation speed is periodically changed in synchronization with the rotation period, and the inconstant speed power transmission is performed. The mechanism is characterized in that the rotational speed of the crankshaft becomes the fastest during the opening period of the intake valve.

  According to the present invention, an inconstant speed power transmission mechanism is interposed between the crankshaft that rotates by the vertical movement of the piston and the output shaft, and the rotational speed of the crankshaft becomes the fastest during the valve opening period of the intake valve. Thus, the gas flow in the cylinder during the intake stroke is enhanced, and combustion can be promoted. As a result, not only the equal volume is improved, but also a high compression ratio can be achieved by avoiding knocking. Furthermore, since the gas flow control valve is not required, the structure can be simplified and the pump loss can be reduced.

Schematic configuration diagram of a gear insertion type internal combustion engine when the piston according to the first embodiment is at exhaust top dead center Schematic configuration diagram of gear insertion type internal combustion engine when intake valve is opened (A) is a timing chart showing a change in in-cylinder pressure, (b) is a timing chart showing an operation of an intake valve, and (c) is a timing chart showing a change in rotational speed of the crankshaft side elliptical gear. Same characteristic diagram showing the relationship between gear phase and intake air average inflow rate increase rate Schematic configuration diagram of a gear insertion type internal combustion engine when the piston according to the second embodiment is at exhaust top dead center Schematic configuration diagram of gear insertion type internal combustion engine when intake valve is opened

  Hereinafter, an embodiment of the present invention will be described with reference to the drawings.

[First Embodiment]
1 to 4 show a first embodiment of the present invention. Reference numeral 1 in FIGS. 1 and 2 denotes a gear insertion type internal combustion engine (hereinafter abbreviated as “internal combustion engine”), and a spark ignition type gasoline combustion engine is shown in the drawings. The internal combustion engine 1 shown in the present embodiment is a four-cycle single-cylinder internal combustion engine, but can be applied to a four-cycle two-cylinder or four-cylinder internal combustion engine.

  A piston 3 is inserted into a cylinder 2 of the internal combustion engine 1 so as to be able to advance and retreat, and a piston pin 3 a of the piston 3 and a crank pin 4 a of the crankshaft 4 are connected via a connecting rod (hereinafter abbreviated as “connecting rod”) 5. Are connected.

  The crankshaft 4 is rotatably supported by a crankcase (not shown) whose crank journal 4 b is connected to the cylinder 2. A combustion chamber 7 is formed in a region defined by the top surface of the piston 3, the cylinder 2 and the cylinder head 6. The cylinder head 6 is provided with an intake port 6a and an exhaust port 6b. An intake valve 8 and an exhaust valve 9 are provided in each of the ports 6a and 6b, and an ignition portion is provided at the top of the cylinder head 6. A spark plug (not shown) that faces the combustion chamber 7 is fixed.

  As shown in FIG. 3B, the valve opening period of the intake valve 8 is set from just before the end of the exhaust stroke to immediately after the start of the compression stroke, and although not shown, the valve opening period of the exhaust valve 9 is the combustion period. It is set immediately before the end of the stroke and immediately after the start of the intake stroke. The valves 8 and 9 are opened / closed by, for example, the operation of an intake cam and an exhaust cam provided on a camshaft that rotates at a half rotation speed in synchronization with the rotation speed of the crankshaft 4. However, each valve 8.9 may be an electromagnetic valve whose opening and closing is electronically controlled.

  Further, the output shaft 11 is disposed in parallel with the crankshaft 4 along the extending direction of the crankshaft 4. Further, a line connecting the axis of the crank journal 4b of the crankshaft 4 and the axis of the output shaft 11 connects the axis of the crank journal 4b and the axis of the crank pin 4a when the piston 3 is at the top dead center. They are arranged in a direction perpendicular to the connecting line (see FIG. 1). Load elements such as a generator and a traveling load are applied to the output shaft 11 via a transmission (not shown) or directly. A flywheel (not shown) is attached to the output shaft 11 so as to rotate the output shaft at a substantially constant speed by inertia force.

  Further, a pair of crankshaft side non-circular gears and output shaft side noncircular gears that mesh with each other are interposed between the crankshaft 4 and the output shaft 11. These non-circular gears are elliptical gears 12 and 13 in the present embodiment. The crankshaft 4 has a crankshaft-side elliptical gear 12 and the output shaft has an output shaft-side elliptical gear 13 with axial core portions. It is attached to the shaft. The elliptical gears 12 and 13 constitute an inconstant speed power transmission mechanism. In this inconstant speed power transmission mechanism, when the output shaft 11 rotates at substantially constant speed, the rotation speed of the crankshaft 4 periodically changes in synchronization with the rotation period.

  The elliptical gears 12 and 13 have the same shape, and the elliptical gears 12 and 13 are meshed with the major axes L1 and L2 and the minor axes S1 and S2 shifted in phase in a direction perpendicular to each other. ing. In this embodiment, as the elliptical gears 12 and 13, a double-leaf gear having a point-symmetrical gear profile with respect to the shaft center is adopted, and the maximum angular velocity ratio of the elliptical gears 12 and 13 is set to 2.0. ing. However, this maximum angular velocity ratio is appropriately set according to the piston velocity change characteristic of the internal combustion engine 1 to be employed.

  When the piston 3 is guided by the cylinder 2 and moves up and down, the crankshaft 4 is rotated, and the rotation of the crankshaft 4 causes the output shaft through the crankshaft side elliptical gear 12 and the output shaft side elliptical gear 13 to mesh with each other. 11 rotates. At this time, when the output shaft 11 is rotated at a constant speed, the crankshaft 4 rotates at a constant speed of 180 [deg] by the rotation of the pair of oval gears 12 and 13 meshing with each other.

  FIG. 1 shows a state in which the piston 3 is at the exhaust top dead center (TDC). The symbol LTDC is an axis connecting the axis of the crank journal 4b of the crankshaft 4 and the axis of the crank pin 4a when the piston 3 is at the top dead center TDC. Further, as indicated by an arrow in the figure, the crankshaft 4 is set to rotate in the clockwise direction, and thus the output shaft 11 rotates in the counterclockwise direction.

  The crank angle at which the rotational speed of the crankshaft 4 becomes the highest when the output shaft 11 is rotated at a constant speed is a line connecting the shaft center of the crank journal 4b and the shaft center of the output shaft 11, as shown in FIG. In this state, the short axis S1 side of the crankshaft side elliptical gear 12 and the long axis L2 side of the output shaft side elliptical gear 13 are in this position. Meshed. Therefore, by synchronizing the position at which the short shaft S1 side of the crankshaft side elliptical gear 12 and the long axis L2 side of the output shaft side elliptical gear 13 are engaged with the crank angle of the piston 3, the piston 3 can move up and down. The moving speed can be the fastest during the intake stroke.

  By the way, by temporarily increasing the descending speed of the piston 3 in accordance with the opening timing of the intake valve 8, it is possible to perform intake in a short time. By completing the intake in a short time, the intake flow velocity during the intake stroke is relatively increased, and the gas flow such as a tumble flow or a swirl flow generated in the cylinder can be strengthened.

  As shown in FIG. 4, according to the experiment, the crank angle (gear phase) θ at which the descending speed of the piston 3 becomes the fastest is set to the retarded direction with respect to the exhaust top dead center, that is, after the exhaust top dead center ( ATDC) The gas flow in the cylinder increases by 10 [%] or more by setting between 25 and 80 [deg]. Among them, by setting the crank angle θ to about 50 [deg] ATDC, It has been found that the gas flow is maximized. In the present embodiment, since the pair of elliptical gears 12 and 13 is a double-leaf gear, the rotational fluctuation of the crankshaft side elliptical gear 12 (crankshaft 4) changes at a cycle of 180 [deg]. Therefore, the crank angle θ that is the fastest appears for each stroke.

  According to the results of this experiment, by making the lowering speed of the piston 3 the fastest in the first half of the intake stroke, the intake flow velocity in the first half of the intake stroke is increased, and a strong gas flow can be generated in the cylinder. Promotion can be aimed at.

  Next, the operation of the present embodiment having such a configuration will be described. In the operating internal combustion engine 1, the crankshaft 4 rotates with the vertical movement of the piston 3, and the rotational force is attached to the output shaft 11 via the crankshaft-side elliptical gear 12. The output shaft 11 is rotated by being transmitted to the elliptical gear 13. A load from a load element such as a traveling load or a generator is applied to the output shaft 11 via a transmission or the like.

  When the piston 3 reaches the end of the exhaust stroke, the intake valve 8 opens, and when the piston 3 moves down to the intake stroke and the piston 3 descends, the intake air flows into the cylinder from the intake port 6a. As shown in FIG. 3C, the crankshaft side elliptical gear 12 that is pivotally attached to the crankshaft 4 is from the long axis L <b> 1 side with respect to the output shaft side elliptical gear 13 that is pivotally attached to the output shaft 11. Since the meshing position gradually moves toward the short axis S1, the descending speed of the piston 3 gradually increases. Then, at the position where the short shaft S1 side of the crankshaft side elliptical gear 12 meshes with the long axis L2 of the output shaft side elliptical gear 13, the descending speed of the piston 3 becomes the fastest.

  In the present embodiment, the crank angle θ at which the rotational speed of the crankshaft side elliptical gear 12 (crankshaft 4) is the fastest is set between exhaust top dead center (ATDC) 25-80 [deg], preferably Since the ATDC is set to about 50 [deg], the rotational speed of the crankshaft side elliptical gear 12 (the descending speed of the piston 3) increases immediately before the start of the intake stroke. As a result, the flow velocity of the intake air flowing into the cylinder immediately after the transition to the intake stroke is increased, and a strong gas flow is generated, so that combustion can be promoted.

  Thus, according to this embodiment, combustion can be promoted without using a gas flow control valve. Further, since the gas flow control valve is not required, the intake resistance when generating the gas flow is small, the pump loss can be relatively reduced, and the structure can be simplified. Moreover, gas flow can be generated even during low-load operation where the intake flow rate is small and combustion tends to become unstable, and combustion during low-load operation can be stabilized.

  By the way, as shown in FIG. 3C, when the crank angle θ at which the crankshaft-side elliptical gear 12 is the fastest is set between ATDC25 and 80 [deg], the crank angle θ is set to the crank angle θ even in the combustion period of the combustion stroke. Until it reaches, the descending speed of the piston 3 is gradually increased, so that the occurrence of knocking is suppressed and the equal volume is improved. Therefore, the ignition timing can be relatively advanced, and the compression ratio can be set high. As a result, the engine output can be further increased and the fuel consumption can be improved.

[Second Embodiment]
5 and 6 show a second embodiment of the present invention. In the first embodiment described above, the line connecting the axis of the crank journal 4b and the axis of the output shaft 11 is the line between the axis of the crank journal 4b and the axis of the crank pin 4a when the piston 3 is at top dead center. In this embodiment, when the piston 3 is at the top dead center, the line connecting the axis of the crank journal 4b and the axis of the output shaft 11 and the crank journal 4b are arranged. And the crankshaft 4 and the output shaft 11 are arranged so that a line connecting the shaft centers of the crankpin 4a and the crankpin 4a is in a straight line.

  Since the elliptical gears 12 and 13 and the meshing relationship between the elliptical gears 12 and 13 are the same as those in the first embodiment, the crankshaft side elliptical gear 12 (when the output shaft 11 is rotated at a constant speed) ( The change characteristics of the rotational speed of the crankshaft 4) are the same as those in FIG.

  As shown in FIG. 6, in this embodiment, on the line connecting the shaft center of the crank journal 4b and the output shaft 11, that is, the shaft center of the output shaft 11 and the shaft center of the piston pin 3a are connected. On the axis LTDC, the short shaft S1 of the crankshaft side elliptical gear 12 and the long shaft L2 of the output shaft side elliptical gear 13 are arranged to coincide with each other, and the crankshaft 4 when the output shaft 11 is rotated at a constant speed in this state. The crank angle θ is set so that is the fastest.

  The fastest crank angle θ is an angle between the above-mentioned axis LTDC and a line connecting the axis of the crank journal 4b and the axis of the crank pin 4a. This angle is as shown in FIG. The same applies to the state in which the short axis S1 of the crankshaft side elliptical gear 12 is advanced by the crank angle θ when the piston 3 is at the exhaust top dead center. Thus, when the short axis S1 is used as a reference, the axis of the crank pin 4a is disposed at a position delayed by the crank angle θ.

  The crank angle θ is set between ATDC 25 and 80 [deg], preferably ATDC 50 [deg], as in the first embodiment described above. In addition, the effect of this embodiment is the same as that of 1st Embodiment mentioned above.

  Further, the present invention is not limited to the above-described embodiment. For example, the applied internal combustion engine may be a diesel combustion engine.

1 ... an internal combustion engine,
3 ... Piston,
4 ... crankshaft,
4a ... crankpin,
4b ... Crank journal,
6 ... Cylinder head,
6a ... intake port,
7 ... Combustion chamber,
8 ... Intake valve,
11 ... Output shaft,
12 ... Crankshaft side elliptical gear,
13 ... Output shaft side elliptical gear θ ... Crank angle L1, L2 ... Long axis S1, S2 ... Short axis

Japanese Unexamined Patent Publication No. 7-119472

Claims (7)

In an internal combustion engine comprising a crankshaft rotating by a vertical movement of a piston and an output shaft rotating by a rotational movement from the crankshaft,
An inconstant speed power transmission mechanism in which when the output shaft rotates at a substantially constant speed between the crank shaft and the output shaft, the rotation speed of the crank shaft is periodically changed in synchronization with the rotation cycle. Intervening,
The internal combustion engine is characterized in that the non-constant speed power transmission mechanism is arranged so that the rotational speed of the crankshaft becomes the fastest within a valve opening period of the intake valve. The inconstant speed power transmission mechanism comprises a crankshaft side non-circular gear and an output shaft side noncircular gear that mesh with each other,
The internal combustion engine according to claim 1, wherein each of the non-circular gears is provided on the crankshaft and the output shaft. The internal combustion engine according to claim 2, wherein each non-circular gear has a point-symmetrical gear profile with respect to the axis of each non-circular gear. The output shaft is provided with a flywheel,
The crankshaft side non-circular gear and the output shaft side non-rotating gear are arranged so that the short shaft side of the crankshaft side noncircular gear meshes with the long shaft side of the output shaft side noncircular gear during the valve opening period of the intake valve. 4. An internal combustion engine according to claim 2, wherein an arrangement of circular gears is set. The internal combustion engine is a spark ignition gasoline combustion engine, and the short shaft side of the crankshaft-side noncircular gear and the long shaft side of the output shaft-side noncircular gear are engaged with each other during the opening period of the intake valve. 5. The internal combustion engine according to claim 4, wherein the crank angle is set between 25 and 80 degrees in the retarding direction with reference to when the piston is at an exhaust top dead center. The short axis side of the crankshaft side non-circular gear is arranged so as to coincide with a line connecting the axis of the crankshaft and the axis of the output shaft during the valve opening period of the intake valve. The internal combustion engine according to claim 4, characterized in that: With respect to the time when the piston is at exhaust top dead center, the short shaft side of the crankshaft-side non-circular gear is retarded with respect to the line connecting the crankshaft axis and the output shaft axis. 7. The internal combustion engine according to claim 6, wherein the internal combustion engine is set between 25 and 80 degrees in the direction.

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