JP2791571B2 - Cooling system for rotor of two cylinder rotary engine - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a cooling device for a rotor of a rotor oil-cooled two-cylinder Wankel rotary engine.

[Conventional technology]

As shown in FIG. 5, a conventional two-cylinder rotary engine includes two rotor housings 1a and 1b having a trochoid-shaped inner peripheral surface (not shown), an intermediate side housing 2 located therebetween, and two Two trochoid-like spaces are formed by the two side housings 3a and 3b (the entire casing composed of these housing members is not shown). The two side housings 3a, 3b include:
A main bearing housing 6 having an external gear 4 on the outer periphery and a main bearing metal 5 mounted on the inner periphery is fixed to form main bearings 7a and 7b. Main shaft portions 9a and 9b at both ends of the eccentric shaft 8 are rotatably supported by the two main bearings. The eccentric shaft has two eccentrics 10
a and 10b are mutually 180 degrees in the eccentric shaft rotation direction.
It is formed with a stagger. The rotors 11a and 11b are rotatably supported by the two eccentric shaft portions by rotor bearings 14a and 14b made of a rotor bearing boss 12 and a rotor bearing metal 13 mounted on the inner periphery thereof. , 11b meshes with the external gear 4 to change the rotational phase of the rotors 11a, 11b in the trochoid-like space, with the tops of the rotors slidingly contacting the trochoid-like inner peripheral surface. Regulate to rotate. Eccentric shaft 8
A counterweight 24 is fixed to the front, a flywheel 25 is fixed to the rear end, and a counterweight 26 is formed on the outer periphery of the flywheel. An axial lubricating oil passage 16 having both ends closed is formed inside the eccentric shaft 8, and the axial lubricating oil passage has a main-shaft branch passage 17 a, 17.
b, the eccentric shaft branch passages 18a and 18b intersect. In the vicinity of the eccentric shaft portions 10a, 10b of the eccentric shaft 8, the main shaft portions 9a,
On the side opposite to 9b, the spout branch path 2 to spouts 19a, 19b
A ball valve 21 is pressed by a spring 22 to close 0a and 20b, and check valves 23a and 23b are configured. The gear 27 attached to the front of the eccentric shaft 8 is engaged with the gear 29 attached to the shaft of the oil pump 28 to drive the oil pump 28. The oil pump pumps and pumps lubricating oil in the oil pan 30. A part of the lubricating oil to be pumped is released by a hydraulic control valve 31, which is a relief valve, and is controlled to an appropriate oil pressure for lubricating various parts of the engine. The side housing 3a, the rotor housing 1a, the intermediate side housing 2, Oil gallery formed through rotor housing 1b and side housing 3b
Introduced in 32. From the oil gallery, it is supplied to front and rear main bearings 7a and 7b through lubricating oil passages 33a and 33b formed in front and rear side housings 3a and 3b, and is supplied to main shaft branch passages 17a and 17b in the eccentric shaft 8. The lubricating oil passing through the axial lubricating oil passage 16 passes through the eccentric shaft branch passages 18a, 18b to the rotor bearings 14a, 14b, and the other passes through the check valves 23a, 23b, and the jet port 19a. , 19b to the rotor hollow portion 34 to cool the rotors 11a, 11b.

In this conventional technique, the opening of the check valves 23a, 23b, that is, the ejection of the lubricating oil from the ejection ports 19a, 19b is performed by the hydraulic pressure determined by the discharge amount of the oil pump 28 and the centrifugal force acting on the ball valve body 21. That is, the load is not related at all by the engine speed alone. Therefore, the rotor 11a at the time of high load,
The check valves 23a and 23b are set to open regardless of the load from the engine speed at which the cooling of 11b is required, and the injection ports 19a and 19b are connected to the rotor 1 at the maximum output.
The cross-sectional area of the flow passage is set so as to secure a sufficient amount of jet for cooling 1a and 11b. Therefore, this technology aims at preventing the rotors 11a and 11b from being overcooled and securing the oil pressure for lubrication of each bearing in the vicinity of idling rotation. Overcooling is inevitable.
In general vehicle engines that are mostly throttled, this leads to poor fuel economy and an increase in exhausted unburned HC.

In recent years, the output of rotary engines has also been increasing, and the amount of heat generated has increased.Therefore, in order to keep the rotor constituting the combustion chamber wall at an appropriate temperature, the amount of lubricating oil injected to cool the rotor has increased. There must be. In that case, the adverse effect of supercooling at a low load is further increased.

To cope with the above problems, Japanese Patent Application Laid-Open No. 63-85223
The technique described in Japanese Patent Application Laid-Open Publication No. H10-26095 has been proposed. This is because, in the two-cylinder rotary engine, the hydraulic control valve 31 is made variable, the oil pressure of the lubricating oil pumped by the oil pump 28 is reduced at a low load, and the ejection of the lubricating oil for cooling the rotors 11a and 11b is reduced. Alternatively, it stops.

[Problems to be solved by the invention]

The rotary engine has no reciprocating mass, and in order to eliminate the unbalanced inertial force of the moving part, it is sufficient to consider only the balancing of the rotating mass, and in the case of the two-cylinder rotary engine, as shown in FIG. The eccentric parts are arranged 180 degrees apart, and a counterweight 2 for this eccentric mass that applies centrifugal force to the center of the eccentric shaft 8
4,26 are arranged and the inertia forces of the moving parts are perfectly balanced.

Therefore, the load applied to the main bearings 7a, 7b is basically only the central component of the combustion gas pressure received by the rotor, and when the load is low, the combustion gas pressure decreases.
Can reduce the oil pressure of the supplied lubricating oil.

However, the rotor bearings 14a and 14b are subjected to the combustion gas pressure received by the rotors 11a and 11b and the centrifugal force due to the planetary motion of the rotor, and the two forces act in substantially opposite directions.
The maximum value of the load of the rotor bearing is mostly influenced by the gas pressure in the low rotation speed region and the centrifugal force in the high rotation speed region with respect to the rotation speed. More specifically, the rotor bearing load becomes maximum around the rotation angle at which the highest combustion force is generated, of the full rotation angle of the eccentric shaft 8 in one rotation in the low rotation range, and the rotation angle before and after the ignition top dead center in the high rotation range. , The centrifugal load is offset by the gas pressure and the rotor bearing load decreases. At other rotation angles, the centrifugal force causes the rotor bearing load to be at or near the maximum value. Therefore, in the high rotation region, the maximum value of the rotor bearing load is hardly affected even if the gas pressure at low load decreases. The centrifugal load works in proportion to the number of rotations squared, and the lubrication condition becomes severe as the load increases.

If the hydraulic pressure is reduced by the variable hydraulic control valve 31 at a low load as in this prior art, the injection ports 19a, 19
Spraying of lubricating oil for cooling the rotor from b is reduced or stopped, and overcooling of the rotor is prevented. However, since the lubricating oil passage in the eccentric shaft 8 is integrated, the oil pressure of the lubricating oil supplied to the rotor bearings 14a, 14b is also reduced. Although there is room for reducing the oil pressure of the lubricating oil supplied to the main bearings 7a and 7b, the reduction of the oil pressure of the lubricating oil supplied to the rotor bearings 14a and 14b is due to the lubrication performance of the bearing at high rotation. There is a problem in securing.

However, if the hydraulic pressure of the variable hydraulic control valve 31 at the time of high load is set to a high value in advance and the hydraulic pressure is reduced at a low load, there is no problem. The resistance increases, causing a decrease in output and a decrease in fuel efficiency.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-described problems, and has been made to prevent the rotor from overheating under a high load and prevent the rotor from being excessively cooled under a low load. It is an object of the present invention to obtain a cooling device for a rotor of a two-cylinder rotary engine that can be realized while ensuring the above.

[Means for solving the problem]

In order to achieve the above object, the present invention provides an eccentric shaft provided with a check valve for preventing outflow of lubricating oil when the hydraulic pressure in a lubricating oil passage provided therein is low, and passing the eccentric shaft through the check valve. In the two-cylinder Wankel type rotary engine in which the rotor is cooled by injecting the lubricating oil into the rotor hollow portion from the spout of the rotor, an intermediate main bearing is provided in the intermediate side housing, and the lubricating oil passage in the eccentric shaft is formed at both ends. The lubricating oil passage from the main bearing to the rotor bearing and the lubricating oil passage from the intermediate main bearing to the jet port are formed independently, and lubrication supplied to the intermediate main bearing among the three main bearings when the engine is under low load. Only the oil pressure of the oil changes from the oil pressure controlled by the oil pressure control valve that controls the oil pressure of the lubricating oil that is pumped to each part of the engine by the oil pump. Means for reducing in the reduced range with no lubrication on trouble of the main bearing is provided.

[Action]

In the case of a two-cylinder rotary engine, the inertia forces of the moving parts are perfectly balanced, so that the load applied to the main bearings at both ends is basically only the central component of the combustion gas pressure received by the rotor. However, the same applies to the intermediate main bearing, and according to the above configuration, the lubricating oil passage in the eccentric shaft is provided with a lubricating oil passage from the main bearings at both ends to the rotor bearing, and a lubricating oil passage from the intermediate main bearing to the injection port. Since the passage is formed independently, when the engine is under low load, the pressure of the combustion gas is reduced and its component in the center direction is also reduced, so the load acting on the intermediate main bearing is reduced, and the lubrication conditions are eased. Only the oil pressure of the lubricating oil supplied to the bearings does not affect the lubrication of the intermediate main bearing from the oil pressure controlled by the oil pressure control valve that controls the oil pressure of the lubricating oil pumped to each part of the engine by the oil pump. By reducing in the reduction range, the lubricating oil from the ejection port becomes the ejection amount decreases or ejection stop, supercooling of the rotor is prevented.

On the other hand, high-pressure lubricating oil is supplied from the main bearings at both ends to the rotor bearing on which a large centrifugal load acts during high rotation, so that the lubrication performance of the rotor bearing during high rotation is ensured.

〔Example〕

The same parts as those in FIG. 5 which explained the prior art are denoted by the same reference numerals, and the embodiment will be described with reference to the drawings.

In FIG. 1, the intermediate main bearing portion 38 of the eccentric shaft 8 is rotatable by an intermediate main bearing 37 formed of an intermediate main bearing boss 35 formed on the intermediate side housing 2 and an intermediate main bearing metal 36 mounted on the inner periphery thereof. It is supported by The intermediate main bearing 37 reduces the load capacity of the main bearings 7a and 7b at both ends and the rigidity of the eccentric shaft 8 when increasing the output by increasing the displacement of the rotor, increasing the supercharging, increasing the rotation, etc. It is definitely necessary if you think about it.

The axial lubricating oil passage inside the eccentric shaft 8 is located between the spout branch passage and the eccentric shaft branch passage in FIG.
split between a and 18a, and between 20b and 18b, 16a,
It is divided in the axial direction into three, 16c and 16b. The lubricating oil passage including the intermediate main bearing branch passage 57, the axial lubricating oil passage 16c, and the outlet branch passages 20a, 20b, the main shaft branch passages 17a, 17b, the axial lubricating oil passages 16a, 16b, and the eccentric shaft The lubricating oil passage including the branch passages 18a and 18b is formed independently. Here, the main shaft portion branch passage, the axial lubricating oil passage, and the eccentric shaft portion branch passage 17a, 16a, 18a and 17b, 16b, 18b are not necessarily independent. Until that was done for convenience.

Eccentric shaft 8 with intermediate main bearing 37
FIG. 2 shows a specific embodiment for enabling the support of the eccentric shaft. The eccentric shaft is a split type using a taper joint known in Japanese Patent Application Laid-Open No. 60-69208.
In the drawing, the left side is the front and the right side is the back.
8b has a rear main shaft portion 9b (FIG. 1), a rear eccentric shaft portion 10b, and an intermediate main shaft portion 38. A tapered outer peripheral surface 39 is formed in front of the intermediate main shaft portion, and the front portion has a small diameter. The portion 40 is formed. On the other hand, the sub member 8a is a cylindrical member formed with a front eccentric shaft portion 10a and a front main shaft portion 9a (FIG. 1), and the inner peripheral surface at the rear end thereof is formed on the tapered outer peripheral surface 39 of the main member 8b. It is formed on the corresponding tapered inner peripheral surface 41. The small diameter portion 40 of the main member 8b penetrates through the front hollow portion 42, and a nut (not shown) is fastened to a screw portion (not shown) formed at the front end of the small diameter portion to form a tapered inner periphery. Tapered outer surface on surface 41
The main member 8b and the sub-member 8a are firmly connected to each other by inserting the main member 8b.

From the rear end of the main member 8b of this known eccentric shaft, the axial lubricating oil passages 16b and 16c are continuously excavated from the rear end thereof to a position where the axial lubricating oil passage 16c intersects with the front outlet branch passage 20a. The axial lubricating oil passage 16a is excavated to a point just before the abutment. Then, the plug 43 is press-fitted between the axial lubricating oil passages 16c and 16b slightly behind the position intersecting with the rear outlet branch passage 20b. Further, the front end of the axial lubricating oil passage 16a and the rear end of the axial lubricating oil passage 16b are sealed to form axial lubricating oil passages 16a, 16c, 16b. For the sake of illustration, the direction of the ejection ports 19a and 19b is
Although it is the same direction as the eccentric direction of a, 10b, actually,
In order to enhance the cooling effect, the eccentric shaft is directed toward the leading side in the rotational direction by about 80 degrees. If this split type eccentric shaft is used, it is not necessary to pass the eccentric shaft portion 10a or 10b through the intermediate side housing 2 at the time of assembling as in the case of the conventional integral eccentric shaft, since it is assembled from both sides of the intermediate side housing, A spring receiving casing 44 having spouts 19a, 19b,
Since 44 may be protruded from the eccentric shaft portions 10a and 10b when viewed from the axial direction of the eccentric shaft, the dimensional constraint of the check valves 23a and 23b is relaxed, and the degree of freedom in setting the characteristics of the spring 22 is increased.

In the embodiment shown in FIG. 2, the ball valve element 21 on the front side must be provided on the side of the sub-member 8a, and the effect of the centrifugal force becomes large because the radius of rotation becomes large. This is not necessarily bad, but as shown in FIG. 3, the axial lubricating oil passage 16c is formed by closing the spring receiving casing 44 from both sides, and the ball valve element 21 is formed. Is placed on the rotation center axis, and the check valve 23a,
The configuration of 23b is not affected by the centrifugal force.

Returning to FIG. 1, a lubricating oil passage 45 from the oil gallery 32 to the intermediate main bearing 37 is formed in the intermediate side housing 2. In the middle of the lubricating oil passage, when the engine is under a high load (when the poppet valve body 48, which will be described later, closes the relief passage 56), there is no shortage of the ejection from the ejection ports 19a and 19b and the lubrication of the intermediate main bearing 37. An orifice 46 having a flow path cross-sectional area which is capable of supplying hydraulic lubricating oil and has no pressure drop downstream thereof is provided. A variable relief valve 47 is provided between the orifice and the intermediate main bearing 37, and the variable relief valve is configured by a spring 50 having a poppet valve element 48 abutted at one end inside a spring receiving casing 49. The spring 52 is pressed against the valve seat 51, and the end surface of the spring receiving casing 49 is a cam sliding surface, to which the cam 52 is in contact. The cam is connected to an intake throttle valve 54 of the intake passage 53 by an appropriate link mechanism 55, and is linked to the opening degree of the intake throttle valve 54. As the opening degree of the intake throttle valve 54 increases, the spring receiving casing 49 The cam 52 is formed in a shape such that the cam 52 is largely pushed in the direction of the body 48.

According to the above configuration, when the load is high, that is, when the opening degree of the intake throttle valve 54 is large, the poppet valve body of the spring 50
Since the force pressing the 48 against the valve seat 51 is increased and the poppet valve element 48 does not open the relief passage 56, the hydraulic lubricating oil controlled by the hydraulic control valve 31 is supplied to the intermediate main bearing 37. In FIG. 4, the vertical axis indicates the hydraulic pressure and the horizontal axis indicates the engine speed. Since the oil pump 28 is driven mechanically by the eccentric shaft 8,
Idling hydraulic low as P _4, it rises fairly rapidly until with increasing engine speed reaches the valve opening pressure P ₁ of the hydraulic control valve 31. In the high speed range than P ₁ point, with increasing engine speed, the relief flow rate from the hydraulic control valve 31 is gradually increased, the spring characteristic of the hydraulic control valve pressure is but approximately constant rise slowly. That is, as the engine speed increases, the hydraulic pressure progresses as indicated by the line a. Therefore, the line a is the check valve 23a, 23b
The engine speed region made above the line c indicating the valve opening pressure P _3, the lubricating oil is ejected spout 19a, the rotor hollow portion from 19b, the rotor 11a, 11b are cooled.

On the other hand, when the engine load is low, for example, the intake throttle valve
At the time of idling 54, the lift by the cam 52 becomes small, the spring receiving casing 49 moves in the direction pushed out by the spring 50, the force of the spring 50 pressing the poppet valve body 48 against the valve seat 51 is weakened, and the relief passage 56 is Since the opening lubricating oil is released, the flow velocity in the orifice 46 increases, and the oil pressure of the lubricating oil supplied to the intermediate main bearing 37 due to the pressure loss is reduced. In FIG. 4, the part with a vertical parallel line in the figure is the intermediate main bearing when the intake throttle valve 54 is idling.
This is a range in which the oil pressure of the lubricating oil supplied to 37 does not hinder the bearing lubrication, and a line e indicates the lower limit. Variable hydraulic lubricating oils such as line b relative to the engine rotational speed than the valve opening pressure P ₂ which reduced the relief valve 47 is supplied to an intermediate main bearing 37, the portion where該線b is denoted by the vertical parallel lines Is set so as to approach the line e. In the engine rotation range where the line b is below the valve opening pressure line c of the check valves 23a and 23b, lubricating oil is not ejected from the injection ports 19a and 19b, and even in the engine rotation range where the line b is above the line c. Since the oil pressure is reduced, the ejection amount is reduced, and the supercooling of the rotors 11a and 11b is prevented. At this time, since the lubricating oil is released from the variable relief valve 47,
The hydraulic pressure controlled by the hydraulic control valve 31, that is, the hydraulic pressure of the lubricating oil supplied to the main bearings 7a, 7b at both ends is delayed in rising as shown by the line d. However, since the load is in the low load and low rotation range, the loads acting on the main bearings 7a and 7b and the rotor bearings 14a and 14b are small, and there is no problem in lubrication.

In the partial load region, the line e moves upward as the opening degree of the intake throttle valve 54 increases, so that the line b approaches the line a so as to be above the line e. As the opening degree of the throttle valve 54 increases, the amount of lubricating oil jetted from the jet ports 19a, 19b is set to increase without hindering lubrication of the intermediate main bearing 37.

The pressures P ₁ , P ₂ , P ₃ , P ₄ and the gradients of the lines a, b, c, d in FIG. 4 should be determined by setting each part according to the purpose and requirements. It is.

As described above, by controlling continuously according to the opening degree of the intake throttle valve 54, the temperature of the rotors 11a and 11b can be appropriately maintained in all engine load states. At the time of full-open acceleration, the engine is actually accelerated and the rotor 11a,
Since 11b is sufficiently cooled even before it is overheated, the knocking prevention effect is enhanced. On the other hand, it is not possible to perform feedback control by sensing the temperature of the lubricating oil discharged by cooling the rotors 11a and 11b.

In the above embodiment, it is possible to reduce the oil pressure of the lubricating oil supplied to the intermediate main bearing 37 by a direct throttle using a variable orifice. However, the pressure control by the variable relief as in the embodiment can easily increase the control accuracy. In addition, variations in cooling performance for each product can be reduced.

Further, if the orifice 46 of the embodiment is a variable orifice and both are used together, it is possible to reduce the delay in the rise of the hydraulic pressure controlled by the hydraulic control valve 31.

In the above-described embodiment, if a torsion coil spring is attached to the shaft 58 of the cam 52 to urge the cam 52 to rotate in a direction in which the lift is increased, the link mechanism 55 may come off. When you are safe.

It is needless to say that the present invention can also be applied to a four-cylinder rotary engine in which two two-cylinder rotary engines are connected in series with a counterweight provided therebetween.

〔The invention's effect〕

The present invention as described above, the oil pressure of the lubricating oil to be supplied for lubrication of the intermediate main bearing required at the time of high output,
The oil pressure for ejecting the lubricating oil from the outlet of the eccentric shaft for cooling the rotor is designed to pay attention to the fact that a change in the same direction is required for a change in the engine load. It is possible to achieve both prevention of overheating of the rotor at the time of rotation and prevention of supercooling of the rotor at the time of low load while securing the lubrication performance of the rotor bearing at the time of high rotation.

In addition, by using an intermediate main bearing required when increasing the output, a controlled hydraulic lubricating oil for jetting from the jet port is introduced into the eccentric shaft.
There is also an advantage that no special dedicated lubricating oil introduction means is required.

[Brief description of the drawings]

FIG. 1 is a structural view of an embodiment of the present invention, FIG. 2 is a longitudinal sectional view showing one embodiment of an intermediate main bearing and an eccentric shaft supported thereon, and FIG. 3 is a longitudinal section showing another embodiment of the eccentric shaft. FIG. 4 is a graph showing an example of setting the oil pressure of each part with respect to the engine speed, and FIG. 5 is a configuration diagram of the prior art. 2. Middle side housing, 7a, 7b Main bearing, 8, 8a,
8b… eccentric shaft, 11a, 11b… rotor, 14a, 14b… rotor bearing, 16, 16a, 16b, 16c… axial lubrication oil passage, 17a, 17b… main shaft branch passage, 18a, 18b…
… Eccentric shaft branch passage, 19a, 19b …… Spout, 20a, 20b…
Injection branch passage, 23a, 23b ... Check valve, 28 ... Oil pump, 31 ... Hydraulic control valve, 34 ... Rotor hollow part, 37
…… Intermediate main bearing, 46 …… Orifice, 47 …… Variable relief valve, 57 …… Intermediate main shaft part branch passage

Claims (1)

(57) [Claims] An eccentric shaft is provided with a check valve for preventing outflow of lubricating oil when the hydraulic pressure in a lubricating oil passage provided therein is low, and a rotor hollow is passed through the check valve from an outlet of the eccentric shaft. In the two-cylinder Wankel type rotary engine that cools the rotor by injecting lubricating oil into the part, an intermediate main bearing is provided in the intermediate side housing, and the lubricating oil passage in the eccentric shaft is moved from the main bearing at both ends to the rotor bearing The lubricating oil passage and the lubricating oil passage from the intermediate main bearing to the injection port are formed independently, and only the oil pressure of the lubricating oil supplied to the intermediate main bearing among the three main bearings at the time of low engine load is used. The oil pressure of the lubricating oil that is pumped to each part of the engine by the pump is controlled by the hydraulic pressure control valve. The rotor of the cooling device for a secondary cylinder rotary engine provided with means for reducing in a reduced range are.