EP1859135B1 - Very compact device for adjusting the compression ratio of an internal combustion engine - Google Patents

Abstract

The device has a cam displacement control unit comprising a kinematic link that does not comprise a lock and connected to a flange ring (21) integrated with a cam (18). A position adjustment mechanism (29) and the link are integrated, outside a crank pin (22), a bearing (27) and a lever (23), in a volume describing a circle whose diameter is the same or less than the diameter of a greater circle. An independent claim is also included for a method for adjusting the compression ratio of an internal combustion engine.

Description

Technical area

The present invention relates to a device for adjusting the compression ratio of an internal combustion engine and to a method enabling the use of such a device.

It relates more particularly to a device that can change the compression ratio of this engine by changing the dead volume of the combustion chamber at the top dead center of the piston.

Prior art

It is already known, by the document EP 0 066 350 , a device for adjusting the compression ratio of an engine in which this engine comprises a crankshaft, a cylinder inside which a piston slides in a reciprocating movement by means of a connecting rod connected to said piston and said crankshaft, this piston delimiting with the top of the cylinder a combustion chamber having a dead volume at the top dead center of the piston, and a rotary eccentric type drawn, interposed between the connecting rod and the piston. This eccentric, in a first position, allows the piston to reduce the dead volume of the combustion chamber while increasing the compression ratio and increase this dead volume, for another position of this eccentric, while obtaining a rate of lower compression. To obtain these different positions, the bore of the big end has axial grooves which cooperate with a locking pin disposed in the eccentric, radially relative to the axis of said eccentric so as to immobilize it in one of the positions corresponding to one of said axial grooves of the connecting rod.

This device for adjusting the compression ratio has many advantages: it is located in the moving equipment and is energy efficient: it is powered by energy supplied directly by the moving equipment. Its implementation is not very constraining: it affects neither the combustion chamber nor the connections with the exhaust, or with the distribution members or with the transmission, nor the weight of the piston. It is nevertheless perfectible.

The big end is very bulky in order to accommodate both the eccentric and the mechanical locking device.

The control of the compression ratio adjustment device shall transit through several movable members relative to each other: the housing, the crankshaft and finally the eccentric linked to the connecting rod.

The mechanical locking system is necessarily subjected to high levels of friction and stress, and even possibly to shocks. This aspect combined with the little space available in the eccentric, affects the service life.

More recent designs have a different implementation of such a mechanical locking system. This is the case of the document JP 3,026,834 where the locking pin is housed in the bearing of the crankshaft or the document DE 10 243 023 where the fingers or locking pawls are housed in the crankpin. These designs also do not have a minimized volume: in fact, these locking devices being integrated almost completely in crankshaft parts through which the engine torque passes, the dimensions of said crankshaft parts are necessarily increased to allow the crankshaft to resist the constraints generated by the couple he transmits.

In addition these mechanical locking systems have a finite number of positions. They therefore do not allow to obtain a continuous adjustment of the compression ratio on a range.

In another type of device for varying the compression ratio, as better described in the patent application EP 1 247 958 , or US 2003/0 029 395 the eccentric is not an eccentric towed type but a motorized eccentric. An electric or hydraulic motor drives an irreversible worm which cooperates with a toothed sector of the eccentric.

This device has a major disadvantage because the electric or hydraulic motor must counteract the various friction and different inertial forces, including those of the mobile engine of the internal combustion engine, to motorize the eccentric. But these friction and these forces of inertia are very important. Said electric or hydraulic motor is therefore necessarily bulky. In addition, the energy to power this motor must be provided by a subsidiary body. The yield is therefore heavily penalized.

Presentation of the invention

The present invention proposes to overcome the drawbacks mentioned above by means of an energy saving compression ratio adjusting device, a compactness at the best level while being easily compatible with a long service life.

To this end, the present invention relates to a device for adjusting the compression ratio of an internal combustion engine comprising at least one cylinder with a combustion chamber, a mobile assembly comprising a piston displaceable in translation under the action of a rod connected by an axis to said piston and connected to a crankpin of a crankshaft, said piston making a race between a top dead center and a bottom dead center leaving a dead volume at the top dead center of said piston, the device comprising between the crankpin and the crankpin of the crankshaft a rotary eccentric to adjust the compression ratio, the device also comprising means for controlling the displacement of the eccentric, characterized in that the control means comprise at least one kinematic connection without a latch between a radial protuberance integral with the eccentric and an adjustment mechanism in position relative to the crankshaft, said m canism being integrated in one of the two blanks of the crankshaft, said kinematic linkage and said adjustment mechanism in position relative to the crankshaft being on the one hand disposed largely or entirely outside the crankpin, the bearing and the lever connecting the crank pin to the bearing of the crankshaft, and secondly integrated into a volume whose point furthest from the axis of the crankshaft describes a circle, during the rotation of the engine, a diameter of the same order of magnitude or less than the diameter of the largest circle described by the big end or the largest circle described by the flanks of the crankshaft.

According to the invention, the adjustment mechanism in position relative to the crankshaft belonging to the control means of the displacement of the eccentric according to the invention comprises at least one linear actuator. The advantage of a linear actuator lies in the simplicity.

The present invention combines several critical advantages, which are never all together simultaneously in the designs described in the prior art.

1. 1- L'excentrique logé dans la tête de bielle peut être dimensionné aux dimensions les plus compactes puisqu'il est solidaire d'une protubérance radiale par laquelle transitent les déplacements et les contraintes mécaniques liées au contrôle de sa position ;
2. 2- L'encombrement du vilebrequin peut être également être dimensionné aux dimensions les plus compactes puisque les moyens de contrôle du déplacement de l'excentrique sont disposés dans les zones disponibles balayées par un vilebrequin conventionnel, en majeure partie ou en totalité en dehors du maneton, du palier et du levier reliant le maneton au palier du vilebrequin, donc lesdits moyens de contrôle selon l'invention ne nécessitent pas d'accroissement de l'encombrement du vilebrequin alors que leur intégration n'affaiblit pas la résistance mécanique du vilebrequin ;
3. 3- L'encombrement nécessaire à l'équipage mobile peut aussi être minimal puisque les moyens de contrôle du déplacement de l'excentrique embarqués sur l'équipage mobile sont intégrés dans le cylindre géométrique définit par la rotation de la tête de bielle ou par la rotation du flan du vilebrequin.

The first advantage is a compactness at the best level on all the component parts of the device. This concerns:

1. The eccentric housed in the connecting rod head can be dimensioned to the most compact dimensions since it is integral with a radial protuberance through which the displacements and the mechanical stresses related to the control of its position transit;
2. 2- The size of the crankshaft can also be dimensioned to the most compact dimensions since the control means of the displacement of the eccentric are arranged in the available areas swept by a conventional crankshaft, largely or entirely outside the crankpin , the bearing and the lever connecting the crank pin to the bearing of the crankshaft, therefore said control means according to the invention do not require an increase in the size of the crankshaft while their integration does not weaken the mechanical strength of the crankshaft;
3. 3- The space required by the mobile crew may also be minimal since the means for controlling the displacement of the eccentric on board the moving element are integrated in the geometric cylinder defined by the rotation of the end of the connecting rod or by the rotation of the crankshaft blank.

The second advantage is an important volume available to house the means for controlling the displacement of the eccentric. This second advantage is compatible with the aforementioned first because according to the invention the main components of said control means: the kinematic connection and the adjustment mechanism in position relative to the crankshaft, are integrated in the crankshaft blank, in space adjacent to the lever, crankpin and bearing of the crankshaft. This volume is important and available in the usual size of the mobile equipment of a traditional engine.

1. a) L'importance du volume disponible pour intégrer le mécanisme d'ajustement en position par rapport au vilebrequin et sa liaison cinématique avec l'excentrique, ce qui permet de dimensionner généreusement ces composants sans pénaliser l'encombrement global ;
2. b) Ladite liaison cinématique ne comporte pas de verrou, elle n'est donc pas fragilisée par ce type de système ;
3. c) La protubérance radiale solidaire de l'excentrique réalise une fonction simple, elle est donc aisée à réaliser de façon robuste ;
4. d) Ladite protubérance radiale participe à la résistance mécanique de l'excentrique et répartit les contraintes dans sa section ;
5. e) L'excentrique est exempt de pièce ou de forme constitutive d'un verrou, il n'est donc pas fragilisé par ce type de système.

The third advantage is the ability to have high durability related to the robustness of the designs possible with the present invention. This third advantage is related to the features of the present invention listed below:

1. a) The importance of the volume available to integrate the adjustment mechanism in position relative to the crankshaft and its kinematic connection with the eccentric, which allows to dimension generously these components without penalizing the overall size;
2. b) Said kinematic connection does not include a lock, it is not weakened by this type of system;
3. c) The integral radial protuberance of the eccentric performs a simple function, so it is easy to achieve robustly;
4. d) Said radial protuberance contributes to the mechanical strength of the eccentric and distributes the stresses in its section;
5. e) The eccentric is free of part or form constituting a lock, it is not weakened by this type of system.

The features of the present invention therefore combine a long service life with a compactness at the best level and a large volume for the integration of the device.

According to a first complementary characteristic, the means for controlling the displacement of the eccentric according to the invention make it possible to carry out a continuous adjustment of the compression ratio over its range of variation. This feature associates with the above advantages of the present invention the ability to adjust at any point the compression ratio to the optimum value.

According to a second complementary characteristic, the means for controlling the displacement of the eccentric according to the invention use energy taken from the moving equipment to move the eccentric. This characteristic associates with the above-mentioned advantages of the present invention the possibility of adjusting the compression ratio with a high reactivity and a high energy consumption of the device Furthermore, this characteristic also combines the advantage of not being able to draw on the peripherals of the device. mobile equipment or the internal combustion engine as the control energy required for the device according to the invention.

According to a third complementary characteristic, the means for controlling the displacement of the eccentric according to the invention comprise two sets placed on either side of the eccentric and each consisting of at least one kinematic link without a lock connected to a protuberance radial integral with the eccentric and an adjustment mechanism in position relative to the crankshaft. This characteristic makes it possible to reinforce the robustness for the highly stressed mobile crews.

According to a fourth complementary characteristic, the two adjustment mechanisms in position relative to the crankshaft belonging to the two sets placed on either side of the eccentric, mentioned in the preceding paragraph, are kinematically linked so that they participate in approximately equal to the control of the displacement of the eccentric. This feature reinforces the robustness of the whole.

According to another complementary feature, said adjustment mechanism in position of the eccentric relative to the crankshaft according to the invention comprises two linear actuators whose axes are distinct. These two actuators can be single-acting. Thus each actuator can work by simple pushing and act in opposite directions on the orientation of the eccentric. This design simplifies the kinematic connections. In addition, the axial size of a single-acting actuator is less than that of a double-acting actuator.

According to another complementary feature of the previous one, the two Linear actuators are placed on each side of the crankpin and bearing of the crankshaft. This design facilitates the integration of said actuators in the crankshaft blank.

According to another complementary feature of the invention, the coefficient of friction between the bore of the eccentric and the crankpin is less than seventeen hundredths. This value of the coefficient of friction has the advantage of allowing the eccentric to be towed for many applications. The definition of a towed eccentric is specified in the description of the preferred embodiment.

According to another complementary feature, the coefficient of friction between the eccentric and the bore of the big end is greater than twenty hundredths. This value of the coefficient of friction has the advantage of allowing the eccentric not to be towed for many applications. The advantage is that the eccentric does not rotate in the absence of a torque generated by a specific actuator. The eccentric therefore maintains its angular position without requiring any specific means to block it or to maintain it.

According to another feature which constitutes a variant of the preceding one, the means for controlling the displacement of the eccentric according to the invention measures a distance using a non-contact measuring sensor, between a fixed position of the motor housing. and one of the parts that moves relative to the crankshaft to adjust the compression ratio. Measuring a distance of this type has the advantage of enabling the device to determine the compression ratio without significant error. In addition, a non-contact sensor provides a long life and high reliability for this measurement.

The integral radial protuberance of the eccentric linked to the kinematic connection belonging to the control means of the displacement of the eccentric may be a collar integral with the eccentric. This design has the advantage of distributing the control constraints of the position of the eccentric over the three hundred and sixty degrees of the eccentric.

The means for controlling the displacement of the eccentric according to the invention have means for controlling the adjustment mechanism in position relative to the crankshaft.

• déterminer le taux de compression souhaité du moteur ;
• déterminer la position que doit atteindre l'une des pièces qui se déplace par rapport au vilebrequin selon une fonction continue du taux de compression, à un angle de rotation déterminé du vilebrequin, pour obtenir le taux de compression souhaité;
• contrôler la distance audit angle de rotation du vilebrequin utilisé à la phase précédente, entre une position solidaire du carter du moteur et ladite pièce qui se déplace par rapport au vilebrequin selon une fonction continue du taux de compression.

The invention also relates to a method of adjusting the compression ratio of an internal combustion engine, said engine comprising at least one cylinder with a combustion chamber, a mobile unit comprising a piston movable in translation under the action of a rod connected by an axis to said piston and connected to a crankpin a crankshaft, said piston making a stroke between a top dead center and a bottom dead center while leaving a dead space at the top dead center of said piston, characterized in that the method comprises:

• determine the desired compression ratio of the engine;
• determining the position to be reached by one of the parts moving relative to the crankshaft according to a continuous function of the compression ratio, at a determined rotation angle of the crankshaft, to obtain the desired compression ratio;
• controlling the distance to said rotation angle of the crankshaft used in the preceding phase, between a fixed position of the engine casing and said part which moves relative to the crankshaft according to a continuous function of the compression ratio.

Brief description of the drawings

• la figure 1 montre, en vue crevée, le moteur à combustion interne selon un plan parallèle aux axes du vilebrequin et du cylindre
• la figure 2 montre le moteur à combustion interne selon un plan de coupe parallèle l'axe du cylindre et perpendiculaire au vilebrequin ;
• la figure 3 montre, en vue crevée selon un plan parallèle aux axes du vilebrequin et du cylindre, le moteur à combustion interne ajusté à son taux de compression maximum équipé du dispositif réalisé selon un mode particulier de réalisation ;
• la figure 4 montre le moteur à combustion interne ajusté à son taux de compression maximum selon un plan de coupe parallèle l'axe du cylindre et perpendiculaire au vilebrequin, équipé de dispositif réalisé selon un mode particulier de réalisation ;
• la figure 5 montre, en vue crevée selon un plan parallèle aux axes du vilebrequin et du cylindre, le moteur à combustion interne ajusté à son taux de compression minimum équipé du dispositif réalisé selon un mode particulier de réalisation ;
• la figure 6 montre le moteur à combustion interne ajusté à son taux de compression minimum selon un plan de coupe parallèle l'axe du cylindre et perpendiculaire au vilebrequin, équipé de dispositif réalisé selon un mode particulier de réalisation ;
• la figure 7 montre, le moteur à combustion interne équipé de ses conducteurs de commande du dispositif ;
• la figure 8 montre les composants mécaniques embarqués sur l'équipage mobile du dispositif d'ajustement du taux de compression;
• la figure 9 montre le vilebrequin et le pied de bielle équipé du dispositif d'ajustement du taux de compression pour les moteurs fortement chargés ;
• la figure 10 montre une variante des éléments de transmission de la commande du dispositif d'ajustement du taux de compression ;
• la figure 11 montre le vilebrequin en coupe pour une variante du dispositif d'ajustement du taux de compression ;
• la figure 12 montre, selon un plan de coupe parallèle l'axe du cylindre et perpendiculaire au vilebrequin, le moteur équipé du dispositif réalisé selon une autre manière de réaliser l'invention en version hydraulique ;
• la figure 13 montre le schéma hydraulique pour une autre manière de réaliser l'invention en version hydraulique ;
• la figure 14 présente un graphique de rotation de l'excentrique en fonction la pression de pilotage ;
• la figure 15 montre le vilebrequin en coupe pour une autre variante du dispositif d'ajustement du taux de compression.

The other features and advantages of the invention will become apparent on reading the following description, given solely by way of illustration and not limitation, and to which are appended:

• the figure 1 shows, in punctured view, the internal combustion engine in a plane parallel to the axes of the crankshaft and the cylinder
• the figure 2 shows the internal combustion engine in a sectional plane parallel to the axis of the cylinder and perpendicular to the crankshaft;
• the figure 3 shows, in a punctured view along a plane parallel to the axes of the crankshaft and the cylinder, the internal combustion engine adjusted to its maximum compression ratio equipped with the device made according to a particular embodiment;
• the figure 4 shows the internal combustion engine adjusted to its maximum compression ratio along a cutting plane parallel to the axis of the cylinder and perpendicular to the crankshaft, equipped with a device made according to a particular embodiment;
• the figure 5 shows, in a punctured view along a plane parallel to the axes of the crankshaft and the cylinder, the internal combustion engine adjusted to its minimum compression ratio equipped with the device made according to a particular embodiment;
• the figure 6 shows the internal combustion engine adjusted to its rate of minimum compression along a cutting plane parallel to the axis of the cylinder and perpendicular to the crankshaft, equipped with a device made according to a particular embodiment;
• the figure 7 shows, the internal combustion engine equipped with its control drivers of the device;
• the figure 8 shows the mechanical components embedded on the mobile equipment of the compression ratio adjustment device;
• the figure 9 shows the crankshaft and small end equipped with the compression ratio adjustment device for heavily loaded engines;
• the figure 10 shows a variant of the transmission elements of the control of the compression ratio adjusting device;
• the figure 11 shows the crankshaft in section for a variant of the adjustment device of the compression ratio;
• the figure 12 shows, according to a sectional plane parallel to the axis of the cylinder and perpendicular to the crankshaft, the engine equipped with the device made according to another way of carrying out the invention in hydraulic version;
• the figure 13 shows the hydraulic diagram for another way to realize the invention in hydraulic version;
• the figure 14 presents a graph of rotation of the eccentric as a function of the pilot pressure;
• the figure 15 shows the crankshaft in section for another variant of the device for adjusting the compression ratio.

Ways to realize the invention

The figures 1 and 2 show an internal combustion engine with at least one cylinder 31 which comprises a bore 16 inside which slides a hollow piston 14 in an alternative translational movement under the impulse of a rod 15. This piston delimits with its upper part, the side wall of the bore 16 and the upper part of this bore, generally formed by a portion of the cylinder head 11, a combustion chamber 10 in which the combustion cycle takes place. The piston carries two diametrically opposed radial bores through which is housed a cylindrical axis 13 which connects the small end 12 to said piston. The connecting rod head 17 is connected by a compression ratio adjusting device 32 to a crank pin 22 of a crankshaft 28. This crankshaft 28 is subjected to a rotational movement about an axis XX. As is known, the piston 14, the axis 13, the connecting rod 15, the crankshaft 28 with its crankpin 22 form the moving element of the engine.

In conventional engines, during the rotational movement of the crankshaft 28, the crank pin 22 passes successively from a high position to a low position. During this movement, the piston 14, which is connected to the crank pin 22 by the connecting rod 15, undergoes an alternative translational movement between a top dead center and a bottom dead center. In these engines, when the piston is at the top dead center, either at the end of the compression phase or at the end of the exhaust phase, there remains a dead volume in the combustion chamber 10. As is well known Those skilled in the art, the compression ratio of an engine is a function not only of the extent of the volume of the cylinder delimited by the stroke of the piston but also the magnitude of the dead volume. To modify the compression ratio, simply modify one of these volumes and more particularly the size of the dead volume.

To do this, the compression rate adjusting device 32 comprises an eccentric 18 housed between the crankpin 22 and a bore 19 provided in the crankshaft head 17. This eccentric 18 has a generally circular shape with a geometric axis X1X1 which corresponds at its middle axis and comprises a bore 20 axis X0X0 non-coaxial with the axis X1X1 but coincides with the axis of the crank pin 22. This eccentric is slidably accommodated in the receiving bore 19 made in the connecting rod head 17 and on the peripheral wall of the crank pin 22.

When the piston 14 is at the top dead center, the dead volume of the combustion chamber 10 is a function continuous of the angular orientation of the eccentric 18. In fact, the axis of the cone head 17 is merged with the X1X1 axis of the eccentric 18 and the axis of the crank pin 22 with the axis X0X0 of the bore 20 of the eccentric 18. Or the axis X1X1 of the eccentric 18 is not coaxial with the axis X0X0 of its bore 20. Thus when the piston 14 is at the top dead center, the distance between the axis of the big end 17 and the cylinder head 11, is a continuous function of the angular orientation of the eccentric, defined by example by the angle between firstly the line passing through its axis X1X1 and the axis X0X0 of its bore 20, secondly the reference line YY perpendicular to the axis of the cylinder 31 and the axis X1X1 of the eccentric 18. The figure 2 presents two angular orientations of the eccentric, one in solid line and the other in dotted line, corresponding to two different compression rates of the internal combustion engine. The angular orientation of the eccentric 18 of angles AH between the straight lines YY and DH, the straight lines DH passing through the axes X1X1 of the eccentric 18 and X0X0 of its bore 20, the top dead center of the piston 14 is PMHmax and corresponds to a dead volume VHmin of the combustion chamber 10. The angular orientation of the eccentric 18 corresponding to the angle AB between the lines YY and DB, when the line DB passes through the axes X1X1 of the eccentric 18 and X0X0 of its bore 20, the top dead center of the piston 14 is PMHmin and corresponds to a dead volume VHmax of the combustion chamber 10. The dead volume VHmax is greater than the dead volume VHmin and corresponds to a lower compression ratio of the internal combustion engine.

The compression rate adjusting device 32 also comprises means for controlling the displacement of the eccentric 18 including on the one hand the kinematic connection 30 kinematically connected to the flange 21 which constitutes the radial protuberance integral with the eccentric 18, on the other hand the adjustment mechanism in position 29 relative to the crankshaft 28. Said adjustment mechanism in position 29 is integrated in the blank 33 of the crankshaft 28, this integration being more particularly carried out almost entirely in the mass of balancing 26 of said crankshaft 28. Said kinematic connection 30 and said adjustment mechanism in position 29 are integrated entirely outside the crankpin 22, the bearing 27 and the lever 23 connecting the crankpin 22 to the bearing 27 of the crankshaft28. They occupy a volume whose point furthest from the axis XX of the crankshaft 28 describes a circle 25, during the rotation of the engine, with a diameter of the same order of magnitude as the largest diameter of the circle 24bi described by the connecting rod 17 as well as the largest diameter of the circle 24vi described by the blanks 33 of the crankshaft 28. The kinematic connection 30 does not include a lock and the eccentric 18 is free of part or form of a constituent lock.

The Figures 3 to 7 presents the first particular preferred embodiment of the invention. The flange 21 secured to the eccentric 18 forms a rocker with two connecting pads 34a, 34b. The adjustment mechanism in position 29 relative to the crankshaft 28 comprises two linear actuators placed on either side of the crank pin 22 and the bearing 27 of the crankshaft 28. These two linear actuators are single acting hydraulic cylinders 36a and 36b of which the axes 37a, 37b are distinct. The kinematic connection 30 between the collar 21 and the adjustment mechanism in position 29 is formed by the top of the rods 30a, 30b of the jacks 36a and 36b which push the connecting studs 34a, 34b of the rocker formed by the collar 21. The kinematic link 30 and the adjustment mechanism in position 29 are completely integrated outside the crankpin 22, the bearing 27 and the lever 23 connecting the crankpin 22 to the bearing 27 of the crankshaft28. They occupy a volume whose point farthest from the axis XX of the crankshaft 28 describes a circle 25, during the rotation of the engine, with a diameter equal to the largest diameter of the circle 24bi described by the connecting rod 17 As well as the circle 24vi described by the flanks 33 of the crankshaft 28.

Each single-acting hydraulic cylinder 36a and 36b operates by simple kinematically opposite thrusts by virtue of the function performed by the rocker articulated around the axis X0X0 of bore 20 belonging to the eccentric 18 and coincide with the axis of the crank pin 22. An additional advantage of this compression rate adjustment device is that it does not apply axial direction force on the eccentric 18.

The circulation of the oil between the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b is controlled by a hydraulic valve 40 placed in the balancing mass 26 of the crankshaft 28.

The first particular preferred embodiment of the invention described above makes it possible to continuously adjust the compression ratio of the internal combustion engine over its range of variation. Indeed, each constituent sub-assembly of the compression ratio adjustment device makes it possible to achieve positioning on any point within the range of variation. These subassemblies are the eccentric 18 secured to its flange 21 and its pads 34a, 34b which can be positioned at any angle within the range of variation of the positioning angle of the eccentric 18 , the kinematic linkage 30a, 30b which is continuous, reversible and free of component parts of a latch, the linear hydraulic cylinders 36a, 36b which can be positioned at any position in their respective range of variation, otherwise they are also reversible, and finally the hydraulic valve 40 which can supply the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b to achieve any positioning within their range of variation.

According to the first particular preferred embodiment of the invention presented on the Figures 3 to 6 , the eccentric 18 is towed during the operation of the engine and the hydraulic valve 40 allows to allow or prohibit at any time and for an adjustable period, via the hydraulic pipe 42a, 42b, the passage of oil between the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b. This eccentric is said towed when subjected, during operation of the motor, to a driving torque in rotation about the axis X0X0, successively in the clockwise direction of the eccentric 18 and in the anti direction. -clockwise, generated by the forces due to the different inertias due to displacements of the moving engine of the rotating engine, conjugated to friction forces and to the resulting forces of the gaseous pressures exerted on said moving element. In order for this driving torque of the eccentric 18 to be positive successively in the clockwise rotation direction of the eccentric 18 and then counter-clockwise, and for it to make it possible to obtain a range of variation S of the Piston altitude 14 at top dead center, it is necessary to meet the following criteria: a limit travel angle of eccentric 18 and the limit values of different influencing coefficients of friction. For the realization presented on the Figures 3 to 6 the variation range S of the piston 14 at the top dead center is five millimeters and the diameter of the crankpin 22 is fifty millimeters. The deflection angle of the eccentric 18 is thirty degrees above and thirty degrees below the reference line YY. The coefficient of friction of the fluid bearing, placed between the bore of the big end 17 and the eccentric 18, is less than five per thousand. The coefficient of friction between the crankpin 22 of the crankshaft 28 and the eccentric 18 is less than one-tenth. The crankpin 22 of the crankshaft 28 is coated with amorphous carbon and lubricated to guarantee this upper limit of coefficient of friction. Thus, when the valve 40 puts in communication the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b, the eccentric 18 is accelerated in rotation, with respect to the crankpin 22 of the crankshaft 28, in a direction which depends mainly on the engine time internal combustion in its operating cycle, suction or compression or exhaust or explosion or other, the angle and speed of rotation of the crankshaft 28, and the load of the internal combustion engine. On the other hand, when the valve 40 blocks the oil transits between the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b, the rotational position of the eccentric 18 with respect to the crankpin 22 and with respect to the crankshaft 28 is stopped because that the oil can not leave the chambers 35a, 35b of the hydraulic cylinders 36a, 36b and that said chambers are free from air. The means for controlling the displacement of the eccentric according to the first preferred embodiment of the invention thus motorize the eccentric with energy taken directly from the moving equipment. Only the energy required to control the control means of the displacement of the eccentric is taken from the peripherals of the internal combustion engine. This feature minimizes the energy required to adjust the compression ratio.

Moreover, in order to eliminate any presence of air in the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b and also in order to compensate for the losses of oil which returns to the cover of the engine casing, the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b are filled with oil permanently by the engine lubrication pump, via the pipes 38a, 38b, the non-return valves 39a, 39b and the pipe 41 of the usual lubrication of the bearing and the crankpin . The mounting direction of the non-return valves 39a, 39b is such that the pipe 41 can supply oil to the chambers 35a, 35b of the hydraulic cylinders 36a, 36b, but the oil returns of said chambers to the pipe 41 are blocked.

In order to position the compression ratio at its maximum value when the engine is stopped, the first preferred embodiment according to the invention provides a hydraulic pipe 44 which connects the chamber 35a of the compression ratio increasing cylinder 36a. , to a means of generating hydraulic pressure while the engine is stopped, via a non-return valve 45 which prevents the oil from returning to said hydraulic pressure generating means. This option has the advantage of allowing the engine to be stopped immediately, for any value of the compression rate, n, at the request of the user, while having the highest compression ratio to facilitate the starting of the internal combustion engine in very cold weather.

According to the first preferred embodiment of the invention, presented on the figure 7 , the hydraulic valve 40 is controlled via an electromagnet. Its electric coil, not shown, is integral with the motor housing and its movable core is embedded on the movable element to enable the hydraulic valve spool to be actuated. The magnetic flux generated by the electric coil transits in the magnetic field conductors 47a, 47b, 47c, 47d integral with the motor housing, and then by air strips 48c, 48d to reach and circulate in the magnetic field conductors 49c, 49d embedded on the mobile unit, specifically for this application, embedded on the balancing mass 26 in the blank 33 of the crankshaft 28 and in the aforementioned movable core. This embodiment has the advantage of a long life because the electrical and electromagnetic control is transmitted without friction. Moreover, a greater choice is possible to place the electric coil without penalizing the overall size. The coil can be powered by continuous electrical connections, without the interface of an electrical collector.

The slide, not shown, of the hydraulic valve 40 in the closed position is firstly pushed in the direction of closure by a spring, on the other hand coupled to a double-acting cylinder whose forces exerted each side of his piston are in equilibrium. When the hydraulic valve 40 is in the closed position, the hydraulic chambers of said double-acting cylinder are supplied with oil under pressure by the chambers 35a, 35b of the cylinders 36a, 36b, via non-return valves, not shown, in order to prevent any communication of hydraulic fluid between the two chambers 35a, 35b of the cylinders 36a, 36b by this control circuit. To open the hydraulic valve 40 and put in communication the two chambers 35a, 35b of the cylinders 36a, 36b, said movable core is moved under the action of the magnetic flux generated by the circuit of control, which opens a valve and causes the laying of a hydraulic chamber of the double-acting cylinder so that the pressure drop in the hydraulic chamber generates a double-acting cylinder force in the direction of opening of the hydraulic valve 40. The pressure drops in the circuit of said tarpaulin under the action of said movable core are much lower than the feed losses of the aforementioned hydraulic chamber of the double-acting cylinder by the chambers 35a. , 35b of the cylinders 36a, 36b. The consequence is a rapid opening movement of the hydraulic valve 40. The oil which thus returns to the cover of the motor housing is replaced in the device for adjusting the compression ratio by oil pressurized by the engine lubrication pump, via the circuit comprising the pipes 41, 38a, 38b and the non-return valves 39a, 39b described above. This makes it possible to obtain hydraulic assistance for the opening and closing of the hydraulic valve 40 and thus to carry out the control of the device for adjusting the compression ratio with a low power of the electromagnetic flux which passes through the magnetic conductors 47a. , 47b, 47c, 47d, 49c, 49d.

A variant of the control of the hydraulic valve 40 is presented on the figure 10 . The hydraulic valve 40 is actuated by the pusher 52. The movable cams 51a, 51b actuated by a device, not shown, make it possible, during the rotation of the motor, to actuate the pusher 52 or not to actuate it according to the control they receive from the control circuit. The cam 51a opens the valve 40 when the piston is close to the top dead center and the cam 51b when the piston is close to the bottom dead center.
According to the first preferred embodiment of the invention, presented on the figures 3 and 5 the compression ratio adjusted by the device is measured by a non-contact distance measuring sensor 43 fixed on the motor housing. This sensor measures the distance that separates it from the highest point reached by the lateral face of the flange 21 secured to the eccentric 18. Said lateral face is of a helicoidal shape so that the non-contact distance measuring sensor is inclined towards the axis of the crankshaft so that the smallest distance measured by said sensor is a continuous function of the angular orientation of the eccentric 18 relative to the crankshaft. This distance is correlated with the compression ratio by the mechanical kinematics of the device. This distance is therefore a reliable image of the compression ratio. This distance is Smax for the minimum compression ratio and Smin for the maximum compression ratio. Advantageously, the non-contact distance measuring sensor is an eddy current sensor. This type of sensor has the advantage of having a very short response time and high accuracy.

According to an addition to the preferred embodiment of the invention, presented on the Figures 8 and 9 , the control means of the displacement of the eccentric 18 are doubled and placed on either side of the eccentric. The eccentric 18 is on the one hand integral with a collar 21c placed on the left and controlled in position in particular by the cylinders 36c, 36d integrated in the blank 33a of the crankshaft 28, secondly secured to a second collar 21a located , to the right of the other side of the connecting rod and controlled in position in particular by the cylinders 36a, 36b integrated in the blank 33b of the crankshaft 28. This construction doubles the capacity of control torque in position of the eccentric. Moreover, the hydraulic chambers 35a, 35c of the jacks 36a, 36c for blocking the rotation of the eccentric in the clockwise direction are in communication via the hydraulic line 46a 46c and the hydraulic chambers 35b, 35d of the jacks 36b, 36d of FIG. blocking the rotation of the eccentric counterclockwise are in communication via the hydraulic pipe 46b 46d. This communication makes it possible to standardize the hydraulic pressures in each pair of jacks which act in the same direction in order to distribute the stresses and thus maximize the robustness of the device.

Another way to realize the invention in the hydraulic version is presented on the figures 12 and 13 . The two hydraulic cylinders 36a and 36b are integrated in a module 80, which module 80 is assembled on the crankshaft 28, positioned relative to its lever 23. The two hydraulic cylinders 36a and 36b are equidistant from the axis X0X0 of the crankpin 22 and the sections of their hydraulic chamber 35a, 35b are identical. This module 80 integrates the hydraulic power circuit comprising the hydraulic cylinders 36a and 36b, the controlled non-return valves 72a and 72b, the booster anti-retow valve 76, the line 75 of communication between the hydraulic cylinders 36a and 36b and the pilot lines 74a and 74b. This module 80 also fulfills the balancing mass function 26. This design has the advantage of making it possible to carry out and test the hydraulic functions for adjusting the compression ratio with the module 80 independently of the crankshaft 28 before assembly of the module 80 on the crankshaft 28. The steering ducts 74a and 74b of the assembled device are connected to annular grooves, not shown, made on the bearing of the crankshaft 28. This design has the advantage of allowing to connect a hydraulic control circuit 81 controlled check valves 72a and 72b, on supports integral with the motor housing. The feeding circuit of the hydraulic cylinders 36a and 36b, controlled by the check valve 76, is supplied by the engine lubrication circuit via the pipes 41 and 77. The return spring 73 makes it possible to put the compression ratio back to the maximum value when the engine is at a standstill.

Moreover, according to the manner of carrying out the invention presented on the figure 12 , the kinematic links 30a, 30b between the flange 21 of the eccentric 18 and the rods of the hydraulic cylinders 36a and 36b are made with the help of rods 70a and 70b. The joints between the rods 70a and 70b and respectively the collar 21 of the eccentric 18 and the rods of the hydraulic cylinders 36a and 36b are formed respectively by the connecting pads 34a, 34b of hemispherical shape and the ball joints 71a and 71b.

The operation of the compression rate adjustment device 32 according to this other way of carrying out the invention in a hydraulic version is controlled by the solenoid valve 79 shown schematically in FIG. figure 13 . When the solenoid valve 79 controls the controlled nonreturn valve 72a, the only possible oil transfers between the two hydraulic cylinders 36a and 36b are those which make it possible to reduce the compression ratio. When the solenoid valve 79 controls opening the controlled nonreturn valve 72b, the only possible oil transfers between the two hydraulic cylinders 36a and 36b are those which make it possible to increase the compression ratio. This design makes it possible to drive continuously one of the piloted check valves 72a or 72b during one or more of any motor while obtaining a variation of the compression ratio always in the same direction. The advantage lies in the fact that the response time of the hydraulic control system can be longer, without penalizing the desired direction of variation of the compression ratio. Moreover, the section of the slide of each controlled non-return valve 72a or 72b on the side of the hydraulic chamber 35a or 35b of the hydraulic cylinders 36a and 36b is greater than the section wetted by the oil on the side of the pipe 75 of communication between the cylinders 36a and 36b. Therefore, the higher the hydraulic pressure in one of the hydraulic cylinders 36a or 36b, the greater the force that tends to close the controlled nonreturn valve 72a or 72b corresponding is high. The controlled nonreturn valve 72a or 72b concerned remains open only if the opening hydraulic control pressure generated by the hydraulic control circuit 81 is sufficient. Consequently, the pilot pressure generated by the hydraulic control circuit 81 is a parameter for regulating the speed of variation of the compression ratio. An example of a graph representing the rotation of the eccentric per engine cycle, for motor rotation speeds of one thousand and two thousand revolutions per minute, when one of the non-return valves 72a or 72b is continuously controlled, is presented on the figure 14 . We find that the amplitude of rotation of the eccentric 18 by motor cycle, when the controlled nonreturn valve 72b is controlled, corresponds to an increase in the compression ratio according to an increasing function of the hydraulic control pressure generated by the hydraulic control circuit 81. analogous to piloting the controlled non-return valve 72a corresponds to a reduction in the compression ratio at each engine cycle according to an increasing function of the hydraulic control pressure generated by the hydraulic control circuit 81. Conversely, when the solenoid valve 79 pilot none piloted check valves 72a or 72b, oil transfers between the two hydraulic cylinders 36a and 36b are blocked and the compression ratio can not vary. The choke 78 is a calibration component by construction of the rate of variation of the compression ratio. The feeding circuit via the hydraulic lines 41 and 77 and the non-return valve 76 makes it possible to fill the oil circuit in the initial phase and then to compensate for any hydraulic leaks in the system. The role of the non-return valve 76 is to prevent any return of oil from the hydraulic power circuit to the engine lubrication circuit.

The figure 11 presents an alternative embodiment of the invention. The fluid bearing is made between the crankpin 22 of the crankshaft and the bore 20 of the eccentric 18. The coefficient of friction of this rotational connection is less than five thousandths. The coefficient of friction between the bore 19 of the big end 17 and the eccentric 18 is greater than thirty five hundredths. This physical characteristic is obtained thanks to a coating of nickel-titanium type deposited under vacuum on the outer diameter of the eccentric 18 and in the bore 19 of the conrod head 17. Moreover, this coating gives a long life treated parts. The variation range S of the piston 14 at top dead center is five millimeters and the diameter of the crankpin 22 is fifty millimeters. The deflection angle of the eccentric 18 is thirty degrees above and thirty degrees below the reference line YY. Given this construction, the eccentric is never towed. The means for controlling the position of the eccentric comprise a flange 21 integral with the eccentric 18 and parts integrated in the blank 33 of the crankshaft 28 composed of a rocker 62 kinematically connected to an actuating device and a skid 60 when in contact with the outer diameter 50 of the flange 21. When said actuating device is controlled by the control means, not shown, the flip-flop 62 pivots clockwise about its axis 61 integral with the blank 33 of the crankshaft 28 and plate 60 pad is on the collar 21. The pad 60 is retained by the hinge 68 integral with the latch 62 and can not rotate with the eccentric. The compound set of the rocker 62 articulated on the axis 61 actuated in rotation clockwise by the actuating device and the pad 60 remains connected to the blank 33 of the crankshaft 28 and induces a direct drive torque between the crankshaft and the eccentric . When this torque is greater than the torque generated by the resultant forces exerted between the connecting rod head 17 and the eccentric 18, the eccentric 18 is accelerated in the direction of rotation of the crankshaft 28. The crankshaft rotates in the direction counterclockwise. When the eccentric is rotated by the crankshaft, the compression ratio increases if the axis X1X1 of the eccentric 18 is placed, with reference to the figure 11 , to the right of the axis X0X0 of its bore and decreases in the opposite case. Thus to reduce or increase the compression ratio, if the axis X1X1 of the eccentric 18 is placed relative to the axis X0X0 of its bore, the side corresponding to the desired direction of variation of the compression ratio, the compression ratio varies in the desired direction as soon as the crankshaft drives the eccentric. In the opposite case, it is necessary to drive the eccentric in rotation with the crankshaft until the position of the axis X1X1 of the eccentric 18 with respect to the axis X0X0 of its bore changes side for obtain a variation of the compression ratio in the desired direction. This construction allows to control the positron in rotation of the eccentric on three hundred and sixty degrees. Consequently the eccentricity between the axis X1X1 of the eccentric and the axis X0X0 of its bore can be reduced to the lowest value since the rotation of the eccentric can be exploited over three hundred and sixty degrees to adjust the rate compression. This construction is therefore the most compact of all.

One way of carrying out the invention in an electric version, according to the variant presented in the preceding paragraph, is to equip the actuating device of the pad 60, via the flip-flop 62, with two piezoelectric actuators 64a, 64b. These two actuators 64a, 64b are on two separate axes, parallel and at the same distance from the axis 61 of the latch 62. They act in opposite directions to pivot the latch 62, via the pushers 63a, 63b. They are plated with the same geometric base towards the rocker 62 by elastic washers 65 which pushes a pusher 66 guided so that the distance differential of the base of the piezoelectric actuators 64a, 64b resting on the pusher 66 does not change. The stroke of the spring washers 65 is more than ten times greater than the stroke of the actuators. Thus, whatever the usual expansion differentials between the piezoelectric actuators 64a, 64b and the surrounding parts, the rotation of the rocker 62 is always a function of the elongation or retraction differential between the two actuators 64a, 64b. Moreover, the two actuators are always controlled simultaneously and in opposition of voltage so that the temperature differences between the two piezoelectric actuators 64a, 64b remains low. The tightness of the two piezoelectric actuators 64a, 64b is ensured by the shutter 67 and by seals, not shown, mounted on the pushers 63a, 63b. The actuators are electrically connected to the control means via electrical wires and rotating joints (not shown). Piezoelectric actuators have the advantage of offering extremely fast response times. This construction is therefore compatible with internal combustion engines whose rotation speed is high. In addition, the functions of the piezoelectric actuators are reversible. Also, the outer diameter 50 of the collar 21 secured to the eccentric 18 to a concentricity defect of three hundredths of a millimeter relative to the bore 20 of the eccentric 18 and a defect of cylindricity less than one hundredth of a millimeter. The relative rotation between the eccentric 18 and the crankshaft 28 thus generates stress variations on the two piezoelectric actuators 64a, 64b that these transform into electrical signals. Said force variations are correlated with the angular position of the eccentric with respect to the crankshaft. The control means in combination with the knowledge of the angular position of the crankshaft, deduces the value of the compression ratio. The piezoelectric actuators thus have the function of actuator for driving in rotation the eccentric with the crankshaft and measuring sensor which allows the control means to know the compression ratio at each revolution of the internal combustion engine.

Another way to realize the invention in electric version, according to the variant presented above in the penultimate paragraph, presented on the figure 15 , consists in equipping the actuating device of the shoe 60, via the flip-flop 62, with a single electric actuator 90 with a large stroke, so that the expansion differentials and the wears are easily compensated.

Other construction combinations are possible within the scope defined by the present invention. Of course, the present invention is not limited to the embodiments described but encompasses all variants and equivalents.

Possibilities of industrial application

The present invention can be applied to any reciprocating piston machine (s) and more particularly to internal combustion engines in order to reduce pollutant emissions as well as fuel consumption.

Claims (10)

1. Device of adjustment of the compression ratio of an internal combustion engine including at least a cylinder (31) with a combustion chamber (10), a moving assembly comprising a removable piston (14) in translation under the action of a rod (15) bound by an axis (13) of the piston (14) and connected to a crank pin (22) of a crankshaft (28), the aforementioned piston (14) carrying out a trajectory between a top dead center and a bottom dead center while letting remain a died volume at the top dead center of piston (14), the device including between the big end (17) and the crank pin (22) of the crankshaft (28) a rotary eccentric (18) allowing to adjust the compression ratio, the device also including means of control of the displacement of the eccentric, characterized in that the means of control include at least a kinematics connection (30) without locking system between a radial protuberance interdependent of eccentric (18) and one mechanism of adjustment in position (29) compared to crankshaft (28) comprising of at least one linear actuator, the aforementioned mechanism being integrated in one of two blanks (33) of the crankshaft, the aforementioned kinematics connection (30) and the aforementioned mechanism of adjustment in position (29) compared to crankshaft (28) being laid out in major part or entirely apart from the crank pin (22), from the rotation guide (27) and from the lever (23) connecting the crank pin (22) to the rotation guide (27) of crankshaft (28).
2. Device according to claim 1 characterized in that the mechanism of adjustment in position (29) compared to the crankshaft (28) is integrated in a volume whose point furthest away from axis XX of the crankshaft describes a circle (25), during the rotation of the engine, of a diameter of the same order of magnitude or lower than the diameter of the largest circle (24bi) described by the big end or of the largest circle (24vi) described by blanks (33) of crankshaft (28).
3. Device according to any of the claims 1 or 2 characterized in that the means of control of the displacement of the eccentric according to the invention include two units placed on both sides of the eccentric (18), each made up of at least one kinematics connection (30) without locking system related to a radial protuberance interdependent of eccentric (18) and of a mechanism of adjustment in position (29) compared to crankshaft (28).
4. Device according to any of the claims 1 or 2 characterized in that the mechanism of adjustment in position (29) of the eccentric (18) compared to the crankshaft (28) comprises two linear actuators.
5. Device according to claim 4 characterized in that the axes (37a), (37b) of the two linear actuators are distinct.
6. Device according to claim 4 characterized in that the mechanism of adjustment in position (29) compared to the crankshaft (28) comprises two hydraulic actuating cylinders for simple purpose (36a) and (36b) placed on both sides of the crank pin (22) and of the rotation guide (27) of the crankshaft (28).
7. Device according to claim 4 characterized in that the two linear actuators are two hydraulic actuating cylinders (36a) and (36b) integrated in a module (80), which module (80) is assembled on the crankshaft (28).
8. Device according to any of the claims 1 or 2 characterized in that the means of ordering of the position of the eccentric include a flange (21) interdependent of the eccentric (18) and parts integrated in the blank (33) of the crankshaft (28) made up in particular of a device of actuation and a friction element (60) in contact with the external diameter (50) of the flange (21), the element of friction which cannot turn with eccentric (18).
9. Device according to claim 8 characterized by the means of control of the displacement of eccentric (18) use at least one piezoelectric actuator (64a), (64b).
10. Device according any of the claims 8 and 9 for which the fluid rotation guide being realized between the crank pin (22) of the crankshaft and boring (20) of the eccentric (18) is characterized in that the friction element (60) makes it possible to induce an input torque between the crankshaft and the eccentric, under the action of the device of actuation, to accelerate the eccentric (18) in the direction of rotation of the crankshaft (28).

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