FR3062188A1 - BALANCING ARRANGEMENT AND BALANCING SYSTEM FOR BALANCING AN INTERNAL COMBUSTION ENGINE - Google Patents

Abstract

A balancing system for balancing an internal combustion engine (10) is formed of two balancing members (24A, 24B) guided in rotation about two coplanar balancing axes (102A, 102B). Each balancing member (24A) comprises a body (25A) provided with one or more bearings (62, 64) for rotating the body (25A), and at least one first balancing mass (26A) having a M1 value and a center of gravity located at a non-zero distance E1 of the balancing axis (102A), fixed in rotation relative to the body (25A) about the balancing axis (102A) and guided relative to the body (25A) so that the distance D1 from the center of gravity of the first balancing mass (26A) to a reference plane (C) containing the center of gravity of the body (25A) is variable.

Description

Holder (s):

NTN-SNR BEARINGS.

O Extension request (s):

(® Agent (s): ALATIS.

® BALANCING MEMBER AND BALANCING SYSTEM FOR BALANCING AN INTERNAL COMBUSTION ENGINE.

FR 3 062 188 - A1 (57) a balancing system for balancing an internal combustion engine (10) is formed by two balancing members (24A, 24B) guided in rotation around two coplanar balancing axes (102A , 102B). Each balancing member (24A) comprises a body (25A) provided with one or more bearings (62, 64) for guiding the body in rotation (25A), and at least one first balancing mass (26A) having a value M- | and a center of gravity located at a non-zero distance E ₁ from the balancing axis (102A), fixed in rotation relative to the body (25A) around the balancing axis (102A) and guided relative to the body (25A) so that the distance D- | from the center of gravity of the first balancing mass (26A) to a reference plane (C) containing the center of gravity of the body (25A) is variable.

i

BALANCING MEMBER AND BALANCING SYSTEM FOR BALANCING AN INTERNAL COMBUSTION ENGINE

TECHNICAL FIELD OF THE INVENTION The invention relates to the balancing of internal combustion engines with one or more reciprocating pistons.

STATE OF THE PRIOR ART [0002] Conventional internal combustion engines use a connecting rod-crank mechanism to convert the reciprocating movement of a piston into a rotational movement of a crankshaft. The piston moves back and forth in a cylinder bore, one end of which is closed by a cylinder head. The reciprocating movement of the piston defines a change in internal volume of the cylinder, the latter being minimum when the piston is close to the cylinder head and maximum when the piston is furthest from the cylinder head.

The piston makes a round trip per revolution of the crankshaft. Each complete stroke, from the minimum volume position to the maximum volume position or vice versa, is called an engine time. In a four-stroke engine, the combustion air is drawn into the volume during the first stroke. The torque is applied from the crankshaft to compress the combustion air during the second stage and form a mixture between the oxidant and the fuel. The mixture is converted into heat and pressure in the combustion chamber when the volume is close to its minimum. The pressure acts on the piston, creating a torque on the crankshaft during the third time. The amplitude of the torque varies with the variable transformation ratio of the connecting rod-crank mechanism and with the pressure in the cylinder, which decreases due to the increase in volume created by the movement of the piston. The combustion gases are removed during the fourth stage.

The four times require two complete rotations of the crankshaft. The torque on the crankshaft generated by a four-stroke engine cylinder is almost zero during the fourth and first stroke, slightly resistant during the second stroke, and strongly engine during the third stroke. This torque pulse generates an output torque of the crankshaft which oscillates around its average, these variations having to be managed by the kinematic chain of transmission downstream of the engine.

The forces transmitted by each connecting rod to the associated crank of the crankshaft correspond to the forces applied to the corresponding cylinder: on the one hand, the pressure exerted in the combustion chamber has a non-zero resultant which is exerted on the cylinder head , in the axis of translation of the piston; on the other hand, the forces exerted by the piston on the jacket generate forces having a substantially radial resultant with respect to the axis of translation of the piston, and oriented, with respect to this axis, opposite to the connecting rod. Forces substantially opposite to the previous ones are applied by the crankshaft to the guide bearings formed in the engine block. All the forces applied to the engine block have a non-zero result and generate torques, this result and these torques being taken up by the supports of the engine block. We will focus in the following essentially on two components transmitted to the engine block supports: on the one hand, the component of the result of the forces which is perpendicular to the axis of revolution and parallel to the axis of translation of the pistons, called pestle, and on the other hand the torque exerted around the axis of revolution of the crankshaft, said tilting torque.

The number of combustion events per revolution of the crankshaft determines the order of the main harmonic of the oscillating output torque, of the tilting torque of the engine block on its supports, and of the pestle. For a four-stroke engine, the number of combustion events per revolution of the crankshaft is equal to half the number of cylinders. Thus, a four-stroke four-cylinder engine has a main harmonic of order two; a four-stroke six-cylinder engine has a third-order main harmonic. As the number of cylinders increases, the frequency of the main harmonic of the output torque and the oscillating movement of the motor increases, while the peak-to-peak amplitudes of the tilting torque and the motor torque decrease, which reduces noise and transmitted vibrations. For a two-stroke engine, the number of combustion events per revolution of the crankshaft is equal to the number of cylinders.

The pestle and the tilting torque have the same frequency, but are out of phase with each other, the maximum of the pestle for a cylinder occurring at the start of the third stage, while the maximum of the tilting torque is found in the middle of the third beat.

[0008] Furthermore, the number of cylinders of an internal combustion engine plays an important role in determining the friction and heat rejection characteristics of the engine. For a given displacement, a lower number of cylinders will generally translate into better thermal efficiency and lower friction, these two factors being combined to reduce fuel consumption. Thus, we see in engine manufacturers a desire to reduce the number of cylinders for a given application. As explained above, this desire of engine manufacturers is in conflict with the desire to control vibrations, noises and variations in engine torque.

Technologies exist to mitigate the effects of displacement on the engine supports, but these technologies generally have a frequency range of attenuation which is low, which does not make it possible to manage the entire speed range of the engine. These technologies are usually applied at the lowest resonant frequency in the system. The remaining frequencies generated by the engine operating range remain problematic.

In order to attenuate the vibrations, the engine torque variations and the noise over the entire engine operating range, it was proposed in document US2014230771 to equip the engine with a balancing device said nutation, which comprises a coupler coupled to the engine crankshaft so as to be rotated by the crankshaft about an axis perpendicular to the axis of revolution of the crankshaft, and a rotating mass coupled to the coupler, the latter being configured to drive the rotating mass in synchronization with the crankshaft. The coupler makes it possible to vary the positioning of the rotating mass in order to bring it closer or away from the axis of revolution of the coupler. However, the mechanism is particularly complex and its cost price very high.

PRESENTATION OF THE INVENTION The invention aims to remedy the drawbacks of the state of the art and to propose means for balancing an internal combustion engine.

To do this is proposed, according to a first aspect of the invention, a method of balancing an internal combustion engine with one or more reciprocating pistons, comprising an engine block, a crankshaft rotating relative to the engine block of the internal combustion engine around a geometric reference axis, method according to which:

a first balancing member is rotated with respect to the engine block about a first tilting balancing axis not parallel to the reference tax, at a speed of revolution having a predetermined ratio with the speed of revolution of the crankshaft tilting having a first dynamic unbalance, and is driven in rotation relative to a motor unit about a second tilting balancing axis not parallel to the reference tax, at a revolution speed of equal amplitude and opposite direction at the speed of revolution of the first tilt balancing member, a second tilt balancing member having a second dynamic unbalance, by synchronizing the rotation of the first tilt balancing member and the second tilt balancing member such that the first tilt balancing member and the second tilt balancing member generate together on the engine block at least one tilting compensation torque around the reference geometric axis.

By dynamic unbalance is meant here a distribution of the masses such that the main axis of inertia of the body considered is neither confused with the axis of rotation (total absence of unbalance) nor parallel to the axis of rotation (static unbalance). It can be a torque unbalance, that is to say a mass distribution which results in a main axis of inertia intersecting with the axis of rotation, or a dynamic unbalance not particular, namely an unbalance superimposing a torque unbalance and a static unbalance, characterized by a main axis of inertia not parallel to the axis of rotation and not intersecting with the axis of rotation.

The rotational movement of the first tilt balancing member about the first tilt balancing axis and the rotational movement of the second tilt balancing member around the second tilt balancing axis being guided relative at the engine block, in practice by bearings, the synchronized rotation of the two tilting balancing members generates at the level of the engine block a compensating torque which at least partially compensates for the tilting torque.

In practice, the speed of rotation of the first tilt balancing axis will be equal to the speed of revolution of the crankshaft or a multiple of this speed. The predetermined ratio between the speed of revolution of the crankshaft and the speed of revolution of the tilt balancing members depends on the type of engine. For a four-stroke engine, this ratio will be two, while for a two-stroke engine it will be one. This ensures that the compensation torque generated by the swing tilt balancing members has the same frequency as the tilt torque.

In practice, the first tilt balancing axis and the second tilt balancing axis are symmetrical to each other with respect to a tilt symmetry plane, the tilt symmetry plane being preferably perpendicular to the geometric reference axis. This avoids generating on the engine block torques or resulting forces not useful for compensating for the tilting torque, and in particular a force in the direction of the reference axis.

Preferably, the first tilt balancing axis and the second tilt balancing axis are parallel. This avoids generating a force resulting in the direction of the bisector of the two tilting balancing axes.

According to a particularly advantageous embodiment, the first tilting balancing member is driven using a first kinematic connection between the crankshaft and the first tilting balancing member and the second is driven tilt balancing member using a second kinematic connection between the crankshaft and the second tilt balancing member or between the first tilt balancing member and the second tilt balancing member. In practice, the first kinematic connection may include a bevel gear.

The alternative consisting in providing an independent auxiliary drive, controlled by speed by measuring the speed of revolution of the crankshaft, for example by electric motor, remains possible.

According to one embodiment, the rotation of the first tilt balancing member and of the second tilt balancing member is synchronized so that the tilting compensation torque generated is maximum, in absolute value, over a turn of the first tilt balancing member, when the crankshaft is in a predetermined angular position making an angle a with respect to a reference position of the crankshaft, which is a top dead center position of a cylinder of the combustion engine internal. The angle a is preferably between 0 ° and 50 ° in the direction of revolution of the crankshaft. The angle a is located around the maximum force of the piston on the jacket due to combustion, and is generally between 20 ° and 40 °, typically about 30 °. This angle is dependent on the combustion parameters and the kinematics of the mobile coupling. We can choose a constant angle a value, which will be a good compromise between the different conditions of use encountered, or we can vary the angle a to refine the attenuation of the tilting torque in all cases Operating. Thus, according to a preferred embodiment, the angle a is modulated as a function of one or more phasing parameters from among the operating parameters of the engine. The phasing parameter or parameters preferably comprise at least one of the following parameters: the speed of revolution of the crankshaft, the engine torque, the power delivered by the engine, the angle of the crankshaft corresponding to a maximum pressure in a cylinder of the engine, or the maximum value of the pressure reached in an engine cylinder. The cylinder pressure information can be estimated or measured.

Preferably, the first tilt balancing axis and the second tilt balancing axis are located in a geometric tilt balancing plane, preferably parallel to the geometric reference tax. Preferably, and to take account of the bulk of the other elements of the propulsion kinematic chain, the geometrical tilt balancing plane is distant from the geometrical axis of revolution of the crankshaft.

According to a preferred embodiment, the first tilt balancing member comprises a first tilt balancing mass having a first value Mi and a first center of gravity located on a first side of a reference plane containing the reference axis and perpendicular to the tilting balancing plane, at a distance £> i from the reference plane and at a distance E \ from the first tilting balancing axis, the second tilting balancing member comprises a second tilting balancing mass having a second value M ₂ and a second center of gravity located on the first side of the reference plane at a distance £> 2 from the reference plane and at a distance £ 2 from the second balancing axis tilt, and Ton is varied at least one of the distances £> i, £ 1, and at least one of the distances £> 2, £ 2, according to one or more control parameters among operational parameters ent of the engine. Preferably, you vary £> i, £ 1, £> 2, and / or £ 2, so that:

My. D _r . Ey = M ₂ D ₂ . E ₂ In practice, and according to a particularly advantageous embodiment, the value Mi of the first tilt balancing mass and the value M2 of the second tilt balancing mass are equal, and the first mass d tilting balancing and the second tilting balancing mass are symmetrical to each other with respect to the plane of symmetry of tilting. Symmetry is thus obtained for the balancing mechanism, which for example makes it possible to minimize the number of different parts and the development time.

According to different embodiments, the control parameter or parameters include one or more of the following parameters: the revolution speed of the crankshaft, the engine torque, the engine power, the angle of the crankshaft corresponding to a maximum pressure in a engine cylinder, the maximum value of the pressure reached in an engine cylinder, and preferably at least the speed of revolution of the crankshaft in combination with at least one second parameter from the following parameters: engine torque, engine power, l angle of the crankshaft corresponding to a maximum pressure in an engine cylinder, the maximum value of the pressure reached in an engine cylinder. In practice, several parameters can be retained and used simultaneously or not.

According to one embodiment, the distance between the first tilt balancing mass and the first tilt balancing axis is varied as a function of the engine operating parameter or parameters. However, this solution requires a significant amount of energy to overcome the centrifugal force generated by the rotation of the tilting balancing member. According to a preferred alternative embodiment, the distance between the first tilt balancing mass and the first tilt balancing axis is constant, and the distance between the first tilt balancing mass and the reference plane as a function of the engine operating parameter (s). This minimizes the energy required to move the first balancing mass, since there is no need to overcome the centrifugal force. The mechanism ensuring the displacement of the first tilting balancing mass is relatively simple, since it only requires a degree of freedom of translation. Naturally, provision is made to vary the position of the second tilt balancing mass in a manner similar to the first tilt balancing mass.

According to a particularly advantageous embodiment, the distance between the first tilting balancing mass and the reference plane is varied so that, for each revolution speed of the crankshaft in a given range of revolution speeds , the distance between the first tilt balancing weight and the reference plane is a monotonic function of the engine torque. This embodiment results in a tilting torque which is itself a monotonous function, in this case an increasing function, of the engine torque, at constant speed of revolution of the crankshaft.

According to a particularly advantageous embodiment, the distance between the first tilt balancing mass and the reference plane is varied so that, for each engine torque in a given range of engine couples, the distance between the first tilting balancing mass and the reference plane is a monotonic function of the speed of revolution of the crankshaft, at least when the speed of revolution of the crankshaft is greater than a predetermined speed threshold. This embodiment results in a tilting torque which is itself, in algebraic value, a monotonic function, in this case a decreasing function, of the speed of revolution of the crankshaft at constant engine torque, at least when the speed of revolution is greater than the predetermined threshold. In this case, the tilting torque is a function of the square of the speed of revolution, and the balancing torque generated by the first tilting balancing mass increases with the square with the speed of revolution of the first mass of tilt balancing around the first tilt balancing axis, at a given distance from the reference plane, and linearly as a function of the distance to the reference plane, at speed of revolution of the first given balancing mass.

In practice, the two laws of variation of the distance between the first balancing mass and the reference plane as a function of the torque at constant revolution speed of the crankshaft and as a function of the revolution speed of the crankshaft at constant torque are used together, defining a law with two variables.

For a four-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the constant ratio between the speed of revolution of the first tilt balancing shaft and the speed of revolution of the crankshaft is preferably equal to N / 2.

ίο For a two-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the constant ratio between the speed of revolution of the first tilt balancing shaft and the speed of revolution of the crankshaft is preferably equal to N.

To symmetrize the balancing system and prevent it from generating a result of parasitic forces adding, where appropriate, to the pestle of the engine, it can advantageously be provided that the first tilt balancing member comprises a third mass d tilting balance of value M3 having a center of gravity located on a second side of the reference plane opposite to the first side, at a distance £> 3 from the geometric reference plane and at a distance £ 3 from the first balancing axis of and that the second tilting balancing member comprises a fourth tilting balancing mass of value M4 having a center of gravity situated on the second side of the reference plane at a distance £> 4 from the geometric reference plane and at a distance E4 from the second tilt balancing axis. In this hypothesis, at least one of the distances £> 3, £ 3, and at least one of the distances £> 4, £ 4, is varied as a function of one or more control parameters from among operating parameters of the engine, preferably so that:

M3-D ₃ .E ₃ = M ₄ .D ₄ .E ₄ When the crankshaft is in the predetermined angular position previously defined, the third tilt balancing mass and the fourth tilt balancing mass are preferably located on one side of the tilt balancing plane opposite to the first tilt balancing mass and to the second tilt balancing mass.

According to a particularly advantageous embodiment, Ton varies the distance between the third tilting balancing mass and the reference plane and / or the distance between the third tilting balancing mass and the first axis of tilt balancing according to the engine operating parameter (s).

According to one embodiment, the distance between the third tilting balancing mass and the reference plane and the distance between the third tilting balancing mass and the first tilting balancing axis are fixed.

Preferably, the values Mi and M3 are equal and the distance between the third tilting balancing mass and the reference plane and / or the distance between the third tilting balancing mass and is varied. the first tilting balancing axis so that the distance between the third tilting balancing mass and the reference plane is equal to the distance between the first tilting balancing mass and the reference plane and the distance between the third tilting balancing mass and the first tilting balancing axis is equal to the distance between the first tilting balancing mass and the first tilting balancing axis.

According to one embodiment, the third tilting balancing mass is positioned in such a way that the projection of the first tilting balancing mass on the reference plane and the projection of the third balancing mass of tilt on the reference plane are aligned with the projection of the first tilt balancing axis on the reference plane.

More generally, the values Mi and M3 and the distances £> i, Ei, and £> 3, E3, are preferably such that the set of two masses Mi and M3 has a main axis of inertia intersecting with the first tilt balancing axis. Similarly, the values M2 and M4 and the distances £> 2, E2, and £> 4, E4, are preferably such that the set of two masses M2 and M4 has a main axis of inertia intersecting with the first axis tilt balancing.

In the case of an in-line engine, the engine block comprises cylinders having axes located in a thrust plane, the reference plane being parallel to the thrust plane.

For internal combustion engines having a pestle force justifying balancing, and in particular for four-stroke four-cylinder engines in line, it can be provided that:

it is rotated around a first pestle balancing axis perpendicular to a pestle reference plane containing the axis of revolution of the crankshaft, at a speed of revolution having a constant relationship with the speed of revolution of the crankshaft , a first pestle balancing mass of value Pi and having a center of gravity located in the pestle reference plane, at a distance F1 from the first pestle balancing axis, it is rotated around a second pestle balancing axis parallel to the first pestle balancing axis and located at a distance from the first pestle balancing axis, at a speed of revolution of equal amplitude and in opposite direction to the speed of revolution of the first pestle balancing mass, a second pestle balancing mass of value Ps and having a center of gravity located in the pestle reference plane at a distance Fs from the second pestle balancing axis, in man ere that:

Pi-F ₁ = P ₂ .F ₂ The two pestle balancing masses, which rotate in opposite directions in the reference plane around the pestle balancing axes, together generate forces whose result is located in the pestle balancing plane, perpendicular to the reference axis, and therefore parallel to the force vector of the pestle to be balanced.

According to a preferred embodiment, the first pestle balancing mass and the second pestle balancing mass are symmetrical to each other with respect to a pestle plane of symmetry perpendicular to the axis revolution of the crankshaft. A desired symmetry is thus obtained for the balancing mechanism, which makes it possible, for example, to minimize the number of different parts and the development time.

The pestle balancing axes can be distant from the tilt balancing axes without harming the quality of the balancing performed. However, and according to a preferred embodiment, it can advantageously be provided that the first pestle balancing axis is merged with the first tilt balancing axis and that the second pestle balancing axis is merged with the second axis tilt balancing, which ensures greater compactness of the balancing mechanism.

For certain engines, and in particular four-stroke four-cylinder engines, it is known that the pestle and the tilting torque vary at the same frequency. It is then advantageously possible to provide that the first pestle balancing mass is rotated at a speed of revolution equal to the speed of revolution of the first tilt balancing member, and that the second is rotated pestle balancing mass at a speed of revolution equal to the speed of revolution of the second tilt balancing member.

There is preferably a non-zero constant phase shift between the rotation of the first pestle balancing mass around the first pestle balancing axis and the rotation of the first tilt balancing mass around the first d axis. 'tilting balancing, this phase shift preferably being between 50 ° and 130 °.

According to a particularly advantageous embodiment, the first pestle balancing mass is an integral part of the first tilt balancing member and the second pestle balancing mass is an integral part of the second tilt balancing member . Thus, the first pestle balancing mass is integral with the first tilting balancing mass in rotation about the first tilt balancing axis merged with the first pestle balancing axis. In practice, the first pestle balancing mass and the first tilting balancing mass are supported by a common rotary body or shaft embodying the first balancing axis. Similarly, the second pestle balancing mass is integral with the second tilting balancing mass in rotation about the second tilt balancing axis coincident with the second pestle balancing axis. In practice, the second pestle balancing mass and the second tilting balancing mass are supported by a common rotary body or shaft embodying the second balancing axis. This produces a device for balancing the tilting torque and the particularly compact pestle.

According to a particularly important embodiment in practice, the internal combustion engine is an in-line engine, preferably a four-stroke engine with four in-line cylinders, the pistons moving parallel to the reference plane of the combustion engine. internal. In this hypothesis, it can advantageously be provided that:

is driven in rotation about the first tilt balancing axis, at a speed of revolution equal to twice the speed of revolution of the crankshaft, a first pestle balancing mass of value Pi and having a center of gravity located in the reference plane of the internal combustion engine, at a distance Fi from the first tilt balancing axis, one rotates around the second tilt balancing axis, at a revolution speed of equal amplitude and opposite direction to the speed of revolution of the first pestle balancing mass, a second pestle balancing mass of value P ₂ and having a center of gravity located in the reference plane of the internal combustion engine at a distance F ₂ of the second tilt balancing axis, so that:

p ₁ .f ₁ = p ₂ .f ₂ In practice, the first pestle balancing mass and the second pestle balancing mass are preferably symmetrical to each other with respect to a plane of pestle symmetry perpendicular to the axis of revolution of the crankshaft.

Preferably there is a non-zero constant phase shift between the rotation of the first pestle balancing mass and the rotation of the first tilt balancing mass around the first tilt balancing axis, this phase shift being preferably between 50 ° and 130 °.

According to another aspect of the invention, it relates to a method of balancing an internal combustion engine with one or more reciprocating pistons, comprising a crankshaft rotating around a geometric reference axis, method by which:

is driven in rotation about a first tilt balancing axis not parallel to the reference tax, at a speed of revolution having a predetermined ratio with the revolution speed of the crankshaft, a first tilt balancing member having a first main axis of inertia not parallel to the first tilt balancing axis, and one rotates about a second tilt balancing axis not parallel to the reference tax, at a revolution speed of equal amplitude and opposite to the speed of revolution of the first tilt balancing member, a second tilt balancing member having a second main axis of inertia not parallel to the second tilt balancing axis, by synchronizing the rotation of the first tilt balancing member and second tilt balancing member such that the first main axis of inertia and the second th main axis of inertia are symmetrical to each other with respect to a tilting plane of symmetry perpendicular to the geometric reference axis.

Preferably, the first tilt balancing axis and the second tilt balancing axis are symmetrical to each other with respect to the tilting plane of symmetry. Preferably, the first tilt balancing axis and the second tilt balancing axis are located in a geometric tilt balancing plane parallel to the geometric reference tax. Preferably, and to take account of the bulk of the other elements of the propulsion kinematic chain, the geometrical tilt balancing plane is distant from the geometrical axis of revolution of the crankshaft.

According to another aspect of the invention, it relates to a balancing system capable of implementing the method described above. In particular, the invention relates to a balancing system for balancing an internal combustion engine with one or more reciprocating pistons comprising an engine block, a crankshaft rotating relative to the engine block of the internal combustion engine about an axis reference geometry, the balancing system comprising:

a first rotary tilt balancing member able to be guided in rotation relative to the engine block around a first tilt balancing axis not parallel to the reference axis, and having a first dynamic unbalance with respect to the first tilting balancing axis, a second rotating tilting balancing member able to be guided in rotation relative to a motor unit around a second tilting balancing axis not parallel to the reference axis, and having a second dynamic unbalance with respect to the second tilt balancing axis, a synchronized drive mechanism for driving the first tilt balancing member in rotation about the first tilt balancing axis at a speed of revolution having a ratio predetermined constant with the revolution speed of the crankshaft, and to drive the second tilting balancing member in auto rotation ur of the second tilt balancing axis at a speed of revolution of equal amplitude and opposite direction to the speed of revolution of the first tilt balancing member, so that the first tilt balancing member and the second tilting balancing member together generate on the engine block a tilting compensation torque around the geometric reference axis.

For a four-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the predetermined constant ratio between the speed of revolution of the first tilt balancing shaft and the speed of revolution of the crankshaft is preferably equal to N / 2.

For a two-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the predetermined constant ratio between the speed of revolution of the first tilt balancing shaft and the speed of revolution of the crankshaft is preferably equal to N.

In practice, the balancing system comprises one or more first bearings to guide the first rotary tilt balancing member in rotation about the first tilt balancing axis with respect to the engine block, and one or more second bearings for guiding the second rotary tilt balancing member in rotation about the second tilt balancing axis with respect to the engine block. If necessary, the bearings can be secured directly to the engine block. Alternatively, the balancing system can be provided with a frame to be fixed to the engine block, and in which the bearings are integrated.

According to one embodiment, the synchronized drive mechanism comprises a transmission mechanism capable of establishing a kinematic connection between the crankshaft and the first tilt balancing member and a kinematic connection between the crankshaft and the second member d 'tilting balancing or between the first tilting balancing member and the second tilting balancing member. The transmission mechanism can be of any suitable type, comprising for example belts, chains, cogwheels and / or countershafts. If a kinematic connection between the first balancing member and the second balancing member is provided, the latter must reverse the direction of rotation between the two balancing members. This solution with a single kinematic link between the balancing system and the crankshaft makes it possible to integrate the kinematic link between the first tilt balancing member and the second tilt balancing member to a common frame of the balancing system . It is then possible to propose a compact and unitary balancing system, the mounting of which to the motor is easy.

According to one embodiment, the balancing system comprises modulation means capable of modulating a phase shift between the rotation of the first tilt balancing member around the first tilt balancing axis and the rotation of the crankshaft around the reference geometric axis, preferably as a function of one or more phasing parameters among the engine operating parameters, the phasing parameter (s) preferably comprising at least one of the following parameters: the speed of revolution of the crankshaft, the engine torque, the power delivered by the engine, the angle of the crankshaft corresponding to a maximum pressure in an engine cylinder, the maximum value of the pressure reached in an engine cylinder. These modulation means make it possible to calibrate the maximum of the compensation torque generated by the balancing system on the maximum of the tilting torque generated by the engine, the phase of which with respect to the rotation of the crankshaft varies with the engine speed, the torque engine, and other parameters, such as, for example, outside temperature, coolant temperature, etc.

According to a particularly important embodiment in practice, the balancing system comprises control means for varying the first dynamic unbalance and the second dynamic unbalance. According to different methods of implementation, if applicable cumulative:

the control means include at least one speed sensor of revolution of the crankshaft to determine at least one of the control parameters;

the control means comprise one or more torque or torque demand sensors or estimators for determining at least one of the control parameters;

the control means include one or more engine cylinder pressure sensors or estimators to determine at least one of the control parameters.

According to one embodiment, the first tilting balancing member comprises a first tilting balancing mass movable at least in a direction parallel to the first tilting balancing axis, the second tilting balancing member comprises a second tilting balancing mass movable at least in a direction parallel to the second tilting balancing axis, and the control means are adapted to vary the position of the first tilting balancing mass parallel to the first axis tilt balancing and the position of the second tilt balancing mass parallel to the second tilt balancing axis according to the control parameter (s).

According to one embodiment, the first tilt balancing member has a center of gravity located on the first tilt balancing axis, and the second tilt balancing member has a center of gravity located on the second tilt balancing axis. In other words, the two tilting balancing members have no static unbalance with respect to their respective axes of rotation. This arrangement will be preferred for an engine which does not generate a pestle force, or whose pestle force has a fundamental frequency which is not equal to that of the tilting torque.

Alternatively, provision may be made for the first tilting balancing member to have a center of gravity located at a distance from the first tilting balancing axis, and the second tilting balancing member to have a center of gravity located at distance from the second tilt balancing axis. In other words, the two tilting balancing members have a static unbalance relative to their respective axis of rotation. This arrangement will be particularly advantageous for an engine generating a pestle force with a fundamental frequency which is equal to that of the tilting torque, and in particular for a four-stroke engine with four cylinders in line.

According to another aspect of the invention, it relates to an internal combustion engine with one or more reciprocating pistons comprising an engine block, a crankshaft rotating relative to the engine block of the internal combustion engine around a geometric reference axis of the engine block contained in a reference plane of the engine block, and a balancing system as described above.

Preferably, the first tilt balancing axis and the second tilt balancing axis are symmetrical to each other with respect to a plane of tilt symmetry, preferably perpendicular to the geometric axis. reference.

According to a particularly advantageous embodiment in practice, the first tilt balancing axis and the second tilt balancing axis are parallel, and preferably perpendicular to the reference plane. This minimizes the parasitic forces and torques generated by the rotation of the tilting balancing members.

Preferably, the first tilt balancing axis and the second tilt balancing axis are located in a geometric tilt balancing plane, preferably parallel to the geometric reference axis.

Preferably, the first tilt balancing member has a center of gravity located in the reference plane and the second tilt balancing member has a center of gravity located in the reference plane.

In particular, the invention is applicable to a four-stroke four-cylinder in-line engine, the balancing system being such that the first tilt balancing member has a center of gravity located at a distance from the first axis tilt balancing, and the second tilt balancing member has a center of gravity located at a distance from the second tilt balancing axis. In this hypothesis, a thrust plane is defined containing the axes of the engine cylinders, which is preferably parallel or coincident with the reference plane defined above. The tilting balancing geometric plane is preferably perpendicular to the thrust plane. The center of gravity of the first tilt balancing member and the center of gravity of the second tilt balancing member are preferably located in the thrust plane.

According to another aspect of the invention, it relates to a tilting balancing member intended to be integrated into an internal combustion engine for the implementation of the method described above.

In particular, the invention relates to a balancing member for balancing an internal combustion engine, the balancing member comprising:

a body provided with one or more bearings for guiding the body in rotation about a balancing axis of the balancing member, the body having a center of gravity situated in a reference plane of the body perpendicular to the balancing axis, and at least a first balancing mass having a value Mi and a center of gravity located at a non-zero distance E \ from the balancing axis, fixed in rotation relative to the body around the balancing axis and guided relative to the body so that the distance £> i from the center of gravity of the first balancing mass to the body's reference plane is variable.

By varying the distance Di, there is a means of varying the lever arm of the balancing member with respect to an axis located in the reference plane.

Preferably, the first balancing mass is guided relative to the body so that the non-zero distance Ei remains constant. We can then move the first tilting balancing mass with a low control energy, because we do not have to oppose centrifugal forces induced by the rotation of the balancing member around the axis d 'balancing.

Preferably, the balancing member further comprises a control device for varying the distance £> i in a range of values between two end-of-travel positions. This control device can be of any suitable type, for example mechanical, electromechanical or hydraulic.

According to a particularly advantageous embodiment, the control device comprises a variable-volume hydraulic chamber, one wall of which is formed by the first balancing mass. In practice, the control device preferably comprises a hydraulic supply circuit between at least one of the bearings and the variable volume hydraulic chamber. This hydraulic circuit preferably comprises channels, bores and / or grooves made in the body of the balancing member, and if necessary in the first balancing mass.

According to one embodiment, the control device comprises an actuator, the actuator having an output member movable relative to the body, preferably in translation parallel to the balancing axis, the control device comprising preferably a rotary connecting stop between the outlet member and the first balancing mass. The rotary stop allows decoupling in rotation around the balancing axis between the body of the balancing member and the actuator. The actuator can thus be mounted fixed in rotation on the support structure on which are also mounted the guide bearing or bearings in rotation of the body. This simplifies the transmission of power to the actuator, whether hydraulic, pneumatic, electric or mechanical.

The actuator can be a unidirectional actuator and the control device also comprises a system for returning the first balancing weight to one of the two end-of-travel positions. It is then the unidirectional actuator which brings the energy to charge the return system by moving the first balancing mass in one direction, and the return system which restores the stored energy to bring back the first balancing mass in the opposite way. The return system may in particular comprise one or more return springs, in particular tension or compression springs, or pneumatic springs. Other reminder systems can also be considered.

Alternatively, the actuator can be a bidirectional actuator, which performs the mechanical work necessary to move the balancing mass alternately in the two opposite directions.

According to one embodiment, the body is equipped with a drive member in rotation about the balancing axis, preferably an annular toothing, intended to mesh with a similar toothing of a second member d balancing or with a transmission member such as a toothed wheel, a belt or a transmission chain.

According to one embodiment, the center of gravity of the body is located on the balancing axis. In other words, the body does not generate any imbalance around the balancing axis.

According to an alternative embodiment, the center of gravity of the body is located at a distance from the balancing axis, the center of gravity of the body preferably being located in a plane containing the balancing axis making a angle δ relative to a plane containing the balancing axis and the center of gravity of the first balancing mass, the angle δ preferably being between 50 ° and 130 °. This unbalance allows balancing of a pestle effort of the internal combustion engine.

The first balancing mass may include one or more parts. According to a particularly advantageous embodiment, the first balancing mass comprises one or more rolling bodies, the body of the balancing member being provided with one or more raceways oriented and turned towards the balancing axis. , the rolling body or bodies being able to roll on the raceway (s) parallel to the balancing axis when the distance £> i from the center of gravity of the first balancing mass to the reference plane varies.

By moving the first balancing mass by rotation, the energy necessary for its positioning in the body of the balancing member is minimized.

The rolling bodies, which may be balls or rollers having a rolling surface constituted by a straight generator (cylindrical or conical rollers), convex (barrel rollers) or concave, are preferably made of a material of high density , in particular in metal, in particular in steel, to constitute a preponderant proportion of the balancing mass. According to a particularly preferred embodiment, the rolling body or bodies are rollers, preferably rollers with generators in an arc of a circle.

Or the raceways have a profile suitable for rolling bodies, which can be flat, concave or convex. According to a particularly simple embodiment, the raceway (s) are cylindrical and concave, and preferably with a center of curvature positioned on the balancing axis.

Preferably, the balancing member comprises a guide cage of the rolling body or bodies, movable in translation parallel to the balancing axis when the rolling body or bodies roll on the race track or tracks.

The guide cage may include, for each rolling body, a housing cell, closed or open. The cage may also include guide surfaces cooperating with complementary guide surfaces formed on the body of the balancing member, in order to guide the cage in translation in the body of the balancing member.

According to one embodiment, the guide cage is driven in translation by the actuator, and constitutes a connection between the actuator and the rolling body or bodies. In particular, provision could be made to integrate a rotary abutment connecting with the actuator into the cage.

Preferably, there is a plane of symmetry of the balancing cage and more generally of the balancing mass, passing through the balancing axis.

Preferably, the number of rolling bodies is even, and the rolling bodies are arranged in equal number on either side of a plane containing the balancing axis.

Preferably, the guide cage has a center of gravity located on the balancing axis, so that the guide friction between the guide cage and the rolling body or bodies is independent of the speed of rotation of the balancing member.

Preferably, the balancing member further comprises at least one additional balancing mass having a value M3 and a center of gravity located at a non-zero distance Es from the first balancing axis, fixed in rotation by relation to the body around the balancing tax and guided with respect to the body so that the distance £> 3 from the center of gravity of the additional balancing mass to the reference plane of the body is variable. In practice, it may advantageously be provided that the first balancing mass and the additional balancing mass have the same value. One or more of the following arrangements may also be provided:

the additional balancing mass is guided relative to the body so that the non-zero distance E3 remains constant;

the balancing member further comprises a control device for varying the distance £> i in a range of values between two end positions and varying the distance £> 3 in a range of values between two positions additional limit switches;

the first additional balancing mass comprises one or more rolling bodies, the body of the balancing member being provided with one or more additional rolling tracks oriented and facing the balancing axis, the rolling body or bodies of the additional balancing mass being able to roll on the additional raceway (s) parallel to the balancing axis when the distance DI from the center of gravity of the additional balancing mass to the reference plane C varies.

the balancing member comprises an additional guide cage of the rolling body or bodies, of the additional balancing mass, the additional guide cage being movable in translation parallel to the balancing axis when the rolling body or bodies the additional balancing weight rolls on the additional raceway (s).

the additional guide cage has a center of gravity located on the balancing axis.

the additional guide cage is symmetrical with respect to a plane of symmetry passing through the balancing axis.

the additional raceway (s) are cylindrical and concave, and preferably with a center of curvature positioned on the balancing axis.

the rolling body or bodies of the additional balancing mass are rollers, preferably rollers with generators in an arc of a circle.

the balancing member further comprises a return spring for the additional balancing mass towards one of its two additional end-of-travel positions;

the control device comprises a variable volume hydraulic chamber, a wall of which is formed by the additional balancing mass and, preferably, a hydraulic supply circuit between one of the bearings and the variable volume hydraulic chamber;

coordination means, for example mechanical, electromechanical or hydraulic, are provided to coordinate the movements of the first balancing mass and of the additional balancing mass, for example to ensure equality between Di and £> 3.

According to one embodiment, the control device is such that, for at least certain values of Di and £> 3, the center of gravity of the additional balancing mass is on the side of the reference plane of the body opposite the center of gravity of the first balancing mass.

Preferably, the center of gravity of the additional balancing mass and the center of gravity of the first balancing mass are in a balancing plane containing the balancing axis, and both 'another of a plane perpendicular to the balancing plane and containing the balancing axis.

What has been said about the structure of the balancing mass also applies to the additional balancing mass.

According to another aspect of the invention, it relates to a tilting balancing system comprising a first tilting balancing member constituted by a balancing member as described above, guided in rotation around a first tilting balancing axis and a second tilting balancing member constituted by a balancing member as described above, guided in rotation about a second coplanar and preferably parallel tilting balancing axis to the first tilt balancing axis. According to one embodiment, the tilting balancing system comprises a frame in which the bearings are housed, at least one member for balancing the chambers and pipes necessary for the action of the control device, see, where appropriate, the control device itself, and which makes it possible to consider the balancing system as a structural and functional unit, which can be attached to an engine block.

Preferably, the tilting balancing system further comprises a kinematic connection between the body of the first tilting balancing member and the body of the second tilting balancing member, such as a rotation of the first member tilt balancing around the first tilt balancing axis in a given direction at a given speed causes a rotation of the second tilt balancing member around the second tilt balancing axis in the opposite direction and at the same speed .

The characteristics disclosed in connection with one aspect of the invention or in one embodiment can naturally be combined with other characteristics disclosed in connection with other aspects of the invention or other embodiments, to make variants.

BRIEF DESCRIPTION OF THE FIGURES Other characteristics and advantages of the invention will emerge on reading the description which follows, with reference to the appended figures, which illustrate:

Figure 1, a block diagram of a system for balancing the tilting torque and an associated internal combustion engine, according to a first embodiment of the invention;

FIG. 2, a block diagram of a system for balancing the tilting torque and an associated internal combustion engine, according to a second embodiment of the invention;

FIG. 3A, a diagram illustrating the variations in the tilting torque as a function of the speed of revolution of the crankshaft when empty, at half load and at full load, of an internal combustion engine with regular ignition;

FIG. 3B, a diagram illustrating the variations of the tilting torque as a function of the engine load, that is to say of the engine torque reduced to the maximum torque, for different speeds of revolution of the crankshaft for the same internal combustion engine with regular ignition;

FIG. 3C, a diagram illustrating the variation of the tilting torque as a function of the angle of the crankshaft, at no load and at half load, for a revolution speed of the crankshaft in the middle of the engine speed range;

Figure 4, a block diagram of the balancing system of Figure 2 provided with a control device according to a first embodiment, executing a first strategy of compensation of the tilting torque, in a torque compensation position of maximum engine tilt;

FIG. 5, a block diagram of the balancing system provided with the control device according to FIG. 4, in a position for compensating for intermediate tilting torque;

Figure 6, a block diagram of the balancing system provided with the control device according to Figure 4, in a neutral position;

FIG. 7, a block diagram of the balancing system provided with the control device according to FIG. 4, in a position of compensation for tilting torque generated mainly by the inertial forces;

Figure 8, a block diagram of the balancing system of Figure 2 provided with a control device according to a second embodiment, executing a second strategy of compensation of the tilting torque, in a torque compensation position of motor tilting;

Figure 9, a block diagram of the balancing system provided with the control device according to Figure 8, in a neutral position;

FIG. 10, a block diagram of the balancing system provided with the control device according to FIG. 8, in a position of compensation for tilting torque generated mainly by the inertial forces;

FIG. 11, a block diagram of the balancing system in FIG. 2 provided with a control device according to a third embodiment, executing a third strategy for compensating the tilting torque, in a torque compensation position of maximum engine tilt;

Figure 12, a block diagram of the balancing system provided with the control device according to Figure 11, in an intermediate tilting torque compensation position;

Figure 13, a block diagram of the balancing system provided with the control device according to Figure 11, in a neutral position;

FIG. 14, a block diagram illustrating a position of the crankshaft of the internal combustion engine associated with a system for compensating the tilting torque according to FIG. 2, corresponding to a maximum on one revolution of the tilting torque, and the corresponding position of the compensation system;

FIG. 15, a device for modulating the phase shift between the rotation of the crankshaft of the internal combustion engine and the rotation of the tilting torque compensation system;

Figures 16 to 22, various possible positions of the tilting torque compensation system relative to the engine block;

FIGS. 23 to 27 various configurations of motors incorporating a system for compensating the tilting torque according to the invention;

FIG. 28, a block diagram of a system for balancing the tilting torque and the pestle force and of an associated in-line four-cylinder four-stroke engine according to an embodiment of the invention;

FIG. 29, a block diagram of a system for balancing the tilting torque and the pestle force and of an associated in-line four-cylinder four-stroke engine, according to another embodiment of the invention;

FIG. 30, a block diagram of an angular offset between a mass for balancing the pestle force and masses for balancing the tilting torque;

FIG. 31, an isometric view of a member for balancing the tilting torque and the pestle force according to an embodiment of the invention, in a neutral position;

FIG. 32, a side view of the member for balancing the tilting torque and the pestle force of FIG. 31;

FIG. 33, a front view of the member for balancing the tilting torque and the pestle force of FIG. 31;

FIG. 34, a view of the member for balancing the tilting torque and the pestle force of FIG. 31 in section along the cutting plane B-B illustrated in FIG. 33;

FIG. 35, a view of the member for balancing the tilting torque and the pestle force of FIG. 31 in section along the cutting plane A-A of FIG. 32;

FIG. 36, a view of the member for balancing the tilting torque and the pestle force of FIG. 31 in section along the cutting plane J-J illustrated in FIG. 32;

FIG. 37, a detailed view of a part of the member for balancing the tilting torque and the pestle force of FIG. 31;

FIG. 38, a view of the member for balancing the tilting torque and the pestle force of FIG. 31 in section along the plane BB illustrated in FIG. 33, but in a position of maximum compensation for the torque of tilting;

FIG. 39, an isometric view of a balancing system according to another embodiment of the invention;

Figure 40, a top view of part of the balancing system of Figure 39;

Figure 41, a sectional view of the balancing system of Figure 39, along a section plane XLI-XLI defined in Figure 40;

Figure 42, a sectional and partial view of the balancing system of Figure 39, along a section plane I-I defined in Figure 41;

Figure 43, a sectional and partial view of the balancing system of Figure 39, along a section plane G-G defined in Figure 41;

Figure 44, a sectional and partial view of the balancing system of Figure 39, along a section plane J-J defined in Figure 41;

Figure 45, an isometric view of a guide cage of the balancing system of Figure 39;

FIG. 46, an isometric view of part of a balancing system according to another embodiment of the invention;

FIG. 47, an isometric view of certain components of the balancing system of FIG. 46;

FIG. 48, an isometric view of certain components of the balancing system of FIG. 46;

Figure 49, a sectional view of the balancing system of Figure 46;

Figure 50, a sectional view of the balancing system according to another embodiment of the invention, in a section plane L-L illustrated in Figure 51;

Figure 51, a sectional view of the balancing system of Figure 50, in a section plane LI-LI illustrated in Figure 50;

Figure 52, a sectional view of the balancing system of Figure 50, in a section plane LII-LII illustrated in Figure 51;

Figure 53, an isometric sectional view of a balancing member for a balancing system according to another embodiment of the invention.

For clarity, identical or similar elements are identified by identical reference signs in all of the figures.

DETAILED DESCRIPTION OF EMBODIMENTS In FIG. 1 is illustrated diagrammatically an internal combustion engine with four in-line cylinders 10, comprising an engine block 12, a crankshaft 14 guided in rotation about a reference geometric axis 100 of the engine block 10, one or more pistons 16 guided in translation in cylinders 18 of the engine block 12, and connecting rods 20 connecting the pistons and the crankshaft 14.

Remarkably, the internal combustion engine 10 is equipped with a balancing system 22 comprising a first tilt balancing member 24A and a second tilt balancing member 24B. The first tilt balancing member 24A is constituted by a body 25A guided in rotation relative to the engine block 12 around a first tilt balancing axis 102A via one or more bearings 62A, and a first tilting balancing mass 26A of value Mi fixed in rotation relative to the body 25A. The second tilt balancing member 24B is constituted by a body 25B guided in rotation relative to the engine block 12 around a second tilt balancing axis 102B parallel to the first tilt balancing axis 102A and distant from the first tilting balancing pin 102A by means of one or more bearings 62B, a second tilting balancing mass 26B of value Ms fixed in rotation relative to the body 25B. In this embodiment, the body 25A is balanced relative to the first tilt balancing axis 102A and that the body 25B is balanced relative to the second tilt balancing axis 102B.

The first tilt balancing axis 102A and the second tilt balancing axis 102B are located in a geometric tilt balancing plane Q parallel to the reference axis 100 and perpendicular to a reference plane P of the internal combustion engine 10 containing the geometric reference axis 100. In the schematic illustration of FIG. 1, the internal combustion engine 10 represented is an in-line engine, in this case a four-cylinder engine 18 in line , and the axes 104 of the cylinders 18 of the engine block 12 are located in a thrust plane PP, which can be parallel to the reference plane P, or even, in the illustration of FIG. 1, coincident with the reference plane P.

The first tilt balancing mass 26A has its center of gravity located on a first side of the reference plane P at a distance Di from the reference plane and at a distance Ei from the first tilt balancing axis 102A . The second tilt balancing mass 26B has its center of gravity located on the same first side of the reference plane P, at a distance D2 from the reference plane P and at a distance E2 from the second tilt balancing axis 102B.

Each tilt balancing member 24A, 24B thus has a dynamic unbalance, in the sense that its main axis of inertia is neither coincident with the corresponding tilt balancing axis 102A, 102B, nor parallel to this axis.

Illustrated in Figure 1 drive means 28 which allow to drive the first tilt balancing member 24A in rotation about the first tilt balancing axis 102A at a speed of revolution having a ratio constant with the speed of revolution of the crankshaft 14, and driving the second tilt balancing member 24B in rotation about the second tilt balancing axis 102B at a revolution speed of equal amplitude and in the opposite direction to the speed of revolution of the first tilting balancing member 24A. In this case, these drive means 28 are here constituted by a bevel gear 30 between the crankshaft 14 and a return shaft 32, which drives a first chain 34 for driving the first balancing member of tilt 24A, and by a second drive chain 36 with reversing function between the first tilt balancing member 24A and the second tilt balancing member 24B.

The transmission ratio between the crankshaft 14 and the first tilting balancing member 24A is chosen so that the period of revolution of the latter corresponds to the period of the tilting torque.

Thus, for the internal combustion engine 10 four-stroke four-cylinder 18 in line illustrated in Figure 1, the constant ratio between the speed of revolution of the first tilt balancing member 24A and the speed of revolution of crankshaft 14 is equal to 2. More generally, for an internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the constant ratio between the speed of revolution of the first tilt balancing member 24A and the speed of revolution of the crankshaft 14 is chosen to be equal to N / 2.

In the case of a two-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the constant ratio between the speed of revolution of the first tilt balancing member 24A and the speed of revolution of the crankshaft 14 is chosen equal to N.

The second tilt balancing member 24B rotates in all cases at the same speed as the first tilt balancing member 24A and in the opposite direction.

The embodiment of FIG. 2 differs from that of FIG. 1 by the fact that the first tilt balancing member 24A comprises a third tilt balancing mass 38A of value M3 located, relative to the first tilting balancing mass 26A, on the other side of the plane P and on the other side of the plane Q. Similarly, the second tilting balancing member 24B comprises a fourth tilting balancing mass 38B of value Ms situated, with respect to the second tilting balancing mass 26B, on the other side of the plane P and on the other side of the plane Q.

The third tilting balancing mass 38A has its center of gravity located on a second side of the reference plane P at a distance D3 from the reference plane P and at a distance Es from the first tilting balancing axis 102A. The fourth tilting balancing mass 38B has its center of gravity located on the same second side of the reference plane P at a distance D4 from the reference plane P and at a distance E4 from the second tilting balancing axis 102B .

In all of the embodiments, the first tilting balancing axis 102A and the second tilting balancing axis 102B are symmetrical to each other with respect to a tilting plane of symmetry S which is perpendicular to the axis of revolution 100 of the crankshaft 14, to the reference plane P and to the tilting balancing plane Q.

To explain the operation of the tilt balancing mechanism 22, it will first be assumed that in the embodiments of Figures 1 and 2, the values Mi of the first tilt balancing mass 26A and the value Mz of the second tilting balancing mass 26B are equal, and that the centers of gravity of the first tilting balancing mass 26A and of the second tilting balancing mass 26B are in symmetrical positions one of the other with respect to the plane of symmetry S. Thus, the distances Ei and E2 are equal, and the distances Di and Dz are equal.

[YES] Thus, the joint rotation of the first tilting balancing mass 24A and the second tilting balancing mass 24B around their respective balancing axis 102A, 102B, in opposite directions, generates a force having a result always located in the plane S, perpendicular to the geometric plane of tilt balancing Q at the distance Di from the reference axis 100, which results in a nonzero moment relative to the reference axis 100. This moment has an amplitude which varies periodically, at the period of rotation of the balancing members 24A, 24B which is also that of the tilting torque. There is thus a first means of generating a torque in the opposite direction to the tilting torque.

For the embodiment of FIG. 2, it will also be assumed that the values M3 and M4 of the third tilt balancing mass 38A and of the fourth tilt balancing mass 38B are equal and that the centers of gravity of the third and fourth tilting balancing masses 38A, 38B are in positions symmetrical to each other with respect to the plane of symmetry S. Thus, the distances Es and E4 are equal, and the distances Ds and D4 are equal.

Thus, the joint rotation of the third and fourth tilting balancing masses 38A, 38B about their respective balancing axis, in the opposite direction, generates a force having a result always located in the plane S, perpendicular to the plane tilting balancing geometry Q at the distance Ds from the reference axis 100, which results in a non-zero moment relative to the reference axis 100. This moment has an amplitude which varies periodically, at the rotation period of the balancing members 24A, 24B which is also that of the tilting torque. There is thus a second means of generating a torque in the opposite direction to the tilting torque.

The torque generated by the third and fourth tilt balancing masses 38A and 38B is perfectly combined with the torque generated by the first and second tilt balancing masses 26A, 26B if the third and fourth balancing masses tilt 38A, 38B rotate in phase opposition with respect to the first and second tilt balancing masses 26A, 26B, in other words if the centers of gravity of the four rotating masses simultaneously cross the geometric plane of balancing of tilting Q, the third mass otherwise being always on one side of the geometric plane of balancing tilting Q opposite to the first mass.

If in addition the values Mi and M3 are equal (and consequently if M2 and M4 are equal), the result of the forces induced on the engine block 12 by the rotation of the masses 26A, 26B, 38A, 38B is zero, their only contribution being to generate a torque at the same frequency and in phase opposition with the tilting torque.

In practice, the tilting torque varies in intensity and in direction with the speed of revolution of the crankshaft and with the engine torque. The diagram in FIG. 3A shows: on the abscissa axis the speed of revolution of the crankshaft and on the ordinate axis the tilting torque exerted on the engine block, for three torque regimes, namely at no load, half load and full load. It can be seen that beyond a threshold of engine revolution speed, here above 2000 rpm, the tilting torque is a monotonous function, in this case decreasing, of the crankshaft revolution speed . On the diagram of FIG. 3B are represented: on the abscissa axis the load of the motor, that is to say the engine torque referred to the maximum torque, and on the ordinate axis and in algebraic value, the torque tilting exerted on the engine block, for different speeds of revolution of the crankshaft. Arbitrarily, the tilting torque has been given negative values when it is mainly due to the inertias of the rotating parts, namely beyond a speed threshold for a given torque, or below a torque threshold for a given speed, and positive values when it is mainly due to the combustion pressure, namely below a speed threshold for a given torque, or above a torque threshold for a given speed. It can be seen that the tilting torque is an increasing monotonic function of the engine torque, at least for a predetermined range of speeds of revolution of the crankshaft. FIG. 3C, for its part, illustrates the variations of the tilting torque in amplitude and in phase, for the same revolution speed of the crankshaft, at no load and at 50% of the maximum torque, for a combustion which starts at around 45 ° in relation to the benchmark chosen.

The balancing systems 22 of the embodiments of FIGS. 1 and 2 can therefore be optimized for a given engine speed and torque by appropriately setting the value and the location of the tilting balancing masses 26A, 26B, and if necessary 38A, 38B. But such a system will have a limited benefit, even be counterproductive, for different engine speeds and torques.

It is therefore advantageous to be able to control the amplitude of the torque generated by the balancing system 22 as a function of the engine speed.

Different devices for modifying the balancing torque generated by the balancing system 22 are illustrated in FIGS. 4 to 13.

In FIGS. 4 to 7, the first balancing member 24A of the balancing system of FIG. 2 has been illustrated, comprising a body 25A guided in rotation relative to the first tilting balancing axis 102A, and which will be assumed here to be dynamically balanced, and two balancing masses 26A, 38A integral in rotation with the body 25A around the first tilting balancing axis 102A and equipped with a control device 40 comprising a control mechanism 42 making it possible to move the balancing masses 26A, 38A relative to the body 25A parallel to the first tilting balancing axis 102A so as to vary the distances Di and Di without varying the distances Ei and E3. The control mechanism 42 is driven here by motor means 44 controlled by a control unit 46 as a function of input signals coming from a unit for determining the speed of the crankshaft 48 and from a unit for determining the torque motor 50.

In this embodiment, the distances Di and D3 are equal and the masses 26A, 38A move together. It is understood that, moreover, the second balancing member 24B is also equipped with a similar control mechanism controlled by the control unit 46 to move the two masses 26B, 38B so that they are always in a symmetrical position with respect to to masses 26A, 38A.

In the position of Figure 4, the compensation of the tilting torque related to the combustion pressures is maximum. In the position of FIG. 5, the two masses 26A, 38A have jointly moved towards the axis of revolution 100 of the crankshaft 14 and the compensation torque generated by the masses 26A, 38A, 26B, 38B is less. In the position of FIG. 6, the tilting balancing masses 26A, 38A, 26B, 38B are in the reference plane P, and no longer generate any compensation torque around the axis of revolution 100. Finally, in the position of FIG. 7, the masses 26A, 38A have crossed the plane P, and are in a position allowing compensation for the tilting torque linked to the inertias.

FIGS. 8 to 10 illustrate the first balancing member 24A of the balancing system of FIG. 2, equipped with a control device 40 comprising a control mechanism 42 making it possible to move the balancing masses 26A, 38A radially relative to the first tilt balancing axis (102A) so as to vary the distances Ei and Es without varying the distances Di and Ds. In this embodiment, the distances Ei and Es are equal and the masses 26A, 38A move together. It is understood that moreover, the second balancing member 24B is also equipped with a control mechanism controlled by the control unit 46 to move the two masses 26B, 38B so that they are always in a symmetrical position relative to the masses 26A, 38A.

In the position of Figure 8, the compensation of the tilting torque related to the combustion pressures is maximum. In the position of FIG. 9, the tilt balancing masses 26A, 38A, 26B, 38B are aligned on the first tilt balancing axis 102, and no longer generate any compensation torque around the axis of revolution 100. Finally, in the position in FIG. 10, the masses 26A, 38A are in the mirror position relative to the position in FIG. 8, thus compensating for the tilting torque linked to the inertias.

It is naturally possible to combine the operating modes of Figures 4 to 7 on the one hand, and 8 to 10 on the other hand, allowing to modify both Di, Ds, Ei and Es. It should be noted that in the embodiments of FIGS. 4 to 10, it is not necessary for the values Mi and M3 of the masses 26A and 38A to be equal. A particular solution is given by Μι = M3, which is characterized by the total symmetry of the mechanism. Furthermore, in the event that Ms = 0, FIGS. 4 to 10 become applicable to the embodiment of FIG. 1.

In Figures 11 to 13, there is illustrated the first balancing member 24A of the balancing system of Figure 2, equipped with a control device 40 comprising a control mechanism 42 for moving the mass d balancing 26A parallel to the first tilting balancing axis so as to vary the distance Di, the mass 38A remaining fixed relative to the body 25A and Ds being constant. It is understood that moreover, the second balancing member 24B is also equipped with means making it possible to move the mass 26B, so that it is always in a symmetrical position relative to the mass 26A.

In the position of FIG. 11, the compensation of the tilting torque linked to the combustion pressures is maximum. In the position of FIG. 12, the mass 26A is in the reference plane, and the compensation torque, corresponding to the action of the masses 26A, 38A, is lower. Finally, in the position of FIG. 13, the mass 26A has joined the mass 38A, the effects of the two masses cancel each other out and the compensation torque is zero.

We can define a repository (O, x, y, z) orthonormal rotating linked to the tilting balancing member 24A, originating from O the center of gravity of the tilting balancing member 24A, its y axis coincides with the axis of rotation of the tilting balancing member 24A, and its z axis positioned so that the plane (O, y, z) contains the centers of gravity of the moving masses Mi and M3. In this frame of reference, the inertia matrix M _A of the tilting balancing member 24A can be expressed as follows:

Ηα, χχ 0 Ia, xz \ ι ^ Α, χχ 0 Ma = Î 0 ^ 4, yy ^ A, yz 1 =, ° ^ 4, yy \ / a, zx ^ A, zy ^ Α, ζζ / \ jA, xz ^ A, yz

According to the theory of inertia tensors, this matrix is symmetrical, so that the components of the matrix outside the diagonal, called products of inertia, are equal two by two: Iaj = Iajî · In the hypothesis where the center of gravity of the tilting balancing member 24A is on the axis of rotation, we will also have Ia, χζ— Ια, ζχ = 0.

In this reference system, the inertia products k, yz and kzy vary as a function of the displacements of the moving masses Mi and M3, the other components being unaffected by the displacements of the moving masses Mi and M3.

Similarly, we can define another frame of reference (O, x, y, z) orthonormal rotating linked to the tilting balancing member 24B, originating from O the center of gravity of the member tilt balancing 24B, its y axis coincides with the axis of rotation of the tilt balancing member 24B, and its z axis positioned so that the plane (O, y, z) contains the centers of gravity of the masses M2 and M4 mobiles. In this frame of reference, the inertia matrix of the tilting balancing member 24B can be expressed as follows:

// β, ΧΧ 0 Ib, xz \ i! Β, χχ 0 Ιβ, χζ m _b = I 0 ^ Byy ^ B, yz 1 = “ ^ Byy ^ B, yz \ Jb, ZX ^ B.zy Ιβ, ζζ / XJB.xz ^ B, yz Ιβ, ζζ

Like the previous one, this matrix is symmetrical. In the chosen frame of reference, the kyz and kzy inertia products vary as a function of the displacements of the moving masses M2 and M4, the other components being unaffected by the displacements of the moving masses M2 and M4.

Preferably, we will choose: I _Byz = I _Ayz .

Naturally, it is advantageous to position the tilting balancing masses 26A, 26B, and if necessary 38A, 38B so that the compensation torque which they generate on a revolution of the balancing members 24A , 24B has its maximum which coincides with the maximum of the tilting torque undergone by the engine block.

In Figure 14, there is illustrated a cylinder 18 of the engine 10 after passing through the top dead center and ignition of the gases. The piston 16 then exerts on the cylinder liner 18 a force whose radial component relative to the axis 104 of the cylinder 18 varies among other things as a function of the gas pressure, the resistive torque of the crankshaft and the angle of the connecting rod 20. It is this effort which is at the origin of the tilting torque, and it is understood that its maximum will be reached at a position of the connecting rod which may vary depending on the operating conditions. If we take the crankshaft rotation reference system as the origin of the crankshaft 14 corresponding to the top dead center of a reference piston at the start of engine time, we observe that the radial force on the cylinder liner 18 reaches its maximum after rotation of the crankshaft 14 which can vary between 0 ° and 60 °, preferably between 10 ° and 50 ° depending on the operating conditions. To obtain a compensating effect in all conditions of use of the engine, it will be advantageous to calibrate the system for balancing the tilting torque 22 at an angle a situated in the central part of this range, between 10 ° and 50 ° , preferably between 20 ° and 40 °, typically at 30 ° or around this value. In practice, this means that the tilt balance weights 26A, 26B, and if necessary 38A, 38B must be at their peak (in the position furthest from the tilt balance plane Q) when the crankshaft is in the angular position a relative to the top dead center of the cylinder.

We can also provide a phase shift modulation device 52, illustrated in Figure 15, to vary the phase shift between the tilt balancing members 24A, 24B on the one hand, and the crankshaft 14 on the other go. For example, if the drive means 28 comprise, as described above, a first chain 34 for driving the first tilt balancing member 24A, a tensioner 54 may be provided, controlled for example by the control unit 46, so as to vary the trajectory of the first drive chain 34 and with it the phase shift between rotation of the crankshaft 14 and rotation of the two balancing members 24A, 24B. Naturally, other devices for modulating the phase shift may be envisaged. In practice, the phase shift range should be at least 10 °, preferably at least 20 °.

To generate the tilting torque compensation torque, it is not necessary for the tilting balancing plane Q, which by definition is the plane which contains the tilting balancing axes 102A, 102B, to be perpendicular to the thrust plane PP containing the axes 104 of the cylinders 18 in line with the engine. Figures 16 to 21 illustrate different possible variants of the positioning of the Q plane. Figure 16 shows the positioning of Figures 1 and 2, namely perpendicular to the thrust plane PP, away from the axis of revolution of the crankshaft 100. The figure 17 illustrates a positioning of the plane Q perpendicular to the thrust plane PP, and containing the axis of revolution. FIG. 18 illustrates a variant in which the plane Q contains the axis of revolution 100, but is oblique with respect to the thrust plane PP. FIG. 19 illustrates a particular variant in which the tilting balancing plane Q and the thrust plane PP are combined. FIG. 20 illustrates a tilting balancing plane Q parallel to the axis of revolution 100 but at a distance therefrom and oblique with respect to the thrust plane PP. Finally, FIG. 21 illustrates a tilting balancing plane Q parallel to the thrust plane PP, and distant from it. In practice, the embodiments of FIGS. 16, 20 and 21, in which the tilting balancing plane Q is distant from the axis of revolution of the crankshaft 100, prove to be easier to implement because they do not require no modification of the architecture of the engine block 12.

The two tilting balancing axes 102A, 102B are not necessarily parallel to each other in the tilting balancing plane Q which contains them. In FIG. 22 is illustrated an embodiment with an orientation of the tilting balancing members along two tilting balancing axes 102A, 102B coplanar but intersecting.

More generally, if we consider that the first tilt balancing member 24A has a first main axis of inertia which is not parallel to the first tilt balancing axis 102A and which rotates around the first tilt balancing axis 102A, and the second tilt balancing member 24B has a second main axis of inertia which is not parallel to the second tilt balancing axis 102B and which rotates around of the second tilt balancing axis 102B, compensation of the tilting torque may be obtained in particular if the first main axis of inertia and the second main axis of inertia are at all times symmetrical with each other with respect to to a tilting symmetry plane S perpendicular to the geometric reference axis.

In all of the previous figures, the engine 10 illustrated is an internal combustion engine with four cylinders in line. It is however advantageously possible to implement the invention for other types of engines generating a tilting torque, and in particular for a single-cylinder engine as illustrated in FIG. 23, a twin-cylinder engine as illustrated in FIG. 24 or a three-engine cylinders as illustrated in FIG. 25. It is analogous to generalize the use of the balancing system 22 of the tilting torque according to the invention to an engine with N cylinders in line, N being any integer greater than or equal to 1. It is also possible to generalize the use of the balancing system 22 of the tilting torque according to the invention to an engine with several V-shaped cylinders as illustrated in FIG. 26 or to a star engine as illustrated in FIG. 27 , and more generally to any internal combustion engine with regular ignition (that is to say at regular intervals of rotation of the crankshaft between two successive combustions).

The tilting torque is not the only source of parasitic forces transmitted by the engine block 12 to its fixing points. An important component on certain in-line engines is the so-called pestle force, which is the component parallel to the axis 104 of the cylinders 18 of the result of the forces exerted by the engine block 12 on its supports. The pestle force has a fundamental frequency which depends on the number of cylinders 18, the ignition sequence and the speed of the crankshaft 14. In the particular case of the four-stroke four-cylinder in-line engine, the fundamental frequency of the pestle force and that of the tilting torque coincide. It is in this context that various embodiments of the invention will now be described, ensuring both the compensation of the tilting torque and the compensation of the pestle force.

The embodiment of Figure 28 differs from that of Figure 1 in that the first tilt balancing member 24A comprises in addition to the elements of Figure 1 a first pestle balancing mass 56A of Pi value, having a center of gravity located in the thrust plane PP of the internal combustion engine, at a distance Fi from the first tilt balancing axis 102A. Similarly, the second tilt balancing member 24B comprises a second pestle balancing mass 56B of value P2, having a center of gravity located in the reference plane P of the internal combustion engine, at a distance Fz from second tilt balancing axis 102B.

The embodiment of FIG. 29 combines the embodiments of FIGS. 2 and 28. Thus, the first tilt balancing member 24A comprises a first tilt balancing mass 26A, a third balancing mass tilt 38A and a first pestle balancing mass 56A, while the second tilt balancing member 24B comprises a second tilt balancing mass 26B, a fourth tilt balancing mass 38B and a second mass d balancing drumstick 56B.

For the embodiments of FIGS. 28 and 29, which are specific to a four-stroke four-cylinder in-line engine, it will also be assumed that the values Pi and P2 of the first and second balancing weights of the drumsticks 56A and 56B are equal and the centers of gravity of the first and second balancing masses of the drumsticks 56A and 56B are in positions symmetrical to each other with respect to the plane of symmetry S. Thus, the distances Fi ctFz are equal . Finally, in these embodiments, the thrust plane PP containing the axes 104 of the cylinders 18 is imperatively parallel to the reference plane P. In practice, the thrust plane PP can be confused with the reference plane P, or offset by a few centimeters from the reference plane P.

Thus, the joint rotation of the first and second pestle balancing masses 56A, 56B about their respective balancing axis, in the opposite direction, generates a force having a result always located on the intersection tax between the plane of symmetry S and the thrust plane PP, and therefore parallel to the axes 104 of the cylinders 18. It is therefore possible to generate a force which, for a four-stroke four-cylinder engine, will have a periodic result, during the period of the pestle generated by the combustion cycles, and that we can stall so as to at least partially compensate for the pestle effort.

In FIGS. 28 and 29, the balancing mass of the pestle 56A, 56B is illustrated perpendicular to the tilting balancing masses 26A, 38A, 26B, 38B of the same balancing member 24A, 24B. This relative positioning generates a compensation effort for the pestle in quadrature with the tilt compensation torque. However, it can be seen that, as a general rule, the phase shift between the pestle force and the tilting torque for a four-stroke engine with four cylinders in lines is constant (at constant speed and torque) and between 65 ° and 115 ° . We will therefore choose for a given motor and for each balancing member 24A, 24B a relative angular positioning δ between the balancing mass of the pestle 56A, 56B and the tilting balancing mass or masses 26A, 38A, 26B, 38B which in an interval of 65 ° to 115 °, will correspond to the characteristic of the motor, as illustrated in figure 30.

An embodiment of a tilting balancing member 24A, and if necessary balancing the pestle force, has been illustrated in FIGS. 31 to 38. The tilting balancing member 24A comprises a body 25A having a domed outer casing 56A and a central plate 66, and equipped with a drive wheel 60 intended to mesh with the belt 34, and two end guide bearings 62, 64. The body 25A makes it possible to accommodate two sliding tilting balancing masses 26A, 38A and two return springs 68, 70. Between the tilting balancing mass 26A and the domed casing 56A is formed a chamber of variable volume 72. Similarly , a variable volume chamber 74 is formed between the tilting balancing mass 38A and the domed casing 56A. The two chambers 72, 74 are in hydraulic connection with each other and with an external pressure source, by means of feed channels opening out at the bearings 62, 64. The hydraulic circuit and the chambers 74 , 72 thus constitute a device 40 for controlling the position of the tilting balancing masses 26A, 38A, allowing the tilting balancing masses 26A, 38A to move from the neutral position illustrated in FIG. 34 to an extreme position illustrated in FIG. 38 against the return force of the springs 70, 68. In the neutral position of FIG. 34, the tilting balancing masses 26A, 38A are in the position shown diagrammatically in FIG. 6. The position of the Figure 38 corresponds to the position illustrated in Figure 4. Finally, it is possible, from the neutral position of Figure 34 and in the absence of pressure in the chambers 72, 74, to bring the mass balancing 26A to the left che at the end of travel stop and bring the balancing mass 38A to the right at the end of travel stop, under the effect of the return springs 70 68, which constitutes the state illustrated schematically in Figure 7 .

The center of gravity of the balancing mass 26A and the center of gravity of the balancing mass 38A are preferably located in the section plane BB located in FIG. 33, at a distance from both sides other of the section plane JJ located in FIG. 32. The center of gravity of the body 25A is located in the section plane AA located in FIG. 32, or in a reference plane C of the body 25A in the immediate vicinity of the section plane AA. In the position of FIG. 8, the center of gravity of the balancing mass 26A and the center of gravity of the balancing mass 38A are on either side of the section plane AA, and on both sides d 'another of a plane containing the center of gravity of the body and perpendicular to the tilting balancing axis 102A.

If the tilting balancing member 24A is not intended to balance a pestle force, the body 25A is formed so as to be dynamically balanced in its rotation about the tilting balancing axis 102A defined by bearings 62, 64.

If on the other hand provision is made for balancing a pestle force, the body 25 A will be shaped so as to have an unbalance. Where appropriate, the domed envelope of the body 25A has a bulged side which creates an unbalance constituting the pestle balancing mass 56A, as illustrated more particularly in FIG. 37. An angle δ of 90 has been illustrated in FIG. 37 ° approximately, but the angle can naturally vary depending on the distribution of the masses at the level of the bulge.

Naturally, many variations on the structure of the tilting balancing member 24A are possible. We can consider a control other than hydraulic, for example by worm or cable. If it is desired to control the balancing masses 26A, 38A jointly, a mechanical connection can be provided between the two, for example a pinion and rack or cable connection. Conversely, one may wish to make independent the control of the tilt balancing mass 26A and the control of the tilt balancing mass 38A, which would lead to providing, in the device of Figures 31 at 38, two independent hydraulic circuits. The return springs 68, 70 can be either compression springs or tension springs. More generally, a pneumatic or cam return system can be provided.

Illustrated in Figures 39 to 45 an embodiment of a balancing system 22 comprising a first tilt balancing member 24A and a second tilt balancing member 24B supported by a common frame 23 , intended to be secured to the engine block 12. The two tilting balancing members 24A, 24B are essentially identical, but have been illustrated in different states of counting to make the internal components visible.

The first tilt balancing member 24A comprises a hollow cylindrical body 25A guided in rotation relative to the engine block 12 around a first tilt balancing axis 102A by means of two end bearings 62A , 64A, and inside which are housed two tilting balancing masses 26A, 38A. The body 25A is equipped with a drive gear 60A comprising a frustoconical portion 60.1A intended to mesh with a drive gear (not shown), itself driven by the crankshaft of the engine, and a cylindrical portion 60.2 A intended to mesh with a drive wheel 60B equipping the body 25B of the second tilt balancing member 24B.

The body 25A internally forms two concave semi-cylindrical walls 125.IA, 125.2A, and two lateral slides 25.3A extending axially and separating the two semi-cylindrical walls 125.IA, 125.2A from one another.

Each tilting balancing mass 26A, 38A is associated with one of the semi-cylindrical walls 125.IA, 125.2A and consists of a set of rollers 26.1A, 38.1A housed in cells 26.21A, 38.21 A of a guide cage 26.2A, 38.2A, so as to roll on the associated semi-cylindrical wall 125.IA, 125.2A. The cage 26.2A, 38.2A further comprises two pairs of guide faces 26.22A, 38.22A which come to bear on the two lateral slides 25.3A, so as to be movable in translation parallel to the first tilting balancing axis 102A relative to the body 25A, but to be integral with the body 25A in rotation about the first tilt balancing axis 102A.

Each tilting balancing mass 26A, 38A, preferably has an even number of rollers 26.1A, 38.1A, preferably divided equally on either side of a plane II passing through the axis d '' tilt balancing 102A. The plane I-I (illustrated in FIG. 41) is preferably a plane of symmetry for the guide cages 26.2A, 38.2A and more generally for the tilting balancing masses 26A, 38A.

The cages 26.2A, 38.2A are shaped so as to be able to circulate axially, overlapping without collision. Remarkably, each cage 26.2A, 38.2A has a center of gravity which is centered on the tilting balancing axis 102A. To do this, the cage 26.2A, 38.2A, which is made of plastic material, is provided, in its part opposite to the cells 26.21A, 38.21A, with metallic inserts 26.23A, 38.23A, which balance the surplus plastic material. in the opposite part. Thus, the cage 26.2A, 38.2A is balanced in its rotation around the tilting balancing axis 102A. The tilting balancing mass 26A, 38A, constituted by the rollers 26.ΙΑ, 38.1A and the cage 26.2A, 38.2A, has its center of gravity located at a non-zero constant distance from the tilting balancing axis 102A, but only because of the rollers 26.1A, 38.1A, made of steel. The advantage of this arrangement is to reduce the centrifugal effects on the cage 26.2A, 38.2A, and in particular to prevent the cage 26.2A, 38.2A from applying to the rollers 26.IA, 38.IA an increasing load with the speed of revolution around the tilting balancing axis 102A. This generally controls the energy required for the translation of the tilting balancing masses 26A, 38A, which is relatively independent of the speed of revolution.

The cages 26.2A, 38.2A, of the tilting balancing masses 26A, 38A, are positioned axially by a control mechanism 42 which comprises a screw 42.IA which extends parallel to the balancing axis tilt 102A and has two threads, namely: a first thread 42.21A for driving a first nut 42.31A connected by a first rotary stop 42.4IA to the cage 26.2A of the tilt balancing member 26A, and a second 42.22A thread for driving a second nut 42.32A linked by a second rotary stop 42.42A to the second cage 38.2A. The rotary stops 42.41A, 42.42A allow decoupling between the movement of revolution of the tilting balancing member 24A around its tilting balancing axis 102 and the control mechanism 42. One or more rods 42.51A, 42.52A integral with the frame 25 extend parallel to the screw 42.IA, but laterally at a distance therefrom, and pass through bores formed in the nuts 42.3ΙΑ, 42.32A, to prevent the nuts 42.3IA, 42.32A to rotate around the axis of rotation of the screw 42.IA, with respect to the frame 25. The two threads 42.21A, 42.22A are in opposite directions and of identical pitch so that the rotation of the screw 42.IA causes simultaneously in translation, in opposite directions and with the same amplitude, the two cages 24.2A, 38.2A.

As indicated above, the tilt balancing member 24B is largely identical to the tilt balancing member 24A, with the exception of the toothed wheel 60B, which does not have a frustoconical portion. The constituent elements of the tilt balancing member 24B have therefore been identified in the figures by using the reference numbers used for the tilting balancing member 24A and by replacing the letter "A" with the letter "B "

A motor 44 drives, via a shaft 44.1, the two screws 42.IA and 42.IB in the same direction and at the same speed.

If the tilting balancing member 24A is not intended to balance a pestle force, the body 25A is formed so as to be dynamically balanced in its rotation about the tilting balancing axis 102A defined by the bearings 62, 64. The same applies to the tilting balancing member 24B. The operation of the balancing system according to this embodiment is essentially in accordance with the embodiment described in FIGS. 4 to 7.

If on the other hand provision is made for balancing a pestle force, the body of each of the balancing members 25A, 25B will be shaped so as to have an unbalance, similar to the unbalance provided in FIG. 37. Operation of the balancing system according to this embodiment will then essentially conform to the embodiment described in FIGS. 29 and 30.

According to a variant not illustrated, we can dispense with the rods 42.51A, 42.52A for sliding the nuts, and obtain an anti-rotation effect by providing that the axis of the screw 42.IA is parallel to the axis tilt balancing 102A, but at a distance therefrom.

In another variant, the two screws 42.IA and 42.IB have threads oriented in opposite directions and the motor 44 drives the two screws 42.IA and 42.IB in opposite directions and at the same speed .

In Figures 46 to 49 is illustrated an embodiment which differs from the previous one by the control mechanism 42 and the shape of the guide cages 26.2A, 38.2A, 26.2B, 38.2B.

The control device 40 in fact comprises an electric motor 44 whose output shaft passing through 44.0, parallel to the tilt balancing axes 102A, 102B, drives by means of two bevel gears, two shafts winding 44.1, 44.2. Each winding shaft 44.1, 44.2 extends perpendicular to the tilting balancing axes 102A, 102B in the plane Q comprising the tilting balancing axes 102A, 102B. Each winding shaft 44.1, 44.2 carries two winding drums 44.21A, 44.21B, 44.22A, 44.22B, each associated with one of the two tilt balancing members 24A, 24B and positioned in the axial extension d one of the ends of the associated tilting balancing member 24A, 24B.

The control mechanism 42 comprises, for each tilting balancing member 24A, 24B, a control cable 42.6A, 42.6B, extending parallel to the tilting balancing axis 102A, 102B and of which the ends are fixed to the winding drums 44.21A, 44.21B, 44.22A, 44.22B associated. In the middle of each control cable 42.6A, 42.6B is fixed the inner ring of a rotary stop 42.3A, 42.3B whose outer ring is fixed to the guide cage 26.2A, 26.2B. This guide cage 26.2A, 26.2B is in kinematic connection with the other cage 38.2A, 38.2B of the tilting balancing member 24A, 24B associated, by means of a transmission mechanism 42.7A, 42.7B with pinions and racks. The racks are formed on the guide cages 26.2A, 26.2B, 38.2A, 38.2B, the pinions being mounted on the body 25A, 25B and meshing with the racks of the two associated guide cages, so that a movement of translation of one of the guide cages 26.2A, 26.2B in a direction parallel to the tilt balancing axis 102A, 102B forces the other guide cage

38.2A, 38.2B a translational movement of the same amplitude in the opposite direction parallel to the tilt balancing axis 102A, 102B.

The engine output shaft drives the two winding shafts 44.1,

44.2 in the same direction and at the same speed, and the winding drums 44.21A, 44.21B, 44.22A, 44.22B are all identical, so that for each tilt balancing member 24A, 24B, the winding of one end of the cable 42.6A, 42.6B is accompanied by an unwinding of the same amplitude from the opposite end, and a translational movement of the same amplitude of the rotary stop 42.3A, 42.3B and the cage guide 26.2A, 26.2B associated inside the body 25A, 25B, parallel to the tilting balancing axis 102A, 102B, which causes, via the transmission 42.7A, 42.7B by pinions and racks, a translation in the opposite direction of the other guide cage 38.2A, 38.2B. Each of the guide cages 26.2A, 38.2A, 26.2B, 38.2B drives the rollers 26.1A, 38.1A, 26.1B, 38.1B which are associated therewith, which roll on the semi-cylindrical wall 125.1A, 125.2A, 125.IB , 125.2B associated.

As in the previous embodiment, the rotary stops 42.3A, 42.3B allow decoupling between the movement of revolution of the tilt balancing members 24A, 24B around their respective tilt balancing axes 102A, 102B and the balancing mechanism 22.

If the tilting balancing member 24A is not intended to balance a pestle force, the body 25A is formed so as to be dynamically balanced in its rotation about the tilting balancing axis 102A defined by the bearings 62, 64. The same applies to the tilting balancing member 24B. The operation of the balancing system according to this embodiment is essentially in accordance with the embodiment described in FIGS. 4 to 7.

If on the other hand provision is made for balancing a pestle force, the body of each of the balancing members 25A, 25B will be shaped so as to have an unbalance, similar to the unbalance provided in FIG. 37. Operation of the balancing system according to this embodiment will then essentially conform to the embodiment described in FIGS. 29 and 30.

The embodiment illustrated in Figures 50 to 52 differs from the previous essentially by the control mechanism 42, which here is hydraulic and comprises, in addition to a hydraulic supply circuit 142.1, a distribution solenoid valve 142.2 to which are connected two bidirectional cylinders 142.3A, 142.3B, each associated with one of the tilt balancing members 24A, 24B. Each cylinder 142.3A, 142.3B can be positioned outside the body 25A, 25B of the tilting balancing member 24A, 24B associated, or preferably and as illustrated, inside, for greater compactness . Each cylinder 142.3A, 142.3B comprises a cylinder 142.4A, 142.4B fixed to the frame, and a piston 142.5A, 142.5B whose rod is fixed, at its free end, to a rotary stop 42.31A, 42.31B, itself fixed , as in the previous embodiment, to one of the cages 26.2A, 26.2B. As in the previous embodiment, each cage 26.2A, 26.2B is kinematically linked to the other cage 38.2A, 38.2B of the associated balancing member 24A, 24B by a transmission with pinions and racks 42.7A, 42.7 B.

According to a variant not illustrated, the cylinder 142.3A, 142.3B can be electromechanical.

According to another embodiment, illustrated in FIG. 53, the control mechanism 42 comprises, identically for each tilting balancing member 24A, 24B (only the balancing member 24A has been illustrated) , two reversible endless screws 242.11, 242.12, here constituted by hollow screws mounted for free rotation on a guide rod 242.2. The threads of the screws 242.11, 242.12 are in the same direction. Each of the screws 242.11, 242.12 cooperates reversibly with a nut 242.31, 242.32 secured to one of the guide cages 26.2, 38.2 for the rolling bodies 26.1, 38.1, the two guide cages 26.2, 38.2 being linked one to the other by a rack and pinion link 42.7. The screws 242.11, 242.12 are each secured to a brake plate 242.41, 246.42. The control mechanism further comprises, associated with each braking plate 242.41, 242.42 and with each screw 242.11, 242.12, a braking member 242.51, 242.52, capable of applying a braking torque to the braking plate 242.41, 242.42.

Insofar as the nut 242.31, 242.32 is integral with the cage 26.2, 38.2 associated not only in translation but also in rotation, the nuts 242.31,

242.32 of the same tilt balancing member 24A, 24B follow the rotation of the body 25 of the tilt balancing member 24A, 24B. The screws 242.11, 242.12 rotate with the nuts 242.31, 242.32, as long as they remain free to rotate. When the screw 242.11 is braked in its rotation by the associated braking member 242.51, the nut 242.31, and the associated cage 26.2 begin to move in translation relative to the body 25. The transmission 42.7 between the guide cages

26.2 and 38.2 impose on the cage 38.2 a translational movement in the direction opposite to the guide cage 26.2, which is made possible by the reversibility of the connection between the nut 242.32 and the associated screw 242.12. The behavior described in FIGS. 4 to 7 is therefore obtained. This embodiment is particular in that the energy required to move the tilting balancing masses 26, 38 is taken directly from the drive of the balancing members. tilt 24A, 24B, without the need to provide specific drive members.

Naturally, other variations of the tilting balancing members are possible, in particular by combining the various embodiments illustrated. One can in particular provide, as a variant of the embodiments of FIGS. 46 to 52, a unidirectional actuator ensuring the displacement of the tilting balancing masses in one direction, preferably towards the most distant position, and elastic return means ensuring the displacement of the tilting balancing masses in the opposite direction, by analogy with the spring 68 of FIG. 34.

The structures illustrated in Figures 31 to 53 are particularly suitable for operation in accordance with the operation described in Figures 4 to 7 in the absence of a pestle balancing mass, or to the operation described in Figures 29 and 30 in the presence of a pestle balancing mass. They are however easily transposable to devices whose third and fourth tilting balancing masses 38A, 38B are fixed (Figures 11 to 13) or absent (Figure 1) · [0178] In the embodiments of Figures 39 to 53 , the rolling bodies 26.1, 26.1A, 26.1B, 38.1, 38.1A, 38.1B are generator rollers in an arc, rolling on concave semi-cylindrical walls 125.1A, 125.2A. It is however perfectly conceivable to provide a cylindrical body 25A, 25B with a polygonal base, defining for the rolling bodies flat raceways, allowing the use of cylindrical rollers. It is also possible to adapt the body structure 25A, 25B and the guide cages to other forms of rolling bodies, in particular balls. The number of rolling bodies can be any, and if necessary limited to one for each tilting balancing mass 26A, 26B, 38A, 38B.

Claims (21)

1. Balancing member (24A) for balancing an internal combustion engine (10), the balancing member (24A) comprising a body (25A) provided with one or more bearings (62, 64) for guiding rotation of the body (25A) about a balancing axis (102A) of the balancing member (24A), the body (25A) having a center of gravity located in a reference plane (C) of the body ( 25A) perpendicular to the balancing axis (102A), characterized in that the balancing member (24A) further comprises at least a first balancing mass (26A) having a value Mi and a center of gravity located at a non-zero distance E \ from the balancing axis (102A), fixed in rotation relative to the body (25A) around the balancing axis (102A) and guided relative to the body (25A) of so that the distance £> i from the center of gravity of the first balancing mass (26A) to the reference plane (C) of the body (25A) is variable. 2. Balancing member according to claim 1, characterized in that the first balancing mass (26A) is guided relative to the body (25A) so that the non-zero distance Ei remains constant. 3. Balancing member according to any one of claims 1 to 2, characterized in that the balancing member (24A) further comprises a control device for varying the distance £> i within a range of values between two end positions. 4. Balancing member according to claim 3, characterized in that the control device comprises a variable volume hydraulic chamber (72), one wall of which is formed by the first balancing mass (26A). 5. Balancing member according to claim 3, characterized in that the control device comprises an actuator, the actuator having a movable output member, preferably movable in translation parallel to the balancing axis (102A), the control device preferably comprising a rotary connecting stop between the outlet member and the first balancing mass (26A). 6. Balancing member according to claim 5, characterized in that the actuator is a unidirectional actuator and the control device further comprises a return system (68) of the first balancing mass (26A) towards the one of the two end positions. 7. Balancing member according to claim 5, characterized in that the actuator is a bidirectional actuator. 8. Balancing member according to any one of the preceding claims, characterized in that the body (25A) is equipped with a member (60) for driving in rotation about the balancing axis (102A), preferably an annular toothing. 9. Balancing member according to any one of claims 1 to 8, characterized in that the center of gravity of the body (25A) is located on the balancing axis (102A). 10. Balancing member according to any one of claims 1 to 8, characterized in that the center of gravity of the body (25A) is located at a distance from the balancing axis (102A), the center of gravity of the body (25A) preferably being located in a plane containing the balancing axis (102A) making an angle δ relative to a plane containing the balancing axis (102A) and the center of gravity of the first mass balancing (26A), the angle δ preferably being between 50 ° and 130 °. 11. Balancing member according to any one of the preceding claims, characterized in that the first balancing mass (26A) comprises one or more rolling bodies (26.IA), the body (25A) of the member d balancing being provided with one or more raceways (125.IA) oriented and turned towards the balancing axis (102A), the rolling body or bodies being able to roll on the raceway (s) (125. IA) parallel to the balancing axis (102A) when the distance D \ from the center of gravity of the first balancing mass (26A) to the reference plane (C) varies. 12. Balancing member according to claim 11, characterized in that the balancing member (24A) comprises a guide cage (26.2, 26.2A) of the rolling body or bodies (26.1, 26.IA), movable in translation parallel to the balancing axis (102A) when the rolling body or bodies (26.1, 26.IA) roll on the raceway (s) (125.1A). 13. Balancing member according to claim 12, characterized in that the guide cage (26.2, 26.2A) has a center of gravity located on the balancing axis (102A). 14. Balancing member according to any one of claims 11 to 13, characterized in that the raceway (s) (125.IA) are cylindrical and concave, and preferably with center of curvature positioned on the axis d 'balancing (102A). 15. Balancing member according to any one of claims 11 to 14, characterized in that the rolling body or bodies are rollers, preferably generator rollers in an arc. 16. Balancing member according to any one of the preceding claims, characterized in that it further comprises at least one additional balancing mass (38A) having a value M3 and a center of gravity located at a non-zero distance E3 of the first balancing axis (102A), fixed in rotation relative to the body (2 5 A) around the balancing axis (102 A) and guided relative to the body (25A) so that the distance D3 from the center of gravity of the additional balancing mass (38A) to the reference plane (C) of the body (25A) is variable. 17. Balancing member according to claim 16 in combination with claim 3, characterized in that the control device is capable of varying the distance £> 3 in a range of values between two additional end positions. 18. Balancing member according to claim 17, characterized in that the control device comprises a mechanical connection (42.7A) between the first balancing mass and the additional balancing mass, so that a displacement of the first balancing mass in a first direction parallel to the balancing axis (102A) drives the additional balancing mass in a second direction parallel to the balancing axis (102A) and opposite to the first direction. 19. Balancing member according to any one of claims 16 to 18, characterized in that the center of gravity of the additional balancing mass (38A) and the center of gravity of the first balancing mass of (26A ) are in a balancing plane (BB) containing the balancing axis (102A), and on either side of a plane (JJ) perpendicular to the balancing plane (BB) and containing the balancing pin (102A). 20. tilting balancing system comprising a first tilting balancing member constituted by a balancing member (24A) according to any one of the preceding claims guided in rotation around a first tilting balancing axis ( 102A) and a second tilting balancing member constituted by a balancing member (24B) according to any one of the preceding claims guided in rotation around a second tilting balancing axis (102B) coplanar and preferably parallel to the first tilt balancing axis (102A). 21. tilting balancing system according to claim 20, characterized in that it further comprises a kinematic connection (36) between the body (25A) of the first tilting balancing member (24A) and the body ( 25B) of the second tilt balancing member [24B], such as a rotation of the first tilt balancing member (24A) about the first tilt balancing axis (102A) in a given direction at a given speed causes the second tilt balancing member (24B) to rotate about the second tilt balancing axis (102B) in the opposite direction and at the same speed. 1/21