FR3096425A1 - Balancing method of an internal combustion engine - Google Patents

Abstract

For balancing an internal combustion engine (10) with one or more reciprocating pistons (16), comprising an engine block (12), a crankshaft (14) rotating relative to the engine block (12) of the internal combustion engine (10) around a geometrical reference axis (100), it is driven in rotation relative to the motor unit (12) around a first tilting balancing axis (102A) not parallel to the reference axis (100), at a speed of revolution having a predetermined ratio with the speed of revolution of the crankshaft, a first tilting balancer (24A) having a first dynamic unbalance, and is rotated relative to the engine block (12) around a second tilting balancing axis (102B) not parallel to the reference axis (100), at a speed of revolution of equal amplitude and in the opposite direction to the speed of revolution of the first balancing member of tilting (24A), a second tilting balancer (24B) having a second th dynamic unbalance, by synchronizing the rotation of the first tilt balancing member (24A) and the second tilt balancing member (24B) so that the first tilt balancing member (24A) and the second tilt balancer member (24A) tilting balancing device (24B) together generate on the engine block (12) at least one tilting compensation torque about the geometric reference axis (100).

Description

Description Title of the invention: Balancing method of 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.

Prior technique

[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 the internal volume of the cylinder, the latter being minimum when the piston is close to the cylinder head and maximum when the piston is farthest from the cylinder head.

[0003] The piston makes one 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 a stroke of the motor.

In a four-stroke engine, the combustion air is drawn into the volume during the first stroke.

Torque is applied from the crankshaft to compress the combustion air during the second stroke and form a mixture between the oxidizer 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 torque on the crankshaft during the third stroke.

The magnitude 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 evacuated during the fourth stroke.

[0004] The four strokes require two complete rotations of the crankshaft.

The torque on the crankshaft generated by a cylinder of a four-stroke engine is almost zero during the fourth and first strokes, slightly resistant during the second stroke, and strongly driving during the third stroke.

This torque pulse generates a crankshaft output torque which oscillates around its average, these variations having to be managed by the transmission kinematic chain downstream of the engine.

[0005] 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 2 on the sleeve generate forces having a resultant substantially radial with respect to the axis of translation of the piston, and oriented, with respect to this axis, opposite the connecting rod .

Efforts substantially opposite to the previous ones are applied by the crankshaft to the guide bearings formed in the engine block.

All of the forces applied to the engine block have a non-zero resultant and generate torques, this resultant and these torques being taken up by the supports of the engine block.

In what follows, we will mainly focus on two components transmitted to the engine block supports: on the one hand the component of the resultant of the forces which is perpendicular to the axis of revolution and parallel to the axis of translation of the pistons, called "drumstick", and on the other hand the torque exerted around the axis of revolution of the crankshaft, called "tilting torque".

[0006] 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 crankshaft revolution is equal to half the number of cylinders.

Thus, a four-cylinder four-stroke engine has a main harmonic of order two; a six-cylinder four-stroke engine has a third-order main harmonic.

As the number of cylinders increases, the frequency of the main harmonic of output torque and motor oscillating motion increases, while the peak-to-peak amplitudes of rocking torque and motor torque decrease, which decreases noise and transmitted vibrations For a two-stroke engine, the number of combustion events per crankshaft is equal to the number of cylinders.

[0007] 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 stroke, whereas the maximum of the tilting torque is observed in the middle of the third beat.

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 result in better thermal efficiency and lower friction, both of which combine to decrease fuel consumption.

Thus, there is a desire among engine manufacturers 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.

[0009] Technologies exist to attenuate the effects of displacement on the motor supports, but these technologies generally have a low frequency range of attenuation 3, which does not make it possible to manage the entire speed range of the motor.

Said technologies are usually applied at the lowest resonant frequency of the system.

The remaining frequencies generated by the motor's operating range remain problematic.

[0010] In order to attenuate the vibrations, the variations in engine torque and the noise over the entire operating range of the engine, it has been proposed in the document US2014230771 to equip the engine with a balancing device called "nutation", which comprises a coupler coupled to the crankshaft of the engine so as to be driven in rotation by the crankshaft around an axis perpendicular to the axis of revolution of the crankshaft, and a rotating mass coupled to the coupler, cc last 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 to bring it closer to or further away from the axis of revolution of the coupler.

But the mechanism is particularly complex and its cost very high.

Disclosure of 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 reference geometric axis, method according to which: rotation is driven relative to the engine block around a first tilting balancing axis not parallel to the reference axis, at a speed of revolution having a predetermined ratio with the revolution speed of the crankshaft, a first tilting balancing member having a first dynamic unbalance, and an engine block is driven in rotation about a second tilting balancing axis not parallel to the reference axis, at a speed of revolution of equal amplitude and opposite direction to the speed of revolution of the first tilting balancing member, a second tilting balancing member having a second dynamic unbalance, by synchronizing the rotation of the first tilt balance member and the second tilt balance member such that the first tilt balance member and the second tilt balance member together generate on the engine block at least a tilt compensation torque around the reference geometric axis, the first tilt balancing axis and the second tilt balancing axis being located in a geometric tilt balancing plane, the first balancing member tipping mechanism comprising a first tipping balance mass having a first value M, and a first center of gravity located on a first side of a reference plane containing the reference axis and perpendicular to the tipping balance plane , at a distance DI from the reference plane and at a distance El from the first tilting balancing axis, the second tilting balancing member comprising a second tilting balancing mass having a second value Mo and a second center of gravity located on the first side of the reference plane at a distance D from the reference plane and at a distance E from the second tilting balancing axis, at least one of the distances Di is varied.

E1, and at least one of the distances D2, E2, as a function of one or more control parameters from among engine operating parameters.

[0013] By dynamic unbalance, we mean here a distribution of the masses such that the main axis of inertia of the body considered is neither coincident with the axis of rotation (total absence of unbalance) nor parallel to the axis of rotation (static unbalance). 11 may be a torque imbalance, that is to say a distribution of masses which results in a main axis of inertia secant with the axis of rotation, or a dynamic imbalance not particular, namely an imbalance superimposing a torque imbalance and a static imbalance, characterized by a main axis of inertia not parallel to the axis of rotation and not intersecting with the axis of rotation.

[0014] The rotational movement of the first tilting balancing member around the first tilting balancing axis and the rotational movement of the second tilting balancing member around the second tilting balancing axis being guided relative in 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 the tilting torque.

[0015] In practice, the speed of rotation of the first tilt balancing axis will be equal to the speed of revolution of the crankshaft or to a multiple of this speed.

The predetermined ratio between the revolution speed of the crankshaft and the revolution speed 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.

It is thus ensured that the compensation torque generated by the rotational tilting balancing members has the same frequency as the tilting torque.

[0016] In practice, the first tilting balancing axis and the second tilting balancing axis are symmetrical to each other with respect to a tilting symmetry plane, the tilting symmetry plane being preferably perpendicular to the reference geometric axis.

This avoids generating torques or resulting forces on the engine block that are not useful for compensating the tilting torque, and in particular a force in the direction of the reference axis.

[0017] Preferably, the first tilting balancing axis and the second tilting balancing axis are parallel.

This avoids generating a resultant force in the direction of the bisector of the two tilting balancing axes.

[0018] 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 link may include a bevel gear.

[0019] The alternative of providing an independent auxiliary drive, speed-controlled by measuring the speed of revolution of the crankshaft, for example by electric motor, remains possible.

[0020] According to one embodiment, the rotation of the first tilting balancing member and of the second tilting balancing member is synchronized such that the tilting compensation torque generated is maximum, in absolute value, on a revolution 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 00 and 50° in the direction of revolution of the crankshaft.

The angle a is located around the maximum force of the piston on the liner due to combustion and is generally between 20° and 40°, typically around 30°.

This angle is dependent on the combustion parameters and the kinematics of the mobile coupling.

We can choose a constant value of the angle a, 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 motor operating parameters.

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 an engine cylinder, or the maximum value of the pressure reached in an engine cylinder.

The cylinder pressure information can be estimated or measured.

[0021] Preferably, the geometric tilting balancing plane is parallel to the reference geometric axis.

Preferably, and to take account of the size of the other elements of the kinematic propulsion chain, the geometric tilting balancing plane is distant from the geometric axis of revolution of the crankshaft.

Preferably, D′ E1, D2 is varied. and/or E,, such that: Mi.D 1.E1= M2.1)2.E2

In practice, and according to a particularly advantageous embodiment, the value MI of the first tilting balancing mass and the value M2 of the second tilting balancing mass are equal, and the first balancing mass tipping mass and the second tipping balancing mass are symmetrical to each other with respect to the tipping symmetry plane.

A 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.

According to different embodiments, the control parameter(s) include one or more of the following parameters: the speed of revolution of the crankshaft, the engine torque, the engine power, the angle of the crankshaft corresponding to a maximum pressure in a cylinder of the engine, the maximum value of the pressure reached in a cylinder of the engine, and preferably at least the speed of revolution of the crankshaft in combination with at least a second parameter among the following parameters: the engine torque, the engine power, the crankshaft angle 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.

[0025] According to one embodiment, the distance between the first tilting balancing mass and the first tilting balancing axis is varied according to the operating parameter(s) of the engine.

But this solution requires a significant energy input to overcome the centrifugal force generated by the rotation of the tilt balancing member.

According to a preferred alternative embodiment, the distance between the first tilting balancing mass and the first tilting balancing axis is constant, and the distance between the first tilting balancing mass and the reference plane according to 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 movement of the first tilt balancing mass is relatively simple, since it only requires one degree of translational freedom.

Naturally, provision is made to vary the position of the second tilt balancing mass analogously to the first tilt balancing mass.

[0026] According to a particularly advantageous embodiment, the distance between the first tilting balancing mass and the reference plane is varied in such a way that, for each speed of revolution of the crankshaft in a given range of revolution speeds , the distance between the first tilt balancing mass and the reference plane is a monotonic function of the motor torque.

This embodiment results in a tilting torque which is itself a monotonic function, in this case an increasing function, of the engine torque, at constant crankshaft revolution speed.

[0027] According to a particularly advantageous embodiment, the distance between the first tilting balancing mass and the reference plane is varied in such a way that, for each motor torque in a given range of motor torques, the distance between the first tilt balance mass and the reference plane is a monotonic function of the crankshaft revolution speed, at least when the crankshaft revolution speed is above 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 from the reference plane, at the speed of revolution of the given first 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 crankshaft revolution speed and as a function of the crankshaft revolution speed at constant torque are used jointly, by defining a two-variable law.

[0029] For an N-cylinder in-line four-stroke internal combustion engine, N being an integer greater than or equal to 1, the constant ratio between the speed of revolution of the first tilting balance shaft and the speed of revolution of the crankshaft is preferably equal to N/2. 8

[0030] 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 tilting balance shaft and the speed of revolution of the crankshaft is preferably equal to N.

To synthesize the balancing system and prevent it from generating a resultant of parasitic forces adding, where appropriate, to the engine rammer, it is advantageous to provide that the first tilting balancing member comprises a third mass of tilting balancing of value M3 having a center of gravity located on a second side of the reference plane opposite the first side, at a distance D3 from the geometric reference plane and at a distance E3 from the first tilting balancing axis, and that the second tilting balancing member comprises a fourth tilting balancing mass of value M4 having a center of gravity located on the second side of the reference plane at a distance D4 from the geometric reference plane and at a distance E4 from the second tilt balance axis.

In this hypothesis, at least one of the distances D3, E:3, and at least one of the distances D4 are varied.

E4, as a function of one or more control parameters from among engine operating parameters, preferably so that: MD 3.E3 = d):14.D 4.E4

[0032] When the crankshaft is in the predetermined angular position previously defined, the third tilting balancing mass and the fourth tilting balancing mass are preferably located on one side of the tilting balancing plane. opposed to the first tilt balance mass and the second tilt balance mass.

[0033] According to a particularly advantageous embodiment, 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 is varied. tipping balancing as a function of the motor 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 is varied and/or the distance between the third tilting balancing mass and the first tilt balance axis so that the distance between the third tilt balance mass and the reference plane is equal to the distance between the first tilt balance mass and the reference plane and the distance between the third tilt balance mass and the first tilt balance axis is equal to the distance between the first tilt balance mass and the first tilt balance axis.

[0036] 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 balance axis on the dc reference plane.

[0037] More generally, the values MI and M3 and the distances Di, El, and /33.

E3, are preferably such that the set of two masses M, and M3 has a main axis of inertia secant with the first tilting balancing axis.

Dc even, the values M, and M4 and the distances D, E2, and D4, ET, are preferably such that the set of the two masses M2 and M4 has a main axis of inertia secant with the first balancing axis tipping.

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 dc plane.

[0039] For internal combustion engines having a pestle force justifying balancing, and in particular for four-cylinder four-stroke in-line engines, provision can be made for: rotation is driven around a first axis of drumstick balancing perpendicular to a drumstick reference plane containing the axis of revolution of the crankshaft, at a speed of revolution having a constant ratio with the speed of revolution of the crankshaft, a first drumstick balance mass of value PI and having a center of gravity located in the pestle reference plane, at a distance FI from the first pestle balancing axis, it is rotated about a second pestle balancing axis parallel to the first axis of drumstick balancing and located at a distance from the first drumstick balancing axis, at a speed of revolution of equal amplitude and opposite direction to the speed of revolution of the first drumstick balance weight, a second balance weight of pestle of value P. and having a center of gravity located in the pestle reference plane at a distance F2 from the second pestle balancing axis, so that: Pi.F1 -

The two pestle balancing masses, which rotate in opposite directions in the reference plane around the pestle balancing axes, together generate forces whose resultant 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.

[0041] According to a preferred embodiment, the first drumstick balancing weight and the second drumstick balancing weight are symmetrical to one another with respect to a drumstick symmetry plane perpendicular to the axis of crankshaft revolution.

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.

[0042] The pestle balancing axes can be spaced apart from the tilting balancing axes without harming the quality of the balancing achieved.

However, and according to a preferred embodiment, provision can advantageously be made for the first pestle balancing axis to coincide with the first tilting balancing axis and for the second pestle balancing axis to coincide with the second axis. tipping balancing mechanism, which ensures greater compactness of the balancing mechanism.

[0043] For certain engines, and in particular four-cylinder four-stroke 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 driven in rotation at a speed of revolution equal to the speed of revolution of the first tilting balancing member, and that the second rammer balance mass at a speed of revolution equal to the speed of revolution of the second tilting balancer.

[0044] TI preferably exists 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 tilting balancing mass around the first axis d tilting balancing, this phase shift preferably being between 50° and 1300°.

According to a particularly advantageous embodiment, the first pestle balancing mass is an integral part of the first tilting balancing member and the second pestle balancing mass is an integral part of the second tilting balancing member .

Thus, the first pestle balancing mass is secured to the first tilting balancing mass in rotation around the first tilting balancing axis coinciding with the first pestle balancing axis.

In practice, the first pestle balancing mass and the first tilting balancing mass are supported by a common rotating 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 tilting balancing axis coinciding with the second pestle balancing axis.

In practice, the second ram balancing weight and the second tilting balancing weight are supported by a common rotating body or shaft embodying the second balancing axis.

A device for balancing the tilting torque and the particularly compact pestle is thus produced.

[0046] According to one embodiment which is particularly important in practice, the internal combustion engine is an in-line engine, preferably a four-stroke engine with four cylinders in-line, the pistons moving parallel to the reference plane of the engine at internal combustion.

In this hypothesis, it can advantageously be provided that: Ion drives in rotation around the first tilting 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 tilting balance axis, Ion rotates about the second tilting balance 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 P2 and having a center of gravity situated in the reference plane of the engine at internal combustion at a distance Ei from the second tilting balancing axis, so that: Pi-Fi - P2--F2

[0047] In practice, the first pestle balancing mass and the second pestle balancing mass are preferably symmetrical to one another with respect to a pestle symmetry plane 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 tilting balancing mass around the first tilting balancing axis, this phase shift being preferably between 50° and 130°.

According to another aspect of the invention, the latter relates to a method of balancing an internal combustion engine with one or more reciprocating pistons, comprising a crankshaft rotating around a reference geometric axis, method according to which: one drives in rotation around a first tilting balancing axis not parallel to the reference axis, at a speed of revolution having a predetermined ratio with the speed of revolution of the crankshaft, a first member of tilt balance having a first main axis of inertia not parallel to the first tilt balance axis, and is rotated about a second tilt balance axis not parallel to the reference axis, a speed of revolution of equal amplitude and opposite direction to the speed of revolution of the first tilting balancing member, a second tilting balancing member having a second main axis of inertia not parallel to the second balancing axis tilting, by synchronizing the rotation of the first tilting balancing member and of the second tilting balancing member such that the first main axis of inertia and the second main axis of inertia are symmetrical with one another. the other relative to a tilting symmetry plane perpendicular to the reference geometric axis.

[0050] Preferably, the first tilting balancing axis and the second tilting balancing axis are symmetrical to each other with respect to the tilting symmetry plane.

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

Preferably, and to take account of the size of the other elements of the kinematic propulsion chain, the geometric tilting balancing plane is distant from the geometric axis of revolution of the crankshaft.

According to another aspect of the invention, the latter 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 rotating tilting balancing member capable of being guided in rotation relative to the engine block around a first tilting 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 with respect to an engine block around a second tilting balancing axis not parallel to the reference axis, and having a second dynamic unbalance relative to the second tilt balance axis, a synchronized drive mechanism for driving the first tilt balance member in rotation about the first tilt axis tilt balancing at a revolution speed having a constant predetermined ratio with the revolution speed of the crankshaft, and for rotating the second tilt balancing member about the second tilt balancing axis at a revolution speed of equal amplitude and in opposite direction to the speed of revolution of the first tilting balancing member, such that the first tilting balancing member and the second tilting balancing member together generate on the engine block a torque of compensation for tilting around the reference geometric axis.

[0052] For a four-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the constant predetermined ratio between the speed of revolution of the first tilting balance shaft and the speed of crankshaft revolution is preferably equal to N/2.

[0053] For a two-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, the constant predetermined ratio between the speed of revolution of the first tilting balance shaft and the speed of crankshaft revolution is preferably equal to N.

[0054] In practice, the balancing system comprises one or more first bearings to guide the first rotating tilting balancing member in rotation around the first tilting balancing axis with respect to the engine block, and one or more second bearings for guiding the second rotating 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 tilting balancing member and a kinematic connection between the crankshaft and the second balancing member. tilt balancing or between the first tilt balancing member and the second tilt balancing member.

The transmission mechanism can be of any suitable type, comprising for example belts, chains, sprockets and/or countershafts.

If a kinematic connection between the first balancing member and the second balancing member is provided, this must reverse the direction of rotation between the two balancing members.

This solution with a single kinematic connection between the balancing system and the crankshaft makes it possible to integrate the kinematic connection between the first tilting balancing member and the second tilting balancing member into a common frame of the balancing system .

It is then possible to propose a compact and unitary balancing system, the assembly of which on the engine is easy.

[0056]According to one embodiment, the balancing system comprises modulating means capable of modulating a phase difference between the rotation of the first tilting balancing member around the first tilting balancing axis and the rotation of the crankshaft around the axis of reference geometry, preferably as a function of one or more phasing parameters from among the motor 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 set 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, the outside temperature, the temperature of the coolant, 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, possibly cumulative: the control means include at least one crankshaft revolution speed sensor to determine at least one of the control parameters; the control means comprise one or more torque or torque demand sensors or estimators to determine at least one of the control parameters; the control means comprise one or more sensors or engine cylinder pressure 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 suitable for varying the position of the first tilting balancing mass parallel to the first axis tilt balancing mass and the position of the second tilt balancing mass parallel to the second tilt balancing axis as a function of the control parameter(s).

According to one embodiment, the first tilting balancing member has a center of gravity located on the first tilting balancing axis, and the second tilting balancing member has a center of gravity located on the second tilt balance axis.

In other words, the two tilt balancing members have no static imbalance with respect to their respective axes of rotation.

This arrangement will be preferred for an engine generating no pestle force, or whose pestle force has a fundamental frequency which is not equal to that of the tilting torque.

[0060] Alternatively, provision could 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 away from the second tilt balancing axis.

In other words, the two tilting balancing members present a static imbalance with respect to their respective axis of rotation. This arrangement will be particularly advantageous for a motor generating a pestle force with a fundamental frequency which is equal to that of the torque. tilting, and in particular for a four-stroke engine with four cylinders in line.

According to another aspect of the invention, the latter 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 reference geometric axis of the engine block contained in a reference plane of the engine block, and a balancing system as described above.

[0062] Preferably, the first tilting balancing axis and the second tilting balancing axis are symmetrical to each other with respect to a tilting symmetry plane, preferably perpendicular to the geometric axis. reference.

According to a particularly advantageous embodiment in practice, the first tilting balancing axis and the second tilting balancing axis are parallel, and preferably perpendicular to the reference plane.

This minimizes the forces and parasitic torques generated by the rotation of the tilt balancing members.

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

[0066] In particular, the invention is applicable to a four-stroke engine with four cylinders in line, the balancing system being such that the first tilting balancing member has a center of gravity situated at a distance from the first axis tilt balance, and the second tilt balance member has a center of gravity located away from the second tilt balance axis.

In this hypothesis, a thrust plane containing the axes of the cylinders of the engine is defined, which is preferably parallel to or merged with the reference plane defined previously.

The tilt balancing geometric plane is preferably perpendicular to the thrust plane.

The center of gravity of the first tilting balancing member and the center of gravity of the second tilting balancing member are preferably located in the thrust plane.

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

[0068] 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 around a balancing axis of the balancing member, the body having a center of gravity located in a reference plane of the body perpendicular to the balancing axis, and at least a first balancing mass having a value M1 -1 and a center of gravity located at a non-zero distance E1 from the balancing axis, fixed in rotation with respect to the body around the balancing axis and guided with respect to the body so that the distance D, from the center of gravity of the first balancing mass to the reference plane of the body is variable.

By varying the distance DI, there is a way to vary the lever arm of the balancing member relative to an axis located in the reference plane.

[0070] Preferably, the first balancing weight is guided relative to the body so that the non-zero distance E i remains constant.

It is then possible to move the first tilting balancing mass with a low control energy, because one does not have to oppose centrifugal forces induced by the rotation of the balancing member around the axis d 'balancing.

[0071] Preferably, the balancing member further comprises a control device for varying the distance DI within 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 one embodiment, the balancing member further comprises a return system, preferably elastic return, of the first balancing mass towards one of the two end-of-travel positions.

The return system may in particular comprise one or more return springs, in particular traction or compression springs, or pneumatic springs.

Other reminder systems can also be considered.

According to a particularly advantageous embodiment, the control device comprises a hydraulic chamber with variable volume, one wall of which is formed by the first balancing weight.

In practice, the control device 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 weight.

[0073] According to one embodiment, the body is equipped with a drive member in rotation around the balancing axis, preferably an annular toothing, intended to mesh with a similar toothing of a second member of 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 generates no unbalance around the balance axis.

According to an alternative embodiment, the body's center of gravity is located at a distance from the balancing axis, the body's center of gravity preferably being located in a plane containing the balancing axis forming an angle h with respect to a plane containing the balancing axis and the center of gravity of the first balancing mass, the angle h preferably being between 50' and 130'.

Preferably, the balancing member further comprises at least one additional balancing weight having a value Met a center of gravity located at a non-zero distance E; of the first balancing axis, fixed in rotation with respect to the body around the balancing axis and guided with respect to the body so that the distance D; of the center of gravity of the additional balancing mass to the dc reference plane of the body is variable.

In practice, provision may advantageously be made for the first balancing weight and the additional balancing weight to have the same value.

Provision may also be made for one or more of the following arrangements: the additional tilting balancing mass is guided relative to the body so that the non-zero distance e′ remains constant; the balancing member further comprises a control device for varying the distance DI within a range of values between two end-of-travel positions and for varying the distance 03 within a range of values between two additional end-of-travel positions ; the balancing member further comprises a return spring for the additional tilting balancing mass towards one of its two additional end-of-travel positions; the control device comprises a variable-volume hydraulic chamber, one wall of which is constituted by the additional tilting 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 second balancing mass, for example to ensure equality between DI and D3.

According to one embodiment, the control device is such that, for at least certain values of DI and /33, the center of gravity of the additional balancing mass is on the side of the reference plane of the opposite body. at the center of gravity of the first balancing mass.

[0078] Preferably, the center of gravity of the additional balancing weight and the center of gravity of the first balancing weight are located in a balancing plane containing the balancing axis, and on both sides other from a plane perpendicular to the balancing plane and containing the balancing axis.

According to another aspect of the invention, the latter relates to a tilting balancing system comprising a first tilting balancing member consisting of a balancing member as described above, guided in rotation around a first tilting balancing axis and a second tilting balancing member consisting of a balancing member as described above, guided in rotation about a second coplanar and preferably parallel tilting balancing axis to the first tilt balance axis.

According to one embodiment, the tilting balancing system comprises a frame in which the bearings are housed, at least one balancing member, the chambers and pipes necessary for the action of the control device, see, where applicable, 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.

[0080] 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 tilt balancing around the first tilt balancing axis in a given direction at a given speed causes the second tilt balancing member to rotate around the second tilt balancing axis in the opposite direction and at the same speed .

[0081] The characteristics disclosed with respect to one aspect of the invention or an embodiment can naturally be combined with other characteristics disclosed with respect to other aspects of the invention or other embodiments, to make variations.

Brief description of the drawings

[0082] Other characteristics and advantages of the invention will become apparent on reading the description which follows, with reference to the appended figures, which illustrate: [fig.1] FIG. 1, a block diagram of a system of balancing the rocking torque and an associated internal combustion engine, according to a first embodiment of the invention; [fig.2] Figure 2 is a block diagram of a tilt torque balancing system and associated internal combustion engine, according to a second embodiment of the invention; [Fig.3A] Figure 3A, a diagram illustrating the variations of the tilting torque as a function of the speed of revolution of the crankshaft at no load, at half load and at full load, of an internal combustion engine with regular ignition; [fig.3-13] figure 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 regular firing internal combustion engine; [fig.3C] 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 speed of revolution of the crankshaft in the middle of the speed range of the engine; [Fig.4] 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 tilting torque compensation strategy, in a position maximum engine tilt torque compensation; [Fig.5] Figure 5, a block diagram of the balancing system provided with the control device according to Figure 4, in an intermediate tilting torque compensation position; [fig.6] Figure 6, a block diagram of the balancing system provided with the control device according to Figure 4, in a neutral position; [fig.7] Figure 7, a block diagram of the balancing system provided with the control device according to Figure 4, in a tilting torque compensation position generated mainly by the forces of inertia; [Fig.8] 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 tilting torque compensation strategy, in a position motor tilt torque compensation; [fig.9] Figure 9, a block diagram of the balancing system provided with the control device according to Figure 8, in a neutral position; [Fig.10] Figure 10, a block diagram of the balancing system provided with the control device according to Figure 8, in a tilting torque compensation position generated mainly by the forces of inertia; [fig.11] figure ii, a block diagram of the balancing system of figure 2 provided with a control device according to a third embodiment, executing a third tilting torque compensation strategy, in a position maximum engine tilt torque compensation; [fig.12] 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; [fig.1.3] Figure 13, a block diagram of the balancing system provided with the control device according to Figure 11, in a neutral position; [fig.14] figure 14, a block diagram illustrating a position of the crankshaft of the internal combustion engine associated with a tilting torque compensation system according to figure 2, corresponding to a maximum over one revolution of the tilting torque, and the corresponding position of the compensation system; [Fig.15] Figure 15, a device for modulating the phase difference between the rotation of the crankshaft of the internal combustion engine and the rotation of the tilting torque compensation system; [fig.16] Figure 16, a possible positioning of the tilting torque compensation system relative to the engine block; [fig.17] Figure 17, another possible positioning of the tilting torque compensation system relative to the engine block; [fig.18] Figure 18, another possible positioning of the tilting torque compensation system relative to the engine block; [fig.19] Figure 19, another possible positioning of the tilting torque compensation system relative to the engine block; [fig.20] Figure 20, another possible positioning of the tilting torque compensation system relative to the engine block; [fig.21] Figure 21, another possible positioning of the tilting torque compensation system relative to the engine block; [fig.22] 1a Figure 22, another possible positioning of the tilting torque compensation system relative to the engine block; [fig.23] FIG. 23, an engine configuration incorporating a tilting torque compensation system according to the invention; [fig.24] FIG. 24, another engine configuration incorporating a tilting torque compensation system according to the invention; [fig.25] Figure 25, another engine configuration incorporating a tilting torque compensation system according to the invention; [fig.26] Figure 26, another engine configuration incorporating a tilting torque compensation system according to the invention; [fig.27] Figure 27, another engine configuration incorporating a tilting torque compensation system according to the invention; 21 Ifig.281 Figure 28 is a block diagram of a rocker torque and ram force balancing system and associated inline four-stroke four-stroke engine according to one embodiment of invention; Fig.291 Figure 29 is a block diagram of a rocker torque and ram force balancing system and associated inline four-stroke four-stroke engine, according to another embodiment of the invention; FIG. 30, a block diagram of an angular offset between a balancing weight dc the pestle force and balancing weights of the tilting torque; [fig.31] Figure 31, an isometric view of a body for balancing the tilting torque and the pestle force according to one embodiment of the invention, in a neutral position; [Fig.32] Figure 32, a dc side view of the dc tilting torque balancing member and the pestle force of Figure 31; [Fig.33] Figure 33, a dc front view of the body for balancing the tilting torque and the pestle force of Figure 31; [Fig.34] Figure 34, a view dc the body for balancing the tilting torque and the effort dc rammer dc Figure 31 in section along the section plane B-B illustrated in Figure 33; [Fig.35] Figure 35, a view dc the body for balancing the tilting torque and the force dc rammer dc Figure 31 in section along the section plane A-A of Figure 32; [Fig.36] Figure 36, a view dc the body for balancing the tilting torque and the force dc rammer dc Figure 31 in section along the section plane J-J illustrated in Figure 32; [fig.37] Figure 37, a detail view of a foot dc the body for balancing the tilting torque and the pestle force of Figure 31; [fig.38] figure 38, a view of the body for balancing the tilting torque and the pestle force of figure 31 in section along the plane B-B illustrated in figure 33, but in a position of maximum tipping torque compensation.

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

[0084] In Figure 1 is schematically illustrated an internal combustion engine with four cylinders in line 10, comprising an engine block 12, a crankshaft 14 guided in rotation about a geometric reference axis 100 of the engine block 10, one or more pistons 16 guided in translation in cylinders 18 of engine block 12, and connecting rods 20 connecting the pistons to crankshaft 14.

Remarkably, the internal combustion engine 10 is equipped with a balancing system 22 comprising a first tilting balancing member 24A and a second tilting balancing member 24B.

The first tilting balancing member 24A consists of a body 25A guided in rotation relative to the engine block 12 around a first tilting balancing axis 102A via one or more bearings 62A, and a first tilting balancing mass 26A of fixed value M in rotation with respect to the body 25A.

The second tilting balancing member 24B is constituted by a body 25B guided in rotation with respect to the engine block 12 around a second tilting balancing axis 102B parallel to the first tilting balancing axis 102A and distant from the first tilting balancing axis 102A via one or more bearings 62B, a second tilting balancing mass 26B of value Mo fixed in rotation with respect to body 25B.

In this embodiment, body 25A is balanced relative to first tilt balance axis 102A and body 25B is balanced relative to second tilt balance axis 102B.

The first tilting balancing axis 102A and the second tilting balancing axis 102B are located in a geometric tilting balancing plane Q parallel to the reference axis 100 and perpendicular to a reference plane P of the internal combustion engine 10 containing the reference geometric axis 100.

In the schematic illustration of Figure 1, the internal combustion engine 10 shown 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 may be parallel to the reference plane P, or even, in the illustration of figure 1, coincident with the reference plane P.

The first tilting balancing mass 26A has its center of gravity located on a first side of the reference plane P at a distance D from the reference plane and at a distance E from the first tilting balancing axis 102A .

The second tilting balancing mass 26B has its center of gravity located on the same first side of the reference plane P, at a distance D 2 from the reference plane P and at a distance E , from the second balancing axis failover 102B.

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

Is illustrated in Figure 1 drive means 28 which 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 revolution speed of the crankshaft 14, and to drive the second tilt balancing member 24B in rotation about the second tilt balancing axis 23 102B at a revolution speed of equal amplitude and opposite direction to the speed of revolution of the first tilt balancing member 24A.

In this case, these drive means 28 are here constituted by a bevel gear 30 between the crankshaft 14 and a countershaft 32, which drives a first chain 34 for driving the first balancing member of tilting 24A, and by a second drive chain 36 with reversing function between the first tilting balancing member 24A and the second tilting 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.

[0090] Thus, for the internal combustion engine 10 with four strokes with four cylinders 18 in line illustrated in FIG. 1, the constant ratio between the speed of revolution of the first tilting balancing member 24A and the speed of revolution of the 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 tilting balancing member 24A and the speed of revolution of the crankshaft 14 is chosen equal to N/2.

[0091] 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 tilting balancing member 24A and the revolution speed of the crankshaft 14 is chosen equal to N.

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

The embodiment of FIG. 2 differs from that of FIG. 1 in that the first tilting balancing member 24A comprises a third tilting balancing mass 38A of value M3 located, with respect to the first tilt balance mass 26A, on the other side of the P plane and on the other side of the Q plane.

Similarly, the second tilting balancing member 24B comprises a fourth tilting balancing mass 38B of value MI located, with respect to the second tilting balancing mass 26B, on the other side of the plane Pet of the other side of the Q plane.

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 E 3 from the first balancing axis of failover 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 D 4 from the reference plane P 24 and at a distance E 4 from the second balancing axis failover 102B.

[0095] In all the embodiments, the first tilting balancing axis 102A and the second tilting balancing axis 102B are symmetrical to one another with respect to a plane of symmetry of tilting 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 tilting balancing mechanism 22, it will be assumed initially that in the embodiments of FIGS. 1 and 2, the values M, of the first tilting balancing mass 26A and the value of the second tilting balancing mass 26B are equal, and the centers of gravity of the first tilting balancing mass 26A and the second tilting balancing mass 26B are in positions symmetrical to each other. the other with respect to the plane of symmetry S.

Thus, the distances E 1 and E 2 are equal, and the distances D1 and D 2 are equal.

Thus, the joint rotation of the first tilting balancing mass 24A and of the second tilting balancing mass 24B around their respective balancing axis 102A, 102B, in opposite directions, generates a force having a resultant always located in the plane S. perpendicular to the geometric tilting balancing plane Q at the distance D i from the reference axis 100, which results in a non-zero moment with respect 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.

A first means of generating a torque in the opposite direction to the tilting torque is thus available.

For the embodiment of FIG. 2, it will further be assumed that the values M and M4 of the third tilting balancing mass 38A and of the fourth tilting balancing mass 38B are equal and that the centers of gravity of the third and fourth tilting balance masses 38A, 38B are in positions symmetrical to each other with respect to the plane of symmetry S.

Thus, the distances E 3 and E 4 are equal, and the distances D 3 and D 4 are equal.

[0099]Thus, the joint rotation of the third and fourth tilting balancing masses 38A, 38B around their respective balancing axis, in opposite directions, generates a force having a resultant always located in the plane S, perpendicular to the geometric tilting balancing plane Q at the distance D 3 from the reference axis 100, which results in a non-zero moment with respect 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.

A second means of generating a torque in the opposite direction to the tilting torque is thus available.

[0100] The torque generated by the third and fourth tilt balancing masses 38A and 38B combines perfectly with the torque generated by the first and second tilt balancing masses 26A, 26B if the third and fourth balancing masses 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 balancing plane of Q-tilt, the third mass being otherwise always on one side of the Q-tilt balancing geometric plane opposite the first mass.

If moreover the values MI and M3 are equal (and consequently if M, and M4 are equal), the resultant 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.

[0102] In practice, the rocking torque varies in intensity and in direction with the speed of revolution of the crankshaft and with the engine torque.

On the diagram of FIG. 3A are represented: 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 speed of revolution of the engine, here beyond 2000 rpm, the tilting torque is a monotonic function, in this case decreasing, of the speed of revolution of the crankshaft .

On the diagram of FIG. 3B are represented: on the abscissa axis the load of the engine, that is to say the engine torque related to the maximum torque, and on the ordinate axis and in algebraic value, the torque rocking 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 inertia 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 rocking torque is a monotonically increasing function of the engine torque, at least for a predetermined range of revolution speed regimes of the crankshaft.

Figure 3C, for its part, illustrates the variations of the rocking torque in amplitude and in phase, for the same speed of revolution of the crankshaft, with no load and at 50% of the maximum torque, for a combustion which is triggered at approximately 45° with respect to the chosen frame of reference.

[0103] The balancing systems 22 of the embodiments of FIGS. 1 and 2 can therefore be optimized for a given engine speed and torque by opportunely fixing 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, or 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.

[0105] Different devices for modifying the balancing torque generated by the balancing system 22 are illustrated in Figures 4 to 13.

In Figures 4 to 7, there is illustrated the first balancing member 24A of the balancing system of Figure 2, 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 respect to 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 D and D 3 without varying the distances E and F3.

The control mechanism 42 is driven here by motor means 44 driven by a control unit 46 as a function of input signals from a crankshaft speed determining unit 48 and a torque determining unit engine 50.

In this embodiment, the distances D and D 3 are equal and the masses 26A, 38A move together.

It will be understood that, moreover, the second balancing member 24B is also equipped with a similar control mechanism driven 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 moved together 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 Figure 6, the tilt balancing weights 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 passed through the plane P, and are in a pet-position providing compensation for the tilting torque linked to the inertias.

[0109] Figures 8 to 10 illustrate 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 balancing masses 26A.

27 38A radially relative to the first tilt balancing axis (102A) so as to vary the distances E 1 and E 3 without varying the distances D1 and D 3 .

In this embodiment, the distances E1 and E 3 are equal and the masses 26A, 38A move together.

It will be understood that, moreover, the second balancing member 24B is also equipped with a control mechanism driven by the control unit 46 to move the two masses 26B, 38B so that they are always in a symmetrical position with respect to the masses 26A, 3M.

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 tilting balancing masses 26A, 38A, 26B, 38B are aligned on the first tilting balancing axis 102, and no longer generate any compensation torque around the axis of revolution 100.

Finally, in the position of FIG. 10, the masses 26A, 38A are in mirror position with respect to the position of FIG. 8, thus allowing compensation for the tilting torque linked to the inertias.

It is of course possible to combine the operating modes of FIGS. 4 to 7 on the one hand, and 8 to 10 on the other hand, by making it possible to modify both D , D 3 , E ctE 3 .

It should be noted that in the embodiments of FIGS. 4 to 10, it is not necessary that the values MI and 413 of the masses 26A and 38A be equal.

A particular solution is given by M, = M', which is characterized by the total symmetry of the mechanism.

Furthermore, assuming that M3=0, Figures 4 to 10 become applicable to the embodiment of Figure 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 of balancing 26A parallel to the first tilting balancing axis so as to vary the distance D i , the mass 38A remaining fixed with respect to the body 25A and D 3 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 with respect to the mass 26A.

In the position of Figure 11, the compensation of the tilting torque related to the combustion pressures is maximum.

In the position of FIG. 12, mass 26A is in the reference plane, and the compensation torque, corresponding to the action of masses 26A, 38A, is lower.

Finally, in the position of FIG. 13, mass 26A has joined mass 38A, the effects of the two masses cancel each other out and the compensation torque is zero.

It is possible to define a rotating orthonormal reference (0, x, y, z) linked to the tilting balancing member 24A, having as its origin 0 the center of gravity of the tilting balancing member 24A, its axis y coinciding with the axis of rotation 28 of the tilting balancing member 24A, and its axis z positioned so that the plane (0, y, z) contains the centers of gravity of the moving masses MI and M3 .

In this frame of reference, the inertia matrix MA of the tilting balancing member 24A can be expressed as follows: IA,xx 0 14 xx i4,xx 0 /A, xz 0 IA0 / /A, yz yy 14)7 /A,z), /4.77 A , X 7 IA,yz r , 7 7 ,

According to the theory of inertia tensors, this matrix is symmetric, so that the components of the matrix outside the diagonal, called inertia products, are equal two by two: / u = / .

Assuming that the center of gravity of the tilt balancing member 24A is on the axis of rotation, there will also be Ia,'..==0.

In this frame of reference, the products of inertia and h vary according to the displacements of the mobile masses MI and M, the other components being unaffected by the displacements of the mobile masses MI and M.

In a similar way, it is possible to define another rotating orthonormal referential (0, x, y, z) linked to the tilting balancing member 24B, having as its origin 0 the center of gravity of the balancing member. tilting balance 24B, its y-axis coinciding with the axis of rotation of the member (the tilting balance 24B, and its z-axis positioned so that the plane (0, y, z) contains the centers of gravity of the mobile masses M2 and M4.

In this frame of reference, it is possible to express the inertia matrix MB of the tilting balancing member 24B as follows: = xx 0 /il, X7 = 1-13, XX 0 IB, IE, yz 0 Isyy IB, yz syy 1B,zx IB,zy IB,zz ,IB,xz B , y z

Like the previous one, this matrix is symmetric.

In the chosen frame of reference, the products of inertia and IB', vary according to the displacements of the mobile masses M., and M4, the other components being unaffected by the displacements of the mobile masses 1\42 and M4.

Preferably, we will choose: / = / , .

A, yz'

[0120] Naturally, it is advantageous to position the tilting balancing masses 26A, 26B, and where appropriate 38A, 38B so that the compensation torque they generate on one 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 29 and ignition of the gas.

The piston 16 then exerts on the liner of the cylinder 18 a force whose radial component with respect to the axis 104 of the cylinder 18 varies among other things according to the gas pressure, the resistive torque of the crankshaft and the angle of the connecting rod 20.

It is this force 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 20 which can vary according to the operating conditions.

If we take as the origin of the reference frame of rotation of the crankshaft the position of the crankshaft 14 corresponding to the top dead center of a reference piston at the start of the engine stroke, we observe that the radial force on the cylinder liner 18 reaches its maximum after a 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 compensation effect under all engine operating conditions, it will be advantageous to set the tilting torque balancing system 22 at an angle a located in the central part of this range, between 10' and 50'. , preferably between 20' and 40', typically at or around 30'.

In practice, this means that the tilting balance masses 26A, 26B, and if applicable 38A, 38B will have to be at their peak (in the position farthest from the tilting balance plane Q) when the crankshaft is in the angular position a provided with respect to the top dead center of the cylinder.

It is also possible to provide a phase shift modulation device 52, illustrated in FIG. 15, making it possible to vary the phase shift between the tilting 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 previously described, a first chain 34 for driving the first tilting balancing member 24A, a tensioner 54 could 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 difference between rotation of the crankshaft 14 and rotation of the two balancing members 24A, 24B.

Naturally, other phase shift modulation devices could be envisaged.

In practice, the phase shift range should be at least 100, 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 that contains the tilting balancing axes 102A, 102B , or 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 plane Q.

Figure 16 repeats the positioning of Figures 1 and 2, namely perpendicular to the thrust plane PP, at a distance from the axis of revolution of the crankshaft 100.

FIG. 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 coincide.

FIG. 20 illustrates a tilting balancing plane Q parallel to the axis of revolution 100 but at a distance from the latter and oblique with respect to the thrust plane PP.

Finally, FIG. 21 illustrates a tilting dc balancing plane Q parallel to the thrust plane PP, and distant from the latter.

In practice, the embodiments of FIGS. 16, 20 and 21, in which the balancing plane dc tilting 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 tilting balancing member 24A has a first main axis of inertia which is not parallel to the first tilting balancing axis 102A and which rotates about first tilt balance axis 102A, and second tilt balance member 24B has a second major axis of inertia that is not parallel to second tilt balance axis 102B and rotates around of the second tilting balancing axis 102B, compensation for the tilting torque can be obtained in particular if the first main axis of inertia and the second main axis of inertia are at all times symmetrical with respect to each other to a tilting symmetry plane S perpendicular to the reference geometric axis.

In all of the preceding figures, the engine 10 illustrated is an in-line four-cylinder internal combustion engine.

However, the invention can advantageously be implemented 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-cylinder engine cylinders as shown in figure 25.

One can similarly generalize the use of the tilting torque balancing system 22 according to the invention to an in-line N-cylinder engine, N being any whole number greater than or equal to 1.

It is further possible to generalize the use of the tilting torque balancing system 22 according to the invention to a V-shaped multi-cylinder engine as illustrated in FIG. 26 or to a radial engine as illustrated in FIG. 27, and more generally to any internal combustion engine with regular ignition (that is to say at intervals of rotation 31 of the regular crankshaft between two successive combustions).

The tilting torque is not the only source of parasitic forces transmitted by the engine block 12 to its attachment 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 resultant of the forces exerted by the engine block 12 on its supports.

The drumstick effort has a fundamental frequency which depends on the number of cylinders 18, the sequence of ignitions and the speed of the crankshaft 14.

In the particular case of the four-stroke engine with four cylinders in line, the fundamental frequency of the pestle force and that of the rocking torque coincide.

It is within this framework that various embodiments of the invention will now be described providing both compensation for the tilting torque and compensation for the pestle force.

The embodiment of FIG. 28 differs from that of FIG. 1 in that the first tilting balancing member 24A comprises, in addition to the elements of FIG. 1, a first pestle balancing weight 56A of value PI, having a center of gravity located in the thrust plane dc PP of the internal combustion engine, at a distance F1 from the first tilt balancing axis 102A.

Similarly, the second tilt balancing member 24B comprises a second ram 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 F2 from the second tilt balancing axis 102B.

The embodiment of Figure 29 combines the embodiments of Figures 2 and 28.

Thus, the first tilting balancing mass 24A comprises a first tilting balancing mass 26A, a third tilting balancing mass 38A and a first pestle balancing mass 56A, while the second balancing member 24B has a second tilt balance mass 26B, a fourth tilt balance mass 38B and a second ram balance mass 56B.

[0130] For the embodiments of Figures 28 and 29, which are specific to an inline four-cylinder four-stroke engine, it will be further assumed that the PI and Po values of the first and second balance masses of the drumsticks 56A and 56B are equal and the centers of gravity of the first and second balancing masses of the pestles 56A and 56B are in positions symmetrical to each other with respect to the plane of symmetry S.

Thus, the distances F and F2 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 with respect to the reference plane P. 32

Thus, the joint rotation of the first and second pestle balancing weights 56A, 56B around their respective balancing axis, in opposite directions, generates a force having a resultant always located on the axis of intersection 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 resultant, at the period of the pestle generated by the combustion cycles, and which can be calibrated so as to compensate at least partially pestle effort.

In FIGS. 28 and 29, the balancing mass of the pestle 56A, 56B has been illustrated perpendicular to the tilting balancing masses 26A, 38A, 26B, 38B of the same balancing member 24A, 24B.

This relative positioning generates a compensation force of the pestle in quadrature with the tilt compensation torque.

However, it can be seen that, as a general rule, the phase difference 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 1150.

One will therefore choose for a given engine and for each balancing member 24A, 24B a relative angular positioning between the balancing mass of the pestle 56A, 56B and the tilting balancing mass or masses 26A, 38A, 26B, 38B which in the range of 65' to 1150, will correspond to the motor characteristic, as shown in figure 30.

[0133] An embodiment of a tilting balancing member 24A, and if necessary for balancing the pestle force, has been illustrated in Figures 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 guide bearings end 62, 64.

The body 25A makes it possible to accommodate two sliding weights for tilting balancing 26A, 38A and two return springs 68, 70.

Between the tilting balancing mass 26A and the domed casing 56A is formed a variable volume chamber 72.

Similarly, a variable volume chamber 74 is formed between the tilting balance 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, via supply channels opening out at the level of 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 on figure 34 to an extreme position illustrated in figure 38 against the return force of the springs 70, 68.

In the neutral position of Figure 34, the tilt balancing masses 26A, 38A are in the position shown schematically in Figure 6.

The position of figure 38 corresponds 33 to the position illustrated in figure 4.

Finally, it is possible, from the neutral position of FIG. 34 and in the absence of pressure in the chambers 72. 74, to bring the balancing weight 26A to the left at the limit stop and to bring the balancing weight 38A to the right at the end of travel, under the effect of the return springs 7068, which constitutes the state schematically illustrated in Figure 7.

[0134] 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 B-B located in Figure 33, at a distance and on both sides other of the section plane J-J located in figure 32.

The center of gravity of the body 25A is in the section plane A-A located in FIG. 32, or in a reference plane C of the body 25A in the immediate vicinity of the section plane A-A.

In the position of Figure 8, the center of gravity of the balancing weight 26A and the center of gravity of the balancing weight 38A are on either side of the cutting plane A-A, and dc part and d the other of a plane containing the body's center of gravity and perpendicular to the tilting balance axis 102A.

[0135] If the dc tilting balancing member 24A is not intended to balance a pestle effort, the body 25A is constituted dc so as to be dynamically balanced in its rotation around the tilting balancing axis 102A defined by bearings 62, 64

If, on the other hand, provision is made for the balancing of a pestle force, the body 25 A will be shaped so as to present an imbalance.

If necessary, the curved envelope of the body 25A has a bulging side which creates an imbalance constituting the balancing weight of the pestle 56A, as illustrated more particularly in Figure 37.

An angle t] of approximately 90° has been illustrated in FIG. 37, but the angle can naturally vary according to the distribution of the masses at the level of the bulge.

101371 Of course, many variations on the structure of the tilt balance member 24A are possible.

It is possible to envisage a control other than hydraulic, for example by endless screw or by cable.

If one wishes to control the balancing weights 26A.

38A jointly, a mechanical connection can be provided between the two, for example a rack and pinion or cable connection.

Conversely, it may be desired to make the control of the tilting balancing mass 26A and the control of the tilting balancing mass 38A independent of each other, which would lead to providing, in the device of FIGS. at 38, two independent hydraulic circuits.

The return springs 68, 70 can either be compression springs or tension springs.

More generally, a pneumatic or cam return system can be provided.

34 [Claim 1]

Claims (1)

CLAIMSMethod for balancing an internal combustion engine (10) with one or more reciprocating pistons (16), comprising an engine block (12), a crankshaft (14) rotating relative to the engine block (12) of the combustion engine internal (10) around a reference geometric axis (100), along which rotation is driven relative to the engine block (12) around a first tilting balancing axis (102A) not parallel to the axis reference (100), at a speed of revolution having a predetermined ratio with the speed of revolution of the crankshaft, a first tilting balancing member (24A) having a first dynamic unbalance, characterized in that one rotates with respect to an engine block (12) around a second tilt balancing axis (102B) not parallel to the reference axis (100), at a speed of revolution of equal amplitude and opposite direction to the speed of revolution of the first tilt balancing member (24A), a second tilt balancing member (24B) having a second dynamic unbalance, by synchronizing the rotation of the first tilt balancing member (24A) and the second tilting balancing member (24B) such that the first tilting balancing member (24A) and the second tilting balancing member (24B) together generate on the engine block (12) at least a torque of tilting compensation around the geometric reference axis, characterized in that the first tilting balancing axis (102A) and the second tilting balancing axis (102B) lie in a geometrical tilting balancing plane (Q), the first tilt balance member (24A) includes a first tilt balance mass (26A) having a first value M1 and a first center of gravity located on a first side of a reference plane (P) containing the reference axis (100) and perpendicular to the tilting balance plane. (Q), at a distance D1 from the reference plane (P) and at a distance El from the first tilting balancing axis (102A), the second tilting balancing member (24B) comprises a second balancing mass (26B) having a second value M2 and a second center of gravity located on the first side of the reference plane (P) at a distance D2 from the reference plane (P) and at a distance E2 from the second balancing axis of tilting (102B), and at least one of the distances D1, El, and at least one of the distances D2 is varied. E2, as a function of one or more control parameters among motor operating parameters. Balancing method according to claim 1, characterized in that the first tilting balancing axis (102A) and the second tilting balancing axis (102B) are symmetrical to each other with respect to a plane of tilting symmetry (S), the plane of tilting symmetry (S) preferably being perpendicular to the reference geometric axis (100). A balancing method according to any preceding claim, characterized in that the first tilt balancing axis (102A) and the second tilt balancing axis (102B) are parallel. Balancing method according to any one of the preceding claims, characterized in that the first tilting balancing member (24A) is driven by means of a first kinematic connection (34) between the crankshaft (14 ) and the first tilting balancing member (24A) and the second tilting balancing member (24B) is driven by means of a second kinematic connection (36) between the crankshaft (14) and the second tilt balancer (24B) or between the first tilt balancer (24A) and the second tilt balancer (24B). Balancing method according to any one of the preceding claims, characterized in that the rotation of the first tilting balancing member (24A) and of the second tilting balancing member (24B) is synchronized in such a way that the tilt compensation torque generated is maximum in absolute value over one revolution of the first tilt balancing member when the crankshaft (14) 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 (18) of the internal combustion engine (10), the angle a preferably being between 00 and 50° in the direction of revolution of the crankshaft (14). Balancing method according to Claim 5, characterized in that the angle a is modulated as a function of one or more phasing parameters from among motor operating parameters, the phasing parameter(s) preferably comprising at least at least one of the following parameters: 36 [Claim 7] [Claim 8] [Claim 9] [Claim 10] [Claim 11] [Claim 12] 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. Balancing method according to any one of the preceding claims, characterized in that the geometric tilting balancing plane (Q) is parallel to the geometric reference axis (100). Balancing method according to Claim 7, characterized in that the geometric tilting balancing plane is distant from the geometric axis of revolution of the crankshaft Balancing method according to any one of the preceding claims, characterized in that the 'D, E1, D0, and/or E, are varied so that: MI.D _NE Balancing method according to any one of the preceding claims, characterized in that the control parameter or parameters include a or more of the following parameters: the speed of revolution of the crankshaft (14), the engine torque, the engine power, the angle of the crankshaft corresponding to a maximum pressure in a cylinder of the engine, the maximum value of the pressure reached in a cylinder of the engine, and preferably at least the speed of revolution of the crankshaft (14) in combination with at least one second parameter among the following parameters: 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. Balancing method according to any one of the preceding claims, characterized in that the distance between the first tilting balancing mass (26A) and the first tilting balancing axis (102A) is constant, and the the distance between the first tilt balance mass (26A) and the reference plane (P) as a function of the engine operating parameter(s). Balancing method according to any one of the preceding claims, characterized in that the distance between the first tilting balancing mass (26A) and the reference plane (P) is varied in such a way that, for each revolution speed of the crankshaft (14) within a given revolution speed range, the distance between the first rocker balancing mass (26A) and the reference plane (P) is a monotonic function of the motor torque. Balancing method according to any one of the preceding claims, characterized in that the distance between the first tilting balancing mass (26A) and the reference plane (P) is varied in such a way that, for each engine torque within a given engine torque range, the distance between the first tilting balance mass (26A) and the reference plane (P) is a monotonic function of the revolution speed of the crankshaft (14), at least when the revolution speed of the crankshaft is greater than a predetermined speed threshold. Balancing method according to any one of Claims 1 to 13 applied to a four-stroke internal combustion engine with N cylinders in line, N being an integer greater than or equal to 1, characterized in that the constant ratio between the revolution speed of the first tilt balancing member (24A) and the revolution speed of the crankshaft (14) is equal to N/2. Balancing method according to any one of Claims 1 to 13 applied to an in-line N-cylinder two-stroke internal combustion engine. N being an integer greater than or equal to 1, characterized in that 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 equal to N. Method balancing device according to any one of Claims 1 to 15, characterized in that: the first tilting balancing member (24A) comprises a third tilting balancing mass (38A) of value M' having a center of gravity located on a second side of the reference plane (P) opposite the first side, at a distance DI from the geometric reference plane (P) and at a distance E' from the first tilting balancing axis (102A), the second tipping balancing member (24B) comprises a fourth tipping balancing mass (38B) of value M4 having a center of gravity located on the second side of the reference plane (P) at a distance D4 from the geometric reference plane ( P) and at a distance E4 from the second tilting balancing axis (102B), and at least one of the distances D3, E3, and at least one of the distances D4, E4 are varied as a function of one or several control parameters from among motor operating parameters, preferably so that: 38 = fr14-1) 4-L