EP1859135B1 - Sehr kompakte vorrichtung zum verstellen des verdichtungsverhältnisses eines brennkraftmotors - Google Patents

Sehr kompakte vorrichtung zum verstellen des verdichtungsverhältnisses eines brennkraftmotors Download PDF

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Publication number
EP1859135B1
EP1859135B1 EP06709379A EP06709379A EP1859135B1 EP 1859135 B1 EP1859135 B1 EP 1859135B1 EP 06709379 A EP06709379 A EP 06709379A EP 06709379 A EP06709379 A EP 06709379A EP 1859135 B1 EP1859135 B1 EP 1859135B1
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EP
European Patent Office
Prior art keywords
crankshaft
eccentric
compression ratio
piston
adjustment
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EP06709379A
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English (en)
French (fr)
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EP1859135A1 (de
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Michel Marchisseau
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/04—Engines with variable distances between pistons at top dead-centre positions and cylinder heads
    • F02B75/048—Engines with variable distances between pistons at top dead-centre positions and cylinder heads by means of a variable crank stroke length

Definitions

  • the present invention relates to a device for adjusting the compression ratio of an internal combustion engine and to a method enabling the use of such a device.
  • It relates more particularly to a device that can change the compression ratio of this engine by changing the dead volume of the combustion chamber at the top dead center of the piston.
  • the bore of the big end has axial grooves which cooperate with a locking pin disposed in the eccentric, radially relative to the axis of said eccentric so as to immobilize it in one of the positions corresponding to one of said axial grooves of the connecting rod.
  • This device for adjusting the compression ratio has many advantages: it is located in the moving equipment and is energy efficient: it is powered by energy supplied directly by the moving equipment. Its implementation is not very constraining: it affects neither the combustion chamber nor the connections with the exhaust, or with the distribution members or with the transmission, nor the weight of the piston. It is nevertheless perfectible.
  • the big end is very bulky in order to accommodate both the eccentric and the mechanical locking device.
  • the control of the compression ratio adjustment device shall transit through several movable members relative to each other: the housing, the crankshaft and finally the eccentric linked to the connecting rod.
  • the mechanical locking system is necessarily subjected to high levels of friction and stress, and even possibly to shocks. This aspect combined with the little space available in the eccentric, affects the service life.
  • the eccentric is not an eccentric towed type but a motorized eccentric.
  • An electric or hydraulic motor drives an irreversible worm which cooperates with a toothed sector of the eccentric.
  • This device has a major disadvantage because the electric or hydraulic motor must counteract the various friction and different inertial forces, including those of the mobile engine of the internal combustion engine, to motorize the eccentric. But these friction and these forces of inertia are very important. Said electric or hydraulic motor is therefore necessarily bulky. In addition, the energy to power this motor must be provided by a subsidiary body. The yield is therefore heavily penalized.
  • the present invention proposes to overcome the drawbacks mentioned above by means of an energy saving compression ratio adjusting device, a compactness at the best level while being easily compatible with a long service life.
  • the present invention relates to a device for adjusting the compression ratio of an internal combustion engine comprising at least one cylinder with a combustion chamber, a mobile assembly comprising a piston displaceable in translation under the action of a rod connected by an axis to said piston and connected to a crankpin of a crankshaft, said piston making a race between a top dead center and a bottom dead center leaving a dead volume at the top dead center of said piston, the device comprising between the crankpin and the crankpin of the crankshaft a rotary eccentric to adjust the compression ratio, the device also comprising means for controlling the displacement of the eccentric, characterized in that the control means comprise at least one kinematic connection without a latch between a radial protuberance integral with the eccentric and an adjustment mechanism in position relative to the crankshaft, said m canism being integrated in one of the two blanks of the crankshaft, said kinematic linkage and said adjustment mechanism in position relative to the crankshaft being on the one hand disposed largely or entirely outside the crankpin
  • the adjustment mechanism in position relative to the crankshaft belonging to the control means of the displacement of the eccentric according to the invention comprises at least one linear actuator.
  • the advantage of a linear actuator lies in the simplicity.
  • the present invention combines several critical advantages, which are never all together simultaneously in the designs described in the prior art.
  • the second advantage is an important volume available to house the means for controlling the displacement of the eccentric.
  • This second advantage is compatible with the aforementioned first because according to the invention the main components of said control means: the kinematic connection and the adjustment mechanism in position relative to the crankshaft, are integrated in the crankshaft blank, in space adjacent to the lever, crankpin and bearing of the crankshaft. This volume is important and available in the usual size of the mobile equipment of a traditional engine.
  • the features of the present invention therefore combine a long service life with a compactness at the best level and a large volume for the integration of the device.
  • the means for controlling the displacement of the eccentric according to the invention make it possible to carry out a continuous adjustment of the compression ratio over its range of variation.
  • This feature associates with the above advantages of the present invention the ability to adjust at any point the compression ratio to the optimum value.
  • the means for controlling the displacement of the eccentric according to the invention use energy taken from the moving equipment to move the eccentric.
  • This characteristic associates with the above-mentioned advantages of the present invention the possibility of adjusting the compression ratio with a high reactivity and a high energy consumption of the device Furthermore, this characteristic also combines the advantage of not being able to draw on the peripherals of the device. mobile equipment or the internal combustion engine as the control energy required for the device according to the invention.
  • the means for controlling the displacement of the eccentric according to the invention comprise two sets placed on either side of the eccentric and each consisting of at least one kinematic link without a lock connected to a protuberance radial integral with the eccentric and an adjustment mechanism in position relative to the crankshaft. This characteristic makes it possible to reinforce the robustness for the highly stressed mobile crews.
  • the two adjustment mechanisms in position relative to the crankshaft belonging to the two sets placed on either side of the eccentric, mentioned in the preceding paragraph, are kinematically linked so that they participate in approximately equal to the control of the displacement of the eccentric. This feature reinforces the robustness of the whole.
  • said adjustment mechanism in position of the eccentric relative to the crankshaft according to the invention comprises two linear actuators whose axes are distinct. These two actuators can be single-acting. Thus each actuator can work by simple pushing and act in opposite directions on the orientation of the eccentric. This design simplifies the kinematic connections. In addition, the axial size of a single-acting actuator is less than that of a double-acting actuator.
  • the two Linear actuators are placed on each side of the crankpin and bearing of the crankshaft. This design facilitates the integration of said actuators in the crankshaft blank.
  • the coefficient of friction between the bore of the eccentric and the crankpin is less than seventeen hundredths. This value of the coefficient of friction has the advantage of allowing the eccentric to be towed for many applications.
  • the definition of a towed eccentric is specified in the description of the preferred embodiment.
  • the coefficient of friction between the eccentric and the bore of the big end is greater than twenty hundredths.
  • This value of the coefficient of friction has the advantage of allowing the eccentric not to be towed for many applications.
  • the advantage is that the eccentric does not rotate in the absence of a torque generated by a specific actuator. The eccentric therefore maintains its angular position without requiring any specific means to block it or to maintain it.
  • the means for controlling the displacement of the eccentric according to the invention measures a distance using a non-contact measuring sensor, between a fixed position of the motor housing. and one of the parts that moves relative to the crankshaft to adjust the compression ratio. Measuring a distance of this type has the advantage of enabling the device to determine the compression ratio without significant error.
  • a non-contact sensor provides a long life and high reliability for this measurement.
  • the integral radial protuberance of the eccentric linked to the kinematic connection belonging to the control means of the displacement of the eccentric may be a collar integral with the eccentric. This design has the advantage of distributing the control constraints of the position of the eccentric over the three hundred and sixty degrees of the eccentric.
  • the means for controlling the displacement of the eccentric according to the invention have means for controlling the adjustment mechanism in position relative to the crankshaft.
  • the figures 1 and 2 show an internal combustion engine with at least one cylinder 31 which comprises a bore 16 inside which slides a hollow piston 14 in an alternative translational movement under the impulse of a rod 15.
  • This piston delimits with its upper part, the side wall of the bore 16 and the upper part of this bore, generally formed by a portion of the cylinder head 11, a combustion chamber 10 in which the combustion cycle takes place.
  • the piston carries two diametrically opposed radial bores through which is housed a cylindrical axis 13 which connects the small end 12 to said piston.
  • the connecting rod head 17 is connected by a compression ratio adjusting device 32 to a crank pin 22 of a crankshaft 28.
  • This crankshaft 28 is subjected to a rotational movement about an axis XX.
  • the piston 14, the axis 13, the connecting rod 15, the crankshaft 28 with its crankpin 22 form the moving element of the engine.
  • the crank pin 22 passes successively from a high position to a low position.
  • the piston 14, which is connected to the crank pin 22 by the connecting rod 15, undergoes an alternative translational movement between a top dead center and a bottom dead center.
  • the compression ratio of an engine is a function not only of the extent of the volume of the cylinder delimited by the stroke of the piston but also the magnitude of the dead volume. To modify the compression ratio, simply modify one of these volumes and more particularly the size of the dead volume.
  • the compression rate adjusting device 32 comprises an eccentric 18 housed between the crankpin 22 and a bore 19 provided in the crankshaft head 17.
  • This eccentric 18 has a generally circular shape with a geometric axis X1X1 which corresponds at its middle axis and comprises a bore 20 axis X0X0 non-coaxial with the axis X1X1 but coincides with the axis of the crank pin 22.
  • This eccentric is slidably accommodated in the receiving bore 19 made in the connecting rod head 17 and on the peripheral wall of the crank pin 22.
  • the dead volume of the combustion chamber 10 is a function continuous of the angular orientation of the eccentric 18.
  • the axis of the cone head 17 is merged with the X1X1 axis of the eccentric 18 and the axis of the crank pin 22 with the axis X0X0 of the bore 20 of the eccentric 18.
  • the axis X1X1 of the eccentric 18 is not coaxial with the axis X0X0 of its bore 20.
  • the distance between the axis of the big end 17 and the cylinder head 11 is a continuous function of the angular orientation of the eccentric, defined by example by the angle between firstly the line passing through its axis X1X1 and the axis X0X0 of its bore 20, secondly the reference line YY perpendicular to the axis of the cylinder 31 and the axis X1X1 of the eccentric 18.
  • the figure 2 presents two angular orientations of the eccentric, one in solid line and the other in dotted line, corresponding to two different compression rates of the internal combustion engine.
  • the angular orientation of the eccentric 18 of angles AH between the straight lines YY and DH, the straight lines DH passing through the axes X1X1 of the eccentric 18 and X0X0 of its bore 20, the top dead center of the piston 14 is PMHmax and corresponds to a dead volume VHmin of the combustion chamber 10.
  • the dead volume VHmax is greater than the dead volume VHmin and corresponds to a lower compression ratio of the internal combustion engine.
  • the compression rate adjusting device 32 also comprises means for controlling the displacement of the eccentric 18 including on the one hand the kinematic connection 30 kinematically connected to the flange 21 which constitutes the radial protuberance integral with the eccentric 18, on the other hand the adjustment mechanism in position 29 relative to the crankshaft 28.
  • Said adjustment mechanism in position 29 is integrated in the blank 33 of the crankshaft 28, this integration being more particularly carried out almost entirely in the mass of balancing 26 of said crankshaft 28.
  • Said kinematic connection 30 and said adjustment mechanism in position 29 are integrated entirely outside the crankpin 22, the bearing 27 and the lever 23 connecting the crankpin 22 to the bearing 27 of the crankshaft28.
  • the kinematic connection 30 does not include a lock and the eccentric 18 is free of part or form of a constituent lock.
  • the Figures 3 to 7 presents the first particular preferred embodiment of the invention.
  • the flange 21 secured to the eccentric 18 forms a rocker with two connecting pads 34a, 34b.
  • the adjustment mechanism in position 29 relative to the crankshaft 28 comprises two linear actuators placed on either side of the crank pin 22 and the bearing 27 of the crankshaft 28. These two linear actuators are single acting hydraulic cylinders 36a and 36b of which the axes 37a, 37b are distinct.
  • the kinematic connection 30 between the collar 21 and the adjustment mechanism in position 29 is formed by the top of the rods 30a, 30b of the jacks 36a and 36b which push the connecting studs 34a, 34b of the rocker formed by the collar 21.
  • the kinematic link 30 and the adjustment mechanism in position 29 are completely integrated outside the crankpin 22, the bearing 27 and the lever 23 connecting the crankpin 22 to the bearing 27 of the crankshaft28. They occupy a volume whose point farthest from the axis XX of the crankshaft 28 describes a circle 25, during the rotation of the engine, with a diameter equal to the largest diameter of the circle 24bi described by the connecting rod 17 As well as the circle 24vi described by the flanks 33 of the crankshaft 28.
  • Each single-acting hydraulic cylinder 36a and 36b operates by simple kinematically opposite thrusts by virtue of the function performed by the rocker articulated around the axis X0X0 of bore 20 belonging to the eccentric 18 and coincide with the axis of the crank pin 22.
  • An additional advantage of this compression rate adjustment device is that it does not apply axial direction force on the eccentric 18.
  • the circulation of the oil between the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b is controlled by a hydraulic valve 40 placed in the balancing mass 26 of the crankshaft 28.
  • each constituent sub-assembly of the compression ratio adjustment device makes it possible to achieve positioning on any point within the range of variation.
  • These subassemblies are the eccentric 18 secured to its flange 21 and its pads 34a, 34b which can be positioned at any angle within the range of variation of the positioning angle of the eccentric 18 , the kinematic linkage 30a, 30b which is continuous, reversible and free of component parts of a latch, the linear hydraulic cylinders 36a, 36b which can be positioned at any position in their respective range of variation, otherwise they are also reversible, and finally the hydraulic valve 40 which can supply the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b to achieve any positioning within their range of variation.
  • the eccentric 18 is towed during the operation of the engine and the hydraulic valve 40 allows to allow or prohibit at any time and for an adjustable period, via the hydraulic pipe 42a, 42b, the passage of oil between the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b.
  • This eccentric is said towed when subjected, during operation of the motor, to a driving torque in rotation about the axis X0X0, successively in the clockwise direction of the eccentric 18 and in the anti direction.
  • the deflection angle of the eccentric 18 is thirty degrees above and thirty degrees below the reference line YY.
  • the coefficient of friction of the fluid bearing, placed between the bore of the big end 17 and the eccentric 18, is less than five per thousand.
  • the coefficient of friction between the crankpin 22 of the crankshaft 28 and the eccentric 18 is less than one-tenth.
  • the crankpin 22 of the crankshaft 28 is coated with amorphous carbon and lubricated to guarantee this upper limit of coefficient of friction.
  • the eccentric 18 is accelerated in rotation, with respect to the crankpin 22 of the crankshaft 28, in a direction which depends mainly on the engine time internal combustion in its operating cycle, suction or compression or exhaust or explosion or other, the angle and speed of rotation of the crankshaft 28, and the load of the internal combustion engine.
  • the valve 40 blocks the oil transits between the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b, the rotational position of the eccentric 18 with respect to the crankpin 22 and with respect to the crankshaft 28 is stopped because that the oil can not leave the chambers 35a, 35b of the hydraulic cylinders 36a, 36b and that said chambers are free from air.
  • the means for controlling the displacement of the eccentric according to the first preferred embodiment of the invention thus motorize the eccentric with energy taken directly from the moving equipment. Only the energy required to control the control means of the displacement of the eccentric is taken from the peripherals of the internal combustion engine. This feature minimizes the energy required to adjust the compression ratio.
  • the chambers 35a, 35b of the two hydraulic cylinders 36a, 36b are filled with oil permanently by the engine lubrication pump, via the pipes 38a, 38b, the non-return valves 39a, 39b and the pipe 41 of the usual lubrication of the bearing and the crankpin .
  • the mounting direction of the non-return valves 39a, 39b is such that the pipe 41 can supply oil to the chambers 35a, 35b of the hydraulic cylinders 36a, 36b, but the oil returns of said chambers to the pipe 41 are blocked.
  • the first preferred embodiment provides a hydraulic pipe 44 which connects the chamber 35a of the compression ratio increasing cylinder 36a. , to a means of generating hydraulic pressure while the engine is stopped, via a non-return valve 45 which prevents the oil from returning to said hydraulic pressure generating means.
  • This option has the advantage of allowing the engine to be stopped immediately, for any value of the compression rate, n, at the request of the user, while having the highest compression ratio to facilitate the starting of the internal combustion engine in very cold weather.
  • the hydraulic valve 40 is controlled via an electromagnet.
  • Its electric coil is integral with the motor housing and its movable core is embedded on the movable element to enable the hydraulic valve spool to be actuated.
  • the magnetic flux generated by the electric coil transits in the magnetic field conductors 47a, 47b, 47c, 47d integral with the motor housing, and then by air strips 48c, 48d to reach and circulate in the magnetic field conductors 49c, 49d embedded on the mobile unit, specifically for this application, embedded on the balancing mass 26 in the blank 33 of the crankshaft 28 and in the aforementioned movable core.
  • This embodiment has the advantage of a long life because the electrical and electromagnetic control is transmitted without friction. Moreover, a greater choice is possible to place the electric coil without penalizing the overall size.
  • the coil can be powered by continuous electrical connections, without the interface of an electrical collector.
  • the slide, not shown, of the hydraulic valve 40 in the closed position is firstly pushed in the direction of closure by a spring, on the other hand coupled to a double-acting cylinder whose forces exerted each side of his piston are in equilibrium.
  • the hydraulic valve 40 is in the closed position, the hydraulic chambers of said double-acting cylinder are supplied with oil under pressure by the chambers 35a, 35b of the cylinders 36a, 36b, via non-return valves, not shown, in order to prevent any communication of hydraulic fluid between the two chambers 35a, 35b of the cylinders 36a, 36b by this control circuit.
  • said movable core is moved under the action of the magnetic flux generated by the circuit of control, which opens a valve and causes the laying of a hydraulic chamber of the double-acting cylinder so that the pressure drop in the hydraulic chamber generates a double-acting cylinder force in the direction of opening of the hydraulic valve 40.
  • the pressure drops in the circuit of said tarpaulin under the action of said movable core are much lower than the feed losses of the aforementioned hydraulic chamber of the double-acting cylinder by the chambers 35a. , 35b of the cylinders 36a, 36b. The consequence is a rapid opening movement of the hydraulic valve 40.
  • the oil which thus returns to the cover of the motor housing is replaced in the device for adjusting the compression ratio by oil pressurized by the engine lubrication pump, via the circuit comprising the pipes 41, 38a, 38b and the non-return valves 39a, 39b described above.
  • a variant of the control of the hydraulic valve 40 is presented on the figure 10 .
  • the hydraulic valve 40 is actuated by the pusher 52.
  • the movable cams 51a, 51b actuated by a device, not shown, make it possible, during the rotation of the motor, to actuate the pusher 52 or not to actuate it according to the control they receive from the control circuit.
  • the cam 51a opens the valve 40 when the piston is close to the top dead center and the cam 51b when the piston is close to the bottom dead center.
  • the compression ratio adjusted by the device is measured by a non-contact distance measuring sensor 43 fixed on the motor housing.
  • This sensor measures the distance that separates it from the highest point reached by the lateral face of the flange 21 secured to the eccentric 18.
  • Said lateral face is of a helicoidal shape so that the non-contact distance measuring sensor is inclined towards the axis of the crankshaft so that the smallest distance measured by said sensor is a continuous function of the angular orientation of the eccentric 18 relative to the crankshaft.
  • This distance is correlated with the compression ratio by the mechanical kinematics of the device. This distance is therefore a reliable image of the compression ratio.
  • This distance is Smax for the minimum compression ratio and Smin for the maximum compression ratio.
  • the non-contact distance measuring sensor is an eddy current sensor. This type of sensor has the advantage of having a very short response time and high accuracy.
  • the control means of the displacement of the eccentric 18 are doubled and placed on either side of the eccentric.
  • the eccentric 18 is on the one hand integral with a collar 21c placed on the left and controlled in position in particular by the cylinders 36c, 36d integrated in the blank 33a of the crankshaft 28, secondly secured to a second collar 21a located , to the right of the other side of the connecting rod and controlled in position in particular by the cylinders 36a, 36b integrated in the blank 33b of the crankshaft 28.
  • This construction doubles the capacity of control torque in position of the eccentric.
  • the hydraulic chambers 35a, 35c of the jacks 36a, 36c for blocking the rotation of the eccentric in the clockwise direction are in communication via the hydraulic line 46a 46c and the hydraulic chambers 35b, 35d of the jacks 36b, 36d of FIG. blocking the rotation of the eccentric counterclockwise are in communication via the hydraulic pipe 46b 46d.
  • This communication makes it possible to standardize the hydraulic pressures in each pair of jacks which act in the same direction in order to distribute the stresses and thus maximize the robustness of the device.
  • the two hydraulic cylinders 36a and 36b are integrated in a module 80, which module 80 is assembled on the crankshaft 28, positioned relative to its lever 23.
  • the two hydraulic cylinders 36a and 36b are equidistant from the axis X0X0 of the crankpin 22 and the sections of their hydraulic chamber 35a, 35b are identical.
  • This module 80 integrates the hydraulic power circuit comprising the hydraulic cylinders 36a and 36b, the controlled non-return valves 72a and 72b, the booster anti-retow valve 76, the line 75 of communication between the hydraulic cylinders 36a and 36b and the pilot lines 74a and 74b.
  • This module 80 also fulfills the balancing mass function 26.
  • This design has the advantage of making it possible to carry out and test the hydraulic functions for adjusting the compression ratio with the module 80 independently of the crankshaft 28 before assembly of the module 80 on the crankshaft 28.
  • the steering ducts 74a and 74b of the assembled device are connected to annular grooves, not shown, made on the bearing of the crankshaft 28.
  • This design has the advantage of allowing to connect a hydraulic control circuit 81 controlled check valves 72a and 72b, on supports integral with the motor housing.
  • the feeding circuit of the hydraulic cylinders 36a and 36b, controlled by the check valve 76, is supplied by the engine lubrication circuit via the pipes 41 and 77.
  • the return spring 73 makes it possible to put the compression ratio back to the maximum value when the engine is at a standstill.
  • the kinematic links 30a, 30b between the flange 21 of the eccentric 18 and the rods of the hydraulic cylinders 36a and 36b are made with the help of rods 70a and 70b.
  • the joints between the rods 70a and 70b and respectively the collar 21 of the eccentric 18 and the rods of the hydraulic cylinders 36a and 36b are formed respectively by the connecting pads 34a, 34b of hemispherical shape and the ball joints 71a and 71b.
  • the operation of the compression rate adjustment device 32 according to this other way of carrying out the invention in a hydraulic version is controlled by the solenoid valve 79 shown schematically in FIG. figure 13 .
  • the solenoid valve 79 controls the controlled nonreturn valve 72a
  • the only possible oil transfers between the two hydraulic cylinders 36a and 36b are those which make it possible to reduce the compression ratio.
  • the solenoid valve 79 controls opening the controlled nonreturn valve 72b the only possible oil transfers between the two hydraulic cylinders 36a and 36b are those which make it possible to increase the compression ratio.
  • This design makes it possible to drive continuously one of the piloted check valves 72a or 72b during one or more of any motor while obtaining a variation of the compression ratio always in the same direction.
  • the advantage lies in the fact that the response time of the hydraulic control system can be longer, without penalizing the desired direction of variation of the compression ratio.
  • the section of the slide of each controlled non-return valve 72a or 72b on the side of the hydraulic chamber 35a or 35b of the hydraulic cylinders 36a and 36b is greater than the section wetted by the oil on the side of the pipe 75 of communication between the cylinders 36a and 36b. Therefore, the higher the hydraulic pressure in one of the hydraulic cylinders 36a or 36b, the greater the force that tends to close the controlled nonreturn valve 72a or 72b corresponding is high.
  • the controlled nonreturn valve 72a or 72b concerned remains open only if the opening hydraulic control pressure generated by the hydraulic control circuit 81 is sufficient.
  • the pilot pressure generated by the hydraulic control circuit 81 is a parameter for regulating the speed of variation of the compression ratio.
  • the amplitude of rotation of the eccentric 18 by motor cycle, when the controlled nonreturn valve 72b is controlled corresponds to an increase in the compression ratio according to an increasing function of the hydraulic control pressure generated by the hydraulic control circuit 81.
  • analogous to piloting the controlled non-return valve 72a corresponds to a reduction in the compression ratio at each engine cycle according to an increasing function of the hydraulic control pressure generated by the hydraulic control circuit 81.
  • the solenoid valve 79 pilot none piloted check valves 72a or 72b, oil transfers between the two hydraulic cylinders 36a and 36b are blocked and the compression ratio can not vary.
  • the choke 78 is a calibration component by construction of the rate of variation of the compression ratio.
  • the feeding circuit via the hydraulic lines 41 and 77 and the non-return valve 76 makes it possible to fill the oil circuit in the initial phase and then to compensate for any hydraulic leaks in the system.
  • the role of the non-return valve 76 is to prevent any return of oil from the hydraulic power circuit to the engine lubrication circuit.
  • the figure 11 presents an alternative embodiment of the invention.
  • the fluid bearing is made between the crankpin 22 of the crankshaft and the bore 20 of the eccentric 18.
  • the coefficient of friction of this rotational connection is less than five thousandths.
  • the coefficient of friction between the bore 19 of the big end 17 and the eccentric 18 is greater than thirty five hundredths.
  • This physical characteristic is obtained thanks to a coating of nickel-titanium type deposited under vacuum on the outer diameter of the eccentric 18 and in the bore 19 of the conrod head 17. Moreover, this coating gives a long life treated parts.
  • the variation range S of the piston 14 at top dead center is five millimeters and the diameter of the crankpin 22 is fifty millimeters.
  • the deflection angle of the eccentric 18 is thirty degrees above and thirty degrees below the reference line YY.
  • the means for controlling the position of the eccentric comprise a flange 21 integral with the eccentric 18 and parts integrated in the blank 33 of the crankshaft 28 composed of a rocker 62 kinematically connected to an actuating device and a skid 60 when in contact with the outer diameter 50 of the flange 21.
  • the flip-flop 62 pivots clockwise about its axis 61 integral with the blank 33 of the crankshaft 28 and plate 60 pad is on the collar 21.
  • the pad 60 is retained by the hinge 68 integral with the latch 62 and can not rotate with the eccentric.
  • the compound set of the rocker 62 articulated on the axis 61 actuated in rotation clockwise by the actuating device and the pad 60 remains connected to the blank 33 of the crankshaft 28 and induces a direct drive torque between the crankshaft and the eccentric .
  • this torque is greater than the torque generated by the resultant forces exerted between the connecting rod head 17 and the eccentric 18, the eccentric 18 is accelerated in the direction of rotation of the crankshaft 28.
  • the crankshaft rotates in the direction counterclockwise.
  • the compression ratio increases if the axis X1X1 of the eccentric 18 is placed, with reference to the figure 11 , to the right of the axis X0X0 of its bore and decreases in the opposite case.
  • One way of carrying out the invention in an electric version, according to the variant presented in the preceding paragraph, is to equip the actuating device of the pad 60, via the flip-flop 62, with two piezoelectric actuators 64a, 64b.
  • These two actuators 64a, 64b are on two separate axes, parallel and at the same distance from the axis 61 of the latch 62. They act in opposite directions to pivot the latch 62, via the pushers 63a, 63b. They are plated with the same geometric base towards the rocker 62 by elastic washers 65 which pushes a pusher 66 guided so that the distance differential of the base of the piezoelectric actuators 64a, 64b resting on the pusher 66 does not change.
  • the stroke of the spring washers 65 is more than ten times greater than the stroke of the actuators.
  • the rotation of the rocker 62 is always a function of the elongation or retraction differential between the two actuators 64a, 64b.
  • the two actuators are always controlled simultaneously and in opposition of voltage so that the temperature differences between the two piezoelectric actuators 64a, 64b remains low.
  • the tightness of the two piezoelectric actuators 64a, 64b is ensured by the shutter 67 and by seals, not shown, mounted on the pushers 63a, 63b.
  • the actuators are electrically connected to the control means via electrical wires and rotating joints (not shown).
  • Piezoelectric actuators have the advantage of offering extremely fast response times. This construction is therefore compatible with internal combustion engines whose rotation speed is high. In addition, the functions of the piezoelectric actuators are reversible. Also, the outer diameter 50 of the collar 21 secured to the eccentric 18 to a concentricity defect of three hundredths of a millimeter relative to the bore 20 of the eccentric 18 and a defect of cylindricity less than one hundredth of a millimeter. The relative rotation between the eccentric 18 and the crankshaft 28 thus generates stress variations on the two piezoelectric actuators 64a, 64b that these transform into electrical signals. Said force variations are correlated with the angular position of the eccentric with respect to the crankshaft.
  • the control means in combination with the knowledge of the angular position of the crankshaft, deduces the value of the compression ratio.
  • the piezoelectric actuators thus have the function of actuator for driving in rotation the eccentric with the crankshaft and measuring sensor which allows the control means to know the compression ratio at each revolution of the internal combustion engine.
  • Another way to realize the invention in electric version consists in equipping the actuating device of the shoe 60, via the flip-flop 62, with a single electric actuator 90 with a large stroke, so that the expansion differentials and the wears are easily compensated.
  • the present invention can be applied to any reciprocating piston machine (s) and more particularly to internal combustion engines in order to reduce pollutant emissions as well as fuel consumption.

Landscapes

  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Output Control And Ontrol Of Special Type Engine (AREA)
  • Shafts, Cranks, Connecting Bars, And Related Bearings (AREA)
  • Road Paving Machines (AREA)

Claims (10)

  1. Vorrichtung zur Anpassung des Verdichtungsverhältnisses einer Brennkraftmaschine versehen mit wenigstens einem Zylinder (31) mit einer Brennkammer (10), einem Bewegungselement mit einem Kolben (14), der unter Einwirkung einer Pleuelstange (15) translatorisch verschiebbar ist und über eine Achse (13) mit dem Kolben (14) verbunden ist sowie mit dem Zapfen einer Kurbelwelle (28) verbunden ist, wobei der Kolben (14) sich zwischen dem oberen und dem unteren Totpunkt eines Verdichtungstaktes bewegt und dabei am oberen Totpunkt des Kolbenhubs (14) ein Totvolumen aufweist, wobei die Vorrichtung zwischen dem Kopf der Pleuelstange (17) und dem Zapfen (22) der Kurbelwelle (28) ein drehbares Exzenterelement (18) zur Anpassung des Verdichtungsverhältnisses aufweist, wobei die Vorrichtung zudem Mittel zur Bewegungssteuerung des Exzenterelements aufweist, dadurch gekennzeichnet, dass die Steuerungsmittel wenigstens eine Bewegungskupplung (30) ohne Schraube aufweisen zwischen einem radialen Einzelvorsprung des Exzenterelements (18) und einem Mechanismus zur Lageregelung (29) im Verhältnis zur Kurbelwelle (28), der wenigstens einen Linearantrieb aufweist, wobei der Mechanismus in eine der Kurbelwellenscheiben (33) integriert ist, wobei die Bewegungskupplung (30) und der Mechanismus zur Lageregelung (29) im Verhältnis zur Kurbelwelle (28) vollständig oder hauptsächlich außerhalb des Zapfens (22), des Lagers (27) und des Hebels (23), der Zapfen (22) und Lager (27) der Kurbelwelle (28) verbindet, angeordnet sind.
  2. Vorrichtung nach Anspruch 1 dadurch gekennzeichnet, dass der Mechanismus zur Lageregelung (29) im Verhältnis zur Kurbelwelle (28) einen Raum einnimmt, dessen entferntester Punkt zur Achse XX der Kurbelwelle im Lauf der Motordrehung einen Kreis (25) beschreibt, dessen Durchmesser gleich oder kleiner als der Durchmesser des größten Kreises (24bi) ist, den der Kopf der Pleuelstange beschreibt, oder als der größte Kreis (24vi), den die Scheiben (33) der Kurbelwelle beschreiben.
  3. Vorrichtung nach Anspruch 1 oder 2 dadurch gekennzeichnet, dass die Mittel zur Bewegungssteuerung des Exzenterelements laut Erfindung mit zwei Einheiten auf jeweils einer Seite des Exzenterelements (18) ausgestattet ist, die jeweils mit wenigstens einer Bewegungskupplung (30) ohne Schraube an einen radialen Einzelvorsprung des Exzenterelements (18) verbunden sind und über jeweils einen Mechanismus zur Lageregelung (29) im Verhältnis zur Kurbelwelle (28) verfügen.
  4. Vorrichtung nach Anspruch 1 oder 2 gekennzeichnet dadurch, dass der Mechanismus zur Lageregelung (29) des Exzenterelements (18) im Verhältnis zur Kurbelwelle (28) zwei Linearantriebe umfasst.
  5. Vorrichtung nach Anspruch 4 gekennzeichnet dadurch, dass die Achsen (37a), (37b) der zwei Linearantriebe verschieden sind.
  6. Vorrichtung nach Anspruch 4 gekennzeichnet dadurch, dass der Mechanismus zur Lageregelung (29) im Verhältnis zur Kurbelwelle (28) zwei einfach wirkende Hydraulik-Zylinder (36a) und (36b) umfasst, die beiderseits des Zapfens (22) und des Lagers (27) der Kurbelwelle (28) angebracht sind.
  7. Vorrichtung nach Anspruch 4 gekennzeichnet dadurch, dass die zwei Linearantriebe zwei Hydraulik-Zylinder (36a) und (36b) sind, die in ein Modul (80) integriert sind, wobei das Modul (80) auf die Kurbelwelle (28) montiert ist.
  8. Vorrichtung nach Anspruch 1 oder 2 gekennzeichnet dadurch, dass die Mittel zur Bewegungssteuerung des Exzenterelements einen Sattel (21) des Exzenterelements umfassen sowie die in die Scheibe (33) der Kurbelwelle (28) integrierten Teile, die insbesondere aus einer Betätigungsvorrichtung und einem Kontaktpuffer (60) mit einem dem Sattel (21) entsprechenden Außendurchmesser (50) zusammengesetzt sind, wobei der Puffer sich nicht mit dem Exzenterelement (18) drehen kann.
  9. Vorrichtung laut Anspruch 8 gekennzeichnet dadurch, dass die Mittel zur Bewegungssteuerung des Exzenterelements (18) wenigstens einen piezoelektrischen Antrieb (64a), (64b) einsetzen.
  10. Vorrichtung nach einem der Ansprüche 8 und 9 mit einem zwischen dem Zapfen (22) der Kurbelwelle und der Nabe (20) des Exzenterelements (18) eingesetzte Fluidlager, das dadurch gekennzeichnet ist, dass der Puffer (60) unter Einwirkung der Betätigungsvorrichtung ein Antriebsmoment zwischen Kurbelwelle und Exzenterelement induzieren kann, um das Exzenterelement (18) in Drehrichtung der Kurbelwelle (28) zu beschleunigen.
EP06709379A 2005-02-28 2006-02-27 Sehr kompakte vorrichtung zum verstellen des verdichtungsverhältnisses eines brennkraftmotors Not-in-force EP1859135B1 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
FR0501983A FR2882575A1 (fr) 2005-02-28 2005-02-28 Dispositif tres compact pour ajuster le taux de compression d'un moteur a combustion interne
PCT/FR2006/000430 WO2006092484A1 (fr) 2005-02-28 2006-02-27 Dispositif tres compacte pour ajuster le taux de compression d’un moteur a combustion interne

Publications (2)

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EP1859135A1 EP1859135A1 (de) 2007-11-28
EP1859135B1 true EP1859135B1 (de) 2009-04-22

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US (1) US20080184966A1 (de)
EP (1) EP1859135B1 (de)
AT (1) ATE429572T1 (de)
DE (1) DE602006006422D1 (de)
FR (1) FR2882575A1 (de)
WO (1) WO2006092484A1 (de)

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US9341110B2 (en) * 2008-07-16 2016-05-17 Wilkins Ip, Llc Internal combustion engine with improved fuel efficiency and/or power output
FR2940362A1 (fr) * 2008-12-22 2010-06-25 Faar Industry Dispositif d'ajustement et procede d'ajustement pour moteur a taux de compression variable.
US8468997B2 (en) 2009-08-06 2013-06-25 Larry C. Wilkins Internal combustion engine with variable effective length connecting rod
US8746188B2 (en) * 2010-03-17 2014-06-10 Larry C. Wilkins Internal combustion engine with hydraulically-affected stroke
ITRM20100155A1 (it) * 2010-04-02 2011-10-03 Matteo Nargiso Motore a combustione interna con albero motore ad eccentricita' variabile
US8851030B2 (en) 2012-03-23 2014-10-07 Michael von Mayenburg Combustion engine with stepwise variable compression ratio (SVCR)
JP2015124635A (ja) 2013-12-25 2015-07-06 三菱自動車工業株式会社 内燃機関の可変圧縮比装置
JP6183610B2 (ja) * 2013-12-25 2017-08-23 三菱自動車工業株式会社 内燃機関の可変圧縮比装置
JP2015124636A (ja) * 2013-12-25 2015-07-06 三菱自動車工業株式会社 内燃機関の可変圧縮比装置
JP6269043B2 (ja) * 2013-12-25 2018-01-31 三菱自動車工業株式会社 内燃機関の可変圧縮比装置
FR3040436B1 (fr) * 2015-08-26 2019-08-02 Psa Automobiles Sa. Ensemble pour moteur a combustion interne comportant un systeme de variation du rapport volumetrique
FR3040437B1 (fr) * 2015-08-26 2019-06-07 Psa Automobiles Sa. Ensemble pour moteur a combustion interne comprenant un systeme de variation de rapport volumetrique
CN108590849B (zh) * 2018-01-09 2023-07-14 西华大学 一种可实现米勒循环的曲柄连杆机构及控制方法
DE102018104292A1 (de) * 2018-02-26 2019-08-29 Avl List Gmbh Sensoreinrichtung für eine längenverstellbare Pleuelstange
FR3081525B1 (fr) * 2018-05-25 2020-05-08 MCE 5 Development Vilebrequin pour un moteur a rapport volumetrique variable pilote
FR3085431B1 (fr) * 2018-08-30 2020-12-04 MCE 5 Development Moteur a rapport volumetrique pilote
CN110375683B (zh) * 2019-07-11 2021-01-12 浙江义利汽车零部件有限公司 一种测量曲轴轴向间隙的方法、系统及车辆
DE102019123601A1 (de) * 2019-09-04 2021-03-04 Bayerische Motoren Werke Aktiengesellschaft Hubkolben-Brennkraftmaschine mit einem variablen Verdichtungsverhältnis
AT524321B1 (de) 2021-03-12 2022-05-15 Roland Kirchberger Dipl Ing Dr Techn Verbrennungskraftmaschine

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US4406256A (en) 1981-05-22 1983-09-27 The United States Of America As Represented By The Administrator Of The National Aeronautics And Space Administration Automatic compression adjusting mechanism for internal combustion engines
PL144411B1 (en) * 1984-11-23 1988-05-31 Politechnika Warszawska Crank mechanism with variable crank radius for a piston-type internal combustion engine
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Publication number Publication date
WO2006092484A1 (fr) 2006-09-08
US20080184966A1 (en) 2008-08-07
FR2882575A1 (fr) 2006-09-01
ATE429572T1 (de) 2009-05-15
DE602006006422D1 (de) 2009-06-04
EP1859135A1 (de) 2007-11-28

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