EP0689642B1 - Improvements to compression or spark ignition four-stroke internal combustion engines having a variable compression ratio enabling high supercharging pressure levels - Google Patents

Abstract

An engine including an assembly of at least two cylinders (2 and 3) with different displacements, and two crankshafts (4 and 5) coupled at the same rotational speed via a gear train (19, 20, 21) and a variably timed transmission having three concentric shafts (17, 28, 32) separable from the drive assembly (18) and designed to reduce the compression ratio as the intake pressure increases. The highest and lowest compression ratios are set within the angular displacement limits between the two crankshafts (4 and 5) by means of the compression ratios between the two cylinders (2 and 3), and between said two cylinders (2 and 3) and the clearance space, so that (a) the start of the variably timed transmission stroke increases the translation of the piston (8) by units of angular displacement between the two crankshafts (4 and 5) at the end of compression phase, and (b) the end of the variably timed transmission stroke combines combustion gas expansion on the piston (8) at least from the maximum torque on the crank of the short-stroke crankshaft (5).

Description

The object of the present invention is the concept of an engine with variable volumetric ratio which consists in varying the volume of the combustion chamber as a function of the density and temperature of the intake air, of the speed of rotation. and the engine temperature, which allows the engine to be supercharged, by pressing a single or double supercharging pressure with intercooling.

WO 89/03476 discloses the compression ignition engine with variable volumetric ratio which represents, by way of example, a preferred embodiment of the invention, relating to a two-stroke cycle engine, with ignition by compression which comprises an axial grouping of two cylinders joined side by side, with a centralization of two crankshafts in the cylinder block, provided with a kinematic chain where the two crankshafts are at the same speed of rotation, the engine is sweeped by equicurrent (l (fresh gas and exhaust are admitted at opposite ends of the larger of the two cylinders). The engine is also provided with a coupler making it possible to synchronize cyclically the positions of the piston of the combustion prechamber in the end of compression phase so as to be able to vary the volumetric ratio according to the different engine speeds. However, these aforementioned means do not allow the two-stroke engine to operate in good conditions, because in the smaller of the two cylinders, in each exhaust cycle, the flue gases are only partially removed from the cylinder, resulting in a loss of efficiency due to the fact that the intake of the fresh air can only be done by equicurrent. According to another form of the invention, it is also specified in this patent that the four-stroke engine can also be provided with a kinematic chain where the small crankshaft is at half speed of the large crankshaft, which has the effect of invalidate the operation of the four-stroke engine, following a desynchronization of the piston of the small cylinder with respect to the phases of the engine, for example, the discharge of the burnt gases from the small cylinder to the large cylinder during the intake phase, and of the burnt gases sucked in by the small cylinder from the large cylinder, in the exhaust phase.

The present invention describes a new combination of a four-stroke variable volume combustion chamber engine. The engine comprises a kinematic chain where the shafts of the two crankshafts are coupled at the same speed by means of the variable-pitch transmission. The angular offset travel between the two crankshafts carried out between the start and the end of travel of the variable-pitch transmission is arranged by an appropriate ratio between the two displacements of the two grouped cylinders and between the volume of the latter and the dead space , which allows to modulate the volumetric ratio of the engine without desynchronization of the piston of the small cylinder compared to the phases of the engine.

Also from patent DE-A-3,616,234 there is known a device for varying the relative angular position of two shafts in drive connection, in particular a crankshaft and a camshaft housed in the housing. a machine, the centerpiece of this device consists of the camshaft on which is mounted all the mechanical parts in rotational movement, as well as one of these parts with the alternative of associated movements of translation and rotation determining the offsets of drive links between the camshaft and the crankshaft. The associated movements of said translation and rotation part are effected by a piston and hydraulic double action cylinder separate from the camshaft. It should be noted that the axial resistance force generated by the displacement of the piston at the change in angular position of the device is proportional by the resistive torque of the camshaft (resistance due to friction of the cams on the valve stems), which has the effect of composing axial forces on the camshaft, the latter being immobilized axially by means of a bearing fitted with a bearing stops. The camshaft generates only weak resistance couples of the order of 1 to 3% compared to the crankshaft torque, this results in weak axial forces on the camshaft, which allows from the point of view mechanical, to use a thrust bearing with low carrying capacity and therefore to be able to have a thrust bearing sufficiently small to make use of it on the camshaft. With regard to the dismantling of the device for varying the angular position outside the engine block, this successively involves the dismantling of a multitude of mechanical parts forming part of this device and supported by the camshaft, the latter counting also as part of this device.

The present invention also describes a novel combination of a variable-pitch transmission suitable for regulating the four-stroke variable volume combustion chamber engine. This new combination of the variable-pitch transmission in coupling connection with the two shafts of the two crankshafts of the engine has the advantage of varying the angular position between these two shafts without any axial force on them, whatever or the torque force on the variable-pitch transmission. According to this new arrangement, the variable-pitch transmission can be removed from the engine block and uncoupled from the two crankshafts as an interchangeable mechanical assembly that can be reassembled on the engine block. Means are also provided on this variable-pitch transmission to have precise angular adjustment between the coupling connection of the two crankshafts and the variable-timing transmission.

By definition, the principle of supercharging piston engines is to increase the air masses without increasing the displacement. The result for engines with a fixed compression ratio is an increase in combustion pressure and greater volumetric power (power per liter of displacement). However, when the boost pressure is increased, the mechanical and thermal stresses increase on the engine components. This major drawback stems from the fact that the volumetric ratio, generated by the combustion chamber and the piston stroke, cannot be modified, being unable to adapt to variations in pressures and temperatures of the intake air and in speeds and temperatures of the motor.

Engineers therefore comply with certain design rules by determining, on the one hand, a limit on the amplitude of the pressure variations at the intake, and on the other hand, by achieving an average compression ratio between the pressure d aspiration atmospheric and boost pressure. As the determination of the average compression ratio is a compromise reconciling at best the different engine speeds, the atmospheric suction regime is located at too low pressures and temperatures, and the boost pressure regime is located at too high pressures and temperatures.

According to the invention, this new engine comprises two lines of crankshafts, one with a long stroke crank, the other with a short stroke crank. The two crankshafts are coupled at the same rotational speed by means of a gear train and a variable-pitch transmission, the coupling pinion of which forms part of the gear train moves angularly relative to the crankshaft. at short stroke, which allows an infinite number of stalls between the two crankshafts without requiring the interruption of the transmission between them.

According to the invention, the variable pitch transmission is designed in such a way that it can be separated from the engine block independently of the crankshaft at short stroke, which has the advantage of being able to quickly and easily replace the defective parts or to a standard exchange of the latter. The cylinders, differentiated by their displacement, are each arranged above one of the two lines of crankshafts. The crankshaft of the short stroke crankshaft operating with the connecting rod of the piston of the smallest cylinder, the crank of the crankshaft of long stroke operating with the connecting rod of the piston of the largest cylinder. The two cylinders are connected one by one, from one row to the other, by a recess in the cylinder head, so as to form a group of two cylinders communicating with each other in order to allow the gases to pass from one to the 'other, regardless of the position of the piston of each of the cylinders.

According to the invention, in the compression ignition version, the engine comprises at least one fuel injector in the dead space, the fuel injection is carried out at half speed with the crankshaft at large stroke.

According to the invention, in a spark-ignition version, the engine comprises at least one spark plug in the dead space, the ignition is carried out by known means in half-speed synchronism with the long-stroke crankshaft.

In accordance with the present invention, the distribution is ensured at least by a camshaft engaged at half speed with the long-stroke crankshaft, putting the group of two cylinders in periodic communication with the intake and exhaust pipes at through the intake and exhaust valves at specific times in the four-stroke cycle. The expansion phase is carried out simultaneously on each piston of the two grouped cylinders making the two crankshafts cooperate with the engine force. The long-stroke crankshaft is connected directly to the external transmission components of the engine, so that the variable-pitch transmission transmits only the engine torque of the short-stroke crankshaft to the long-stroke crankshaft, the engine force on the variable-pitch transmission is therefore dependent on the smaller displacement of the two grouped cylinders.

The different angular offsets of the variable-pitch transmission between the two crankshafts have the effect of modifying, in the end of compression phase (top dead center of the piston of the largest displacement), an additional space generated in the smallest displacement. This additional space being defined with the dead space, so as to modify the volumetric ratio of the engine in the maximum direction at the start of travel of the variable-timing transmission, and in the minimum direction at the end of travel of the variable-timing transmission .

According to the present invention, a hydraulic force amplifier whose slave cylinder acts on the variable-pitch transmission, modifies the additional volume of the small displacement in proportion to the boost pressure, so as to maintain the engine in optimal operating conditions with the minimum of pollution.

Also according to the invention, a preset program on a pre-production engine eliminates the excessive stresses of pressures and temperatures. Each engine speed is stored in a point progression scale, so as to encompass all of the engine's capabilities. Each storage point is a combination formed by the measurements of four sensors: the pressure of the intake air. intake air temperature, engine speed, and engine temperature. Each combination is recorded simultaneously with the position of the variable timing transmission control cylinder. This program allows the automatic piloting of the series engine identical to that of the engine produced on the test bench. Fuel specifications must also be identical to reproduce exactly the same operating conditions on the series engine, thanks to high-frequency monitoring of the measurements of the four sensors.

• la figure 1 est une vue en coupe longitudinale partielle d'un moteur à quatre temps, à chambre de combustion à volume variable avec un rapport de 5 entre les deux cylindres groupés, représenté en position de début de course de la transmission à calage variable en phase fin de compression. On peut voir les cannelures hélicoïdales appariées entre les premier et troisième éléments concentriques qui sont à hélices circulaires contraires de celles des cannelures hélicoïdales appariées entre les deuxième et troisième éléments concentriques ;
• la figure 2 représente une vue en éclaté du moteur de la figure 1 montrant la transmission à calage variable démontée des deux vilebrequins ;
• la figure 3 représente le moteur de la figure 1, suivant une variante de l'invention, montrant en détail les cannelures droites appariées entre les premier et troisième éléments concentriques et les cannelures hélicoïdales appariées entre les deuxième et troisième éléments concentriques :
• la figure 4 est une vue schématique en coupe transversale d'un moteur à quatre temps suivant l'invention. à chambre de combustion à volume variable avec un rapport de 5 entre les cylindres groupés, représenté en position de fin de compression en début de course de la transmission à calage variable avec 36° d'avance angulaire de la manivelle du vilebrequin à petite course par rapport à la manivelle du vilebrequin à grande course ;
• la figure 5 est une vue schématique en coupe transversale du même moteur que celui de la figure 4, représenté en position de fin de compression en fin de course de la transmission à calage variable avec 69° d'avance angulaire de la manivelle du vilebrequin à petite course par rapport à la manivelle du vilebrequin à grande course :
• la figure 6 est une vue en plan du fond de culasse des deux cylindres groupés du même moteur que celui représenté aux figures 4 et 5:
• la figure 7 est une vue schématique en coupe transversale d'un moteur à quatre temps suivant l'invention , à chambre de combustion à volume variable avec un rapport de 2.5 entre les deux cylindres groupés, représenté en position de fin de compression en début de course de la transmission à calage variable avec 30° d'avance angulaire de la manivelle du vilebrequin à petite course par rapport à la manivelle du vilebrequin à grande course ;
• la figure 8 est une vue schématique en coupe transversale du même moteur que celui de la figure 7, représenté en position de fin de compression en fin de course de la transmission à calage variable avec 70° d'avance angulaire de la manivelle du vilebrequin à petite course par rapport à la manivelle du vilebrequin à grande course ;
• la figure 9 est une vue en plan du fond de culasse des deux cylindres groupés du même moteur que celui représenté aux figures 7 et 8 ;
• la figure 10 représente les diagrammes superposés d'un moteur à rapport volumétrique de 5 entre les deux cylindrées des deux cylindres groupés, montrant les rapports volumétriques par degré de rotation angulaire du vilebrequin à grande course (5) dans les phases de compression et de détente sans allumage. en début ou en fin de course de la transmission à calage variable avec le rendement volumétrique correspondant.
• la figure 11 représente les diagrammes superposés d'un moteur à rapport volumétrique de 2,5 entre les deux cylindrées des deux cylindres groupés. montrant les rapports volumétriques par degré de rotation angulaire du vilebrequin à grande course (5) dans les phases de compression et de détente sans allumage, en début ou en fin de course de la transmission à calage variable avec le rendement volumétrique correspondant.

The invention will be described in more detail with the aid of the following description and with reference to the accompanying drawings by way of non-limiting example and in which:

• FIG. 1 is a view in partial longitudinal section of a four-stroke engine with a variable-volume combustion chamber with a ratio of 5 between the two grouped cylinders, shown in the start-of-travel position of the variable-pitch transmission in end of compression phase. We can see the helical grooves paired between the first and third concentric elements which are circular helices opposite to those of the helical grooves paired between the second and third concentric elements;
• FIG. 2 represents an exploded view of the engine of FIG. 1 showing the disassembled variable timing transmission of the two crankshafts;
• FIG. 3 represents the motor of FIG. 1, according to a variant of the invention, showing in detail the straight grooves paired between the first and third concentric elements and the helical grooves paired between the second and third concentric elements:
• Figure 4 is a schematic cross-sectional view of a four-stroke engine according to the invention. with variable volume combustion chamber with a ratio of 5 between the grouped cylinders, represented in the end of compression position at the start of the travel of the variable-pitch transmission with 36 ° angular advance of the crankshaft at short travel by compared to the crankshaft with long stroke;
• Figure 5 is a schematic cross-sectional view of the same engine as that of Figure 4, shown in the end of compression position at the end of travel of the variable-pitch transmission with 69 ° angular advance of the crankshaft crankshaft small stroke compared to the crankshaft crank:
• FIG. 6 is a plan view of the cylinder head bottom of the two grouped cylinders of the same engine as that shown in FIGS. 4 and 5:
• FIG. 7 is a schematic cross-sectional view of a four-stroke engine according to the invention, with a combustion chamber with variable volume with a ratio of 2.5 between the two grouped cylinders, represented in the end of compression position at the start of stroke of the variable-pitch transmission with 30 ° angular advance of the crankshaft at short stroke relative to the crankshaft at long stroke;
• Figure 8 is a schematic cross-sectional view of the same engine as that of Figure 7, shown in the end of compression position at the end of the travel of the variable-pitch transmission with 70 ° angular advance of the crankshaft crankshaft small stroke relative to the crankshaft crank with large stroke;
• Figure 9 is a plan view of the cylinder head bottom of the two grouped cylinders of the same engine as that shown in Figures 7 and 8;
• FIG. 10 represents the superimposed diagrams of a volumetric ratio engine of 5 between the two displacements of the two grouped cylinders, showing the volumetric ratios by degree of angular rotation of the long-stroke crankshaft (5) in the compression and expansion phases without ignition. at the start or end of travel of the variable-pitch transmission with the corresponding volumetric efficiency.
• FIG. 11 represents the superimposed diagrams of an engine with a volumetric ratio of 2.5 between the two displacements of the two grouped cylinders. showing the volumetric ratios by degree of angular rotation of the long-stroke crankshaft (5) in the compression and expansion phases without ignition, at the start or at the end of the travel of the variable-pitch transmission with the corresponding volumetric efficiency.

Referring to Figures 1 to 9, the cylinder block (1) comprises two crankshafts (4 and 5) arranged in parallel, one with a long stroke crank (4), the other with a short stroke crank (5 ), the two cylinders (2 and 3) provided with pistons respectively (6 and 8) and connecting rods respectively (7 and 9) are arranged each above of the two lines of crankshafts (4 and 5). The crankshaft of the short stroke (5) operating with the connecting rod (9) of the piston (8) of the smallest cylinder (3), the crank of the long stroke crankshaft (4) operating with the connecting rod (7) of the piston ( 6) of the largest cylinder (2). The two cylinders (2 and 3) are connected one by one, from one row to another, by a recess in the cylinder head (10), so as to form a group of two cylinders (2 and 3) communicating with each other .

In the compression ignition version, the engine includes at least one fuel injector (not shown) in the dead space. The fuel is injected by known means (not shown) in half-speed engagement with the crankshaft with long stroke crank (4).

In the spark ignition version, the engine comprises at least one spark plug (not shown) in the dead space. The ignition is carried out by known means (not shown) in half-speed synchronism with the long-stroke crankshaft (4).

Distribution is ensured at least by a camshaft (not shown) engaged at half-speed with the long-stroke crankshaft (4). The part of the cylinder head (10) overhanging the largest cylinder (2) comprises the intake and exhaust valves respectively (13 and 14), putting the group of the two cylinders (2 and 3) in periodic communication with the pipes. intake and exhaust respectively (11 and 12) at specific times in the four-stroke cycle.

For very large displacement engines, a second camshaft (not shown) engaged at half speed with the long stroke crankshaft (4) can be provided in the part of the cylinder head (10) overhanging the smallest cylinder ( 3), so as to ensure second periodic opening and closing of the intake and exhaust at the same time as the opening and closing of the four-stroke cycle carried out in the largest cylinder (2). The ratio between the displacements of the two grouped cylinders (2 and 3) is at least between 2.5 and 5 allowing the engine to be adapted to boost pressure rates from 1 to 7.

The variable setting transmission is formed by three superposed concentric elements: the first element consists of the transmission shaft (17) located in the internal part, the second element consists of the sleeve (28) of the gear (20 ) located in the external part and the third element is constituted by the sliding tube (32) located in the intermediate part between the two other aforementioned elements. Said sleeve (28) is held in an applied bearing (15) by means of a two-row angular contact bearing (16) suitable between the applied bearing (15) and the sleeve (28). Said applied bearing (15) is fixed to the engine block (1) so that the variable-pitch transmission can constitute a separate assembly from the shaft (18) of the short-stroke crankshaft (5). To this end, the variable setting transmission and the short stroke crankshaft (5) are each made with their respective shaft (17 and 18). The contiguous ends between the shaft (17) of the variable-pitch transmission and the shaft (18) of the short-stroke crankshaft (5) are shaped with corresponding straight male and female splines allowing their coupling in the engine block (1 ) by axial sliding when the applied bearing (15) is applied in an orifice provided in the engine block (1). The applied bearing (15) is centered on the shaft (18) of the short-stroke crankshaft (5), so as to allow the self-centering of the shaft (17) on said shaft (18), the latter also serving as free bearing on the shaft (17) when applying the applied bearing (15) on the engine block (1); this means allowing the disassembly of the variable setting transmission outside the engine block (1) without having to disassemble the crankshaft at short stroke (5).

The transmission shaft (17) and the sleeve (28) are advantageously held concentrically and axially with respect to each other by means of a bearing (22) integral with the shaft (17). The bearing (22) is provided with a bearing (23) with axial and radial abutment allowing the free rotation of the shaft (17) independently of the sleeve (28). The bearing (22) is an integral part of the shaft (17) at the point where the straight grooves of the mating coupling ends between the shaft (17) and the shaft (18) of the short stroke crankshaft are limited. (5). The bearing (22) and the sleeve (28) are located inside the engine block (1). The bearing (22) is made in the form of a disc also acting as a flywheel, the periphery of this flywheel is regularly pierced with holes (24) allowing the bolting of a ring (25) located on the side face opposite the side where the straight grooves are limited. The application of the ring (25) on the flywheel of the bearing (22) is used to form a housing allowing the fixing of the outer ring (26) of the bearing (23) with axial and radial forces, while the ring inner (27) of the bearing (23) is fixed on the sleeve (28) against a spacer (29) in the form of a ring surrounding the sleeve (28), the spacer (29) is intended to make up for the separation space between the inner ring (27) of the bearing (23) and the inner ring of the angular contact bearing (16), the latter being held axially against a shoulder provided on the sleeve (28) by fixing all the aforementioned parts by means of a single nut (30) on the sleeve (28).

The gear (20) of the sleeve (28) is located outside the engine block (1) coupled at the same speed with the long-stroke crankshaft (4) by means of a gear (19) integral with this last and an intermediate gear (21) between the two aforementioned gears (19 and 20).

The transmission shaft (17) comprises, on the side of the bearing (22) facing the applied bearing (15), helical grooves (31) on which the sliding tube (32) is fitted. This sliding tube (32) has on its internal periphery grooves (33) matched to the helical grooves (31), so that the sliding tube (32) can slide helically on the drive shaft (17) and allow angular offset between the said first and third elements.

The sliding tube (32) also has on its outer periphery helical grooves (34) whose helix is in the opposite direction to that of the grooves (33) produced inside the sliding tube (32). The sleeve (28) has on its internal periphery helical grooves (35) matched to the external helical grooves (34) of the sliding tube (32), so that the latter can slide helically in the sleeve (28) and allow angular offset between the said second and third elements at the same time as the helical sliding between the first and third abovementioned elements, the sleeve (28) again becomes integral in rotation with the shaft (17) when the sliding tube (32) is not in axial translation.

The length of the sliding tube (32) is established inside the sleeve (28) when the end of said sliding tube (32) is at the stop limit defined by the obstruction of the bearing (22), l 'other end of the sliding tube (32) is released outside the sleeve (28) through the gear (20) out of the engine block (1) to allow, by appropriate means, the fixing of the inner ring of the two-row angular contact bearing (36). Said inner ring of the bearing (36) is made integral with the rotation movement of the sliding tube (32), while the outer ring of the bearing (36), without rotational movement, is secured to the attachment piece (37 ).

A memory of the compression ratio program decision acting by a hydraulic control system allows the movement of the attachment piece (37) and the sliding tube (32) to modify the timing between the two crankshafts (4 and 5).

The start of travel of the variable-pitch transmission is arranged so that the sliding tube (32) is in the exit stop position (not shown) of the sleeve (28) (low torque) which corresponds to the minimum advance angle of the crankshaft with short stroke (5) relative to the crankshaft with long stroke (4).

The limit switch of the variable-pitch transmission is arranged such that the sliding tube (32) is in the re-entry stop position (not shown) of the sleeve (28) (high torque) corresponding to the maximum angular advance of the short stroke crankshaft (5) relative to the long stroke crankshaft (4).

According to the invention, to clarify and facilitate the timing of the two crankshafts (4 and 5) between the variable-timing transmission, the teeth of the gear (20) are in even number, when the paired grooves (34 and 35) respectively of the sliding tube (32) and of the sleeve (28), the paired grooves (31 and 33) of the shaft (17) and of the sliding tube (32) respectively, as well as the splines joining between the two shafts (17 and 18 ) are each in odd number and vice versa.

According to a variant of the invention, the shaft (17) of the variable-pitch transmission comprises, on the side of the bearing (22) facing the applied bearing (15), straight grooves (38) in substitution for the helical grooves ( 31) on which the sliding tube (32) is fitted, which comprises on its internal periphery straight grooves (39) in substitution for helical grooves (33), the straight grooves (39) being matched with the straight grooves (38 ) of the tree (17).

According to the invention, the minimum and maximum volumetric ratios selected for the type of engine to be designed are produced as a function of the dimensions of the different elements of the engine, namely on the one hand, the ratio between the displacement of the two grouped cylinders (2 and 3 ) and on the other hand, the ratio formed by the total volume of the two displacements of these cylinders (2,3) with the volume formed by the dead space (40), the latter ratios are arranged in such a way that the maximum angular advance of the crankshaft of the short stroke (5) relative to the crank of the long stroke (4), defined by the end position of the transmission to variable setting, match, at the end of compression phase (top dead center of the piston 6), the positioning of the piston (8) in relation to the additional space necessary for the dead space (40) to define said minimum volumetric ratio of the engine with an angle of at least 90 ° between the connecting rod (9) and the crankshaft of the short stroke crankshaft (5).

• en phase de détente, avec les gaz de combustion sur le piston (8) associés au moins à partir du couple maximun instantané sur la manivelle du vilebrequin à petite course (5) ;
• en phase de détente, en limitant la remontée du piston (8) antérieurement à l'ouverture de la soupape d'échappement (14) source de contre-pressions des gaz de combustion sur le dit piston (8);
• en phase fin d'admission, en limitant la remontée du piston (8) source de diminution du volume de remplissage dans le cylindre (3).

The aforementioned angular adjustment arrangements between the two crankshafts in the end-of-travel position of the variable-pitch transmission, in relation to the appropriate dimensions between the various elements of the engine, allow the latter to operate:

• in the expansion phase, with the combustion gases on the piston (8) associated at least from the instantaneous maximum torque on the crank of the short-stroke crankshaft (5);
• in the expansion phase, by limiting the rise of the piston (8) prior to the opening of the exhaust valve (14) which is a source of combustion gas back pressures on said piston (8);
• at the end of intake phase, limiting the rise of the piston (8), a source of reduction in the filling volume in the cylinder (3).

These operations have the advantage of ensuring that optimum engine performance is maintained at full load speed.

The maximum volumetric ratio selected is produced on the same database as the dimensional values defined for the minimum volumetric ratio, in such a way that the minimum angular advance of the crankshaft at short stroke (5) relative to the crank of the long-stroke crankshaft (4), defined by the start-of-travel position of the variable-pitch transmission, at the end of the compression phase (top dead center of the piston 6), the positioning on the piston (8) with the additional space necessary for the dead space (40) to define the maximum volumetric ratio of the engine with the connecting rod (9) of the crankshaft of the short stroke (5) spaced from its top dead center, so that the said connecting rod ( 9) forms an angle with the crankshaft of the short stroke (5).

• en phase fin de compression, en assurant un mouvement de translation plus important sur le piston (8) par degré unitaire de décalage angulaire entre les manivelles des deux vilebrequins (4 et 5).

The aforementioned angular adjustment provisions between the two crankshafts in the start position of the variable-pitch transmission in relation to the appropriate dimensions between the various elements of the engine allow the latter to operate:

• in the end of compression phase, ensuring a greater translational movement on the piston (8) per unit degree of angular offset between the cranks of the two crankshafts (4 and 5).

This operation has the advantage of speeding up the process of modifying the volumetric ratio of the engine at low load.

=  rapport volumétrique entre les deux cylindrées des deux cylindres groupés.
α =  avance angulaire de la manivelle du vilebrebrequin à petite course.
ve =  volume de l'espace mort des deux cylindres groupés nécessaire pour le transfert des gaz sans laminage excessif.
(α minimum) =  avance angulaire de la manivelle du vilebrequin à petite course. en début de course de la transmission à calage variable.
(α maximum) =  avance angulaire de la manivelle du vilebrequin à petite course. en fin de course de la transmission à calage variable.
V a (α minimum) =  volume additionnel s'ajoutant au volume de l'espace mort. en début de course de la transmission à calage variable, défini par l'angle minimum de l'avance angulaire de la manivelle du vilebrequin à petite course lorsque la manivelle du vilebrequin à grande course se situe au point mort haut, en phase fin compression.
V a (α maximum) =  volume additionnel s'ajoutant au volume de l'espace mort. en fin de course de la transmission à calage variable. défini par l'angle maximum de l'avance angulaire de la manivelle du vilebrequin à petite course lorsque la manivelle du vilebrequin à grande course se situe au point mort haut, en phase fin de compression.
Vr (α minimum) =  volume de refoulement d'air en début de course de la transmission à calage variable. défini par l'angle minimum de l'avance angulaire de la manivelle du vilebrequin à petite course lorsque la manivelle du vilebrequin à grande course se situe au point mort bas. en phase fin d'admission.
Vr (α maximum) =  volume de refoulement d'air en fin de course de la transmission à calage variable. défini par l'angle maximum de l'avance angulaire de la manivelle du vilebrequin à petite course lorsque la manivelle du vilebrequin à grande course se situe au point mort bas. en phase fin d'admission.Nature of symbols adopted
P = volumetric ratio.
V1 = displacement of the larger of the two grouped cylinders.
V2 = displacement of the smaller of the two grouped cylinders.
$\frac{\mathit{\text{V1}}}{\mathit{\text{V2}}}$

= volumetric ratio between the two displacements of the two grouped cylinders.
α = angular advance of the crankshaft at short stroke.
ve = volume of the dead space of the two grouped cylinders necessary for the transfer of gases without excessive rolling.
(α min imum) = angular advance of the crankshaft at short stroke. at the start of the variable-pitch transmission.
(α max imum) = angular advance of the crankshaft at short stroke. at the end of the variable-pitch transmission.
V a (α min imum) = additional volume added to the volume of the dead space. at the start of travel of the variable-pitch transmission, defined by the minimum angle of the angular advance of the crankshaft at short travel when the crankshaft of the long stroke crankshaft is in top dead center, in the compression end phase.
V a (α max imum) = additional volume added to the volume of the dead space. at the end of the variable-pitch transmission. defined by the maximum angle of the angular advance of the crankshaft at short stroke when the crankshaft at long stroke is in top dead center, at the end of compression phase.
Vr (α min imum) = air delivery volume at the start of the variable-pitch transmission. defined by the minimum angle of the angular advance of the crankshaft at short stroke when the crankshaft at long stroke is in bottom dead center. at the end of admission phase.
Vr ( α max imum) = air delivery volume at the end of travel of the variable pitch transmission. defined by the maximum angle of the angular advance of the crankshaft with short stroke when the crankshaft with long stroke is in bottom dead center. at the end of admission phase.

$$\mathit{\text{V1 +}}\text{[}\mathit{\text{V2 - Vr (α)}}\text{] x nomb. de grp. de 2 cyl. = cylindrée du moteur définie par le calage de la transmission à calage variable.}$$

$$\frac{\mathit{\text{V1+}}\text{[}\mathit{\text{V2-Vr(α)}}\text{]}\mathit{\text{+ve}}}{\mathit{\text{ve+Va(α)}}}\text{= P. théorique}$$

caractéristique volumétrique théorique du moteur avec définition des rapports volumétriques agencés par le calage de la transmission à calage variable. $$\frac{\mathit{\text{V1+}}\text{[}\mathit{\text{V2-Vr(α}}\text{min}\mathit{\text{imum)}}\text{]}\mathit{\text{+ve}}}{\mathit{\text{ve+Va(α}}\text{min}\mathit{\text{imum)}}}\text{= P max}\mathit{\text{imum}}$$

définition du rapport volumétrique maximum en début de course de la transmission à calage variable. En pratique, on peut considérer que Vr (α minimum) ne doit pas se déduire de V2 car trop négligeable. $$\frac{\mathit{\text{V1+}}\text{[}\mathit{\text{V2-Vr(α}}\text{max}\mathit{\text{imum)}}\text{]}\mathit{\text{+ve}}}{\mathit{\text{ve+Va (α}}\text{max}\mathit{\text{imum)}}}\text{=}\mathit{\text{P}}\text{min}\mathit{\text{imum}}$$

Characteristics and formulas of the volumetric ratios of the engine with combustion chamber with variable volume. $$\mathit{\text{(V1 + V2)}}\text{x number. of grp. of 2 cyl. = engine displacement.}$$

$$\mathit{\text{V1 +}}\text{[}\mathit{\text{V2 - Vr (α)}}\text{] x number. of grp. of 2 cyl. = engine displacement defined by the timing of the variable timing transmission.}$$

$$\frac{\mathit{\text{V1 +}}\text{[}\mathit{\text{V2-Vr (α)}}\text{]}\mathit{\text{+ ve}}}{\mathit{\text{ve + Va (α)}}}\text{= Theoretical P.}$$

theoretical volumetric characteristic of the engine with definition of the volumetric ratios arranged by the timing of the variable-timing transmission. $$\frac{\mathit{\text{V1 +}}\text{[}\mathit{\text{V2-Vr (α}}\text{min}\mathit{\text{imum)}}\text{]}\mathit{\text{+ ve}}}{\mathit{\text{ve + Va (α}}\text{min}\mathit{\text{imum)}}}\text{= P max}\mathit{\text{imum}}$$

definition of the maximum volumetric ratio at the start of travel of the variable-pitch transmission. In practice, we can consider that Vr (α min imum) should not be deduced from V2 because it is too negligible. $$\frac{\mathit{\text{V1 +}}\text{[}\mathit{\text{V2-Vr (α}}\text{max}\mathit{\text{imum)}}\text{]}\mathit{\text{+ ve}}}{\mathit{\text{ve + Va (α}}\text{max}\mathit{\text{imum)}}}\text{=}\mathit{\text{P}}\text{min}\mathit{\text{imum}}$$

Definition of the minimum volumetric ratio at the end of travel of the variable-pitch transmission. In practice, we can consider that Vr (α max imum) must not be deduced from V2 because the mass admitted in V1 and V2 is dependent on the calibration memorized at the maximum boost pressure.

We can accept a simplified formula of the volumetric ratio depending on whether Va (α) is located at any angular position between the start and the end of the travel of the variable-pitch transmission, that is: $$\frac{\mathit{\text{V1 + V2 + ve}}}{\mathit{\text{ve + Va (α)}}}\text{= P}$$

In accordance with the invention, the minimum volumetric ratio selected can be achieved between two limit switches of the variable-pitch transmission. The first limit is achieved with a maximum angular advance of the crankshaft of the short stroke (5) relative to the crank of the long stroke crankshaft (4) so as to determine at the end of compression (top dead center of the piston 6) positioning the piston (8) in relation to the additional space necessary for the dead space (10) to define said minimum volumetric ratio with an angle of at least 90 ° between the connecting rod and the crankshaft of the short-stroke crankshaft (5), the second limit is achieved with a lower angular advance of the crankshaft of short stroke (5) compared to the crank of long stroke crankshaft (4) and this in proportion to the decrease in the ratio between the two displacements of the two cylinders (2 and 3) up to the tolerance limit generated by the working space of the two crankshafts (4 and 5) defined by the parallel and close positions of the two grouped cylinders (2 and 3 ) according to the formula of the minimum volumetric ratio below. $$\frac{\mathit{\text{V1 +}}\text{[}\mathit{\text{V2-Vr (α}}\text{max}\mathit{\text{imum)}}\text{]}\mathit{\text{+ ve}}}{\mathit{\text{ve + Va (α}}\text{max}\mathit{\text{imum)}}}\text{=}\mathit{\text{P}}\text{min}\mathit{\text{imum}}$$

One can calculate a greater volumetric ratio between the two cubic capacities of the two cylinders grouped in order to reduce the stresses of forces on the transmission with variable timing on the motors with smaller cubic capacity, conversely one can calculate a smaller volumetric ratio between the two displacements of the two grouped cylinders (2 and 3) in order to increase the speed of the larger displacement motors.

In practice. we can consider that Vr (α max imum ) should not be deduced from V2, because the mass admitted in V1 and V2 is dependent on the calibration memorized between the volumetric boost and the boost pressure.

The maximum volumetric ratio selected is achieved on the basis of the data of the dimensional values defined for the minimum volumetric ratio, in such a way that at the start of the travel of the variable-pitch transmission, the minimum angular advance of the crankshaft at small stroke (5) relative to the crankshaft of the long-stroke crankshaft (4) determines, at the end of compression (top dead center of the piston 6), the positioning of the piston (8) in relation to the additional space necessary for the dead space (1O) to define a maximum volumetric ratio with the connecting rod (9) of the crankshaft of the short-stroke crankshaft (5) separated from its top dead center, so that said connecting rod (9) forms an angle with the crank of the short stroke crankshaft (5). We can therefore define the maximum volumetric ratio according to the formula: $$\frac{\mathit{\text{V1 + [V2-Vr (α}}\text{min}\mathit{\text{imum)}}\text{]}\mathit{\text{+ ve}}}{\mathit{\text{ve + Va (α}}\text{min}\mathit{\text{imum)}}}\text{=}\mathit{\text{P}}\text{max}\mathit{\text{imum}}$$

In practice, we can consider that V _r (α min imum) must not be deduced from V2, because the mass of air admitted in V1 and V2 is dependent on the calibration memorized between the volumetric ratio and the atmospheric depression in the intake pipe.

a =

point mort haut du petit cylindre

b =

sommet du petit piston

s =

surface du petit piston

l =

longueur de la petite bielle

r =

longueur du petit vilebrequin

A =

point mort haut du grand cylindre

B =

sommet du grand piston

S =

surface du grand piston

L =

longueur de la grande bielle

R =

longueur du grand vilebrequin

Vm =

volume mort

α =

rotation angulaire (0° au point mort haut) (sens anti-trigonométrique)

ϕ =

avance angulaire du petit vilebrequin par rapport au grand vilebrequin

The diagrams in Figures 10 and 11 are based on the following formula:

a =

top dead center of the small cylinder

b =

top of the small piston

s =

small piston surface

l =

length of small connecting rod

r =

length of small crankshaft

A =

top dead center of the large cylinder

B =

top of large piston

S =

large piston surface

L =

length of large connecting rod

R =

length of large crankshaft

Vm =

dead volume

α =

angular rotation (0 ° at top dead center) (anti-trigonometric direction)

ϕ =

angular advance of the small crankshaft relative to the large crankshaft

Example to make the engine functional and efficient according to one of the many applications.

The above formula recorded in a computer spreadsheet allows you to manage and select the dimensional values between the different elements of the engine, that is to say, the volumetric ratios between the two displacements of the two grouped cylinders (2 , 3) and the ratio formed by the total volume of the two displacements of these cylinders (2,3) with the volume formed by the dead space (40), the calculation is established so that the specifications which have been provided for the maximum and minimum volumetric ratios of the engine may coincide with the corresponding degrees of the minimum and maximum angular advances of the crankshaft at short stroke relative to the crankshaft at long stroke respectively of the start and end of travel of the stalled transmission variable. The graphs in FIGS. 10 and 11 show examples of curves of variations in the volumetric ratio and in the volumetric efficiency of the two grouped cylinders (2,3) over 360 ° degrees of angular rotation of the crankshaft of the long stroke crankshaft (4).

According to a particular embodiment of the invention, in a high capacity generator set, but not exclusively, the two crankshafts (4 and 5) are each mechanically connected to a generator, the electrical circuits of the two generators are connected in parallel. The capacity of each of the two generators is defined as a function of the power of their respective crankshaft in cruising speed of the engine, therefore, the variable-pitch transmission and the corresponding couplings between the two crankshafts (4 and 5) are limited to efforts to compensate couples.

• augmentation du rendement volumétrique;
• augmentation de la puissance massique ;
• diminution des pertes par frottements mécaniques;
• adaptation du moteur à l'indice de cétane;
• définition avec précision d'une température de fin de compression idéale pour l'autoinflammation du carburant dans toutes les circonstances envisageables ( du démarrage à froid jusqu'aux hautes pressions de suralimentation);
• meilleures performances du moteur en altitude;
• minimisation des rejets d'oxyde d'azote et d'hydrocarbures imbrûlés.

Advantages for the four-stroke engine with compression ignition.

• increased volumetric efficiency;
• increased mass power;
• reduction of losses by mechanical friction;
• adaptation of the engine to the cetane number;
• definition with precision of an ideal end-of-compression temperature for auto-ignition of the fuel in all conceivable circumstances (from cold start up to high boost pressures);
• better engine performance at altitude;
• minimization of nitrogen oxide and unburnt hydrocarbon releases.

• augmentation du rendement volumétriques ;
• augmentation de la puissance massique ;
• diminution des pertes par frottements mécaniques et par pompages ;
• augmentation du rendement du moteur en charges partielles, du fait de l'augmentation du
• taux de compression proportionnellement à la dépression dans la pipe d'admission. ( fermeture du papillon du gaz )
• adaptation du moteur à l'indice d'octane ;
• meilleures performances du moteur en altitude ;
• meilleure homogénéité du mélange ;
• minimisation des rejets de monoxyde de carbone, d'oxydes d'azote et d'hydrocarbures imbrûlés.

Avantages et conditions d'utilisation du moteur à quatre temps à allumage par compression à haut taux de pression de suralimentation sur les véhicules tracteurs routiers.Advantages for the four-stroke spark ignition engine.

• increase in volumetric efficiency;
• increased mass power;
• reduction of losses by mechanical friction and by pumping;
• increased efficiency of the engine at partial loads, due to the increase in
• compression ratio proportional to the vacuum in the intake pipe. (closing the throttle valve)
• adaptation of the engine to the octane number;
• better engine performance at altitude;
• better homogeneity of the mixture;
• minimization of releases of carbon monoxide, nitrogen oxides and unburnt hydrocarbons.

Advantages and conditions of use of the four-stroke compression-ignition engine with high boost pressure rate on road tractor vehicles.

The reduction in the displacement of each cylinder of the engine according to the criterion of the average speed of the pistons, allows an increase in engine speed and a consistent decrease in low frequencies. A greater reduction will be provided on the gearbox - drive shaft assembly until the second reduction of the engine axle. As the mechanical friction is proportional to the displacement and not very sensitive to the load, the efficiency is improved. The engine brake can be maintained by considering an increase in the power of the engine with the support of a speed limiter on the vehicle.

Claims (6)

1. A four-stroke internal combustion engine having an intake phase, a compression phase, an expansion phase and an exhaust phase, said engine including :

   - reciprocating pistons (6, 8), provided with self-ignition or spark ignition means;

   - two crankshafts, the first crankshaft (4) having a long-stroke crank and the second crankshaft (5) having a crank with a stroke shorter than that of the first crankshaft, said crankshafts (4, 5) being coupled at the same rotational speed via a gear train (19, 20, 21) and a variably timed transmission;

   - a number of cylinders (2, 3), arranged each above one of the two crankshafts (4, 5), said arrangement including small cylinders (3) with a displacement smaller than that of the larger cylinders (2), each larger cylinder communicating with a smaller cylinder (3) via a clearance space (40), so as to form a group of two cylinders (3, 4) in communication with each other, so as to enable gases to flow from one cylinder to the other, irrespective of the position of the pistons (6, 8) in each of said cylinders (2, 3), each piston being associated to a connecting rod (9) coacting with a crank of a crankshaft, the crank of the second crankshaft (5) coacting with the connecting rod (9) of the piston (8) in the smaller cylinder (3), and the crank of the first crankshaft (4) coacting with the connecting rod (7) of the piston (6) in the larger cylinder (2);

   - a camshaft in mesh, at half speed, with the first crankshaft (4), so as to connect periodically groups of two cylinders (2, 3) with intake and exhaust pipes (11, 12) via intake and exhaust valves (13, 14), at definite moments of the four-stroke cycle,

   in which the variably timed transmission has a control mechanism to vary the lead angle between the crank of the second crankshaft (5) and the crank of the first crankshaft (4), by means of a hydraulic force amplifier having a controlled thrustor acting on the transmission, said transmission altering at the end of the compression phase of the piston (6) in the larger cylinder (2), the compression ratio of the engine between a minimum and a maximum, said minimum and maximum compression ratios depending on :

   a) the ratio between the displacement of the larger cylinder (2) and the displacement of the smaller cylinder (3), and

   b) the ratio between, on the one hand, the total volume of the smaller cylinder and the larger cylinder and between, on the other hand, the volume of the clearance space (40) and an additional volume created in the smaller cylinder (3) at the end of the compression phase of the piston (6) in the larger cylinder (2), the variably timed transmission adjusting the lead angle between the crank of the second crankshaft (5) and the crank of the first crankshaft (4), so as to obtain said compression ratios, said lead angle varying between a maximum so that an angle of at least 90° is obtained between the connecting rod (9) of the piston (8) in the smaller cylinder (3) and the crank of the second crankshaft (5), at the end of the compression phase of the piston (6) in the larger cylinder (2), in order to define the minimum compression ratio, and a minimum so that the lead angle corresponds, at the end of the compression phase of the piston (6) in the larger cylinder (2), to the appropriate position of the piston (8) in the smaller cylinder (3) to create the additional volume required to obtain the maximum compression ratio, the crank of the second crankshaft (5) forming an angle with the connecting rod (9) of the piston (8) in the smaller cylinder (3).
2. A four-stroke internal combustion engine according to claim 1, for which the crank of the first crankshaft (4) passes through a top dead center and a bottom dead center during its rotation, in which the paired cylinders (2, 3) with parallel and close positions and the two crankshafts (4, 5) are arranged to define a minimum working space of the two crankshafts, so that a minimum compression ratio between the paired cylinders (2, 3) is obtained, and in which the variably timed transmission travels between a start-of-travel position and an end-of-travel position, the minimum compression ratio of the paired cylinders (2, 3) being obtained at the end-of-travel of the variably timed transmission and calculated according to the following formula : $$\frac{\mathit{\text{V1 +[V2 - Vr (α maximum)] +ve}}}{\mathit{\text{ve +Va (α maximum)}}}\text{=}\mathit{\text{P minimum}}$$

   

   in which

   V1 :   displacement of the larger cylinder (2) of the paired cylinders (2, 3)

   V2 :   displacement of the smaller cylinder (3) of the paired cylinders (2, 3)

   ve :   volume of the clearance space (40) of the paired cylinders (2, 3) required for gas transfer between the cylinders (2, 3), without excessive lamination.

   α maximum :   lead angle of the crank of the second crankshaft (5), at the end-of-travel of the variably timed transmission

   Vr (α maximum) :   compressed air volume at the end-of-travel of the variably timed transmission, defined by the lead angle of the crank of the second crankshaft (5), when the crank of the first crankshaft (4) is located at its bottom dead center, at the end of the intake phase

   Va (α maximum) :   additional volume added to the clearance space volume (40), at the end-of-travel of the variably timed transmission, defined by the lead angle of the crank of the second crankshaft (5), when the crank of the first crankshaft (4) is located at its top dead center, at the end of the compression phase.
3. A four-stroke internal combustion engine according to claim 1 or 2, in which the variably timed transmission includes three superposed concentric members, i.e. an inner member constituted by a drive shaft (17), an outer member constituted by a sleeve (28) supporting a gear (20) for coupling the two crankshafts (4, 5), and an intermediate member located between said inner and outer members and constituted by a tube (32) which slides with respect to said inner and outer members, the sleeve (28) being held in a bearing plate (15) by means of a double-row angular contact bearing (16),

   in which the shaft (18) of the second crankshaft (5) has one end which abutes one end of the drive shaft (17), said ends having straight male splines and corresponding female splines, so as to enable their coupling and self-centering of the three members with respect to the shaft (18) of the second crankshaft (5) when the bearing plate (15) is engaged in an opening of the engine unit, and enable the transmission to be removed without having to remove the second crankshaft (5),

   in which a bearing (22) has a mounting ring (25) which forms the housing of the outer ring (26) of a bearing (23), the inner ring (27) of which is mounted on the sleeve (28) so as to hold the drive shaft (17),

   in which a spacer (29) extends between the inner ring (27) of the bearing (23) and the inner ring of the angular contact bearing (16), said spacer serving to take up the space between said rings and holding axially the ring of the angular contact bearing (16) against a shoulder of the sleeve (28),

   in which a single nut (30) holds the inner rings of the bearing (23) and of the angular contact bearing (16) and the spacer (29) on the sleeve (28),

   in which the drive shaft (17) has, on the side of the mounting ring (25), helical or straight splines (31) onto which the sliding tube (32) is engaged, the inner surface of which has helical or straight splines (33) so as to enable said tube (32) to travel helically or linearly along the drive shaft (17),

   in which the inner surface of the sleeve (28) has helical splines (35), the helix of which is contrary to that of the splines of the drive shaft when the latter are helical,

   in which the sliding tube (32) has one end permanently free outside the sleeve (28), said end being held by an inner ring of a double-row angular contact bearing (36), the outer ring of said bearing (36) being rigidly connected to a holding member (37) of the thrustor, and

   in which the helical splines are arranged so that when the sliding tube (32) travels out of the sleeve, said tube reduces the lead angle between the crank of the second crankshaft (5) and the crank of the first crankshaft (4).
4. A four-stroke internal combustion engine according to any claim 1 to 3, in which the spark ignition means includes at least one spark plug in the clearance space (40), the ignition being effected in sychronism, at half speed, with the first crankshaft (4).
5. A four-stroke internal combustion engine according to any claim 1 to 4, in which the ratio between the displacements of the paired cylinders (2, 3) is between 2.5 and 5.
6. A four-stroke internal combustion engine according to claim 3, in which the gear (20) supported by the sleeve (28) has respectively an even or uneven number of teeth when the number of teeth of the splines between the drive shaft (17) and the sliding tube (32) and the number of teeth of the splines at the abuting ends of the drive shaft (17) and of the shaft (18) of the second crankshaft (5) are respectively uneven or even.

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