ITMI20110330A1 - ALTERNATIVE ENDOTHERMAL ENGINE WITH MOTORCYCLE CONVERSION MECHANISM INCLUDED IN THE PISTON - Google Patents

Description

â € œAlternative endothermic engine with motion conversion mechanism included in the pistonâ €

Field of application of the invention

The present invention is applied to the motor industry, and more precisely to a reciprocating internal combustion engine with motion conversion mechanism included in the piston.

Review of the known art

It is useful to underline how the invention of the internal combustion engine has marked technical progress and changed the way of life and work of all humanity. The continuous evolution in the sector has produced internal combustion engines of the most different types. By restricting the field of interest to the only technique closest to the invention that will be described, we could group these engines into two broad categories: the reciprocating engine category and that of rotary engines. The more traditional cylinder and piston automotive engines belong to the first category, the latter driven into translational motion by the pressurized gases generated during the rapid combustion of a mixture of compressed air and fuel within the cylinder chamber, so as to operate a connected piston. to a thrust crank mechanism which rotates a motor shaft. Continuous rotation engines such as gas turbines belong to the second category, but we will not consider them below as they are not relevant for the purposes of the invention, and so-called rotary piston engines, such as the Wankel, which we will consider instead. In reciprocating engines the piston completes a complete cycle in two or four revolutions of the crankshaft; Combustion can be spontaneous, as in the Diesel type, or caused by the spark sparked between the electrodes of a spark plug at the end of the compression phase. In the four-stroke engine, a valve distribution opens and closes with appropriate synchronism of the lights in the combustion chamber head, while in the two-stroke engine it is the piston itself that, during its stroke, blocks the outflow and outflow ports present in the wall. side of the cylinder.

For a better understanding of the invention that will be described, and of the problems present in the known art in this regard, it is necessary to study the two types of motors mentioned above in greater depth.

Figure 1 is a schematic of a Wankel 1 engine in longitudinal section. In the figure you can see a stator 2 with a rotor 3 inside. The internal profile of the stator is that of a two-lobed epitrochoidal curve, corresponding to a nephroid epicycloidal curve. The side wall of the stator 2 is extruded from the geometric profile shown in the figure, as is the body of the rotor 3. The stator casing is closed by two bases crossed by the motor shaft. A pinion 4 is rigidly connected to the center of a base of the stator case 2, internally to it. The external profile of rotor 3 is that of an equilateral triangle with convex sides (Rouleaux triangle). The rotor 3 has an internal toothed crown 5 which meshes with the pinion 4, having a double diameter with respect to the external toothed crown of the latter. A driving shaft 6 is free to rotate within the pinion 4. A circular eccentric 7 is keyed to the driving shaft 6, free to rotate within a large bushing 8 integral with the rotor 3 in a central position. On one side of the stator 2 open the mouths of two ducts 9 and 10, respectively for outflow and outflow. On the opposite side with respect to the mouths 9 and 10, two spark plugs 11 and 12 partially cross the stator wall in correspondence with a combustion chamber 13a, 13b which is formed in this position.

In operation, the Wankel engine, like the two-stroke engine, has no distribution valves; It is in fact the rotor 3 that cyclically opens the mouth of the outflow duct 9 and closes the mouth of the outflow duct 10. Following the direction of rotation indicated by the arrow, the intake phase of the air-petrol mixture begins when the volume of the chamber where the intake duct 9 opens, it begins to increase, thus creating a depression that draws the mixture back into the stator chamber. Continuing its movement, the rotor 3 causes a reduction in the space between its wall and that of the stator, so that the sucked mixture is compressed. When the compression has reached the optimal value, the spark is fired through the spark plug electrodes; thus begins the combustion phase and, therefore, the expansion of the gases; the consequent pressure increase causes forces to act on the rotor, forcing it to continue in its rotary motion. The stator pinion has the task of forcing the rotor to follow an eccentric orbit, such as to ensure constant contact of the apical sealing elements against the internal walls of the stator. The eccentric 7 acts as a crank set in rotation by the translational component inherent in the eccentric rotation of the rotor 3, transforming it into a rotational component transferred to the motor shaft 6. Once the expansion phase is completed, another volume reduction occurs during which the burnt gases are pushed out of the engine through the exhaust duct 10. From the kinematic point of view, the Wankel rotor 3 performs a planetary movement, that is, it is mobile around a non-fixed axis, which in turn rotates around to the fixed axis of the crankshaft 6.

The Wankel motor is very simple as it consists of only two moving parts, the rotor and the crankshaft. The axis of symmetry of the rotating piston 3 rotates around the axis of the stator 2, resulting in an eccentric mass to which a centrifugal inertia force is applied. However, balancing this force does not pose any problems, as the second order inertia forces (typical of reciprocating engines) are absent. Despite its kinematic simplicity, the Wankel engine has not proved capable of competing with the reciprocating engine in the automotive sector, due to construction complications and higher fuel consumption. In addition to this, the main disadvantages consist in the poor life of the rotor sealing elements; in the quite contained torque at low revs; in the problematic lubrication of the apical segments; and in the very high rate of unburnt hydrocarbons.

A two-stroke reciprocating engine free from these problems is instead shown in the succession of figures 2, 3, 4, and 5 which illustrate its operation in four different configurations assumed during a 360 ° rotation of the crankshaft. The term â € œtwo-strokeâ € derives from the fact that a complete Otto cycle uses only one piston rise and fall period, corresponding precisely to the aforementioned 360 ° rotation. As regards the transfer of power to the crankshaft, it is on a par with the Wankel engine which has three active phases every three revolutions of the crankshaft.

Referring to Figure 2, it can be seen a two-stroke reciprocating engine 20, consisting of a cylinder 21 within which a sliding piston 22 equipped with seal rings (solid) has reached the top dead center position (TDC). In such a position, a combustion chamber 23 is created between the piston crown 22 and the internal wall of the cylinder head 21. The chamber 23 is filled with a mixture of air and fuel strongly compressed during the upward motion of the cylinder. piston 22. A spark plug 24 passes through the head of cylinder 21 to ignite combustion when a spark produced by its electrodes protruding inside the combustion chamber 23. The piston 22 is hollow and contains the upper end (the foot) of a connecting rod 25 hinged to a pin 26 integral with the piston 22. The lower end (the head) of the connecting rod 25 is hinged to the pin 27 of a thrust crank 28 integral with the crankshaft 29. The shape of the crank 28 is suitable for balancing its rotational mass. The wall of the cylinder 21 is contiguous to the wall of a crankcase 30 inside which the crank 28 and a large part of the connecting rod 25 are housed. An intake duct 31 is open in the lower part of the crankcase 30 for the inlet of the air-fuel mixture. The piston 22 at the TDC occludes a port 32 open in the wall of the cylinder 21 in the area where a duct 33 for the flow of the air-fuel mixture, previously sucked into the crankcase-pump 30 by the rising of the piston 22 ends. In the same way the piston 22 occludes an open port 34 in the wall of the cylinder 21, on the opposite side of the port 22 and in line with it, where a flue gas discharge duct 34 begins. The figure does not show a one-way reed valve inserted at the inlet of the crankcase-pump 30 which opens automatically when a vacuum forms in the crankcase, and then closes again when atmospheric pressure is restored. This valve is made up of a usually triangular support with openings on which elasticized "petals" (lamellae) in steel, carbon fiber or glass fiber rest, and have a thickness in the order of tenths of a millimeter. The reed valve allows you to automatically adjust the intake timing to the needs of the engine.

Figure 3 shows the position of the piston 22 reached during the combustion gas expansion phase, after a quarter turn of the crankshaft 29 in the direction indicated by the arrow placed on the crank 28. In this position the inflow ports 33 and 34 are still closed, but this is not mandatory as the distance of these ports from the cylinder head 21 is established by the timing chosen in the design. The piston 22 during its descent carries out the prestressing of the air-fuel mixture previously sucked into the pump-casing 30 during the upward motion. During its descent, the piston 22 by means of the connecting rod 25 and the crank 28 supplies the crankshaft 29 with an active moment.

Figure 4 shows the position reached by the piston 22 at bottom dead center (PMI). As the ports 32 and 33 are opened by the piston 22 moving towards the PMI, the fresh charge enters the inside of the cylinder 21 through the duct 33 and the burnt gases exit from the duct 34. This phase is also called washing as the fresh charge contributes to the outflow of the burnt gases and, except for particular precautions, is partly dispersed through the exhaust. These measures consist in appropriately shaping the piston crown but above all in the use of an efficient resonant exhaust system, able to favor the extraction of the burnt gases and to generate a retrograde pressure wave near the mouth outlet of the muffler which pushes the unburned mixture back into the cylinder.

Figure 5 shows the position reached by the piston 22 after three quarters of a turn of the crankshaft 29. The piston 22 during its inertial rise closes the ports 32 and 33 and begins to compress the air-fuel mixture trapped in the upper part of the cylinder 21. At the same time a new charge is sucked into the pump casing 30 through the intake duct 31 due to the effect of the depression which is created during the rising of the piston 22.

In operation, when the piston 22 reaches the dead points PMS and PMI its speed is zero and then reverses. Stall is avoided thanks to the flywheel consisting of the circular expander of the crank 28 which, thanks to the accumulated kinetic energy, makes the piston run continue beyond the dead point The connecting rod 25 which connects the piston 22 to the crank 28 performs a movement rototranslatory which allows the transformation of the rectilinear motion of the piston into the rotary motion of the crank, and therefore of the crankshaft 29.

In two-stroke engines with dry sump functioning as pump casing, mainly used in some categories of motorcycles and in tools for various outdoor works, for example, gardening, chainsaws, etc., the lubrication of the cylinder walls in contact with the piston, as well as the crank mechanism inside the crankcase, it is mostly carried out by dispersing a percentage of lubricating oil in the air-fuel mixture which is sucked into the crankcase-pump and then sent to the transfer port. This â € œto loseâ € lubrication considerably simplifies the engine but is rather polluting. Two-stroke engines with separate lubrication are also known, for example Diesel with Root compressor (volumetric lobe) for pumping the washing air into an open seat in the cylinder liner near the PMI. In four-stroke engines the lubrication of the cylinder walls in contact with the piston, as well as of the crank mechanism inside the crankcase, is carried out by circulating pressurized oil by means of a volumetric pump through special channels present in the crankshaft and in the connecting rod; after having completed its task, the oil falls by gravity into a collection cup where it is fished out by the pump. This lubrication is called wet sump, it is easy to carry out but has some known drawbacks that need not be listed. The most sophisticated four-stroke engines use a small dry sump and at least a second pump to take the oil that settles there and immediately put it back into circulation.

Highlighting of technical problems

The reciprocating engine is mechanically and dynamically very different from the Wankel rotary; in fact, unlike the latter, during the reciprocating movement of the piston additional forces of inertia of the second order occur which, like those of the first order, act on the mechanical constraints, in particular the journals of the crankshaft. More precisely, called Ï ‰ the angular speed of rotation of the crankshaft, the mathematical expression of inertial forces highlights first sinusoidal components at pulsation Ï ‰, and second components at pulsation 2Ï ‰, respectively called inertial forces of the first and second order (although the more appropriate term would be that of inertial moments). The two components are additive and, if not perfectly balanced, they induce anomalous stresses on the main journals capable of deforming the crankshaft and propagating through the frame in the form of noisy vibrations. These vibrations strain the moving mechanical parts, increasing their wear and reducing their operating life. It is therefore essential to neutralize the aforementioned dynamic forces so that only the weight of the piston and the relative thrust crank mechanism are applied to the main journals. Neutralization generally consists in balancing the moments produced by the various moving parts with respect to the axis of the crankshaft. Balancing by adding eccentric masses to the rotating parts. The balancing technique does not pose serious problems in the neutralization of the inertial forces of the first order, as the movement of the parts involved in the thrust crank mechanism is cyclical at the Ï ‰ pulsation, but not the same happens in the neutralization of the inertial forces of the second order, requiring of countershafts rotating at angular speed -2Ï ‰. Obviously, the difficulties in neutralizing the inertial forces generally decrease as the number of cylinders increases, since it is possible to choose a particular arrangement of the same along the crankshaft together with an appropriate sequence of the combustion phases designed to compensate between their various inertial contributions.

Another intrinsic problem of reciprocating engines consists in their overall dimensions in the direction of the longitudinal axis of the cylinder. This fixed footprint can be immediately evaluated from the PMI configuration shown in figure 2; it is approximately equal to the sum of the following contributions: a) the height of piston 22; b) the length of the connecting rod 25 decreased by the internal portion of the piston, c) approximately double the crank arm 28. Such a size cannot be reduced more than that by resorting to shorter connecting rods for the following reasons: a) it should also be increased the diameter of the cylinder to contain the greater inclination of the connecting rod when the crank is horizontal, b) as will be seen shortly, the shorter length of the connecting rod involves an increase in the inertial forces of the second order, more difficult to neutralize, and it will therefore be necessary to achieve a compromise between the smaller mechanical dimensions and the lower inertial forces.

The contribution of the connecting rod to the inertial forces is usually divided into two contributions, of which, a first of them purely translational is that of a concentrated mass located in the small end, pin side, while a second purely rotary contribution is that of a concentrated mass located in the big end, crank side. Knowing the length of the connecting rod and its mass, the two concentrated masses are calculated by canceling the respective moments with respect to the center of gravity. The first contribution is added to the inertial forces of the first and second order generated by the piston, while the second contribution is added to the inertial forces generated by the crank. Since the motion of the crank is purely rotary, the total inertial moment including the contribution coming from the connecting rod has only inertial components of the first order that can be easily balanced.

The inertia forces Fbp of the mass connecting rod foot mbp which moves together with the piston with acceleration a, are given by the expression:

(1)

in which:

â € ¢ It is the angle between the cylinder axis and the crank axis at the moment;

â € ¢ It is the angular pulsation of the crank;

â € ¢ It is the crank radius;

where is the length of the connecting rod. The parameter generally has a low value; in the example with results

The forces of inertia of the first order are given by:.

Second order forces of inertia are given by:

which at the same angular pulsation show a double pulsation cyclic trend. The parameter that appears with a multiplicative value in the expression of the inertial forces of the second order only, demonstrates that they decrease as the length of the connecting rod increases with respect to the crank radius, but at the same time increases the size of the engine in the direction of the longitudinal axis of the cylinder.

In the alternative endothermic engine technique, the use of connecting rods connected to a 2: 1 hypocycloidal crank mechanism is known, capable of imparting a linear motion to the connecting rod capable of keeping the connecting rod aligned with the axis of the cylinder, so as to exclude the onset of second order inertial forces within the connecting rod. The connecting rods used in these examples do not change their traditional elongated shape nor the type of connection to the piston, which is why a simplification of the mechanics or a reduction in overall dimensions is not obtained. Aims of the invention

The purpose of the present invention is to reduce the inertial forces generated in the reciprocating endothermic engines, in particular by the connecting rods, simultaneously simplifying the mechanics and reducing the overall dimensions.

Summary of the invention

To achieve these purposes, the present invention relates to an internal combustion engine comprising:

∠’at least one hollow body, hereinafter called a cylinder, whose cavity closed at one end of the head extends along a longitudinal axis;

∠'a hollow piston that can move inside the cylinder in contact with the wall, delimiting a variable volume chamber in contrast to the cylinder head in which to cyclically trigger the combustion of a gaseous mixture of air and fuel and exploit the burnt gases to accelerate the piston ;

∠’a mechanism for converting the translational motion of the piston into the rotary motion of the crankshaft, and vice versa,

in which according to the invention:

â 'the said mechanism is contained in the piston, the crankshaft crossing the wall of the cylinder and of the piston orthogonally to the said longitudinal axis, the wall of the piston having opposite openings that allow it to move across the crankshaft during the entire run, as described in claim 1.

Further advantageous characteristics of the present invention, in its various embodiments considered innovative, are described in the dependent claims.

According to one aspect of the invention, both the cylinder and the piston have a parallelepiped shape with the faces crossed by the crankshaft almost square, the piston being closed by a base wall.

According to a first example of embodiment of the invention, the said mechanism for converting the translational motion of the piston into the rotary motion of the crankshaft includes:

∠’a pinion integral and coaxial to a crank pin of the crankshaft, the primitive circumference of the pinion teeth having a radius equal to the distance between the pinion axis and the crankshaft axis;

∠’a toothed crown obtained in a septum integral with the piston arranged so that the internal teeth of the crown can mesh with the teeth of the pinion, the radius of the pitch circumference of the internal tooth crown being double the radius of the pitch circumference of the pinion teeth;

â ’a circular eccentric free to rotate on the crank pin remaining in contact with the piston wall during the entire stroke so as to keep the pinion constantly engaged with the crown.

The first example of embodiment of the invention applies the geometry of hypocycloidal curves with a ratio of 2: 1 between the radius of a fixed circumference and the radius of a mobile circumference that rolls within the fixed one. With this ratio it is known that a point on the mobile circumference runs through a diameter (roller) of the fixed circumference in both directions during a complete roll. The novelty consists in the fact that the internal toothed crown (corresponding to the fixed circumference in the geometric representation) is now forced to translate along its own diameter alternately, forcing the pinion, with the help of an eccentric, to roll within the crown by turning a crank. In practice, the internal toothed crown is obtained in a wall inside the piston in the form of a septum, while the eccentric equipped with annular bands of anti-friction metal can rotate within a seat obtained beyond the septum of the toothed crown.

The PMS â † ’PMI translation is controlled by the flue gases while the PMI â †’ PMS translation by the inertia of a flywheel rotated by the crankshaft, as happens in traditional engines. When the piston moves from TDC to PMI, the piston wall exerts a force on the edge of the circular eccentric forcing it to rotate in one direction (or indifferently in the other depending on the initial position) around the crank pin, the eccentricity by compensating for the distance variation that would otherwise occur between pinion and crown during piston translation; with this the pinion is kept constantly engaged with the crown.

It should be remembered that the known examples of use of the 2: 1 hypocycloidal geometry are quite different from the one described, above all because the crank mechanism adopted in these examples is strictly external to the piston.

According to an aspect of the invention, the circular eccentric is a wheel or a disc with its own half-part of mass comparable to the mass of the piston, said mass rising and lowering due to the eccentric rotation in synchrony with the alternate translation of the piston , but in the opposite direction, it balances the inertial forces caused by the acceleration of the piston.

Regarding the manufacturing methods of the piston, it would be advantageous if the walls crossed by the crankshaft have large openings that allow the insertion of a gooseneck obtained in a single piece. In this way the toothing of the pinion can be obtained by removing chips from an extremity with an excess diameter of the crank pin. The eccentric can be made in two parts, each with a semicircular half-seat to be placed in contact with the crank pin. The half-sleeves that house the two half-seats will be joined together by means of bolts. As far as the cylinder is concerned, it too must allow the crankshaft to pass through and this is possible in different ways, for example, by building the cylinder in two parts that can be joined with respective half seats for the main bushings. .

According to one aspect of the invention:

â ’the walls of the sliding faces of the piston have outlet holes for the lubricating oil circulated under pressure inside the piston, and inlet holes for the oil collected from the walls;

∠’the wall of each sliding face of the piston has grooves for the insertion of sheets coupled to resilient means capable of pushing them against the cylinder wall to distribute the lubrication oil and then sweep it;

â ’first sheets being arranged along the sides of the walls crossed by the drive shaft to form perimeter bands;

â ’second sheets being arranged transversely to the adjacent walls and including seats on the sides for interlocking with the perimeter bands.

According to one aspect of the invention, the transverse sheets are in pairs of superimposed sheets, each having a rectangular window arranged longitudinally in an asymmetrical position with respect to the center offset from one sheet to the other, for the inclusion of resilient means suitable for exert pressure against the misaligned sides of the windows making the sheets translate in opposite directions against the perimeter bands.

According to an aspect of the invention, the resilient means coupled to the perimeter bands exert a pressure in an orthogonal direction on the longer sides of the transversal sheets, transmitted to the latter by the joints.

According to one aspect of the invention, the internal combustion engine is a two-stroke engine with separate lubrication from the fuel and inlet and exhaust ports at the same level of opposite cylinder walls, the corresponding piston walls including, each, three transversal plates, of which, two located at the ends respectively upper and lower with respect to the piston crown and the third intermediate, and said oil outlet and inlet holes being included between the intermediate plate and the lower plate, the intermediate plate being further away from the upper plate than the stroke length, placing itself beyond the inlet and exhaust ports when the piston is at top dead center, thereby preventing the entry of oil into the combustion chamber.

According to an aspect of the invention referring to the two-stroke engine, the portion of each wall of the piston between the intermediate transverse leaf and the upper transverse leaf is surface hardened, and the cylinder walls are furrowed by micro-lines capable of to retain a certain quantity of oil even after having been swept by the transversal blades rising towards the top dead center, then yielded to the surface hardened area during the descent to the bottom dead center.

The mechanism inside the piston for converting the motion from alternating to rotary, and vice versa, is simple to make and has several advantages, however the developed displacement could be too high in relation to the stroke, especially when the internal gooseneck to the piston is maximally sizing. This happens because both the area of the square section of the piston, as well as the stroke, are strictly linked to the radius of the pitch circumference of the pinion. In the case of minimum sizing of the gooseneck in which the crankshaft has the same radius R as the pinion, the stroke and the side of the square section of the piston are respectively equal to 4R and 6R, developing a displacement of 24R < 2> d, where d is the depth of the piston. It is clear that in order to limit the displacement and make it fall within the values of correct dimensional proportions, the parameter d and therefore the depth of the piston should be reduced, consequently penalizing the area of the washing and exhaust ports in the two-stroke engine.

The problem just described is solved by a second example of realization of the invention in which the motion transformation mechanism located inside the piston does not exploit the coupling between toothed wheels, but only exploits the interference exerted by the piston wall on at least one eccentric coupled to a thrust crank, as already occurs in the piston described above, but only for the purpose of maintaining meshing between pinion and crown.

Therefore, in accordance with a second example of embodiment of the invention, the aforesaid mechanism inside the piston includes at least a first crank pin on which a circular eccentric is free to rotate while remaining in contact with the piston wall during the entire stroke. , the circular eccentric having an eccentricity value equal to the distance of the axis of the crank pin from the axis of the crankshaft.

The second example of embodiment describes the simplest mechanism for converting the motion inside the piston, which is not optimal as regards the symmetry of the accelerations acting along the stroke, and therefore does not guarantee the absence of second order inertial forces, the advantages (with the same developed displacement) abundantly compensate for the gap, since it was possible to make pistons that allow the use of crankshafts and crank pins that are more dimensioned in relation to the stroke of the piston compared to the hypocycloidal gear piston 2 : 1. A variant will try to remedy the gap highlighted at the expense of a complication in the internal mechanics of the piston.

It can be appreciated how the most significant technical characteristics of the engine of the invention remain unchanged, namely that the piston includes the motion conversion mechanism inside which forces it to move astride the piston.

The example of construction based on the eccentric differs considerably from traditional alternative engines in which the connecting element between piston and crank is precisely the connecting rod, universally understood in the motoring context as a rod equipped with a circular seat at each end for as many pins: the pin integral with the piston and the crank pin integral with the crankshaft. The traditional connecting rod performs a rototranslatory motion of limited angular excursion and wide translation, which involves the problems already mentioned; furthermore the way it is made is almost completely external to the piston, as obviously the crank. Conversely, the internal eccentric of the piston performs a rototranslatory motion whose rotational component is prevalent with respect to translation (a complete rotation at each translation equal to double the eccentricity) and therefore the ideal condition of absence of inertial components of the second order is well approximated.

For a better understanding of the second example of realization of the invention it is useful to consider a Cartesian reference system in a plane transversal to the axis of the crankshaft, in which the axis of the ordinates coincides with the axis of the piston. In this reference, the rotation of the generic point belonging to the geometric axis of the crank pin is described by the functions of time:, which define a circumference, where is the length of the crank arm corresponding to the distance between the axis of the crank relative pin and the axis of the crankshaft; ; and is the angular speed of the crankshaft. The internal surface element of the piston head in contact with the edge of the circular eccentric transmits a longitudinal force to the latter, this force has a non-zero arm with respect to the axis of rotation of the eccentric and can then generate a moment that triggers an infinitesimal rotation of the eccentric around the crank pin. The elementary rotation of the eccentric, being asymmetrical, exerts a pressure against the lateral wall of the piston, whose orthogonal reaction on the edge of the eccentric is transmitted to the crank pin. The resultant of the forces acting on the crank pin has two variable components along the Cartesian axes, which also involve two respective translations subject to the rigid constraint of the crank pin with the axis of the crankshaft. This constraint, together with the dimensional choices made, involve the rotation of the crank arm around the axis of the crankshaft caused by the motion of the piston. Ultimately, the eccentric rotates around a moving axis which in turn rotates around a fixed axis. The acts of infinitesimal motion produce the macroscopic effect observable in each geometry characterized by its own value of the ratio between the length of the crank arm and the eccentricity of the circular eccentric. Geometries characterized by the value 2 allow the planetary motion of the eccentric, other values may not always be compatible with a cyclic movement of the piston. The constraint imposed by the fixedness of the distance between the walls of the piston, together with the other geometric constraints, allows the eccentric rotation to be completed when it is the crankshaft that gives the moment of thrust to be transformed into translation of the piston.

It has been seen that to generate the rotational moment of the crankshaft at any instant of the operating cycle starting from a unidirectional propulsive force, it is necessary to produce a component of the force orthogonal to this direction and of alternating direction. This derives from the fact that the point of application of the force rotates on the circumference that the crank pin must travel. In the piston of the first example of embodiment, the orthogonal component is generated by the continuous rolling of the pinion on the crown. The meshing between the two mediated by the eccentric creates a particularly robust mechanical constraint capable of transferring the necessary power to the orthogonal component. In the piston of the second embodiment example, the aforementioned orthogonal component is instead generated only by geometric constraints between the wall and the eccentric and between these and the crank pin, these constraints not as direct as the meshing, and moreover more subject to Friction due to the sliding of the eccentric on the piston walls, as usually occurs in a bushing coupling.

The problem that arises is therefore that of how to generate a force component orthogonal to the direction of translation of the piston more effectively in the absence of a `` direct '' mechanical constraint between piston and crank pin of the meshing or pivoting rod type with two ends hinged as in traditional connecting rods. More generally, the problem is how to convert the translational motion of the piston into the rotational motion of the crankshaft by means of eccentrics, and vice versa, with uniformity and symmetry of the accelerations along the entire stroke such as to minimize if not cancel the forces second order inertials.

This problem is solved by a variant of the second example of embodiment of the invention in which the internal mechanism of the piston further includes: ∠'a second crank pin with a diameter smaller than the diameter of the first crank pin, the axis of the second crank pin crank being distant from the axis of the first crank pin than the first crank pin is from the axis of the crankshaft;

∠’a double slide similar to a frame within which the second crank pin can rotate while remaining in contact with the sides of the slide;

â 'two pairs of tracks, each pair integral with a respective wall of the piston crossed by the drive shaft, the tracks having the same length as the application wall and being arranged orthogonally to the translation axis of the piston, mutually spaced to accommodate the slide and guide it during the translation in both directions performed at each turn of the second crank.

In practice, the eccentric develops the reciprocating rectilinear motion of the piston along the axis of the cylinder, while the second crank pin coupled to the slide helps to develop the alternating rectilinear motion orthogonal in the direction congruent to the motion of the piston, and this solves the announced technical problem and helps the rapid overcoming of the dead points of the race in proximity of which the orthogonal component of the force inherently assumes the lowest values. The configuration of the variant has some salient features in common with the alternating rectilinear hypocycloidal function, such as the fact that the piston stroke is equal to four times the eccentricity, and that at every 90 ° of rotation of the crankshaft a translation of the piston corresponds to the eccentricity value. The symmetry of movement that allows similar results is evident.

Advantages of the invention

The present invention in its first example of embodiment combines the advantages of the Wankel engine with those of the more traditional reciprocating engines. In fact, as for the Wankel, the inertial forces of the second order that are difficult to neutralize are absent, while as for the alternative engines the feasibility and reliability problems are minor. Furthermore, the internal mechanics of the piston are extremely simple. It should be noted that the hypocycloidal reciprocating motion of the piston, and of the eccentric that translates with it, follows a law different from that of a traditional piston bound to a classic connecting rod. In fact, in a traditional piston due to the finite length of the connecting rod, the speed at TDC is higher than the speed at PMI, and with them the relative accelerations, causing the onset of second order accelerations to equal the two contributions to the ™ overall acceleration. The piston described in the first embodiment generates only first order inertial forces, which are easier to neutralize. In fact, starting from the parametric equation of the generic hypocycloid:

(2)

where a is the radius of the circumference with the largest diameter (the ring gear) within which the circumference with the smallest diameter b (the pinion) rolls without crawling; while it is the angle swept by a point of the circumference of diameter b at time t.

By setting, we obtain:

(3)

which represents the parametric equation of a straight line corresponding to the diameter of the circumference of diameter traveled by the piston with cosine law. The corresponding acceleration is:

(4)

entirely devoid of second-order components of the type.

The internal eccentric of the piston performs a roto-translation whose translational component on the circumference traveled by the crank pin around the crankshaft is synchronous to the hypocycloidal motion of the piston, so it generates only inertial forces of the first order that are easy to balance. The eccentric therefore differs from a traditional connecting rod both in form and in operation, even if like a connecting rod it is interposed between the â € œcrankâ € and the piston but differently coupled. As already pointed out, a good part of the balancing of the inertial forces due to the piston is already advantageously obtained inside it, thanks to a studied non-uniformity in the mass of the eccentric.

The second embodiment of the present invention allows, for the same displacement, to reduce the center distances between the drive shaft and the other elements involved in the motion conversion, making it possible to reduce the overall dimensions in the mechanics with respect to both the hypocycloidal gear motor and to traditional reciprocating engines, and this allows the center of gravity to be lowered, increasing the stability of the car. The variant, at the expense of a complication of the piston, allows the minimization if not the cancellation of the inertial forces of the second order.

Lastly, since there are no problems in reaching high compression values, the engine of the present invention can operate according to the Diesel cycle without having to use a Root or other type of compressor. The complex system of bands guarantees the tightness of the oil circulated by a pump within the piston for the separate lubrication of the moving elements and an effective cooling of the same. In two-stroke engines, the belt system avoids contamination of the intake air, or of the air-fuel mixture, with the lubricating oil.

Brief description of the figures

Further objects and advantages of the present invention will become clear from the following detailed description of an example of its embodiment and from the annexed drawings given purely for explanatory and non-limiting purposes, in which: ∠'figure 1 is an elevation view of a Wankel motor with the stator open to show the rotor;

∠’Figures 2 to 5 represent four different instants of an operating cycle of a two-stroke reciprocating engine of a known type;

∠’Figure 6 is a geometric schematic of how to generate a rectilinear hypocycloid with two directions of travel, as the guiding principle of the present invention;

∠’figure 7 is a schematization of the invention that realizes a hypocycloid movement of the piston as in figure 6;

∠’figure 8 is a front longitudinal sectional view of a two-stroke endothermic engine made according to the present invention;

∠’Figures 9A, 9B, 9C, 9D, 9E, 9F are plan views of the various component parts of the piston of Figure 8;

∠’figure 10 is a sectional view according to the A-A axis of figure 8.

∠’figure 11A is a top view of a single transverse sealing plate visible in figure 9C;

- figure 1B is a perspective view of a pair of transverse sealing plates of figure 11A;

∠’figure 12A is a sectional view along the B-B axis of figure 11B;

- figure 12B shows the working configuration of the sheets of figure 12A; ∠’figure 13 is an exploded view of the frame formed by the transverse sealing plates and the perimeter bands with which the piston in figure 8 is equipped;

∠’figure 14 is a perspective view of the frame formed by foils and sealing bands of figure 13;

∠’figure 15 is a perspective view of the piston of figure 8 complete with the frame of figure 15;

∠’Figures 16A, 16B, 16C, 16D show in sequence the configurations assumed by the piston of Figure 8 in four instants of the operating cycle;

∠’figure 17 is a schematic of the piston of figure 9A which will be used in the following figures;

∠’figure 18 shows the configuration of the moving elements inside the piston of figure 17 just after the TDC where the arrows indicate the corresponding initial direction of rotation;

∠’figures 18A, 18B, 18C, 18D repeat the configurations of the moving parts inside the piston taken at the instants indicated in figures 16A, 16B, 16C, 16D;

∠’Figures 19A, 19B, 19C, 19D schematize the configurations of the moving elements, assumed inside a piston made in accordance with a second example of the invention, in the four instants of the previously indicated operating cycle;

- figure 20 is a front view of a piston according to a variant of the second example of embodiment of the invention;

∠’figure 21 is a longitudinal section view according to the C-C plane of figure 20, flanked to the front view so as to facilitate the identification of corresponding elements;

∠’figures 22A, 22B, 22C, 22D schematize the configurations of the moving elements inside the piston of figure 21 taken in the four instants of the previously indicated operating cycle.

Detailed description of some preferred embodiments of the invention

In the following description, identical elements appearing in different figures may be indicated with the same symbols. The scale and proportions of the various elements depicted do not necessarily correspond to the real ones.

Two preferred examples will be described, of which the first relates to an internal combustion engine whose piston includes a toothed couple with a 2: 1 gear ratio to achieve a straight hypocycloidal movement, while the piston of the second example performs a similar movement using a or more eccentric.

Preliminary to the first example are the schematizations of figures 6 and 7 which deal with the rectilinear hypocycloidal motion.

Figure 6 schematizes the rectilinear hypocycloid traveled in alternating directions from the generic point of a mobile circumference that rolls on a circumference with a double diameter. The point to track has been highlighted with a thick radius marked along a 360 ° rolling trajectory divided into 30 ° values. By observing the line curve of figure 6 which joins the positions gradually occupied by the center of the mobile circumference during a 360 ° rolling, it can be seen that the mobile circumference completes a complete rotation with respect to its center during a 360 ° rolling, and therefore during an alternate cycle of translation of the fixed circumference along its own diameter. Therefore, by rigidly constraining the center of the mobile circumference to any point of a third circumference having the same value as the radius and free to rotate around a fixed axis, the rotational movement of the mobile circumference will be transferred to it during rolling.

Figure 7 shows the cross section of an assembly which schematises inside a piston (not shown): the circular edge (40) of the crankshaft; the primitive (41) of the pinion, the primitive (42) of the crown gear integral with the piston; the circular edge of an eccentric (coinciding with primitive 42 for the sole purpose of simplifying the drawing, since the edge of the eccentric is always in contact with the wall of the square section piston). Referring to figure 7, it is possible to notice three parallel lines called As, A0, Ai mutually spaced by a value equal to the diameter D of the primitive (42) of the ring gear integral with the piston, in which: the line A0 represents the fixed dimension maintained from the center of the circumference (40) during the rotation of the crankshaft in steps of 30 ° in the complete cycle of 360 °. The upper line As represents the height of the top dead center TDC reached by the primitive (42). The lower line Ai represents the PMI lower dead center height reached by the primitive (42). As can be seen in figure 7, the primitive (41) of the pinion has a diameter equal to D / 2, and the eccentricity of the primitive (41) with respect to the center of the circumference (40) is equal to D / 4. The figure illustrates the movement of the piston, represented by the primitive (42) of the crown, whose center translates by alternating motion making a distance D in each direction of translation, equal to the stroke between the two dead points. This movement is controlled by the primitive (41) of the pinion which rolls within the primitive (42) of the crown, constantly meshing it, forcing its center to complete the aforementioned alternating straight trajectory. The condition of constant meshing indispensable for the 2: 1 planetary motion is made possible by the eccentric interposed between the extension of the pinion axis and the internal wall of the piston cavity, avoiding the latter to â € œ fallâ € literally on the pinion. The movement of the eccentric is necessarily rototranslatory, having to participate simultaneously in the alternating rectilinear movement of the piston that contains it and in the rotational movement around the axis of the pinion on which it is free to rotate. The kinematics of such a movement make it rotate in the opposite direction to the rotation of the crankshaft (as can be seen from the arrows).

Figure 8 is a front longitudinal sectional view of a two-stroke reciprocating internal combustion engine 45, comprising a cylinder 46 with inside a piston 47 of new conception. In the figure, the piston has reached the top dead center. The two-stroke engine in the example exploits the thrust exerted by the gases produced by the combustion of a gaseous mixture of air and petrol introduced into the cylinder, the so-called â € œfresh charge, produced by a carburetor that is not shown, where the ignition of combustion is caused by the spark between the electrodes of a spark plug. The example does not limit the invention as regards the following variants: a) the introduction of air into the cylinder and the injection of petrol into the compressed air from the piston to obtain the mixture directly in the cylinder; b) the introduction of air into the cylinder and the injection of high pressure diesel into the compressed air from the piston (or from an external compressor) which undergoes self-combustion according to the Diesel technique; c) the applications of the techniques of the previous points to a four-stroke engine. The variants at points a), b), c) in fact require modest changes to the cylinder only according to the known technique, without thereby having to modify the structure of the piston. From the following figures it will be evident that both the cylinder 46 and the piston 47 are parallelepipeds, so the term â € œcylinderâ € used in the example favors the meaning assumed in the motoristic context rather than the geometric shape. The cylinder 46 is an internally hollow parallelepiped body, extended longitudinally, without a base to allow the insertion of the piston 47 which can slide while remaining in contact with the walls of the cylinder cavity by means of a system of sealing bands, better described in the subsequent figures. In this way a variable volume chamber is formed between the piston crown and the cylinder head with the reciprocating translational motion of the piston which will be first of compression and then of combustion. The edge of the base opening of the cylinder 46 extends outwards like a flange for fixing to the mouth of a crankcase 48. Two openings 49 and 50 are formed along opposite walls of the cylinder 46 at almost the same height . The opening 49 corresponds to the inlet port of the gaseous mixture in the combustion chamber The mixture reaches the inlet port 49 through a transfer duct 51 obtained in the same wall of the cylinder 46. The opening 50 corresponds to the flue gas vent port. This opening communicates with a short duct 53 made in the same wall of the cylinder 46 to convey the exhaust gases into a manifold (not shown) ending in a muffler, preferably of a so-called resonant shape to avoid the loss of unburned mixture. The walls of the cylinder 46 include air spaces 54 for the passage of the cooling water. The cylinder 46 with the crankcase 48 bolted rests on a frame 55. The crankcase 48 is divided into two separate compartments 57a and 57b, placed one under the other. The upper compartment 57a communicates with the cylinder 46, while the lower compartment 57b is a sump for collecting the lubricating oil circulated by a pump (not shown). The shape of the wall of the upper compartment is such as to delimit a curved pipe having an inlet 56 communicating with a central section open at the top towards the lower base of the cylinder 46, and a terminal section 52 for conveying towards the inlet of the transfer duct 51 of the mixture or of the air sucked into the crankcase through the port 56. One-way valves 58 of the reed type are applied to the inlet port 56.

The opposite walls of the cylinder 46 orthogonal to the walls where the intake ports 49 and exhaust ports 50 are present are crossed by the crankshaft 59 at a distance from the cylinder head 46 greater than the distance of the latter from the ports 49 and 50. The methods of making the cylinder 46 to allow such a crossing have already been mentioned. The cylinder head has a central threaded hole for screwing an electric spark plug 60, or alternatively a diesel injector.

The piston 47 is also parallelepiped in shape and internally hollow, it includes a head wall 64, a base wall 65, two opposite side walls 66, 67, and two other opposite side walls crossed by the drive shaft 59 in correspondence of a large circular opening (on a plane parallel to that of the section). This is instrumental to the fact that the piston 47 must be free to translate astride the crankshaft 59 which controls its movement. The four walls 64, 65, 66 and 67 have the same length, thus determining a piston whose longitudinal section orthogonal to the axis of the crankshaft is square in shape. The walls are equipped with a system of sealing strips which will be described in the following figures. The external surface of the head 64 of the piston 47 is the so-called â € œ skyâ €, hardened on the surface since it is destined to receive the violent thrust of the hot gas produced at each combustion.

Inside the piston 47 the mechanism that allows the hypocycloidal translation is visible in the longitudinal section. The mechanism is operated by the crankshaft 59 by means of a crank pin 68, which is in turn integral with a coaxial pinion 69. The distance between the axis of the pinion 69 and the axis of the crankshaft 59 is ̈ set equal to R. The toothed crown (schematized) of the pinion 69 meshes with a toothed crown with internal teeth 70 (schematized). The pitch circumference of the teeth of the pinion 69 has radius R, while the pitch circumference of the teeth of the sprocket 70 has a radius 2R. An eccentric 72 made in two modular parts is free to rotate on the axis of the pinion 69. The axis of the eccentric 72 is close to the center of the piston 47 while the axis of rotation coincides with the axis of the pinion 69. The eccentric 72 has a radius greater than the radius 2R of the ring gear 70 , having to rotate while remaining in contact with the internal faces of the piston 47 which are spaced apart more than the diameter of the aforesaid crown.

The elements inside the piston as well as the lateral faces of the cylinder in contact with the piston are subjected to forced lubrication (separated from the mixture of air and fuel) through a channel inside the crankshaft 59 which leads oil to the eccentric 72, and to the ring gear 70.

Figure 9A reproduces the section of figure 8 relating to the piston 47 alone, placing in the foreground the septum 71 within which the ring 70 with internal teeth is obtained in a central position which engages the crown (with external teeth) of the pinion 69. The number of teeth of the crown 70 is double that of the pinion 69.

Figure 9B frontally shows a square wall 95 of the piston 47 having a large central opening 100a which, together with a similar opening (100b) in the opposite face (not shown), allows the piston 47 to translate astride the Crankshaft 59. The figure highlights the system of oil containment bands included in each closed wall 64, 65, 66, 67. As you can see, the open walls resemble circular joints between edges converging in a vertex . In each side wall 66 and 67 orthogonal to the square walls there are included and partially protruding three containment bands arranged transversely to the longitudinal axis of the piston 47; of these an intermediate transverse lamina is closer to the center of the face while the other two are at the two ends. The transverse laminae protruding from the wall 67 are indicated with 74a, 74b, 74c; the transverse laminae protruding from the wall 66 are indicated by 74d, 74e, 74f. Along the two sides of the same walls 66 and 67 parallel to the longitudinal axis of the piston 47 are also included and partially protruding respective longitudinal containment bands wedged at the ends of the transverse laminae. The figure shows a longitudinal band 75a of the torque inherent in the wall 67 and a band 75c of the torque inherent in the wall 66. Along the two sides of each wall 64 and 65 perpendicular to the walls 66, 67 are also included and partially protruding respective side bands. The figure shows a longitudinal band 76a of the torque inherent in the wall 66 and a band 76b of the torque inherent in the wall 65. The forced lubrication of the open walls is schematized by drops of oil coming out of holes 77a visible in the upper part near the head and reentering in holes 77b visible at the bottom near the base. The bands 75a, 75b, 76a, 76b that surround the open face of the piston shown in the figure, and the corresponding other four bands of the other open face, constitute a set of perimeter bands that make the oil flowing out of the holes 77a translate in contact with the corresponding wall of the cylinder for the entire stroke of the piston and for the entire surface delimited by them. Figure 9C shows the four walls of the piston 47 orthogonal to the four sides of the two open walls overturned on the sheet plane, in order: the head wall 64, the base wall 65, and the two side walls 66 and 67. The four walls have the same rectangular shape and the same dimensions and the solid delimited by the set of these walls has a depth less than its height and length in the direction of the motor shaft axis. The side walls 66 and 67 have a series of holes 66b, 67b just below the transverse containment lamina in an intermediate position, and another series of holes 66c, 67c just above the transverse containment lamina placed in proximity to the base. The remaining portion 66a, 67a of the walls 66, 67 comprised between the intermediate transverse lamina and the transverse lamina placed in proximity to the head is devoid of holes. The lubrication of the walls 66 and 67 is schematised by drops of oil emerging from the holes 66b and reentering in the holes 66c. These openings allow the oil to exit the piston to lubricate the cylinder and then re-enter without accumulating and putting pressure on the affected area. They also allow the conveying of the burnt gases infiltrated through the upper bands into the sump. The upper portion 66a, 67a of the respective walls 66, 67 is an apparently lubrication-free zone (dry zone), and in fact the corresponding upper zone of the cylinder wall would also be free, but in reality this is not the case as lubrication takes place using a mechanism that will be illustrated later. Such a `` dry '' area in the upper part of the piston is necessary in the two-stroke engine of the example, without disposable lubrication, otherwise the oil carried on that portion of the face would contaminate the mixture, or the € ™ air, circulating in the transfer duct.

Figure 9D shows the assembly of the driving shaft 59 and the pinion 69 integral with it at the distance R between the two axes. The two elements are joined together by a crank pin (not visible) with which they form a single piece.

Figure 9E highlights only the crown with internal teeth 70 of diameter 4R obtained on the circular edge of a hole in the center of the internal wall constituting the septum 71. As can be seen in the figure, the diameter of the crown 70 allows it to be crossing by the assembly of figure 9D, thus avoiding having to break the drive shaft 59 or the septum 71 into two sections.

Figure 9F highlights the circular eccentric 72 with a circular hole 72c in an eccentric position for the insertion of a crank pin (visible in the following figure) around which it can rotate. Insertion is possible as the eccentric 72 is broken along the center line of the hole 72c into two unequal modular parts. Eccentric 72 has symmetry with respect to a diameter. A half part 72d delimited by a diameter orthogonal to the previous one is solid, while the other half part essentially reduced to a circular edge and a solid central race 72b that separates two empty quadrants 72a. The diameter of the eccentric 72 is equal to the distance between the opposite internal faces of the square section piston 47.

Figure 10 shows the longitudinal section along the plane A-A of the piston 47 of figure 8. The figure highlights the crossing of the cylinder 46 and the piston 47 by the crankshaft 59 which rests on two rolling bearings 61 constrained to the frame 55. Comparing with the front view of figure 8 it is possible to notice the smaller depth of the cylinder and piston with respect to their height. On the cross-sectional plane it is possible to note the almost absence of walls on the front and rear sides of the piston 47 in contact with the internal surface of the cylinder 46. The crankshaft 59 integrates in a single piece the crank pin 68, the pinion 69, and a collar 80 adjacent to the pinion 69. The crown gear of the pinion 69 is obtained by removing chips from one end of the crank pin having a larger diameter. The cylinder 46 is made in two parts that can be joined to allow such a crossing. The crankshaft 59 has two grooves spaced apart from each other more than the distance between the walls of the cylinder 46 for the insertion of two counterweights 81 and 82 which extend beyond the crankshaft 59 orthogonally to it on the opposite side with respect to the crank pin 68. Above the figure is a miniature representation of the crankshaft and relative counterweights. The two parts that make up the cylinder 46 each have two respective semicircular seats on the facing edges to accommodate two bushings 83 and 84 which support the crankshaft 59 while preserving the seal of the cylinder 46. A third bush 85 composed of two half-shells à It is placed astride the crank pin 68 to support the eccentric 72. The latter has a full 72d half part and the other lightened half part. The ring gear of the pinion 69 meshes with the internal toothed ring 70 having a double number of teeth. Top left in the figure shows a Diesel injector 62 which can replace spark plug 60, with or without a Root compressor. The arrows in the crankcase 48 indicate the direction of the mixture towards the transfer duct 51 when the reed valve 58 closes as soon as the piston 47 begins to descend. In the specific case, the displacement is 324 cc, the piston crown 47 has an area of 30 x 100 mm and a stroke of 100 mm.

Figure 11A is a top view of the transverse containment plate 74a representative of all the others: 74b, 74c, 74d, 74e, 74f. In the profile of the lamina 74a it is possible to note a rectangular portion contiguous to a trapezoidal portion, with two recesses 88, 89 at the corners between the two parts. Two transversely aligned rectangular windows 90, 91 are displaced with respect to the center laterally and towards the side of the trapezium opposite the base of the lamina. Figure 11B is a perspective view of the transverse containment lamina 74a showing its superposition on a second transverse containment lamina 74aa, completely identical to the previous one, held in position by two helical springs 92, 93 respectively housed in the rectangular windows 90 , 91. The two laminae are longitudinally misaligned since the second lamina is rotated by 180 ° with respect to the first in a longitudinal direction. The two spring configuration does not preclude the possibility of using a single spring.

Figure 12A is a longitudinal section along the plane B-B of figure 11B showing the spring 93 in the equilibrium configuration in which each end is in contact with both blades, and the same is true for the spring 92. In the configuration of equilibrium the spring 93 does not work and the plates 74a, 74aa have the maximum misalignment equal to the difference between the length of the two sections in contact; in this case the distance between the outermost ends of the two springs is greater than the distance between the walls of the cylinder 46.

Figure 12B shows the working configuration of the spring 93 compressed between the sheets 74a, 74aa in contact with the walls CL1, CL2 of the cylinder 46. As it can be seen, the distance between the outermost ends of the sheets has decreased by about half with respect to the maximum misalignment and consequently the shorter section of each lamina loses contact with the end of the spring due to the shortening suffered by it.

Figure 13 shows an exploded view that better highlights the complex system of plates that allow the lubrication of the surfaces in contact with the cylinder and piston and the containment inside the piston of the oil circulated forcibly to lubricate and cool the moving parts. . Each sheet shown in the figure is inserted in its own seat obtained in the corresponding wall of the piston 47, from which it can partially exit to come into contact with the wall of the cylinder 46 under the direct or indirect action of springs also inserted within seats inside the walls. The function of these plates has already been illustrated above. In the figure, two systems of four plates each can be distinguished, which once inserted in their seats form two perimeter bands around the open faces of the piston, and two groups of three plates each arranged transversely to the two remaining side walls. perimeter the foils 76a, 75a, 76b, 75b. The plates 76c, 75c, 76d, 75d belong to a second perimeter band. The laminae 74a, 74b, 74c belong to a first group of transverse laminae. The laminae 74d, 74e, 74f belong to a second group of transverse laminae. The sheets 75a, 75b, 75c, 75d belonging to the perimeter bands are narrow rectangular bars each having three rectangular slits 95 at the two ends and an intermediate point near the center; the slots are accessible from one side for the interlocking of the lamina within three similar slots of the type 88, 89 present in three transverse laminae. The seats which house the laminae 75a, 75b, 75c, 75d include a corrugated wire spring 96 acting on the lamina so as to push it in a direction orthogonal to the transverse laminae.

Figure 14 ideally recomposes the exploded view of figure 13 to form the frame that is present in the walls of the piston 47 of figure 15. Figure 14 highlights the twelve joints that connect the two perimeter bands to the two groups of transversal laminae, and the various springs that provide expansion capacity to the elements of the frame. Due to the joints, the transverse expansion operated by the helical springs 92, 93 pushes the perimeter bands against the square walls of the cylinder 46; similarly, the expansion operated by the undulated wire springs 96 pushes the transverse plates against the rectangular walls of the cylinder. Figure 15 shows the piston 47 complete with the frame of figure 14, highlighting the path of the oil along the external surfaces in relation to the various transverse plates and perimeter containment bands. In the figure it is possible to notice the large opening 100a in the square wall which, together with the opening 100b in the opposite face, allows the piston 47 to slide inside the cylinder 46 remaining astride the crankshaft. 59. These openings are framed by the respective perimeter bands that grease the walls of the cylinder with which they are in contact and then sweep the oil back into the openings. More generally, the various plates and bands force the oil to circulate within the boundary they define.

The sequence of figures from 16A to 16C reproposes for the hypocycloidal two-stroke engine (with separate lubrication from the mixture or air) the same instants of the operating cycle illustrated in the sequence of figures from 2 to 5 which concerned a two-stroke petrol engine of known type with disposable lubrication. The configuration assumed by the mechanics inside the piston 47 in each of the instants represented is better schematized in the sequence of figures from 18A to 18D, preceded by a more complete schematization in figure 17, and by figure 18 which shows the initial unbalance of the internal mechanics just beyond the TDC that establishes the direction of rotation of the crankshaft 59 under the translational thrust exerted by the burnt gases on the crown 44 of the piston 47. What is also interesting to show during a complete cycle of the piston stroke is the mechanism lubrication of the upper portion of the cylinder 46 in correspondence with the â € œdryâ € area of the piston. The oil drops visible in the sequence of figures 16A-16C indicate the lubrication of the portion below the â € œdryâ € area of the piston faces on both sides of the face shown in the figures. Starting arbitrarily from the TDC instant of figure 16A, it can be seen that the lubricating oil of the walls 66, 67 present in the areas below the portions of the wall 66a, 67a does not reach the upper part of the cylinder directly, since the intermediate transverse plate holds it below it, preventing it from entering the combustion chamber through the inlet ports 49 and exhaust ports 50 which open just above. Lubrication will therefore exploit an indirect mechanism which will be illustrated in the following figures. To avoid overheating and deformation of the walls 66 and 67 of the piston 46, the portions 66a, 67a are covered with a micrometric layer of silicon carbide which increases their surface hardness, thermal resistance, corrosion resistance, and decreases the coefficient of clutch. The head of the piston, equipped with the two upper transversal plates and with two other plates of the perimeter bands, blocks the way to the burnt gases. The base of the piston, equipped with the two lower transverse plates and two other plates of the perimeter bands, prevents oil from leaking into the crankcase 48. The surface of the cylinder 46 will be greased by the piston in the areas between the intermediate and lower transverse plate, as well as in the areas opposite the perimeter bands. Making further reference to the corresponding figure 18A for the numbering of the parts, it can be noted that at TDC the heavier half-part of the eccentric 72 occupies the position furthest from the crown of the piston 47.

The following figure 16B shows an intermediate instant of the stroke of the piston 47 towards the PMI due to the expansion of the burnt gases. The inlet and exhaust ports 49, 50 are still occluded by the piston. The reed valves are closed and the mixture present in the crankcase 48 undergoes a prestress. During the descent of the piston the â € œsiccâ € areas 66a and 67a of the walls 66 and 67 gradually come into contact with areas of the cylinder 46 which had been greased during the piston ascent and currently swept by the intermediate transverse plates. Despite this, a transfer of oil from the swept surface of cylinder 46 to these areas 66a and 67a without direct lubrication is made possible by a surface machining of the inner wall of the cylinder called â € œbrushingâ €. This process impresses on the smooth surface some inclined micro lines crossed to form a net which retains a small quantity of oil even after the oil scraper bands have passed. The metal that will subsequently come into contact with the surface thus machined may be greased by the trapped and released residual oil. The intermediate instant of the piston 47 descending stroke, in which it is lowered by D / 2, corresponds to a 90 ° clockwise rotation of the crankshaft 59 (see also figure 18B) and simultaneously to a 90 ° anticlockwise rotation ° of the circular exception 72. This rotation modifies the mass displacement of the eccentric 72, raising half of its heavier half-part above the median line of the previous dislocation, partially compensating for the inertial acceleration of the piston 47.

The following figure 16C shows the instant of reaching the PMI in which the piston 47 has completed its downward stroke and the `` dry '' areas 66a and 67a of the walls 66 and 67 have been entirely greased by the oil previously trapped by the swept surface of the cylinder 46. The inlet and exhaust ports 49, 50 are no longer blocked by the piston 47 and the transfer of the fresh charge can take place in concomitance with the so-called â € œflushingâ € of the cylinder chamber. The instant at BDC at the end of the descending stroke of length D corresponds to a clockwise rotation of a further 90 ° of the crankshaft 59 (see also figure 18C) and simultaneously to a counterclockwise rotation of a further 90 ° of the eccentric circular 72 which completes the raising of its heavier half part, partially compensating for the inertial acceleration of piston 47.

The following figure 16D last of the series shows an intermediate instant of the upward stroke of the piston 47 towards the TDC favored by the inertia of the flywheel. The inlet and exhaust ports 49, 50 are again occluded by the piston 47, which compresses the mixture entering the cylinder chamber and at the same time creates a vacuum in the crankcase 48 which opens the reed valves present at the inlet port 56 so that new mixture can enter the crankcase. During the ascent of the piston 47, the areas 66a and 67a of the walls 66 and 67 now entirely greased by the residual oil by means of the highlighted mechanism slide in contact with the surface of the cylinder 46, reducing the volume of the combustion chamber. The instant highlighted corresponds to a clockwise rotation of a further 90 ° of the crankshaft 59 and a corresponding rise of the piston 47 of D / 2 (see also figure 18D) and contextually to a counterclockwise rotation of a further 90 ° of the Circular eccentric 72, which involves a lowering beyond the median line of half of the heavier half part, partially compensating for the inertial acceleration of piston 47. The configuration assumed at the end of the last instant of the cycle coincides with the initial one of the Figure 16A and 18A, during this instant the piston 47 completes its stroke and therefore the compression of the combustible mixture within the combustion chamber, simultaneously creates a depression in the crankcase 48 with the intake of new mixture. A two-stroke endothermic engine with separate lubrication is now described, whose piston differs from that of the previous figures 8 and 9 due to the lack of the pinion and crown which, together with the eccentric, generated a 2: 1 hypocycloidal translational motion. The parallelepiped shape of the cylinder and piston remains unchanged, although the dimensions change, just as the frame of the plates and bands for the oil seal remains unchanged. The resulting piston is extremely simple, and can be described schematically in the sequence of figures from 19A to 19D depicting the configurations assumed by the internal mechanics of the piston in the instants of operation characterized by the following angles of rotation of the crankshaft: 0 < or> (360 °), 90 °, 180 °, 270 °. The endothermic cycle does not differ from the traditional one. Figure 19A is a longitudinal section of the piston at TDC taken according to a plane orthogonal to the crankshaft. It schematises a piston 105 free to translate in both directions astride the drive shaft 107 which crosses two opposite square walls. The crankshaft 107 includes, in a single piece, a crank pin 117 whose longitudinal axis is L from the axis of the crankshaft 107. The circular section of the crank pin 117 is externally tangent to the circular section of the Crankshaft 107. An eccentric 119 is inserted on the crank pin 117 whose geometric axis is L from the axis of the crank pin 117. The three axes pass through the centers of the three circles. The circular section of the eccentric 119 is tangent externally to the circular section of the crank pin 117. The condition of mutual tangency does not, however, limit the invention but is advantageous in reducing the size of the piston 105. The compatibility of the movement of the piston 105 with the kinematic constraints imposed by the cylinder 46 and the internal mechanics, appears in the following figures, where it can be seen that the eccentric rotates in the opposite direction with respect to the crank pin 117 and the crankshaft 107, and that at the PMI the stroke is equal to 2L. However, there is no symmetry in the positions reached by the piston during the instants indicated with the same angles of rotation of the crankshaft 107. In fact, passing from 0 <o> to 90 °, the piston 105 makes a longer journey than the one it makes when passing from 90 ° to 180 ° and this induces inertial components of the second order to balance the condition of zero speed at the dead points.

The translational movement of the piston 100 of figure 19A equipped with a single eccentric 119 is made symmetrical during the stroke by introducing a further thrust crank, as shown in the following figures relating to a variant which has no impact on the shape of the cylinder, and not even a system of transversal plates and oil containment bands present in the piston walls. The following figures all refer to the aforementioned variant.

Figure 20 shows the square face of the piston 105 made according to the variant. Figure 21 is a central section view along a plane orthogonal to the face of figure 20 and parallel to the narrower faces of the parallelepiped 105 piston. Referring to both figures 20 and 21 side by side, it can be seen that along the edges of the walls 64, 65, 66, 67 of the piston 105 there are the same transverse plates and perimeter bands for containing the oil visible in figure 9B. Said walls form a frame which allows a glimpse of a wide opening 106a for the introduction of the drive shaft 107 and of the elements that will be mentioned. A similar opening 106b has a beautiful opposite face. Inside the piston 105 there are two L-shaped tracks 108 and 109, specular with respect to the center line, fixed by screws 112 and 113 to respective protrusions towards the interior 110 and 111 of the head and base of the piston 105. The tracks 108 and 109 have the shorter side resting on its protrusion which runs along the entire length of the face 109, from one side wall to the other. A square-shaped slide 114 similar to a small frame can translate along said tracks, with one side in contact with the track 108 and the opposite side in contact with the track 109; the sides of the slide in contact with the rails have a longitudinal groove to accommodate the longitudinal edge of the respective rail during translation. The crankshaft 107 passes through the piston 105 supported by two main bushings 115 and 116 inserted in corresponding holes in the walls of the cylinder 46. The crankshaft 107 forms a sort of neck inside the piston 105. goose in a single piece, comprising a first crank 117 and a second crank 118 contiguous to the first. The two cranks are in practice offset sections of the drive shaft 107, therefore reduced to just the crank pins of different diameters. A circular eccentric 119 is free to rotate on the crank pin 117 while remaining in contact with the wall of the piston 10. The circular eccentric 119 has a lightened half-part of reduced thickness towards the outer circular edge. The edge of the crank pin 117 exceeds the remainder for a short distance starting from the part of the crank 118 and in contact with the edge of the eccentric 119.

The first crank pin 117 has a diameter of approximately 3/2 that of the crankshaft 107. The second crank pin 118 has a diameter approximately equal to that of the crankshaft 107. The cross sections of the three elements are circles with their centers aligned along a common axis, of which the circle corresponding to the crankshaft 107 is internally tangent to the circle of the first crank pin 117, while the circle corresponding to the second crank pin 118 is internally tangent to the circle corresponding to the first crank pin 117. The distance L of the axis of the crankshaft 107 from the axis of the first crank pin 117 is equal to the distance L of the axis of the latter from the axis of the latter. axis of the second crank pin 118. This allows an abundant sizing of both the crankshaft 107 and the first crank pin 117 for that given stroke of the piston 105. In practice it was possible to adopt a crankshaft diameter ug equal to five times the value of the eccentricity against the two of the solution with gears. Since the second crank pin 118 is aligned with the first crank pin 117 at the point of maximum extension beyond the crankshaft, there remains a 1/2 diameter high compartment of the crankshaft above and below the second crank pin 118 for the slide e114 d the relative tracks 108 and 109. Although the diameter of the eccentric 119 is approximately double that of the crankshaft 107, the area of the square face of the piston 105 does not increase excessively since the overall dimensions of the eccentric coincides for three quarters with that of its own pin 117. The two figures also highlight the channels for the forced circulation of the lubrication and cooling oil inside the walls of the cylinder 46 and of the crankshaft 107. More precisely , in the opposite walls of the cylinder 46 of figure 20 swept by the transverse containment laminae, respective ducts 120 and 121 are obtained for draining the oil collected by these laminae. In the opposite walls of the cylinder 46 of figure 21, swept by the perimeter bands, there are respective ducts 122 and 123 for the delivery of the pressurized oil coming from the pump, and of the respective ducts 124 and 125 for the drainage of the exhausted oil towards the sump 57b. Each delivery duct communicates with a hole present in a respective main bushing 115 and 116. The duct 123 continues beyond the bushing 115 in a duct 126 inside the drive shaft 107 in a radial direction, leading to a duct 127 in the axial direction. The axial duct 127 penetrates into the first and the second crank pin, 117 and 118, to branch into respective ducts 128 and 129 directed towards the circular edge of the same to lubricate the contact with the eccentric 119 and the contact of a shoe of the slide 114 with the track 113. From the axial duct 127 two other ducts branch off, aligned with the previous ones 128 and 129 but directed towards opposite points of the circular edge of the two respective crank pins to lubricate the contact with the eccentric 119 and the sliding block of the other shoe of the slide 114 with the rail 108. The oil brought to the eccentric 119 passes through two radial ducts in the wall of the same to lubricate the sliding contact against the internal head and base walls of the piston 47.

The configurations assumed by the elements inside the piston 105 in the four most significant instants of an operating cycle are schematized in the sequence of figures 22A to 22D. The instants considered correspond to the following angles of rotation of the motor shaft: 0 <o> (360 °), 90 °, 180 °, 270 °. The endothermic cycle does not differ from the traditional one. Figures 22A to 22D are aligned with figures 19A to 19D so that, since the arrangement of the common elements and the eccentricity value L are the same, a direct comparison between the two embodiments is possible. The sequence of figures 22A to 22D clarifies the contribution of the sliding slide 114 together with the piston 105 as it is rigidly connected to it. In figure 22A with the piston 105 at TDC, the pin 118 is in the center of the slide 11, the axis of the eccentric 119 is aligned with the axis of the first crank pin 117 and the second crank pin 118, so that no component orthogonal to the axis of the piston is possible to generate with the piston at TDC, which ideally should remain immobile as in any other type of reciprocating engine. However, the stall condition does not occur at all thanks to the flywheel and a phase shift in the ignition. The figure in question should therefore be imagined after a very short instant from exceeding the TDC in the direction indicated by the arrows. Figure 22B shows the internal configuration when the piston 105 has performed a translation towards the PMI of length 2L and the crankshaft 107 has performed a rotation of 90 °. It can be seen in the figure the rigid rotation of the gooseneck consisting of the elements 107, 117, 118 in one piece, and the 90 ° roto-translation of the eccentric 119 which resulted in the 2L lowering of the piston 105 and with of the slide 114. In the variant, the further kinematic constraint constituted by the slide 114 with the second crank pin 118 generates an additional moment on the motor shaft 107 of a sense congruent to the moment generated by the eccentric 119 alone. of the additional moment is that of synchronizing the translation of the piston with the rotation of the crankshaft, which will be evident in the following figures; in fact, the translation 2L of the piston 105 in figure 22B is greater than the translation found in the analogous figure 19B, and the difference is such that this translation corresponding to the 90 ° rotation of the crankshaft 107 can be expressed through integer multiples The above is equivalent to a synchronization which will be maintained at each 90 ° increment. Also in figure 22B it can be seen that while the piston 105 is translated towards the PMI of 2L, the second crank pin 118 is translated towards the outside within the slide 114 of 2L. The joint effect of the two translations along orthogonal directions is to make the drive shaft 107 perform a 90 ° rotation. The previous synchronization is maintained in the configuration of figure 22C where the crown of the piston 105 is at BDC 4L from the cylinder head, and the second crank pin 118 is shown in the center of the slide 114. The configuration shown in the following figure 22D is the one in which the piston 105 in the stroke towards the TDC has reduced the distance from the cylinder head by 2L, the crankshaft 107 is rotated by another 90 °, and the second crank pin 118 is translated within the slide 114 of 2L in the opposite direction to the previous one. The next instant completes the endothermic cycle by proposing the configuration of figure 22A. At the end of this instant the piston 105 has completed the stroke towards the TDC and the crankshaft a 360 ° turn, and the second crank pin 118 has returned to the center of the slide 114.

On the basis of the description provided for a preferred embodiment example, it is obvious that some changes can be introduced by the person skilled in the art without thereby departing from the scope of the invention as shown in the following claims.

Claims (11)

R I V E N D I C A Z I O N I 1. Internal combustion engine comprising: ∠’at least one hollow body, later called cylinder (46), whose cavity closed at one end of the head extends along a longitudinal axis (A-A; C-C); ∠'a hollow piston (47, 105) which can move within the cylinder in contact with the wall, delimiting a variable volume chamber in contrast to the cylinder head within which to cyclically trigger the combustion of a gaseous mixture of air and fuel and exploit the gases fuel to accelerate the piston; ∠’a mechanism for converting the translational motion of the piston into the rotary motion of the crankshaft (59, 107) and vice versa, characterized by the fact that: ∠'the said mechanism is contained in the piston (47, 105), the crankshaft (59, 107) crossing the cylinder wall (46) and the piston perpendicular to the longitudinal axis, the piston wall having openings opposed that allow it to translate astride the crankshaft during the entire stroke. 2. The engine of claim 1, characterized in that both the cylinder (46) and the piston (47, 105) have a parallelepiped shape with the faces crossed by the crankshaft (59, 107) almost square, the piston being closed by a base wall (65). 3. The engine of claim 2, characterized in that said mechanism includes: ∠'a pinion (69) integral and coaxial to a crank pin (68) of the crankshaft (59), the pitch circumference of the pinion teeth having a radius equal to the distance between the pinion axis and the Axis of the drive shaft (59); ∠'a toothed crown (70) obtained in a septum (71) integral with the piston (47), arranged so that the internal teeth of the crown can mesh with the teeth of the pinion (69), the radius of the primitive circumference of the toothed crown internal (70) being double the radius of the pitch circumference of the pinion teeth (69); ∠’a circular eccentric (72) free to rotate on the crank pin (68) remaining in contact with the wall of the piston (47) during the entire stroke so as to keep the pinion constantly engaged with the crown. 4. The engine of claim 3, characterized by the fact that the circular eccentric (72) is a wheel or disc with its own half-part (72d) of mass comparable to the mass of the piston (47), said mass rising and lowering due to the effect of the eccentric rotation in synchrony with the reciprocating translation of the piston, but in the opposite direction, it balances the inertial forces caused by the acceleration of the piston. 5. The engine of claim 2, characterized in that said mechanism inside the piston (105) includes at least a first crank pin (117) on which a circular eccentric (119) is free to rotate while remaining in contact with the wall of the piston during the entire stroke, the eccentric (119) having an eccentricity value equal to the distance of the axis of the crank pin (117) from the axis of the crankshaft (107). 6. The engine of claim 5 characterized in that said internal piston mechanism further includes: ∠'a second crank pin (118) with a diameter smaller than the diameter of the first crank pin (117), the axis of the second crank pin being as far away from the axis of the first crank pin as the first crank pin crank is located from the axis of the crankshaft (107); ∠’a double slide (114) similar to a frame within which the second crank pin (118) can rotate while remaining in contact with the walls of the slide; ∠'two pairs of tracks (108, 109), each pair integral with a respective wall of the piston crossed by the motor shaft (107), the tracks having the same length as the application wall and being arranged orthogonally to the translation axis piston (105), mutually spaced to accommodate the slide (114) and guide it during the translation in both directions performed at each revolution of the crankshaft (107). 7. The engine according to any one of claims 2 to 6, characterized in that: ∠'the walls of the sliding faces of the piston have outlet holes (66b, 67b, 77a) for the lubricating oil circulated under pressure inside the piston, and inlet holes (66c, 67c) for the € ™ oil collected from the walls; ∠'the wall of each sliding face of the piston has grooves for the insertion of laminae (74a-e, 75a-d) coupled to resilient means (92, 93, 95) capable of pushing them against the cylinder wall to distribute the lubricating oil and then sweep it; ∠’of the first sheets (75a-d) being arranged along the sides of the walls crossed by the drive shaft, to form perimeter bands; ∠’of the second sheets (74a-e) being arranged transversely to the adjacent walls and including seats (88, 89) on the sides for interlocking with the perimeter bands (75a-d) 8. The engine of claim 7, characterized in that the said transverse sheets are in pairs of superimposed sheets (74a, 74aa), each having a rectangular window (90, 91) arranged longitudinally in an asymmetrical position with respect to the center, offset by a lamina on the other, for the inclusion of resilient means (92, 93) able to exert pressure against the misaligned sides of the windows making the laminae translate in opposite directions against the perimeter bands (75a-d). 9. The engine of claim 8, characterized in that the resilient means (96) coupled to the perimeter bands (75a-d) exert a pressure in an orthogonal direction on the longer sides of the transverse laminae (74a-e), transmitted to the latter from the joints. 10. The engine of claim 9, characterized in that it is a two-stroke engine with separate fuel lubrication and inlet (49) and exhaust (50) ports at the same level of opposite cylinder walls (47, 105), the corresponding walls (66, 67) of the piston each including three transversal plates (74a-c; 74d-f), two of which are located at the ends respectively upper and lower than the piston crown and the third intermediate , and the said oil outlet and inlet holes (66b, 67b being included between the intermediate lamina and the lower lamina, the intermediate lamina being more distant from the upper lamina than the length of the stroke, being located beyond the Inlet and exhaust (49, 50) when the piston is at top dead center, thereby preventing oil from entering the combustion chamber. 11. The engine of claim 10, characterized in that the portion (66a) of each piston wall comprised between the intermediate transverse foil and the upper transverse foil is surface hardened, and the walls of the cylinder (46) are furrowed by micro crossed lines capable of retaining a certain quantity of oil even after having been swept by the transversal blades rising towards the top dead center, then yielded to the surface hardened area and from there to the cylinder during the descent to the bottom dead center.