Field of application of the invention

The present invention concerns the automotive sector and in particular an endothermic reciprocating engine with a motion-converting mechanism inside the piston.

Review of the known art

Mention should be made of the fact that arrival of the endothermic engine marked a turning point in technical progress for the whole of humanity. Further evolution in this field has produced internal combustion engines of many types, but narrowing the field of invention to that of interest here, these engines may be grouped into two main categories: reciprocating engines and rotary engines. The more conventional automobile engines with cylinders and pistons belong to the first category, those operated by pressurised gas during the rapid combustion of an air-fuel mixture compressed in the cylinder chamber, activating a piston joined to a crank mechanism of thrust that rotates a drive shaft. Continuous-rotation engines belong to the second category such as gas turbines, not described in this case as they do not concern the invention, and those known as rotary piston engines such as the Wankel engine that will be discussed. In reciprocating engines each the piston makes a complete cycle in one or two revolutions of the drive shaft; combustion can be spontaneous as in the Diesel engines, or provoked by sparks between the electrodes of a spark plug at the end of a stage of compression. In the four-stroke engines a suitably synchronised valve device opens and closes the openings in the top of the combustion chamber while, in two-stroke engines, piston movement itself closes the efflux and defluxion ducts present in the lateral wall of the cylinder.

For a clearer understanding of the invention to be described and of the problems existing in the known art, it will be useful to give a more detailed account of the two types of engines referred to above.

Figure 1 is a diagram showing a longitudinal section of a Wankel engine 1. The figure shows a stator 2 with a rotor 3 inside it. The internal profile of the stator is a two-lobed epitrochoidal curve corresponding to that of an epicycloidal nephroid. The lateral wall of the stator 2 is extruded from the geometrical profile in the figure, as is also the body of the rotor 3. The drive shaft passes through the two bases that close the stator. A pinion 4 is constrained to the centre of a base of the stator 2 inside it. The external profile of the rotor 3 is that of an equilateral triangle with convex sides (Rouleaux triangle). Inside the rotor 3 there is a crown gear 5 that engages with the pinion 4, the diameter of the crown 5 is double that of the external gear. A drive shaft 6 freely rotates inside the pinion 4. A circular cam 7 is fitted onto the drive shaft 6, free to rotate inside a large bushing 8 fixed in a central position to the rotor 3. The mouths of two ducts 9, 10, respectively for efflux and defluxion, open at one side of the stator 2. On the opposite side two spark plugs 1 1, 12 partially pass through the stator wall in correspondence of a combustion chamber 13a, 13b.

From the operative standpoint, the Wankel engine, like the two-stroke engine, has no distribution valves; it is in fact the rotor 3 that cyclically opens efflux mouth 9 and closes defluxion mouth 10. Following the direction of rotation indicated by the arrow, the suction stage of the air-fuel mixture begins when the volume of the chamber into which the suction duct 9 starts increasing, so creating a depression that attracts the mixture into the stator chamber. Continuing its movement, the rotor 3 reduces the space comprised between its wall and that of the stator so compressing the mixture that has been sucked in. When compression has reached the optimum level, the spark plug electrodes emit a spark so initiating the phase of combustion and therefore expansion of the gas. The resulting increase in pressure causes forces to act on the rotor constraining it to continue its rotational motion. The function of the stator pinion is to constrain the rotor to execute an eccentric orbit and ensure constant contact between the apical seal elements and the internal walls of the stator. The cam 7 acts as a crank rotated by the translational component inherent in eccentric rotation of the rotor 3, transforming it into a rotational component transferred to the drive shaft 6. On conclusion of the phase of expansion there is a further reduction of volume during which the burnt gas is expelled through the defluxion duct 10. From the kinematic standpoint, the Wankel 3 rotor executes a sun-and-planet motion, meaning that it is mobile around an unfixed axis, this in turn rotating round the fixed axis of the drive shaft 6.

The Wankel is a simply designed engine consisting as it does of only two moving parts, the rotor and the drive shaft. The axis of symmetry of the rotating piston 3 turns around the axis of the stator 2 producing an eccentric mass to which a centrifugal inertial force is applied. No problems arise over balancing this force because there are no inertial forces of the second order which are typical with reciprocating engines. In spite of its kinematic simplicity the Wankel engine has not been able to compete with the reciprocating engine in the automobile field due to constructional complications and to a higher consumption of fuel. Apart from this the main drawbacks consist in the short life of the rotor sealing elements, in the rather low torque at low rotating speeds, in problems over lubricating the apical segments and in the high rate of unburnt hydrocarbons. A reciprocating two-stroke engine free of these problems is shown in the following Figures 2, 3, 4, 5 that explain how it functions in four configurations assumed during rotation of the drive shaft at 360°. The expression "two-stroke" is used because a complete Otto cycle of functions uses only one period of upward movement of the piston and two downward, corresponding to the above rotation at 360°. Regarding transfer of power to the drive shaft, this is the same as for the Wankel engine in which there are three active stages in every three revolutions made by the drive shaft.

Referring to Figure 2, this shows a two-stroke reciprocating engine 20 consisting of a cylinder 21 in which a sliding piston 22, fitted with gas rings (thick line), has reached the top dead centre (PMS). In a position such as this a combustion chamber 23 is created between the crown of the piston 22 and the internal wall at the top of the cylinder 21. Chamber 23 is filled with a mixture of air and fuel, highly compressed during upward movement of the piston 22. A spark plug crosses the top of the cylinder 21 to provoke combustion when there is a spark from its electrodes projecting inside the loading chamber 23. The piston 22 is hollow and contains the upper end (foot) of a connecting rod 25 hinged to a pin 26 fixed to the piston 22. The lower end (head) of the connecting rod 25 is hinged to the pin 27 of a crank 28 fixed to the drive shaft 29. The shape of crank 28 is suitable for balancing its own rotational mass. The wall of cylinder 21 is contiguous to the wall of a sump-pump 30 inside which is the crank 28 and most of the connecting rod 25. A suction duct 31 is open in the lower part of the sump 30 for entry of the air-fuel mixture. At its upper dead centre (PMS) the piston occludes an opening 32 in the wall of cylinder 21 at the terminal point of a duct 33 bringing the air-fuel mixture, previously sucked into the sump-pump 30 by upward movement of the piston 22. In the same way, piston 22 occludes an opening 34 in the wall of cylinder 21, on the side opposite opening 32 and aligned with it, at the starting point of an outflow duct 34 for burnt gas. The figure does not show a one-way slat valve at the entry to the sump-pump 30, that opens automatically when a depression forms in the sump, then closes again when atmospheric pressure is re-established. This valve consists of a support, usually triangular, with openings in which elastic "slats" of steel, carbon fibre or glass fibre, some tens of a millimetre thick are laid. The slat valve enables the suction timing to be automatically suited to the needs of the engine.

Figure 3 shows the position reached by the piston 22 while the burnt gas is expanding, after a quarter turn has been made by the drive shaft 29 in the direction indicated by the arrow on the crank 28. In that position the inflow 33 and outflow 34 openings are still closed but this is not obligatory as the distance of said openings from the top of the cylinder 21 is established by phase times decided in the design stage. During its downward movement the piston 22 pre- compresses the air-fuel mixture previously sucked into the sump-pump 30 while rising. The piston 22 during its descent provides the drive shaft 29 with an active torque by means of the connecting rod 25 and the crank 28. Figure 4 shows the position reached by the piston 22 at the bottom dead centre (PMI). As little by little the piston 22 moving to the PMI causes openings 32 and 33 to appear, fresh mixture enters the cylinder 21 through duct 33 and the burnt gas is discharged through duct 34. This is also known as the 'washing' stage as the fresh mixture helps to expel the burnt gas that is partially dispersed through the drain, unless special precautions are taken to prevent it. These precautions consist in suitably shaping the crown of the piston, but above all in adopting an efficient resonant discharge system able to assist extraction of the burnt gas and generate a retrograde wave of pressure close to the outlet mouth of the silencer that returns the unburnt mixture into the cylinder.

Figure 5 shows the position reached by the piston 22 after three quarters of a revolution has been made by the drive shaft 29. During its inertial upward movement the piston 22 closes openings 32 and 33 and begins to compress the air-fuel mixture trapped in the upper part of the cylinder 21. A new load is simultaneously drawn into the sump-pump 30 through the suction duct 31 due to the depression created during upward movement of the piston 22.

Operatively, when the piston 22 reaches the top (PMS) and bottom (PMI) dead centres its speed is zeroed prior to reversal. Stalling is avoided thanks to the flywheel formed of a circular expansion of the crank 28 that, because of accumulated kinetic energy, carries the piston beyond the dead centre. The connecting rod 25 that joins the piston 22 to the crank 28 makes a roto- translational movement transforming the rectilinear motion of the piston into the rotary movement of the crank and therefore of the drive shaft 29.

In two-stroke engines with a dry sump functioning as a sump-pump, mainly used in some types of motorcycles and in the tools for a wide variety of outdoor activities, lubrication of the cylinder walls in contact with the piston, as also of the crank mechanisms inside the sump, is mostly done by diffusing a percentage of lubricating oil in the air-fuel mixture sucked into the sump-pump and then passed to the transfer opening. This total loss lubrication simplifies considerably the engine though somewhat polluting. Other known engines are the two-stroke with separate lubrication, such as the Diesel with a Root (lobe-type) compressor for pumping the cleaning air into a seat open in the cylinder chamber near the bottom dead centre (PMI). In the four-stroke engines lubrication of the cylinder walls in contact with the piston, as for the crank mechanism in the sump, is effected by causing oil to circulate under pressure in a positive-displacement pump through channels in the drive shaft and in the connecting rod; after that the lubricating oil fall down for gravity into an oil sump from where it is picked up by the pump. This type, called damp-sump lubrication, is easy to realize but has certain well-known drawbacks that it would be superfluous to describe. The more sophisticated four-stroke engines use a small dry sump, and at least one other pump to recover the oil that deposits there and immediately put it back into circulation.

Technical problems

The reciprocating engine differs both mechanically and dynamically from the Wankel rotary engine; unlike the Wankel, during the piston's reciprocating movement additional inertial forces of the second order, similarly to those of the first order, act on the mechanical constraints, particularly on the main journals of the drive shaft. To put it more precisely, said ω angular speed of rotation of the drive shaft, mathematical expression of the inertial forces, highlights first sinusoidal components at ω pulsation, and second components at 2ω pulsation, respectively called inertial forces of the first and second order (though the term inertial moments would be more appropriate). The two components are additives and, though not perfectly balanced, induce abnormal stresses in the main journals; these can deform the drive shaft and spread through the frame in the form of noisy vibrations that cause wear on the moving mechanical parts increasing the wear and shorten their working lives. It is therefore of fundamental importance to neutralize these dynamic forces so that only the piston and related mechanisms weigh on the main journals. Neutralization generally consists in balancing the moments produced by the various moving parts in relation to the axis of the drive shaft. Balancing consists in adding eccentric masses to the rotating parts; it does not create serious problems over neutralizing the inertial forces of the first order as the movement of parts involved in thrust mechanisms is cyclic to the ω pulsation, but the same does not happen when neutralizing the forces of the second order that require countershafts rotating at an angular speed of -2co. Difficulties over neutralization of the inertial forces usually lessen with an increase in the number of cylinders because a particular setup can be chosen for them along the drive shaft together with a suitable sequence of ignition phases to compensate the different inertial contributions one with another.

Another problem inherent in reciprocating engines is their bulk along the longitudinal axis of the cylinder. This fixed bulk can immediately be seen from the bottom dead centre (PMI) configuration in Figure 2, roughly equal to the sum of these contributions: a) height of the piston 22; b) length of the connecting rod 25 less the section inside the piston; c) about double the arm of the crank 28. Making use of shorter connecting rods cannot significantly reduce a bulk such as this for the following reasons: a) the diameter of the cylinder would have to be increased to an equal extent to contain the greater inclination of the connecting rod when the crank is horizontal; b) as will shortly be explained, the shorter connecting rod involves an increase in the inertial forces of the second order, harder to neutralize, so that a compromise would have to be made between less mechanical bulk and reduced inertial forces.

The contribution by the connecting rod to the inertial forces is usually divided into two, a first of these purely translational, is a concentrated mass located in the foot of the connecting rod on the side of the piston pin; while a second purely rotary contribution, is that of a concentrated mass located in the head of the connecting rod on the crank side. Knowing the length of the connecting rod and its mass, the two masses together are calculated by cancelling the respective moments in relation to the centre of gravity. The first contribution is added to the inertial forces of the first and second orders generated by the piston, while the second contribution is added to the inertial forces generated by the crank. As the movement made by the crank is purely rotary, the total inertial moment, including the contribution by the connecting rod, contains only inertial components of the first order that are easily balanced.

The inertial forces F _\ , _p of the foot of the connecting rod of mass that translates together with the piston at acceleration a, are given by the expression: Fbp - -w-bv ' ^a - ^~m bv ' cos or * if - ω ¹ -rn _bp ^■ ! · cos 2c? · R - eo ^s (i ) wherein:

• a is the angle between the axis of the cylinder and the axis of the crank at instant t ;

· ω = ά is the angular pulsation of the crank;

• R is the radius of the crank;

R

, =

L where L is the length of the connecting rod. The value of parameter λ is generally low; in the example: R = 15 mm and L = 58 mm gives a = 0.2586. The inertial forces of the first order give: ^~m p ^{* COs} « - R ^■ ω ³

The inertial forces of the second order give: ^~m bp · λ ^■ cos 2a ^■ R - ω" _{mat? ω} angular pulsations .being equal, shows a double-pulsation cyclic trend. The λ parameter, that appears with multiplicative value in the expression of inertial forces of the second order only, proves that said forces in fact diminish with an increase in the length of the connecting rod in relation to the radius of the crank, but at the same time increases the bulk of the engine in the longitudinal axis of the cylinder.

The technique of reciprocating endothermic engines uses connecting rods joined to a 2: 1 hypocycloidal crank mechanism able to impress a linear motion on the connecting rod to keep it aligned with the axis of the cylinder and exclude any creation of inertial forces of the second order in the rod itself. The connecting rods used in such examples do not alter their conventional length, not even the type of connection to the piston, so that there is neither mechanical simplification nor any reduction of bulk.

Purposes of the invention

Purposes of the present invention are to reduce the inertial forces generated in endothermic reciprocating engines, especially by the connecting rods, at the same time simplifying the mechanics and reducing bulk.

Further purpose of the invention applied to a separate lubrication two-stroke endothermic engine, is that to make reliable the lubrication of the sliding parts and other moving elements without contaminating the combustive mixture. Summary of the invention

To achieve these purposes, subject of the present invention is an endothermic reciprocating engine comprising:

at least one hollow body, hereinafter called a cylinder, whose cavity, closed at its upper end, extends along a longitudinal axis;

a hollow piston able to translate inside the cylinder in contact with the wall delimiting, in opposition to the cylinder head, a chamber of variable volume in which combustion is cyclically effected of a gaseous mixture of air and fuel, exploiting the burnt gas to accelerate the piston;

- a mechanism to convert the translational motion of the piston into a rotating motion of the drive shaft and vice versa,

wherein, according to the invention:

said mechanism is contained in the piston, the drive shaft passing through the wall of the cylinder and of the piston orthogonally to said longitudinal axis, there being opposite apertures in the wall of the piston enabling it to translate across the drive shaft throughout its stroke, as described in claim 1.

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

According to a main aspect of the invention applied to a separate lubrication two-stroke endothermic engine, having inflow and outflow openings at the same quote of opposing walls of the cylinder, both cylinder and piston are parallelepipeds, the faces through which the drive shaft passes being substantially square, the piston being closed by a bottom wall, so to make reliable the lubrication of the mutually sliding walls of cylinder and pistons and the moving parts inside the piston, as the cooling of them.

In accordance with a first embodiment of the invention, said mechanism for converting the translational motion of the piston into a rotary motion of the drive shaft includes:

a pinion fixed to and coaxial with a crank pin of the drive shaft, a radius of the pitch circle of the pinion teeth being equal to the distance between the axis of the pinion and the axis of the drive shaft; an internal gear inside a baffle fixed to the piston, so placed that the internal teeth of the gear can engage with the teeth of the pinion, the radius of the pitch circle of the internal gear being double the radius of the pitch circle of the teeth of the pinion;

- a circular cam free to rotate eccentrically on the crank pin maintaining contact with the wall of the piston throughout its stroke to keep the pinion constantly engaged with the internal gear.

The first 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 rotates inside the fixed one. With that ratio, a point on the mobile circumference covers a diameter (roulette) of the fixed circumference in both directions when a complete turn is made. The novelty consists in the fact that the internal gear (representing the fixed circumference from the geometrical standpoint) would be constrained to make alternating translations along its diameter, obliging the pinion, aided by a cam, to roll inside the internal gear and causing a crank to rotate. In practice, the internal gear is made in an internal wall of the piston similar to a baffle, while the cam, fitted with anti-friction metal annular bands, can rotate in a seat beyond the baffle of the internal gear. Top and bottom dead centre (PMS → PMI) translation is controlled by the burnt gas while bottom and top dead centre (PMI→ PMS) translation is controlled by a flywheel rotated by the drive shaft, as happens in conventional engines. When piston translation is from PMS to PMI the piston wall exerts pressure on the edge of the circular cam constraining it to rotate round the crank pin in one direction or the other (according to its initial position), the eccentricity making up for the difference in distance which would otherwise be created between pinion and internal gear during translation of the piston; in this way the pinion remains constantly engaged in the internal gear.

It should be remembered that the known examples of using hypocycloidal geometry 2: 1 differ somewhat from the one described, mainly because the crank mechanism adopted in those examples is outside the piston.

According to one aspect of the invention, the circular cam is a wheel or a disk a part of which is a mass comparable to the mass of the piston; said mass, rising and falling due to eccentric rotation in synchrony with alternating translation of the piston but in the opposite direction, balances the inertial forces produced by acceleration of the piston .

With regard to how the piston is made, the walls through which the drive shaft passes should have ample openings in them to allow for insertion of a gooseneck made in a single piece. In this way the pinion teeth can be cut by chip machining from one end of diameter beyond the crank pin. The cam can be made in two parts, each with a semicircular half-seat to be put in contact with the crank pin. The half-couplings that contain the half-seats will be bolted together. As regards the cylinder, the drive shaft must be able to pass in different ways through this too, for example by making the cylinder in two parts joinable with respective half-seats for the bushings.

According to one aspect of the invention:

the walls of the sliding faces of the piston contain outlet holes for the lubricating oil circulating under pressure inside the piston, and inlet holes for oil collected from the walls;

the wall of each sliding face of the piston contains grooves for insertion of laminas coupled to resilient means able to press them against the wall of the cylinder to spread lubricating oil and then remove it;

- first laminas disposed along the sides of the walls through which the drive shaft passes, to form perimetric bands;

second laminas disposed transversally to the adjacent walls and including seats at the sides for fixing the perimetric bands.

According to one aspect of the invention, the transversal laminas are in pairs of superimposed laminas, in each of which is a rectangular window placed longitudinally in an asymmetrical position in relation to centre, offset between one lamina and another, to include resilient means able to exert pressure against the misaligned sides of the windows causing the laminas to translate in opposite directions against the perimetric bands.

According to one aspect of the invention, the resilient means coupled to the perimetric bands exert orthogonal pressure on the longest sides of the transversal laminas, said pressure being transmitted to the laminas due to the fact that they are fixed.

According to one aspect of the invention, referred to a two-stroke endothermic engine with lubrication separated from fuel and having inflow and outflow openings at the same level in opposite walls of the cylinder, the corresponding walls of the piston each including three transversal laminas, two of which are respectively placed at the upper and lower ends in relation to the crown of the piston, the third being intermediate, said output and input holes for the oil being comprised between the intermediate and lower laminas, the intermediate lamina being spaced from the upper lamina at a distance longer than the stroke, situated beyond the inflow and outflow openings when the piston is at the top dead centre position, thus preventing entry of oil into the combustion chamber.

According to one aspect of the invention relating to said two-stroke engine, the part of each piston wall comprised between the intermediate transversal lamina and the upper transversal lamina is surface-hardened, the walls of the cylinder being grooved by crossed micro lines able to retain a certain quantity of oil even after cleaning by the transversal laminas when rising towards the top dead centre, said quantity of oil being then released to the surface-hardened area and from there to the cylinder during downward movement to the bottom dead centre. The mechanism inside the piston for converting the alternating motion to a rotary motion, and vice versa, is simple to make and possesses several advantages, although there might be too much piston displacement in relation to the stroke especially if the gooseneck inside the piston were of maximum size. This could happen because both the area of the square section of the piston and the stroke are closely linked to the radius of the pinion's pitch circle. If the gooseneck is of minimum size where the drive shaft radius R is the same as that of the pinion, the stroke and side of the square section of the piston are respectively equal to 4R and 6R, developing a displacement of 24R d where d is the depth of the piston. To limit piston displacement and bring it back to correct dimensional proportions, the d parameter should be reduced and therefore the depth of the piston, but this would penalise the area of the washing and discharge openings in the two-stroke engine. The problem raised here is solved by a second realized example of the invention wherein the mechanism inside the piston for transforming the motion does not exploit the coupling between crown gear wheels but only interference by the piston wall on at least one cam coupled to a crank, as for that matter already happens in the piston described, though only for keeping the pinion engaged with the internal gear.

Therefore in accordance with a second embodiment of the invention, the above mechanism inside the piston includes at least one first crank pin on which a circular cam is free to rotate eccentrically remaining in contact with the piston wall throughout its stroke, the value of eccentricity of the circular cam being equal to the distance between the axis of the crank pin and the axis of the drive shaft.

The second embodiment describes the simplest motion converting mechanism internal to the piston; this is not ideal as regards symmetry of accelerations acting along the stroke and does not therefore ensure absence of inertial forces of the second order. Piston displacement being the same, the advantages easily outweigh the drawbacks, it being possible to produce pistons that can operate with larger drive shafts and crank pins in relation to piston stroke compared with the hypocycloidal 2:1 pistons and gears. Another version will attempt to remedy the drawback described but at the cost of creating mechanical complications inside the piston.

It will be appreciated how the motion-converting mechanism inside the piston, constrains it to translate across the drive shaft, remains unaltered.

The example of a realization based on the cam differs considerably from the conventional reciprocating engines where piston and crank are joined by the connecting rod, universally known in the engine sector as a rod in which there is a circular seat at each end to house two pins: the pin fixed to the piston and the crank pin fixed to the drive shaft. The conventional connecting rod makes a roto- translational movement of limited angular amplitude and of ample translation that involves the problems already referred to; its particular form also means that it lies almost completely outside the piston, as of course does the crank. Conversely, the cam inside the piston makes a roto-translational movement, the rotational component prevailing over the translational (a complete rotation at each translation equal to double the eccentricity) so that the ideal condition of absence of inertial components of the second order comes close to achievement. For a clearer understanding of the second example of realization of the invention it is worth considering a system of Cartesian reference in a plane transversal to the axis of the drive shaft, in which the axis of the ordinates coincides with the axis of the piston. In this reference, rotation of the x(t),y(t) generic point belonging to the geometrical axis of the crank pin is described by the time functions: (Π = Λί cos or ; y(t) = M sin or , that define a circumference where the length of the crank arm M is the same as the distance between the axis of the relative pin and the axis of the drive shaft; ^a = ^ω£ and ω is the angular speed of the drive shaft. The inner surface element of the piston head in contact with the edge of the circular cam transmits a longitudinal force to said cam; this force has an arm in relation to the axis of rotation of the cam and can therefore generate a moment that initiates an infinitesimal rotation ^Acf of the cam round the crank pin. Being unsymmetrical, an elementary rotation of the cam exerts pressure against the lateral wall of the piston, whose orthogonal reaction on the edge of the cam passes to the crank pin. The resultant combination of forces acting on the crank pin has two variable components along the Cartesian axes, that also involve two respective translations subjected to the rigid constrained of the crank pin with the drive shaft. This constraint, together with the dimensional choices made, involves the rotation of the crank arm round the axis of the drive shaft caused by movement of the piston. The cam in fact rotates round a moving axis that in turn rotates round a fixed axis. Such infinitesimal movements produce the macroscopic effect observable in each geometrical diagram characterized by its own particular ratio between length of the crank arm and eccentricity of the circular cam. Geometrical diagrams characterized by value 2 allow the cam to make a sun-and-planet motion, other values may not always be compatible with cyclic movement of the piston. The constraint imposed by a fixed distance between the piston walls is fixed, along with other geometrical constraints, permits that eccentric rotation can be completed if it is the drive shaft that provides the moment of thrust to be transformed into translation of the piston. It has been seen that to create the rotational movement of the drive shaft, at any instant in the working cycle and starting from a one-way propulsive force, it will be necessary to produce a component of the force orthogonal to that way and alternating in direction. This is because the point where the force is applied rotates on the circumference that the crank pin must cover. In the piston included in the first embodiment, the orthogonal component is created by continuous roll of the pinion on the internal gear. Meshing between the two mediated by the cam creates a particularly strong mechanical constraint able to transfer the necessary power to the orthogonal component. In the piston of the second embodiment, the orthogonal component is created only by the geometrical constraints between wall and cam and between these and the crank pin, but these constraints are not so direct as meshing and, further, are more subject to friction caused by the cam sliding against the piston wall, as usually happens in a bush coupling.

The problem therefore is how more effectively to create a component of force orthogonal to the direction of translation of the piston if there is no 'direct' mechanical constraint of the meshing type between piston and crank pin, or of the rod type with two pivoting ends as in conventional connecting rods. More generally speaking, the problem is how to use cams to convert the translational movement of the piston into a rotary movement of the drive shaft, and vice versa, with uniform and symmetrical accelerations along the entire stroke so as to minimise, if not cancel, any inertial forces of the second order.

This problem is solved by a different version of the second embodiment of the invention, wherein the internal mechanism of the piston also includes:

a second crank pin of a diameter lesser than that of the first crank pin, the distance of the axis of the second crank pin from the axis of the first crank pin being the same as the distance of the first crank pin from the axis of the crank shaft;

- a double slide similar to a frame in which the second crank pin can rotate while remaining in contact with the walls of the slide; - two pairs of rails, each pair being fixed to a respective piston wall through which the drive shaft passes, the rails being of a length equal to said respective wall and disposed orthogonally to the axis of translation of the piston, reciprocally spaced to receive the slide and to guide it during the translation made in both directions, at each turn of the drive shaft.

In practice the cam develops the alternate rectilinear movement of the piston along the axis of the cylinder, while the second crank pin coupled to the slide aids development of the orthogonal alternate rectilinear movement in the direction appropriate to piston movement. This solves the announced technical problem and helps to overcome the dead centre positions of the stroke in proximity of which the orthogonal component of the force is at its lowest value. The different configuration of the variant has some salient characteristics in common with the alternate rectilinear hypocycloidal function, such as the fact that piston stroke equals four times the eccentricity and that, at every 90° rotation of the drive shaft, the piston makes a translation equal to the value of eccentricity. The symmetry of movement required to achieve such a result is evident.

Advantages of the invention

The present invention can be particularly appreciated in a separate lubrication two-stroke endothermic engine, to say total loss lubrication is not used. This is mainly due to the parallelepiped shape of piston that, thanks just to this shape, differently from the cylindrical shape allows to separate the sliding closed walls from the sliding ones crossed by the drive shaft, hence to be able to apply the whole system of perimetric and transversal elastic bands which is suitable to grant the oil seal inside the boundary swept by the piston, simplifying the forced lubrication of the moving elements inside the piston for an efficient cooling of the same and the implementation of an hydrodynamic bearing around the cam.

The first embodiment of the present invention combines the advantages of the Wankel engine with those of the more conventional reciprocating types. As in the Wankel, the inertial forces of the second order, difficult to neutralize, are absent while as, in the case of the reciprocating engines, there are fewer problems of feasibility and reliability. Added to this, the internal mechanics of the piston are extremely simple. It must be stated that the alternate hypocycloidal movement made by the piston, and by the cam that translates with it, obeys a law different from that guiding conventional pistons constrained to a classical type of connecting rod. In the case of a conventional piston and due to the finite length of the connecting rod, speed at the top dead centre (PMS) is greater than that at the bottom dead centre (PMI), as well as related accelerations, creating accelerations of the second order to equalize the two contributions to overall acceleration. The piston described in the first embodiment of the invention only generates inertial forces of the first order, easier to neutralize. Starting from the

wherein is the radius of the circumference of the greater diameter (internal gear) within which the circumference of the lesser diameter b (pinion) smoothly rolls; while a = cot is the angle swept by a point on the circumference of diameter b at time t.

Putting a = 2b = 2R, gives:

(x(a) = ZR cos a I

I y(a) = 0 1 (3)

that represents the parametric equation of a straight line corresponding to the diameter of the circumference of diameter 2R covered by the piston under cosine law. Corresponding acceleration is:

(a) = -2R cQs (4)

entirely without components of the second order of type cos 2 cr.

The cam inside the piston performs a roto-translation the translational component of which, on the circumference covered by the crank pin round the drive shaft, is synchronous with the hypocycloidal motion of the piston that consequently generates only inertial forces of the first order easy to balance. The cam therefore differs from a conventional connecting rod both in form and in function even if, like a connecting rod, it lies between the 'crank' and the piston though coupled differently. As already mentioned, a fair part of balancing the inertial forces from the piston is already done with advantage inside it, thanks to a desired lack of uniformity within the mass of the cam.

The second embodiment of the present invention, piston displacement being equal, makes it possible to lessen the centre distances between drive shaft and the other parts involved in motion conversion; in this way mechanical bulk can be reduced both in hypocycloidal and in conventional reciprocating engines so as to lower the centre of gravity making the car more stable. At the cost of making the piston more complex, the variant reduces to the minimum, perhaps even eliminates, the inertial forces of the second order.

Finally, having no difficulties in achieving high compression, the engine of the present invention can function by the Diesel cycle without needing a Root or some other compressor.

Short description of the figures

Further purposes and advantages of the present invention will be made clear by the following detailed description of an example of it and by the attached drawings given for purely explanatory reasons in no way limitative, wherein: Figure 1 is an elevation view of a Wankel engine with the stator open to show its rotor;

Figures 2 - 5 represent four instants in an operative cycle of a known type of reciprocating two-stroke engine;

Figure 6 is a geometrical diagram of how to generate a two-way rectilinear hypocycloid, as being the principle inspiring the present invention;

Figure 7 is a diagram of the invention that realizes as hypocycloid movement of the piston as in Figure 6;

Figure 8 is a frontal longitudinal section of a two-stroke endothermic engine built according to the present invention;

Figures 9A, 9B, 9C, 9D, 9E, 9F are plan views of the various parts of the piston in Figure 8.

Figure 10 is a section view through the axis A-A in Figure 8.

Figure 11A is a view from above of a single transversal seal lamina shown in Figure 9C. Figure 1 IB is a perspective view of a pair of transversal seal laminas in Figure 11 A.

Figure 12 A is a section view through plane B-B in Figure 1 IB.

Figure 12B shows the working configuration of laminas in Figure 12 A.

Figure 13 is an exploded view of the framework formed by the transversal laminas and by the perimetric bands for the piston in Figure 8.

Figure 14 is a perspective view of the frame formed by laminas and bands in Figure 13.

Figure 15 is a perspective view of the piston in Figure 8 complete with the frame in figure 14.

Figures 16A, 16B, 16C, 16D show the series of configurations assumed by the piston in Figure 8, in four instants of its working cycle.

Figure 17 is a diagram of the piston in Figure 9 A to be used in subsequent figures.

Figure 18 shows the configuration of mobile parts inside the piston in Figure 17 immediately after reaching the top dead centre (PMS) where the arrows indicate the direction of initial rotation.

Figures 18A, 18B, 18C, 18D repeat the configuration of mobile parts inside the piston at the instants indicated in Figures 16A, 16B, 16C, 16D.

Figures 19A, 19B, 19C, 19D show in diagrammatic form the configurations of mobile parts inside the piston realized in accordance with a further embodiment of the invention, in the four previously shown instants of the working cycle.

Figure 20 is a front view of a piston in accordance with a variant of the second embodiment of the invention.

Figure 21 shows a longitudinal section on plane C-C in Figure 20, together with the front view to help identification of corresponding parts.

Figures 22A, 22B, 22C, 22D are diagrams of the configurations of mobile parts inside the piston in Figure 21, assumed in the four instants previously indicated of the working cycle.

Detailed description of some preferred forms of realization of the invention In the following description identical parts that appear in different figures may be marked with the same symbols. The scale and proportions of the various parts are not necessarily the same as the real ones.

Two preferred examples will be described, the first of which concern an endothermic engine in whose piston there is an internal gear with meshing ratio of 2:1 to create a rectilinear hypocycloidal movement, while the piston in the second example realizes a similar movement using one or more cams.

Preparatory for the first example are the diagrams in Figures 6 and 7 that concern the rectilinear hypocycloidal movement.

Figure 6 shows in diagrammatic form the rectilinear hypocycloid covered in alternate directions by the generic point on a mobile circumference that rolls on a circumference of a diameter double the size. The point to be traced is shown up by a radius of marked thickness along a 360° rolling trajectory divided in 30° sections. Observing the dotted curve that joins the positions gradually occupied by the centre of the mobile circumference during the 360° roll, it will be seen that the mobile circumference executes a complete rotation in relation to its own centre during the 360° roll, and therefore during an alternate cycle of translation of the fixed circumference along its diameter. It follows that, by constraining the centre of the mobile circumference to any point in a third circumference of the same radius and free to rotate round a fixed axis, the rotary movement of the mobile circumference will be transferred to said third circumference during the roll.

Figure 7 shows the transversal section of an ensemble that, inside a piston (not shown) schematizes: the circular edge 40 of the drive shaft; the pitch line 41 of the pinion; the pitch line 42 of the internal gear fixed to the piston; the circular edge of a cam (coinciding with the pitch line 42 only to simplify the drawing since the edge of the cam is always in contact with the wall of the square section piston). It will be seen from the figure that there are three parallel lines marked As, AO, Ai reciprocally spaced by a value equal to the diameter D of pitch line 42 of the internal gear fixed to the piston, wherein: line AO represents the fixed level held by the centre of the circumference 40 during rotation of the drive shaft at steps of 30° in the complete cycle of 360°; the upper line As represents the quota of the top dead centre (PMS) reached by pitch line 42; the lower line Ai represents the quota of the bottom dead centre (PMI) reached by pitch line 42. In this figure it will be seen that the diameter of pitch line 41 of the pinion is equal to D/2 and eccentricity of pitch line 41, in relation to the centre of the circumference 40, is equal to D/4. The figure explains the movement of the piston represented by pitch line 42 of the internal gear whose centre translates in an alternate motion covering a distance D in both directions of translation, equivalent to the stroke between the two dead centres. This movement is controlled by pitch line 41 of the pinion that rolls within pitch line 42 of the internal gear meshing constantly with it, constraining the centre to accomplish the above alternate rectilinear trajectory. Constant meshing, essential to the hypocycloid movement 2: 1, is made possible by the cam being situated between the extended axis of the pinion and the inner wall of the cavity of the piston which it prevents from literally 'falling' onto the pinion. Cam movement is necessarily roto-translational in that it must simultaneously participate in the alternate rectilinear movement of the piston, that contains it, and in the rotary movement round the axis of the pinion on which it is free to rotate. The kinematics of such a movement cause it to rotate in the direction opposite to rotation of the drive shaft (as shown by the arrows).

Figure 8 is a view of a front longitudinal section of a two-stroke reciprocating endothermic engine 45 comprising a cylinder 46 with piston 47 of a new design inside it. The figure shows the piston at the top dead centre (PMS). The two- stroke engine in this example exploits the thrust exerted by gas from combustion of a gaseous mixture of air and petrol put into the cylinder, the so-called 'fresh mixture' produced by a carburettor not shown, where combustion is provoked by the spark between the electrodes of a spark plug. This example in no way limits the invention as far as concerns the following versions: a) introduction of air into the cylinder and injection of petrol into the air compressed by the piston to obtain the mixture directly in the cylinder; b) introduction of air into the cylinder and injection of Diesel oil at high pressure in the air compressed by the piston (or by an external compressor) subjected to spontaneous combustion according to the Diesel technique; c) application of the techniques in the preceding points to a four-stroke engine. The versions in points a), b), c) in fact require small alterations by known techniques to the cylinder only without having to alter the structure of the piston. The following figures will show that both the cylinder 46 and the piston 47 are parallelepipeds so that the term 'cylinder' used in the example adopts the meaning current in the automotive field rather than describing its geometrical shape. The cylinder 46 is a hollow parallelepiped body, lying longitudinally, having no base to allow insertion of the piston 47 able to slide while maintaining contact with the walls of the cylinder cavity by means of a system of sealing bands, more clearly described in subsequent figures. A chamber of variable volume forms between the crown of the piston and the cylinder head with alternating translational movement by the piston, first compression then combustion. The edge of the opening at the base of the cylinder 46 extends outward like a flange for fixing it to the mouth of a sump 48. There are two openings, 49 and 50, along the opposite walls of the cylinder 46 substantially at the same height. Opening 49 is for introduction of the gaseous mixture in the combustion chamber. The mixture reaches the entry 49 through a transfer duct 51 in the wall of the cylinder 46. Opening 50 is a vent for burnt gas and communicates with a short duct 53 in the wall of the cylinder 46 for conveying the discharged gas to a collector (not shown) ending in a silencer, preferably of 'resonant' form to avoid loss of non-burnt mixture. In the walls of the cylinder 46 there are cavities 54 for the passage of cooling water. The cylinder 46, with the sump 48 bolted to it, rests on a frame 55. The sump 48 is divided into two separate compartments 57a and 57b, one below the other. The upper compartment 57a communicates with the cylinder 46 while the lower compartment 57b is a cup for collecting lubricating oil circulated by a pump (not shown). The shape of the wall of the upper compartment is such as to delimit a curved tube with an entry mouth 56 communicating with a central section open above towards the lower base of the cylinder 46, and an end section 52 for conveying mixture or air towards the entry to the transfer duct 51 into the sump through the mouth 56. One-way slat-type valves 58 are mounted on the entry mouth 56. The drive shaft 59 passes through the opposite walls of the cylinder 46, orthogonal to the walls containing openings 49 and 50 for introduction and discharge, at a distance from the head of the cylinder 46 greater than that of the cylinder head from openings 49 and 50. The ways of building the cylinder 46 to allow for such a passage have already been described. In the cylinder head there is a central threaded hole in which to screw an electric spark plug 60 or alternatively a Diesel oil injector.

The piston 47 is also a parallelepiped and hollow inside; it comprises an upper wall 64, a bottom wall 65, two opposite side walls 66, 67 and a further two opposite side walls through which the drive shaft 59 passes at a wide circular opening (at a level along which a section is made). This is necessary because the piston 47 must be free to translate across the drive shaft 59 that controls its movement. The four walls 64, 65, 66, 67 are of the same length, so forming a piston whose longitudinal section, orthogonal to the axis of the drive shaft, is square in shape. The walls are fitted with a series of seal bands that will be described later. The external surface of the head 64 of the piston 47 is the so- called 'crown', the surface of which is hardened to withstand the strong pressure from the hot gas produced whenever combustion takes place.

Inside the piston 47 a longitudinal section can be seen of the mechanism that permits its hypocycloidal translation. This mechanism if activated by the drive shaft 59 by a crank pin 68 in turn fixed to a coaxial pinion 69. The distance between the axis of the pinion 69 and the axis of the drive shaft 59 is that of R. The internal gear (schematized) of the pinion 69 meshes in a crown gear 70 with internal teeth (schematized). Radius of the pitch circle of the teeth of the pinion 69 is R while the radius of the pitch circle of the teeth of the internal gear 70 is 2R. A circular cam 72, made in two parts, freely rotates on the axis of the pinion 69. The axis of circular cam 72 is close to the centre of the piston 47 while its axis of rotation coincides with the axis of pinion 69. The radius of circular cam 72 is greater than the 2R radius of the internal gear 70 as it must rotate while maintaining contact with the internal faces of the piston 47, faces and piston being spaced at a distance greater than the diameter of said internal gear. Both the internal parts of the piston and the lateral faces of the cylinder in contact with the piston receive forced lubrication (separate from the air and fuel mixture) through a channel inside the drive shaft 59 that carries oil to the cam 72 and to the internal gear 70.

Figure 9A reproduces the section in Figure 8 relating to the piston 47 only, showing the baffle 71 within which is the internal gear 70 in a central position that meshes with the external teeth of the pinion 69. The internal gear 70 has twice the number of teeth as the pinion (69).

Figure 9B shows a square wall 95 of the piston 47 with a large central opening 100a; this together with a similar opening (100b) in the opposite face (not shown) enables the piston 47 to translate across the drive shaft 59. The figure brings out the system of oil-containing bands in each closed wall 64, 65, 66, 67. It will be seen that the open walls are similar to circular junctions between edges meeting at a vertex. Included in, and partially emerging from each side wall, 66, 67, orthogonal to the square walls, are three bands placed transversally to the longitudinal axis of the piston 47; of these an intermediate transversal lamina is closer to the centre of the face while the other two are at the two ends. The transversal laminas emerging from the wall 67 are marked 74a, 74b, 74c; the transversal laminas emerging from the wall 66 are marked 74d, 74e, 74f. Also included and partially emerging along the two sides of the same walls 66, 67 parallel to the longitudinal axis of the piston 47, are respective longitudinal oil- containing bands constrained at the ends of the transversal laminas. The figure shows a longitudinal band 75a of the pair in the wall 67, and a band 75c of the pair in wall 66. Also included and partially emerging along the two sides of each wall 64, 65, orthogonal to walls 66, 67, are respective lateral bands. The figure shows a longitudinal band 76a, of the pair in wall 66, and a band 76b of the pair in wall 65. Forced lubrication of the open walls is illustrated by 'drops of oil' from holes 77a shown uppermost near the head and returning into holes 77b below near the base. Bands 75a, 75b, 76a, 76b that surround the open face of the piston seen in the figure, and the other four corresponding bands in the other open face, form an ensemble of perimetric bands that translate the oil from holes 77a in contact with the corresponding cylinder wall throughout the piston stroke and over the whole face bounded by said walls.

Figure 9C shows the four walls of the piston 47, reversed in relation to the sheet, orthogonal to the four sides of the two open walls: the top wall 64, the base wall 65 and the two lateral walls 66, 67. The four walls are all of the same rectangular shape, of the same size and, towards the drive shaft, the depth of the solid delimited by all four walls is less than its height and length. In the lateral walls 66, 67 there is a set of holes 66b, 67b slightly below the transversal oil-containing lamina in an intermediate position, and another set of holes 66c, 67c slightly above the transversal containing lamina close to the base. The remainder 66a, 67a of the walls 66, 67, comprised between the intermediate transversal lamina and the transversal lamina close to the head, is devoid of holes. Lubrication of walls 66, 67 is schematically indicated by drops of oil from holes 66b that reenter in holes 66c. Through these holes the oil leaves the piston to lubricate the cylinder and then returns without accumulating or creating undesired pressure. They also permit the burnt gas, that has permeated the upper bands, to be conveyed into a cup. The upper part 66a 67a of the respective walls 66, 67 is apparently deprived of lubrication (dry zone), so the same would apply to the corresponding upper part of the cylinder wall, but in actual fact this is not so as lubrication is done through a mechanism to be described later. This 'dry' zone in the upper part of the piston is necessary for the two-stroke engine described in the example without total loss lubrication, otherwise the oil conveyed on that part of the face would contaminate the mixture, or the air, circulating in the transfer duct. Figure 9D shows the drive shaft 59 and the pinion 69 fixed to it at distance R between the two axes. These two parts are firmly held together by a crank pin (not shown) so forming a single piece.

Figure 9E shows the internal gear 70 only, of a diameter 4R, made on the circular edge of a hole in the centre of the inner wall forming the baffle 71. It will be seen from the figure that the diameter of the gear 70 is sufficient for part of the ensemble in Figure 9D to pass through it so that neither the drive shaft 59 nor the baffle 71 have to be broken in two. Figure 9F shows up the circular cam 72 with an eccentrically placed circular bole 72c for insertion of a crank pin (seen in the next figure) around which it can rotate. Insertion is possible because the cam 72 is divided into two unequal parts across the hole 72c. The cam 72 is symmetrical in relation to a diameter. One part, 72d, delimited by a diameter orthogonal to the preceding one, is solid while the other part is reduced to a circular edge and a central solid spoke 72b that separates two void quadrants 72a. The diameter of cam 72 is equal to the distance between opposite internal faces of square-section piston 47.

Figure 10 shows the longitudinal section along a plane A-A of piston 47 in Figure 8. It will be seen how the drive shaft 59, resting on two roller bearings 61 fixed to the frame 55, passes through the cylinder 46 and piston 47. A comparison with the front view in Figure 8 shows that the depth of the cylinder and piston is less than their height. Front and back walls of piston 47 in contact with the internal surface of cylinder 46 are almost entirely absent. The drive shaft 59 combines in a single piece the crank pin 68, the pinion 69 and a collar 80 adjacent to pinion 69. The crown gear of pinion 69 is obtained by chip machining one end of the crank pin of a larger diameter. The cylinder 46 is made in two joinable parts to permit the above passage. In the drive shaft 59 are two notches spaced farther apart than the walls of the cylinder 46 for insertion of two counterweights, 81, 82 that extend orthogonally beyond the drive shaft 59 from the opposite side in relation to the crank pin 68. Above the top of the figure there is a miniature reproduction of the drive shaft and its counterweights. Each of the two parts of the cylinder 46 has two semicircular seats on the edges to receive two bushings, 83, 84 that support the drive shaft 59 preserving the seal of cylinder 46. A third bushing 85, in two parts, lies across the crank pin 68 to support the circular cam 72, one part, 72d, of which is solid and the other part is lighter. The crown gear of the pinion 69 engages with the internal gear 70 that has twice as many teeth as the pinion. Above the main figure to the left is a drawing of a Diesel 62 injector that can replace the spark plug 60 with or without a Root compressor. The arrows in the sump 48 indicate the direction taken by the mixture towards the transfer opening 51 when the slat valve 58 closes as soon as the piston 47 starts its downward stroke. In this specific case piston displacement is 324 cc, the crown of piston 47 measures 30x100 mm and the stroke is 100 mm, Figure 11 A is a view from above of the transversal containing lamina 74a and is representative of all the others: 74b, 74c, 74d, 74e, 74f. In the profile of the lamina 74a, a rectangular part is contiguous to a trapezoidal part with two notches 88, 89 at the angles between them. Two aligned rectangular transversal windows, 90, 91 are situated to one side of centre towards that of the trapeze opposite the base of the lamina.

Figure 1 IB is a perspective view of transversal containing lamina 74a showing how it is superimposed over a second transversal containing lamina 74aa identical with the first, held in place by two helical springs 92, 93 lodged respectively in the rectangular windows 90, 91. The two laminas are longitudinally misaligned as the second one is rotated at 180° in relation to the first in a longitudinal direction. The configuration with two springs does not rule out the possibility of using one spring only.

Figure 12 A is a longitudinal section along plane B-B in Figure 11B that shows the spring 93 in a balanced configuration in which each end is in contact with both laminas; the same applies to the spring 92. In the balanced configuration the spring 93 is not engaged, the laminas 74a, 74aa are misaligned to the maximum which is equal to the difference between the length of the two sections in contact; in this case the distance between the outermost end of the two springs is greater than the distance between the walls of the cylinder 46.

Figure 12B shows spring 93 in its operative configuration compressed between laminas 74a, 74aa in contact with walls CL1, CL2 of the cylinder 46. It will be seen that the distance between the outermost ends of the laminas is reduced by about half in relation to maximum misalignment so that the shortest section of each lamina loses contact with the end of the spring as it has become shorter. Figure 13 gives an exploded view that better illustrates the complex system of laminas that make possible lubrication of surfaces in contact in both cylinder and piston, and containment inside the piston of oil in forced circulation to lubricate and cool the moving parts. Each lamina shown in the figure fits into its own seat in the corresponding wall of the piston 47, from where it can partially move to make contact with the wall of the cylinder 46 directly or indirectly pressed by springs, these too being in seats on the insides of the walls. The function of these laminas has already been explained. The figure distinguishes two systems of four laminas each; once in their seats they form two perimetric bands around the open faces of the piston, and two groups of three laminas each transversal to the other two lateral walls. The laminas 76a, 75a, 76b, 75b belong to a first perimetric band: laminas 76c, 75c, 76d, 75d belong to a second perimetric band. Laminas 74a, 74b, 74c, belong to a first transversal group; laminas 74d, 74e, 74f belong to a second transversal group. Laminas 75a, 75b, 75c, 75d, belonging to the perimetric band, are narrow rectangular bars there being in each one three rectangular fissures 95, one at each end and one near the centre; the fissures are accessible from one side for fitting the lamina into three similar fissures of type 88, 89 in three transversal laminas. The seats housing laminas 75a, 75b, 75c, 75d include a wavy wire spring 96 acting on the lamina to press it in a direction perpendicular to the fixed transversal laminas.

Figure 14 is an ideal recomposition of Figure 13 forming a framework present in the walls of the piston 47 in Figure 15. Figure 14 shows the twelve constraints that join the two perimetric bands to the two groups of transversal laminas, and the various springs that enable the parts of the frame to expand. On account of the constraints, transversal expansion worked by the helical springs 92, 93 presses the perimetric bands against the square walls of the cylinder 46. Similarly, expansion worked by the wavy wire spring 96 presses the transversal laminas against the rectangular walls of the cylinder. Figure 15 shows the piston 47 complete with the framework in Figure 14, making clear the movement of oil over the outer surface in relation to the various transversal laminas and perimetric containment bands. The Figure shows the wide opening 100a in the square wall that, together with the opening 100b in the opposite face, allows the piston 47 to slide inside the cylinder 46 while remaining across the drive shaft 59. Said openings are framed by respective perimetric bands that lubricate the walls of the cylinder with which they are in contact and then clean off the oil returning it into the openings. Generally speaking, the various laminas and bands force the oil to circulate within the limits they set. The sequence of Figures 16A to 16 C repeats the same instants in the operative cycle for the hypocycloidal two-stroke engine (lubrication kept separate from the mixture or from the air) as those in Figures 2 to 5 for a petrol-driven, two-stroke engine of a known type with total loss lubrication. The configuration assumed by the mechanics inside the piston 47 at each instant is better schematized in the sequence of Figures 18A to 18D; this is preceded by a more complete schematization in Figure 17, and by Figure 18 showing the initial imbalance of the internal mechanics just beyond the top dead centre (PMS) that establishes the direction of rotation of drive shaft 59 under translational thrust exerted by burnt gas on the crown 44 of piston 47. What should also be made clear, during a complete piston stroke, is the mechanism for lubricating the upper part of cylinder 46 corresponding to the 'dry' zone in the piston. The 'drops of oil' seen in the sequence of Figures 16A-16C indicate how the part below the 'dry' zone is lubricated on both sides of the face shown in the figures. Taking as an arbitrary starting point the instant at the top dead centre (PMS) in Figure 16A, it will be seen that oil lubricating walls 66, 67 in the area below wall parts 66a, 67a, does not directly reach the upper part of the cylinder because the intermediate transversal lamina holds it back below preventing it from entering the combustion chamber through entry and discharge openings 49, 50 just above. Lubrication therefore adopts an indirect mechanism that will be explained in the following figures. To avoid overheating and deforming the walls 66, 67 of piston 47, parts 66a and 67a are coated with a micrometric layer of silicon carbide that increases its surface hardness, its thermal resistance and resistance to corrosion, and lowers the coefficient of friction. The piston head, fitted with the two upper transversal laminas and with two other laminas in the perimetric bands, blocks the way to the burnt gas. The piston base, fitted with two lower transversal laminas and two more laminas in the perimetric bands, prevents penetration of oil into the sump 48. The surface of cylinder 46 is oiled by the piston in the areas between the intermediate and lower transversal laminas, in addition to those opposite the perimetric bands. With further reference to corresponding figure 18A, for numeration of the parts, it will be seen that, at the upper dead centre (PMS) the heavier part of cam 72 is farthest from the crown of piston 47. The next figure 16B shows an intermediate instant in the stroke of piston 47 towards the lower dead centre (PMI) due to expansion of burnt gas, Entry and discharge openings 49, 50 are still occluded by the piston. The slat valves are closed and the mixture in the sump 48 is subjected to precompression. During the piston's downward movement the 'dry' areas 66a and 67a on walls 66 and 67 gradually come into contact with the areas on cylinder 46 that were oiled when the piston was rising and are at present 'swept' by the intermediate transversal laminas. In spite of this, transfer of oil from the 'swept' surface of cylinder 46 to areas 66a and 67a devoid of direct lubrication is made possible by a process carried out on the surface of the inner wall of the cylinder called 'brushing'. This process impresses a criss-cross of sloping micro lines on the smooth surface forming a network that retains a small quantity of oil even after passage of the oil-scraping bands. The metal that then makes contact with this processed surface can be oiled by the residual oil entrapped and then released. The intermediate instant in the downward stroke of piston 47, during which it is lowered by D/2, corresponds to a clockwise rotation of 90° by the drive shaft 59 (see also Figure 18B) and at the same time to an anticlockwise rotation by the circular cam 72 of 90°. This rotation modifies dislocation of the mass of cam 72, lifting half of its heavier part above the median line of the preceding dislocation, partially compensating inertial acceleration of the piston 47.

The next figure 16C shows the instant when the piston 47 has completed its down stroke reaching bottom dead centre (PMI), and when the 'dry' areas 66a and 67a of walls 66, 67 have been entirely oiled by the oil previously entrapped in the swept surface of cylinder 46. Piston 47 no longer occludes entry and discharge openings 49, 50, and transfer of fresh mixture can proceed simultaneously with 'washing' the cylinder chamber. The instant spent at the bottom dead centre (PMI), on completion of the down stroke of D length, corresponds to a clockwise rotation of a further 90° by the drive shaft 59 (see also Figure 18C) and simultaneously to an anticlockwise rotation of a further 90° by the circular cam 72 that completes the uplift of its heavier part, partially compensating inertial acceleration of piston 47. The next figure 16D, the last of the series, shows an intermediate instant in the upstroke of piston 47 towards the top dead centre (PMS) favoured by inertia of the flywheel. The piston 47, once more occluding the entry and discharge openings 49, 50, compresses the mixture in the cylinder chamber and simultaneously creates a depression in the sump 48 that opens the slat valves at the entry mouth 56 to allow fresh mixture to enter the sump. While piston 47 is rising, areas 66a and 67a on walls 66, 67, now entirely oiled with residual oil by means of the mechanism described, slide into contact with the surface of cylinder 46 reducing the volume of the combustion chamber. The instant described corresponds to a further 90° clockwise rotation made by the drive shaft 59, to a corresponding rise of piston 47 by D/2 (see also Figure 18D) and simultaneously to an anticlockwise rotation of a further 90° by the circular cam 72 that lowers half the heavier part below the median line, partially compensating inertial acceleration of the piston 47. The configuration assumed at the end of the last instant of the cycle coincides with the initial one in figures 16A and 18 A; during this instant piston 47 completes its stroke and therefore compression of the mixture in the combustion chamber, simultaneously creating a depression in the sump 48 and suction of fresh mixture.

A two-stroke endothermic engine with separate lubrication will now be described, its piston differing from that in the preceding figures 8 and 9 for lack of the pinion and internal gear that, together with the circular cam, generated a hypocycloidal 2: 1 translational movement. The parallelepiped shape of the cylinder and of the piston is the same, though of different dimensions, as also unaltered is the framework of laminas and bands for oil seal. The resulting piston is extremely simple and is schematically described in the sequence of Figures 19A to 19D showing the configurations assumed by the piston's internal mechanics in the operative instants characterized by the following angles of rotation of the drive shaft: 0° (360°), 90°, 180°, 270°. The endothermic cycle is the same as the conventional one. Figure 19A is a longitudinal section of the piston at the top dead centre (PMS) according to a plane orthogonal to the drive shaft. It schematizes a piston 105 free to translate in both directions across the drive shaft 107 that passes through two opposite square walls. The drive shaft 107 includes, in a single piece, a crank pin 117 whose longitudinal axis is at an L distance from the axis of the drive shaft 107. The circular section of the crank pin 117 is externally tangent to the circular section of the drive shaft 107. Inserted on the crank pin 117 is a circular cam 1 19 whose geometrical axis lies at an L distance from the crank pin 1 17. The three axes pass through the centres of the three circles. The circular section of the circular cam 119 is externally tangential to the circular section of the crank pin 117. Reciprocal tangency does not however limit the invention but is advantageous in that it reduces the size of the piston 105. Compatibility of movement of the piston 105 with the kinematic constraints imposed by the cylinder 46 and by the internal mechanics, appears in subsequent figures where it will be seen that the cam rotates in the opposite direction in relation to the crank pin 1 17 and to the drive shaft 107, and that the stroke is 2L on reaching the bottom dead centre (PMI). There is however a lack of symmetry in the positions reached by the piston during the instants indicated, rotation of angles of the drive shaft 107 being equal. Passing from 0° to 90° the piston 105 follows a path longer than the one from 90° to 180° and this induces inertial components of the second order to balance the condition of nil speed at the dead centres.

The translational movement of the piston 100 in Figure 19 A, fitted with a single cam 119, becomes symmetrical during the stroke introducing a further crank, as shown in the following figures illustrating a variant, that has no effect on the shape of the cylinder, nor has it a system of transversal laminas and oil- containing bands present in the walls of the piston. The following figures all concern this variant.

Figure 20 shows the square face of piston 105 according to the variant. Figure 21 is a central section view along a plane orthogonal to the face in Figure 20 and parallel to the narrow faces of the parallelepiped shaped piston 105. Referring to both Figures 20 and 21, it will be noted that along the edges of walls 64, 65, 66, 67 of piston 105 there are the same transversal laminas and perimetric bands containing oil as seen in Figure 9B. Said walls form a frame through which can be seen a wide opening 106a for introduction of the drive shaft 107 and of other parts to be mentioned later. A similar opening 106b is present in the opposite face. Inside piston 105 are two L-shaped rails 108, 109, symmetrical to the centre line, fixed by screws 112, 113 to respective protrusions towards the inside 1 10, 1 11 of the head and base of piston 105. The shorter side of each rail 108, 109 is supported by its own protrusion that runs across the whole length of rail 109 from one lateral wall to the other. A square slide 1 14, similar to a small frame, can translate along said rails, one side in contact with rail 108 and the opposite side with rail 109; in the sides of the slide in contact with the rails is a longitudinal groove to receive the longitudinal edge of the respective rail during translation. The drive shaft 107 passes through piston 105 supported by two bushings 1 15, 116 inserted in corresponding holes in the walls of the cylinder 46. Inside the piston 105, the drive shaft 107 forms a sort of gooseneck in a single piece, comprising a first crank 117 and a second crank 118 contiguous to the first. In practice the two cranks are misaligned sections of the drive shaft 107, reduced to crank pins of different diameters. A circular cam 119, free to rotate on crank pin 1 17, maintains contact with the wall of the piston 105. One part of the circular cam 119 is lighter, thinner towards the external circular edge. The edge of crank pin 117 exceeds the remaining part for a short length starting at the crank 118 side and in contact with the edge of the cam 119.

The diameter of the first crank pin 117 is about 3/2 that of the drive shaft 107. The diameter of the second crank pin 118 is about equal to that of the drive shaft 107. The transversal sections of the three elements are circles with their centres aligned along a common axis, of which the circle corresponding to drive shaft 107 is internally tangential to the circle of the first crank pin 1 17 while the circle corresponding to the second crank pin 118 is internally tangential to the circle corresponding to the first crank pin 117. The distance L of the axis of drive shaft 107 from the axis of the first crank pin 117 is equal to the distance L of the axis of crank pin 117 from the axis of the second crank pin 1 18. This means that both the drive shaft 107 and the first crank pin 1 17 for that given stroke of the piston 105 can be of a larger size. In practice, it has been possible to adopt a drive shaft diameter of five times the value of eccentricity against the two of a model with gears. As the second crank pin 118 is aligned with the first crank pin 1 17 at the point of maximum extension beyond the drive shaft, there is a space half the height of the drive shaft diameter, above and below the second crank pin 118, for slide 114 and its rails 108, 109. Although the diameter of circular cam 119 is about double that of drive shaft 107, the area covered by the square face of the piston 105 does not increase overmuch because three-quarters of the bulk of the circular cam coincides with that of its pin 1 17. The two figures also show the channels, for forced circulation of lubricating and cooling oil, inside the walls of cylinder 46 and of drive shaft 107. For greater clarity, ducts respectively 120 and 121, for draining the oil collected by the transversal containing laminas, are made in the opposite walls of cylinder 46 of Figure 20 swept by these laminas. In the opposite walls of cylinder 46 of Figure 21, swept by the perimetric bands, are respective ducts 122, 123 for delivery of oil under pressure from the pump, and respective ducts 124, 125 for defluxion of exhausted oil towards the cup 57b. Each delivery duct communicates with a hole present in a respective bushing 115, 116. Duct 123 extends in a radial direction, beyond bushing 115, in an internal duct 126 of the drive shaft 107, entering a duct 127 in an axial direction. The axial duct 127 enters the first and second crank pins 117, 118 separating into respective ducts 128, 129 towards their circular edge to lubricate the contact with circular cam 1 19 and that between a sliding block of the slide 114 and the rail 113. Two further ducts depart from the axial duct 127, these being aligned with the previous two 128, 129 but directed towards opposite points on the circular edge of the two respective crank pins to lubricate the contact with circular cam 119 and the sliding block between slide 114 and rail 108. The oil carried to circular cam 119 passes through two radial ducts in its wall to lubricate the sliding contact with the internal walls of the head and base of piston 47.

The configurations assumed by the internal parts of the piston 105 in the four most significant instants of an operative cycle are schematized in the sequence of figures from 22 A to 22D. These instants correspond to the following four angles of rotation of the drive shaft: 0° (360°), 90°, 180°, 270°. The endothermic cycle is the same as the conventional one. Figures 22A to 22D are aligned with figures 19A to 19D so that, disposition of the parts and the value L of eccentricity being common to both sets of figures, a direct comparison may be made between the two types realized. The sequence of figures 22A to 22D clarifies the contribution of the slide 114 translating together with the piston 105 to which it is constrained. In Figure 22 A, with the piston 105 at the top dead centre (PMS), pin 1 18 is in the centre of the slide 114, the axis of circular cam 119 is aligned with the axis of the first crank pin 1 17 and with that of the second crank pin 118; in this way no component orthogonal to the axis of the piston can be created when the piston is at the upper dead centre (PMS). Ideally it should remain immobile as happens with any other type of reciprocating engine. There is however no stalling on account of the flywheel and of a difference of phase in the ignition. A figure such as this should therefore be imagined at a microscopic interval after exceeding the upper dead centre (PMS) in the direction indicated by the arrows. Figure 22B shows the internal configuration when piston 105 has completed a translation of 2L to the bottom dead centre (PMI) and the drive shaft 107 has completed a rotation of 90°. The figure shows how rigidly the gooseneck, made in one piece of parts 107, 1 17, 118 rotates, and rotary translation at 90° of circular cam 119 that has provoked a 2L lowering of piston 105 and, with it, of slide 114. In the variant, the further kinematic constraint formed by the slide 114 with the second crank pin 118 creates an additional moment on the drive shaft 107 in a direction appropriate to the moment created by circular cam 1 19 alone. The effect produced by the extra moment is to synchronise piston translation with rotation of the drive shaft and this is clearly seen in subsequent figures; translation of 2L by piston 105 in figure 22B is in fact greater than that seen in the similar figure 19B, and the difference is such that this translation, corresponding to a 90° rotation by drive shaft 107, can be expressed in whole multiples and is equivalent to a degree of synchronization that will be maintained at each increase of 90°. Figure 22B also shows that, while piston 105 has translated by 2L to bottom dead centre (PMI), the second crank pin 118 has translated outwards by 2L into the slide 114. The joint effect of the two translations in orthogonal directions is to cause drive shaft 107 to make a 90° rotation. The preceding synchronization is maintained in the configuration shown in figure 22C where the crown of piston 105 is at the lower dead centre (PMI), a length of 4L away from the cylinder head, and the second crank pin 1 18 has returned to the centre of slide 114. The configuration shown in the next figure 22D is that in which piston 105, in its stroke towards top dead centre (PMS), has reduced the distance from the cylinder head by 2L, the drive shaft 107 has rotated a further 90° and the second crank shaft 118 has translated by 2L into slide 114 in the direction opposite to the previous one. The next instant completes the endothermic cycle repeating the configuration in flgure 22A. At the end of this instant piston 105 has completed its stroke towards top dead centre (PMS), the drive shaft its revolution of 360° and the second crank pin 1 18 has returned to the centre of slide 114.

Based on the description given of a preferred example of realization of the invention, some changes can clearly be introduced by a specialist in the field without thereby departing from the sphere of the invention as will be shown by the following claims.