WO2005012692A1

WO2005012692A1 - Multicylinder barrel-type engine

Info

Publication number: WO2005012692A1
Application number: PCT/EP2004/051484
Authority: WO
Inventors: Giulio Martinozzi
Original assignee: VOGLAIRE, Hélène
Priority date: 2003-07-25
Filing date: 2004-07-14
Publication date: 2005-02-10
Also published as: ATE448385T1; EP1658417B1; EP1658417A1; DE602004024082D1

Abstract

A Multi-cylinder barrel engine having a centre line (C/L), and a centre point O, comprising: -a casing (12) on which a plurality (N) of cylinders (11) are parallely disposed in a circular row around C/L; - a corresponding number (N) of pistons (13), each mounted inside each of said cylinders; - a crankshaft (16) comprising two co-axial end sections (17,18) and an intermediate section (19) whose axis (w-w) intersects the common axis (C/L) of the end sections at the centre point O with a tilt angle α, whereby said crankshaft is rotatably mounted on the casing through a first couple of bearings (20,20') operating between the casing and the two end sections, the latter being solidly connected to the extremities of the intermediate section by at least one connecting crank (21); - a wobbling body (15), substantially axial-symmetric, rotatably mounted, through a second couple of bearings (22,22'), on the crankshaft's intermediate section (19) so that its axis (w-w) coincides with the central axis of said body, said pistons being connected and carried at the peripheral edge (14) of said wobbling body, whereby said pistons and wobbling body connected together form a single piece called Wobbling Unit (WU); - a first couple of flat bevel gears (26,27) having identical primitive cones with a vertex angle µ = π - α/2, said gears being fixedly mounted the one (26) on the casing (12) and the other (27) on the WU, in opposed position and such that the vertexes of their primitive cones coincide with the centre point O of the engine, whereby the axis of the one gear (26) coincides with the C/L, and the axis of the other gear (27) coincides with the common axis (w-w) of the WU and intermediate section (19), said first couple of gear providing, by their meshing, a constraint against the rotation of the WU around the C/L.

Description

Multicylinder barrel-type engine

The invention relates to a multicylinder engine with the cylinders disposed in a circle and parallel to the centre line C/L, to assume the so-called barrel-type configuration, wherein the pistons are mounted and supported by a wobbling mechanism. Said mechanism is conceived to reversibly convert the reciprocating movement of the multiplicity of pistons (disposed in a circular row), into the rotation of a crankshaft around the C/L, and for reversibly convert at the same time, the forces insisting on said pistons, into a corresponding couple insisting on the same central crankshaft.

Several mechanisms of this type are known in the art since long, usually called "Wobbler" or "swash disk mechanism", which are able to accomplish such a conversion of movement and forces. Most pertinent examples are disclosed by the following documents: - D1 : US-A-2 702483 (to Girodin, 1965)

- D2 : US-A-2 737 055 (to Dauben, 1956) - D3 : US-A-3 180 159 (to Girodin, 1965) - D4 : FR-A-2 251 220 (to Girodin, 1975)

- D5 : US-A-4491 057 (to Ziegler, 1985) None of these disclosures, with the exception of D2, describes the mechanism as adapted to operate as the basis for an internal combustion engine. However, they all have in common an essentially axial-symmetrical core element, whose movement can be described also as a spherical movement, that is a movement in which all the points of the element always remain on a sphere, whose centre represents the central point O of the mechanism: see the example of figure 1 and its vectorial representation of figure 2.

The spherical movement of such a core element, which from now on will be referred to as "the wobbling unit" and abbreviated in WU, is essentially determined by two types of constraints: a) - an axial constraint, formed by supporting bearings mounted on a tilted shaft portion of a generally Z-shaped main crankshaft, whereby the axis (w-w) of the WU is bound to form a fixed angle α with the axis (C/L) of the crankshaft, thus describing a circular double- cone having as aperture angle the angle α (corresponding to the tilt angle of the shaft portion with respect to the axis of the main crankshaft) and having its vertex coinciding with the central point O of the spherical movement; b) - an angular constraint against the rotation around its own axis, determined by the meshing of a bevel gear (m), having an aperture of 180°- α/2, integrated to the WU, over an identical bevel gear (n), symmetrically opposed and integrated to the casing of the mechanism: see, the prior art examples illustrated in figure 16 and 17 (where the couple of bevel gears are referred to with 25, 26 and 15, 22 respectively).

It is to be noted however that said angular constraint can alternatively, but less efficiently, be obtained by means of a spherical coupling and a cardan joint between the wobbling element and the casing of the mechanism, as illustrated by the known mechanisms of D1 and D 5.

A deeper and theoretical description of this type of wobbling movement, which is the essential cinematic basis for the present invention, is to be found in a further patent publication: US-A-3 864 982 dated 1975. The "swash disk mechanism" of D2, which is the only known mechanism allegedly adapted for an internal combustion engine, hints to an admission and discharge system for the operating fluid which would be readily controllable by means of "a single rotary element mounted on the crankshaft". However, no details are given in D2 about such rotary element or any indication on its functional mode. It is also known, from the above prior arts, to movably connect at the peripheral edge of the wobbling unit, an undetermined number of rods and pistons to define a circular row of pistons with relevant cylinders within which the pistons are driven, with a cyclic circular sequence dictated by the cyclic revolution of the Z-crankshaft, to move up and down, so as to generating a displacement for a volumetric machine. However, it is up to date not known to provide a fluid admission and exhausting system capable of efficiently serving an axially disposed circular row of pistons and cylinders, driven by a wobbling unit as defined above, so as to allow the performance of the classic 4-stroke thermodynamic cycle of the current internal combustion engines.

No such type of engine, in fact, ever appears to have been built, within the state of the art. It is therefore the main scope of the present invention to provide a multi-cylinders engine having a layout configuration based on cylinders disposed on a circular, axial row with the corresponding pistons supported on a wobbling unit, said engine being provided with a simple and efficient fluid admission and exhaust system, so as to allow the construction of a compact, light weight and efficient multi-stroke internal combustion engine, said engine being also characterised by a very reduced number of moving parts.

This scope can be met by an internal combustion engine of the type specified in the preamble of the independent claim 1 and further comprising, according to the definition of the characterising portion of claim 1, a fluid distributor ring which is composed by a number S of identical circular sectors, each comprising one admission opening and one exhaust opening as well relevant ducts leading to admission and exhaust collectors respectively, said number S depending on the number N of the cylinders, said ring being mounted directly over the heads of the cylinders in a sliding contact therewith, along a plane perpendicular to the central axis, so that the each of said openings is able to successively coincide with a unique hole provided in the cylinder heads, thereby acting as an admission or an exhaust valve alternatively, the rotational speed ω of the ring being proportional to the rotational speed ΩM of the crankshaft and function of the number of cylinders, according to the relations: S = (N + 1)/2 and ω_d = Ω_M / (N + 1) respectively, or to the alternative relations: S = (N - 1)/2 and ω_d = - Ω_M / (N - 1) where N is an uneven integer > 3. Another object of the invention is to select the value of certain geometrical parameters of the wobbling unity, in particular the tilt angle α < 10°, such as to minimise the lateral displacement of the piston, and optimise the space availability for the cylinders in the circular row, in relation to the piston's stroke. Another object is to minimise the friction forces around the piston. This is achieved by fixedly mounting the pistons on the wobbling unit so as to form a single solid piston unit.

Another object is to alleviate the torsional stresses on the crankshaft and the cyclic forces on the crankshaft bearings (against the wobble unit and against the casing) so as to allow an overall lighter structure and the lower level of vibrations. This is achieved by providing a further couple of bevel gears having the vertex of their primitive cones coinciding with the centre point O of the mechanism.

Another object is to shape the internal rotating ducts of the distributor ring so as to generate a pre-compression of the operating fluid at the entrance of the cylinders.

This can be achieved by giving the walls of the ducts the shape of centrifugal turbine blade.

Other objects and related advantages of the invention will be recognized from the following detailed description taken in conjunction with the accompanying drawing's sheets, in which:

Figure 1 is a vertical axial section of an internal combustion engine according to a preferred embodiment of the invention.

Figure 2 represents the vectorial diagram governing the cinematic parameters of the wobbling movement in relation to the rotating parts of the engine.

Figure 3 is a perspective view of a type of distributor ring, featuring 5 sectors and adapted to govern the 9-cylinder engine of figure 1.

Figure 4 is a schematic perspective representation of the 9-cylinders engine main body, which can be mated with fluid distributor ring of Figure 3. Figure 5 is a perspective representation of a partially cut-away engine according to the configuration of figure 1.

Figure 6 represents an axial section of a three-dimensional model capable of physically demonstrating the dynamic valve function of the distributor ring, by its synchronic rotating versus the pistons strokes. Figure 7 comprises 8 schemes (Fig.7a to 7h) which reproduce a 7-piston- engine axial-cylindrical sections, developed on a plane, taken at the radial level of the cylinder ports and at 8 different points in time (like photo shuts), all spaced a part by a regular time interval ΔT, during one complete 4-stroke cycle. Figure 8 is an over-view and a cylindrical section of a cylinder-head's port provided with one possible type of sealing (against the distributor ring). Figure 9 represents the classic Diesel (full line) and Otto (dash line) thermodynamic cycles in a p/v plane, along a 4-stroke cycle. Figure 10 represents the same cycles of fig. 9 but on a T/S plane, with the assumption of an ideal gas working in frictionless and pure adiabatic conditions. Figure 11 similar cycles as above, on a p/v plane, with additional dashed lines and area to represent a double expansion phase along an 8-stroke cycle. Figure 12 thermo-dynamical cycles as above, on a T/S plane, with additional dashed areas to represent a double compression phase and a double expansion phase, along a 12 stroke cycle. Figure 13 is a bottom view, completed by a circular elevation section (A-A) of one sector (A) of a distributor ring adapted for a 4-stroke cycle to be operated in a 7-cylinder engine.

Figure 14 comprises a horizontal section (A - A) and a radial elevation section (B - B) of one sector (A) of a distributor ring adapted for an 8-stroke cycle to be operated in a 7-cylinder engine. Figure 15 represents the trajectory of a point P(r,λ) solid to the WU, when λ=

90°, during one entire wobbling cycle. Figs 16, 17 represent two examples of the prior art. Various embodiments of the present invention will be described with reference to accompanying figures. It is to be noted that the same or similar reference numerals are applied to the same or similar parts and elements throughout the drawings. Figures 1 to 5 show a first embodiment of the claimed internal combustion engine whose general layout is formed by a plurality of combustion chambers 11 or cylinders assembled together in a circular row over an essentially circular casing 12 with their axis generally parallel to the centre line (C/L) of the casing, i.e. of the engine.

Such a layout authorises the conventional denomination of "Barrel Engine". Within each of said cylinders a corresponding plurality of pistons 13 are able to move up and down in a sealed contact, so as to generate a displacement volume, the pistons being supported and connected, through their stems, to the peripheral edge 14 of a wobbling unit 15 (WU), which is in turn rotatably supported, through bearings (not shown), by a Z-shaped crankshaft 16. This crankshaft is mounted on the casing, through other bearings (not shown), in order to rotate around said C/L. On has to stress here that, contrary to the known configurations, the present engine comprises an uneven number of cylinders: this is a necessary condition in order to allow the synchronisation with a rotary distributor 23 which governs the admission and exhaust of the operative fluid into and from each combustion chamber, as later on further explained.

Said rotary distributor is formed by a solid ring composed by a number S of identical circular sectors, see figure 3, each sectors comprising one admission opening (Aa, ..., Ea) and one exhaust opening (Ae, ..., Ee), as well as relevant ducts 24, 25 leading to corresponding admission and exhaust external openings which communicate with an admission collector and an exhaust collector respectively.

Said number S depends on the number N of the cylinders, according to the relation S = (N + 1)/2. Said ring is rotatably mounted around and perpendicular to the C/L of the engine, directly over the heads of the cylinders in a sliding contact therewith so that the each of said openings is progressively brought into register with a single hole or port (P1, ... , P9) provided in the cylinder heads, thereby said ring functioning as admission and alternatively exhaust valves, contemporary over all the cylinder heads with a cyclic frequency equal to Ω_M/2, as it is necessary for controlling the well known four-stroke thermodynamic cycle of the internal combustion engines. To explain the function of the distributor ring, the attention is drawn to an example of a 9 cylinders engine (N=9), as schematically represented in figures 1 , 3 and 4, wherein the 5-sector distributor ring 23 rotates and slides over the 9 cylinders heads according to the arrows, wherein each sector (A, B, C, D and E) spans an angle of 2π/(N+1 )/2= 2π/5= 72° It will be recognised that in a four-stroke cycle and supposing to start the cycle in cylinder n°1 , with the piston n°1 at its upper dead point, the aspiration phase begins through the admission port Aa of the sector A (see figure 2) which is initially (almost) in register with the port P1 of the cylinder-head n° 1. After a crankshaft rotation σ =180° (π radiants) the piston reaches its lower dead point, the aspiration phase terminates and the sector A rotates over an angle = π/N+1=π/10.

In this position, the port P1 is out of register and the sector A faces said port with a flat wall which seals the cylinder head, thus allowing the piston to begin the compression phase, which is followed, during a successive shaft rotation of 2π radiants (σ = 360°), by the ignition and the expansion phase. At the end of the expansion (piston again in the lower dead point) the shaft has completed a 3π radiants of rotation while the sector A has rotated by an angle δ = 3π/10, so as to bring its second port, the exhaust port Ae, progressively in front of the same opening 1 of cylinder 1. During a further rotation of π radiants of the shaft, the piston with its fourth stroke can then expel the exhaust gases through the second port Ae, thus concluding the entire four stroke cycle: the piston n°1 is again in its upper dead point, the shaft having completed two revs, (i.e. σ = 4π), the distributor having rotated by δ = 4 π l 0, that is δ =72°. Thus the sector B is now ready to take over the distribution on the cylinder n° 1 , for the following cycle.

In figure 13 another more detailed example of a single sector is given, as part of a four-sectors distributor ring adapted to govern a 4-stroke cycle on a 7 cylinder engine. Examining now the sequence of figures 7a to 7h ( figure 7), which have been drawn in the case of a 7-piston engine with a 4-sectors (A,B,C,D) distributor ring, it becomes evident how the sector A, while controlling the in-and-out gas flow on the cylinder n° 1 , as above explained, finds itself in the correct position in order to simultaneously control, in an identical manner, the in and out flow in the following cylinders n° 2 and 3 although in a different time phase, the time shift being given by the angle between two cylinders (: 2π/N= 2π/7 radiants = 51,4°) divided by the rotational speed of the ring: ω_d = Ω / (N + 1)

The same consideration is applicable to the other sectors B, C and D which are synchronously and simultaneously governing the in-and-out gas flow over the other cylinders, although with an engine angle shift of 2π/4 radiants for each sector. Coming back to the illustrations of figures 3 and 4, it is finally to remark that in this way, when the engine's shaft runs clockwise, each phase (e.g. the ignition) of the 4-stroke cycle taking place through all 9 cylinders, invests successively said cylinders with the following sequence: 1 , 8, 6, 4, 2, 7, 5, 3, 1, and 1 , 8, 6, ... again (according to the cylinder numbering given in figure 4), that is regularly jumping from one cylinder to the after-next.

The above sequence clearly explains why the selection of an uneven number N of cylinders is a necessary condition for a regular succession of all the thermodynamic phases over the cylinders, i.e. for the harmonic and smooth operation of the engine.

Such a sequence is further illustrated by the series of schematic vertical sections, taken on a 7-cylinder-engine, at a certain number of subsequent points in time of a 4-stroke cycle, as reported in the already cited figure 7. The associated distributor ring, in this example, is that partially (one sector only, spanning 90°) reproduced in figure 13. The combination of the two figures allows to visualise the synchronisation and succession of the 4 phases, guaranteed by the illustrated angular position of the admission (Aa, Ba, Ca, Da) and exhaust (Ae, Be, Ce, De) openings provided in the 4 sectors. From the cinematic point of view one has also to observe, from these figures, that in order to correctly synchronise the opening/closing phases of the N valves (the ports P-i, .., P_n of the cylinder heads) with the pistons strokes in each cylinder, the distributor ring must rotate, every full cycle (i.e. every 2 revs of the engine), by the angle corresponding to a single sector: 90° in the present example. That means engine must rotate 4x2= 8 (= N+1 ) times for every full revolution of the ring. This reasoning brings to the simple mathematical relation already quoted on page 3: G>_d = Ω_M / (N + 1), (with N being an uneven integer > 3). By extending the above reasoning, which are based on a mere geometric or cinematic analysis, it has been found that it is possible to generalize the governing function of the ring distributor of the present invention, also to new types of thermodynamic cycles characterised by a higher number of piston strokes, e.g. : a 6-stroke or in a 8-stroke or even in a 12-stroke cycle. This will become apparent by following this reasoning:

If one define with C the number of strokes characteristic of a given thermodynamic cycle, C/2 obviously representing the number of engine shaft's revs needed to accomplish one full cycle within each piston, and if one maintains the same speed ratio ωd = Ω_M / (N + 1) between the shaft and the ring, it is an immediate step to find the following general relation: S=(N+1)/(C/2) where S is the number of identical sectors needed to form a distribution ring capable of synchronize and govern an N-cylinder engine in a C-stroke cycle. In fact, such a distributor ring, given said speed ratio, will have to slide over the port of each cylinder, by the span of one full sector during the time the shaft accomplish C/2 revs ( necessary for the piston to accomplish a C-stroke cycle).

Therefore, one revolution of the ring will corresponds to S x C/2 revs (of the shaft) but also to N+1 revs, due said established speed ratio. Hence the general relation: SxC/2=N+1

Of course, this relation is also valid for the 4-stroke cycle:

S₄= (N+1 )/2 (e.g.: for N=7 , S₄ = 4 )

Focusing, instead, on a 6-stroke cycle the design of the distribution ring must provide for a number S₆ of identical sectors: S₆= (N+1)/3.

It appears soon that each sector completes a full sweep over one of the N cylinders every 3 revolutions (σ = 6π) of the shaft, that is while each piston moves up and down 3 times. It is not sufficient, in such a design, that N be an uneven number (for the reason of synchronisation) but, since S₆ must clearly be a whole number, also N+1 must be a multiple of 3.

Thus, one has to select N=5 with two sectors, as a minimum, or N=11 with 4 sectors, or N=17 with 6 sectors, in order to operate a 6-stroke cycle engine. Similarly for an 8-stroke cycle S₈ = (N+1)/4, and each sector completes, in operation, a full sweep, over one of the N cylinders, every 4 revolutions (σ = 8π) of the shaft, that is while each piston moves up and down 4 times. Also in this case the number of cylinder N must be odd (for the reason of synchronisation) but also N+1 must be a multiple of 4.

Thus, one has to select N=3, as a theoretical minimum with a single sector (Ss = 1 ), or N=7 with two sectors, or N=11 with 3 sectors, or N=15 with 4 sectors, etc... A particularly advantageous ring design, is shown in figure 14, wherein N= 7 and S₈= 2 . In fact the ring is formed by two sectors, only the sector A being completely drawn with full lines, and sector B is drawn with dash lines. They are however identical and provided each with six ports, in specific angular positions, which allow the correct control of the gas flow in-and-out flow with a synchronic sequence, parallel to that explained in relation to figure 13, but extended to the case of the 8-stroke cycle. Further aspects on this example will be discussed later on.

A still further extension of the distributor ring function is also conceivable for a 12- stroke cycle, with a design featuring a number of identical sectors: Sι₂=(N+1)/6, always with the same speed ratio: ω_d = Ω / (N + 1) between shaft and ring. It is soon to understand that each sector completes, in operation, a full sweep over one of the N cylinders, every 6 revolutions (σ = 12π) of the shaft, that is while each piston moves up and down 6 times. Also in this case the number of cylinder N must be odd (for the reason of synchronisation) but also N+1 must be a multiple of 6. Thus, one has to select N=5, as a theoretical minimum with S₁₂= 1 (a single sector spanning full 360°), or N=11 with S₁₂=2 (each spanning 180°), or N=17 with 3 sectors, or N=23 with 4 sectors, etc...

The thermodynamic nature and associated advantages of a 6- or 8-or 12-stroke cycle is unknown in the state of the art, at least within the field of the volume- displacement engines, having the classic multi-cylinder V or in-line configuration, supposedly because considered too complex and impracticable to realize, so as to leave them unexplored. Instead, it has been found, analysing the present barrel engine configuration, that very useful applications can be accomplished, taking advantage of the above multi- stroke cycles capability, and still maintaining a very simple mechanical layout. Contrary to the 6-stroke cycle, which does not appear to find any useful application, the 8-stroke and the 12 stroke cycles, have shown to be of particular interest, as will be soon explained.

In order to analyse and evaluate the thermodynamic phenomena associated to the above mentioned multi-stroke cycles, it is necessary to recall some theoretical notion at the basis of the well-known four-stroke "Otto" and "Diesel" cycles, which are therefore represented by the superimposed diagrams of Figures 9, and 10. In figure 9, for the ideal case of absence of losses and of perfect gas, the "Otto cycle" (dash lines) and "Diesel cycle" (full lines) are represented on the Cartesian plane of the "pressure versus volume" (p/v plane). In figure 10 the same cycles are to be seen on the Cartesian plane of the "absolute temperatures versus entropy" (T/S plane).

Therein, the three main phases of these "open" cycles, are shown by the lines 1-2, 2-3 and 3-4, which indicate the adiabatic compression phase, the heat supply phase (combustion by spark ignition or by progressive injection) and the adiabatic expansion phase, respectively. The line 4-1 conventionally represents the "heat extraction" phase which takes place externally of the engine, since notoriously the "open cycles" start with fresh gas (air) sucked from atmosphere (at the conditions of point 1) and terminates with expanded gas discharged into atmosphere (at the conditions of point 4). It is known how the area internal to the diagram of figure 9 represents the "Work" transformed into mechanical energy by the engine during every cycle. Similarly in figure 10, i.e. on the T/S plane the internal area of the diagrams represents the equivalent "Heat" (Qi -Q₂) to said Work, transformed into mechanical energy. Furthermore the energetic efficiency of the cycles is measured by the ratio η = (Qι-Q₂)/Q₂. It can be seen, in both figures, that the Diesel cycle is characterised by a bigger diagram area, forthe same initial conditions and for the same max temperature T₃, than the Otto cycle. This explains, at least in considerable part, the higher energetic efficiency of the Diesel engines over the Otto engines. In both cycle, however, one must observe that the conditions at which the gases are discharged from the exhaust valve (represented by the point 4) are determined by the available volume V for the expansion which is inevitably identical, for mere geometrical reasons, to the initial volume \ : in the lower part of the figure 9 a schematic cylinder is drawn with the piston in both high and low extreme positions corresponding to the points 2 and 4 (or 1) of the cycles. This brings to evidence that in all types of conventional volumetric engines, the volume displaced by the piston during the compression is inevitably identical to the volume displaced during expansion.

The equally inevitable consequence is that the pressure and temperature levels of the gas discharged are much higher than those of the gas sucked at the initial point 1 , with a considerable loss of energy (thermal and mechanical) dispersed into the atmosphere (or into the muffler).

If one could continue the expansion process in a further expansion chamber, such energy (i.e. power), corresponding to the dashed area in the diagram of figure 11 , would be recovered and transformed in an additional portion of mechanical power, with a considerable increase of the thermodynamic efficiency of the cycle. However, a side of the turbo-compressed engines which partially utilise such a lost gas energy at the discharge, i.e. trough a turbine, none of the known volumetric engine uses a double expansion chamber or system in order to exploit such an energy portion of energy of the Otto or Diesel cycles. It has been found, that utilising the above mentioned 8- or 12-stroke cycles, corresponding to 6 or 8 thermodynamic phases respectively, said double volumetric expansion can be accomplished with an astonishing simplicity and mechanical efficiency, just as if the expansion travel of the piston could be prolonged beyond the lower dead point (see the schematic drawing in the lower part of figure 11), which is clearly impossible with a classic engine configuration. This becomes instead feasible with the presently claimed multi-cylinder engine configuration, by means of two particular embodiments of the distributor ring, which are hereafter described.

The first of these embodiments, which is defined by claims 6 and 8, will be described on the basis of an example, schematically illustrated in figure 14, i.e. a particular ring configuration adapted for an 8-stroke cycle and for a 7 cylinders engine. Through a sequence of 8 strokes within the same cylinder (e.g. the n° 1 , having a port P1 ) a very interesting thermodynamic cycle, composed of 6 phases can be accomplished, which is represented by the cycle of figure 11. With the selected parameters (N=7 and C=8), only two identical sectors A and B are necessary, as a consequence of the already explained relation S= (N+1)/(C/2).

These two sectors, as it can be seen from figure 14, are provided, beyond the admission opening Aa, Ba, for the aspiration (similar to the 4-stroke ring of figure 13), each with an expansion transfer chamber (ETC), these chambers being identically shaped in the two sectors so that a toroidal intermediate pressure transfer chamber (IPTC) is formed inside the distributor ring 23. Said chamber is then provided with extensions (25', 25", 25"' ), in the number of 3 for each sector, said extensions opening into the bottom circular surface 23' of the ring. The relevant chamber's openings are located in a determined order and angular positions, among the positions of the admission openings Aa, Ba and of the exhaust openings Ae, Be in such a way that said toroidal expansion chamber enters in communication with the various cylinders, as the ring slides over them, during predetermined time intervals coinciding with the appropriate time phases of the 6-phases thermodynamic cycle illustrated in figure 11.

Such a synchronism can be verified by observing on figure 14 the phases indicated at the external periphery of the ring and confronting them with the contemporary travel direction of the piston, as indicated at the internal periphery of the ring, as the ring slides over a single cylinder (in the example the cylinder 1 and tits port P1).

It will be understood that, as the various openings pass over the port P1 of a cylinder head, the partially expanded gas (point 4' of figure 11) is pushed by the piston to transfer from one cylinder, trough the toroidal chamber, to two cylinders which are diametrically opposed within the circular row of the barrel engine, whereby said same chamber IPTC performs also the function of a buffer for the intermediately expanded fluid. A careful analysis of figure 13 allows to understand also how each single cylinder becomes progressively the seat of all the 6-phases of the thermodynamic cycle of figure 11 , as the series of openings of a single sector (A or B) progressively slides over its port P1 in perfect co-ordination and synchrony with the other cylinders. It will be understood that such an arrangement allows doubling the volume of expansion with respect to that of the compression, which in the conventional crank and piston configuration could be obtained by the virtual but impossible prolongation or duplication of the expansion travel of the piston. A second very interesting embodiment of the distribution ring, which is defined by claims 6 and 9, is adapted to govern a 12-stroke cycle which offer the possibility of accomplishing another thermodynamic cycle composed by 8 phases, as represented in figure 12.

Such a distributor ring is very similar to that illustrated in figure 14 (see also sect. B -B in this figure). More specifically, this ring is also formed by a number S-ι₂ of identical sectors (e.g.: Sι₂=(N+1)/6= 2 for an engine of N=11 cylinder) but each sector is additionally provided with a second intermediate pressure chamber , such as to form a second toroidal "intermediate pressure compression chamber (IPCC)" inside the distributor, when said S sectors are assembled together (see also figurel). Said compression chamber is also provided with three extensions (24',24"₎24"'), for each sector, which open into said bottom circular surface 23' of the ring, at an appropriate angular position to allow the compressed gas flow to enter and leave the cylinders during the first six strokes of each piston, in phase with the following six strokes (total 12-stroke) which are then responsible for the double expansion phase, as already explained above, with the figure 13, in the case of a 6 phase thermodynamic cycle. In fact, similarly to the above described "toroidal expansion transfer chamber" this additional IPCC compression chamber is enabled to transfer the fluid, compressed to an intermediate pressure, from one cylinder to two cylinders which are diametrically opposed within the cylinders circular row of the barrel engine, whereby the compression chamber performs also the function of a buffer for the intermediately compressed fluid.

The resulting "double compression phase" is represented by the lines between the points 1 and 2' (first compression) and 2"-2 (second compression) in figure 12. In order to better define some further interesting features of the present invention, it is convenient to establish a conventional denomination for certain geometrical references of the present basic mechanism and its wobbling movement, see figs 1 , 2 and 6: a) - The static equatorial plane E is the virtual plane in sympathy with the casing of the mechanism, perpendicularto its C/L and containing the centre point O of the spherical movement. b) - The mobile equatorial plane W is the virtual plane in sympathy with the moving wobbling unit, perpendicular to the axis (w-w) of the same unit and also passing through the centre point O of the mechanism. c) - The equatorial straight line R is the intersecting virtual line of the planes E and W, which also goes through the centre point O.

It will be understood that during the wobbling movement said straight line R rotates on the plane A around the point O at the same speed Ω_M of the engine's main crankshaft.

It will also appear that every point P (r,λ) solid to the wobbling unit (where r and λ are polar coordinates with respect to the centre O and the axis w-w of the wobbling unit) describes on a sphere of radius r, for every revolution of the crankshaft, a cyclic trajectory which varies from a circle, when λ = 0°(i.e. P lays on the axis w-w), to a kind of lemniscate when λ = 90° (i.e. P lays on the plane W), with all sort of intermediate oval shapes when the points P assumes intermediate values of the coordinate λ.

It is evident that such a lemniscate, whose projection on a vertical cylinder, is reported in figure 13, has the slimmest profile amongst all possible trajectories (for a constant radius r and for a constant tilt angle α of the Z-crankshaft) when.

Supposing the axis (c-c) of the mechanism in a vertical direction, one can define the slimness of the trajectories by the factor g/s = max. width / stroke length of said trajector , that is by the ratio of maximum horizontal versus maximum vertical displacement of the point P, over one complete cycle (see figure 15. It is also interesting to observe that said factor g/s decreases, at a constant radius, with the value of the tilt angle α. This means that by selecting a very low value of tilt angle α one and correspondingly positioning the top part of a piston with respect to the equatorial plane E the lateral gap between a wobbling piston (in sympathy with its guided movement by the wobbling unit) and the wall of the corresponding fixed cylinder can be kept to very little value relative to the stroke of the piston, for example in the order of 6.5% for α = 15°, and down to 4% for α = 8° see figure 15, to be compared to the 100% of the circle trajectory followed by rod base in the classic rod /crank configuration.

This observation enable the design of an integral wobbling unit (IWU), that is a wobbling unit which integrally incorporates also the pistons (see the schematic design in Figs 1 and 5), wherein the heads of the piston are sensibly disposed at the level of the plane W of the wobbling unit, and wherein the sealing function between pistons and cylinders can be relied upon simple circular segments carried, in a floating but tight contact, by the lateral surface of the pistons. Should the sealing problem reveal itself to be technologically critical and/or not sufficiently effective, the problem can be relaxed by providing a swivel joint at the foot of the piston rod, instead of a fixed connection, for example, by means of a rod-end-bearing.

A further preferable characteristic is formed by a second couple of bevel gears, the vertex of which is also placed at the centre O of the engine. The duty of said bevel gears is that of taking the torque (thus the power) out of the wobbling unit and transmitting it to an output shaft co-axial, but distinct from the main engine shaft. The cited torque is that generated by the cyclic succession of forces exerted by the pistons in their expansion strokes all around the periphery of the wobbling unit. The first (28) of these bevel gear is an internal bevel gear mounted under the bottom of the wobbling unit and it meshes progressively over a second external bevel gear or conical pinion (29) which is rotatably supported by the casing (see figure 1 ) around the centreline C/L of the engine and coupled to the output shaft of the engine. This first internal gear 28, fixedly and co-axially mounted to the WU (15), has a primitive cone with an angle β at its vertex, said vertex being placed on the centre point O, and the second (external) pinion gear (29), rotatably mounted around the C/L on the bottom part of casing (12), has a primitive cone with an angle γ at its vertex which coincide with the vertex of the first gear (28) on the centre point O of the engine so that the are in an appropriate meshing condition, during the wobbling movement of the WU. Of course the value of the angles β and γ must verify certain following relations: β - γ = α and Ωp Ω_M - (sin β-sin γ) / sin γ which are easily derived from the diagram of figure 2 , and wherein

Ω_P Ω_M is a desired value of a reductive transmission ratio between the wobbling frequency, i.e. the main-shaft rotary speed: Ω_M, and the pinion's (29), i.e. the engine-output-shaft's, rotary speed: Ω_P. It is possible to recognise, analysing figures 1 and 5, that with such a disposition, said second couple of bevel gears (28, 29), is adapted to extract the forces, generated by the pistons, out of the WU and to transform them into a couple on the axis of the pinion, i.e. onto the output-shaft (30) of the engine, advantageously discharging the engine's main crank-shaft (17+19+21) from all torque stresses. Summarizing the above description, the suggested engine configuration comprises the following major and essential components see figures 1 and 5:

A) - a Z-crankshaft (or main engine shaft 17+19+21 )

B) - a substantially circular wobbling unit 15 (WU), carrying at its peripheral circle an uneven number of pistons (integral or movably jointed to the WU)

C) - a substantially barrel-shaped casing 12 incorporating a corresponding number of cylinders and cylinder's heads

D) - a fluid distributor ring 23, rotatably mounted around the C/L and disposed immediately above the cylinder heads

E) - a couple of identical flat bevel gears (26+27), one opposed to the other, which are to ensure the WU does not rotate around the C/L, and, as non essential but much preferred component:

F) - a second couple of bevel gears (28+29), for the extraction of the torque (or power) developed by the engine. It is to be underlined here that only two of the above essential components, or parts, are rotating when the engine is in operation: part A) and part D), whereby A) rotates at the nominal engine speed (e.g. Ω = 5.000 rpm) and D) rotates at a drastically reduced speed: ω_d - Ωl N+1. At the same time, the part B) is guided to a wobbling movement at a cycling frequency Ψ equivalent to Ω (i.e. in the example: Ψ = 5.000 cycle/min). All the remaining parts are static parts.

This constitutes a dramatic decrease when comparing with the classic configurations of a 4-strokes-multicylinder engine, whether in-line or V- configuration or even in a star configuration.

The following considerations are also to be noted, which add to the list of advantages and improvements to be obtained with the present invention: a) the Z-crankshaft even for a very high number of cylinder (say N =15), needs only two short crank-levers and two bank-bearings, whereas a comparable classic 12 cylinder engine needs 24 levers and at least 6 bank bearings. b) The centrifugal forces on said crank shaft are reduced in proportion to the reduced radial dimensions, and the interacting forces among the various pistons are integrated within the wobbling unit, which transmit to the shaft only the compensated resultant of all the piston's forces. Moreover, if the presence of the second couple of bevel gears (28 and 29) is selected, the main shaft is substantially if not totally relieved from the torque stresses and from all sorts of torsional vibrations which represent a major problem in the classic crankshaft of multi-cylinders engines. c) The wobble unit can be considered as a three-dimensional structure which is far more rigid compared to the mono-dimensional structure of the classic crankshaft.

Consequently the WU can support and integrate the internal stresses much more efficiently and can smooth out all sources of vibrations. Moreover, considering the absence of centrifugal forces in the pistons and in the WU it is predictable that such a three-dimensional structure can accept a very high rotational regime with a comparatively lower stress and vibrational levels, so as to allow a considerably lighter design with a consequent saving in the overall weight of the engine. Thus, the engine of the invention can be expected to reach much higher rotational speed than the conventional engines d) The perfect angular symmetry of the configuration allows for a modular construction, with a basic angular sector or module spanning an angle of 360°/N each representing a cylinder. This simplify the manufacture and the assembly of the engine. e) The annular shape of the distributor ring, which integrates in a single piece the function of managing the gas admission and exhaust of the gases in all the cylinders in circle, can be efficiently coupled to a fan or a compressor wheel at the centre of the ring concentric to, and driven by the main shaft so as to enhance a pre-com pressed four stroke cycle, without turbine. f) In view of the fact that movement of the pistons is exclusively determined and guided by the WU, the piston's side forces insisting against the cylinder's wall are totally eliminated (or substantially minimized in case of a swivel joint between the piston and WU) with the consequent drastic decrease of the friction drag and parallel increase of engine organic efficiency. g) An overall volume saving can be obtained compared to any classical configuration, thanks to the compact and integrated geometrical combination of the cylinder, the WU and the distribution ring which requires a very flat space over the head of the cylinders. A rough calculation allows to predict a volume (i.e. weight) saving of 40 to 60 % depending on the size of the engine total displacement. Although the invention has been described in terms of specific embodiments and applications, the persons skilled in the arts could arrive, in the light of this teaching, at further embodiments or applications without departing from the scope of the invention, which scope is defined by the following claims.

Claims

1. A Multi-cylinder barrel engine having a centre line (C/L), and a centre point O, comprising:

- a casing (12) on which a plurality (N) of cylinders (11 ) are paralleling disposed in a circular row around C/L;

- a corresponding number (N) of pistons (13), each mounted inside each of said cylinders; - a crankshaft (16) comprising two co-axial end sections (17,18) and an intermediate section (19) whose axis (w-w) intersects the common axis (C/L) of the end sections at the centre point O with a tilt angle α , whereby said crankshaft is rotatably mounted on the casing through a first couple of bearings (20,20') operating between the casing and the two end sections, the latter being solidly connected to the extremities of the intermediate section by at least one connecting crank (21 );

- a wobbling body (15), substantially axial-symmetric, rotatably mounted, through a second couple of bearings (22,22'), on the crankshaft's intermediate section (19) so that its axis (w-w) coincides with the central axis of said body, said pistons being connected and carried at the peripheral edge (14) of said wobbling body, whereby said pistons and wobbling body connected together form a single piece called Wobbling Unit (WU);

- a first couple of flat bevel gears (26, 27) having identical primitive cones with a vertex angle μ = π - α/2, said gears being fixedly mounted the one (26) on the casing (12) and the other (27) on the WU, in opposed position and such that the vertexes of their primitive cones coincide with the centre point O of the engine, whereby the axis of the one gear (26) coincides with the C/L, and the axis of the other gear (27) coincides with the common axis (w-w) of the WU and intermediate section (19), said first couple of gear providing, by their meshing, a constraint against the rotation of the WU around the C/L, characterized in that: - a synchronized rotary distributor (23) is provided, having substantially the shape of a ring, being rotatably mounted about the C/L of the engine, thus co-axially mounted with the N cylinders row, so as to slide over the heads (H₁,...,H_N) of the cylinder and being adapted to contemporary control the cyclical and intermittent flow of the operative fluid into and out of said N cylinders, through corresponding ports (Pi, .... P_N) provided in said heads, its rotational speed (ω_d) being opportunely synchronised with the engine speed (Ω ) so as to allow the performance of a four-stroke thermodynamic cycle.

- said rotary distributor ring is mounted directly over the heads of the cylinders in a sliding contact therewith, along a plane perpendicular to the central line, so that as the distributor rotates, each of said admission or exhaust ports comes in register with, and successively swings over, to alternatively open and close said ports provided in the cylinder heads, thereby acting as a series of admission or exhaust valves for the cylinders. - the tilt angle α has a value smaller than 15°, preferably smaller than 8° - said cylinder heads are provided with one single port (Pi P_N) for the input or the output of compressible fluid (air or gases respectively), - said rotary distributor is composed by a number S of identical circular sectors (A, B, C,...) integrally formed together, each comprising at least one admission opening (Aa, Ba, Ca, ...) and one exhaust opening (Ae, Be, Ca...) as well as relevant ducts (24, 25) leading to admission and exhaust collectors respectively, said number S depending on the number N of the cylinders.

2. The Multi-cylinder engine according to claim 1 , further characterized in that: - said number S depends on the number N of the cylinders according to the relation: S = (N + 1)/2 ,

- said distributor rotates in the same direction (e.g. clockwise) of the crankshaft,

- the rotational speed (ω_d) of said distributor is synchronized, through a reduction gear mechanism, with the engine's speed ( Ω_M ) according to the relations: ω_d = + Ω_M / (N + 1), where N is an uneven integer > or = 3.

3. The Multi-cylinder engine according to claim 1 , further characterized in that:

- said number S of identical sector depends on the number N of the cylinders according to the relation: S = (N - 1)/2 ,

- said distributor rotates in the reverse direction (e.g. anti-clockwise) of the crankshaft,

- the rotational speed (ω_d) of the ring distributor is synchronized, through a reduction gear mechanism, with the engine's speed (ΩM) according to the relations: ω_d = - Ω_M / (N - 1), where N is an uneven integer > or = 3.

4. The Multi-cylinder engine according to claim 2 or claim 3, further characterized in that a second couple of bevel gears is provided which is formed by:

- one conical internal gear (28) which is fixedly and co-axially mounted to the WU (15), and has a primitive cone with an angle β at its vertex, said vertex being placed on the centre point O, and

- one conical pinion gear (29) which is rotatably mounted, around the C/L, on the bottom part of casing (12) and has a primitive cone with an angle γ at its vertex said vertex being placed on the centre point O, whereby the couple of values of said angles is selected so as to verify the following relations : β - γ - a and Ωp /Ω_M = (sin β-sin γ) / sin γ , where Ω_P /Ω_M is a desired value of the transmission ratio between the wobbling frequency, i.e. the main-shaft rotary speed (Ω_M), and the pinion's (29) rotational speed, i.e. the eπgine-output-shaft's speed (Ω_P), said second couple of conical gear being so adapted, by meshing, to extract forces, generated by the pistons, out of the WU and to transform them into a couple on the axis of the pinion, i.e. onto the output- shaft of the engine.

5. The Multi-cylinder engine according to claims 2 or 3 or 4, further characterized in that the design and the position of the WU with respect to the casing and the cylinders are such that the upper heads of the pistons describe, during their strokes, a trajectory which is crossing twice the static equatorial plane (E) of the engine, thus allowing to solidly connect the pistons to WU.

6. A rotary distributor (23), having substantially the shape of a ring, capable of being rotatably and co-axially mounted over the cylinder-head of a barrel engine, as claimed in claim 1 , thus controlling the intermittent fluid flow in and out of the N cylinders of said engine, by sliding with its planar circular bottom surface (23") over and along a common top surface plane (11') of the cylinders heads,

wherein said distributor ring comprises a determined number S of identical circular sectors (A,B,C,D,E, ...), each of which includes at least two openings (Aa, Ae) in said bottom surface (23') and relevant admission and exhaust ducts (24, 25) which when brought in register with each of the engine cylinders, are able to act as an admission or an exhaust valve, in such a synchronous sequence as to allow an at least 4-stroke thermodynamic cycle to be performed by said engine, the numbers having an integral value of: S= (N+1)/(C/2) where N is the number of cylinders in the engine and C is the number of the full strokes necessary to each piston in order to perform a predetermined thermodynamic cycle, and wherein:

- in the event of C = 8 or more, an expansion/transfer duct is provided within each sector, these ducts being identically shaped in all S sectors so that a toroidal expansion transfer chamber (IPTC) is formed inside the distributor ring (23) formed by said S sectors assembled together, and being further provided with extension ducts (24ip, 24a', 24a"), in the number of three for each sector, whereby said extensions open into said bottom circular surface (23') in correspondence to such angular positions, among the positions of said two openings (Aa, Ae), that said expansion chamber is enabled to transfer the fluid, expanded to an intermediate pressure, from one cylinder to two cylinders which are diametrically opposed within the circular row of the barrel engine, whereby the said expansion transfer chamber (ETC ) performs also the function of a buffer for the intermediately expanded fluid; - in the event of C = 12, or more, a compression duct (IPCD ), is additionally provided within each sector, which further ducts being identically shaped in all S sectors so that a toroidal compression transfer chamber (IPCC ) is formed inside the distributor ring (23) formed by said sectors S assembled together and being further provided with extensions (25a^,,25a",25"')_! in the number of three for each sector, whereby said extensions open into said bottom circular surface (23') in correspondence to such angular positions, among the positions of said openings (Aa.Ae), that said compression chamber (CTC) is enabled to transfer the fluid, compressed to an intermediate pressure, from one cylinder to two cylinders which are diametrically opposed within the cylinders circular row of the barrel engine, whereby the compression duct has also the function of a buffer for the intermediately compressed fluid.

7. The rotary distributor according to claim 6 characterised in thatthe numbers is equal to (N+1 )/2, i.e. C = 4 , and each of the S sectors comprises two ports, an admission opening (Aa, Ba, ...)and an exhaust opening (Ae, Be, ...), so as to allow the engine to perform a 4-stroke cycle.

8. The rotary distributor according to claim 6 characterised in that the numbers is equal to (N+1 )/4, i.e. C=8, and each of the S sectors comprises 6 ports: one low pressure (LP) admission port (Aa, Ba,...), one intermediate pressure (IP) transfer ports (24'), two IP admission ports (24",24"') and two LP exhaust ports (Ae',Ae", Be', Be",...), the distributor ring (23) further comprising, internally all around the ring, the toroidal expansion chamber (ETC) which is in communication with the IP transfer opening and with the two IP transfer/admission openings, so as to allow the engine to perform an 8-strokes cycle, (i.e. a 4-strokes cycle with double expansion strokes and two transfer strokes).

9. The rotary distributor according to claim 6 characterised in that the number S is equal to (N+1)/6, i.e. C=12, and each sector comprises 10 ports: two low pressure (LP) admission ports (Aa'.Aa"), two intermediate pressure (IP) compression-transfer openings, one IP transfer-admission opening, one IP exhaust-transfer opening, two IP transfer-expansion openings and two LP exhaust openings, the ring distributor internally comprising:

- a toroidal compression chamber (IPCC) for the IP compressed fluid, which is in communication with the two IP compression-transfer openings and the IP transfer- admission opening, and

- a toroidal expansion chamber (IPTC) for the IP exhausted fluid, which is in communication with the IP transfer-exhaust opening and with the two IP transfer- expansion openings, so as to allow the engine to perform a 12-stroke cycle, (i.e. a 4-strokes cycle with two compression strokes and two expansion strokes plus four transfer strokes).