A loop-scavenged two-stroke internal combustion engines with an intake valve (7) engaging a seat (10) for fresh air intake, and an exhaust valve (8) engaging a seat (13) for combustion gas exhaust, is disclosed. The valves are arranged in such a way that the fresh air intake scavenges a substantial part of the burnt gases. In at least one of the valves, the valve surface (21) located downstream from the valve face (9) and the surface (23) of the downstream extension of the seat (10) are configured in such a way that they form a substantially isentropic diffuser.

FIELD OF THE INVENTION
 The present invention concerns an improvement to two-stage internal
 combustion engines with loop scavenging of the type
 having at least one variable volume working chamber delimited by a
 cylindrical wall in which a piston slides, the mobile top face of the
 piston and a fixed cylinder head,
 operating in accordance with the two-stroke cycle, with a loop scavenging
 system via the cylinder head, controlled by at least one inlet valve
 cooperating with a seat, preferably of generally conical shape, to cause
 the working chamber to communicate cyclically with an inlet cavity
 communicating with means for supplying air to the engine and by at least
 one exhaust valve cooperating with a seat, preferably of generally conical
 shape, to cause the working chamber to communicate cyclically with an
 exhaust cavity communicating with the combustion gas exhaust system of the
 engine,
 and in which said inlet and exhaust valves are disposed so that the air
 entering the working chamber through the inlet valve causes scavenging of
 at least a substantial part of the burned gases in the chamber and their
 evacuation via the exhaust valve.
 BACKGROUND OF THE INVENTION
 In engines of the above type the difference AP in the gas pressure between
 the means supplying the engine with air at pressure P and the engine
 combustion gas exhaust system is relatively low and in practice is imposed
 by the specifications of the supercharged air supply means.
 Scavenging can take place only during a limited part of each cycle, another
 and large part of the cycle being devoted to compression and expansion of
 the gases renewed in the chamber.
 As a result the geometry and the operation of the inlet and exhaust valves
 play a decisive role in the efficiency, the power and the speed of the
 engine.
 Increasing the size of the valves rapidly runs up against a geometrical
 limit imposed by the dimensions of the cylinder head while increasing the
 valve lift, that is to say the distance the valve moves away from its
 seat, and the lift speed, which are determined by the profile of the cams
 opening and closing the valves, is rapidly limited by mechanical
 constraints imposed by the permissible contact pressure between the nose
 of the cam and the components that it actuates.
 This limits performance, i.e. the permeability of the cylinder head, the
 efficiency of use of the air passing through the cylinder head, i.e. the
 ratio between the mass of air enclosed in the working chamber at the end
 of scavenging to the mass of air passing through the cylinder head, and
 the scavenging efficiency, i.e. the ratio between the mass of air and the
 total mass of gas enclosed in said chamber at the end of scavenging.
 Patent application EP-A-0 673 470 (or U.S. Pat. No. 5,555,859 or WO
 95/08052) has made it possible to improve significantly on the above
 limitations by providing a single inlet valve and a single exhaust valve,
 said inlet and exhaust valves being coaxial circular cylinders, preferably
 coaxial with the cylindrical wall of the working chamber, the coaxial
 arrangement being such that the inlet valve is outside the exhaust valve,
 the seat of the inlet valve being attached to the cylinder head and
 oriented so that the pressure of the drive fluid contained in the working
 chamber exerts a force that tends to press said valve onto its seat, said
 seat being in the immediate vicinity of the periphery of the upper part of
 said cylindrical wall inside which the piston slides, and in contact with
 the cylinder head,
 said exhaust valve having a tubular part the inside wall of which slides on
 a fixed hub carried by the cylinder head, to which it is sealed by sealing
 means, and the end of which towards the chamber has a bearing surface
 coaxial with said tubular part so that it can cooperate with a seat
 provided inside the lower part at the same end of the chamber as said
 inlet valve, enabling communication of said exhaust cavity with the
 working chamber, by virtue of the annular space delimited radially by the
 inside wall of the inlet valve and by the outside wall of the exhaust
 valve.
 This arrangement optimises and controls scavenging and doubles the actual
 lift of the exhaust valve because the inlet and exhaust valves lift in
 opposite directions.
 If means are provided to cause the inlet air passing through the inlet
 valve to rotate, axi-symmetrical centrifugal layering can be achieved and
 the fuel can be injected into a hot central area that is relatively
 impoverished in oxygen, to obtain the advantages described in the above
 patent and in patent application FR-A-2 690 951.
 BRIEF SUMMARY OF THE INVENTION
 The present invention proposes to improve further the performance of
 engines having concentric inlet and exhaust valves, in particular as
 defined above, enabling a choice between reducing the scavenging time and
 consequently increasing the usable expansion stroke of the engine and
 therefore its efficiency, or, for a given angular duration of scavenging
 during the upward travel of the piston, a high permeability of the
 cylinder head enabling the rotation speed of the engine to be increased.
 Another objective of the invention is to reduce significantly wear of the
 valves and their seat and in particular the inlet valve and its seat, the
 exhaust valve and its seat being generally better protected by the effects
 of lubrication by the carbon deposits caused by combustion gases moving
 towards the exhaust.
 The invention consists in a two-stroke internal combustion engine with loop
 scavenging, preferably of the compression ignition type, of the kind
 described in the preamble, and preferably in which the inlet and exhaust
 valves are coaxial circular cylinders, preferably coaxial with said
 cylindrical wall, preferably with the inlet valve outside the exhaust
 valve, and preferably having the other features of the valves of
 concentric valve engines of the type described hereinabove, characterised
 in that, for at least one of said inlet and exhaust valves the surface of
 the valve downstream of the bearing surface of said valve, in the
 direction of flow through it, on the one hand, and the surface of a part
 extending the seat of said valve, with which said bearing surface
 cooperates, and also situated downstream of said seat, on the other hand,
 are configured to constitute a substantially isentropic diffuser
 discharging into the cavity downstream of the valve.
 "Isentropic diffuser" means a divergent nozzle in which the flow of gas
 through the nozzle is slowed and compressed virtually isentropically.
 The outlet into said downstream cavity has a discharge section greater than
 the geometrical flow section of the valve in the fully open position
 between the bearing surface of said valve and its seat.
 "Geometrical flow section of the valve" means the minimum unrestricted flow
 section between the lifted valve and its seat. This section remains in the
 vicinity of the bearing surface, i.e. the areas of contact between the
 closed valve and its seat, but its axial position can vary with the lift
 of the valve.
 Accordingly, flow downstream of the valve is effected in a diffuser whose
 section increases progressively, at least on approaching the cavity
 downstream of the valve, i.e. the working chamber in the case of an inlet
 valve and the exhaust cavity in the case of the exhaust valve, the
 discharge section of the diffuser, i.e. the section through which the
 diffuser opens into said cavity, being greater than the flow section of
 the valve in the fully open position between the bearing surface of the
 valve and its seat.
 In a preferred embodiment, the ratio between the discharge section where
 the flow from the valve enters the cavity downstream of the latter, in the
 direction of flow, and said geometrical flow section of the valve in the
 fully open position is at least equal to the critical ratio calculated for
 the value of the ratio of the pressures of the fluid flowing in said valve
 on either side of the latter during normal operation of the engine.
 The critical ratio is defined as that for which the speed of the flow
 reaches the speed of sound at the throat of the fluid flow in the vicinity
 of the valve bearing surface; it can easily be calculated using the
 equations of isentropic diffusion.
 If the valve forming the diffuser is a tubular inlet valve opening into a
 cylindrical chamber, the surface profile of the valve downstream of its
 bearing surface is preferably configured so that it progressively becomes
 substantially parallel to the direction of the cylindrical wall of the
 chamber in which the piston slides.
 The bearing surface of the inlet valve, in conjunction with the seat, can
 then advantageously define an angled path progressively widening in the
 meridian plane and with advantage initially oriented outwards, i.e.
 towards the wall of the chamber, so as progressively to become parallel to
 the wall of the chamber.
 In the case of a tubular exhaust valve the profile of a part of the valve
 downstream of its bearing surface is preferably configured at its outlet
 so as to be substantially parallel to the interior cylindrical part of the
 inlet valve that forms the seat of the exhaust valve.
 In a preferred embodiment deflector means are provided in the air supply
 passage that delivers air to the tubular inlet valve to impart to the flow
 of air a rotary component so as to send through the flow area of the valve
 and then the part forming the diffuser a substantially isentropic flow of
 air with a rotary movement procuring a centrifugal effect tending to hold
 the air against the wall of the cylindrical chamber in which the piston
 slides to obtain the advantages described in patent application FR-A-2 690
 951.
 The access passages to the inlet valve are preferably inclined to the
 geometrical axis of the cylinder at the exit from said inlet cavity and
 towards the piston to reduce deflection of the flow in a meridian plane in
 order to minimise head losses.
 Deflector means such as fixed deflector blades can be disposed either in
 these passages or even on the inlet valve itself, if necessary. By
 increasing the effective permeability of the cylinder head by virtue of
 the isentropic diffusion of the flow, the invention limits the
 disadvantageous head loss that inevitably results from the rotation
 imparted to the air by the deflector means, which are generally inclined
 at an angle near 45.degree..
 The invention also consists in an engine as defined in the preamble
 preferably including a tubular inlet valve having a bearing surface that
 is preferably generally conical in shape cooperating with a valve seat
 carried by the cylinder head, the bearing surface of the valve bearing on
 the seat along a circular line in a plane perpendicular to the axis of
 translation of the valve, characterised in that the valve and the seat
 downstream of said circular line of bearing engagement between the bearing
 surface and the seat as defined when the pressure in the chamber is low or
 nil are adapted so that on the occasion of cyclic deformation of the valve
 by forces due to the pressure of the gases the diameter of the circular
 line of contact decreases so that the bearing surface of the valve pivots
 about its bearing engagement with its seat and rolls without sliding on
 the latter.
 The deformation of the valve due to the action of the pressure of the gases
 can then be exploited with advantage to prevent sliding leading to wear of
 the bearing surface of the valve by achieving an effect of rolling against
 the surface of the seat, by virtue of the deformation of the valve, when
 the pressure increases as a result of compression and then combustion.
 This rolling contact without sliding is possible only because of the hollow
 structure of the valve on each side of the line of contact.
 To achieve this result the conjugate surface at the bearing surface and
 preferably also the surface at the seat advantageously have profiles
 having a point of inflection, the line of contact, i.e. of bearing
 engagement, moving in the vicinity of this point of inflection when the
 pressure varies.
 With surfaces of the above kind, for example S-shape surfaces, when the
 valve is not loaded by the pressure, or only slightly loaded, the line of
 contact is below the point of inflection. When the valve is highly loaded,
 when the gases are at the maximal pressure, for example, it is on or
 slightly above the point of inflection.
 The conjugate surfaces are preferably adapted to form, downstream of the
 bearing surface and the seat in the direction of flow of the fluid, a
 diffuser procuring substantially isentropic flow as defined hereinabove.

DETAILED DESCRIPTION OF THE INVENTION
 Refer first to FIGS. 1 and 2.
 The prior art engine described in application EP-A-0 673 470 is a
 two-stroke diesel engine with loop scavenging comprising a cylinder 1 in
 which slides a piston 2 and which is closed at the top by a cylinder head
 5.
 The cylinder, the piston and the cylinder head delimit a variable volume
 working chamber 3 in which combustion takes place when the piston is near
 top dead centre, as shown in FIG. 1.
 The cylinder head 5 has a circular cylindrical fixed central hub 6 attached
 to it the axis 23 of which is preferably coincident with that of the
 cylinder and the piston and inside which there is a fuel injector, not
 shown, on the axis of said hub and discharging along the axis of the
 combustion chamber 4 forming part of the working chamber 3.
 The engine also includes a generally tubular inlet valve 7 and a generally
 tubular exhaust valve 8, said valves being concentric along the common
 axis 23 and the exhaust valve being inside the inlet valve. The inlet
 valve 7 has at the bottom a bearing surface 9 cooperating with a seat 10
 formed in the lower part of the cylinder head 5. Concentric with and
 outside this bearing surface, an inlet cavity 11 distributes air to the
 valve from a conventional air supply device (not shown). The cavity 11
 advantageously communicates with the inlet valve 7 via passages 12
 oriented to impart to the inlet air flow a rotation component about the
 common geometrical axis 23 of the various components of the engine.
 These passages can be delimited by two coaxial conical surfaces in the
 cylinder head, fixed deflector blades being disposed as close as possible
 to the exit into the working chamber.
 The tubular inlet valve 7 has a circular cylindrical inside surface,
 preferably with a conical surface in its lower part coaxial with the axis
 23 of the valve and forming a seat 13 that cooperates with the bearing
 surface 14 of the tubular exhaust valve 8.
 The tubular inlet valve 7 also has a cylindrical tubular body 15 which
 slides in a bore in the cylinder head 5. The tubular exhaust valve 8 also
 has a cylindrical tubular body which slides on the fixed central hub 6.
 Oil passages are formed between the inside cylindrical surface of the body
 of the exhaust valve 8 and the outside cylindrical surface of the fixed
 central hub 6 to enable cooling and lubrication of the facing components.
 Because of their different diameters, the two coaxial circular tubular
 valves delimit an annular passage 16 through which the exhaust gases are
 conducted from the working chamber 3 to the exhaust cavity 17 which
 communicates with the exhaust system, not shown, of the engine.
 The two valves 7 and 8 are operated hydraulically, as described in
 application EP-A-0 673 470, by virtue of cyclic variations in the pressure
 of a hydraulic liquid enclosed in two cavities 50 or 60 of constant volume
 but having variable surface areas delimited by a drive piston 51 or 61
 cooperating with a camshaft 53 coupled to the main shaft of the engine and
 by a receiving piston 52 or 62 respectively attached to the inlet valve
 and the exhaust valve, and which includes adequate return spring means.
 The receiving pistons 52 and 62 are disposed so that the inlet and exhaust
 valves are actuated in opposite directions, the tubular inlet valve 7
 opening downwards, i.e. towards the piston, and the tubular exhaust valve
 opening upwards.
 With the bearing surface 14 of the exhaust valve cooperating with its seat
 13 on the inside surface of the inlet valve, and with the two valves
 moving in opposite directions, the opening of the exhaust valve is clearly
 increased by the lift of the inlet valve.
 An engine of the above type operates in the following manner:
 When the piston 2 is propelled towards bottom dead centre by the gases in
 the working chamber after combustion of the fuel, and therefore at the end
 of expansion of the working chamber, the exhaust valve is opened to enable
 the pressure in the working chamber to fall below the pressure in the
 inlet cavity 11 ("exhaust puff") to prevent the gases flowing back towards
 the inlet circuit ("counter-scavenging"). The inlet valve is then opened
 to scavenge the working chamber, which consists in substituting air for
 the combustion gases.
 The inlet air enters the flow section of the inlet valve delimited between
 the bearing surface 9 and the seat 10, having had rotation imparted to it
 previously in said passages 12. The air therefore enters the working
 chamber in the space delimited by the lateral wall of the cylinder and the
 lower part of the valve. Because of the rotation of the air about the axis
 23, the air streams entering the working chamber are inclined to said axis
 23, forming a layer of air along the cylindrical wall moving towards the
 piston and rotating about this axis.
 At the same time the hot gases which are concentrated near the axis of the
 chamber 3 and of the combustion chamber 4 escape via the flow section
 between the seat 13 and the bearing surface 14 of the exhaust valve.
 During the first phase of the upward movement of the piston 2 the
 combustion gases are therefore largely evacuated and replaced by air. In
 the second part of the upward movement of the piston 2 the valves are
 closed and all of the gases contained in the chamber 3 are then
 progressively compressed to the state of maximum compression in the
 combustion chamber 4 into which fuel is then injected under pressure,
 which ignites the fuel and starts a new engine cycle.
 The times at which the valves close can advantageously be adjusted to
 obtain inside the available volume above the piston 2 a mass of air that
 is rotating and therefore centrifuged towards the periphery and
 surrounding a smaller mass of hot combustion gas in the central part, from
 the previous cycle and retained in the chamber during scavenging, with the
 result that injection takes place into this central part, which will
 procure the advantages described in patent application FR-A-2 690 951.
 In accordance with the invention, and as shown in FIGS. 3 and 4, the lower
 part of the tubular inlet valve 7 downstream of the bearing surface 9
 cooperating with the conical seat 10 (in the direction of flow of the air
 through it) is extended downwards, i.e. towards the piston, by a skirt 21
 having symmetry of revolution and coaxial with said valve, the outside
 surface of which, i.e. the surface at the greatest distance from its axis
 23, is a circular cylindrical surface the meridian profile of which merges
 at the upstream end tangentially with the surface of the bearing surface
 9. At the downstream end it preferably terminates parallel to the axis 23.
 In the same manner the conical seat 10 is extended regularly by a circular
 cylindrical surface 22 around the axis 23 the meridian profile of which is
 tangential--at the upstream end--to the conical seat and parallel--at the
 downstream end--to the axis 23.
 The facing circular surfaces 21 and 22 therefore delimit an annular passage
 having symmetry of revolution and the discharge section of which increases
 regularly from the minimal flow section Sc (with the valve in the
 maximally open position) called the throat of the valve and in line with
 the seat 10 and the bearing surface 9 to the maximal flow section Sd at
 the bottom of the skirt 21.
 In accordance with the invention, the meridian profiles of the facing
 surfaces 21 and 22 are designed to constitute an "isentropic diffuser"
 when the valve is in the fully open position. By "isentropic diffuser" is
 meant a passage whose flow section, increasing in the direction of flow,
 is such that the flow through it, having been previously accelerated and
 expanded until it passes the throat of the valve, is there decelerated and
 recompressed quasi-isentropically [i.e. with no thermal losses at the wall
 and with conservation of the total pressure (cut-off pressure) all along
 the flow], to the static pressure downstream of said valve.
 The progressive increase in the flow section in the diffuser must be
 neither too small--because friction at the walls would then become
 excessive, leading to a drop in the total pressure of the flow--nor too
 large--because the flow then separates from the wall, also leading to a
 drop in the total pressure. For a conical diffuser, for example, the
 optimal angle characterising the progressive increase in the flow section
 is known to be around 7.degree. to the axis of said cone.
 An arrangement of the above kind has the following advantages:
 The flow, which is deflected relative to the axis of the cylinder on
 passing through the throat of the valve, is regularly straightened so as
 to be directed parallel to the axis of the cylinder towards the piston
 (with a tangential component, if momentum is imparted to this flow on
 passing said valve).
 For a given pressure different AP between the inlet cavity and the
 cylinder, the value of which depends on the efficiency characteristics of
 the turbocharger, and a given temperature T and a given pressure P in the
 inlet cavity 11, the flowrate Q through the valve is increased in the
 ratio S.sub.d /S.sub.c of the maximal discharge section S.sub.d at the
 exit from the diffuser to the minimal section S.sub.c at the throat
 relative to the flowrate Q* that would pass through the same valve without
 its diffuser.
 The increase in the flowrate is nevertheless limited by the fact that,
 expanding on passing through the throat of the valve, the flow is
 accelerated and is then recompressed and decelerates in the diffuser.
 However, for a given section at the throat (determined by the geometry and
 the maximal lift of the valve) there is a maximal value of the outlet
 section (S.sub.d).sub.m of the diffuser for which the speed of the flow at
 the throat of the valve reaches the speed of sound. If the exit section of
 the diffuser were greater than this critical section (S.sub.d).sub.m the
 flow would separate from the wall beyond said critical section and there
 would no longer be an isentropic diffusion of the flow beyond the critical
 section and consequently no increase in the flowrate through the valve.
 The critical section of course depends on the expansion ratio
 .omega.=P(P-.DELTA.P) between the inlet cavity and the cylinder. The
 theoretical expression for this quantity is:
EQU [S.sub.c /S.sub.d ].sub.critical
 ={2/(.gamma.-1).multidot.[(.gamma.+1)/2].sup.(.gamma.+1)/
 (.gamma.-1).multidot..omega..sup.n.multidot.(.omega..sup.n -1)}.sup.1/2
 /.omega.
 with: n=(.gamma.-1)/.gamma. and .gamma.=C.sub.p /C.sub.v
 C.sub.p and C.sub.v being the specific heats at constant pressure and at
 constant volume of the gaseous fluid concerned.
 For example: .gamma.=1.404:

.DELTA.P/P [S.sub.d /S.sub.c ].sub.critical
 0.05 2.228
 0.10 1.622
 0.15 1.366
 0.20 1.222
 This means that the flow through the valve of the invention can be
 increased 62% relative to a conventional valve if the pressure difference
 on either side of the valve is 10% of the total upstream pressure. On the
 other hand, there is no point in increasing the exit section of the
 diffuser beyond this critical ratio, as the flow would then be stuck at
 the speed of sound on passing through the throat of the valve.
 The diffuser is said to be matched to the throat of the valve if the speed
 of sound is reached at said throat and if the flow diffuses reversibly,
 i.e. without separating from the wall, as far as the exit section.
 Thus if the diffuser is matched to the valve in the fully open position
 (with a discharge section increased 62% relative to the section at the
 throat if the pressure difference across the valve is 10% of the pressure
 upstream of the valve) it would clearly not be matched for smaller lifts
 of the valve. If the valve is lifted halfway, for example, the flow will
 separate from the wall of the diffuser in the section of the diffuser
 whose flow section is equal to half its exit section. However, the facing
 profiles 21 and 22 can be organised so that the diffusion is as near
 perfect as possible even during lifting of the valve. This considerably
 increases the velocity of the flow at the throat and this considerably
 increases the effective permeability of the inlet valve for a pressure
 difference between the inlet pressure due to the supercharging means and
 the pressure in the exhaust, which remains constant. The flow of air
 entering via the inlet valve is therefore considerably increased.
 It should also be pointed out that the diffuser is extremely efficient,
 including at the instant the inlet valve begins to move away from its seat
 and the flow area at the level of the seat is still very small. The
 efficiency, i.e. the increase in permeability, is immediately obtained,
 even at low engine operating speeds and when starting the engine, in other
 words at time when, in the conventional arrangement, scavenging is most
 difficult.
 Likewise, in the exhaust valve of the invention, as represented in FIGS. 3
 and 4, the circular cylindrical surfaces constituting the inside wall 25
 of the inlet valve and the outside wall 24 of the exhaust valve, situated
 downstream of the bearing surface 14 of the exhaust valve and the seat 13
 of said valve formed on the inside wall of the tubular inlet valve,
 delimit an annular passage 26. The meridian profiles of these facing
 surfaces 24 and 25 are designed so that the annular passage 26 constitutes
 a divergent passage from the minimal section Sc at the throat of the valve
 to the maximal value Sd where it joins the exhaust passage 17 and so that
 this divergent passage is an isentropic diffuser in the sense defined
 above.
 Similarly, the ratio between the exit section Sd of the diffuser and the
 section Sc at the throat of the valve in the fully open position is
 preferably at most equal to the critical ratio calculated for the nominal
 value of the relative pressure difference across the valve.
 The position of the throat of the valve, defined by the minimal geometrical
 flow section of the annular passage, can vary relative to the position of
 the bearing surface 14 and the seat 13 of said valve and in accordance
 with the degree to which the valve is open.
 In FIG. 3, for example, in which the exhaust valve is shown slightly open,
 it can be seen that the throat of the valve, with the minimal flow section
 Sc, is in the immediate vicinity of the bearing surface 14 of the valve
 and its seat 13.
 In FIG. 4, on the other hand, which shows the two valves in the fully open
 position (and with the flow section of the valve increased because of the
 downward movement of its seat 13 on the inlet valve), it can be seen that
 the position of the throat of the valve of minimal section Sc is well
 above its bearing surface, in the rectilinear part of the annular passage,
 which is highly favourable to obtaining good diffusion of the flow (it is
 well known that it is particularly difficult to obtain perfect diffusion
 in a curved passage).
 The skilled person can determine conjugate profiles 21 and 22 for the inlet
 valve and 24 and 25 for the exhaust valve either by calculation or by
 experiment. The profiles are preferably designed so that the divergent
 passage following on from the throat of the valve constitutes as near
 perfect as possible an isentropic diffuser, in the sense defined above,
 when the valve concerned is in the fully open position. To determine the
 ideal profile it is necessary to take into account the fact that the flow
 has an axial component ("discharge velocity") and a tangential component
 imparted to it on passing the deflector means such as blade 6.
 Refer now to FIG. 5.
 The inlet valve has a large diameter and centres naturally, when closed, on
 its seat on the cylinder head. The smaller diameter exhaust valve bears on
 the conical seat 13 on the inlet valve and can therefore be off-centre by
 a non-negligible amount relative to the central hub 6 on which it slides.
 In the case of large bore engines, given the tolerances for the
 manufacture and stacking of the parts, this eccentricity can be as much as
 several tenths of a millimeter.
 Under these conditions, to assure a good seal between the inside
 cylindrical surface of the exhaust valve 8 and the outside surface of the
 central hub 6, a floating ring 28 can advantageously be fitted that is
 able to move laterally relative to the hub. The floating ring 28 can be
 accommodated in such a fashion as to be able to move laterally with a
 small clearance in a groove formed between a shoulder 29 on the hub 6 and
 a counter-shoulder 36 on a part 31 which is also part of the hub 6, the
 outside cylindrical surface of the ring 28 providing a sliding track for a
 sliding seal or packing 32 in the lower inside part of the exhaust valve
 8. The sliding track obviously has an outside cylindrical surface of
 sufficient height to enable sliding of the packing 32 throughout the
 lifting of the exhaust valve.
 Refer now to FIG. 6.
 Any known valve actuation means can be used to actuate the valves, for
 example the inlet valve 7, such as mechanical actuation by a camshaft, for
 example, or electromagnetic actuation synchronised with rotation of the
 main shaft of the engine. In any event, hydraulic actuation means can
 advantageously be used, as in patent application EP-A-0 673 470, which
 consist in a deformable cavity of constant volume filled with a hydraulic
 liquid such as the lubricating oil of the engine, for example, and having
 a first chamber 34 of variable volume delimited by a cylinder head and in
 which slides an actuator piston 35 cooperating with a camshaft 36 coupled
 to the main shaft of the engine and which communicates via a passage 37
 with a second chamber 38 of variable volume delimited by the bore in which
 the cylindrical outside surface of the inlet valve 7 slides, which has a
 shoulder 39 serving as a receiver piston so that when the nose of the cam
 36 actuates the drive piston 35 the oil, deemed to be an incompressible
 liquid, expelled through the passage 37 from the first chamber 34 into the
 second chamber 38, causes the receiver piston 39 to descend and thereby
 the valve 7 to be opened. Return movement in the upward direction can be
 effected by return means, for example a spring or preferably pneumatic
 return means consisting of the compression of the air contained in a
 cavity 40 one face of which is also delimited by a shoulder 41 of the
 valve 7 acting as a return piston surface. If the volume formed by the
 cavities 34 and 38 and the passage 37 is too full of oil, for example
 after an oil leak into the cavity or thermal expansion of the oil, the
 valve will not drop back onto its seat. If there is insufficient oil, for
 example through leakage to the exterior, contact between the cam 36 and
 the roller of the piston 35 will be interrupted, which will cause impacts
 in the actuating means.
 The invention avoids these drawbacks by means of an automatic device for
 taking up clearance providing a small diameter passage 42 that can
 discharge into the cavity 34 and is connected to the low-pressure oil
 supply means 43, the outlet from the narrow passage 42 into the chamber
 being disposed so that it is cyclically covered and uncovered by the
 movement of the piston 35, its location being such that, when the piston
 35 with its roller is released to return to its initial position after
 actuation by the cam 36, the outlet is uncovered and places the cavity
 filled with oil in communication with said low-pressure oil supply means
 43, whereas it is very quickly covered when the piston 35 begins to be
 moved downwards by the cam to start lifting the valve.
 Refer now to FIG. 7.
 In the engine in accordance with the invention the inlet valve 7 has a
 large area exposed to the combustion gases with the result that it is
 important to cool the valve effectively. Likewise, because of the high
 performance that can be obtained with an engine of the above kind, the
 exhaust valve can advantageously be rigorously cooled.
 In accordance with the invention, the valve can advantageously be made with
 an elongate annular cavity inside it substantially exposing the shape of
 the valve and descending to a point near the free end 27 of the valve 7 so
 as to extend largely inside the cylindrical tubular part 15 of the valve.
 This cavity is partly filled with a fluid 44 that is a good conductor of
 heat, for example sodium which is in the liquid state when the valve has
 reached its operating temperature. In this way heat can be evacuated
 outwards into an area in which it is easy to cool the valve. Moreover, the
 large surface area swept by the air during scavenging enables transfer of
 heat from the inside surface of the inlet valve during combustion to the
 outside surface of said valve during compression. This transfer can be
 obtained either by conduction or by convection in the heat-conducting
 fluid.
 A similar cavity can be provided in the exhaust valve in order to convey
 heat from the lower end of the valve to the water or oil cooling means of
 the valve.
 Note that this technique, using a heat-conducting fluid such as sodium
 partly filling a cavity within the thickness of the valve and more or less
 espousing its outside surface and consisting in extracting heat in the hot
 part of the head of the valve in order to transfer it to the stem of the
 valve where cooling means are disposed, is known in itself but of very low
 efficiency. With a valve of conventional shape there is a disproportion
 between the surface area that receives the heat (the "tulip" or valve
 head) and the surface area where the heat can be evacuated (the valve
 stem).
 On the other hand, with a tubular valve as used in the invention, these
 proportions are reversed and there is a very large tubular surface area
 for evacuating heat extracted from the head of the valve by means of an
 appropriate cooling system.
 Refer now to FIG. 8.
 In a tubular inlet valve, such as a valve 7, for example, associated with
 the tubular exhaust valve 8 of which it carries the seat 13, the forces
 due to the pressure of the gases, especially when the piston is near its
 top dead centre, are exerted on the inside face of the inlet valve, mainly
 facing the chamber 4. The valve part above its bearing engagement with the
 seat 10 will be subjected to tensile stresses and the free end of the
 valve below this bearing engagement will be subject to compression
 stresses. The resultant of these forces is exerted on the valve 7 between
 where it bears on the fixed seat 10 and where it bears on the mobile
 bearing surface 14 of the exhaust valve 8. The combination of this
 resultant of forces due to the action of the gases and bearing engagement
 reaction forces (represented by the solid line arrows) exerts a tilting
 torque (the forces of which are represented by chain-dotted arrows) on the
 inlet valve 7 which therefore pivots about its fixed bearing point, i.e.
 the seat 10, anticlockwise in FIG. 8.
 The conjugate profiles of the bearing surface 9 and the seat 10 of the
 valve 7 can be calculated allowing for the radial compression strength of
 the part of the valve under the bearing surface 9 and the radial tensile
 strength of the part of the valve between the bearing surfaces 9 and 13,
 so that the bearing surface 9 of the valve 7 bears on its seat 10 at a
 circular contact line the plane of which is perpendicular to the axis of
 said valve, and which can roll without sliding of the seat 10 when the
 pressure of the combustion gases cyclically deforms the valve 7. This
 effect of rolling without sliding can be obtained by imparting a rounded
 profile to the surface of the valve at its bearing surface 9 and to the
 surface of the cylindrical head at the seat 10, whether the valve is a
 tubular valve of the type defined in the present invention, that is to say
 one in which the surfaces downstream of the seat form a quasi-isentropic
 diffuser, or a conventional tubular valve.
 In the tubular inlet valve of the invention, which includes a
 quasi-isentropic diffuser downstream of its seat, the fact that the latter
 is hollow and has an S-shape profile on opposite sides of its bearing
 surface 9 lends itself particularly well to obtaining this rolling without
 sliding of the circular line of contact of the bearing surface 9 on the
 seat 10, which line will migrate upwards (i.e. towards the cylinder head)
 in a plane perpendicular to the axis of said valve because of the cyclic
 deformation of the valve by the pressure of the gases in the combustion
 chamber. The conjugate profiles of the bearing surface 9 and the seat 10
 of the valve can be determined by experiment, seeking to minimise friction
 and therefore wear of the parts in contact, or by calculation, using the
 evolution of the thickness and the shape of the valve (and therefore its
 inertia) as a function of the axial position of the section plane and of
 the stiffness of the material from which it is made.