EP1606505A2

EP1606505A2 - Reciprocating engine with burnt gas recirculation, which is designed to drive a motor vehicle, and method of turbocharging said engine

Info

Publication number: EP1606505A2
Application number: EP04742343A
Authority: EP
Inventors: Jean Frédéric Melchior
Original assignee: Individual
Current assignee: Individual
Priority date: 2003-03-26
Filing date: 2004-03-24
Publication date: 2005-12-21
Also published as: WO2004088115A3; US20070271919A1; FR2853011A1; FR2853011B1; WO2004088115A2; US7313918B2

Abstract

The invention relates to a reciprocating engine which is used between a minimum rotation speed Nmin and a maximum speed Nmax, comprising a turbocharging unit (2) which is dimensioned such as to operate autonomously when: (i) supplying air to the intake manifold (8) of the engine via a coolant, (ii) being supplied with gas by the exhaust manifold (9, CR and CT) of the engine at the exhaust temperature, and (iii) the turbine supply pressure (P3) is essentially equal to the compressor discharge pressure (P2). In this way, at a constant air temperature and with fixed geometry, the turbcharging system supplies an essentially-constant volume of cooled air Vc when the pressure varies, and volume Vc is essentially proportional to the discharge section Sd offered to the hot gases. The invention is characterised in that the turbine pressure (P3) is maintained essentially equal to the compressor pressure (P2) by an EGR bypass (3) between the intake manifold (8) and the exhaust manifold (9), the latter being dimensioned such as to transfer the exhaust gas flow towards the intake manifold without any significant pressure drop. Moreover, the volume of cooled air Vc is smaller than the volume ingested by the engine at speed Nmax, such that a hot gas flow is reingested by the engine via the bypass (3) above speed Na, known as the compression adaptation speed, where the ingested volume is equal to Vc, and an air flow is diverted towards the turbine below speed Na.

Description

Recirculating burnt gas reciprocating engine for the propulsion of motor vehicles and turbocharging process for this engine.

In the past, the thermodynamic cycles of automobile engines were optimized for efficiency and specific power.

Today, to the performance and specific power criteria are added the pollution control constraints and particularly the elimination of NOX. These constraints are today limited to the conditions of urban engine use and to low-power road journeys. The foreseeable tightening of the regulations leads to extending the field of use cleaned of the engine.

The clean-up area is today limited to 50% of maximum speed and 50% of maximum torque, the engine being supplied with fresh air in the area of high torques and power.

Many decontamination techniques by gas treatment post released into the atmosphere are used or under development, such as redox catalysts and particulate filters and NOx regenerable. Among the pollutants, the most difficult to treat post in the presence of oxygen are NOX sought to eliminate at source by diluting the fresh air by exhaust gases (EGR) recirculated externally or internally recycled. To sufficiently limit the flame temperature, the mass rate of EGR must reach 50% of the thermodynamic mass present in the cylinder.

The disadvantage of this process is that it reduces the volume available in the cylinder by 50% to accommodate the fresh air necessary for combustion. The restoration of power therefore requires double the pre compression of the oxidizer load turbocharging. In addition, a turbocharged car engine must provide the starting torque of the vehicle clutch system and its maximum torque at a speed as low as possible. The boost pressure must therefore be established very quickly when the engine changes from idle speed to clutch speed.

The industry is looking for turbomachinery capable of delivering a variable volume of air at a constant air pressure of about 2.5 bars over the entire useful range of engine speeds (number of revolutions / minute) which is currently expanding. about 1 to 4 hours. This pressure level responsibility of a single compression stage with a characteristic pattern as wide as possible.

The flow section of the turbine must vary substantially in the same proportions as the airflow.

The most effective solution to date is the variable turbine distributor which may cover a range of 1 to 3 homogeneous with the maximum width of the compressor map.

The ends of this range compressor efficiency is approximately 60% and that of the turbine by 50%. These yields are improving to the center of the beach to respectively reach 75% and 65%. A quarter of the engine speed range is therefore not covered by the compressor. The low-end torque is generally favored and power decreases from 75% of maximum speed.

Another solution consists in bypassing the turbine by a piloted valve called waste cake. The speed range is only 1 to 2. The yield decreases of relaxation between the minimum speed and maximum speed.

To compensate for the energy dissipation and extend the flow range, we are forced to increase the exhaust pressure at the cost of increased pumping losses.

These solutions with low energy efficiency are sufficient to moderate pressure when the enthalpy available in the exhaust gas is excessive. For double pressure, the overall performance of turbocharger must be improved.

The external recycling of cooled gases is managed by an EGR valve (exhaust gas recirculation) controlled which drifts towards the intake a cooled fraction of the gas flow emitted by the engine in the exclusively cleaned area. When this fraction exceeds a limit, the exhaust temperature becomes insufficient to ensure the compressor turbine energy balance.

To compensate for this lack of temperature at low depolluted regimes, the expansion rate is increased by reducing the section of the turbine at the cost of a degradation in the indicated yield. When the speed increases, the distributor or waste spoils open gradually to reduce the EGR rate and limit the back pressure.

This maneuver is only possible above a certain speed which depends on the size of the turbine.

These operations are carried out with poor energy efficiency due to energy losses in the waste tank or in the variable distributor of the turbine. In addition, the back pressure increases the consumption of the engine over the cleanup operating range widely used in urban driving. The variable geometry of the turbine is highly stressed in urban driving.

To improve the expansion efficiency, it is necessary to stick to turbines with fixed geometry and to limit the operations with rolling of the flows. The compression in two stages makes it possible to generate high pressures by taking advantage of the refrigeration between stages which reduces the work of compression.

To generate the necessary pressure for the torque takeoff of the vehicle, the section of the high pressure turbine HP must be small enough to effectively relax the flow of gas emitted by the engine Clutch system, about 20% of the volume flow at maximum speed. To limit the exhaust back pressure at high speeds must increase the section offered to the gases when the speed increases.

The R2S process from 3K WARNER provides for series connection of the two compressors and the two turbines. To increase the section offered to the gases is gradually transfers the small HP turbine flow of gas to the large low-pressure turbine BP at the cost of energy loss in the bypass regulated by the high pressure turbine. The section offered to increase gas is limited to the LP turbine section. In addition, the opening of the bypass cancels the rate of expansion of the HP turbine which no longer drives the HP compressor which constitutes a throttling which must be bypassed.

The sequential turbocharger provides the parallel turbochargers and turbines. A single turbocharger is active at low speed while the two compressors are active at high speed.

The transition is very used to intermediate regimes with a fall turbocharging efficiency.

This solution has the advantage of offering a maximum gas section equal to the sum of the two turbines. As before the transition takes place with an energy loss by rolling in a widely used area.

In addition, the air pressure is limited as in the case of the single turbocharger.

In both previous cases some transitions involve the acceleration of a rotor which may be too slow in the fast transients of urban driving.

To avoid discontinuities, the patent application No. WO / 0248510 discloses a turbocharger method unregulated two-stage mounted fixed geometry in series wherein the pressure in the cylinder is limited by the pressure drop created by the inlet openings undersized. This solution simple improves performance at low speed at the expense of high-rpm performance when pumping losses are high, the exhaust pressure is proportional to the regime.

The present invention relates to a turbocharging process using the advantages of the series and parallel configurations in an original gas recycling strategy.

The object of the invention is to increase the EGR rate and / or the power and the torque of the engine by increasing the boost pressure by the series connection of two compressors with air cooling between the compressors, remove the power drop at high speed, reduce the torque gain time from idle and on occasions, improve combustion and catalytic cleanup after cold start and at very low power, exploit the mature technology of waste spoils by limiting operating modes with gas flow valve control, reduce the volumetric ratio to respect the maximum authorized pressure in the cylinder while retaining the ability to start and idle quietly in a cold environment, extend the clean-up of the engine's fields of use.

The invention relates to an engine with double turbocharging, recycling of burnt gases and variable valve timing comprising a large number of operating parameters which interact with each other.

The structure includes a limited number of control bodies to create relationships between the parameters managed by the engine control computer. These relationships make it possible to generate a very large number of engine operating modes.

The invention consists of new inter-parameter relationships and the means of implementing them.

Given the complexity of the interactions concerned, the invention will be described by its general concepts and several elementary non-limiting modes of implementation. To facilitate the presentation, the invention will be described on a typical engine whose figures are approximate and have not been the subject of a numerical simulation.

It goes without saying that the numerical values are given to illustrate the description of the invention and that they have no limiting character.

The basic principle is to introduce into the cylinder the mass of fresh air necessary for combustion at a temperature (Tadm) and at a pressure (Padm) such that the volume of this air is always less than the volume of the trap door at closing. valves, to make room for an EGR mass, preferably at least equivalent in most of the engine's operating range.

In the following, the volume Vm drawn by the motor is defined as the product of the volume trapped by the number of cycles per minute that the cycle comprises 2 or 4 times. The invention provides for fulfilling this condition either by adjusting the volume Vc delivered by the compressors, or by adjusting the hatch volume if the valve timing is variable, or by the two adjustments made, preferably, successively so as not to multiply the regulations. concurrent. To optimize the turbocharging efficiency and avoid geometry variations in fast transients, the invention provides several configurations with fixed geometry offering several power levels with the same maximum torque. We can thus consider a city configuration, a road configuration and a highway configuration. For engines with fixed valve timing, the invention describes turbocharging units making it possible to modulate the volume Vc of air cooled between 1 and 3 at constant compressor outlet pressure P2 chosen by way of example for the purposes of the description to 4.5 bars.

For engines with variable valve timing, the invention provides for covering the mass flow range between 1 and 2 by modulating the pressure of air P2 between 4.5 bars and 9 bars with fixed geometry therefore at substantially constant volume Vc, then by increasing the turbine flow section Sd by only 50% (instead of 300% previously) to cover the mass flow range from 2 to 3 with constant P2 and variable Vc. In most modes of the invention provides to maintain approximately equal pressure P2 to P3 compressor outlet and turbine entry in order to perform the exhaust gas recirculation with minimum losses by pumping.

This relationship also has the advantage of positioning on stationary turbomachines lines of their characteristic diagrams for a given geometry.

In the case of the 2-stroke cycle where the sweep imposes P2> P3, the volume of the exhaust manifold is chosen small enough so that

P3 undergoes pulsations around an average value equal to P2. The scanning is then carried out in the fraction of time when P2> P3 and recycling takes place when P3> P2.

The volume Vc of cooled air delivered by the compressors depends in these conditions only on the flow section offered to the gases, that is to say on the geometry of the turbines and their by-passes. For a given geometry of this volume is substantially constant for all loads and all engine speeds.

The invention provides for dimensioning the turbomachines so that Vc is always less, and preferably less than half the volume drawn in by the engine in its depolluted domain. For the specifications of modern engines, this volume corresponds substantially to the volume drawn by the engine at idle in the most closed configuration of the turbines. We will set the idle speed at 700 rpm for the purposes of the presentation.

For engines with fixed valve timing the maximum air volume will therefore be the volume Vm drawn by the engine at 2100 rpm and for engines with variable pitch, where P2 is no longer limited by the motor, the volume Vm at 1050 t / mn.

In propulsion mode, the volume drawn in by the engine varies very quickly with the speed and timing of the valves. The present invention provides for instantly filling the volume not occupied by fresh air Vm - Vc, with a mass of burnt gases, preferably at least equal to the mass of fresh air, recycled at a temperature compatible with the energy balance. between turbines and compressors.

This temperature is defined by the upper isochore of the entropy diagram.

In summary, unlike the current art which imposes on the compressor the volume of air drawn in by the engine, the invention compensates by EGR for the difference between the volume drawn in by the engine and the constant volume delivered by the turbomachines in a fixed geometric configuration. Current art therefore provides Vc = Vm outside the depolluted domain, while the invention provides Vm = Vc + Vegr (volume of EGR gases) throughout the field of use.

This strategy makes it possible to cover the entire field of use of the engine situated under the curve at maximum P2 authorized for the engine considered, without having to modify the geometry of the turbocompression group. This curve is described at constant motor power proportional to Vc and P2.

When the temperature of the combustion mixture increases, the cycle moves to the right of the diagram entropy / temperature. In atmospheric suction, the EGR is not cooled to place the engine cycle to the right of the entropy / temperature diagram, the partial charges being carried out at constant end of combustion temperature and at variable start of combustion temperature. The exhaust temperature and keeps the maximum level supported the establishment of the turbocharger and the catalytic pollution.

Depending on the case, mixing between the fresh air and the gases takes place in the intake duct, in the cylinder or in both. The invention preferably provides devices for homogenizing the oxidizing charge.

The invention is based on two-stage turbocharging with air cooling upstream and downstream of the high pressure compressor.

This compression method imposes certain relationships between the flow rate and the pressure of the air delivered which depend on the characteristic diagrams of the compressors.

Modern automotive engines must provide their maximum torque at 25% of maximum speed.

The ideal would therefore be a volume flow Vc varying from 1 to 4 delivered at double pressure and at a volume half the volume sucked by the engine.

The turbines with variable distributor allow to cover a range from 1 to 3 with an efficiency of 50% at the ends of the range.

The compressors associated with these turbines pay in yield for this flow flexibility. The yield of end of range does not exceed 60%.

In addition to the fact that the desired range is not reached, the overall turbocharging efficiency is insufficient to double the pressure of 2.25 bars currently required for the maximum torque without gas recycling.

Between 25% and 75% of the maximum speed, the turbocharging efficiency is sufficient to maintain the nominal intake pressure, but between 75% and 100% it collapses and with it, the power.

The generalization of a mass EGR rate of 50% would reduce by 50% torque at low revs and would see the collapse of the power from 37.5% of maximum speed. The use of fixed geometry turbomachinery with a narrower operating range provides a gain of 10 to 15% on the efficiency of compressors and turbines.

To cover the flow range from 1 to 3 at constant pressure, the low pressure compressor according to the invention delivers its minimum flow at low pressure and its maximum flow at high pressure.

The role of the high pressure compressor is to complete the compression up to the required level set at 4.5 bars to fix ideas. The pressure ratio of the HP compressor must therefore vary according to the flow rate by means of a variable geometry in the supply of the turbines.

To cover the range from 1 to 1, 5 it is enough to partially bypass the HP turbine or a small opening of its distributor less penalizing in yield.

The efficiency of the HP compressor is higher in this second case, as shown in diagrams 9 and 10.

To avoid yield destructive geometry regulations the present invention provides a fixed geometry operation modes including full charge takes place in part at constant power and variable volume fraction of EGR. On constant power curves, the compressors park at a single operating point when the speed varies.

The partial loads are carried out at variable pressure and flow rate according to a unique curve in the diagram of each compressor chosen in the zones of good yields. The mass rate of EGR is then regulated by its temperature.

When the power of a mode becomes insufficient, the invention plans to activate the variable geometry to increase either to switch to another mode with fixed geometry or to tighten the EGR bypass to increase the pressure P3. For example, the goal that achieves the invention may be summarized as follows for a diesel engine today cleared:

- the idle speed is 700 rpm,

- the clutch is made at 1000 rpm,

5 - the maximum torque is available between 1200 rpm and 3400 rpm,

- the maximum power is available between 3400 rpm and 5000 rpm,

- the intake and exhaust pressures are limited to 4.5 bars, 0 - the mass rate of EGR in the depolluted area is 40% at 1200 rpm and more than 50% between 1400 and 5000 t / min.

The description now lists the principles of the invention and examples of implementation with reference to the figures below:

Fig 1: functional diagrams A, B, C of the power supply structure of the motor,

Fig 2 diagrams regime / volume of transfers by variable valve timing and limited variation of the geometry of the turbines in an extreme case where the maximum torque is available at 20% of the maximum regime,

Fig 3: diagram of examples of operating modes of the invention with o (3A) and without (3B) variable valve timing in the extreme case of Figure 2, Fig 4: Position of hot EGR cycles in the diagram 11 S, Fig 5: example of a feed structure for a 4-stroke engine with fixed valve timing,

Fig 6: example of a 2-stroke engine structure with 5 variable valve timing and single exhaust manifold,

Fig 7: example of a feed structure for a 2-stroke engine with variable valve timing and two exhaust manifolds, Fig 8: operating diagram of the double waste trash, Fig 9: compressor adaptation diagram for the fixed setting of the valves, Fig 10: Adaptation diagram compressors for the variable valve timing,

Fig 11: two examples of directive intake chapel (A and B) for mixing the introduced gases and the residual gases as well as for organizing the scanning of the two-stroke cycle.

1. The subject of the invention is an alternating engine used between a minimum rotation speed Nmin and a maximum speed Nmax, which comprises a turbocharging unit sized to be in autonomous operation when:> It supplies air to the intake manifold of the engine via coolant

> It is supplied with gas by the engine exhaust manifold

> The turbine supply pressure P3 is substantially equal to the compressor discharge pressure P2.

It is known that under these conditions, at constant air temperature and with fixed geometry, turbocharging delivers a volume of cooled air Vc substantially constant when the pressure varies.

It is also known that the volume Vc is substantially proportional to the flow section of the turbine Sd offered to the hot gases.

This engine is characterized in that the turbine pressure P3 is maintained substantially equal to the compressor pressure by a bypass between the intake manifold and the exhaust manifold sized to transfer the flow of exhaust gas to the manifold intake without significant pressure drop, and the air volume Vc is less than the volume drawn in by the engine at Nmax speed so that a flow of hot gases is drawn back in by the engine via the bypass above of the Na regime where the aspirated volume is equal to Vc, and an air flow is diverted towards the turbine below the Na regime.

In the following, the bypass between the collectors will be called the EGR duct and the Na regime will adapt the turbocompression regime. 2. The motor according to 1. above can provide that the bypass or conduit

EGR includes an EGR valve to increase the turbine pressure above the compressor pressure.

3. The engine according to 1. above can provide that the turbocharging unit includes an intake valve located on the discharge of the compressed air compressor for increasing the pressure above the turbine pressure.

4. Advantageously, the engine according to 1. can provide that the EGR duct includes a gas coolant with adjustable temperature, preferably up to a temperature close to that of fresh air.

5. The engine according to 2. above can provide that the temperature is adjusted by controlling a refrigerant bypass.

This general principle refers to Figure 1.

6. A method of supplying an engine according to 4. above may provide that the EGR temperature is controlled to create the desired excess air for combustion in the engine.

7. This supply method according to 6. can in particular be characterized by the fact that the EGR temperature is controlled so that the mass of recycled gases remains substantially equal to the mass of fresh air until the regime where this temperature reaches the exhaust temperature. Beyond this regime, the recycled mass becomes greater than the mass of fresh air.

8. This supply method according to 6 can also be characterized by the fact that the air cooler is completely bypassed when the engine does not deliver propellant power. 9. This supply method according to 8 can in particular be characterized by the fact that for cold starting and idling operation the adjustment of the turbine valves (6 and 7) and / or the setting of the valves is adjusted so that the 'excess combustion air is minimal for the desired level of pollution control. These processes relate to the different modes of piloting the bypass.

EGR depending on whether you want to optimize smoke, NOX, noise or engine recovery capacity.

10. The engine according to 1. or 4. above can be further characterized by the fact that the adaptation regime Na is substantially equal to Nmin / 2 so that the volume of recycled gas is at least equal to that of l fresh air, and the minimum temperature of the recycled gases is preferably close to the temperature of fresh air so that the mass of the recycled gas is at least equal to that of fresh air to the minimum speed N min used for clean up the whole area of engine use. 11. The motor according to 1. may also be characterized in that the turbocharger unit comprises a low-pressure turbocharger BP and a high pressure compressor which HP compressors working in series with preferably one cooling air between compressors and the exhaust flow section Sd can be adjusted between a minimum Sd min and a maximum Sd max by one or a combination of the following means:

- adjustment of the variable section of the gas distributor of the turbines,

- opening of a bypass between the inlet and outlet of the turbines, - transition from a series configuration to a parallel configuration of the turbines.

The adaptation regime of the turbocharging Na thus becomes adjustable, continuously or discontinuously, between two values Na min and

Na max. In the following, a bypass between the inlet and the outlet of a turbine will be called waste cake.

This structure refers to Figure 5.

12. The engine according to 11. above can also be characterized by the fact that the minimum flow section Sd min offered to the gases is constituted by the two turbines mounted in series at maximum closing if their distributor is variable and all waste spoils are closed. 'there are.

These methods relate to all the modes chosen to describe the invention.

13. The engine according to 12. can be arranged so that it operates on a 4-stroke cycle and that the valve timing is fixed.

14. The engine according to 13. can also be characterized by the fact that the maximum flow section Sd max offered to the gases consists of two turbines with fixed distributors mounted in parallel.

To switch the turbines from the serial configuration to the parallel configuration of the means, the following maneuvers can be carried out successively:

- gradual partial opening of the HP trash waste - progressive and simultaneous partial opening of HP and BP waste treats

- simultaneously and quickly: total opening of the HP trash waste, total closing of the BP trash waste, communication of the output of

5 the HP turbine with the BP turbine outlet

15. The engine according to 13. can also be characterized by the fact that the maximum flow section Sd max offered to the gases consists of a BP turbine with fixed distributor and an HP turbine with variable distributor mounted in parallel, the HP distributor being in full opening. 0 To switch the turbines from the serial configuration to the parallel configuration of the means, the following maneuvers can be carried out successively:

- gradual opening of the HP turbine distributor,

- gradual partial opening of the BP trash waste 5 - simultaneously and quickly: total closure of the BP trash waste and communication of the HP turbine outlet with the BP turbine outlet

These methods relate to the mode designated below A4.

16. A method of supplying an engine according to 2., 3. or 11. 0 above can also be characterized by the fact that to limit the frequency of configuration changes, the geometry is immobilized for a type of driving which implements a limited range of power, for example the series configuration for city driving and the parallel configuration for road driving, the power thresholds corresponding to 5 each configuration can be crossed for short duration maneuvers, such as accelerations, overtaking, speed spikes, etc.

The thresholds can be crossed as follows:

- closure of the EGR valve if the pressure in the exhaust manifold o can be increased. - by opening a waste gate or two if the exhaust temperature can be increased.

- by closing the inlet valve if the maximum cycle pressure is reached or if the compressors are close to their maximum flow. This process concerns the modes designated A1.1, A2, A3, B2, B4, C3,

D3.

17. An engine according to 14. can also be characterized by the fact that the waste container BP has a second seat for simultaneously closing the bypass input / output turbine PB and placing in communication output turbine HP output turbine BP.

This process concerns the A4 mode below.

18. An engine according to 14. can also be characterized by the fact that the two waste spoils are concentric and include stops so that their simultaneous movements are actuated by one of them and communicated to the other by said stops. . This structure refers to Figure 8.

19. The motor according to 13. above can be characterized by the fact that the maximum flow section Sd max is constituted by two turbines with variable distributors in full opening mounted in series, and the distributors are opened simultaneously to maintain the pressure of admission to its maximum set value on the full load curve.

This very expensive solution was not chosen as an example. It can nevertheless replace all the modes presented.

20. An engine according to 12. above can also be characterized in that the valve timing can be controlled for moving the closing of the cylinder between the vicinity of PMB and mid stroke of the piston, the maximum débitante Sd section consists by the HP turbine in distributor series configuration in full opening if it is variable, waste spoils HP in full opening in the opposite case and the turbines are sized for allow the compressors to simultaneously reach their maximum pressure ratios.

This concerns modes B, C, and D below.

21. A method of supplying an engine according to 20. above can be characterized by the fact that the curve of full load as a function of the speed is executed as follows: from Nmin to 2 Nmin the closing of the intake Fa passes of the PMB (bottom dead center) at around 90 dv after the PMB so as to maintain the cycle pressure below its set value. The HP waste distributor or waste is closed, from 2 Nmin to approximately 3 Nmin the HP distributor or waste HP is open and possibly the waste waste BP, to maintain the inlet pressure at its maximum set value, FA is maintained at 90 dv (crankshaft degrees) after PMB, from 3 Nmin to Nmax the overall fuel flow is kept constant to keep the intake pressure at its limit value, at partial load the FA setting will be controlled according to a memorized map by the engine control computer.

This process described by Figure 2 relates to modes B1, C2, D2 below.

22. An engine according to 12. above may be characterized in that it works on the cycle time 2, the inlet openings are closed by valves, the exhaust ports are closed by valves and communicate with a single exhaust manifold, the external recycling phase precedes the scan, the valve timing can be controlled for moving the closing of the cylinder between the vicinity of the PMB and the mid-stroke of the piston, the maximum débitante Sd section consists of the HP turbine in serial configuration fully open dispenser if it is variable, wastegate HP fully open in the contrary case, the turbines are sized to allow the compressor to simultaneously achieve their maximum pressure conditions, the EGR valve is replaced by a check valve or a closable aerodynamic diode. 23. A method of supplying an engine according to 22. above can be characterized by the fact that the full load curve as a function of the speed is executed as follows:

- from Nmin to 2 Nmin the cylinder closure changes from PMB to around 90 dv after PMB so as to maintain the cycle pressure at its set value.

- the distributor or the HP waste trash are closed.

- from 2 Nmin to approximately 3 Nmin the HP distributor or the HP waste tank are open and possibly the BP waste waste, to maintain the inlet pressure at its maximum set value. FA is maintained at 90 dv after PMB.

- from 3 Nmin to Nmax the overall fuel flow rate is kept constant to keep the intake pressure at its limit value.

To maximize the cooled external EGR, the depolluted partial loads can be carried out as follows:

- the cylinder remains closed in the vicinity of the PMB and the turbines remain in the closed configuration until the limit P2 for this setting.

- the turbines are then opened to maintain P2 at its limit value. - the aerodynamic diode when the external recycling flow is canceled.

This process refers to Figure 6 and is about the modes C below.

24. An engine according to 12. above can be characterized by the fact that:

- it works on the cycle time to 2, - it comprises two exhaust ports per cylinder, closed by, valves, which respectively communicate with an exhaust manifold connected to the turbine and an exhaust manifold connected to the EGR conduit and / or to the turbine via a piloted distributing valve, - calibration of the affected valve EGR can be controlled to move the closure of the cylinder between the vicinity of the PMB and the mid-stroke of the piston,

- the external recycling phase precedes scanning when the cylinder closes in the vicinity of the PMB and follows it when the cylinder closes at mid-stroke of the piston,

- the maximum flow section Sd is constituted by the HP turbine in the distributor series configuration in full opening if it is variable, waste spoils HP in full opening in the opposite case, - the turbines are sized to allow the compressors to simultaneously reach their maximum pressure ratios,

- the EGR valve is replaced by a non-return valve or a closable aerodynamic diode.

25. The process for supplying an engine according to 24. above can be characterized by the fact that the pressure P2 is less than the authorized limit for this setting, the distributor valve is in the recycling position, the cylinder is closed at in the vicinity of the PMB, the distributor or the HP waste spoiler are closed, the pressure reaches the authorized limit value for this setting, the cylinder closure is moved to the mid-stroke of the piston to substantially double the authorized limit P2, the distributor valve remains in the recycling position, the distributor or the HP waste container remains closed, the P2 pressure reaches the new limit authorized for this new setting, the distributor valve blocks recycling, the distributor or the HP waste container opens to maintain the P2 at its new authorized limit, the transition can be done gradually in both directions or quickly with a hysteresis.

This process refers to FIG. 7 and relates to modes D below.

26. The method according to one of the preceding modes can also be characterized by the fact that at full load the variable geometry is controlled to maintain a parameter at its set limit value, at partial load the variable geometry is controlled to optimize pollution control and / or performance according to a map stored in the engine control computer.

27. An engine according to 1. above comprising a flat cylinder head carrying valves whose chamber side faces are coplanar with the cylinder head and substantially tangent to the cylinder can be characterized by the fact that the intake pipe or pipes terminate in a oblong nozzle defined by an upper half-cylinder resting on and tangent to the upper edge of the conical seat along its generatrix situated in a plane substantially perpendicular to the plane passing through the axis of the seat and through the axis of the cylinder and a lower cylinder covering half of the head opposite to said valve generator.

The nozzles will also be directed to create a tangential speed in the same direction. The angle of the seats are chosen to optimize the stratification of the combustive load.

28. An engine according to 1. comprising a flat cylinder head carrying valves whose chamber side faces are coplanar with the cylinder head and substantially tangent to the cylinder can also be characterized by the fact that the conical sealing surface of the intake valves extends towards the piston by a cylindrical part of height slightly greater than the lifting of said valves, that the conical seats of said valves are arranged at the bottom of cylindrical housings arranged in the cylinder head to receive said cylindrical parts of said valves so that the underside plane of the valves are in the plane of the cylinder head when they rest on their seats, the clearance between the housings and the valves being minimal, that recesses are made in the cylinder head which do not exceed the following borders: - two cylindrical portions concentric with the bore and tangent externally and internally to the cylindrical recess of each valve,

- a conical surface extending the half-seat of the valve delimited by a plane passing through its axis and the axis of the cylinder,

- the recesses will also be oriented to create a tangential speed in the same direction,

- the seat angle is chosen to optimize the stratification of the oxidizing charge. 29. An engine according to 27. or 28. above can be characterized by the fact that it comprises two diametrically opposite intake valves.

These structures described in FIG. 11 relate to the modes where residual burnt gases are retained in the cylinder.

30. An engine according to 1. may be characterized in that a fraction of the recycled gases is retained in the cylinder at closing of the latter, the fresh gases are introduced through the directional inlet ducts in order to organize stratification of temperatures and concentrations in the chamber combustion top dead center, fuel is vaporized in the fresh gas. An advantageous solution for a radial stratification is to confine the combustion bowl in a central small diameter which fills hot gas concentrated in the center of cylinder during compression when the fuel mixture of fresh air is centrifuged in the peripheral space between the piston and the cylinder head until its transfer to the bowl begins. If the richness of the fuel mixture is between 60 and 70%, the flame initiated by hot gases in the bowl will not spread to the peripheral fresh gas but will grow by mixing turbulent with the hot gases already in the bowl . The excess air present in all parts of the chamber guarantees complete combustion without NOX or particles. The layer of carburetted air remains in the clearance between the piston and the head that not taking part in the combustion in the bowl will be burned at the beginning of relaxation or suivant.Les cycle gas recirculated used in this case to initiate and support combustion of a lean mixture during his turbulent transfer in the combustion bowl . The initial ignition can be ensured by an electric spark.

31. Such an engine according to 30. can be further characterized in that the fuel is introduced into the clean air between the compressor and the external EGR mixer.

32. 30. An engine as claimed may also be characterized in that the fuel is introduced into the mixing between the fresh air and external EGR.

33. An engine according to 30. can also be characterized in that the ignition point is controlled by setting the valves when the cylinder is closed.

34. 30. An engine as claimed may also be characterized in that the ignition point is controlled by the temperature of the external EGR.

35. An engine according to 30. can also be characterized by the fact that the first ignition is electrically controlled or is triggered spontaneously by the injection of high pressure fuel at top dead center. 36. A motor according to 30. can also be characterized by the fact that the gas working chamber has a geometry of revolution around the axis of the cylinder, the stratification has a geometry of revolution around the axis of the cylinder created by the orientation of the intake orifices, the temperature of the oxidizing charge increases between the periphery and the axis so that the self-ignition propagates from the center to the periphery.

37. An engine according 36. may be further characterized in that the meridian profile of the combustion chamber is selected to optimize the energy release rate escalation isothermal surfaces to the reactive load. These processes primarily concern the variably timed engine valves and particularly to two cycles time.

Reference is now made to diagram A of FIG. 1 which describes the functional diagram of the power supply structure of the engine. The engine 1, used between a minimum speed Nmin and a maximum speed Nmax, is supplied with gas at the temperature Tadm and at the pressure Padm by an intake manifold 8.

After combustion, it rejects them at temperature T3 and at pressure P3 in an exhaust manifold 9. The collectors 8 and 9 are interconnected by a recycling conduit

3, dimensioned to be able to bypass all the burnt gases discharged by the engine without significant pressure losses.

The recycling conduit 3 comprises a gas cooler 4, of the gas / water type for example, in order to be able to cool the burnt gases recycled EGR to a temperature Tegr adjustable between T3 and a minimum temperature preferably close to the temperature of the cooling water.

The temperature can advantageously be adjusted by controlling a refrigerant bypass.

In this case, it is advantageously possible to use a bypass short-circuiting all or part of the refrigerant exchanger 4, as shown in diagram B of FIG. 1.

The EGR pipe 3 includes a valve 6 located upstream or downstream of the refrigerant 4. and called the EGR valve.

The EGR duct 3 is preferably connected to the intake manifold 8 via a mixer 5 to homogenize the oxidizing charge sucked in by the engine.

An intake valve (7) may be provided on the compressor discharge to increase the compression pressure over the turbine pressure. The assembly constituted by the motor 1 and the conduit 3 is supplied with fresh air at the pressure P2 cooled to the temperature T2 by a turbocharger unit 2, preferably via the mixer 5.

The turbocharging unit is activated by the gases emitted by the assembly 1 and 3 at the pressure P3 and at the temperature T3.

The turbocharging group 2 can comprise one, or preferably, two turbochargers with fixed or variable geometry, one, or preferably, two air coolers, one or more discharge valves called waste tank and an inlet valve 7 for control the pressure P2 located upstream or downstream of the last air cooler.

Group 2 is supplied with atmospheric air by an air filter and expels its gases in an exhaust line which can include post-treatments and silent pressure drop generators.

To illustrate the depollution potential of the invention, we now refer to FIG. 2.

This diagram describes the full load curve of a fully cleaned extreme engine which presents its maximum torque at 20% of its maximum speed.

The turbocharging in series configuration is therefore adapted to half this value, i.e. 0.1 Nmax

The compressors deliver a maximum P2 pressure of 9 bars.

The flow section Sd of the turbines varies only from 1 to 1.5.

The engine has a variable valve timing that is used to control the closing of the cylinder between the BDC where the pressure P2 is limited to 4.5 bar and the mid-stroke of the piston 90 after the PMB dv where P2 is limited at 9 bars.

The volume unit in this diagram is equal to the displacement and the speed unit the maximum speed.

All curves show the amount of gas present in the cylinder at its closed except the curve C represents changes in the total volume Vc of cooled air supplied by group 2 without taking the units into account. It should be understood that this volume flow is constant up to 0.4 Nmax, then increases linearly by 50% between 0.4 and 0.6 Nmax to stay at this value up to Nmax. 5 On this graph the unit volume is the cylinder, Vc is controlled to 0.1 to 0.15 Nmax for a two-stroke engine and of 0.05 to 0.075 Nmax for a 4-cycle engine, when the speed goes from 0.4 to 0.6 Nmax.

We first describe the preferred operating modes where valve 6 and valve 7 are in full opening and where Sd (flow section offered to 0 exhaust gases) is fixed at its minimum value.

The turbine inlet pressure P3 is in these conditions substantially equal to the compressor outlet pressure P2.

When P3 exceeds 2 bars, the sonic flow rate passing through the turbines is proportional to P3 and inversely proportional to the root of T3 which varies little as a function of P3.

The assembly constituted by the motor 1 and the duct 3 is supplied with fresh air at the pressure P2 cooled to the temperature T2 by a turbocharging unit 2, preferably via the mixer 5.

The turbocharging unit is activated by the gases emitted by o the assembly 1 and 3 at the pressure P3 and at the temperature T3.

The turbocharging group 2 can comprise one or preferably two turbochargers with fixed or variable geometry, one or preferably two air coolers, one or more discharge valves called waste tank and an inlet valve 7 for controlling the pressure P2 located upstream 5 or downstream of the last air cooler.

To illustrate the pollution potential of the invention Referring now to Fig o 2. This diagram describes the full load curve of a fully cleaned extreme motor which presents its maximum torque at 20% of its maximum speed (Nmax).

The turbocharging in series configuration is therefore adapted to 0.1 Nmax.

The compressors deliver a maximum P2 pressure of 9 bars.

The flow section Sd of the turbines varies only from 1 to 1.5.

All the curves represent the volumes of gas present in the cylinder when it is closed, except curve C which represents the total volume Vc of cooled air delivered by group 2.

On this diagram, Vc is therefore controlled from 0.1 to 0.15 Nmax for a two-stroke engine and from 0.05 to 0.075 Nmax for a 4-stroke engine, when the speed changes from 0.4 to 0.6 N.sub.max. We first describe the preferred operating modes where valve 6 and valve 7 are in full opening and where Sd (flow section offered to exhaust gases) is fixed at its minimum value.

The turbine inlet pressure P3 is in these conditions substantially equal to the compressor outlet pressure P2. When P3 exceeds 2 bars, the sonic flow rate passing through the turbines is proportional to P3 and inversely proportional to the root of T3 which varies little as a function of P3.

If T2 is maintained more consistently by the air cooler, the volume of cooled air Vc delivered by the group 2 is substantially constant when P2 varies. This volume is also proportional to the flow section Sd offered to the gases which expand in the turbocharging group 2.

When the speed increases, the fraction of this volume retained in the cylinder (curve D) decreases inversely proportionally. The calculations also show that the mass flow of fresh air is substantially proportional to the flow of fuel burned in the engine, itself proportional to the speed when the load is constant.

The volume of fresh air retained in the cylinder is therefore inversely proportional to the speed whereas the pressure P2 of this air is proportional to it. This results in the remarkable fact that the mass of fresh air retained per cycle is independent of the speed and depends only on the fuel flow. The combustion is thus carried out at substantially constant wealth.

This is a consequence of the stability of T3 imposed on the engine by this type of turbocharging. The maximum torque to be reached at 0.2 Nmax, the volume hatches

(curve A) is equal to 1 and the maximum fuel notch is reached at this minimum operating speed where P2 (curve B) has reached the limit value of 4.5 bars for this hatch volume.

The volume of fresh air retained in the cylinder (curve D) is then half of the hatch volume, the difference being occupied by the EGR cooled to temperature T2 to achieve a mass recycling rate of 50%.

When N goes from 0.2 to 0.4 Nmax at full fuel notch, P2 goes from 4.5 bars to 9 bars. To respect the maximum pressure of the cycle, the setting must be simultaneously modified to bring the hatch volume to 0.5 for N = 0.4 Nmax. The volume of fresh air retained having undergone the same reduction by half the mass rate of EGR remained stable at 50%.

When N goes from 0.4 to 0.6 Nmax, P2 remains at its 9 bar limit, the hatch volume remains at its minimum 0.5, the variable geometry of the turbocharging group then takes over from the variable valve timing to increase 50% of the volume Vc fresh air and therefore its mass flow. The mass of fresh air burned per cycle (curve E) therefore remained constant between 0.2 and 0.6 Nmax, the range where the maximum torque was available.

Between 0.6 Nmax and Nmax the torque decreases at constant power and the EGR temperature must be increased to limit the mass rate of EGR to 50%.

The entire area of use of the engine is thus cleaned up by a mass EGR rate of 50%.

We now refer to Figure 3 which compares in the field PME / regime (PME is the effective average pressure of the engine): the previous operating modes where we share the geometric adaptation between the engine and the turbomachines by successively exploiting the variable valve timing to modulate the volume Vm drawn by the motor and the turbine section to modulate the volume Vc delivered by the compressor and the fixed valve valve operating modes where all of the geometric adaptation is carried out by the section Sd turbine.

We see that each geometrical configuration makes it possible to reach all the PME and all the regimes located under the constant power hyperbola corresponding to the maximum air flow in its configuration.

These hyperbolas are also curves with maximum constant P2 in the configuration.

It is therefore advisable to choose P2 as a parameter to control the geometry.

For example in the diagram 3 A, if the maximum power is Wmax, we can choose 3 clean-up operating modes. W <0.33 Wmax which corresponds to city driving:

The valve timing and turbine geometry remain fixed for rapid transients in city driving.

Only the EGR temperature is regulated at high speed.

P2 varies with the power up to 4.5 bar. 0.33 Wmax <W <0.67 Wmax which corresponds to driving on the road: the geometry of the turbines remains fixed, the valve timing is programmed as a function of P2, the EGR temperature is regulated at high speed, P2 varies from 4 , 5-9 bar. 0.67 Wmax <W <Wmax which corresponds to highway driving: the valve timing stops, a waste spoil or distributor opens to limit P2 to 9 bars, the EGR temperature is controlled at high speed, P2 remains constant at 9 bar.

Diagram 3B shows the 2 possible modes without any variable geometry thanks to the series and parallel configurations of two turbines:

W <0.33 Wmax which corresponds to clean city driving: the turbines are in series configuration, the turbine geometry remains fixed for rapid transients in urban driving, only the EGR temperature is regulated at high speed, P2 varies with the power up to 4.5 bars.

0.33 Wmax <W <Wmax which corresponds to driving on a partially depolluted highway: the turbines are in parallel configuration, the geometry of the turbines remains fixed, the EGR temperature is regulated at high speed, P2 varies up to 4.5 bars . The transition from one configuration to another is explained in more detail below:

We now refer to FIG. 5 which describes a structure well suited to current engines with fixed valve timing which comprises:

> A small high-pressure turbocharger (HP) fixed geometry or variable 102, supplying air to the intake manifold through an optional inlet valve 105, and preferably an air refrigerant HP 108 to reduce the volume of air entering the engine. The HP turbine is always powered by the gases from the exhaust manifold of the engine discharges into a conduit 111 that can communicate with the inlet and the LP turbine outlet as well as the exhaust manifold.

> A large low-pressure (BP) turbocharger with fixed geometry 101, supplying air to the HP compressor, preferably through a BP 107 air cooler to reduce the work of the HP compressor. The BP turbine is supplied by conduit 111.

> A double waste spade 103 comprising a waste waste HP and a waste waste BP, for example coaxial to be driven together, controlled to pass from the series configuration to the parallel configuration via if necessary a series / parallel configuration which can be pneumatically actuated by the pressure of the compressors for rapid maneuvers and by a hydraulic or electric actuator for fine regulation.

> A flue gas recirculation conduit connecting the exhaust manifold to the intake manifold through a gas cooler 109. This conduit is provided with an EGR valve pilot 105 at its junction with the exhaust manifold and a refrigerant bypass regulated by a controlled distributor flap 106. The oxidant mixture is homogenized by the mixer 110 located upstream or in the intake manifold.

> The hot gases mix with the cold gases through the perforated partition of the mixing tube 110. The mixture obtained then mixes with fresh air in the perforated wall of the mixing tube.

The compressors are dimensioned to be able to deliver at pressure 4.5 bars the volume of cooled air drawn in by the engine between 700 and 2100 rpm.

The HP turbine is dimensioned to receive, in serial mode, gas 4.5 bar emitted by the engine when operating at 700 rev / min and is supplied at 4.5 bar. The BP turbine has a volumetric capacity such that the two turbines mounted in series drive the compressors at their minimum volume flow rate (700 rpm 4.5 bars) and that mounted in parallel they drive the compressors at their volume flow rate corresponding to (2100 t / min, 4.5 bar), this for a feed pressure of 4.5 bar and a temperature compatible with good combustion. These conditions are met when the capacity of the turbine is between 1, 5 and 2 times that of the HP turbine.

This first structure is well suited to current engines. It includes control valves whose control allows three groups of operating modes of two turbochargers depending on whether the engine is equipped with a variable timing distribution or not and whether it operates on the 2 or 4 stroke cycle.

When the valve timing is variable, the parallel configuration is no longer necessary. In the majority of cases, only one regulatory body is active within a mode and no mode provides for more than two simultaneous regulations.

This structure works as follows:

When the valve timing is fixed, the turbines operate in series mode in the low-power depolluted area and in parallel mode in the high-power area not subject to legislation. The Sd section available for gas evacuation passes from approximately 1 to 3 between these two modes. The non-evacuated fraction of the gases emitted by the engine is re-aspirated via the recycling duct. In series mode the two turbines are traversed by the entire flow of the compressors and the HP turbine delivers a power greater than that of the BP turbine. The pressure ratios of the compressors are fixed by this power ratio.

Mode parallel each turbine receives a fraction of the flow compressors proportional to its flow section. The HP turbine delivers while a power lower than that of the LP turbine which leads to a pressure ratio of the HP compressor lower than the BP. An HP turbine variable valve allows a finer optimization of pressure ratios. It is also possible to winnow admission into the LP turbine to increase the gas flow through the HP turbine.

The transition can be made instantaneously in the rapid operation of a double waste trash 103 between two sealing seats located in the exhaust duct 111 of the HP turbine, without loss of energy by winnowing of the fluids. The transition is accompanied by a sudden change in gas pressures and the speed of the turbochargers.

These discontinuities can be eliminated by two mixed series / parallel transition modes controlled by the double tank waste shown in figure N ° 8.

These mixed modes also make it possible to considerably extend the clean-up area.

The invention is also based on an original strategy of external recirculation of the burnt gases:

The compressors always work in series with air cooling upstream and downstream of the high pressure compressor. For example, according to diagram C of FIG. 1, the refrigerant 4 is located in front of the cylinder intake, on the intake duct downstream of the mixer 5. In the example below, the pressure delivered by the group is limited to 4.5 bars.

The EGR recycling duct is dimensioned to create a low pressure drop, when the EGR valve is open, in order to be able to impose equal pressure between the intake manifold and the exhaust manifold.

This solution minimizes decanting losses in the recycled phase. This solution concentrates the operating points of the compressors on two OB and PC lines of the characteristic diagrams corresponding respectively to the serial and parallel modes (Figures 9 and 10).

In mixed mode, the flow in the diagrams is controlled by the double tank waste and the variable distributor of the HP turbine when it exists, as well as by the EGR valve.

The recycling duct includes a gas / water refrigerant and its pilot by-pass which allows the EGR temperature to be adjusted between the temperature of the fresh air and the exhaust temperature. Only this mapped regulation is active in parallel and serial modes, the EGR valve remaining in full opening.

The adjustment of the fraction passed by of a gas / water refrigerant whose walls remain in the vicinity of the water temperature allows precise programming of the intake temperature depending on the speed, the injection rate and the case any other parameters.

Programming has the advantage of avoiding the difficult measurement of a rapidly varying temperature, a measure necessary for complete regulation.

This structure also makes it possible to control the ratio between the compressor outlet pressure P2 and the supply pressure of the turbines P3:

P3 = P2 = Pad valve 104 and 105 open

P3> P2 = Pad valve 104 partially closed and valve

105 open

P3 = Padmission <P2 Valve 104 open and valve 105 partially closed

P3> or <P2; P3 and P2> Padmission Valves 104 and 105 partially closed

It will be noted that all previous configurations allow a gas recycle because P3 is always greater than or equal to the admission. The invention favors the modes where the intake = P3 to allow recycling of the burnt gases while minimizing pumping losses.

The EGR valve 104 is only used in a 4-stroke cycle to extend the above modes for short periods and to avoid frequent activation of the double spoil waste 103.

The setting P3> P2 makes it possible to compensate for a lack of efficiency of the turbomachines at the cost of an increase in pumping losses.

We can cite for example the transient regime between idling and clutch, city acceleration and speed spikes. Similarly, the intake valve is only used to improve the positioning of the compressors in their characteristic diagrams near the maximum power or reduce the maximum pressure of the cycle at the cost of a reduction in the EGR rate.

Most of the pipe is therefore EGR valve and inlet valve blocked in full opening.

For engines with fixed valve timing whose maximum cleared torque is 25% of maximum speed, the range of variation of the flow sections is approximately 1 to 3. The full cleared load is therefore carried out at constant torque up to 75% of maximum speed and constant power between 75% and 100%, which is perfectly suited for road propulsion.

The invention is also based on a new positioning of the partial load cycles in the temperature entropy diagram 11 S as shown in Figure No. 4: At very low loads aspirated, the current art maximizes the combustive mass by reducing its temperature. These cycles located to the left of diagram 11 S lead to exhaust temperatures that are too low to initiate the turbocharging, which is installed only from a minimum power. This makes it difficult to switch from idling to take-off conditions of the vehicle during the clutch where the quantity of air is insufficient to provide the desired torque.

When the power increases, the cycle moves to the right of the T / S diagram to stabilize at the turbine-compressor equilibrium point.

The present invention provides for executing very low loads and high speeds on the right of the T / S diagram by re-aspirating hot burnt gases to which the quantity of air necessary for combustion is added.

This strategy maintains an exhaust temperature level which prevents cooling of the exhaust manifold and scrolls of the turbines at idle or during non-propelled phases, cooling which delays the acceleration of the turbochargers during subsequent recoveries.

It also improves the efficiency of catalytic depollution. Finally, it makes it possible to reduce the effective compression rate.

Operation without load in atmospheric suction is carried out at a high rate of hot EGR to maintain the thermal level of the reaction zone, the walls of the exhaust duct and catalytic post-treatments. When the load increases from idle, priority is given to establishing turbocharging, the engine behaving like a generator of hot gases.

The EGR valve can be temporarily closed during acceleration of the turbochargers to benefit from pulsed pressure and reach the turbocharged phase at the clutch point without EGR.

After the clutch, the EGR rate and temperature return to the program selected for pollution control, which preferably maintains the cycle on the right of the 11 S diagram to minimize the response time on occasions. In fact, the higher the speed of the turbochargers at partial load, the higher the excess air available for recovery.

We now refer to FIGS. 3A, 3B, 9 and 10.

The invention has the advantage of concentrating the operating points of the compressors on lines of their characteristic diagrams.

Diagram 3A describes three operating modes in serial configuration represented in diagram 10:

- all the city mode points are on the OA segment, - all the road mode points are on the AB segment,

- all the points of the highway mode are on the BC segment.

The boundaries with constant P2 between the modes are respectively represented by points A, B and C where the aerodynamic regime of the compressors is stationary. Likewise for diagrams 3B and 9:

- all the points of the city mode are on OA,

- all the points of the route mode are on ABC,

- all the points of the highway mode are on PC.

The air flow delivered to 4.5 bar by the compression group depends on the positioning of the operating points in the characteristic diagrams of the compressors. A good adaptation allows a variation of flow between 1 and 3 approximately. Such an adaptation implies precise management of the power of each turbine and of the flow section offered to the exhaust gases. The flow rate of the compressors is set by the flow section of the exhaust system. It depends on the pressure and temperature of the gases.

In order to minimize transfer losses, the present invention plans to operate mainly with an exhaust pressure P3 equal to the intake pressure P2. The turbines are therefore supplied at a substantially constant pressure. This relationship requires, for a fixed geometry of the exhaust system, that the compressors operate on a single line of their operating diagrams.

The driving pleasure implies that the maximum cleaned-up torque is available at 25% of the maximum speed. We see that for P2 = 4.5 bars, the compressor flow range corresponds to the speed range at constant maximum torque between 12, 5% and 37.5% of the maximum speed.

Between 37.5% and 100% the motor at full load can only operate at constant air flow and therefore at constant power. With a fixed valve timing, the 4-stroke engine draws in a volume proportional to the speed which becomes more than double the volume delivered by the compressors above 25% of the maximum speed.

Reference is also made to patent application WO 02/48510, filed on December 14, 2001 under the PCT number FRO1 / 04006 under French priority N ° 0016422 of December 15, 2000.

This request provides for filling this excess volume by expanding the compressed air by rolling the air flow between the compression group and the cylinders.

It plans to perform this expansion by rolling the intake valves or by early closing of the latter.

The present invention provides for controlling this expansion by an intake valve located at the outlet of the compression group in order to be able to fill with fresh air a fraction of the excess volume under certain exceptional operating conditions. Under normal conditions, the present invention provides to occupy the excess volume from the emitted by the engine exhaust gas via a conduit between the exhaust manifold and the intake manifold sufficiently permeable so that the transfer takes place substantially at P2 = P3. The clean-up area therefore extends over the entire range of regimes. The exhaust pressure is fixed, the flow rate of the compressor depends only on the permeability of the exhaust system.

The present invention provides several solutions depending on the extent of the range at constant pressure and the quality of the turbomachines used. In the order of increasing ranges, the following structures will be mentioned without limitation:

1) Turbines in series with waste waste on the HP turbine (1 to 1, 8),

2) Turbines in series with variable distributor of the HP turbine (1 to 2),

3) Turbines in series / parallel with double feed waste between the turbines (1 to 2.4),

4) Turbines in series with variable distributors of HP and BP turbines (1 to 3),

5) Turbines in parallel (1 to 3),

6) Turbines in parallel series with variable distributor of the HP turbine (1 to

4). The invention also provides for extending the above ranges by approximately

30% by winnowing the recycling duct if the exhaust manifold can withstand a pressure of 6 bars. Under these conditions P3 = 1.33 P2.

This extension is no longer realistic above 3, the flow range is essentially limited by the efficiency of the compressors at maximum flow.

For engines with fixed valve timing, the invention favors structures 3 and 6 which includes structures 1, 2 and 5.

Structure 3 benefits from the good efficiency and simplicity of fixed geometry turbines. . The structure 6 allows a better series / parallel transition.

The volume of recycled gas depends on the speed and the flow section of the exhaust system.

At maximum speed the recycled gases occupy 87.5% of the displacement Sd for minimum and maximum 62.5% for Sd. For an exhaust temperature of 990 ° K and a fresh air temperature of 330 ° K, the corresponding mass ratios of uncooled EGR are 70% and 37.5% respectively.

The corresponding temperatures of the admitted mixture are 518 ° C and 304 ° C. We therefore see that the rate of 50% at maximum Sd implies refrigeration of the gases from 990 ° K to 550 ° K which gives an inlet temperature of 167 ° C after mixing.

The NOX production decreases with the oxygen concentration of the oxidizing mixture. The present invention plans to operate mainly with the oxygen concentration and the richness which give the best NOX-particle compromise with the post treatment used. It is therefore necessary to control the mass rate and the EGR temperature according to the best particle NOX compromise stored in the map of the computer which controls the engine.

In the prior art, the excess air is regulated by the exhaust section which acts on P2 and the EGR rate by the EGR valve which acts on P3 / P2.

The present invention where P3 / P2 is predominantly constant plans to act on the temperature of the recycled gases. Means are provided for cooling the recycled gases preferably to the temperature of the fresh air. Above a certain mass rate of EGR (about 50%), this cooling too reduces the enthalpy available for the turbines.

The invention therefore provides for limiting the energy dissipated in the EGR refrigerant by a pilot bypass of said refrigerant.

For an exhaust section, a torque and a given regime, P2 increases and the recirculation rate decreases when the EGR temperature increases. These two effects accumulate to increase the excess air and the oxygen concentration are therefore bound by a relationship. Temperature EGR is therefore an effective parameter for choosing the NOX-particle compromise.

To simplify the regulations and optimize the use of the exhaust enthalpy, the present invention provides operating modes with fixed exhaust geometry and without flow lamination.

Only the EGR temperature is controlled according to an imposed EGR map. This configuration will always give the best compromise between excess air and consumption for the EGR rate chosen.

For example, urban driving where accelerations and decelerations follow one another at reduced power is done with turbines with fixed geometry working in series, by controlling only the temperature of the re-aspirated gases according to an EGR map.

The following description relates, by way of examples, several operating modes selected for city driving, road and highway.

The list is not exhaustive.

GROUP A. 4-STROKE ENGINE WITH FIXED VALVE TIMING.

The numerical example is the driving force behind diagrams 3A, 3B, 9 and 10.

Mode A 1: Clean-up urban driving.

Power <0.33 Wmax turbines working in series

Padm = P3 <4.5 bars

N> 0.2 Nmax The double spoiler waste 103 is in the high position.

The inlet valve 105 is fully open

The EGR 105 valve is fully open to maintain the relationship

Padmission = P2 = P3.

The guillotine 106 of bypass the EGR cooler is in regulation. The bypass refrigerant EGR inlet temperature program for a mapping EGR set in advance, for example in memory of the computer which controls the engine.

The compressors deliver under these conditions a minimum volume of cooled air which reaches 4.5 bar the volume drawn by the motor at 0.1 Nmax. When the speed increases the engine completes with a volume of recycled gas which reaches 2 times the volume of air at 0.2 Nmax and 10 times at Nmax.

The OA adaptation curve of the compressors in the characteristic diagrams is shown in Figure 9. The mass fraction recycled depends on the temperature of the gases. To obtain 50% at 0.2 Nmax, the gases must be cooled to air temperature by closing (106) the EGR refrigerant bypass.

To maximize the enthalpy available for the turbines when the speed increases, the temperature of the recycled gases must be increased by gradually opening the guillotine 106 according to the speed, the quantity of fuel burned and the desired EGR rate.

The EGR 106 bypass is fully open for starting, idling and operating at very low load to maximize the thermal level in the combustion chambers to reduce noise and unburned and maximize the speed and temperature of the turbochargers to optimize their recovery capacity.

If necessary, the recycling rate in steady state can be refined by a fixed or variable half-open position of the EGR valve 104.

As soon as the engine loads, the EGR valve fully opens and the guillotine switches to inlet temperature regulation which increases appreciably from 60 to 450 ° C when the speed increases from 0.2 Nmax to Nmax the mass rate of EGR going from 50% to 70%.

The operating range of the cleared mode, which implies that the continuous control of the guillotine only 106, is the city area of the SME chart / system 3B. It covers the entire city driving where transients are common. The absence of discontinuity has a definite advantage in this context.

The inlet pressures and temperatures map in the domain is a function of the map introduced in the engine control computer for piloting the guillotine 106.

The upper limit of the domain corresponds to the inlet and exhaust pressure limit set here at 4.5 bars.

Inside this mode, the adaptation of the turbochargers can be slightly modified by partially closing the EGR valve 104 to have P3 / P2> 1. This may prove necessary to improve the pumping guard of the compressors very low in this fashion.

The domain of this mode can be extended to high powers by closing the EGR valve (104) so that P2 = 4.5 bars and P3> 4.5 bars. The limit of this extension is the P3 authorized by the exhaust manifold technology.

This extension mode is simple and increases excess air at the expense of consumption. It is therefore interesting in urban driving where the large area is little used and where the transients are rapid.

The passage from P3 = 4.5 to P 3 = 6 bars makes it possible to increase the area covered by mode A1 by 30% under the following conditions.

We now refer to FIG. 8 which describes the operation of the double spoil waste.

Mode A11:

Clean-up city accelerations. Power between 0.33 and 0.44 Wmax

Turbines working in series Pperm = 4.5 bar 4.5 bar <P3 <6 bar N> Nmax 0.27 The double spoil waste 103 is in the high position.

The inlet valve 105 is fully open

The EGR valve 104 is in regulation to maintain the intake = 4.5 03 rs. The bypass 106 of the EGR refrigerant programs the intake temperature to obtain an EGR map fixed in advance, for example on the memory of the computer which controls the engine.

Mode A12 Turbines working in series

Padm = 4.5 bars N> 0.25 Nmax

A2 mode: extension of A1 mode limited by turbine flow

Clean-up road driving Power between 0.33 W max and 0.5 Wmax

Turbines working in parallel series regulated by HP waste only. Padm = P3 = 4.5 bars.

The double spoil waste 103 is in the high position.

The EGR valve 104 is open. The inlet valve 105 is fully open

The guillotine 106 is programming the intake temperature.

The HP part of the spoiled waste opens to maintain P2 = P3 = 4,

5 bars

We now refer to Figure 8. To fix the ideas, let's start from the border of mode A1 where the initial conditions are as follows:

P2 = P3

The pressure between turbines P4 = 2 bars

The compressors are at points A on their diagrams. Compressor flow = Q

When the HP waste container opens gradually to regulate P2 = P3 = 4.5 bars until contact with the BP waste container. The compressor flow increases to increase the power delivered by the engine. During this maneuver, the exhaust temperature T3 has increased to compensate for the energy dissipated by rolling; the pressure P4 between turbines increased from approximately 2 to 3 bars; the total throughput of the compressors passed through the BP turbine, the expansion rate of which went from 2 to 3. The throughput of the compressors therefore went from Q to 1.5 Q and their operating points went from A to B; the HP turbine flow rate decreased when its expansion rate went from 2.25 to 1.5; the LP compressor accelerated and the HP compressor slowed down to a point where the flow rate no longer increased.

A3 mode: extension of A2 limited by the exhaust temperature Clean-up driving.

Power between 0.5 Wmax and 0.8 Wmax

Turbines working in parallel series regulated by HP and BP waste gates controlled simultaneously.

Padm = P3 = 4.5 bars The double spoiler waste 103 is in the high position.

The EGR valve is open.

The inlet valve is fully open

The guillotine 106 is programming the intake temperature.

Waste spoils HP and BP open to maintain P2 = P3 = 4.5 bars

When the fuel flow continues to increase, the two waste tanks, of the double waste tank, pressed one on the other open simultaneously to maintain P2 = P3 = 4.5 bars. The rolling ramp (see Figure 8) of waste spoils on this part of their course determine the change in pressure between turbines P4 during this phase.

P4 determines the operating points of the compressors in their diagrams. The ramps will be defined for a progressive path between points B and C. Let us assume for simplicity that this path leads to P4 = 2.7 bars.

During this maneuver, the temperature T3 increased further to compensate for the energy dissipated in the discharge of the waste tank BP. The flow rate of the two turbines did not change significantly. The compressor flow increased from the discharge by the waste tank BP limited by the maximum admissible T3. Set this flow rate at 2.4 Q at point C.

If T3 exceeds the limit value imposed by the richness, one can act on the EGR valve to increase P3 and decrease T3. This operating mode can be achieved with two turbochargers equipped with conventional waste gates which limit the pressure delivered by the compressor. We can adjust the HP waste gat to 4.5 bars and the BP waste gat to 2.5 bars for example. method positions the BP compressor on a horizontal of its characteristic diagram.

Transition between A3 mode and A4 mode.

The initial conditions of the transition are constant on the border A4 A2. For the understanding of the presentation they are fixed as follows: P2 = P3 = 4,5 bars Waste treats HP and BP in maximum discharge P4 = 2,7 bars

Compressor flow = 2.4 Q LP turbine flow = 1.5 Q HP turbine flow = 0.75 Q LP expansion rate = 2.75 HP expansion rate = 1.64 The final conditions of the transition are:

P2 = P3 = 3 bars

Compressor flow = 2.4 Q

N furbo HP unchanged BP turbine flow rate = 1.6 Q

Turbine flow HP = 0.8 Q

BP expansion rate = 3

HP expansion rate = 3

The transition takes place by the rapid and simultaneous tilting of the waste trash to the lower seat of the waste trash BP. This maneuver is programmed on the limit T3 curve. To prevent the motor from stabilizing on the transition, the tilt line A2 / A5 will be different from the tilt line A5 / A2.

During this maneuver, the P2 and the P3 go from 4.5 to 3 bars. The compressor flow remains unchanged, the EGR flow decreases, the turbine flow varies little, the BP expansion rate varies from 2.75 to 3, the expansion rate

HP goes from 1, 64 to 3, the speed of the LP turbo increases slightly, the speed of the HP turbo decreases slightly.

Direct transition between A1 and A4 mode It will be done by a short passage through A4 and A5 to limit discontinuities.

Mode A 4:

Driving on low-pollution economic motorways

Power between 0.33 Wmax and Wmax Turbines working in parallel.

Padm = P3 <4.5 bars.

N> 0.35 Nmax

The double spoil waste 103 is in the low position.

The intake valve is fully open The EGR 104 valve is fully open to impose the relationship

P2 = P3.

The EGR refrigerant bypass programs the intake temperature to obtain a map of EGR fixed in advance, for example in the memory of the computer which controls the engine.

The compressors operate on the PC adaptation curve of Figure 9 with a different distribution of pressure ratios due to the fact that the powers developed by the turbines are in the constant ratio of their permeabilities. The area covered by this mode is the highway area of the diagram

3B. As in the previous mode, its upper limit corresponds to the inlet and exhaust pressure limit set at 4.5 bars. It covers the essentials of economical driving on motorways without operating discontinuity. This mode with a high expansion rate of the two turbines allows post-treatments with high pressure drop.

GROUP B: 4-STROKE ENGINES; VARIABLE SETTING OF

VALVES Group A brings together processes where the variable geometry is located at the level of the turbocharging group to modulate from 1 to 3 the volume Vc of cooled air delivered by the compressors at a pressure limited to a set point of 4.5 bars imposed by the high compression ratio of the engine.

As mentioned above, controlling the valve timing allows the volume drawn in by the engine to be modulated from 1 to 2 approximately at a given speed by positioning FA (intake closure) at mid-stroke of the piston and FE (exhaust closure) at TDC.

The decrease in the volume flap is accompanied by a decrease in the compression ratio that can double the set pressure 9 bar. In addition, a modulation of only 1 to 1.5, of the volume Vc makes it possible to cover the entire speed range in the series configuration, and to avoid the series / parallel discontinuity (FIG. 10).

In a 4-stroke cycle the volume of air drawn in is substantially the difference between the total volume of gas present in the cylinder defined by the volume of the FA chamber and the volume of residual burnt gas defined by the volume of the FE.

The recycling mechanism then breaks down into hot internal recycling and cooled external recycling, the mixing taking place inside the cylinder. The proportion between the hot EGR and the cold EGR can be controlled by the timing of the valves.

Piloting FE is therefore another means of modulating the volume drawn in by the engine without modifying the compression ratio.

NOx reduction implies a reduction in temperature during combustion which starts at the end of compression temperature. It is therefore favorable to minimize the compression ratio and the EGR temperature.

Furthermore, the triangular loss of cycle which accompanies the Joule expansion when the exhaust is opened can be recovered by two turbines in series which take care of part of the work of compression of the piston.

To maximize the torque at low speed and facilitate cold starting, FA must intervene in the vicinity of the bottom dead center of the piston to maximize the trap volume at low speeds. To increase the intake pressure while respecting the pressure limit in the cylinder, it is therefore necessary to reduce the effective compression ratio by advancing or delaying FA.

The easiest way is to delay FA by accepting the reflux of part of the air sucked into the intake manifold. Two series compressors deliver naturally a pressure proportional to the power that can reach 8 to 10 at full power bars.

If the maximum inlet and exhaust pressures go, for example, from 4.5 to 9 bars, the area in A1 mode is doubled at the cost of a high pressure technology for HP ducts facilitated by the reduction of 50% cutting sections.

The delay FA must then go from 20 to 90 degrees crankshaft (dv) approximately to divide the hatch volume by 2. This strategy makes it possible to operate the compressors in the restricted range of 1 to 1.5, where the yields can be optimized to tolerate a bypass rate of the HP turbine close to 30% at full power.

Under these conditions, the extended A1 mode covers the entire domain of the modes A2, A3 with a total pollution reduction rate and a regulation limited to the control of the EGR bypass and the angle of FA.

HP waste waste can be managed as a simple pressure relief valve.

This Blet mode and its extension differ from A1 mode and its extensions by the FA control which regulates the EGR volume for each engine speed and therefore the pressure P2.

The FA control can replace the EGR refrigerant by-pass control which modulates the EGR temperature. Maintaining the piloted bypass nevertheless has the following advantages:

Maintaining the A1 mode when P2 <4.5 bars, to avoid the phase shift of the camshafts in the rapid transients of urban driving, FA can then be controlled more slowly at high power to limit the cylinder pressure when P2> 4.5 bars.

Can set two or three angle fixed values FA and maintaining control based on the bypass controlled. Maintains hot EGR at very low power. Reduced ignition delay very delayed FA where the effective compression ratio is very low.

B1 mode: extended A1 mode replaces A2 and A3

Cleared city and road driving; Power between 0.33 Wmax and 0.67 Wmax

Turbines in series

4.5 <Padm = P3 <9 bars

FE to TDC

Closure of programmed intake to limit cylinder pressure N> 0.2 Nmax

The double spoil waste 103 is in the high position.

The inlet valve is fully open

The EGR 104 valve is fully open to maintain the relationship

Padmission = Exhaust (P2 = P3). Guillotine 106 is under regulation.

The EGR refrigerant bypass programs the intake temperature to obtain an EGR map fixed in advance, for example on the memory of the computer which controls the engine.

The angle of FA is controlled discreetly or continuously to limit the maximum pressure in the cylinder.

B2 mode: extended B1 mode external EGR only.

Driving on a clean highway

Power between 0.67 Wmax and Wmax

Turbines in series / parallel Padm = P3 = 9 bars

FE to TDC

FA halfway.

N> 0.4 Nmax

The dual wastegate 103 is in the high position. The intake valve 105 is fully open The EGR valve 104 is fully open to maintain the relationship between intake and exhaust (P2 = P3).

The EGR refrigerant bypass programs the intake temperature to obtain an EGR map fixed in advance, for example on the memory of the computer which controls the engine. FA remains in the delayed position.

The waste spoils HP opens to maintain P2 = P3 = 9 bars. The usual variable distribution device consists of a controlled phase shift of the camshafts which generally control two valves per shaft.

The phase shift of a valve therefore results in the same phase shift on another valve, the effect of which is negligible in the vicinity of the piston dead centers.

For example, the variation of FA only is possible for a cylinder head with 4 valves and two camshafts which actuate respectively an intake valve and an exhaust valve with the diagrams below:

A fixed camshaft controls the opening of the intake phase as well as the opening and closing of the exhaust phase with for example: (R = delay, A = advance, O = opening, F = closing, 1 = fixed camshaft, 2 = phase-shifting camshaft). Values are in crankshaft degrees.

ROA1 = 0 AOE 1 = 20

RFA1 = 20. AFE1 = 0 The other phase-shifting camshaft only controls the closing of the intake phase with for example: P2 <4.5 bars

ROA2 = 0 AOE2 = 20

RF A2 = 20 AFE2 = 80 P2 = 9 bars ROA2 = 80 AOE2 = - 60

RFA2 = 100 AFE2 = 0 B3 mode: extended A1 mode. External and internal EGR. Replaces B2 Variable intake closure, exhaust closure and intake opening.

Turbines working in series 4.5 bars <P2 = P3 <9 bars The double waste 103 is in the high position. The inlet valve 105 is in full opening The EGR valve 104 is in full opening to maintain the relationship

Padmission = P2 = P3.

The EGR refrigerant bypass 106 is closed or eliminated The hot internal EGR is controlled by FE The cooled external EGR is controlled by FA OA intervenes approximately when the cylinder pressure crosses

padm

FA is controlled to limit the maximum pressure of the cycle. The AF delay is accompanied by an equal delay of OA which occurs 80 degrees after TDC at high speed. To avoid pumping losses, FE must be simultaneously advanced to retain hot gases, the pressure of which crosses the inlet pressure at OA.

The oxidizing mixture is carried out in the cylinder during filling with a possibility of stratification described in US Pat. No. 5,551, 954. The intake valve (s) are actuated by a phase shifted intake camshaft.

The exhaust valve or valves are actuated by a second phase-shifted exhaust camshaft.

The rate of hot EGR is controlled by FE. The additional cold EGR is managed by FA. The phase shifters are actuated according to a mapping of the EGR rate hot and cold.

If the exhaust section is undersized, the pressure at the TDC crossover naturally increases with the speed of the pistons. It will therefore suffice for a more reduced phase shift of the exhaust shaft. We can choose for example: P2 = P3 <4.5 bars ROA = 0 AOE = 0

RFA = 20 AFE = 25 P2 = P3 = 9 bars

ROA = 80 AOE = 30 RFA = 100 AFE = 55

Part of the gases are thus retained inside the cylinder and mixing with the cold gases takes place during the filling phase with the possibility of stratification. The exhaust phase shift then replaces the refrigerant bypass to regulate the EGR temperature.

These cycles, which exhibit gas recompression at the TDC crossover, can create known piston axis lubrication problems on both stages.

B4 mode: extended B2 or B3 mode

Turbines in series / parallel

P2 = P3 = 9 bars

The double spoil waste 103 is in the high position.

The inlet valve 105 is in full opening The EGR valve 104 is in full opening to maintain the relationship

P2 = P3.

EGR refrigerant bypass 106 is closed or eliminated

FA remains in delayed position

FE is delayed to reduce the internal EGR when the fuel flow increases, The waste spoils HP opens slightly to maintain P2 = P3 = 9 bar.

Mixture of hot EGR and fresh gases in the cylinder.

Feeding methods which leave residual burnt gases after closing the cylinder have the advantage of allowing stratification of temperatures and concentrations in the combustion chamber.

This advantage is exploited by the invention described in US Pat. No. 5,517,954 to reduce the ignition time of diesel engines with low compression ratio, the ignition point of which is regulated by direct injection of the liquid fuel at high pressure. .

US Patent 5,517,954 is incorporated here by reference for obtaining the stratification.

Recently, homogeneous premixed combustion processes have appeared, in which the fuel is vaporized in the oxidizing charge before or during the compression phase which precedes combustion in the cylinder, such as spark-ignition engines.

These new processes are distinguished by the fact that self-ignition is triggered by heating due to compression. The pitfall of these solutions is the explosive nature of the simultaneous self-ignition at all points of a volume of the combustion chamber.

The present invention makes it possible to control the supply of the engine with fresh air, cooled external EGR and uncooled internal EGR in order to create a stratification of the temperatures and concentrations in the chamber at high combustion neutral point.

For engines wherein the fuel is vaporized into the oxidizing charge before or during the compression, the invention provides to arrange the lamination so that the autoignition by compression extends gradually warm to cold areas of the combustion chamber zones when the gas pressure increases. Simultaneous autoallumage thus relates to surfaces of the reactive load rather than detonating volumes.

The instantaneous energy release rate is proportional to the value of the surface being self-ignited and the concentration of fuel vapor on that surface.

In addition to the temperature and concentration gradients, there are therefore geometric parameters of the combustion chamber to adjust the progressivity over time of the release of energy.

The ignition point is advantageously controlled by the angle of the crankshaft at the closing of the cylinder which simultaneously regulates the quantity of hot gases recycled and the effective compression rate which brings them to the self-ignition temperature.

Another mode of piloting the ignition point is the EGR temperature. Fuel can be vaporized in clean air between the compressor and the external EGR mixer for precise adjustment of the richness of the reactive charge.

Fuel can also be introduced into the external air / EGR mixture upstream of the cylinder, in the intake ports or inside the cylinder before and / or during compression.

This compression self-ignition of a stratified carburetor charge can lead to non-explosive combustion without particles or NOX if the local temperature and concentration conditions are well controlled.

The first ignition can be triggered by an electric spark or high pressure injection.

To guarantee a symmetry of revolution of the oxidizing charge, it is advantageous to direct the flows of cold gases introduced into the cylinder to create a rotational movement of the oxidizing charge.

The most suitable structure for controlling the stratification is the axisymmetric 2-stroke engine described in US Patent No. 5,555,859, being understood that in the present invention the fuel can be vaporized as described above

US Patent 5,555,859 is also here incorporated by reference for obtaining the stratification. A more conventional architecture nevertheless allows an organization of revolution around the axis of the cylinder by the following arrangements:

Modes B2 and B21 provide for an intake volume limited to the volume drawn in by the engine at 1400 rpm. The cross-section of the intake ports can therefore be reduced to 28% of their normal value. This makes it possible to use only one sector of the variable orifice constituted by the valve and its seat.

Two well-suited geometries are described below for a 4-valve cylinder head:

FIG. 11 represents the conventional architecture of a plane cylinder head carrying four valves with axes perpendicular to the cylinder head plane and whose cylinder side faces are in the cylinder head plane in the closed position in order to respect the revolution geometry of the gas work.

The two intake valves A are identical and diametrically opposite to create a flow symmetrical with respect to the axis of the cylinder which will acquire the symmetry of revolution at the end of compression. They are placed as close as possible to the cylinder.

The two exhaust valves E are identical and diametrically opposed on a diameter which can be offset by 70 to 90 degrees compared to that of the intakes to avoid interference between the air jets and the heads of the exhaust valves in situation of crossing (case of 2 times).

The injector is conventionally located in the center of the valve pattern.

The toroidal combustion chamber is located in the piston and it is coaxial. Its collar is chosen to create the desired layering. Filling phases and discharge are largely separated there is no risk of interference between the piston and valves. The piston thus includes no recesses for the valves that would destroy the geometry of revolution. To generate a rotary flow, the intake pipes direct the flow towards the piston tangentially to the cylinder in the vicinity of the valve seats. Two architectures are possible depending on the valve lift:

Inlet nozzle suitable for heavy lift. The intake pipes 151 terminate upstream of the seat by an oblong convergent nozzle, the neck 152 of which is defined by an upper half-cylinder resting on the upstream edge of the conical seat and tangent to the latter along its generatrix located in a plane substantially perpendicular to the plane passing through the axis of the seat and through the axis of the cylinder and through a lower cylinder covering half of the valve head opposite to said generator.

The lifting of the valves is such that they do not interfere with the fluid flow coming from the nozzle, this at least at full opening.

The nozzles will also be oriented to create a tangential speed in opposite directions. Inlet nozzle suitable for low lift. The conical sealing surface of the intake valves is extended by a cylindrical part of height slightly greater than the lifting of said valves, the conical seats of said valves are arranged at the bottom of cylindrical housings 153 arranged in the cylinder head to receive said cylindrical parts of said valves, so that the flat underside of the valves is in the plane of the cylinder head when they rest on their seats. The diametral clearance between the housings and the cylindrical extensions of the valves is minimal.

Recesses are made in the cylinder head within the following borders: 1) Two cylinder portions concentric with the bore and tangent externally and internally to the obviously cylindrical of each valve, the external cylindrical portion being able to be confused with the cylinder. 5 2) A conical surface extending the half-seat of the valve delimited by a plane passing through its axis and the axis of the cylinder.

3) The recesses will also be oriented to create a tangential speed in opposite directions.

The angle of the intake seats is chosen between 90 and 120 degrees to 0 optimize the stratification of the oxidizing charge.

GROUP C: SELF-SCANNING 2-STROKE ENGINES WITH VARIABLE VALVE TIMING.

In a recycling context, the two-stroke cycle with variable self-swept distribution has the following advantages:

> Natural hot internal recycling and disappearance of the historical 2-stage filling handicap.

> Doubling of the specific power, the volumetric efficiency of the two cycles being identical. 0> Reduction of the delay between the generation of active radicals present in the burnt gases retained in the cylinder and their use in the following cycle.

> Reduction in the time of residual gas presence in the cylinder and associated thermal losses. 5 Disappearance of the 4 stroke pumping losses.

> Less friction losses in the absence of a sweep compressor coupled to the motor.

> Double rate of use of high pressure injection units.

> Large and adjustable variation of the ratio between expansion rate and o compression ratio. > Excellent transition between the atmospheric phase and the turbocharged phase. On the other hand, the cooling of the EGR implies special arrangements. To simplify the comparison, the displacement of the 2-stroke engine is half that of the 4-stroke engine described above and the specifications are identical.

It does not have a sweep pump driven by the motor shaft.

In order to decrease the temperature at the start of combustion and to increase, the turbocharged propulsion phases are executed on an asymmetrical cycle with a high expansion rate and a low compression rate.

This also makes it possible to push the intake pressure while respecting the authorized limit as in modes B.

The initial part of the compression stroke is then used for gas transfers.

The evacuation of the burnt gases is ensured by one or more exhaust orifices located in the cylinder head and closed by valves controlled by one or more cam shafts which are phase-shifted relative to the crankshaft. The filling is ensured by lights at the bottom of the cylinder opened by the piston in the vicinity of the bottom dead center or, preferably, by at least one inlet orifice located in the cylinder head and oriented towards the piston and closed by at least one valve. intake controlled by one or more fixed or phase shifting cam shafts. The directive valve architecture already described in FIG. 11 can be used to sweep a 2-stroke cycle.

The feeding philosophy is similar to modes B. We now refer to Figure 6.

The feed structure is that of 4 times of which the intake valve and the EGR bypass have been removed. The EGR valve 104 has been replaced by a check valve or an aerodynamic diode 204.

The turbocharging is adapted to 0.1 Nmax and capable of a maximum pressure of 9 bars. The turbocharging unit is of the series configuration type defined for processes B.

The filling with cold gases is carried out when the pressure in the cylinder is lower than the intake pressure after the discharge of the pressure Po at the end of expansion. External recycling implies that the gas pressure is higher than the intake pressure.

This condition is met periodically in a pulsed exhaust manifold.

The pressure Pf in the cylinder at the closing of the cylinder is the end of sweeping pressure, close to the inlet pressure.

The fraction of gas that can be recycled externally against the intake pressure is therefore substantially equal to Po / Pf where Po is the pressure in the cylinder at the opening of the exhaust phase.

The mass of gas being constant between FE and OE we can write: Po / Pf = Vf / VO x To / Tf

Where Vo and Vf are the volumes occupied by the gases at OE and FE Where To and Tf are the absolute temperatures of the gases at OE and FE The ratio Vf / Vo, which is none other than the ratio between the expansion rate and the compression ratio, can be controlled by phase shift of the camshafts.

As the thermodynamic efficiency improves when Vf increases, the OE settings are favored in the vicinity of the PMB.

Tf is the total temperature at the inlet of the turbine which varies little between 900 and 1100 ° K for a turbocharger in two stages where P3 neighboring rest of P2 To is the mixing temperature of fresh air at 330 ° K, cold EGR at 330 ° K and internal EGR at Tf.

For 50% of cold EGR it is necessary that: Po / Pf = 2

For the maximum expansion rate: Vo = C (cylinder volume at

Let us choose for example; Tf = 990 ° K

In cold EGR: To = 330 ° K

So it comes: Vf / Vo = Po / Pf x Tf / To = 2/3

We can therefore say that a valve timing respecting these conditions generates a cold EGR rate close to 50% when the excess volume is occupied by clean air.

This percentage varies rapidly around this setting and can therefore be controlled easily.

For OE at PMB and FE at 70 dv after PMB we have: The external EGR occupies, at pressure P2, a volume substantially equal to C and C / 3 after cooling.

The hatch volume being 2C / 3, the volume available for fresh air and residual gases is equal to C / 3.

The turbocharging adapted to 0.1 Nmax will therefore give an external EGR rate of 50% 0.3 Nmax

The internal EGR increases linearly above this regime by increasing Tf.

It follows that Pf / Po = 2/3 x 990 / To decreases with the regime.

The internal EGR gradually replaces the external EGR when the speed increases with fixed valve timing.

To obtain the rate of 50% at 0.2 Nmax it is necessary to advance the valve timing to increase Vf / Vo until obtaining Pf = 3 Po

We made the assumption that the entire exhaust breath is recycled externally. This is not realistic with the present architecture. A more detailed description of exhaust manifold pulsed modes of operation is given below.

The non-return EGR valve may be noisy, we choose the aerodynamic diode for silent engines. The diode can be summed up in a calibrated orifice comprising a non-converging inlet pavilion on the collector side ensuring a flow coefficient close to 1 and an outlet with sharp edges offering a coefficient of 0.5 at the reflux flow.

The exhaust puff is shared between the turbine and the recycling duct in proportion to their flow sections.

The flow entering the diode benefits from a high expansion rate which can reach 3. The reflux flow generated by an expansion rate limited to 1, 2 is, in addition, divided by 2 by its flow coefficient.

The diode can be dimensioned to absorb the mass of gas sufficient for external recycling. This mass can be controlled by FE and / or OE which act on the pressure Po at the end of expansion.

Fine adjustment is ensured by the variable section of the diode controlled for example by a conical central body connected to an actuator capable of completely closing the diode. FIG. 6 describes the simplest case of a single phase-shifting camshaft 200 controlling the intake and the exhaust with the following sequence:

OA = 30 dv after FE

FA = FE = 70 dv after OA Cold start, acceleration of turbochargers, engine brake.

Turbines working in series Diode 204 closed OE = 100 dv before PMB OA = 70 dv before PMB FE = FA = PMB After OE the gases at high pressure and temperature are expanded on the HP turbine via valve E and leave a vacuum in the cylinder supplemented by fresh air at OA. A very hot mixture is recompressed in the second cycle to quickly install a sufficient thermal level for silent ignition and efficient catalysis.

Economic slowdown

Turbines working in series

Diode 204 closed

OE = 80 dv before PMB OA = 50 dv before PMB

FE = FA = 20 dv after PMB

C1 mode: depolluted urban driving power <0.33 Wmax Fixed turbines working in series

P3 pulsed around P2 <4.5 bars

External EGR rate controlled by the diode section.

OE = 50 dv before PMB

OA = 20 dv before PMB FE = FA = 5O dv after PMB

The exhaust puff carries a fraction of the mass of hot gases present in the cylinder which is a function of Pc / Pad and which varies very quickly with the setting of FE and OE distant by 100 dv in the present case.

We have previously seen that in cold EGR this fraction goes from 3 for a symmetrical setting FE = 50 dv after PMB to 1 for FE = 110 dv after PMB.

The advance of the camshaft thus simultaneously increases the flow rate in the turbine and the recycling conduit in proportions which can be set by the section of the diode. A fixed section of the diode, an advanced timing P2 increases and the rate of external EGR.

With fixed valve timing, the opening of the diode decreases P2 and increases the rate of external EGR. 5 With a fixed diode and setting section, increasing the speed decreases the fresh air delivery rate which is replaced by the hot internal EGR with movement of the cycle to the right of the entropy diagram.

The Po / Pf ratio decreases when the speed increases from 3 to 1400 rpm to 1, 3 for 5000 1 / min. 0 In summary, the hot EGR gradually replaces the cold EGR when the speed increases.

When the average flow through the diode is canceled, it must be closed to avoid an exhaust intake bypass.

Mode C2: 5 Driving on partially depolluted road.

Power varying from 0.33 to 0.66 Wmax

Stationary turbines working in series

4.5 bars <P2 <9 bars

N> 0.2 Nmax o Active diode at low speed.

FE programmed between 60 and 90 dv after PMB depending on P2 to limit the maximum pressure of the cycle

When P2 goes from 4.5 bars to 9 bars, it is necessary to delay setting to respect the limit pressure in the cylinder. 5 The external EGR decreases in favor of the internal EGR X.

C3 mode:

Driving on low-pollution motorways

Power varying between 0.66 and wmax

P2 = 9 bar controlled by the opening of the HP turbine (waste gate or variable o distributor). N> 0.4 Nmax

Closed diode

FE = 90 dv after PMB

GROUP D: STRUCTURE ADAPTED TO FUTURE ENGINES For future 4 or 2 stroke engines, the invention provides a second supply structure described below in the case of 2 stroke (Fig N ° 7):

It has the advantage of externally recycling the entire exhaust puff. The engine comprises two exhaust valves per cylinder, one of which (ER) is assigned to the external recycling duct and the other (ET) to the supply of the turbine, the engine then carrying two exhaust manifolds CT and CR.

A distributor valve VD makes it possible to distribute the flow of the collector CR between the collector CT and the recycling conduit.

When the two valves are open that the valve is in neutral position and that the pressure in the cylinder is higher than P2, the cylinder feeds in parallel the turbine and the recycling, the turbine taking off as a priority its flow. When only one valve is open and the valve isolates the two manifolds, the cylinder supplies one or the other circuit.

The valve must close the recycling pipe when the cylinder pressure is below P2 and ER is open to avoid an exhaust intake bypass. The two valves then supply the turbine. In the case of a cylinder with intake lights, recycling is carried out by puffing before the power-based scanning and recycling by discharge after the scanning at high speed.

In the case of valve intake, the position of the Scanning at BDC, before recycling, has the advantage to limit the air loss in the exhaust.

In fact, the jets of fresh gas directed towards the piston are less disturbed by the speed of the gases being evacuated. In addition, the air mixed with the recycled volume 5 is not lost for combustion.

The specialization of the intake valves can also be considered when looking for a stratification of the oxidizing mass.

The description below relates to the second case of intake by valves with recycling before sweeping. 0 Take as an example a fixed sweeping camshaft 210 which controls the intake and AND and a phase shifting recycling camshaft 211, with phase shifter 212 which controls ER. The cams are calibrated as follows:

Scanning shaft: 5 OA = OET = 20 dv before PMB

FA = FET = 50 dv after PMB

Recycling tree

FER = 60 dv after OER

Cold start, acceleration of turbochargers, engine brake. o Turbines working in series

VD valve blocks recycling

OER = 90 dv before PMB

FER = 30 dv before PMB

OET = OA = 20 dv before PMB 5 FA = FET = 50 dv after PMB

After RJO high pressure gases and temperature are relaxed about the HP turbine through the ER valve and leave a vacuum in the cylinder supplemented by fresh air to PMB. A very hot mixture is recompressed in the second cycle to quickly install a thermal level o sufficient for silent ignition and efficient catalysis. Slow economic Turbines working in series The valve VD blocks recycling OER = 70 dv before BDC OET = OA = 20 dv before BDC

FER = 10 dv before PMB FA = FET = 50 dv after PMB

D1 mode

Clean-up urban driving. Power <Wmax / 3 Turbines working in series

The VD valve is in the recycling position.

OER = 60 dv before PMB

OET = OA = 20 dv before PMB

FER = PMB FA = FET = 50 dv after PMB

P2 <4.5 bars

N> to, 2 Nmax

The VD valve is in the recycling position.

There is a recycling rate whatever the regime.

D2 mode

Clean-up driving.

Power between 1/3 and 2/3 Wmax

Turbines working in series The VD valve is in the recycling position.

OET = OA = 20 dv before PMB

OER = 30 dv after PMB

FA = FET = 50 dv after PMB

FER = 90 dv after the PMB 4.5 bar <P2 <9 bars N> 0.4 Nmax

EGR cold

During the rapid transition, the ER and ET valves are opened simultaneously for a short time. 5 When Pcylindre> P2 a flow is established from the cylinder towards the recyling duct.

When Pcylindre <P2 a reflux replaces the cold EGR present in the purge air without disturbing the operation of the turbines.

Mode D3 o Driving on the highway.

Power varying between 2/3 and 3/3 Wmax with little pollution control. Turbines working in series.

The VD valve blocks recycling. The ER and ET valves supply the turbine with increased flow. 5 OET = OA = 20 dv before PMB

OER = 30 dv after PMB FA = FET = 50 dv after PMB FER = 90 dv after PMB P2 = 9 bars regulated by the opening of the HP turbine (Waste spoiler or variable dispenser).

N> 0.4 Nmax internal EGR hot.

Claims

1. Reciprocating engine used between a minimum rotation speed Nmin and a maximum speed Nmax which includes a turbocharging unit (2) sized to be in autonomous operation when:

> It supplies air to the intake manifold (8) of the engine via a refrigerant

> It is supplied with gas by the exhaust manifold (9, CR and CT) of the engine at the exhaust temperature> The turbine supply pressure (P3) is substantially equal to the compressor discharge pressure (P2)

In such a way that, at constant air temperature and with fixed geometry, the turbocharging delivers a volume of cooled air Vc that is substantially constant when the pressure varies. and that the volume Vc is substantially proportional to the flow section Sd offered to the hot gases, characterized in that the turbine pressure (P3) is maintained substantially equal to the compressor pressure (P2) by an EGR bypass (3) between the intake manifold (8) and the exhaust manifold (9) sized to transfer the flow of exhaust gas to the intake manifold without significant pressure drop, the volume of cooled air Vc is less than the volume drawn in by the engine at Nmax speed so that a flow of hot gases is drawn back in by the engine via the bypass (3) above the Na speed, known as the compression adaptation speed, where the volume sucked is equal to Vc, and an air flow is diverted towards the turbine below the Na regime.

2. Engine according to claim 1 characterized in that the EGR bypass (3) includes an EGR valve (6) for increasing the turbine pressure above the compressor pressure.

3. Engine according to claim 1 characterized in that the turbocharging group comprises an intake valve (7) located on the compressed air delivery pipe for increasing the compressor pressure above the turbine pressure.

5

4. Engine according to claim 1 characterized in that the EGR bypass pipe (3) comprises a gas cooler (4) at adjustable temperature preferably up to a temperature close to that of fresh air.

0 5. Motor according to claim 4 characterized in that the temperature is adjusted by controlling a bypass of the refrigerant.

6. A method of feeding an engine according to claim 4 characterized in that the EGR temperature is controlled to create the excess air 5 desired for combustion in the engine.

7. Supply method according to claim 4 characterized in that the EGR temperature is controlled so that the mass of recycled gas remains substantially equal to the mass of fresh air until the regime where this 0 temperature reaches the temperature d 'exhaust, beyond this regime the recycled mass becoming greater than the mass of fresh air.

8. Feeding method according to claim 5 characterized in that the air cooler is completely bypassed when the engine does not deliver propellant power.

9. A method of feeding according to claim 8 characterized in that for cold start and idling, the setting of the valves (6) and (7) and / or the valve timing is adjusted so that the o excess combustion air is minimal for the desired level of pollution control.

10. Engine according to claims 1 and 4 characterized in that the adaptation regime Na is substantially equal to Nmin / 2 so that the volume of recycled gas is at least equal to that of fresh air, the minimum temperature of recycled gases being preferably close to the temperature of the fresh air so that the mass of the recycled gases is at least equal to that of the fresh air at the minimum speed used N min to clean up the entire field of use of the engine .

11. Engine according to claim 1 characterized in that the turbocharger group comprises a low pressure turbocharger BP and a high pressure compressor HP whose compressors work in series with, preferably, air refrigeration between the compressors, the exhaust flow section Sd which can be adjusted between a minimum Sd min and a maximum Sd max by one or a combination of the following means:

- adjustment of the variable section of the gas distributor of the turbines,

- opening of a bypass between the inlet and the outlet of the turbines,

- transition from a series configuration to a parallel configuration of the turbines,

- the adaptation regime of the turbocharging Na thus becoming adjustable, continuously or discontinuously, between two values Na min and Na max. In the following, a bypass between the inlet and the outlet of a turbine will be called waste cake.

12. Engine according to claim 11 characterized in that the minimum flow section Sd min offered to the gases is constituted by the two turbines mounted in series at maximum closing if their distributor is variable and all waste g te closed if there is one.

13. Engine according to claim 12 characterized in that it operates on a 4-stroke cycle and that the timing of the valves is fixed.

14. Motor according to claim 13 characterized in that the maximum flow section Sd max offered to the gases is constituted by two turbines with fixed distributors mounted in parallel and that to pass the turbines from the series configuration to the parallel configuration, the following maneuvers:

- gradual partial opening of the HP trash waste, - progressive and simultaneous partial opening of the HP trash waste and

BP

- simultaneously and quickly: total opening of the HP trash waste, total closing of the LP trash waste, communication between the output of the HP turbine and the output of the LP turbine.

15. Motor according to claim 13 characterized in that the maximum flow section Sd max offered to the gases is constituted by a BP turbine with fixed distributor and an HP turbine with variable distributor mounted in parallel, the HP distributor being in full opening and that to switch the turbines from the serial configuration to the parallel configuration, the following maneuvers are carried out successively:

- gradual opening of the HP turbine distributor,

- gradual partial opening of the BP trash waste,

- simultaneously and quickly: total closure of the BP waste tank and communication between the HP turbine outlet and the BP turbine outlet.

16. A method of powering an engine according to one of claims 2, 3 or 11 characterized in that to limit the frequency of configuration changes, the geometry is immobilized for a type of conduct which implements a limited range of power, for example configuration set for city driving and setup parallel to highway driving, the power thresholds for each configuration that can be crossed for maneuvers short periods, such acceleration, overtaking, speed spikes, etc. the crossing of the thresholds which can be done as follows:

- by closing the EGR valve if the pressure in the exhaust manifold can be increased,

- by opening one or two waste spoils if the exhaust temperature can be increased,

- by closing the inlet valve if the maximum cycle pressure is reached or if the compressors are close to their maximum flow.

17. Engine according to claim 14 characterized in that the waste tank BP has a second seat for simultaneously closing the bypass entered PB turbine outlet and the communication HP turbine outlet BP turbine outlet.

18. Engine according to claim 14 characterized in that the two waste spoils are concentric and have stops so that their simultaneous movements are actuated by one of them and communicated to the other by said stops.

19 Motor according to claim 13 characterized in that the maximum flow section Sd max is constituted by two turbines with variable distributors in full opening mounted in series, the distributors being opened simultaneously to maintain the intake pressure at its maximum set value on the full load curve.

20. Engine according to claim 12 characterized in that the timing of the valves can be controlled to move the closing of the cylinder between the vicinity of the PMB and the mid-stroke of the piston, the maximum flow section Sd being constituted by the HP turbine in distributor series configuration in full opening if it is variable, waste spoils HP in full opening otherwise. The turbines are dimensioned to allow the compressors to simultaneously reach their maximum pressure ratios.

21. A method of supplying an engine according to claim 20 characterized in that the full load curve as a function of the speed is carried out as follows:

- from Nmin to 2 Nmin FA goes from PMB to around 90 dv after PMB so as to maintain the cycle pressure at its set value, - the distributor or the HP waste tank are closed;

- from 2 Nmin to approximately 3 Nmin the HP distributor or the HP waste tank are open and possibly the BP waste waste, to maintain the inlet pressure at its maximum set value,

FA is maintained at 90 dv after PMB, - from 3 Nmin to Nmax the overall fuel flow rate is kept constant to keep the intake pressure at its limit value,

- at partial load the FA setting will be controlled according to a map memorized by the engine control computer.

22. Engine according to claim 12 characterized in that it operates on the 2-stroke cycle,

- the intake ports are closed by valves,

- the exhaust ports are closed by valves and communicate with a single exhaust manifold, - the external recycling phase precedes the scan, - the valve timing can be controlled for moving the closing of the cylinder between the vicinity of PMB and mid stroke of the piston,

- the maximum flow section Sd is constituted by the HP turbine in the distributor series configuration in full opening if it is variable, waste spoils HP in full opening in the opposite case,

- the turbines are dimensioned to allow the compressors to simultaneously reach their maximum pressure ratios,

The EGR valve is replaced by a non-return valve or a closable aerodynamic diode.

23. A method of feeding an engine according to claim 22 characterized in that the full load curve as a function of the speed is carried out as follows:

- from Nmin to 2 Nmin, the closing of the cylinder changes from PMB to approximately 90 dv after PMB so as to maintain the cycle pressure at its set value,

- the distributor or the HP waste trash are closed,

- from 2 Nmin to around 3 Nmin the HP distributor or the HP waste tank are open and possibly the BP waste waste, to maintain the inlet pressure at its maximum set value,

FA is maintained at 90 dv after PMB,

- from 3 Nmin to Nmax the overall fuel flow rate is kept constant to keep the intake pressure at its limit value,

To maximize the cooled external EGR, the depolluted partial loads are executed as follows:

- the cylinder remains closed in the vicinity of the PMB and the turbines remain in the closed configuration until the limit P2 for this setting,

- the turbines are then opened to maintain P2 at its limit value, - the aerodynamic diode when the external recirculation rate is canceled.

24. Engine according to claim 12 characterized in that it operates on the 2-stroke cycle, that it has two exhaust ports per cylinder, closed by valves, which communicate respectively with a connected exhaust manifold to the turbine and an exhaust manifold connected to the EGR duct and / or to the turbine via a piloted distribution valve. 0 The timing of the valve assigned to the EGR can be controlled to move the cylinder closure between the vicinity of the PMB and the mid-stroke of the piston.

The external recycling phase precedes the sweep when the cylinder closes in the vicinity of the PMB and follows it when the cylinder closes at mid-stroke of the piston.

The maximum Sd flow section is constituted by the HP turbine in distributor series configuration in full opening if it is variable, waste spoils HP in full opening in the opposite case.

The turbines are dimensioned to allow the compressors to simultaneously reach their maximum pressure ratios.

25. A method of supplying an engine according to claim 24, characterized in that the pressure P2 is less than the limit authorized for this setting,

- the distributor valve is in the recycling position,

- the cylinder is closed in the vicinity of the PMB,

- the distributor or the HP waste trash are closed, The pressure reaches the authorized limit value for this setting: the cylinder closure is moved at the mid stroke of the piston to substantially double the authorized limit P2,

The dispensing valve remains in the recycling position, The dispenser or the HP waste container remains closed.

The pressure P2 reaches the new authorized limit for this new setting: the distributing valve blocks recycling, the distributor or the waste treats HP open to maintain the P2 at its new authorized limit, the transition can be made gradually in both directions or quickly with hysteresis.

26. Method according to one of claims 6, 7, 8, 9, 16, 21, 23, 25 characterized in that at full load the variable geometry is controlled to maintain a parameter at its setpoint limit value; at partial load, the variable geometry is controlled to optimize pollution control and / or performance according to a map stored in the engine control computer.

27. Engine according to claim 1 comprising a flat yoke carrier valves whose chamber side faces are coplanar with the cylinder head and substantially tangent to the cylinder characterized in that the pipes or the inlet end with an oblong nozzle defined by a half upper cylinder is pressing the upper edge of the conical seat and tangent to the latter along its generatrix situated in a plane substantially perpendicular to the plane passing through the axis of the seat and through the axis of the cylinder and by a cylinder lower covering half of the head valve opposite to said generatrix. The nozzles will also be directed to create a tangential speed in the same direction. The angle of the seats are chosen to optimize the stratification of the combustive load.

28. Engine according to claim 1 comprising a plane cylinder head carrying valves whose faces on the chamber side are coplanar with the cylinder head and substantially tangent to the cylinder, characterized in that the

5 conical sealing surface of the intake valves is extended towards the piston by a cylindrical portion of height slightly greater than the lifting of said valves, that the conical seats of said valves are arranged at the bottom of cylindrical housings arranged in the cylinder head for receive said cylindrical parts of said valves so that the 0 flat underside of the valves are in the plane of the cylinder head when they rest on their seats, the clearance between the housings and the valves being minimal, that recesses are made in the cylinder head which do not exceed the following borders:

two cylindrical portions concentric with the bore and tangent 5 externally and internally to the cylindrical recess of each valve,

- a conical surface extending the half seat of the valve delimited by a plane passing through its axis and the axis of the cylinder,

The recesses will also be oriented to create a tangential speed o in the same direction.

The seat angle is chosen to optimize the stratification of the oxidizing charge.

29. Engine according to one of claims 27 and 28 characterized in that it comprises two diametrically opposed intake valves.

30. Engine according to claim 1 characterized in that a fraction of the recycled gases is retained in the cylinder at closing of the latter. The fresh gases are introduced by directional inlet ducts 0 in order to arrange a temperature stratification and concentrations in the neutral top combustion chamber. Fuel is sprayed into the fresh gas.

31. Engine according to claim 30 characterized in that the fuel is introduced into the clean air between the compressor and the external EGR mixer.

32. Engine according to claim 30 characterized in that the fuel is introduced into the mixture between clean air and the external EGR.

33. Engine according to claim 30 characterized in that the ignition point is controlled by the timing of the valves when the cylinder is closed.

34. Engine according to claim 30 characterized in that the ignition point is controlled by the temperature of the external EGR.

35. Engine according to claim 30 characterized in that the first ignition is electrically controlled or is spontaneously triggered by the injection of high pressure fuel at top dead center.

36. Engine according to claim 30 characterized in that the gas working chamber has a geometry of revolution around the axis of the cylinder. The stratification has a geometry of revolution around the axis of the cylinder created by the orientation of the intake orifices. The temperature of the oxidizing charge increases between the periphery and the axis so that the self-ignition propagates from the center to the periphery.

37. Motor according to claim 36, characterized in that the meridian profile of the combustion chamber is chosen to optimize the rate of release of energy by the progressive increase of the reactive load isothermal surfaces.