SE1250994A1 - Super turbo compressor which has a fixed high speed gearbox and a continuously variable transmission - Google Patents

Abstract

SUPER TURBO COMPRESSOR WHICH HAS A FIXED HIGH SPEED SWITCH AND ONE. CONTINUOUS VARIABLE TRANSMISSION. A super-turbocharger using a fixed high speed gearbox connected to a continuously variable transmission to allow high speed operation is provided. A fixed high-speed gearbox is used to provide speed reduction from the high-speed turbine shaft. A second gear provides infinitely variable speed ratios with a continuously variable transmission. Gas recirculation and the super-turbocharger are also published. Figure 1

Description

WO 2011/096936 PCTHJS 2010/23398 starts to provide more energy than is required to drive the compressor, then the super-turbocharger recovers the excess energy by applying the extra power to the piston engine, usually via the crankshaft. The result is that the super-turbocharger provides the benefits of both low speed and high torque and the added value of high speed with many horsepower, all from one system.

SUMMARY OF THE INVENTION

An embodiment of the present invention may therefore comprise a super-turbocharger connected to an engine, comprising: a turbine which generates the mechanical energy of the turbine rotation generated by the turbine from enthalpy in the exhaust gases produced by the engine; a compressor that compresses incoming air and provides compressed air to the engine in response to the mechanical energy of the turbine rotation generated by the turbine and the mechanical energy of the engine rotation transmitted from the engine; a shaft having end portions connected to the turbine and compressor, and a central portion having a surface driving against the shaft; a friction drive unit located around the central part of the shaft, the friction drive unit comprising: a plurality of planetary gears having a plurality of planetary gear drive surfaces connected to the drive surface of the shaft so that a first number of drive couplings are located between a plurality of planetary gear drive surfaces and the shaft driving surface; a ring roller rotated by a plurality of planetary gears through a second plurality of driving connections; a continuously variable transmission which is mechanically coupled to the friction drive unit and to the engine, which transmits the mechanical energy of the turbine rotation to the engine and the mechanical energy of the engine rotation to the super-turbocharger at the operating speed of the engine.

An embodiment of the present invention may further comprise a method of transferring mechanical rotational energy between the super-turbocharger and an engine, comprising: generating the mechanical energy of the turbine rotation from enthalpy in the exhaust gases produced by the engine; compressing the air intake air to provide compressed air to the engine in response to the mechanical energy of the turbine rotation generated by the turbine and the mechanical energy of the engine rotation generated by the engine; providing a shaft having end portions connected to the turbine and compressor, and a central portion having a drive surface for the shaft; WO 2011/096936 PCTHJS 2010/23398 mechanical connection of the friction drive unit to the drive surface of the shaft; placing a plurality of planetary gear drive surfaces in contact with the shaft drive surface so that a plurality of planetary roller drive surfaces so that a plurality of first drive connections are created between a plurality of planetary rollers so that a plurality of other drive connections are created between a plurality of planetary rollers and the ring roller; mechanical coupling of a continuously variable gearbox to the friction drive unit and to the engine to transmit the mechanical energy of the turbine rotation to the engine and the mechanical energy of the engine rotation to the super-turbocharger at the operating speeds of the engine.

An embodiment of the present invention may further comprise a method of facilitating exhaust gas circulation in an explosion engine with a super-turbocharger comprising: providing a high pressure exhaust outlet with a different predetermined size in the explosion engine; providing a low pressure exhaust outlet with a second predetermined size in the explosion engine, this second predetermined size being substantially larger than the first predetermined size; providing a super-turbocharger with at least a first proportion of high pressure exhaust gases from the high pressure exhaust outlet; providing at least a second proportion of high pressure exhaust gases from the high pressure exhaust outlet to an inlet manifold on the explosion engine; operating a low-pressure super-turbocharger from the low-pressure exhaust outlet; providing compressed air from an outlet of the low pressure compressor to the air inlet of the high pressure compressor, at a predefined pressure, to the inlet manifold of the explosion engine; to open the high-pressure outlet of the exhaust gases when the pressure therein is higher than the predetermined pressure so that the other part of the exhaust gases under high pressure recirculates through the explosion engine.

An embodiment of the present invention may further comprise a method of facilitating exhaust gases to be recirculated in an explosion engine with super-turbocharger, comprising: providing a high pressure exhaust outlet having a first predetermined size in the explosion engine; providing an exhaust outlet of a second predetermined size in an explosion engine, this second predetermined size substantially larger than the first predetermined size; operating a super-turbocharger with high-pressure exhaust gases from the high-pressure exhaust outlet; operating a low pressure super-turbocharger with lower pressure exhaust from the low pressure exhaust outlet; providing compressed air from the low pressure compressor to an air intake for the high pressure compressor; WO 2011/096936 PCTHJS 2010/23398 providing compressed air outlet of the high pressure compressor, at a predetermined pressure, to an inlet manifold for the explosion engine; conducting high-pressure exhaust gases from an outlet on the high-pressure super-turbocharger to an intake manifold for the explosion engine; to open the high-pressure exhaust outlet when the pressure in the high-pressure exhaust outlet is greater than the predetermined pressure so that the high-pressure exhaust gases from the outlet of the super-turbocharger recirculate through the explosion engine.

An embodiment of the present invention may further comprise a method for facilitating exhaust gas recirculation in a super-turbocharger explosion engine, comprising: providing a high pressure exhaust outlet having a first predetermined size; providing a low pressure exhaust outlet having a second predetermined size in the explosion engine, the second predetermined size being substantially larger than the first predetermined size, providing high pressure exhaust gases from a high pressure exhaust outlet to an inlet manifold; for the explosion engine; operating a low-pressure super-turbocharger with lower-pressure exhaust gases from the exhaust outlet of a low-pressure exhaust outlet; the supply of compressed air from an outlet of the low pressure compressor, at a predetermined pressure, to an inlet manifold for the explosion engine; to open the high pressure drain outlet when the pressure in the high pressure drain outlet is greater than the predetermined pressure so that the other part of the high pressure exhaust gases recirculates through the explosion engine.

BRIEF DESCRIPTION OF THE DRAWINGS

Figure 1 is a side illustration of an embodiment of the super-turbocharger.

Figure 2 is a transparent isometric view of the embodiment of the super-turbocharger in Figure 1.

Figure 3A is a transparent side view of an embodiment of the super-turbocharger illustrated in Figures 1 and 2.

Figure 3B is a view with parts cut away of another embodiment of the super-turbocharger.

WO 2011/096936 PCTHJS 2010/23398

Figure 3C is a transparent side view of modifications of the design of the super-turbocharger illustrated in Figures 1, 2 and 3A.

Figures 4-9 are various drawings of a super-turbocharger using an embodiment with a multi-diameter planetary gear driver.

Figure 10 is an illustration of another embodiment of a high speed friction drive unit.

Figures 11 and 12 are illustrations of an embodiment of a driving continuously variable gearbox.

Figure 13 is a view with parts cut away of another embodiment.

Figure 14A is a schematic view of another embodiment of a gas recirculation device with super-turbocharger.

Figure 14B is a schematic view of another embodiment of a gas recirculation device with super-turbocharger.

Figure 14C is a schematic view of another embodiment of a gas recirculation device with super-turbocharger.

Figure 14D is a graph of valve lift, fate rate and cylinder pressure shown versus piston position for the embodiments of Figures 14A-C

Figure 14E is a pressure / volume diagram showing cylinder pressure versus Cylinder volume for the embodiments of Figures 14A-C

Figure 15 is a graphical illustration of the simulated improvement of BSFC (Brake Specific Fuel Consumption).

DETAILED DESCRIPTION OF THE INVENTION

Figure 1 is a schematic illustration of an embodiment of the super-turbocharger 100 using a high speed friction drive 114 and a continuously variable transmission 116.

As shown in Figure 1, the super-turbocharger 100 is connected to the engine 101. The super-turbocharger includes a turbine 102 which is connected to the engine 101 via an exhaust conductor 104. The turbine 102 receives the hot exhaust gases from the exhaust conductor 104 and generates mechanical rotational energy before blowing the exhaust gases to an exhaust outlet 112. A catalytic filter for particles from diesel engines (not shown) can be connected between the exhaust conductor 104 and the turbine 102. Alternatively, the catalytic filter for particles from diesel engines (not shown) can be connected with the exhaust outlet 112. The mechanical rotational energy generated by the turbine 102 is transmitted to the compressor 106 via a turbine / compressor shaft L such as shaft 414 in Figure 1, to rotate a compressor kt properly located in the compressor 106, which compresses the air intake 110 and transmits the compressed air to a conductor 108, which is connected to an inlet manifold (not shown) for motor l0l. As shown in the above application referred to herein, super-turbochargers, in contrast to turbochargers, are connected to a driveline to transfer energy to and from the driveline. The driveline referred to in this document may include the engine 101, the gearbox of a vehicle in which the engine 101 is located, or other applications of the mechanical rotational energy generated by the engine 101. In other words, mechanical rotational energy may be connected to or transmitted from the super-turbocharger to the engine via at least one intermediate mechanical device such as a gearbox or driveline in the vehicle, and vice versa. In the embodiment of Figure 1, the mechanical rotational energy of the super-turbocharger is coupled directly to the crankshaft 122 in the engine 101 via a shaft 118, a pulley 120 and a drive belt 124. As also illustrated in Figure 1, a high speed friction drive 114 is mechanically connected to the continuously variable gearbox 116.

In operation, the high speed friction drive unit 114, of Figure 1, is a high speed friction drive unit with a fixed relationship mechanically connected to the turbine / compressor shaft using a traction clutch to transmit mechanical rotational energy to and from the turbine. / compressor shaft. The high speed friction drive 114 has a fixed gear ratio which may be different depending on the size of the motor 101. For smaller engines, a large gear ratio is required for the high-speed friction drive 114.

For smaller engines, the compressor and turbine of a super-turbocharger must necessarily be smaller to maintain a small engine size and to meet the needs of the compressor and turbine. For the smaller turbine and the smaller compressor to work well, they must rotate at a higher speed. For example, WO 2011/096936 PCTHJS 2010/23398 smaller motors may require the compressor to rotate at a speed of 300,000 RPM. For very small engines, such as with an engine volume of 500 cc, the super-turbocharger may need to rotate at a speed of 900,000 RPM. One of the reasons that smaller engines require compressors that operate at a higher speed is to avoid pressure surges ("surge"). In addition, in order to operate efficiently, the speed of the compressor fan tips must be just below the speed of sound. Because the spikes are not as long in smaller compressors, the tips do not move as fast as the tips of larger compressors at the same speed. As the size of the compressor decreases, the speed required to operate efficiently increases exponentially. Since gears are limited to approximately 100,000 RPM, normal gear systems cannot be used to achieve the power consumption at the higher speeds necessary for a super-turbocharger for a car engine. Therefore, various embodiments using a high speed friction drive 114 are used to provide and receive power from the turbine shaft.

The mechanical rotational energy from a high-speed friction drive unit 114 is therefore reduced to a speed which is variable depending on the rotational speed of the turbine / compressor, but at a speed which is within the operating range of the Continuously Variable Transmission (CVT). 116. For example, the high speed friction drive 114 may have an output speed that varies between zero and 7000 RPM, while the input speed from the turbine / compressor shaft may vary from zero to 300,000 RPM or higher. The continuously variable gearbox 116 adjusts the speed of the high speed friction drive unit 114 to the speed of the crankshaft 122 and the pulley 120 to apply mechanical rotational energy to motor 110, or extract mechanical rotational energy from motor 101 at the correct speed. In other words, the continuously variable gearbox 116 includes an interface for transmitting mechanical rotational energy between the engine 101 and the high speed friction drive 114 at the correct speed, which varies depending on the engine speed and the turbine / compressor speed. The continuously variable gearbox 116 may comprise any desired type of continuously variable gearbox which can operate at the required speeds and has a gear ratio corresponding to the speed of the crankshaft 122 or other mechanisms connected, directly or indirectly, to the engine 101. To For example, in addition to the embodiments shown here, CVTs with two rollers can, as well as traction ball friction drives and pressing steel belt CVTs.

WO 2011/096936 PCTHJS 2010/23398

As an example of a continuously variable gearbox suitable for use as a continuously variable gearbox 116, shown in Figure 1, the continuously variable gearbox is shown in Figures 11 and 12. Other examples of continuously variable gearboxes which can be used as the continuously variable gearbox 116 in Figure 1 includes US Patent Serial No. 7,540,881 issued June 2, 2009, to Miller et al. Miller's patent is an example of a friction driven, continuously variable gearbox that uses a planetary ball bearing.

The Miller friction drive is limited to about 10,000 RPM so Miller's continuously variable gearbox is not useful as a high speed friction drive, such as the high speed friction drive 114. However, the Miller patent discloses a continuously variable gearbox using a friction drive and is suitable as an example of a continuously variable gearbox that could be used as a continuously variable gearbox 116 as illustrated in Figures 1-3. Another example of a continuously variable gearbox is discussed in U.S. Pat. Patent Serial No. 7,055,507, issued June 6, 2005 to William R. Kelley, Jr., and given to Borg Warner. Another example of a continuously variable gearbox is published in U.S. Pat. Patent No. 5,033,269 issued July 23, 1991 to Smith. In addition, U.S. Patent No. 7,491,149 also discloses a continuously variable gearbox which would be suitable for use as a continuously variable gearbox 116. U.S. Pat. Patent No. No. 7,491,149 issued February 17, 2009 to Greenwood et al., And given to Torotrak Limited, discloses an example of a continuously variable transmission using a friction drive unit which can be used as the continuously variable transmission 116. All of these patents are specifically incorporated by reference. for everything they publish and teach. European application No. 92830258.7, published 1995-08-09 as publication No. B1 also illustrates another continuously variable gearbox 3 which is suitable for use as the continuously variable gearbox 116.

Various types of high speed friction drives can be used as the high speed friction driver 114. For example, the high speed planetary friction drive 406 shown in Figures 4-9 and the high speed planetary friction drive in Figure 10 can be used as the high speed friction drive 114.

Examples of high speed friction drives using gears are published in U.S. Pat. Patent No. 2,397.94l issued April 4, 1946 to Birgkigt and U.S. Pat. Patent No. 5,729,978 WO 2011/096936 PCTHJS 2010/23398 issued March 24, 1998 to Hiereth et al. Both of these patents are specifically incorporated by reference for everything they publish and teach. Both of these references use standard gears and do not use friction drives. Thus, with carefully polished, specially designed gear systems, the gears in these systems are limited to speeds of about 100,000 RPM or less. U.S. Patent No. 6,960,147 issued 2005-11-01 to Kolstrup and given to Rulounds Roadtracks Rotrex A / S publishes a planetary gear that can give gears of 13: 1. Kolstrup ° s planetary gear is an example of a friction drive (114) with high speed in Figure 1. US

Patent No. 6,960,14 is also specifically incorporated by reference for all that it publishes and teaches.

Figure 2 is a schematic transparent side view of the super-turbocharger 100. As shown in Figure 2, turbine 102 has an exhaust exhaust line 104 which receives exhaust gases applied to turbine shaft 130. Compressor 106 has a compressed air line 108 which supplies the constituent manifold with air. The compressor housing 128 encloses the compressor shaft 126 and is connected to the compressed air line 108. As shown above, the high speed friction drive 114 is a fixed gear friction drive connected to a continuously variable gearbox 116. The continuously variable gearbox 116 drives a shaft 118 and a pulley 120.

Figure 3A is a schematic transparent side view of the embodiment of the super-turbocharger 100 illustrated in Figures 1 and 2. Again, as shown in Figure 3A, turbine 102 includes a turbine fl genuine 130, while compressor 106 includes a compressor fan 126. A shaft (not shown) which connects the turbine shaft 130 and the compressor shaft 126 are connected to the high speed friction drive unit 114. Mechanical rotational energy is transmitted from the high speed friction drive 114 to a transmission gear 132 which transmits the mechanical rotational energy to a CVT gear and to the continuously variable gearbox (CVT) 116. The continuously variable gearbox 116 is connected to a shaft 120 and a pulley 120.

Figure 3B is a schematic cut-away view of another example of a super-turbocharger 300 connected to a motor 304. As shown in Figure 3B, the turbine 302 and the compressor 306 are mechanically connected via a shaft 320. The friction drive unit 308 for high speed transmits mechanical rotational energy to, and receives mechanical rotational energy from, WO 2011/096936 PCTHJS 2010/23398 the transmission gear 322. A specific example of a friction drive 308 for is illustrated in Figure 3B. Transmission gears 322 transfer mechanical rotational energy between the friction drive unit 308 and the continuously variable gearbox 310. A specific example of a continuously variable gearbox 310 is also illustrated in Figure 3B. Shaft 312, pulley 314 and pulley 316 transfer mechanical rotational energy mechanical rotational energy between crankshaft 318 and the continuously variable gearbox 310.

Figure 3C is a schematic cut-away view of modifications to the design of the super-turbocharger 100 illustrated in Figures 1, 2 and 3A. As shown in Figure 3C, turbine 102 and compressor 106 are connected to each other via a shaft (not shown). The high speed friction drive 114 is connected to the shaft. Mechanical rotational energy is transferred from the high speed friction drive 114 to a transmission gear 132 which transmits the mechanical rotational energy to the gearbox 134. The high speed friction drive 114, the transmission gear 132 and the gearbox gear 134 may all be located in the same housing. Transmission gears 134 are connected to a gearbox 140 which may include a manual gearbox, a CVT, a straight shaft, an automatic gearbox or a hydraulic transmission.

The gearbox 140 is then connected to shaft 118 which is connected to a pulley 120.

Pulley 120 is connected to the driveline. In an alternative embodiment, the pulley 120 is connected to an electric motor / generator 142.

Figure 4 is a schematic transparent view of another embodiment of super-turbocharger 400 using a high speed friction drive 416 connected to a continuously variable gearbox 408. As shown in Figure 4, the turbine 404 is mechanically coupled to the compressor 402 with a compressor / turbine shaft 414. Mechanical rotational energy is transferred between the compressor / turbine shaft 414 and the friction drive unit 416 with fl your diameters in the manner shown in more detail below. Transmission gear 418 transfers mechanical rotational energy between the friction drive unit 416 with its diameters and the CVT gear 420 in the continuously variable gearbox 408. The shaft 410 and the pulley 412 are coupled to the continuously variable gearbox 408 and transmit power 8 between the variable gearbox and the variable gearbox.

Figure 5 is a cut-away side view of the more diameter friction drive unit 416 connected to the transmission gear 418, which in turn is connected to the CVT gear WO 2011/096936 PCTHJS 2010/23398 420. The compressor / turbine shaft 414 has a polished, hardened surface on a central part, as shown in more detail below, which acts as a solar drive in the friction drive unit 416 with fl your diameters.

Figure 6 is an exploded view 600 of the embodiment of the super-turbocharger 400 illustrated in Figure 4 as shown in Figure 4. As shown in Figure 6, the turbine housing 602 includes a turbine fl genuine 604. The cover plate 606 on the hot side is located adjacent turbine shaft 604 and head restraint 608 for housing. A ring gasket 610 seals the exhaust gases on the hot side, cover plate 606.

Ring roller bearing 612 is mounted in the ring roller 614. The compressor / turbine shaft 414 passes through the headrest 608 for the housing. The hot side cover plate 606 is connected to the turbine shaft 604.

Planetary ball bearings 618 are mounted on the planet carrier 620. Oil supply tubes 624 are used to supply traction fluid to the traction surface. The planet carrier 626 is mounted on the planet carrier 620 and uses planet bearing ball bearings 628. Fixed ring 630 is then mounted on the outside of the planet carrier 626. Cage 632 is mounted between fixed ring 630 and the cool side cover plate 636. Compressor fl genuine 638 is coupled to the compressor / turbine shaft 414. The compressor housing 640 includes the compressor fl genuine 638. The housing headrest 608 also supports the continuously variable gearbox and the transmission gear 418. Various bearings 646 are used to mount the transmission gear 418 and the housing headrest 608. The continuously variable gearbox includes a cover plate 642 and a CVT bearing plate 644. The CVT cover plate 654 covers the various parts of the CVT device. Shaft 410 is connected to the continuously variable gearbox.

Pulley 412 is mounted on shaft 410 and transmits mechanical rotational energy between shaft 410 and a driveline.

Figure 7 is a perspective view of individual key components of the friction drive unit 416 with fl your diameters, and also of the turbine fl shaft 604 and the compressor fan 638. As shown in Figure 7, the compressor / turbine shaft 414 is connected to the turbine fl shaft 604 and the compressor, 6 through the friction drive unit 416 with several diameters, center.

The friction drive unit 416 with fl your diameters includes planetary rollers 664, 666 with fl your diameters (Figure 9), 668. These planetary rollers with fl your diameters are rotationally coupled to a planet carrier 626 (Figure 9). The balls 656, 658, 660, 662 rest on a sloping plane for ball ramps on the fixed ring 630. The ring roller 614 is driven by an inner diameter of the planetary rollers 664, 666, 668 as shown in detail below. 11 WO 2011/096936 PCTHJS 2010/23398

Figure 8 is a cut-away side view of the friction drive unit with fl your diameters 416. As shown in Figure 8, the compressor / turbine shaft is hardened and polished to form a traction surface used as a solar roller 674 having a traction interface 676 to the planetary roller 664. .

The planetary roller 664 with fl your diameters rotates about the axis 672 of the planetary roller with fl your diameters. The planetary roller 664 with its diameters touches the fixed ring 630 at the interface 690 of the planetary roller 664 and the fixed ring 630. The planetary roller 664 with its diameters touches the ring roller 614 at the interface 691, which is a different radial distance from the planetary roller shaft 672 with its diameters 691. Figure 8 also illustrates the planet carrier 626 and the ball ramp 631 overlapping with the ball 660. The balls 656, 658, 660, 662 are wedged between a housing (not shown) and the ball ramp, such as the ball ramp 630 on the fixed ring 664. When torque is applied to the ring roller 614 this causes the fixed ring 664 to move slightly in the direction of rotation of the ring roller 614. This causes the balls to surface along the various ball ramps, such as the ball ramps 630, 631, which in turn causes the fixed ring 630 to press against the planetary rollers 664. , 666, 668 with a number of diameters. Since interface 691 for planetary roller 664 and fixed ring 630 is inclined, and the interface between planetary roller 664 and ring roller 690 is inclined, an inward force is generated on planetary roller 664 with fl your diameters, which creates a force on tensile interface 676 which increases the tensile force at tensile interface 676 between planetary roller 664 with fl your diameters and the solar roller 674.

In addition, a force is created at interface 691 for the multi-diameter planetary roller 664 and ring roller 614, which increases the traction at interface 691. As also shown in Figure 8, both the compressor shaft 638 and the turbine shaft 604 are connected to the compressor / turbine shaft 414.

Ring roller 614 is connected to the transmission gear 418, as also shown in Figure 8.

Figure 9 is a cut-away side view of the friction drive unit 416 with fl your diameters. As shown in Figure 9, the solar roller 674 rotates clockwise, as shown by the direction of rotation sensor 686.

The planetary rollers with fl your diameters 664, 666, 668 have roller surfaces with an outer diameter, such as outer diameter roller surface 688 for the planetary roller 664 with fl your diameters. These outer diameter roller surfaces touch the solar roller 674 which causes the planetary rollers 664, 666, 668 to rotate counterclockwise, such as the direction of rotation 684 of the planetary roller 666 with fl your diameters. The planetary rollers 664, 666, 668 also have an inner diameter rolling surface, such as the inner diameter rolling surface 680 of planetary roller 664 having fl diameters. The inner diameter rolling surface of each planetary roller with its diameters touches the rolling surface 687 of the ring roller 614. Accordingly, the interfaces 678 of the planetary roller 664 with the rolling surface 687 of the ring roller 614 form a tensile interface which transmits mechanical rotational energy when a 12 WO 2011/096936 PCTHJS 2010/ The interface between each of the planetary rollers 664, 666, 668 with their diameters and the solar roller 674 also forms a tensile interface which transmits mechanical rotational energy when a tensile fluid is applied.

As indicated above with respect to Figures 8 and 9, the fixed ring 630 generates a force which presses the planetary rollers 664, 666, 668 with fl your diameters against the solar roller 674 to generate traction. Each of the planetary rollers 664, 666, 668 is rotationally coupled to the planetary carrier 626 with planetary roller shafts, such as the planetary roller shaft 672 for the planetary roller 664 having its diameters.

These shafts have a slight play so that the planetary rollers 664, 666, 668 can move slightly and create a force between the solar roller 674 and the outer diameter roller surface of the planetary rollers 664, 666, 668, so that the outer diameter of the roller surface 688 of the planetary roller 664. The movement of the planetary roller 664 towards the solar roller 674 also increases the tensile force at the interface of the planetary rollers 664, 666, 668 with their diameters and the ring roller 614, since the interface between the planetary rollers 664, 666 and 668 and the ring roller 614, such as the interface 678, is inclined.

The contact between the planetary rollers 664, 666, 668 and the roller surface 687 of the ring roller 614 for the planet carrier 626 to rotate clockwise, such as the direction of rotation 682, illustrated in Figure 9. As a result, the ring roller 614 rotates counterclockwise, as the direction of rotation 687, and drives the transmission gear 418 clockwise. ,

Figure 10 is a schematic cross-sectional view of another embodiment of a high speed friction drive 1000. As shown in Figure 10, a shaft 1002, which is a shaft that connects a turbine to a compressor in a super-turbocharger, act as a solar roller in friction drive 1000 for high speed. Planet roller 1004 touches the shaft 1002 at traction interface 1036. Planet roller 1004 rotates about a shaft 1006 using the bearings 1008, 1010, 1012, 1014. As also shown in Figure 10, the gear 1016 is located and connected to the outer surface of the carrier 1018.

The carrier 1018 is connected to a housing (not shown) via the bearings 1032, 1034, which allows the carrier 1018 and the gear 1016 to rotate. The fixed rings 1020, 1022 include ball ramps 1028 and 1030, respectively. The ball ramps 1028, 1030 are similar to the ball ramps 630 illustrated in Figures 7 and 8. When the gear 1016 is separate, the balls 1024, 1026 move in the ball ramps 1028 and 1030, respectively, forcing the fixed rings 1020 , 1022 inwards towards each other. A force is generated between the fixed rings 1020, 1022 and the surface of the planetary roller 1004 at the traction surfaces 1038, 1040 when the balls 1024, 1026 force the fixed rings inwards towards each other. The force generated by the fixed rings 1020, 1022 also forces the planetary roller 1004 downward, as illustrated in Figure 10, so that a force is created between the shaft 1002 and the planetary roller 1004 at the traction interface 1036. As a result, greater traction is achieved. at the traction interface 1036 and the traction surfaces 1038, 1040.

Traction fluid is applied to these surfaces, which becomes adhesive and increases the friction at the traction surfaces as the traction fluid heats up as a result of the friction created at the traction interfaces 1036, 1038, 1040.

The high speed traction unit 1000, illustrated in Figure 10, can rotate at high speeds above 100,000 RPM, which cannot be achieved by gear systems. For example, the friction drive unit 1000 for high speed could be capable of rotating at speeds higher than 300,000 RPM. However, friction drive unit 1000 is limited to a gear ratio of about 1011 due to the physical limitations in size. The high speed friction drive unit 1000 can use three planetary rollers, such as planetary roller 1006, which are located radially around axis 1002. As illustrated in Figure 9, the size of the planetary rollers is limited with respect to the solar roller. If the diameter of the planetary rollers in Figure 9 increases, then the planetary rollers will collide. Thus, only gears of about 1:10 can be achieved with a planetary friction drive, as illustrated in Figure 10, whereas planetary friction drives connected to a planet carrier, as illustrated in Figures 7-9, can have gears up to 47: 1 or more. Consequently, if a compressor is needed for a smaller engine, and the compressor must rotate at a speed of 300,000 RPM to be efficient, then a friction drive with a gear ratio of 47: 1, as illustrated in Figures 7-9, can reduce the maximum speed 300 000 RPM to about 6 400 RPM.

Standard gearboxes with gears or continuously variable gears can then be used to transfer the mechanical rotational energy between the high-speed friction drive unit and the engine driveline.

As shown above, the high speed friction drive unit 1000, illustrated in Figure 10, may have a gear ratio up to 10: 1. Assuming that the speed of the shaft 1002 is 300,000 RPM for a super-turbocharger for a small engine, then the speed of 300,000 RPM can be reduced to 30,000 RPM at 1016 gears. Various types of continuously variable gearboxes 116 can be used, those which can operate up to 30,000 RPM using standard gear technologies. Continuously variable friction drives, such as the continuously variable friction drive illustrated in Figures 11 and 12, can also be used as the continuously variable gearbox 116, illustrated in Figure 1. In addition, gears up to 100: 1 may be achievable with the friction drive unit 416 with fl your diameters, illustrated in Figures 4-9.

Thus, small 0.5 liter engines, which may require a compressor operating at a speed of 900,000 RPM, can be reduced to 9,000 RPM, which is a speed that can be easily used by various gearboxes 116 with continuously variable to couple mechanically rotational energy between a driveline and a turbine-compressor shaft.

Figures 11 and 12 illustrate an example of a friction drive / transmission that is continuously variable and can be used as the continuously variable gearbox 116 in Figure 1.

The continuously variable friction drive unit illustrated in Figures 11 and 12 operates by moving the bearing rails 1116, 1118 in a lateral direction on bearing rail surfaces having a radius of curvature which causes the ball bearing contact points to move, which in turn causes the balls to rotate at a different spin angle to drive bearing rail 1122 at different speeds. In other words, the contact point of each of the bearings on the surfaces of the bearing rails changes as a result of the lateral displacement of the bearing rails 1116, 1118, which changes the speed at which the bearings rotate at the contact point, as explained in more detail below.

As shown in Figure 11, the input shaft 1102 is coupled to the transmission gear 132 (Figure 3A). For example, splines 1104 may be splined to the CVT gear 134, illustrated in Figure 3A. Therefore, the spline input 1104 at the input shaft 102 can be connected to the super-turbocharger via a high speed friction drive 114, as illustrated in Figure 3A. In this way, the input torque from the drive line is used to drive the spline input wheel 1104 at the input shaft 1102. The torque on the spline input 1104 provides a spin direction of rotation 1112 on both the input shaft 1102 and its associated structure including the input bearing rail 1114. spun around the axis of rotation 1106 in response to the torque applied by spline 1166 from the input shaft 1102 to the input bearing rail 1116. The rotation of the input shaft 1102, the input bearing rail 1114 and the input bearing rail 1116 produces a spin on a number of ball bearings 1132 because the stationary the bearing rail 1120 impedes the rotation of the ball bearings at the point of contact with the stationary bearing rail 1120. The input bearing rail 1114 and the input rail 116 rotate at the same angular velocity because they are interconnected via spline 1116. The input rail 1114 and the entrance rail 1116 cause the ball bearings 1132 to rotate vertically WO 2011/096936 PCTHJS 2010/23398 because the ball bearings 1132 are in contact with the stationary bearing rail 1120. The contact between the ball bearings 1132 with the stationary bearing rail 1120 also causes the ball bearings 1132 to precede around the circumference of the bearing rails 1114, 1 116, 11 18, 1 120. The rotation of the ball bearings 1132 as a result of being driven by the input bearing rail 114 and the input bearing rail 1116 creates a tangential contact between the ball bearings 1132 on the output bearing rail 1118. Depending on the point of contact between the ball bearings 1132 and the output bearing rail 1118, the gear ratio between the input bearing rails 11 114, 11 16 and output varies. The output bearing rail 118 is coupled to the output gear 1122. The output gear 1122 drives the output gear 1124, which in turn is coupled to the output shaft 1126.

The manner in which the continuously variable gear ratio 1100 of the friction drive unit, illustrated in Figure 11, changes the gear ratio between the input shaft 1102 and the output shaft 1126 is obtained by changing the relative position of the contact point between the four bearing rails 1114, 1116, 1118, 1120 in contact with the ball bearings 1132. The manner in which the contact points between the bearing rails 1114, 1116, 1118, 1120 and the ball bearings 1132 are changed is by shifting the position of the transfer cramp 1152. The surface cramp 1152 is moved horizontally, as illustrated in Figure 11, in response to electric actuator 1162. Electric actuator 1162 has a shaft that drives the telescopic shifter 1158 and rotates the telescopic shifter 1158.

Telescopic shifter 1158 has different thread types on the inside and outside. A difference in thread pitch of the different thread types causes the superficial cramp 1152 to move horizontally in response to rotation of the axis of the electric actuator 1162, which causes rotation of the telescopic shifter 1158. Lateral movement of the superficial cramp 1152, which is in contact with 1152. the bearing cramp 1164, causes lateral movement of the input bearing rail 1116 and the output bearing rail 1118. Lateral movement of the input bearing rail 1116 and the output bearing rail may vary, in the embodiment illustrated in Figure 11, by about 2-3 mm. For the displacement of the input bearing rail 1116 and the output bearing rail 1118, the contact angle between the ball bearings 1132 and the output bearing rail 1118 changes, which changes the gear ratio, or the speed at which the ball bearings 1132 fl move the bearing rails due to a change in contact angle between stationary bearing rail 1120 and input bearing.

The combination of change in angle between the bearing rails allows the contact speed, or contact point between the ball bearings 1132 and the output bearing rail 1118, to vary, resulting in a variation in speed between 0 percent of the rotational speed of the input shaft 1102 up 16 WO 2011/096936 PCTHJS 2010/23398 to 30 percent of the rotational speed of the input shaft 1102. The variation in speed of the output bearing rail 1102 provides a wide range of adjustable rotational speeds that can be achieved at the output shaft 1126.

To ensure proper holding of the ball bearings 1132 between the bearing rails 1114, 1116, 1118, 1120, the springs 1154, 1156 are provided. Spring 1154 creates a holding force between the input bearing rail 1114 and the fixed bearing rail 1120. Spring 1156 creates a holding force between the input bearing rail 1116 and the output bearing rail 1118. These retaining forces against the ball bearings 1132 are maintained over the entire fl removable area of the fl removable clamp 1152. The telescopic shifter 1158 has threads on the inner surface which are connected to the threads of the fixed threaded device 1160. The fixed threaded device 1160 attached to the frame 1172 and provides a fixed position relative to the frame 1172, so that the superficial cramp 1152 can move in the horizontal direction as a result of the difference between the threads on the two sides of the telescopic shifter 1158.

As also illustrated in Figure 11, all the rotating components of the friction drive unit with continuously variable gear 1100 rotate in the same direction, i.e. direction of rotation 1112 and output rotation 1128 of output gears 1122. Clamp nut 1168 holds spring 1156 in place and preloads spring 1156 to create the correct diagonal pressure between stationary bearing rail 1120 and input bearing rail 1114. When the fl surface cramp 1152 fl is flattened horizontally, as illustrated in Figure 11 , there is a slight movement of the input shaft 1102, based on the angles of the bearing rails 1114-1120 which are in contact with the ball bearings 1132.

The spline input gear 1104 allows lateral movement in the directions 1108, 1110 based on the points where the ball bearings 1132 are in contact with the bearing rails 1114-1120 and the specific contact angle of the bearing rails with respect to the ball bearings 1132. The frame 1170 is tightly bolted to the frame 1172 to hold spring. 1154, which creates the proper holding strength between the input bearing rail 1114 and the fixed bearing rail 1120. The ball bearings 1132 as illustrated in Figure 11 have progressive rotation 1131 in the four bearing rails 1114, 1116, 1118, 1120.

The direction of rotation 1112 of the shaft 1102 causes the gear 1122 to rotate in the direction of rotation 1128, as illustrated in Figure 11.

Figure 12 is a close view of the bearing rails 1114-1120 and ball 1132, illustrating the mode of operation of the traction unit with continuously variable gear 1100. As shown in Figure 17 WO 2011/096936 PCTHJS 2010/23398 12, the bearing rail 1,114 contacts with force ball 1 132 at contact point 1134. Bearing rail 11 16 contacts with force ball 1132 at contact point 1136. Bearing rail 1118 contacts with force ball 1132 at contact point 1138. Bearing rail 1120 contacts with force ball 1132 at contact point 1140. Each of the contact points 1134, 1136, 1138, 1140 are placed on a common great circle on the surface of the ball 1132. The great circle is located in a plane comprising the center of the ball 1132 and the central line ("axis") 1106 of the axis 1102. The ball 1132 spins around a virtual spinning axis 1142 which passes through the center of the ball 1132 and divides the great circle comprising the contact points 1134, 1136, 1138, 1140 The virtual spin axis 1142 through the ball 1132 is inclined at an angle 1146 to the virtual vertical axis 1144. The angle of inclination 1146 is the same for each of the balls located in the bearing rails around the circumference of the friction drive unit 1100. The angle of inclination 1146 establishes a mathematical relationship between a distance ratio and a peripheral speed ratio. The distance ratio is the ratio between the first distance 1148, which is the perpendicular distance between the spin axis 1142 and the contact point 1134, and a second distance 1150, which is the perpendicular distance between the spin axis 1142 and the contact point 1136. This distance ratio is equal to the ratio between the peripheral speeds. The ratio of the peripheral velocities is the ratio between the first peripheral velocity and the second peripheral velocity, where the first peripheral velocity is the difference between the peripheral velocity of the ball 1132 at the bearing rail 1114 and a common orbital peripheral velocity of the ball 1132 and the other balls of the bearing rails, while the second peripheral velocity is the difference the ball 1132 at the bearing rail 1116 and the common circumferential speed of the ball 1132 as well as the other balls placed in the bearing rails. The radius of curvature of each of the bearing rails 1114-1120 is greater than the radius of curvature of ball 1132. In addition, the radii of curvature of each of the bearing rails 1114-1120 need not be a constant radius of curvature, but may vary. In addition, the radii of curvature of each of the four bearing rails need not be the same.

When the bearing rails 1116, 1118 are moved simultaneously laterally, such as the direction of movement 1108, the speed ratio changes between the rotation of the shaft 1102 and the direction of rotation 1112 with respect to gears 1122 and direction of rotation 1128. For the movement of the bearing rails 1116, 1118 laterally in the direction 1108 the first distance 1148 becomes larger and the second distance 1150 becomes smaller. Therefore, the ratio between the distances, and 18 WO 2011/096936 PCTHJS 2010/23398 also changes the ratio between peripheral speeds, which changes the speed of gears 1122 compared to shaft 1022.

As stated above, the output of the continuously variable speed gearbox is in gear contact with the friction drive for speed reduction which is coupled to the shaft of the turbine compressor. As stated above, there are at least two or three different types of friction drives for speed reduction that can be used. The common type is a planetary type of friction drive for high speed reduction, as shown in Figures 6-9, and in Figure 10. If a large speed difference is desired between the turbine shaft and the planetary roller, then the embodiment in Figure 10 can use only two rollers instead of three, to obtain the desired gear ratio.

With three rollers, there is a limit of about 10: 1 in speed reduction and there may be a need for something more such as the 2011 gear to get a high speed like 250,000 RPM to anything below the 24,000 RPM as a 10: 1 gear would require. Therefore, a two-roll planetary friction drive can be used instead of a three-roll system in Figure 10, to achieve the speed reduction required for the smallest high-speed systems. Two rollers also provide less inertia, as each roller adds some amount of inertia to the system. For the slightest inertia, two rollers should be sufficient. The width of the draw roller is slightly larger than in a version with three rollers.

The multi-diameter planetary rollers rolling against the shaft are made of resilient material, e.g. either resilient steel or other material that allows some deformation of the roller within the outer drum. The application of a spring-loaded roller can provide the necessary pressure on the shaft, but does not limit the shaft's ability to find its ideal center of rotation.

When a turbocharger operates at extremely high speeds, it has balance limitations that cause the shaft to have to find its own center of rotation. The balance will be compensated by movements of the central axis. This movement can be compensated with spring-loaded rollers. The spring-loaded rollers can also be made very light to manufacture them from a thin steel strip which allows them to work against a shaft with very little inertia. The thickness of the belt must be large enough to apply sufficient pressure to the traction surface to provide a perpendicular force necessary to pull. A cam follower can be placed inside the roller, which will place each roller 19 WO 2011/096936 PCTHJS 2010/23398 and maintain this position within the system. Rollers need to work in a very rectilinear position between the outer drum and the turbine / compressor shaft, but the key to low inertia is low weight.

One or two cam followers can be used to hold the steel strip in place, so that the steel strip remains in its rectilinear position in the system.

The ring roller 614 is connected to a gear on the outer surface so that the ring roller can transmit the force into or out of the friction drive unit with fl your diameters 416. The ring roller 614 can be designed in many ways. Ring roller can simply be a solid steel article or other suitable material capable of transmitting the torque into and out of the multi-diameter friction drive unit 416. Ring roller 614 must be made of a material that allows ring roller 614 to be lightweight, but ring roller 614 must be made of a material that can be used as a friction drive surface on the roller surface 687. A correct roller surface 787 allows the planetary rollers 664, 666, 668 to transmit torque via traction on the surfaces.

In addition, the compressor shaft 414 must be kept in a very accurate position. The rectilinear position of the turbine / compressor shaft shaft 414 within the housing allows the tolerance to be maintained between the tips of the compressor blades and the compressor housing. A narrower tolerance increases the efficiency of the compressor. A more precise location reduces the likelihood of the turbine compressor fan 638 and the compressor housing 640. A method of controlling the linear force against the compressor wheel, a force arising from compressing the gas, is necessary to ensure that there is at least a minimum of clearance. This can be done by using an thrust bearing (not shown) which is oil fed or a thrust bearing which is a ball bearing type or roller bearing type.

Typically, a turbocharger, the bearings are, for reasons of reliability, plain bearings that have oil layers on both the inside and outside to allow the turbine shaft to center itself in its harmonic rotation. The balancing requirements for a turbocharger that is manufactured in large volume are reduced by using a bearing with two clearances. These types of bearings have been used because of the requirement for narrower clearance and more accurate fitting of the turbocharger shaft. A ball bearing is used both to hold the compressor and the turbine and to maintain better fit to the housing from a side-to-side movement perspective. This can be achieved with one or two ball bearings. Alignment of bearings within an outer area that is under pressure with oil allows the bearings to flow and allows the bearing to find a center. This does not affect the clearance between the outermost parts of the housing, the turbine and the compressor, but it allows the axial clearance to remain small.

Turbine shaft bearings provide a third limitation point for maintaining the alignment of the rollers.

Cam followers in the middle of the rollers can keep the rollers at 120 degrees apart. Two small cam followers can be used for each roller to eliminate backlash when the force changes direction.

In addition, a larger turbine can be used. The turbine wheel can be made larger than normal in diameter. It is possible to make the outer diameter of the turbine even larger than the compressor wheel without reaching the critical speed where the tips come close to the speed of sound, because the density of the exhaust gases is lower than the incoming air and therefore the speed of sound is higher.

This allows the exhaust gases to apply more torque to the turbine / compressor shaft without higher back pressure. Having more torque causes the turbine to regain more energy than is required to compress the incoming air. This produces more energy than can be extracted and sent to the engine. More energy from the same exhaust fate, which is not needed for compression, is transferred to the crankshaft and creates less fuel consumption.

In addition, turbine efficiency can be improved by using control vanes that control the incoming angle at which the exhaust gases reach the turbine wheel. This makes the peak efficiency higher, but makes the speed range smaller within which this efficiency is achieved. A narrow speed range is not good for a normal turbocharger, but is not a problem for a super-turbocharger where the controller can provide the necessary speed control.

Higher back pressure over the turbine compared to the pressure over the compressor can also create an unbalanced super-turbocharger. For a normal turbocharger, this pressure difference goes in the other direction. Having more back pressure causes the turbine to recover more energy than is required to compress incoming air. This produces more energy that can be recovered and transferred to the engine. High back pressure is necessary for EGR (exhaust gas recycling, exhaust gas recirculation) returns on diesel engines. High back pressure normally requires a valve or a restriction, so high back pressure is normally lost energy because a normal turbocharger can not be made unbalanced without overspeed. Increasing back pressure is not good for petrol and natural gas engines, as it increases the amount of exhaust gases that get stuck in the cylinder, which makes the engine more likely to have detonation problems. 21 WO 2011/096936 PCTHJS 2010/23398

According to another embodiment, a second turbine wheel can be placed on the turbine / compressor shaft to increase the energy recovered by the turbine and improve the fuel efficiency of the engine system. In addition, a second compressor wheel can be placed on the same shaft to increase the pressure increase capability of the super-turbocharger and allow cooling between steps. This makes the inlet temperature lower for a given pressure increase and thus produces less nitrous gases.

In addition, cooling of the turbine blades can be provided through the blade tips to reduce the temperature in high temperature applications. This can be done with hollow blade tips at the outer edge of the turbine. This special design of the tips increases the efficiency of the turbine and provides a way for cooling air to pass through the blades. Turbine blade cooling can also be provided by compressed air from the compressor, side-by-side over the housing to the rear of the turbine wheel. In addition, a heating tube can be used to cool the turbine wheel and blades.

In addition, a rotating emollient device can be used in the path of force.

Crankshaft energy or mechanical rotational energy from the driveline can be brought via a flexible shaft or device to soften pulses (either a spring-loaded or bendable) in such a way that torque pulses from the motor or driveline are removed without loss of this energy, before entering the housing . By not affecting the transmission with strong torque peaks, the requirement for peak torque is reduced. By eliminating these torque peaks, friction drives are more reliable, as the traction requirements are limited to the system's maximum torque. By minimizing these torque peaks on the friction drives, the size and surface contact area of the friction drives can be minimized. Minimal surface contact surfaces maximize system efficiency, and can still achieve the torque required to transmit the continuous force.

Alternatively, and in accordance with another embodiment, a design of a variable speed friction drive with hydraulic pumps with fixed displacement can be used instead of a shaft, belt or gear drive. This makes the system easier to package, which could be especially useful for very large engines that have a number of turbochargers. 22 WO 2011/096936 PCTHJS 2010/23398

In a further embodiment, illustrated in Figure 13, a second super-turbocharger is driven from a gearbox as a means of obtaining a higher pressure ratio, and as a means of obtaining cooler intake air to the engine by using a second intercooler. This is possible with a fixed speed ratio between the two super-turbochargers. The first super-turbocharger 1302 has an air intake line 1308 and compresses air, which is then supplied to the engine via the compressed air line 1310. Exhaust line 1314 receives exhaust gases from the engine to drive the turbine in the first super-turbocharger 1301. The exhaust gases leave the output exhaust line 1310. Exhaust line 1314 receives exhaust gases from the engine to drive the turbine in the first super-turbocharger 1302. The exhaust leaves the outlet line 1312.

The first super-turbocharger 1302 is coupled to the second super-turbocharger 1304 with a transmission gear 1306.

Figure 14A illustrates another embodiment of an implementation of the use of two super-turbochargers, such as a low-pressure super-turbocharger 1402 and a high-pressure super-turbocharger 1404. A standard super-turbocharger does not do a good job of recovering. the high pressure pulse that comes out of the cylinder when the exhaust valve is first opened. To improve this pulsed pressure recovery, as illustrated in Figure 14A, the high pressure exhaust valves 1406, 1408 are separated from the low pressure exhaust valves 1410, 1412 in a four valve engine. The exhaust gases from the high pressure exhaust valves 1406, 1408 are sent to the high pressure turbine 1434 via high pressure outlet manifolds 1430, while the exhaust gases from the low pressure exhaust valves 1410, 1412 are sent to the low pressure turbine 1420 via the low pressure outlet manifolds. via the outlets 1406, 1408 are opened first and the exhaust gases are sent to the high pressure turbine 1434, then the pulse energy is better recovered. The valves on the high pressure outlets 1406, 1408 close quickly, and then the valves on the low pressure outlets 1410, 1412 are opened during the entire blow-out stroke of the piston. The exhaust gases from the valves on the low pressure outlets 1410, 1420 are sent to the low pressure turbine 1420. This process reduces the work required by the piston to empty the cylinder. This process improves fuel efficiency at idle, or at least eliminates parasite losses at idle. The outlet from the high pressure turbine 1434 is also connected to the low pressure turbine 1420. Catalyst filters for particles from diesel engines (not shown) can also be placed before the low pressure turbine. 23 WO 2011/096936 PCTHJS 2010/23398

As also illustrated in Figure 14A, an EGR line 1438 is connected to the high pressure outlet manifold 1430. The EGR line 1438 allows some of the exhaust gases from the high pressure outlet manifold 1430 to be returned to the inlet manifold 1444, via the radiator 1440 and the EGR valve 1442. The exhaust gases from the high pressure outlet manifold 1430 sent through the EGR line 1438 are sent to the inlet manifold 1438 the purpose of recirculating exhaust gases. The exhaust gases passing through the exhaust gas recirculation line 1438 help to lower the combustion temperature in the combustion chamber, especially after cooling in the radiator 1440.

The exhaust gases contain moisture and other liquids that help to cool the temperature in the combustion chamber and thereby reduce the emission of nitrous gases from the engine. The amount of recycled exhaust gas is controlled by the EGR valve 1442. The EGR valve 1442 can be fixed, such as by using a restriction valve, or can be varied, depending on the monitored nitrous oxide emissions from the engine.

As also shown in Figure 14A, high pressure air is sent through the high pressure compressor manifold 1446 from the high pressure compressor 1432 to the inlet manifold 1444. Thus, the inlet manifold 1444 is held at a predetermined high pressure based on that coming out of the high pressure compressor. the recirculated gases must pass through the EGR line 1438, then the pressure in the high pressure manifold 1430 must be higher than the pressure in the inlet manifold 1444, as determined by the outlet pressure of the high pressure compressor 1432. In this regard, the valves in the high pressure outlets 1406, 1408 open sufficiently early , when residual pressure is still present in the cylinder to create sufficiently high pressure in the high pressure outlet manifold 1430 to drive the exhaust gases from the high pressure outlet manifold 1430 through the EGR line 1438. As shown below, the valves in the high pressure outlets 1406, 1408 open at a time when it is still present. a small amount of energy loss t in the process of driving the pistons downwards. The opening point for the high pressure valves is before the lower dead center but after the point with maximum torque from the piston to the crankshaft, which is the point where the connecting rods are substantially at 90 °. This point is located at about l00 °. The torque is proportional to the cosine of the angle of the connecting rods, so that the lower the piston is when the high pressure valves are opened, the less energy is lost to drive the pistons. However, there is a significant amount of residual pressure in the cylinder, which can be released by the high pressure valves before the lower dead center is reached, and this pressure can be used to drive the exhaust gases in the EGR line 1438 into the high pressure turbine 1434. By pre-24-WO 2011 / 096936 PCTHJS 2010/23398 release gas from the cylinder using the high pressure valves at the high pressure outlets 1406, 1408, then a large amount of the residual pressure in the cylinder is released before the low pressure outlets 1410, 1412 are opened. When opened, the low pressure outlets 1410, 1412 are capable of releasing most of the pressure from the cylinders. In this way, the residual pressure in the cylinders is used to direct exhaust gases both through the EGR line 1438 to reduce nitrous oxide emissions and to drive the high pressure turbine 1434, which adds additional power and efficiency to the ITIOÉOITI.

As also shown in Figure 14A, the exhaust gases from the low pressure outlet manifold are used to drive the low pressure turbine 1420 in the low pressure super-turbocharger 1402. Exhaust gases coming from the high pressure turbine 1434 are combined with the low pressure exhaust gases from the low pressure emissions 1410, 1412 to drive the low pressure turbine 1420. the low pressure turbine 1420 is discharged through the outlet 1436. The low pressure turbine 1420 is connected to the low pressure compressor 1418, which compresses the inlet air 1422 to a predetermined level. Line 1424 sends the compressed air from the low pressure compressor 1418 to the inlet of the high pressure compressor 1432, which works to further compress the compressed air in 1424 to produce air which is compressed to higher pressure, which is sent to the inlet manifold 1444 through the high pressure manifold 1446.

Figure 14B illustrates a variation of the embodiment of Figure 14A. As illustrated in Figure 14B, the high pressure outlets 1406, 1408 have been combined into a high pressure outlet manifold connected to the high pressure turbine 1434. In other words, all of the high pressure exhaust manifold 1430 is sent to the high pressure turbine 1434 to drive the high pressure turbine 14 in the high pressure turbine.

The high pressure compressor 1432 receives compressed air in line 1424 from the low pressure compressor 1418 in the low pressure super-turbocharger 1402 which compresses air from the inlet 1422. The output air from the high pressure compressor 1432 is sent to the input manifold 1444 via the compressor low pressure manifold 14. , in the low pressure outlet manifold 1428, which comes from the low pressure outlets 1410, 1412. Exhaust gases from the low pressure turbine 1420 are discharged through the exhaust gas outlet 1436. The high pressure gases from the high pressure outlet manifold 1430, which drive the high pressure turbine 1434, are sent to line recirculation 1496. PCTHJS 2010/23398 to the inlet manifold 1444. The high pressure gases from the high pressure outlet manifold 1430, which drive the high pressure turbine 1434, do not have a significant pressure drop and have a sufficiently high pressure to inject the exhaust gases from the EGR line 1426 into the inlet manifold 144 Figure 14B provides the largest reduction for nitrous gases, since substantially all of the exhaust gases from the high pressure outlet manifold 1430 are recycled to the inlet manifold 1444.

As also illustrated in Figure 14B, an overflow valve 1448 can be used to allow high pressure exhaust from the high pressure outlet manifold 1430 to bypass the EGR line 1426 directly.

The high pressure exhaust may occasionally be too hot and / or may be at a pressure that will overload the high pressure turbine 1434. In this case, the overflow valve 1448 may be opened to send high pressure exhaust from the high pressure outlet manifold 1430 directly to the EGR line 1426.

In addition, an EGR valve 1450 can be added, the bay is connected to the EGR line 1426 to the low pressure outlet manifold 1428. If sufficient exhaust gases are sent through the EGR line 1426, then some of these gases can be sent from the EGR line 1426 to the low pressure outlet manifold 1428 via the EGR valve 1450. The excess gases from the EGR line 1426 can then be used to drive the low pressure turbine 1420 to add extra power to the engine by increasing the pressure in the inlet manifold 1444. Use of the EGR valve 1450 provides a further way in which recirculated gases can be recycled to add additional power to the engine and increase the efficiency of engine operation.

Figure 14C illustrates another modification of the embodiments of Figures 14A and 14B.

As shown in Figure 14C, incoming air 1422 is compressed by the low pressure compressor 1418.

The compressed air from the low pressure compressor 1418 is sent via line 1424 to the inlet manifold 1444. As also illustrated in Figure 14C, the second turbine, the high pressure turbine, is not used and all recirculated gas is recycled from the high pressure outlets 1406, 1408 via the EGR line 1426 to inlet 44. from the low pressure outlets 1410, 1412 are combined in line 1428 to drive the low pressure turbine 120. The exhaust gases are then released via the exhaust 1436. Therefore, all the gases from the high pressure outlets 1406, 1408 are fed back to the inlet manifold 1444 to create a large reduction of nitrous gases. Alternatively, an EGR valve 1450 may be used to send some of the exhaust gases in the EGR line 1426 to the low pressure outlet manifold 1428, which provides additional power to the low pressure turbine 1420 and reduces the amount of recirculated gases in the EGR line 1426. The EGR valve 1450 may be adjusted 26 WO 2011/096936 PCTHJS 2010/23398 to adjust the amount of exhaust gases sent from EGR line 1426 to low pressure outlet manifold 1428. This process may be beneficial if a sufficient amount of exhaust gases is recirculated in EGR line 1426 to reduce engine emissions of nitrous gases.

Figure 14D is a diagram showing valve lift, cylinder pressure and velocity of velocity with respect to the position of the piston after the upper dead center. As shown in Figure 14D, the cylinder pressure 1450 decreases steadily after the top dead center, all the way through the stroke of the piston. The lifting of the high pressure valve 1456 creates the high pressure fate 1452. The lifting of the high pressure valve 1456 takes place around 100 ° s rotation and creates a large exhaust of high pressure exhaust 1452, which is blown out through the high pressure outlets 1406, 1408 (Figures 14-a, 14B and 14C). Lifting the low pressure valve creates low pressure flow 1458 through the low pressure outlets 1410, 1412. As a result, the cylinder pressure 1450 drops further in the cylinder.

Figure 14E is a pressure / volume diagram of cylinder pressure versus Cylinder volume as the piston moves downward and then upward in the cylinder. Near zero represents the upper dead center, while 1 represents the lower dead center of the rotation of the cylinder. Two curves are shown in Figure 14E. Curve 1464 represents the cylinder pressure curve for an engine that does not use the Riley cycle. Curve 1462 is a graph illustrating cylinder pressure versus volume in a cylinder for a Riley cycle device, as illustrated in Figures 14A-C. At point 1466, the high pressure valve on the Riley cycle device opens, as illustrated in Figures 14A-C, and the pressure is reduced. The area 1468, between points 1466, 1470, represents the energy lost by opening the high pressure valve. However, as shown in Figure 14E, at point 1472, the pressure in the Riley cycle device falls below the pressure in a non-Riley cycle device and remains below the pressure of a non-Riley device all the way to point 1474. Between point 1472 and point 1474 is the lower pressure in the cylinder, which results in less back pressure on the cylinder when the cylinder fl moves from point 1472 to point 1474. The large area between the Riley cycle curve 1462 and a normal curve 1464 between points 1472 and 1476 , as indicated by 1478, indicates the amount of energy saved when moving the piston in the cylinder at the lower pressure.

In an alternative embodiment, a super-turbocharger can be used as an after-treatment air pump, as well as for the engine, eliminating the need for a burner-only pump. 27 WO 2011/096936 PCTHJS 2010/23398

In another embodiment, a regulator (not shown) is included to prevent excessive speed, which keeps the compressor away from a pressure surge condition and which regulates the maximum efficiency of the turbine and compressor. A super-turbocharger can be unique compared to a normal turbocharger because the speed (for the turbine / compressor) where the turbine efficiency is highest and the speed where the compressor's efficiency is highest can be the same speed. Controlling this maximum efficiency for a given power increase requirement can be modeled and programmed into an electronic controller. An electric actuator can provide control, but an actuator is not necessary for the electric transmission.

In another embodiment, the cooling system of the super-turbocharger draws a vacuum into the housing, thereby reducing the aerodynamic losses in the high-speed components.

In another alternative embodiment, a two-clutch super-turbocharger includes an automatic transmission manual transmission. This type of gearbox shifts very smoothly because it has a coupling at both ends. Figure SC illustrates that the gearboxes can be of many different types.

In another embodiment, friction drives are used for both the transmission and the speed reduction from the turbine shaft. With ball bearings, the traction fluid also works as a lubricant. During super-turbo compression, the system improves load acceptance, reduces soot emissions, and provides up to 30% increase in torque at low speeds and up to 10% increase in maximum power. During turbo assembly, the system provides 30% more torque at low pressure, which means that the engine can be 30% -50% smaller, has less engine weight and improved fuel economy by up to 17% or more. Figure 15 illustrates the simulated BSFC (Brake Specific Fuel Consumption) improvement for a natural gas engine.

In addition, a catalyst, a DPF (Diesel Particle Filter) or even a burner plus a DPF, can be placed in front of the super-turbocharger turbine to heat the exhaust gases to a higher temperature than the engine heat. Higher temperatures expand the gas even more, which makes the speed of fate through the turbine higher. Approximately 22% of this heat can be converted to mechanical work by the super-turbocharger, assuming 80% 28 WO 2011/096936 PCTHJS 2010/23398 turbine efficiency. Normally, feeding higher volumes from the exhaust gases would slow down the turbine's response and create an even greater turbo delay, but the super-turbocharger cures this problem with the friction drive 114 and the continuously variable gearbox 116 driving the pressure response. Similar techniques using catalytic converters are disclosed in Patent Application No. PCT / US 2009/051741 filed 2009-07-24 by Van Dyne et al., Entitled "Improving Fuel Efficiency for a Piston Engine Using a Super-Turbocharger", which is specifically incorporated herein by reference, for all that it publishes and teaches out.

Figure 16 is a single line simplified illustration of an embodiment of a highly efficient, super-turbocharged engine system 1600. As will be apparent to those skilled in the art through the following description, such a super turbocharged engine system 1600 special applicability in diesel engines and in some spark-ignited gasoline engines used in both passenger and commercial vehicles, and therefore the illustrative examples discussed herein use such an environment to aid in the understanding of the invention. However, it should be noted that these embodiments of system 1600 are applicable to other operating environments such as land-based power generators and other land-based motors, and the above examples are to be taken as illustrative and not restrictive in any way.

As shown in Figure 16, the super-turbocharger 1604 includes a turbine 1606, a compressor 1608 and a gearbox 1610 which is coupled to the crankshaft 1612 in the engine 1602 or other parts of the driveline. Although not required in all embodiments, the embodiment of Figure 16 also includes an intercooler 1614 to increase the density of the air supplied to the engine 1602 from the compressor 108 to further increase the power available from the engine 1602.

Super-turbochargers have certain advantages over turbochargers. A turbocharger uses a turbine that is powered by exhaust gases from the engine. This turbine is connected to a compressor that compresses the intake air which is then sent to the engine cylinders. As such, the engine experiences a delay in this addition of power until there is enough exhaust gas to spin up the turbine to drive the compressor, which is mechanically coupled to the turbine, to generate sufficient power increase. To minimize the delay, smaller and lighter turbochargers are typically used. The minor inertia of a lightweight turbocharger 29 WO 2011/096936 PCTHJS 2010/23398 allows them to spin up very quickly, thereby minimizing the delay of its operation.

Unfortunately, such smaller and lighter turbochargers can be run too fast when the engine is operating at high speed and a large amount of exhaust de fate and temperature are produced. To prevent times at too high a speed, typical turbochargers include an over-valve that is installed in the exhaust pipe before the turbine. The exhaust valve is pressure operated and releases some of the exhaust gases around the turbine when the outlet pressure from the compressor exceeds a predetermined limit. This limit is set to a pressure that indicates that the turbocharger is close to going too fast. Unfortunately, this results in some of the energy available from the engine exhaust being lost.

In view of the fact that conventional turbochargers sacrifice performance at low speeds to gain power at high speeds, devices known as super-turbochargers have been developed.

Such a super-turbocharger is described in U.S. Pat. 7,490,594 entitled "Super-Turbocharger", issued 2009-02-17, which is specifically incorporated herein by reference for all it publishes and teaches.

As discussed in the application referred to above, in a super-turbocharger the compressor is driven by the engine crankshaft via a gearbox connected to the engine at low speeds when sufficiently heated exhaust gases from the engine are not available to drive the turbine.

The mechanical energy supplied by the engine to the compressor reduces the turbo delay problem associated with conventional turbochargers, and allows a larger or more efficient turbine and compressor to be used.

The super-turbocharger 1604, illustrated in Figure 15, is there to provide compressed air from compressor 1608 to engine 1602 without suffering from the turbo delay problem of a conventional low speed turbocharger and without wasting available energy from engine exhaust gases sent to the turbine 1606 at high speed.

These benefits come from including the super-turbocharger gearbox 1610 which can both extract power from and transmit power to the engine crankshaft. 1612 to both drive the compressor 1608 and drive the turbine 1606, respectively, under different operating conditions for the engine 1602. WO 2011/096936 PCTHJS 2010/23398

During start-up, when conventional turbochargers suffer a delay due to the lack of sufficient power from the engine exhaust heat to drive the turbine, the super-turbocharger provides a supercompressor function in which the power is taken from the crankshaft 1612 via the super-turbocharger gearbox 1610 to drive the compressor 1608 to provide sufficient power to the engine 1602. When the engine reaches normal speed and the amount of power available from the engine exhaust heat is sufficient to drive the turbine 1606, the amount of power taken from the crankshaft 1612 by the gearbox 1610 is reduced. 1606 to provide power to the compressor 1608 to compress the engine intake air for use by the engine 1602.

As the engine speed increases, the amount of power available from the heat of the engine exhaust increases to the point where the turbine 1606 would run too fast in a conventional turbocharger.

However, with the super-turbocharger 1604, the excess energy - which comes from the heat of the engine exhaust to the turbine 1606 - is sent through the gearbox 1610 to the engine crankshaft 1612 while the compressor 1608 is maintained at the right speed to provide the ideal power to the engine 1602. is available from the engine's exhaust heat, the more power is generated by turbine 1606, which is sent via gearbox 1610 to crankshaft 1612 while maintaining optimal power supply available from compressor 1608. This loading of turbine 1606 from gearbox 1610 prevents turbine 1606 from going too fast, and maximizes the efficiency of the power extracted from the engine exhaust. As such, a conventional over-fate valve is not required.

While the amount of power available to drive the turbine 1606 in a conventional super-turbocharged application is strictly limited to the amount of power available from the engine exhaust, the turbine 1606 can generate significantly more power if the thermal energy or flow provided to the turbine blades can be fully utilized and / or can be increased. However, the turbine 1606 cannot operate at a temperature higher than a certain value without being damaged, and the mass flow is conventionally limited to the exhaust gases coming out of the engine 1602.

In view of this, the design of system 1600 protects the turbine 1606 from high temperature transients by placing a catalytic filter for diesel particles 1616 in front of the turbine 1606. In one embodiment, the catalytic filter for diesel particles is located in front of the turbine near the outlet manifold, enabling an exothermic reaction resulting in an increase in the exhaust gas temperature during prolonged high speed or load work for the engine. Using a diesel particulate catalytic filter, energy can be recovered from the soot, hydrocarbons and carbon monoxide burned in the 1616 diesel particulate catalytic filter to add power to the super-turbocharger located after (downstream of) the diesel particulate filter. Energy recovery can be obtained from either a conventional catalytic filter for diesel particles having a very limited flow capacity, with almost 100% soot capture, or by using a flow-through catalytic filter for diesel particles. A flow-through catalytic filter for diesel particles is that catalytic filter for diesel particles that only captures about half of the soot and lets the other half through. Both types of filters for diesel particles are catalyzing to be able to have emissions burning at a relatively low temperature. The catalysis in the filter for diesel particles is accomplished by providing a platinum plating on the elements of the particulate filter, which ensures that soot, hydrocarbons and carbon monoxide burn at low temperature. In addition, it is possible to use filters for diesel particles and a burner to burn the soot from the filter for diesel particles upstream of the super-turbocharger. Gasoline engines generally do not have enough soot to require a filter for diesel particles. However, some direct injection gasoline engines produce enough soot and other particles that the use of a particulate filter can be advantageous, and the use of a catalytic filter for diesel particles can be done in the manner described herein.

To cool the exhaust gases, before they reach the turbine, some of the compressed air generated by the compressor is fed directly into the exhaust gases upstream of the turbine, via a control valve 1618, and added to the engine exhaust gases which leave the catalytic filter for diesel particles 1616 The cooler intake air expands and cools the exhaust gases and adds additional weight to the exhaust gas exhaust, adding additional power to the 1606 turbine as described in more detail below. When cooler air is given to the hot exhaust gases to maintain the temperature in the combined flow of the 1606 turbine to the optimum temperature, the energy and mass delivered to the turbine blades also increase. This significantly increases the power delivered by the turbine to the engine crankshaft.

In order not to affect the stoichiometric reaction in the catalytic filter for diesel particles 1616, the compressor return air is injected downstream from the catalytic filter 32 WO 2011/096936 PCTHJS 2010/23398 for diesel particles 1616. In such an embodiment the exhaust gases pass through a catalytic filter for diesel particles 1616 and the temperature of the exhaust gases is increased by the exothermic reaction. The compressed return air is then added and expanded so that the total mass delivered to the turbine increases. Embodiments of the present invention control the amount of compressed air delivered to cool the exhaust gases and to drive the turbine to ensure that the combination of the cooler compressed return air and engine exhaust gases is delivered to the turbine at optimum pressure for turbine blade operation.

Since the catalytic filter for diesel particles 1616, illustrated in Figure 16, has a large thermal mass compared to the exhaust gases from engine 1602, the catalytic filter 1616 for diesel particles acts as a thermal damper in the beginning, which prevents a thermal peak with a high temperature from reaching the turbine 1606. However, since the reactions in the catalytic filter 1616 for diesel particles are exothermic in nature, the temperature of the exhaust gases leaving the catalytic filter for diesel particles is higher than the temperature of the exhaust gases entering the catalytic filter 1616 for diesel particles. As long as the temperature of the exhaust gases entering the turbine remains below the maximum operating temperature of the turbine 1606, there is no problem.

However, during long-term high speed and heavy load operation of the engine 1602, the outlet temperature of the catalyzed exhaust gases from the catalytic filter for diesel particles 1616 may exceed the maximum operating temperature of the turbine 1606. As described above, the temperature of the exhaust gases leaving the diesel particulate catalytic filter 1616 reduced by sending a portion of the compressed air from the compressor 1608 via the return valve 1618, and mixing it with the exhaust gases exiting the diesel particulate catalytic filter 1616. Significantly improved fuel economy is achieved by not using fuel as a refrigerant under these circumstances, as is done in conventional systems. In addition, the operation of the gearbox is controlled to allow the compressor 1608 to provide a sufficient amount of compressed air to provide optimum power amplification to the engine 1602 and the compressed return air to the turbine 1606 via the return valve 1618. The excess power generated by the turbine 1616 resulting from increased mass consumption of compressed air through the turbine is sent via the gearbox 1610 to the crankshaft 1612, which also increases the fuel efficiency. 33 WO 2011/096936 PCTHJS 2010/23398

The outlet temperature of the compressed air 1608 is typically between about 200 ° C and 300 ° C. A conventional turbine can operate optimally to extract power from gases at approximately 950 ° C, but not at higher temperatures without distortion or possible failure. Due to material limitations of the turbine blades, optimum power is achieved at approximately 950 ° C.

Since the material limits the temperature of the exhaust gases to approximately 950 ° C, the supply of more air increases to increase the mass flow through the turbine at the temperature limit, e.g. 950 ° C, turbine performance.

With such a fate of compressed return air at 200 ° C to 300 ° C is helpful in reducing the temperature of the exhaust gases coming out of the catalytic filter 1616 for diesel particles, it is known that maximum power from the turbine 1606 can be supplied when the temperature and mass flow are maximized within the thermal limits of the turbine 1606. As such, in one embodiment, the amount of return air is controlled so that the combination of exhaust and return air is maintained at or near the maximum operating temperature of the turbine so that the amount of power delivered to the turbine is maximized or significantly increased . Since all of this excess power is not normally required by the compressor 1608 to provide optimum extra power to the engine 1602 and to send the compressor return air to via the return valve 1618, excess power can be transmitted from the gearbox 1610 to the crankshaft 1612 in the engine 1602 to thereby increase overall efficiency or the power of engine 1602.

As discussed above, in one embodiment, the coupling between the compressor return air via return valve 1618 uses a catalytic filter 1616 for diesel particles as the thermal buffer between the engine 1602 and the turbine 1606. As such, air is supplied from the compressor downstream of the catalytic filter for diesel particles so as not to interfere with the stoichiometric reaction within the 1616 diesel particulate catalytic filter. That is, in embodiments that use a catalytic filter 1616 for diesel particles that inject the compressor return air upstream of the catalytic filter 1616 for diesel particles, it would result in too much oxygen being supplied to the catalytic filter 1616 for diesel particles, and thereby the catalytic filter 1616 for diesel particles would be prevented from producing a stoichiometric reaction required for proper operation.

Since optimum efficiency for power generation of the turbine 1606 is achieved when the temperature of the gas mixture - the compressor return air and exhaust gases - which comes to the turbine blades is maximized (within the material limits of the turbine itself), the amount of return air is from the compressor passed through by the return valve 1618 is limited so as not to significantly reduce the temperature below such an optimized temperature. When the catalytic filter 1616 for diesel particles produces more thermal energy via an exothermic reaction and the temperature of the converted exhaust gases from the catalytic filter 1616 for diesel particles increases to a temperature exceeding the maximum operating temperature of the turbine 1606, more return air from the compressor can be provided via the return valve 1618 which increases the mass fl and the energy sent to the turbine 1606. As the amount of thermal energy generated by the catalytic filter 1616 for diesel particles decreases, the amount of return air from the compressor sent via return valve 1618 can also be reduced to avoid providing more air than necessary. results in maintaining the temperature of the gas mixture at the optimum operating temperature.

In another embodiment, the system uses the return valve 1618 to send the cooler compressor air back to the exhaust gases before the turbine, when operating at low speed and high load, to avoid overloading the compressor ("surging"). Compressor overload occurs when the compressor pressure becomes high but the mass flow into the engine is low due to the engine having a low speed and not requiring much air flow in the intake. Surging (or aerodynamic stall) in the compressor as a result of low air leakage through the compressor blades causes the compressor efficiency to fall very quickly. In the case of a normal turbocharger, sufficient surge can cause the turbine to stop rotating. In the case of a super-turbocharger, it is possible to use power from the engine crankshaft to drive the compressor to a surge. Opening the return valve 1618 allows some of the compressed air to return around the engine.

This return fate brings the compressor out of surge and allows higher boost pressure to reach the engine 1602, thereby allowing the engine 1601 to generate more power than would normally be possible at low engine speeds. Injection of the compressed air into the exhaust gases before the turbine minimizes the force required from the engine to supercompress to a high amplified pressure level.

In another embodiment, an additional cold start control valve 1620 may be included for operation during cold start with rich gas landing. During such a cold start, the exhaust gases from the 1602 engine typically contain an excess of unburned fuel. Since this rich mixture is not stoichiometric WO 2011/096936 PCTHJS 2010/23398, the catalytic filter 1616 for diesel particles is not capable of completely reducing the unburned hydrocarbons (UHC) in the exhaust gases. During this time, the cold start control valve 1620 can be opened to provide the compressor return air to the inlet of the catalytic filter for diesel particles to provide extra oxygen to bring the rich mixture down to stoichiometric levels. This allows the catalytic filter for diesel particles to start up faster and more efficiently reducing emissions during a cold start. If the engine is idling, a normal turbocharger would not have any extra pressure to provide return air. However, the gear ratio of the gearbox 1610 can be adjusted to provide sufficient speed for the compressor to generate the pressure needed for air to flow through the valve 1620. In this regard, control signal 1624 may be used to adjust the gear ratio of the gearbox 1610 so that sufficient speed can be provided from the engine crankshaft 1612 to the compressor 1608 at idle, especially at cold start, to compress sufficient loft to flow through the cold start valve 1620 and ignite the catalytic filter for diesel particles 1616 with a sufficient amount of oxygen.

The requirement for the additional oxygen is typically limited to a cold start, and often lasts only 30 to 40 seconds. Many vehicles currently include a separate air pump to supply this oxygen during a cold start, at a significant cost and weight if one considers the short time that such an air pump needs to operate. By replacing the separate air pump with a simple control valve for cold state 1620, significant savings in cost, weight and complexity can be realized. Since the super-turbocharger 1604 can control the speed of the compressor 1608 via the gearbox 1610, the control valve 1620 for cold start can consist of a simple on / off valve.

The amount of air provided during the cold start can then be controlled by controlling the speed of the compressor 1608 via the gearbox 1610 using a control signal 1624.

The cold start control valve 1620 can also be used during periods of very high temperature operation if fuel is used as the coolant within the engine and / or for the catalytic filter 1616 for diesel particles, despite the negative effect on fuel economy. In such situations, the cold start control valve 1620 will be able to provide the extra oxygen necessary to bring the rich exhaust gases back to stoichiometric levels to enable the diesel particulate catalytic filter 1616 to properly reduce the amount of PCTHJS 36 WO 2011/096936 PCTHJS 2010/23398 unburned hydrocarbon emissions in the exhaust gases. This provides a significant advantage for the environment compared to previous systems.

In embodiments where the cold start control valve 1620 is an on / off valve, the system may modulate the cold start control valve 1620 to vary the amount of compressed air provided, to bring the exhaust gases down to stoichiometric levels. Other types of variable flow control valves can also be used to fulfill this same function.

Figure 16 also shows a control unit 1640. Control unit 1640 controls the operation of the return valve 1618 and the cold start valve 1620. Control unit 1640 works to optimize the amount of air passing through the return valve 1618 under different conditions. The amount of air passing through the return valve 1618 is the minimum amount of air necessary to achieve a specific desired condition, as described above. There are two specific conditions in which control unit 1640 controls return valve 1618, and they are: 1) when the surge limit of the compressor for a given gain requirement is close to low speed, and, 2) the temperature of the gas mixture entering the turbine 1606 is close at high speeds and high loads.

As shown in Figure 16, the control unit 1640 receives a signal 1630 for the temperature of the gas mixture from a temperature sensor 1638 which senses the temperature of the gas mixture in the cooling air supplied from compressor 1608 and mixes with hot exhaust gases produced by the catalytic filter 1616 for diesel particles. In addition, the controller 1640 detects the signal 1632 concerning the pressure of the compressed air inlet from the pressure sensor 1636 located in the compressed air line coming from compressor 1608. In addition, a signal 1626 for engine speed and a signal 1628 for engine load are sent from the engine or from a throttle. to the control unit 1640.

With respect to controlling the temperature of the gas mixture sent to the turbine 1606 under high speed and high load conditions, the control unit 1640 limits the temperature of the gas mixture to a temperature which maximizes the operation of the turbine 1606 without being so high as to damage the mechanisms in the turbine. turbine 1606. In one embodiment, a temperature around 925 ° C is an optimal temperature for the gas mixture to drive turbine 1606. Once the temperature of the gas mixture fed into turbine 1606 begins to exceed 900 ° C, the return valve 1618 is opened, to allow compressed air from the compressor 1608 to cool the hot exhaust gases from the catalytic filter 1616 for diesel particles before entering the turbine 1606.

The control unit 1640 can be designed to aim for a temperature of approximately 925 ° C, with an upper limit of 950 ° C and a lower limit of 900 ° C. The limit of 950 ° C is that at which damage to the turbine 1606 can occur using conventional materials. Of course, the control unit can be designed for other temperatures, depending on the specific type of components and materials used in the turbine 1606. A conventional proportional integrating derivative (PID) control logic unit can be used in the control unit 1606 to produce these controlled results.

The advantage of controlling the temperature of the gas mixture entering the turbine 1606 is that the use of fuel in the exhaust gases to limit the inlet temperature of the gas mixture is eliminated. The use of cooler compressed air to cool the hot exhaust gases from the diesel particulate catalytic filter 1616 requires a large amount of air, which has a large mass to achieve the desired cooler temperature of the gas mixture. The amount required to cool the hot exhaust gases coming from the catalytic filter 1616 for diesel particles is large because cooler compressed air from the compressor 1608 is not that good coolant, especially when compared to a liquid fuel injected into the exhaust gases. The hot exhaust gases from the outlet of the diesel particulate catalytic filter 1616 cause the cooler compressed gas from the compressor 1608 to expand to create the gas mixture.

Since a large amount of the cooler compressed air from compressor 1608 is required to lower the temperature of the hot exhaust gases from the catalytic filter 1616 for diesel particles, a large amount of gas mixture flows through turbine 1606, which greatly increases the yield from turbine 1606. The turbine force increases with the difference between the force created by the mass and the work required to compress the compressed air passing through the return valve 1618. By receiving the gas mixture temperature signal 1630 from the temperature sensor 1638 and controlling the addition of compressed air via the return valve 1618, the maximum temperature is not exceeded.

Control unit 1640 also controls return valve 1618 to limit surge in compressor 1618.

The surge limit is a boundary line that varies as a function of the gain pressure, the air flow through the compressor and the design of the compressor 1608. Compressors, such as compressor 1608, which are typically used in turbochargers, exceed the surge limit when the flow of inlet air 1622 is small and the pressure difference between the inlet air 1622 and the compressed air 38 WO 2011/096936 PCTHJS 2010/23398 is large. In conventional supercompressors, the fl fate of inlet air 1622 is low when the engine speed 1626 is low. At low speeds, when the compressed air is not used in large volumes of the engine 1602, the mass flow of inlet air 1622 is low and surge occurs because the rotary compressor 1508 cannot force air into a high pressure line without a reasonable flow of inlet air 1622. The return valve 1618 allows fate through the compressed air line 1609 and prevents or reduces surge in the compressor 1608. When a surge in the compressor 1608 takes place, the pressure in the compressed air line 1609 cannot be maintained. Therefore, at low speed and high load of motor 1602, the pressure of the compressed air in the compressed air line 1609 may drop below desired levels.

By opening the return valve 1618, the flow of inlet air 1622 through the compressor 1608 increases, especially at low speeds and high loads on the engine, allowing the desired amount of gain to be achieved in the compressed air line 1609. Return valve 1618 can simply be opened until the desired pressure in line 1609 is reached. However, by simply detecting the boost pressure in the compressed air line 1609, surge will occur before the return valve 1618 is opened to take the compressor out of the surge state.

However, it is preferable to set the surge limit and open the return valve 1618 in advance, before the surge situation arises. For a given speed and desired gain level, a surge limit can be set. The return valve 1618 may begin to open before compressor 1608 reaches a calculated surge limit. Opening the valve early enables the compressor to spin faster to a higher boost pressure, because the compressor remains closer to the higher efficiency points of the compressor's operational parameters. Rapid increase of the gain pressure at low speed can then be achieved. By opening the valve before a surge occurs, a more stable control system can also be achieved. 39 WO 2011/096936 PCTHJS 2010/23398

Opening the return valve 1618 in such a way that improvement of the response of the motor 1602 is achieved by allowing 1602 to obtain higher gain pressure faster when motor 1602 has low speed. Compressor 1608 is also more efficient, resulting in less work for gearbox 1610 to achieve supercompression. A model of control of control limit can be done within standard model-based control simulation code, such as MATLAB. Model building in this way allows simulation of the controller 1640 and automatic programming of algorithms for the controller 1640.

A model-based control system, as described above, is unique in that the use of the gearbox 1610 to control the rotation of the turbine 1606 and compressor 1608 generates boost pressure without turbo delay. In other words, the gearbox 1610 can extract rotational energy from the crankshaft 1612 to drive compressor 1608 to achieve the desired gain from compressed air line 1609 very quickly and before turbine 1606 generates sufficient energy to drive compressor 1608 at such a desired level. In this way, control units in a conventional turbocharger are reduced or eliminated to reduce delay. The model-based control of the controller 1640 should be designed to maintain optimal efficiency of compressor 1608 within the operational parameters of compressor for compressor 608.

The control model for control unit 1640 should also be carefully modeled with respect to the operational parameters of the pressure, such as depending on the mass fl allowed by the engine for a given target in speed and load, in which the target for speed and load can be defined relative to the position of the vehicle. gas control. As shown in Figure 16, the motor speed signal 1626 can be obtained from motor 1602 and sent to control unit 1640. Similarly, the motor load signal 1628 can be obtained from motor 1602 and sent to control unit 1640. Alternatively, these parameters can be obtained from sensors located on motor motor throttle control (not shown).

The return valve 1618 can then be controlled in response to a control signal 1642 generated by control unit 1640. Pressure sensor 1636 generates the signal 1632 for incoming air pressure sent to control unit 1640, which calculates the control signal 1642 in response to engine speed signal 1626, engine pressure signal 16 and compressor 1606 input. 40 WO 2011/096936 PCTHJS 2010/23398

Under operating conditions of engine 1602 in which the temperature in compressor 1608 does not approach a surge limit, and the temperature of the gas mixture, as detected by the temperature sensor 1638, is not reached, the return valve 1618 is closed so that the system functions as a conventional super-turbocharged system. This applies during a majority of operational parameter values for the motor 1602. When high load and low speed occur, the return valve 1618 is opened to prevent surge. Similarly, when high speed and high load for engine 1602 occur, high exhaust gas temperatures are generated at the outlet of the diesel particulate catalytic filter 1616, so that the return valve 1618 must be opened to reduce the temperature of the fuel mixture sent to turbine 1606 to a temperature under the one that would cause damage to turbine 1606.

Figure 17 is a detailed diagram of the embodiment of the high efficiency super-turbocharged system 1600 illustrated in Figure 16. As shown in Figure 17, engine 1602 includes a super-turbocharger system that has been modified as described above with respect to Figure 16. to provide overall higher efficiency than conventional super-turbocharged engines, and also provide high optimum efficiency at low speeds and high loads, and high optimum efficiency at high speeds and high loads. The super-turbocharger includes a turbine 1606 which is mechanically connected to a shaft of compressor 1608. Compressor 1608 compresses inlet air 1622 to line 1704. Line 1704 is connected to return valve 1618 and intercooler 1614. As shown above, intercooler 1614 cools the compressed air which gets heated during the compression process. The intercooler 1614 is connected to the compressed air line 1726, which in turn is connected to the inlet manifold (not shown) in motor 1602. Pressure sensor 1636 is connected to the compressed air line 1704 to detect the pressure and provide a pressure reading via the pressure signal 1632 in the compressed inlet air, which is sent to control unit 1640. The return valve 1618 is controlled by the control signal 1642 for the return valve 1618, generated by the control unit 1640, as shown above. Under certain operating conditions, the return valve 1618 is opened to provide compressed air from the compressed air line 1704 to a mixing chamber 1706. 41 WO 2011/096936 PCTHJS 2010/23398

As shown in the embodiment of Figure 17, mixing chamber 1706 simply includes a series of openings 1702 in the outlet line 1708 of the catalytic filter for diesel particles, which is surrounded by the line 1704 for compressed air, so that the compressed air provided from the line 1704 for compressed air passes through the openings 1702 to mix with the exhaust gases in the outlet line 1708 from the catalytic filter for diesel particles.

Any desired type of mixing chamber can be used to mix the cooler compressed air with the exhaust gases to lower the temperature of the exhaust gases.

Temperature sensor 1638 is located in the diesel particulate catalytic filter outlet line 1708 to measure the temperature of the exhaust gas exhaust line 1708 from the diesel particulate catalytic filter. Temperature sensor 1638 provides a temperature signal 1630 for the temperature of the gas mixture to the control unit 1640, which controls the return valve 1618 to ensure that the temperature of the exhaust gases in the outlet line 1708 from the catalytic filter for diesel particles does not exceed a maximum temperature that would damage the turbine 1606. The catalytic filter 1616 for diesel particles is connected to the outlet manifold 1710 via the inlet line 1714 to the catalytic filter for diesel particles. By placing the catalytic filter 1616 for diesel particles near the outlet manifold 1710, the hot exhaust gases from the engine go directly into the catalytic filter 1616 for diesel particles, which helps to activate the catalytic filter 1616 for diesel particles. In other words, the close placement of the diesel particulate catalytic filter 1616 to the exhaust for the engine exhaust does not allow the exhaust gases to cool significantly before entering the diesel particulate catalytic filter 1616, which increases the performance of the diesel particulate catalytic filter 1616. As the exhaust gases pass through the diesel particulate catalytic filter 1616, the diesel particulate catalytic filter 1616 adds additional heat to the exhaust gases. These very hot exhaust gases at the outlet of the diesel particulate catalytic filter 1616 are sent to the particulate catalytic filter outlet line 1708 and are cooled in the mixing chamber 1706 by the compressed inlet air from the compressed air line 1704. Depending on the temperature of the very hot exhaust gases produced at the outlet of the diesel particulate catalytic filter 1616, which varies depending on the operating conditions of engine 1602, different amounts of compressed intake air will be added to the exhaust gases under high speed and high load conditions. Under conditions of low engine speed and high load, the return valve 1618 also functions to allow inlet air to enter to pass through the compressor 42 WO 2011/096936 PCTHJS 2010/23398 to avoid surge. Surge is similar to the aerodynamic stable for the compressor blades, which arises as a result of conditions with low fl fate through the compressor at low speed. When surge occurs, the pressure drops into the inlet manifold (not shown) because the compressor 1608 is not capable of compressing inlet air. By allowing air to pass through the compressor as a result of the return valve 1618 being opened, pressure can be maintained in the inlet manifold so that, when high torque is at low speed, the high torque can be achieved due to the high pressure in the inlet manifold.

As shown above, when the engine is operating at high speed and high load, the catalytic filter 1616 for diesel particles causes a large amount of heat to be generated in the exhaust gases supplied to the exhaust line 1708 from the catalytic filter for diesel particles.

By providing compressed, cooler inlet air to the outlet line 1708 from the catalytic filter, the hot exhaust gases are cooled under high speed and high load conditions. As the load on and speed of the engine increases, hotter gases are produced and more compressed air from line 1704 is required. If the turbine 1606 does not produce sufficient rotational energy to drive the compressor, such as at low speed and high load, then the engine crankshaft 1612 may provide rotational energy to the compressor 1608 via drive belt 1722, pulley 1718, shaft 1724, the continuously variable gearbox 1716 and the gearbox 1728. Again then any part of the driveline can be used to provide rotational energy to compressor 1608 and Figure 17 shows an embodiment in accordance with a published embodiment.

As also illustrated in Figure 17, a cold start valve 1620 is also connected to the compressed air line 1704, which in turn is connected to the cold start line 1712.

The cold start line 1712 is connected to the input line 1714 for the catalytic filter for diesel particles, which is upstream of the catalytic filter 1616 for diesel particles. The purpose of a cold start valve is to provide compressed inlet air to the inlet of the catalytic filter 1616 under start conditions, as discussed above. Under starting conditions, before the diesel particulate catalytic filter 1616 reaches full operating temperature, extra oxygen is supplied via the cold start line 1712 to initiate the catalytic process. The extra oxygen provided via the cold start line 172 assists in the initiation of the catalytic process. Control unit 1640 controls cold start valve 1620 via control unit control signal 1644 which 43 WO 2011/096936 PCTHJS 2010/23398 responds to engine speed signal 1626, engine load signal 1628 and temperature signal 1630 for gas mixture temperature.

Therefore, the high-efficiency super-turbocharged engine 1600 operates in a manner similar to a super-turbocharger, except that the return valve 1618 provides some of the compressed air from the compressor to the turbine inlet for two reasons. One reason is to cool the exhaust gases before they enter the turbine, so that all the energy in the exhaust gases can be utilized and an overflow valve is not needed under conditions of high speed and high load. The second reason is to provide flow of air through the compressor to prevent surge at low speeds and high loads. In addition, the catalytic filter for diesel particles can be connected to the exhaust stream before the exhaust gases reach the turbine so that the heat generated by the catalytic filter 1616 for diesel particles can be used to drive the turbine 1606, and expand the compressed intake air mixed with the hot gases from the catalytic filter 1616 for diesel particles, which greatly increases the efficiency of the system. In addition, the cold start valve 1620 can be used to initiate the catalytic process in the catalytic filter 1616 for diesel particles by supplying oxygen to the exhaust gases under start conditions.

Therefore, a unique super-turbocharger using a high speed friction drive with a fixed gear ratio that reduces the mechanical rotational energy of a turbine / compressor shaft to a speed that can be used by a continuously variable gearbox that couples the energy between a driveline and turbine compressor shaft. One of the unique features of the design of the super-turbocharger is that the gearbox is located within the system. The continuously variable gearbox is located in the lower part of the housing for the super-turbocharger. The continuously variable gearbox 1116 provides the infinitely variable gears needed to transfer mechanical rotational energy between the super-turbocharger and the engine. Either a continuously variable gearbox with gears can be used as the continuously variable gearbox 1116 or a friction drive unit which is continuously variable can be used. Therefore, friction drives can be used for both the high speed friction drive 114 and for the continuously variable gearbox 1116. 44 WO 2011/096936 PCTHJS 2010/23398

The above description of the invention has been presented for the purpose of illustrating and describing. It is not intended to be exhaustive or to limit the invention to exactly the forms disclosed, and other modifications and variations may be possible in light of the above teachings. The embodiment was chosen and described in order to best explain the principle of the invention and its practical application, thereby enabling others skilled in the art how best to use the invention in various embodiments and various modifications suitable for the particular use contemplated.

It is intended that the appended claims be construed as including other alternative embodiments of the invention except to the extent limited by prior art. 45

Claims (33)

WO 2011/096936 PCTHJS 2010/23398 PATENT CLAIMS 1. A super-turbocharger coupled to an engine, comprising: a turbine which generates the mechanical energy of the turbine rotation from enthalpy of exhaust gases produced by said engine; a compressor compressing the inlet air and providing compressed air to said engine in response to said turbine rotation mechanical energy generated by said turbine and the engine rotation mechanical energy transmitted from said engine; a shaft having end portions coupled to said turbine and said compressor, and a central portion having a traction surface on the shaft; a friction drive unit located around said central portion of said shaft, comprising: a plurality of planetary rollers having a plurality of planetary roller traction surfaces that interact with said traction surface of said shaft so that a first plurality of traction couplings are present between said plurality of planetary roller traction surfaces and said shaft traction surface; a ring roller rotated by said planet number of planetary rollers via a second plurality of traction couplings; a continuously variable gearbox mechanically coupled to said friction drive unit and said engine, which transmits the mechanical energy of the turbine rotation to said engine and the mechanical energy of the engine rotation to said super-turbocharger at operating speed of said engine. A super-turbocharger according to claim 1 wherein said continuously variable gearbox comprises a friction drive unit with continuously variable gearbox. The super-turbocharger of claim 2 wherein said continuously variable gearbox comprises a planetary gear unit with ball bearings and friction drive unit which is a continuously variable gearbox. WO 2011/096936 PCTHJS 2010/23398 The super-turbocharger according to claim 2, wherein said friction drive unit comprises a planetary gear unit with friction drive unit having at least two planetary rollers. The super-turbocharger according to claim 4, wherein said planetary gear with friction drive unit has at least three planetary rollers. A super-turbocharger according to claim 4, wherein said planetary gear with friction drive has a planet carrier on which the planetary rollers are mounted. A super-turbocharger according to claim 6, wherein said planetary gear unit with friction drive unit has planetary rollers with fl your diameters. The super-turbocharger according to claim 6, wherein said ring roller has a ring roller traction surface that interacts with said number of traction surfaces on planetary rollers to create said second number of traction couplings. The super-turbocharger according to claim 7, wherein said ring roller has a ring roller traction surface that interacts with a number of additional traction surfaces on planetary rollers having a diameter smaller than the traction surfaces of said number of planetary rollers to create said number of traction surfaces. A method of transferring mechanical rotational energy between a super-turbocharger and an engine, comprising: generating the mechanical energy of the turbine rotation in an turbine from enthalpy of exhaust gases produced by said engine; compressing inlet air using a compressor to provide compressed air to said engine in response to said turbine rotation mechanical energy generated by said turbine and the engine rotation mechanical energy generated by the engine; providing a shaft having end portions coupled to said turbine and said compressor, and a central portion having a traction surface on the shaft; mechanical coupling of a friction drive unit to the traction surface of said shaft; WO 2011/096936 PCTHJS 2010/23398 placing a plurality of planetary roller traction surfaces on a plurality of planetary roller contacts with the traction surface of said shaft so that a plurality of first traction couplings are created between said plural planetary rollers and said traction surface on the shaft; placing a ring roller in contact with said plurality of planetary rollers so that a number of other traction couplings are created between said number of planetary rollers and said ring roller; mechanical coupling of a continuously variable gearbox to said friction drive unit and said engine to transmit the mechanical energy of the turbine rotation to said engine at operating speed of said engine and the mechanical energy of the engine rotation to said shaft at operating speed of said compressor and said turbine. The method of claim 10 wherein said process of transmitting mechanical rotational energy between said super-turbocharger and said engine comprises transmitting mechanical rotational energy via at least one mechanical device. A method according to claim 11, wherein said process of transmitting mechanical rotational energy through at least one mechanical device comprises transmitting mechanical rotational energy through a gearbox in a vehicle. The method of claim 11 wherein said process of transmitting mechanical rotational energy through at least one mechanical device comprises transmitting mechanical rotational energy to a driveline of a vehicle. The method of claim 10 wherein said process of placing said ring roller in contact with said fl number of planetary rollers comprises: placing a ring roller 'traction surface from said ring roller in contact with said fl number of traction surfaces on planetary rollers to create said fl number of other traction couplings. The method of claim 10 wherein said process of placing said ring roller in contact with said planetary rollers comprises: WO 2011/096936 PCTHJS 2010/23398 placing a ring roller 'traction surface from said ring roller in contact with a number of additional traction surfaces on planetary rollers having a diameter which is smaller than said number of traction surfaces on planetary gears, in order to create a number of other traction couplings. The method of claim 10, wherein said process of mechanically coupling a continuously variable gearbox to said friction drive unit comprises: mechanically coupling a continuously variable gearbox having a friction drive unit to said friction drive unit. 17. l7. The method of claim 16 wherein said process of mechanically coupling a continuously variable gearbox with friction drive unit to said friction drive unit comprises: mechanically coupling a continuously variable gearbox with planetary ball bearings to said friction drive unit. The method of claim 16, wherein said process of mechanically coupling a friction drive unit to the traction surface of said shaft, comprising: mechanically coupling a planetary friction drive unit having at least three planetary rollers having a plurality of diameters. A method of enabling exhaust gas recirculation in a super-turbocharged explosion engine comprising: providing a high pressure exhaust outlet having a first predetermined size in said explosion engine; providing a low pressure exhaust outlet having a second predetermined size in said explosion engine, said second predetermined size substantially larger than said first predetermined size; operating a high pressure super-turbocharger with at least a first portion of high pressure exhaust gases from said high pressure exhaust outlet; WO 2011/096936 PCTHJS 2010/23398 to supply at least a second part of said exhaust gases at high pressure from said high pressure exhaust outlet to an inlet manifold in said explosion engine. operating a low pressure super-turbocharger with exhaust gases at lower pressure from said low pressure exhaust outlet; providing compressed air from the outlet of said low pressure compressor to an air inlet of said high pressure compressor; providing compressed air from the outlet of said high pressure compressor, at a predetermined pressure, to an inlet manifold for said explosion engine; opening said high pressure exhaust outlet while the pressure in said high pressure exhaust outlet is greater than said second portion of said high pressure exhaust gases so that said second portion of said high pressure exhaust gases recirculates through said explosion engine. The method of claim 19 further comprising: controlling said amount of said second portion of said high pressure exhaust gases with respect to said first portion of said high pressure exhaust gases using a valve located in a conduit providing said second portion of said exhaust gases with high pressure to said inlet manifold. A method of enabling exhaust gas recirculation in a super-turbocharged explosion engine comprising: providing a high pressure exhaust outlet having a first predetermined size in said explosion engine; providing a low pressure exhaust outlet having a second predetermined size in said explosion engine, said second predetermined size substantially larger than said first predetermined size; WO 2011/096936 PCTHJS 2010/23398 operating a high pressure super-turbocharger with high pressure exhaust gases from said high pressure exhaust outlet; operating a low pressure super-turbocharger with lower pressure exhaust gases from said low pressure exhaust outlet; providing compressed air from the outlet of said low pressure compressor to an air inlet for said high pressure compressor; providing compressed air from the outlet of said high pressure compressor at a predetermined pressure, to an inlet manifold in said explosion engine; directing said high pressure exhaust gas from an outlet on said high pressure super-turbocharger to an inlet manifold on said explosion engine; opening said high pressure exhaust outlet while the pressure in said high pressure exhaust outlet is greater than said predetermined pressure so that said high pressure exhaust gases from said outlet of said super-turbocharger recirculate through said explosion engine. The method of claim 21, further comprising: providing a portion of said high pressure exhaust gases from said outlet of said high pressure super-turbocharger to said lower pressure exhaust gases to assist in operating said low pressure super-turbocharger. A method of enabling exhaust gas recirculation in a super-turbocharged explosion engine comprising: providing a high pressure exhaust outlet having a first predetermined size in said explosion engine; providing a low pressure exhaust outlet having a second predetermined size in said explosion engine, said second predetermined size substantially larger than the first predetermined size; WO 2011/096936 PCTHJS 2010/23398 to provide high pressure exhaust gases from said high pressure exhaust outlet to an inlet manifold in said explosion engine; operating a low pressure super-turbocharger with lower pressure exhaust gases from said low pressure exhaust outlet; providing compressed air from an outlet of said low pressure compressor, at a predetermined pressure, to an inlet manifold in said explosion engine; opening said high pressure outlet while the pressure in said high pressure outlet is higher than said predetermined pressure so that said second portion of said high pressure exhaust gases recirculates through said explosion engine. The method of claim 23 further comprising: providing a portion of said high pressure exhaust gases to said low pressure exhaust gases to assist in operating said low pressure super-turbocharger. A method of improving the efficiency of a turbo-supercharged engine system comprising: providing an engine; providing a catalytic filter for diesel particles connected to an exhaust outlet near said engine, which receives engine exhaust gases from said engine, which activates an exothermic reaction in said catalytic filter for diesel particles which adds additional energy to said engine exhaust gases and produces catalyzed exhaust gases at the outlet of said catalytic filter for diesel particles, gases hotter than said engine exhaust gases; providing a fate of compressed air to an inlet of said motor using a compressor; mixing a portion of said compressed air with said catalyzed exhaust gases in a mixing chamber downstream of said catalytic filter for WO 2011/096936 PCTHJS 2010/23398 diesel particles to produce a gas mixture of said catalyzed exhaust gases and said compressed air; controlling said fate of said compressed air into said mixing chamber using a control valve to keep said gas mixture below a maximum temperature and maintain a fate of said compressed air through said compressor during the operational phases of said engine when a surge in said compressor otherwise would arise; providing said gas mixture to a turbine which produces the mechanical energy of the turbine rotation as a result of the fate of said gas mixture; sending the mechanical energy of said turbine rotation from said turbine to said compressor using said mechanical energy of turbine rotation to compress air to produce said compressed air when said fate of said gas mixture through said turbine is sufficient to drive said compressor; extracting at least a portion of said mechanical energy of turbine rotation from said turbine and applying said portion of said mechanical energy of turbine rotation to a driveline when said portion of said mechanical energy of turbine rotation is not needed from said turbine to drive said compressor; to provide the mechanical energy of the driveline rotation from said driveline to said compressor to prevent turbo delay when said flow of gas mixture through said turbine is not sufficient to drive said compressor. The method of claim 25 wherein said maximum temperature of said gas mixture is lower than the temperature at which said gas mixture would otherwise cause damage to said turbine. The method of claim 26 wherein said maximum temperature of said gas mixture is lower than about 950 ° C. WO 2011/096936 PCTHJS 2010/23398 The method of claim 26 wherein the efficiency of said engine is improved by not using an over-fl valve to discharge over fl excess gases of said gas mixture. The method of claim 28, wherein said process of extracting excess mechanical energy from turbine rotation from said turbine and providing mechanical rotation energy from said driveline to said compressor comprises: using a gearbox that couples said turbine rotation over unnecessary mechanical energy and said driveline mechanical energy between said turbine rotation. driveline and a shaft connecting said turbine and said compressor. The method of claim 29 wherein said process of maintaining a flow of compressed air during operational phases of said engine comprises: maintaining a flow of said compressed air through said compressor when said engine is operating at low speed and requires high torque by open said return valve to reduce surge. A method according to claim 30 wherein said process of mixing said compressed air with said catalyzed exhaust gases in a mixing chamber comprises: providing at least one opening in an exhaust line which is connected to a line of compressed air so that said compressed air fl surfaces through said at least one opening and mixes with said hotter exhaust gases in said exhaust line. The method of claim 31 further comprising: mixing a portion of said compressed air with said exhaust gases upstream of said diesel particulate catalytic filter at cold start of said engine to provide oxygen to said diesel particulate catalytic filter assisting said diesel particulate catalytic filter to start said exothermic reaction. A super-turbocharged engine system comprising: an engine; WO 2011/096936 PCTHJS 2010/23398 a catalyzed filter for diesel particles connected to an exhaust line near an exhaust outlet from said engine so that hot exhaust gases from said engine activate an exothermic reaction in said catalytic filter for diesel particles, which adds energy to said hot exhaust gases and produces catalyzed exhaust gases; a compressor coupled to a source of air, which provides compressed air having a pressure greater than the pressure level of said exhaust gases; a conduit supplying said compressed air to said catalyzed exhaust gases so that at least a portion of said compressed air is mixed with said catalyzed exhaust gases to produce a gas mixture; a turbine mechanically coupled to said compressor and generating the mechanical energy of the turbine rotation from said gas mixture; a valve which controls the flow of said portion of said compressed air through said conduit to maintain said gas mixture below a predetermined maximum temperature and to maintain a flow of air from said air source through said compressor during operational phases of said engine when a surge in said compressor would otherwise occur; a gearbox which provides the mechanical energy of the driveline rotation from a driveline to said compressor to reduce turbo delay when said fate of said exhaust gases through said turbine is not sufficient to propel said compressor to a desired