Note: Descriptions are shown in the official language in which they were submitted.
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
SUPER-TURBOCHARGER HAVING A HIGH SPEED TRACTION DRIVE AND
A CONTINUOUSLY VARIABLE TRANSMISSION
CROSS-REFERENCE TO RELATED APPLICATION
[00011 This patent application is a continuation-in-part of U.S. Application
Serial No.
12/536,421, filed August 5, 2009, which application claims the benefit of U.S.
Provisional
Patent Application Serial No. 61/086,401, filed August 5, 2008, the entire
teachings and
disclosure of which are incorporated by reference thereto.
BACKGROUND OF THE INVENTION
[00021 Conventional turbochargers are driven by waste exhaust heat and gases,
which are
forced through an exhaust turbine housing onto a turbine wheel. The turbine
wheel is
connected by a common turbo-shaft to a compressor wheel. As the exhaust gases
hit the
turbine wheel, both wheels simultaneously rotate. Rotation of the compressor
wheel draws
air in through a compressor housing, which forces compressed air into the
engine cylinder to
achieve improved engine performance and fuel efficiency. Turbochargers for
variable
speed/load applications are typically sized for maximum efficiency at torque
peak speed in
order to develop sufficient boost to reach peak torque. However, at lower
speeds, the
turbocharger produces inadequate boost for proper engine transient response.
[0003] To overcome these problems and provide a system that increases
efficiency, a super-
turbocharger can be used, which combines the features of a supercharger and a
turbocharger.
Super-turbochargers merge the benefits of a supercharger, which is primarily
good for high
torque at low speed, and a turbocharger, which is usually only good for high
horsepower at
high speeds. A super-turbocharger combines a turbocharger with a transmission
that can put
engine torque onto the turbo shaft for supercharging and elimination of turbo
lag. Once the
1
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
exhaust energy begins to provide more work than it takes to drive the
compressor, the super-
turbocharger recovers the excess energy by applying the additional power to
the piston
engine, usually through the crankshaft. As a result, the super-turbocharger
provides both the
benefits of low speed with high torque and the added value of high speed with
high
horsepower all from one system.
SUMMARY OF THE INVENTION
[0004] An embodiment of the present invention may therefore comprise a super-
turbocharger
that is coupled to an engine comprising: a turbine that generates turbine
rotational
mechanical energy from enthalpy of exhaust gas produced by the engine; a
compressor that
compresses intake air and supplies compressed air to the engine in response to
the turbine
rotational mechanical energy generated by the turbine and engine rotational
mechanical
energy transferred from the engine; a shaft having end portions that are
connected to the
turbine and the compressor, and a central portion having a shaft traction
surface; a traction
drive disposed around the central portion of the shaft, the traction drive
comprising: a
plurality of planetary rollers having a plurality of planetary roller traction
surfaces that
interface with the shaft traction surface so that a first plurality of
traction interfaces exist
between the plurality of planetary roller traction surfaces and the shaft
traction surface; a ring
roller that is rotated by the plurality of planet rollers through a second
plurality of traction
interfaces; a continuously variable transmission, that is mechanically coupled
to the traction
drive and the engine, that transfers turbine rotational mechanical energy to
the engine and
engine rotational mechanical energy to the super-turbocharger at operating
speeds of the
engine.
[0005] An embodiment of the present invention may further comprise a method of
transferring rotational mechanical energy between a super-turbocharger and an
engine
2
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
comprising: generating turbine rotational mechanical energy in a turbine from
enthalpy of
exhaust gas produced by the engine; compressing intake air to supply
compressed air to the
engine in response to the turbine rotational mechanical energy generated by
the turbine and
engine rotational mechanical energy generated by the engine; providing a shaft
having end
portions that are connected to the turbine and the compressor, and a central
portion having a
shaft traction surface; mechanically coupling a traction drive to the shaft
traction surface of
the shaft; placing a plurality of planetary roller traction surfaces of a
plurality of planetary
rollers in contact with the shaft traction surface so that a plurality of
first traction interfaces
are created between the plurality of planetary roller traction surfaces and
the shaft traction
surface; placing a ring roller in contact with the plurality of planetary
rollers so that a
plurality of second traction interfaces are created between the plurality of
planet rollers and
the ring roller; mechanically coupling a continuously variable transmission to
the traction
drive and the engine to transfer the turbine rotational mechanical energy to
the engine and
engine rotational mechanical energy to the super-turbocharger at operating
speeds of the
engine.
[00061 An embodiment of the present invention may further comprise a method of
facilitating exhaust gas recirculation in a super-turbocharged internal
combustion engine
comprising: providing a high pressure exhaust port of a first predetermined
size in the
internal combustion engine; providing a low pressure exhaust port of a second
predetennined
size in the internal combustion engine, the second predetermined size being
substantially
larger than the first predetermined size; driving a high pressure super-
turbocharger with a
least a first portion of high pressure exhaust gases from the high pressure
exhaust port;
providing at least a second portion of the high pressure exhaust gases from
the high pressure
exhaust port to an intake manifold of the internal combustion engine; driving
a low pressure
super-turbocharger with lower pressure exhaust gases from the low pressure
exhaust port;
3
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
providing compressed air from an output of the low pressure compressor to an
air input of the
high pressure compressor; providing compressed air from an output of the high
pressure
compressor, at a predetermined pressure, to an intake manifold of the internal
combustion
engine; opening the high pressure exhaust port while pressure in the high
pressure exhaust
port is greater than the predetermined pressure so that the second portion of
the high pressure
exhaust gases recirculate through the internal combustion engine.
[00071 An embodiment of the present invention may further comprise a method of
facilitating exhaust gas recirculation in a super-turbocharged internal
combustion engine
comprising: providing a high pressure exhaust port of a first predetermined
size in the
internal combustion engine; providing a low pressure exhaust port of a second
predetermined
size in the internal combustion engine, the second predetermined size being
substantially
larger than the first predetermined size; driving a high pressure super-
turbocharger with high
pressure exhaust gases from the high pressure exhaust port; driving a low
pressure super-
turbocharger with lower pressure exhaust gases from the low pressure exhaust
port; providing
compressed air from an output of the low pressure compressor to an air input
of the high
pressure compressor; providing compressed air from an output of the high
pressure
compressor, at a predetermined pressure, to an intake manifold of the internal
combustion
engine; channeling the high pressure exhaust gases from an output of the high
pressure super-
turbocharger to an intake manifold of the internal combustion engine; opening
the high
pressure exhaust port while pressure in the high pressure exhaust port is
greater than the
predetermined pressure so that the high pressure exhaust gases from the output
of the high
pressure super-turbocharger recirculate through the internal combustion
engine.
[00081 An embodiment of the present invention may further comprise a method of
facilitating exhaust gas recirculation in a super-turbocharged internal
combustion engine
comprising: providing a high pressure exhaust port of a first predetermined
size in the
4
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
internal combustion engine; providing a low pressure exhaust port of a second
predetermined
size in the internal combustion engine, the second predetermined size being
substantially
larger than the first predetermined size; providing high pressure exhaust
gases from the high
pressure exhaust port to an intake manifold of the internal combustion engine;
driving a low
pressure super-turbocharger with lower pressure exhaust gases from the low
pressure exhaust
port; providing compressed air from an output of the low pressure compressor,
at a
predetermined pressure, to an intake manifold of the internal combustion
engine; opening the
high pressure exhaust port while pressure in the high pressure exhaust port is
greater than the
predetermined pressure so that the second portion of the high pressure exhaust
gases
recirculate through the internal combustion engine.
BRIEF DESCRIPTION OF THE DRAWINGS
[0009] Figure 1 is a side view illustration of an embodiment of a super-
turbocharger.
[0010] Figure 2 is a transparency isometric view of the embodiment of the
super-
turbocharger of Figure 1.
[0011] Figure 3A is a side transparency view of an embodiment of the super-
turbocharger
illustrated in Figures 1 and 2.
[0012] Figure 3B is a side cutaway view of another embodiment of a super-
turbocharger.
[0013] Figure 3C is a side transparency view of modifications to the
embodiment of the
super-turbocharger illustrated in figures 1, 2 and 3A.
[0014] Figures 4-9 are various drawings of a super-turbocharger using an
embodiment of a
multi-diameter planetary roller traction drive.
[0015] Figure 10 is an illustration of another embodiment of a high speed
traction drive.
[0016] Figures 11 and 12 are illustrations of an embodiment of a traction
continuously
variable transmission.
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0017] Figure 13 is a side cutaway view of another embodiment.
[0018] Figure 14A is a schematic view of an embodiment of a super-turbocharged
gas
recirculation device.
[0019] Figure 14B is a schematic view of another embodiment of a super-
turbocharged gas
recirculation device.
[0020] Figure 14C is a schematic view of another embodiment of a super-
turbocharged gas
recirculation device.
[0021] Figure 14D is a graph of valve lift, flow rate and cylinder pressure
versus piston
position for the embodiments of Figures 14A-C.
[0022] Figure 14E is a PV graph of cylinder pressure versus cylinder volume
for the
embodiments of Fgiures 14A-C.
[0023] Figure 15 is a graphical illustration of simulated BSFC improvement.
DETAILED DESCRIPTION OF THE INVENTION
[0024] Figure 1 is a schematic illustration of an embodiment of a super-
turbocharger 100 that
uses a high speed traction drive 114 and a continuously variable transmission
116. As shown
in Figure 1, the super-turbocharger 100 is coupled to the engine 101. The
super-turbocharger
includes a turbine 102 which is coupled to engine 101 by an exhaust conduit
104. The
turbine 102 receives the hot exhaust gases from the exhaust conduit 104 and
generates
rotational mechanical energy prior to exhausting the exhaust gases in an
exhaust outlet 112.
A catalyzed diesel particulate filter (not shown) can be connected between the
exhaust
conduit 104 and turbine 102. Alternatively, the catalyzed diesel particulate
filter (not shown)
can be connected to the exhaust outlet 112. The rotational mechanical energy
generated by
6
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
the turbine 102 is transferred to the compressor 106 via a turbine/compressor
shaft, such as
shaft 414 of Figure 4, to rotate a compressor fan disposed in the compressor
106, which
compresses the air intake 110 and transmits the compressed air to a conduit
108, which is
coupled to an intake manifold (not shown) of the engine 101. As disclosed in
the above
referenced application, super-turbochargers, unlike turbochargers, are coupled
to a propulsion
train to transfer energy to and from the propulsion train. The propulsion
train, as referred to
herein, may comprise the engine 101, the transmission of a vehicle in which
the engine 101 is
disposed, the drive train of a vehicle in which the engine 101 is disposed, or
other
applications of the rotational mechanical energy generated by engine 101. In
other words,
rotational mechanical energy can be coupled or transferred from the super-
turbocharger to the
engine through at least one intermediate mechanical device such as a
transmission or drive
train of a vehicle, and vice versa. In the embodiment of Figure 1, the
rotational mechanical
energy of the super-turbocharger is coupled directly to a crankshaft 122 of
engine 101
through a shaft 118, a pulley 120 and a belt 124. As also illustrated in
Figure 1, a high speed
traction drive 114 is mechanically coupled to a continuously variable
transmission 116.
[0025] In operation, the high speed traction drive 114, of Figure 1, is a
fixed ratio, high speed
traction drive that is mechanically coupled to the turbine/compressor shaft
using a traction
interface to transfer rotational mechanical energy to and from the
turbine/compressor shaft.
The high speed traction drive 114 has a fixed ratio which may differ in
accordance with the
size of the engine 101. For small engines, a large fixed ratio of the high
speed traction drive
114 is required.
[0026] For smaller engines, the compressor and turbine of a super-turbocharger
must
necessarily be smaller to maintain a small engine size and to match the flow
requirements of
the compressor and turbine. In order for a smaller turbine and a smaller
compressor to
function properly, they have to spin at a higher rpm. For example, smaller
engines may
7
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
require the compressor and turbine to spin at 300,000 rpm. For very small
engines, such as
half liter engines, the super-turbocharger may need to spin at 900,000 rpm.
One of the
reasons that smaller engines require compressors that operate at a higher rpm
level is to avoid
surge. In addition, to operate in an efficient manner, the tip velocity of the
compressor must
be just short of the speed of sound. Since the tips are not as long in smaller
compressors, the
tips of a smaller compressor are not moving as fast as the tips on larger
compressors at the
same rpm. As the size of the compressor decreases, the rotational speed
required to operate
efficiently goes up exponentially. Since gears are limited to approximately
100,000 rpm,
standard gear systems cannot be used to achieve the power take off at the
higher speeds
necessary for a car engine super-turbocharger. Therefore, various embodiments
use a high
speed traction drive 114 to add and receive power from the turbo shaft.
100271 The rotational mechanical energy from the high speed traction drive 114
is therefore
reduced to an rpm level that is variable depending upon the rotational speed
of the
turbine/compressor, but at an rpm level that is within the operating range of
the continuously
variable transmission (CVT) 116. For example, the high speed traction drive
114 may have
an output that varies between zero and 7,000 rpm while the input from the
turbine/compressor shaft may vary from zero to 300,000 rpm, or greater. The
continuously
variable transmission 116 adjusts the rpm level of the high speed traction
drive 114 to the
rpm level of the crankshaft 122 and pulley 120 to apply rotational mechanical
energy to
engine 101, or extract rotational mechanical energy from engine 101 at the
proper rpm level.
In other words, the continuously variable transmission 116 comprises an
interface for
transferring rotational mechanical energy between engine 101 and the high
speed traction
drive 114 at the proper rpm level which varies in accordance with the engine
rotational speed
and the turbine/compressor rotational speed. Continuously variable
transmission 116 can
comprise any desired type of continuously variable transmission that can
operate at the
8
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
required rotational speeds and have a ratio to match the rotational speed of
the crankshaft 122
or other mechanisms coupled, directly or indirectly, to the engine 101. For
example, in
addition to the embodiments disclosed herein, two roller CVTs can be used as
well as traction
ball drives and pushing steel belt CVTs.
[0028] An example of a continuously variable transmission that is suitable for
use as
continuously variable transmission 116, disclosed in Figure 1, is the
continuously variable
transmission disclosed in Figures 11 and 12. Other examples of continuously
variable
transmissions that can be used as the continuously variable transmission 116
of Figure 1
include U.S. Patent Serial No. 7,540,881 issued June 2, 2009, to Miller et al.
The Miller
patent is an example of a traction drive, continuously variable transmission
that uses a
planetary ball bearing. The traction drive of Miller is limited to about
10,000 rpm so that the
Miller continuously variable transmission is not usable as a high speed
traction drive, such as
high speed traction drive 114. However, the Miller patent does disclose a
continuously
variable transmission that uses a traction drive and is suitable for use as an
example of a
continuously variable transmission that could be used as continuously variable
transmission
116 as illustrated in Figures 1-3. Another example of a suitable continuously
variable
transmission is disclosed in U.S. Patent Serial No. 7,055,507 issued June 6,
2006, to William
R. Kelley, Jr., and assigned to Borg Warner. Another example of a continuously
variable
transmission is disclosed in U.S. Patent Serial No. 5,033,269 issued July 23,
1991 to Smith.
Further, U.S. Patent No. 7,491,149 also discloses a continuously variable
transmission that
would be suitable for use as continuously variable transmission 116. U.S.
Patent No.
7,491,149 issued February 17, 2009 to Greenwood et al. and assigned to
Torotrak Limited
discloses an example of a continuously variable transmission that uses a
traction drive that
can be used as the continuously variable transmission 116. All of these
patents are
specifically incorporated by reference for all that they disclose and teach.
European
9
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
Application No. 92830258.7, published August 9, 1995, as Publication No.
0517675B1, also
illustrates another continuously variable transmission 3 that is suitable for
use as the
continuously variable traction drive 116.
[0029] Various types of high speed traction drives can be used as the high
speed traction
drive 114. For example, the high speed planetary traction drive 406 disclosed
in Figures 4-9
and the high speed planetary drive of Figure 10 can be used as high speed
traction drive 114.
[0030] Examples of high speed drives that use gears are disclosed in U.S.
Patent No.
2,397,941 issued April 9, 1946 to Birgkigt and U.S. Patent No. 5,729,978
issued March 24,
1998 to Hiereth et al. Both of these patents are specifically incorporated
herein by reference
for all that they disclose and teach. Both of these references use standard
gears and do not
use traction drives. Hence, even with highly polished, specially designed
gearing systems,
the gears in these systems are limited to rotational speeds of approximately
100,000 rpm or
less. U.S. Patent No. 6,960,147 issued November 1, 2005 to Kolstrup and
assigned to
Rulounds Roadtracks Rotrex A/S discloses a planetary gear that is capable of
producing gear
ratios of 13:1. The planetary gear of Kolstrup is an example of a high speed
drive that could
be used in place of a high speed traction drive 114 of Figure 1. U.S. Patent
No. 6,960,147 is
also specifically incorporated herein by reference for all that it disclosed
and teaches.
[0031] Figure 2 is a schematic side transparency view of the super-
turbocharger 100. As
shown in Figure 2, turbine 102 has an exhaust conduit 104 that receives
exhaust gases that
are applied to the turbine fan 130. Compressor 106 has a compressed air
conduit 108 that
supplies compressed air to the intake manifold. Compressor housing 128
encloses the
compressor fan 126 and is coupled to the compressed air conduit 108. As
disclosed above,
high speed traction drive 114 is a fixed ratio traction drive that is coupled
to a continuously
variable transmission 116. The continuously variable transmission 116 drives
shaft 118 and
pulley 120.
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0032] Figure 3A is a side transparency 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 fan 130, while compressor 106 includes a compressor fan 126. A shaft
(not shown)
connecting the turbine fan 130 and compressor fan 126 is coupled to a high
speed traction
drive 114. Rotational mechanical energy is transferred from the high speed
traction drive 114
to a transfer gear 132 that transfers the rotational mechanical energy to a
CVT gear 134 and
the continuously variable transmission (CVT) 116. The continuously variable
transmission
116 is coupled to the shaft 118 and pulley 120.
[0033] Figure 3B is a schematic cutaway view of another example of a super-
turbocharger
300 that is coupled to an engine 304. As shown in Figure 3B, the turbine 302
and the
compressor 306 are mechanically coupled by shaft 320. High speed traction
drive 308
transfers rotational mechanical energy to, and receives rotational mechanical
energy from,
transfer gear 322. A specific example of a high speed traction drive 308 is
illustrated in
Figure 3B. Transfer gear 322 transfers rotational mechanical energy between
the traction
drive 308 and the continuously variable transmission 310. A specific example
of a
continuously variable transmission 310 is also illustrated in Figure 3B. Shaft
312, pulley 314
and belt 316 transfer rotational mechanical energy between the crankshaft 318
and the
continuously variable transmission 310.
[0034] Figure 3C is a side schematic cutaway view of modifications to the
embodiment of
the super-turbocharger 100 illustrated in Figures 1, 2 and 3A. As shown in
Figure 3C,
turbine 102 and compressor 106 are coupled together by a shaft (not shown).
High speed
traction device 114 is coupled to the shaft. Rotational mechanical energy is
transferred from
the high speed traction device 114 to a transfer gear 132 that transfers the
rotational
mechanical energy to transmission gear 134. High speed traction drive 114,
transfer gear 132
and transmission gear 134 may all be housed in the same housing. Transmission
gear 134 is
11
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
connected to a transmission 140 that can comprise a manual gear box, a CVT, a
straight shaft,
an automatic gear box, or a hydraulic transmission. Transmission 140 is then
connected to a
shaft 118 which is connected to a pulley 120. Pulley 120 is coupled to the
propulsion train.
In an alternative embodiment, pulley 120 is coupled to an electric
notor/generator 142.
[0035] Figure 4 is a schematic transparency view of another embodiment of
super-
turbocharger 400 that utilizes a high speed traction drive 416 that is coupled
to a continuously
variable transmission 408. As shown in Figure 4, the turbine 404 is
mechanically coupled to
the compressor 402 with a compressor/turbine shaft 414. Rotational mechanical
energy is
transferred between the compressor/turbine shaft 414 and the multi-diameter
traction drive
416 in the manner disclosed in more detail below. Transfer gear 418 transfers
rotational
mechanical energy between the multi-diameter traction drive 416 and the CVT
gear 420 of
the continuously variable transmission 408. Shaft 410 and pulley 412 are
coupled to the
continuously variable transmission 408 and transfer power between the
continuously variable
transmission 408 and a propulsion train.
[0036] Figure 5 is a side cutaway schematic view of the multi-diameter
traction drive 416
that is coupled to the transfer gear 418, which is in turn coupled to the CVT
gear 420. The
compressor/turbine shaft 414 has a polished, hardened surface on a central
portion, as
disclosed in more detail below, that functions as a sun drive in the multi-
diameter traction
drive 416.
[0037] Figure 6 is an exploded view 600 of the embodiment of the super-
turbocharger 400
illustrated in Figure 4. As shown in Figure 6, turbine housing 602 houses a
turbine fan 604.
The hot side cover plate 606 is mounted adjacent to the turbine fan 604 and
the main housing
support 608. A ring seal 610 seals the exhaust at the hot side cover plate
606. Ring roller
bearing 612 is mounted in the ring roller 614. Compressor/turbine shaft 414
extends through
the main housing support 608. The hot side cover plate 606 connects with the
turbine fan
12
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
604. Planet carrier ball bearing 618 is mounted on the planet carrier 620.
Multi-diameter
ring rollers 622 are rotationally connected to the planet carrier 620. Oil
feed tubes 624 are
used to supply traction fluid to the traction surface. Planet carrier 626 is
mounted to the
planet carrier 620 and uses a planet carrier ball bearing 628. Fixed ring 630
is then mounted
outside of planet carrier 626. Cage 632 is mounted between the fixed ring 630
and the cool
side cover plate 636. Compressor fan 638 is coupled to the compressor/turbine
shaft 414.
Compressor housing 640 encloses the compressor fan 638. The main housing
support 608
also supports the continuously variable transmission and the transfer gear
418. Various
bearings 646 are used to mount the transfer gear 418 and the main housing
support 608. The
continuously variable transmission includes a CVT cover 642 and a CVT bearing
plate 644.
CVT gear 420 is mounted inside the main housing support 608 with bearings 650.
CVT
bearing plate 652 is mounted on the opposite side of the CVT gear 420 from the
CVT bearing
plate 644. CVT cover 654 covers the various portions of the CVT device. Shaft
410 is
coupled to the continuously variable transmission. Pulley 412 is mounted on
shaft 410 and
transfers rotational mechanical energy between shaft 410 and a propulsion
train.
[00381 Figure 7 is a perspective view of isolated key components of the multi-
diameter
traction drive 416, as well as the turbine fan 604 and compressor fan 638. As
shown in
Figure 7, the compressor/turbine shaft 414 is connected to the turbine fan 604
and
compressor fan 638, and passes through the center of the multi-diameter
traction drive 416.
The multi-diameter traction drive 416 includes multi-diameter planet rollers
664, 666 (Figure
9), 668. These multi-diameter planet rollers are rotationally coupled to a
planet carrier 626
(Figure 9). Balls 656, 658, 660, 662 rest on an incline surface for ball ramps
on the fixed ring
630. Ring roller 614 is driven by an inner diameter of the multi-diameter
planet rollers 664,
666, 668, as disclosed in more detail below.
13
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0039] Figure 8 is a side cutaway view of the multi-diameter traction drive
416. As shown in
Figure 8, the compressor/turbine shaft 414 is hardened and polished to form a
traction surface
that is used as a sun roller 674 that has a traction interface 676 with the
multi-diameter planet
roller 664. The multi-diameter planet roller 664 rotates along the multi-
diameter planet roller
axis 672. The multi-diameter planet roller 664 contacts the fixed ring 630 at
the interface 690
of the planet roller 664 and the fixed ring 630. The multi-diameter planet
roller 664 contacts
the ring roller 614 at interface 691, which is a different radial distance
from the multi-
diameter planet roller axis 672, than the interface 691. Figure 8 also
illustrates the planet
carrier 626 and the ball ramp 630 that intersects with the ball 656, and ball
ramp 631 that
intersects with ball 660. The balls 656, 658, 660, 662 are wedged in between a
housing (not
shown) and the ball ramp, such as 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 the rotation of the ring roller 614. This causes the balls to move up the
various ball ramps,
such as ball ramps 630, 631, which in turn causes the fixed ring 630 to press
against the
multi-diameter planet rollers 664, 666, 668. Since the interface 691 of the
planet roller 664
and fixed ring 630 is sloped, and the interface of the planet roller 664 and
ring roller 690 is
sloped, an inward force on the multi-diameter planet roller 664 is generated,
which generates
a force on the traction interface 676 to increase the traction at the traction
interface 676
between the multi-diameter planet roller 664 and the sun roller 674. In
addition, a force is
created at the interface 691 of the multi-diameter planet roller 664 and the
ring roller 614,
which increases traction at interface 691. As also shown in Figure 8, the
compressor fan 630
and the turbine fan 604 are both coupled to the compressor/turbine shaft 414.
Ring roller 614
is coupled to the transfer gear 418, as also shown in Figure 8.
[0040] Figure 9 is a side cutaway view of the multi-diameter traction drive
416. As shown in
Figure 9, the sun roller 674 rotates in a clockwise direction, as shown by
rotation direction
14
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
686. The multi-diameter planet rollers 664, 666, 668 have outer diameter
roller surfaces,
such as outer diameter roller surface 688 of multi-diameter planet roller 664.
These outer
diameter roller surfaces contact the sun roller 674 which cause the multi-
diameter planet
rollers 664, 666, 668 to rotate in a counter-clockwise direction, such as
rotational direction
684 of multi-diameter planet roller 666. The multi-diameter planet rollers
664, 666, 668 also
have an inner diameter roller surface, such as inner roller diameter roller
surface 680 of
multi-diameter planet roller 664. The inner diameter roller surface of each
multi-diameter
planet roller contacts the roller surface 687 of the ring roller 614. Hence,
the interface 678 of
planet roller 664 with the roller surface 687 of ring roller 614 constitutes a
traction interface
that transfers rotational mechanical energy when a traction fluid is applied.
The interface
between each of the multi-diameter planet rollers 664, 666, 668 and sun roller
674 also
constitutes a traction interface that transfers rotational mechanical energy
upon application of
a traction fluid.
[0041] As indicated above with respect to Figures 8 and 9, the fixed ring 630
generates a
force, which pushes the multi-diameter planet rollers 664, 666, 668 towards
the sun roller 674
to generate traction. Each of the multi-diameter planet rollers 664, 666, 668
is rotationally
attached to the planet carrier 626 with planet roller axes, such as the multi-
diameter planet
roller axis 672 of the multi-diameter planet roller 664. These axes have a
slight amount of
play so that the multi-diameter planet rollers 664, 666, 668 can move slightly
and create a
force between the sun roller 674 and the outer diameter roller surface of the
multi-diameter
planet rollers 664, 666, 668, such as the outer diameter of the roller surface
688 of the planet
roller 664. The movement of the multi-diameter planet roller 664 towards the
sun roller 674
also increases the traction at the interface of the multi-diameter planet
rollers 664, 666, 668
and the ring roller 614, since the interface between the multi-diameter planet
rollers 664, 666,
668 and the ring roller 614, such as interface 678, is sloped. The contact
with the multi-
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
diameter planet rollers 664, 666, 668 with the roller surface 687 of ring
roller 614 causes the
planet carrier 626 to rotate in a clockwise direction, such as the rotational
direction 682,
illustrated in Figure 9. As a result, the ring roller 614 rotates in a counter-
clockwise
direction, such as rotational direction 687, and drives the transfer gear 418
in a clockwise
direction.
[0042] Figure 10 is a schematic cross sectional view of another embodiment of
a high speed
traction drive 1000. As shown in Figure 10, a shaft 1002, which is a shaft,
that connects a
turbine and a compressor in super-turbocharger, can act as a sun roller in the
high speed
traction drive 1000. Planet roller 1004 contacts the shaft 1002 at traction
interface 1036.
Planet roller 1004 rotates on an axis 1006 using bearings 1008, 1010, 1012,
1014. As also
shown in Figure 10, gear 1016 is disposed and connected to the outer surface
of the carrier
1018. Carrier 1018 is coupled to a housing (not shown) via bearings 1032,
1034, which
allow the carrier 1018 and gear 1016 to rotate. Fixed rings 1020, 1022 include
ball ramps
1028, 1030, respectively. Ball ramps 1028, 1030 are similar to the ball ramps
630 illustrated
in Figures 7 and 8. As the gear 1016 moves, the balls 1024, 1026 move in the
ball ramps
1028, 1030, respectively, and force the fixed rings 1020, 1022 inwardly
towards each other.
A force is created between the fixed rings 1020, 1022 and the surface of the
planet roller
1004 at traction surfaces 1038, 1040 as the balls 1024, 1026 force the fixed
ramps 1020, 1022
inwardly towards each other. The force created by the fixed rings 1020, 1022
also forces the
planet roller 1004 downwardly, as illustrated in Figure 10, so that a force is
created between
the shaft 1002 and the planet roller 1004 at the traction interface 1036. As a
result, greater
traction is achieved at a traction interface 1036 and the traction surfaces
1038, 1040.
Traction fluid is applied to these surfaces, which becomes sticky and
increases friction at the
traction interfaces, as the traction fluid is heated as a result of the
friction created at the
traction interfaces 1036, 1038, 1040.
16
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0043] The high speed traction drive 1000, illustrated in Figure 10, is
capable of rotating at
high speeds in excess of 100,000 rpm, which is unachievable by gearing
systems. For
example, the high speed traction drive 1000 may be able to rotate at speeds
greater than
300,000 rpm. However, high speed traction drive 1000 is limited to a gear
ratio of
approximately 10:1 because of the physical limitations of size. The high speed
traction drive
1000 may utilize three planet rollers, such as planet roller 1006 that are
disposed radially
around the shaft 1002. As illustrated in Figure 9, the size of the planet
rollers is limited with
respect to the sun roller. If the diameter of the planet rollers in Figure 9
increases, the planet
rollers will abut each other. Hence, gear ratios of only about 10:1 can be
reached with a
planetary traction drive, such as illustrated in Figure 10, while the multi-
diameter planet
drives that are connected to a planet carrier, such as illustrated in Figures
7-9, may have ratios
of as much as 47:1 or greater. Accordingly, if a compressor is required for a
smaller engine
that must rotate at 300,000 rpm to be efficient, a 47:1 ratio traction drive,
such as illustrated
in Figures 7-9, can reduce the maximum rotational speed of 300,000 rpm to
approximately
6,400 rpm. Standard geared or traction continuously variable transmissions can
then be used
to transfer the rotational mechanical energy between the high speed traction
drive and the
propulsion train of the engine.
[0044] As disclosed above, the high speed traction drive 1000, illustrated in
Figure 10, may
have a ratio as large as 10:1. Assuming a rotational speed of the shaft 1002
is 300,000 rpm
for a super-turbocharger for a small engine, the 300,000 rpm rotational speed
of the shaft can
be reduced to 30,000 rpm at gear 1016. Various types of continuously variable
transmissions
116 can be used that operate up to 30,000 rpm using standard gearing
techniques. Traction
drive continuously variable transmissions, such as the traction drive
continuously variable
transmission illustrated in Figures 11 and 12, can also be used as the
continuously variable
transmission 116, illustrated in Figure 1. Further, ratios of up to 100:1
maybe achievable
17
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
with the multi-diameter traction drive 416, illustrated in Figure 4-9.
Accordingly, small
engines of .5 liters, which may require a compressor that operates at 900,000
rpm, can be
reduced to 9,000 rpm, which is a rotational speed that can be easily utilized
by various
continuously variable transmissions 116 to couple rotational mechanical energy
between a
propulsion train and a turbine/compressor shaft.
[0045] Figures 11 and 12 illustrate an example of a continuously variable
traction drive
transmission that can be used as the continuously variable transmission 116 of
Figure 1. The
traction drive continuously variable transmission illustrated in Figures 11
and 12 operates by
translating races 1116, 1118 in a lateral direction on race surfaces that have
a radius of
curvature that causes contact locations of the ball bearings to move, which,
in turn, causes the
balls to rotate with a different spin angle to drive race 1122 at different
speeds. In other
words, the contact location of each of the bearings on the race surfaces is
changed as a result
of the lateral translation of the races 1116, 1118, which alters the speed at
which the bearings
are rotating at the contact location, as explained in more detail below.
[0046] As shown in Figure 11, input shaft 1102 is coupled to the transfer gear
132 (Figure
3A). For example, splines 1104 may be splined to the CVT gear 134, illustrated
in Figure
3A. Hence, the spline input gear 1104 of the input shaft 1102 can be coupled
to the super-
turbocharger through a high speed traction drive 114, as illustrated in Figure
3A. In this
manner, input torque from the propulsion train is used to drive the spline
input gear 1104 of
the input shaft 1102. The input torque on the spline input gear 1104 imparts a
spin in
rotational direction 1112 on both the input shaft 1102 and its associated
structure including
input race 1114. Input race 1116 is also spun around the axis of rotation 1106
in response to
the torque imparted by spline 1166 from the input shaft 1102 to the input race
1116. The
rotation of the input shaft 1102, input race 1114 and input race 1116 impart a
spin on the
plurality of ball bearings 1132 because the stationary race 1120 impedes the
rotation of the
18
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
ball bearings at the contact point with stationary race 1120. Input race 1114
and input race
1116 rotate at the same angular speed since they are coupled together through
spline 1116.
Input race 1114 and input race 1116 cause the ball bearings 1132 to spin in a
substantially
vertical orientation since the ball bearings 1132 contact the stationary race
1120. The contact
of the ball bearings 1132 against the stationary race 1120 also causes the
ball bearings 1132
to precess around the perimeter of the races 1114, 1116, 1118, 1120. In the
embodiment
illustrated in Figure 11, there may be as many as 20 ball bearings 1132 that
rotate on the
surfaces of the races 1114, 1116, 1118, 1120. The rotation of the ball
bearings 1132 as a
result of being driven by input race 1114 and input race 1116 creates a
tangential contact of
the ball bearings 1132 on the output race 1118. Depending upon the contact
position of the
ball bearings 1132 on the output race 1118, the ratio of the rotational speed
of the input races
1114, 1116 with respect to the output race 1118 can be varied. Output race
1118 is coupled
to output gear 1122. Output gear 1122 engages output gear 1124, which in turn
is connected
to the output shaft 1126.
[0047] The manner in which the traction drive continuously variable
transmission 1100,
illustrated in Figure 11, shifts the ratio between the input shaft 1102 and
the output shaft 1126
is accomplished by changing the relative position of the contact point between
the four races
1114, 1116, 1118, 1120 that are in contact with the ball bearings 1132. The
manner in which
the contact surfaces of the races 1114, 1116, 1118, 1120 with the ball
bearings 1132 is
changed is by shifting the position of the translating clamp 1152. The
translating clamp 1152
is moved horizontally, as illustrated in Figure 11, in response to electric
actuator 1162.
Electric actuator 1162 has a shaft that engages the telescopic shifter 1158
and rotates the
telescopic shifter 1158. Telescopic shifter 1158 has different thread types on
an inside
portion and an outside portion. A difference in thread pitch of the different
thread types
causes the translating clamp 1152 to translate horizontally in response to
rotation of the shaft
19
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
of the electric actuator 1162, which imparts rotation in the telescopic
shifter 1158. Lateral
translation of the translating clamp 1152, which is in contact with bearing
clamp 1164, causes
lateral transition of input race 1116 and output race 1118. Lateral
translation of the input race
1116 and output race 1118 may vary, in the embodiment illustrated in Figure
11, by
approximately one-tenth of an inch. The translation of the input race 1116 and
the output
race 1118 changes the angle of contact between the ball bearings 1132 and the
output race
1118, which changes the ratio, or speed at which the ball bearings 1132 are
moving in the
races because of a change in contact angle between the stationary race 1120
and input race
1114 and input race 1116. The combination of the change in angle between the
races allows
the contact velocity, or the point of contact between the ball bearings 1132
and output race
1118, to vary which results in a variation of speed of between 0 percent of
the rotational
speed of the input shaft 1102 up to 30 percent of the rotational speed of the
input shaft 1102.
The variation of speed in the output race 1118 of 0 percent to 30 percent of
the rotational
speed of the input shaft 1102 provides a wide range of adjustable rotational
speeds that can be
achieved at the output shaft 1126.
[0048] To ensure proper clamping of the ball bearings 1132 between the races
1114, 1116,
1118, 1120, springs 1154, 1156 are provided. Spring 1154 generates a clamping
force
between input race 1114 and stationary race 1120. Spring 1156 generates a
clamping force
between input race 1116 and output race 1118. These clamping forces against
the ball
bearings 1132 are maintained over the entire translating distance of the
translating clamp
1152. The telescopic shifter 1158 has threads on an inside surface that
connect to the threads
on the fixed threaded device 1160. The fixed threaded device 1160 is fixed to
housing 1172
and provides a fixed position relative to the housing 1172 so that the
translating clamp 1152
is able to translate in a horizontal direction as a result of the differential
threads on the two
sides of the telescopic shifter 1158.
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[00491 As also illustrated in Figure 11, the rotating components of the
traction drive
continuously variable transmission 1100 all rotate in the same direction, i.e.
rotational
direction 1112 and output rotation 1128 of the output gear 1122. Clamping nut
1168 holds
spring 1156 in place and preloads the spring 1156 to create the proper
diagonal pressure
between stationary race 1120 and input race 1114. When the translating clamp
1152 is
horizontally translated, as illustrated in Figure 11, there is a slight
translation of the input
shaft 1102 based upon the angles of the races 1114-1120 that contact the ball
bearings 1132.
The spline input gear 1104 allows translational movement in directions 1108,
1110 based
upon the points at which the ball bearings 1132 contact the races 1114-1120
and the
particular contact angle of the races with respect to the ball bearings 1132.
Housing 1170 is
bolted tightly to housing 1172 to contain the spring 1154, which creates the
proper amount of
clamping force between input race 1114 and stationary race 1120. Ball bearings
1132, as
illustrated in Figure 11, have a rotational progression 1131 in the four races
1114, 1116,
1118, 1120. The rotational direction 1112 of the shaft 1102 causes the gear
1122 to rotate in
a rotational direction 1128, as illustrated in Figure 11.
[00501 Figure 12 is a closeup view of the races 1114-1120 and ball 1132,
illustrating the
operation of the traction drive continuously variable transmission 1100. As
shown in Figure
12, race 1114 forcibly contacts ball 1132 at contact location 1134. Race 1116
forcibly
contacts ball 1132 at contact location 1136. Race 1118 forcibly contacts ball
1132 at contact
location 1138. Race 1120 forcibly contacts ball 1132 at contact location 1140.
Each of the
contact locations 1134, 1136, 1138, 1140 is located on a common great circle
on the surface
of the ball 1132. The great circle is located in a plane that contains the
center of the ball 1132
and the axis 1106 of the shaft 1102. Ball 1132 spins about a spin axis 1142
passing through
the center of the ball 1132 and bisects the great circle containing contact
locations 1134,
1136, 113 8, 1140. The spin axis 1142 of the ball 1132 is inclined at an angle
1146 with the
21
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
vertical axis 1144. The inclination angle 1146 is the same for each of the
balls disposed in
the races around the circumference of the traction drive 1100. The inclination
angle 1146
establishes a mathematical relationship between a distance ratio and a
circumferential
velocity ratio. The distance ratio is the ratio between the first distance
1148, which is the
orthogonal distance from the spin axis 1142 to the contact location 1134, and
a second
distance 1150, which is the orthogonal distance from the spin axis 1142 to
contact location
1136. This distance ratio is equal to the circumferential velocity ratio. The
circumferential
velocity ratio is the ratio between the first circumferential velocity and the
second
circumferential velocity, where the first circumferential velocity is the
difference between the
circumferential velocity of ball 1132 at race 1114 and a common orbital
circumferential
velocity of ball 1132 and the other balls in the races, while the second
circumferential
velocity is the difference between the circumferential velocity of the ball
1132 on the race
1116 and the common orbital circumferential velocity of the ball 1132, as well
as the other
balls disposed in the races. The radius of curvature of each of the races 1114-
1120 is larger
than the radius of curvature of ball 1132. In addition, the radius of
curvature of each of the
races 1114-1120 need not be a constant radius of curvature, but can vary.
Further, the radius
of curvature of each of the four races does not have to be equal.
[0051] When races 1116, 1118 translate simultaneously in a lateral direction,
such as lateral
translation direction 1108, the speed ratio of the rotation of shaft 1102 and
the rotational
direction 1112 change with respect to the rotation of the gear 1122 and
rotational direction
1128. Translation of races 1116, 1118 in lateral translation direction 1108
causes the first
distance 1148 to be larger and the second distance 1150 to be smaller. Hence,
the ratio of
distances, as well as the circumferential velocity ratio, changes, which
changes the rotational
speed of the gear 1122 with respect to shaft 1102.
22
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0052] As indicated above, the continuously variable transmission output is in
gear contact
with the traction drive speed reduction mechanism that connects to the turbine
compressor
shaft. As indicated above, there are at least two or three different types of
traction drive
speed reduction systems that may be used. The typical type is a planetary type
traction drive
for high speed reduction, which is disclosed in Figures 6-9, and Figure 10. If
a large speed
differential between the turbine shaft and the planetary roller is desired,
the embodiment of
Figure 10 may utilize only two rollers instead of three, in order to get the
gear ratio change
that is desired.
[0053] With three rollers, a limit of about a 10:1 reduction in speed exists
and there may be a
need for more like a 20:1 transmission to get the high speed 250,000 rpm
operation below the
25,000 rpm to which a 10:1 transmission would require. Therefore, a two roller
planetary
traction drive can be used in place of a three planetary drive system, in
Figure 10, in order to
achieve the speed reduction required of the smallest highest speed systems.
Two rollers also
provide for lower inertia, as each roller adds some amount of inertia to the
system. For the
lowest inertia, two rollers should be sufficient. The width of the traction
roller is slightly
wider than a three roller embodiment.
[0054] The multi-diameter planet rollers that roll against the shaft are made
of a springy
material, e.g., either a spring steel or another 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 not restrict the shaft's ability to find
its ideal center of
rotation.
[0055] When a turbocharger operates at extremely high speeds, it has balance
constraints that
cause the shaft to need to find its own center of rotation. The balance will
be compensated by
the movement of the center shaft. This movement can be compensated by spring-
loaded
rollers. The spring-loaded rollers can also be made extremely light weight by
making them
23
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
out of a thin band of steel that allows them to operate against the shaft with
very low inertia.
The band thickness must be thick enough to put sufficient pressure on the
traction surfaces to
provide the normal force needed for traction. A cam follower can be disposed
inside the
roller that will position each roller and hold that position within the
system. Rollers need to
operate in a very straight alignment between the outer drum and the
turbine/compressor shaft,
but the key to low inertia is lightweight. One or two cam followers can be
utilized to hold the
steel band in place, such that the steel band stays in alignment in the
system.
[00561 The ring roller 614 is connected to a gear on the outside surface so
that the ring roller
can transmit the power in or out of the multi-diameter traction drive 416. The
ring roller 614
can be made in numerous ways. Ring roller 614 can simply be a solid piece of
steel or other
appropriate material that is capable of transmitting the torque in and out of
the multi-diameter
traction drive 416. Ring roller 614 can be made of numerous materials that
allow ring roller
614 to be lightweight, but ring roller 614 has to be from a material that can
be used as a
traction drive surface on the roller surface 687. A proper roller surface 687
allows the planet
rollers 664, 666, 668 to transmit the torque through traction.
[00571 Also, turbine/compressor shaft 414 needs to be held in very accurate
alignment. The
alignment of the turbine/compressor shaft 414, within the housing, allows the
clearances to
be held between the tips of the blades of the compressor and the compressor
housing. A
tighter clearance increases the compressor efficiency. A more accurate
position decreases the
chance of touching between the turbine compressor fan 638 and the compressor
housing 640.
A method of controlling the thrust load that comes from compressing the gas
against the
compressor wheel is necessary to ensure that there is a minimum of clearance.
This can be
done using a thrust bearing (not shown) that is oil fed or a thrust bearing
that is a ball bearing
or roller bearing type of bearing.
24
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0058] Typically, in a turbocharger, the bearings are, for reliability
purposes, sleeve bearings
that have an oil clearance both on the inside and the outside in order to
allow for the turbine
shaft to center itself in its harmonic rotation. The balancing requirements
for a high volume
manufactured turbocharger are reduced by using a double clearance bearing.
These bearing
types have been used because of the requirement of tighter clearances and more
accurate
alignment of the shaft of the turbocharger. A ball bearing is used for both
holding the
compressor and turbine and for maintaining better alignment to the housing
from a side-to-
side motion perspective. This can be accomplished with one or two ball
bearings. Alignment
of bearings within an outer area that is pressurized with oil allows the
bearings to float and
allows the bearing to find a center. This does affect the clearance between
the housing,
turbine and compressor outside edges, but allows thrust clearance to remain
small. Turbo
shaft bearings provide a third point of constraint to maintain alignment of
the rollers. Cam
followers in the middle of the rollers can keep the rollers at 120 degrees
from one another.
Two small cam followers can be used for each roller to eliminate backlash when
power
changes direction.
[0059] Also, a larger turbine can be used. The turbine wheel can be made
larger in diameter
than normal. It is possible to make the turbine outer diameter even larger
than the
compressor wheel, without hitting the critical speed where tips come close to
the speed of
sound, because the density of the exhaust is lower than inlet air and
therefore the speed of
sound is higher. This allows the exhaust to generate more torque on the
turbine/compressor
shaft without higher backpressure. Having higher torque causes the turbine to
recover more
energy than is required to compress the intake air. This produces more energy
than can be
recovered and transmitted to the engine. More energy from the same exhaust gas
flow that is
not needed for compression gets transferred to the crankshaft and creates
lower fuel
consumption.
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0060] Further, turbine efficiency can be improved by using guide vanes that
control the
angle of incidence which exhaust gases impact the turbine wheel. This makes
the peak
efficiency higher, but narrows the speed range upon which that efficiency is
achieved. A
narrow speed range is bad for a normal turbocharger, and is not a problem for
a super-
turbocharger where the governor can provide the necessary speed control.
[0061] Higher backpressure across the turbine compared to the pressure across
the
compressor can also create an unbalanced super-turbocharger. For a normal
turbocharger,
this pressure difference is the other way around. Having higher backpressure
causes the
turbine to recover more energy than is required to compress the intake air.
This produces
more energy that can be recovered and transmitted to the engine. Higher
backpressure is
needed for high pressure EGR loops on diesel engines. High backpressure
normally requires
a valve or a restriction, so high backpressure is normally lost energy because
a normal
turbocharger cannot be unbalanced without over-speeding. Increasing
backpressure is bad
for gasoline and natural gas engines, because it increases the amount of
exhaust gas that gets
trapped in the cylinder, which makes the engine more likely to have detonation
problems.
[0062] In accordance with another embodiment, a second turbine wheel can be
positioned on
the turbine/compressor shaft to increase the energy recovered by the turbine
and improve the
fuel efficiency of the engine system. Also, a second compressor wheel can be
positioned on
the same shaft to increase the boost pressure potential of the super-
turbocharger and allow
intercooling between the stages. This makes the intake temperature cooler for
a given boost
and therefore lowers NOx.
[0063] In addition, turbine blade cooling can be provided through the wing
tips to reduce
temperatures in high temperature applications. This can be done with hollow
wing tips at the
outer edge of the turbine. This special tip design increases turbine
efficiency and provides a
path for cooling air to get through the blades. Turbine wing cooling can also
be provided by
26
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
compressed air from the compressor side fed across the housing to the back
side of the
turbine wheel. In addition, a heat pipe can be used to cool the turbine wheel
and blades.
[0064] In addition, a torsional softening device can be used on the power
path. Crankshaft
energy or rotational mechanical energy from a propulsion train can be brought
through a flex
shaft or an impulse softening device (either spring loaded or flexing) in such
a way that
torque impulses from the engine or propulsion train are removed without loss
of that energy,
before entering the housing. By not impacting the transmission with high
torque spikes on
the traction drive, the peak torque requirement is reduced. By eliminating
these torque
spikes, traction drives are more reliable, because the traction requirements
are limited by the
maximum torque on the system. By minimizing these torque spikes on the
traction drives,
the size and surface contact areas of the traction drives can be minimized.
Minimal surface
contact areas maximize efficiency of the system, and can still achieve the
torque required for
transmitting the continuous power.
[0065] Alternatively, and in accordance with another embodiment, a variable
speed traction
drive design with fixed displacement hydraulic pumps in place of the shaft,
belt or gear drive
may be utilized. This makes the system easier to package, which could be
especially useful
on very big engines having multiple turbochargers.
[0066] In a further embodiment, illustrated in Figure 13, a second super-
turbocharger is run
off one transmission as a way to get a higher pressure ratio, and as a way to
get cooler intake
temperatures 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 conduit
1308 and compresses air, which is supplied to the engine from compressed air
conduit 1310.
Exhaust air conduit 1314 receives exhaust gas from the engine to run the
turbine of the first
super-turbocharger 1302. The exhaust gas exits the exhaust exit conduit 1312.
The first
27
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
super-turbocharger 1302 is coupled to the second super-turbocharger 1304 with
a transfer
gear 1306.
[00671 Figure 14A illustrates another embodiment of an implementation of the
use of two
super-turbochargers, such as low pressure super-turbocharger 1402 and 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
first opens. To
improve this impulse pressure recovery, as illustrated in Figure 14A, the high
pressure
exhaust valve ports 1406, 1408 are separated from the low pressure exhaust
valve ports 1410,
1412 of a four-valve engine. The high pressure exhaust ports 1406, 1408 are
directed to high
pressure turbine 1434 via high pressure exhaust manifold 1430, while low
pressure exhaust
ports are directed to low pressure turbine 1420, via low pressure exhaust
manifold 1428. By
changing valve timing of the valves in the high pressure exhaust ports 1406,
1408, such that
valves on the high pressure exhaust ports 1406, 1408 are opened first and
ported to the high
pressure turbine 1434, the pulse energy is recovered better. The valves on the
high pressure
exhaust ports 1406, 1408 are closed quickly, and then the valves on the low
pressure exhaust
ports 1410, 1412 are opened for the duration of the exhaust stroke. The valves
on the low
pressure exhaust ports 1410, 1412 are ported to a low pressure turbine 1420.
This process
reduces the work required by the piston to exhaust the cylinder. This process
improves idle
fuel efficiency, or at least eliminates parasitic losses at idle. The outlet
of the high-pressure
turbine 1434 is also connected to the low-pressure turbine 1420. A catalyzed
diesel
particulate filter (not shown) can also be disposed before the lower pressure
turbine.
100681 As also illustrated in Figure 14A, an EGR conduit 1438 is connected to
the high
pressure exhaust manifold 1430. The EGR conduit 1438 allows a portion of the
exhaust from
the high pressure exhaust manifold 1430 to be channeled back to the intake
manifold 1444,
via cooler 1440 and EGR valve 1442. The exhaust from the high pressure exhaust
manifold
28
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
1430, that is channeled through the EGR conduit 1438, is channeled to the
intake manifold
1444 for the purpose of the recirculation of exhaust gases. The exhaust gases
flowing
through the exhaust gas recirculator conduit 1438 assist in lowering the
combustion
temperature in the combustion chamber, especially after being cooled in cooler
1440. The
exhaust gases contain moisture and other liquids that assist in lowering the
temperature of the
combustion chamber to thereby reduce NOx emissions from the engine. The amount
of
recirculated exhaust gas is controlled by the EGR valve 1442. EGR valve 1442
can be fixed,
such as through the use of a restrictor valve, or can be varied, depending
upon the monitored
NOx emissions of the engine.
[0069] As also shown in Figure 14A, high pressure air is funneled through the
high pressure
compressor manifold 1446 from the high pressure compressor 1432 to the intake
manifold
1444. Hence, the intake manifold 1444 is maintained at a predetermined high
pressure level
dictated by the output of the high pressure compressor 1432. In order for the
recirculated
gases to flow through the EGR conduit 1438, the pressure in the high pressure
manifold 1430
must be higher than the pressure in the intake manifold 1444, as dictated by
the output
pressure of the high pressure compressor 1432. In that regard, the valves in
the high pressure
exhaust ports 1406, 1408 are opened sufficiently early during the downstroke
of the piston,
when residual pressure still exists in the piston to create a sufficiently
high pressure in the
high pressure exhaust manifold 1430 to drive the exhaust gases from the high
pressure
exhaust manifold 1430 through the EGR conduit 1438. As disclosed below, the
valves in the
high pressure exhaust ports 1406, 1408 open at a point at which there is a
small amount of
energy loss in the process of driving the pistons downwardly. The opening
point of the high
pressure valves is prior to bottom dead center, but beyond the point of
maximum torque of
the piston on the crankshaft, which is the point at which the rods are at
substantially 90 .
This point occurs at approximately 100 . The amount of torque is proportional
to the cosine
29
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
of the angle of the rods, so that the lower the piston is when the high
pressure valves open,
the less energy that is lost in driving the pistons. However, there is a
substantial amount of
residual pressure left in the cylinder chamber, which can be exhausted from
the cylinder
chamber by the high pressure valves prior to reaching bottom dead center, that
can be used to
drive the exhaust gases in the EGR conduit 1438 into the high pressure turbine
1434. By pre-
exhausting the cylinder, using the high pressure valves of the high pressure
exhaust ports
1406, 1408, a large amount of the residual pressure in the cylinder is
exhausted prior to
opening of the low pressure exhaust ports 1410, 1412. When opened, the low
pressure
exhaust ports 1410, 1412 are capable of exhausting most of the pressure from
the cylinders.
In this manner, the residual pressure in the cylinders is used to channel
exhaust gas through
both the EGR conduit 1438, to reduce NOx emissions and to drive the high
pressure turbine
1434, which adds additional power and efficiency to the engine.
[0070] As also shown in Figure 14A, the exhaust gases from the low pressure
exhaust
manifold are used to drive a low pressure turbine 1420 of the low pressure
superturbocharger
1402. Exhaust gases emitted by the high pressure turbine 1434 are combined
with the low
pressure exhaust gases from the low pressure exhaust ports 1410, 1412 to drive
the low
pressure turbine 1420. Exhaust gases from the low pressure turbine 1420 are
exhausted by
exhaust outlet 1436. The low pressure turbine 1420 is coupled to the low
pressure
compressor 1418, which compresses the inlet air 1422 by a predetermined
amount. Conduit
1424 channels the compressed air from the low pressure compressor 1418 to the
input of the
high pressure compressor 1432, which functions to further compress the
pressurized air in
1424 to produce higher pressure compressed air, which is channeled to the
inlet manifold
1444 by high pressure compressor manifold 1446.
10071 ] Figure 14B illustrates a variation of the embodiment illustrated in
Figure 14A. As
illustrated in Figure 14B, the high pressure exhaust ports 1406, 1408 are
combined into a
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
high pressure exhaust manifold that is coupled to the high pressure turbine
1434. In other
words, all of the high pressure exhaust from the high pressure exhaust
manifold 1430 is
applied to the high pressure turbine 1434 to drive the high pressure turbine
1434, which in
turn drives the high pressure compressor 1432. The high pressure compressor
1432 receives
compressed air in conduit 1424 from the low pressure compressor 1418 of low
pressure
super-turbocharger 1402 that compresses the inlet air 1422. The output of high
pressure
compressor 1432 is fed to the input manifold 1444 via high pressure compressor
manifold
1446. The low pressure compressor 1418 is driven by the low pressure turbine
1420 that is
driven by the low pressure exhaust gases, in the low pressure exhaust manifold
1428, that are
emitted by the low pressure exhaust ports 1410, 1412. Exhaust gases from the
low pressure
turbine 1420 are exhausted through exhaust outlet 1436. The high pressure
gases from the
high pressure exhaust manifold 1430, that drive the high pressure turbine
1434, are coupled
to the exhaust gas recirculation (EGR) conduit 1426 and transmitted back to
the intake
manifold 1444. The high pressure gases from the high pressure exhaust manifold
1430, that
drive the high pressure turbine 1434, are not substantially reduced in
pressure and have a
sufficiently high pressure to insert the exhaust gases from the EGR conduit
1426 into the
intake manifold 1444. Figure 14B provides the greatest reduction in NOx gases,
since
essentially all of the exhaust gases from the high pressure exhaust manifold
1430 are
recirculated to the intake manifold 1444.
[00721 As also illustrated in Figure 14B, a waste gate 1448 may be utilized to
bypass high
pressure exhaust gases from the high pressure exhaust manifold 1430 to the EGR
conduit
1426. The high pressure exhaust gases, at times, may be too hot and/or may
provide exhaust
gases at a pressure that will overdrive the high pressure turbine 1434. In
that instance, the
waste gate 1448 can be opened to feed a portion of the high pressure exhaust
gas from the
high pressure exhaust manifold 1430 directly to the EGR conduit 1426. In
addition, an EGR
31
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
valve 1450 may be added, which connects the EGR conduit 1426 to the low
pressure exhaust
manifold 1428. If a sufficient amount of exhaust gases are being fed through
the EGR
conduit 1426, a portion of those gases may be directed from the EGR conduit
1426 to the low
pressure exhaust manifold 1428 via EGR valve 1450. The excess gases from the
EGR
conduit 1426 can then be used to run the low pressure turbine 1420 to add
additional power
to the engine by increasing the intake manifold pressure 1444. Use of the EGR
valve 1450
provides an additional manner in which recirculated gases can be recovered to
add additional
power to the engine and increase the efficiency of the operation of the
engine.
[00731 Figure 14C illustrates another modification of the embodiments of
Figures 14A and
14B. As shown in Figure 14C, inlet air 1422 is compressed by low pressure
compressor
1418. The compressed air from the low pressure compressor 1418 is fed by
conduit 1424 to
the intake manifold 1444. As also illustrated in Figure 14C, the second high
pressure turbine
is not utilized and all of the recirculation gas is recirculated from the high
pressure exhaust
ports 1406, 1408 via EGR conduit 1426 to the intake manifold 1444. Exhaust
gases from the
low pressure exhaust ports 1410, 1412 are combined in conduit 1428 to operate
low pressure
turbine 1420. The exhaust gases are then exhausted at exhaust outlet 1436.
Hence, all of the
blow down gases from the high pressure exhaust ports 1406, 1408 are fed back
into the intake
manifold 1444 to create a large reduction in NOx gases. Alternatively, an EGR
valve 1450
can be used to channel a portion of the exhaust gases in the EGR conduit 1426
to the low
pressure exhaust manifold 1428, which adds further power to the low pressure
turbine 1420
and reduces the amount of recirculated gases in the EGR conduit 1426. The EGR
valve 1450
can be adjusted to adjust the amount of exhaust gases that are fed from the
EGR conduit 1426
to the low pressure exhaust manifold 1428. This process may be beneficial if a
sufficient
amount of exhaust gases are recirculated in the EGR conduit 1426 to reduce NOx
output of
the engine.
32
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0074] Figure 14D is a graph of the valve lift, cylinder pressure and flow
rate versus the
piston position after top dead center. As shown in Figure 14D, the cylinder
pressure 1450
steadily decreases after top dead center, all the way through the stroke of
the piston. The lift
of the high pressure valve 1456 creates the high pressure flow 1452. The lift
of the high
pressure valve 1456 occurs around 1000 rotation and creates a large blow down
surge of the
high pressure flow 1452 that is exhausted through the high pressure exhaust
ports 1406, 1408
(Figures 14A, 14B and 14C). The lift of the low pressure valve is illustrated
at curve 1454.
The low pressure valve lift creates the low pressure flow 1458 in the low
pressure exhaust
ports 1410, 1412. As a result, the cylinder pressure 1450 is further reduced
in the cylinder.
[0075] Figure 14E is a PV graph of the cylinder pressure versus the volume in
the cylinder,
as the piston moves downwardly and then upwardly in the cylinder. Near zero
represents the
top dead center, while 1 represents the bottom dead center of the rotation of
the cylinder.
Two curves are shown in Figure 14E. Curve 1464 represents the curve of the
cylinder
pressure versus the volume for an engine that does not employ the Riley cycle.
Curve 1462 is
a curve that illustrates the cylinder pressure versus volume in the cylinder
for a Riley cycle
device, such as illustrated in Figures 14A-C. At point 1466, the high pressure
valve is
opened on the Riley cycle device, as illustrated in Figures 14A-C, and the
pressure is
reduced. The area 1468, between points 1466, 1470, is representative of the
energy lost by
opening the high pressure valve. However, as indicated 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 the non-Riley cycle device all the way through
to point 1474.
Between 1472 and point 1474, there is less pressure in the cylinder, which
results in less
backpressure on the cylinder as the cylinder moves from point 1472 to point
1474. The large
amount of area between the Riley cycle curve 1462 and the nonnal curve 1464,
between
33
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
points 1472 and 1476, as indicated by 1478, is indicative of the energy saved
by movement of
the piston in the cylinder at the lower pressure.
[0076] In an alternate embodiment, a super-turbocharger may be used as an air
pump for
after treatment, as well as for the engine and eliminates the need for a
separate pump just for
the burner.
[0077] In another embodiment, a governor (not shown) is provided to prevent
over-speeding,
keeping the compressor out of a surge condition and controlling to the maximum
efficiency
of the turbine and compressor. A super-turbocharger can be unique from a
normal
turbocharger because the peak of the turbine efficiency and the peak of the
compressor
efficiency can be at the same speed. Controlling to this peak efficiency speed
for a given
boost requirement can be modeled and programmed into an electronic governor.
An actuator
can provide governing, although an actuator is not needed for the electric
transmission.
[0078] In another embodiment, the oiling system for the super-turbocharger
pulls a vacuum
inside the housing, and therefore reduces aerodynamic losses of the high speed
components.
[0079] In another alternate embodiment, a dual clutch super-turbocharger
includes an
automatically shifted manual transmission. This type of transmission shifts
very smoothly
because it has a clutch on both ends. Figure 3C illustrates that the
transmission could be of
many different types.
[0080] In another embodiment, traction drives for both the transmission and
the speed
reduction from the turbo shaft are used. With ball bearings, the traction
fluid works as the
lubricant as well. During supercharging, the system improves load acceptance,
reduces soot
emissions, provides up to 30% increase in low end torque and up to 10%
increase in peak
power. During turbo-compounding, the system provides improved fuel economy of
up to
10% and controls backpressure. For engine downsizing, the system provides 30%
more low
end torque that allows the engine to be 30 to 50% smaller, having lower engine
mass and
34
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
improved vehicle fuel economy of 17% or more. Figure 15 illustrates the
simulated BSFC
improvement for a natural gas engine.
[0081] Also, a catalyst, a DPF or even a burner plus DPF can be positioned in
front of the
turbine of the super-turbocharger to heat the exhaust gas to a higher
temperature than the heat
of the engine. Higher temperatures expand the air even further making the flow
rate across
the turbine higher. Approximately 22% of this heat addition can be turned into
mechanical
work across the super-turbocharger, assuming 80% turbine efficiency. Normally,
higher
volume in the exhaust that is fed to the turbine would slow the turbine
response and create
even bigger turbo lag, but the super-turbocharger overcomes this problem with
the traction
drive 114 and continuously variable transmission 116 driving the pressure
response. Similar
techniques using a catalytic converter are disclosed in International Patent
Application No.
PCT/US 2009/051742 filed 24 July 2009 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 discloses and teaches.
[0082] Figure 16 is simplified single line form illustration of one embodiment
of a high
efficiency, super-turbocharged engine system 1600. As will become apparent to
those skilled
in the art from the following description, such a super-turbocharged engine
system 1600 finds
particular applicability in diesel engines and some spark ignited, gasoline
engines that are
used in passenger and commercial vehicles, and therefore the illustrative
examples discussed
herein utilize such an environment to aid in the understanding of the
invention. However,
recognizing that embodiments of system 1600 have applicability to other
operating
environments such as, for example, land based, power generation engines, and
other land
based engines, such examples should be taken by way of illustration and not by
way of
limitation.
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[0083] As shown in Figure 16, the super-turbocharger 1604 includes a turbine
1606, a
compressor 1608, and a transmission 1610 that is coupled to the crank shaft
1612 of the
engine 1602 or other portions of the propulsion train. While not required in
all embodiments,
the illustrated 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.
[0084] Super-turbochargers have certain advantages of turbochargers. A
turbocharger
utilizes a turbine that is driven by the exhaust of the engine. This turbine
is coupled to a
compressor which compresses the intake air that is fed into the cylinders of
the engine. The
turbine in a turbocharger is driven by the exhaust from the engine. As such,
the engine
experiences a lag in boost when first accelerated until there is enough hot
exhaust to spin up
the turbine to power a compressor, which is mechanically coupled to the
turbine, to generate
sufficient boost. To minimize lag, smaller and/or lighter turbochargers are
typically utilized.
The lower inertia of the lightweight turbochargers allows them to spin up very
quickly,
thereby minimizing the lag in performance.
100851 Unfortunately, such smaller and/or lighter weight turbochargers may be
over-sped
during high engine speed operation when a great deal of exhaust flow and
temperature is
produced. To prevent such over speed occurrences, typical turbochargers
include a waste
gate valve that is installed in the exhaust pipe upstream of the turbine. The
waste gate valve
is a pressure operated valve that diverts some of the exhaust gas around the
turbine when the
output pressure of the compressor exceeds a predetermined limit. This limit is
set at a
pressure that indicates that the turbocharger is about to be over-sped.
Unfortunately, this
results in a portion of the energy available from the exhaust gases of the
engine being wasted.
100861 Recognizing that conventional turbochargers sacrifice low end
performance for high
end power, devices known as super-turbochargers were developed. One such super-
36
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
turbocharger is described in US Patent No. 7,490,594 entitled "Super-
Turbocharger," issued
February 17, 2009, which is specifically incorporated herein by reference for
all that it
discloses and teaches.
[00871 As discussed in the above-referenced application, in a super-
turbocharger the
compressor is driven by the engine crank shaft via a transmission that is
coupled to the engine
during low engine speed operation when sufficiently heated engine exhaust gas
is not
available to drive the turbine. The mechanical energy supplied by the engine
to the
compressor reduces the turbo lag problem suffered by conventional
turbochargers, and allows
for a larger or more efficient turbine and compressor to be used.
100881 The super-turbocharger 1604, illustrated in Figure 16, operates to
supply compressed
air from the compressor 1608 to the engine 1602 without suffering from the
turbo-lag
problem of a conventional turbocharger at the low end and without wasting
energy available
from the engine exhaust gas heat supplied to the turbine 1606 at the high end.
These
advantages are provided by inclusion of the super-turbocharger transmission
1610 that can
both extract power from, and supply power to, the engine crank shaft 1612 to
both drive the
compressor 1608 and load the turbine 1606, respectfully, during various modes
of operation
of the engine 1602.
[00891 During start up, when conventional turbochargers suffer a lag due to
the lack of
sufficient power from the engine exhaust heat to drive the turbine, the super-
turbocharger
1604 provides a supercharging action whereby power is taken from the crank
shaft 1612 via
the super-turbocharger transmission 1610 to drive the compressor 1608 to
provide sufficient
boost to the engine 1602. As the engine comes up to speed and the amount of
power
available from the engine exhaust gas heat is sufficient to drive the turbine
1606, the amount
of power taken from the crank shaft 1612 by the transmission 1610 is reduced.
Thereafter,
37
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
the turbine 1606 continues to supply power to the compressor 1608 to compress
the intake air
for use by the engine 1602.
[0090] As the engine speed increases, the amount of power available from the
engine exhaust
gas heat increases to the point where the turbine 1606 would over speed in a
conventional
turbocharger. However, with the super-turbocharger 1604, the excess energy
provided by the
engine exhaust gas heat to the turbine 1606 is channeled through the
transmission 1610 to the
engine crank shaft 1612 while maintaining the compressor 1608 at the proper
speed to supply
the ideal boost to the engine 1602. The greater the output power available
from the exhaust
gas heat of the engine 1602, the more power generated by the turbine 1606 that
is channeled
through the transmission 1610 to the crank shaft 1612 while maintaining the
optimum boost
available from the compressor 1608. This loading of the turbine 1606 by the
transmission
1610 prevents the turbine 1606 from over speeding and maximizes the efficiency
of the
power extracted from the engines exhaust gases. As such, a conventional waste
gate is not
required.
[0091] While the amount of power available to drive the turbine 1606 in a
conventional
super-turbocharged application is limited strictly to the amount of power
available from the
engine exhaust, the turbine 1606 is capable of generating significantly more
power if the
thermal energy and mass flow supplied to the turbine blades can be fully
utilized and/or can
be increased. However, the turbine 1606 cannot operate above a certain
temperature without
damage, and the mass flow is conventionally limited to the exhaust gases
coming out of the
engine 1602.
[0092] Recognizing this, the embodiment of the system 1600 protects the
turbine 1606 from
high temperature transients by placing a catalyzed diesel particulate filter
1616 upstream of
the turbine 1606. In one embodiment, the catalyzed diesel particulate filter
is placed
upstream from the turbine near the exhaust manifold which enables exothermic
reactions that
38
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
result in an increase in exhaust gas temperature during sustained high speed
or load operation
of the engine. Using a catalyzed digital particulate filter, energy can be
recovered from the
soot, hyrocarbons and carbon monoxide that is burned on the catalyzed diesel
particulate
filter 1616 to add power to the super turbo charger which is located
downstream from the
catalyzed digital particulate filter 1616. Energy recovery can be achieved
from either a
conventional diesel particulate filter that has a very restricted flow-through
capacity, with
nearly 100% soot collection, or by using a flow-through catalyzed digital
particulate filter. A
flow-through catalyzed digital particulate filter is a diesel particulate
filter that only collects
about half of the soot and lets the other half pass through. Both types of
digital particulate
filters are catalyzed in order to have emissions burn at a reasonably low
temperature.
Catalyzing of the digital particulate filter is accomplished by providing a
platinum coating to
the particulate filter elements that ensures that soot, hydrocarbons and
carbon monoxide burn
at low temperatures. Additionally, it is possible to use a diesel particulate
filter and a burner
to burn the soot off of the digital particulate filter upstream from the super-
turbocharger.
Gasoline engines typically do not have enough soot to require a diesel
particulate filter.
However, some gasoline direct injection engines produce sufficient soot and
other
particulates so that the use of a particulate filter may be beneficial, and
the use of a catalyzed
diesel particulate filter may be deployed in the manner disclosed herein.
10093] To cool the exhaust gas, prior to reaching the turbine, a portion of
the compressed air
generated by the compressor is fed directly into the exhaust upstream from the
turbine, via a
control valve 1618, and added to the engine exhaust gases leaving the
catalyzed diesel
particulate filter 1616. The cooler intake air expands and cools the exhaust
gas and adds
additional mass to the exhaust gas flow, which adds additional power to the
turbine 1606 as
described in more detail below. As more cooler air is provided to the hot
exhaust gases to
maintain the temperature of the combined flow to the turbine 1606 at the
optimum
39
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
temperature, the energy and the mass flow that is delivered to the turbine
blades also
increases. This significantly increases the power supplied by the turbine to
drive the engine
crank shaft.
[0094] So as to not interfere with the stoichiometric reaction within the
catalyzed diesel
particular filter 1616, the compressor feedback air is added downstream of the
catalyzed
diesel particulate filter 1616. In such an embodiment, the engine exhaust gas
is passed
through the catalyzed diesel particulate filter 1616 and temperature of the
exhaust gas is
increased by the exothermic reaction. The compressed feedback air is then
added and
expands so that the total mass flow supplied to the turbine is increased.
Embodiments of the
present invention control the amount of compressed feedback air supplied to
cool the exhaust
and to drive the turbine to ensure that the combination of the cooler
compressed feedback air
and the engine exhaust gases are delivered to the turbine at an optimum
temperature for
turbine blade operation.
[0095] Since the catalyzed diesel particulate filter 1616, illustrated in
Figure 16, has a large
thermal mass than the exhaust gases from engine 1602, the catalyzed diesel
particulate filter
1616 operates as a thermal damper initially, which prevents a high temperature
thermal spike
from reaching the turbine 1606. However, since the reactions in the catalyzed
diesel
particulate filter 1616 are exothermic in nature, the temperature of the
exhaust gases leaving
the catalyzed diesel particulate filter 1616 are higher than that of the
exhaust gas entering the
catalyzed diesel particulate filter 1616. So long as the temperature of the
exhaust gas
entering the turbine remains below the maximum operating temperature of the
turbine 1606,
there is no problem.
[0096] However, during sustained high speed and high load operation of the
engine 1602, the
exit temperatures of the converted exhaust gas from catalyzed diesel
particulate filter 1616
can exceed the maximum operating temperature of turbine 1606. As set forth
above, the
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
temperature of the exhaust gases exiting the catalyzed diesel particulate
filter 1616 are
reduced by supplying a portion of the compressed air from the compressor 1608
via a
feedback valve 1618, and mixed with the exhaust gas exiting the catalyzed
diesel particulate
filter 1616. Significantly improved fuel economy is achieved by not using fuel
as a coolant
during such conditions, as is done in conventional systems. Additionally, the
operation of the
transmission is controlled to allow the compressor 1608 to supply a sufficient
amount of
compressed air to provide optimum boost to the engine 1602 and the compressed
feedback air
to the turbine 1606 via the feedback valve 1618. The excess power generated by
the turbine
1606 resulting from the increased mass flow of the compressed air through the
turbine is
channeled via the transmission 1610 to the crank shaft 1612, yet further
increasing fuel
efficiency.
[0097] The output temperature of the compressed air from the compressor 1608
is typically
between about 200 C to 300 C. A conventional turbine can operate optimally to
extract
power from gases at approximately 950 C, but not higher without distortion or
possible
failure. Because of the material limits of the turbine blades, the optimal
power is achieved at
approximately 950 C. Since the materials limit the exhaust gas temperatures to
about 950 C,
supplying more air to increase the mass flow across the turbine at the
temperature limit, e.g.,
950 C, increases the performance of the turbine.
[0098] While such a flow of compressed feedback air at 200 C to 300 C is
helpful in
reducing the temperature of the exhaust gas coming out of the catalyzed diesel
particulate
filter 1616, it is recognized that maximum power from the turbine 1606 can be
supplied when
the temperature and the mass flow is maximized within the thermal limits of
the turbine 1606.
As such, in one embodiment, the amount of feedback air is controlled so that
the combination
of exhaust gas and feedback air is maintained at or near the turbine's maximum
operating
temperature so that the amount of power delivered to the turbine is maximized
or
41
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
significantly increased. Since all of this excess power is normally not
required by the
compressor 1608 to supply the optimum boost to engine 1602 and to supply the
compressor
feedback air via feedback valve 1618, the excess power may be transferred by
the
transmission 1610 to the crank shaft 1612 of the engine 1602 to thereby
increase the overall
efficiency or power of the engine 1602.
[0099] As discussed above, in one embodiment, the connection of the compressor
feedback
air via feedback valve 1618 employs a catalyzed diesel particulate filter 1616
as the thermal
buffer between the engine 1602 and turbine 1606. As such, the supply of air
from the
compressor is provided downstream of the catalyzed diesel particulate filter
1616 so as to not
disrupt the stoichiometric reaction within the catalyzed diesel particulate
filter 1616. That is,
in embodiments that utilize a catalyzed diesel particulate filter 1616,
supplying the
compressor feedback air upstream of the catalyzed diesel particulate filter
1616 would result
in excess oxygen being supplied to the catalyzed diesel particulate filter
1616, thereby
preventing the catalyzed diesel particulate filter 1616 from generating a
stoichiometric
reaction that is required for proper operation.
[00100] Since optimum efficiency of power generation by the turbine 1606 is
achieved
when the temperature of the gas mixture of the compressor feedback air and
exhaust gas on
the turbine blades is maximized (within the material limits of the turbine
itself), the amount
of compressor feedback air admitted by the feedback valve 1618 is limited so
as to not reduce
the temperature significantly below such an optimized temperature. As the
catalyzed diesel
particulate filter 1616 produces more thermal energy via an exothermic
reaction and the
temperature of the converted exhaust gases from the catalyzed diesel
particulate filter 1616
increases to a temperature above the maximum operating temperature of the
turbine 1606,
more compressor feedback air may be supplied via feedback valve 1618 which
increases the
mass flow and energy supplied to the turbine 1606. As the amount of thermal
energy
42
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
generated by catalyzed diesel particulate filter 1616 is reduced, the amount
of compressor
feedback air supplied by feedback valve 1618 can also be reduced so as to
avoid supplying
more air than necessary, which results in the maintenance of the temperature
of the gas
mixture at the optimum operating condition.
[00101] In another embodiment, the system utilizes the feedback valve 1618 for
feeding
back the cooler compressor air into the exhaust ahead of the turbine at low
speed, high load
operating conditions to avoid surging the compressor. Compressor surge occurs
when the
compressor pressure gets high but the mass flow allowed into the engine is low
as a result of
the engine turning at a slow rpm and not requiring much intake air flow.
Surging (or
aerodynamic stalling) of the compressor resulting from low airflow across the
compressor
blades causes the efficiency of the compressor to fall very rapidly. In the
case of a normal
turbocharger, enough surge can stop the turbine from spinning. In the case of
a super-
turbocharger it is possible to use power from the engine crank shaft to push
the compressor
into surge. Opening the feedback valve 1618 allows a portion of the compressed
air to
feedback around the engine. This feedback flow brings the compressor out of
surge and
allows higher boost pressure to reach the engine 1602, thereby allowing the
engine 1602 to
generate more power than would normally be possible at low engine speeds.
Injecting the
compressed air into the exhaust ahead of the turbine conserves the total mass
flow through
the compressor so that all the flow reaches the turbine which minimizes the
power needed
from the engine to supercharge to a high boost pressure level.
[001021 In another embodiment, an additional cold start control valve 1620 may
be
included for operation during rich engine cold starts. During such an engine
cold start, the
exhaust gases from the engine 1602 typically include excess un-burnt fuel.
Since this rich
mixture is not stoichiometric, the catalyzed diesel particulate filter 1616 is
unable to fully
reduce the un-burnt hydrocarbons (UHC) in the exhaust gas. During such times,
the cold
43
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
start control valve 1620 may be opened to provide compressor feedback air to
the input of the
catalyzed diesel particulate filter 1616 to supply the extra oxygen necessary
to bring the rich
mixture down to stoichiometric levels. This allows the catalyzed diesel
particulate filter 1616
to light off faster and more efficiently reduce the emissions during the cold
start event. If the
engine is idling, a normal turbocharger would have no boost pressure to be
able to supply the
feedback air. However, the transmission ratio of transmission 1610 can be
adjusted to give
enough speed to the compressor to generate the pressure needed for the air to
flow through
valve 1620. In that regard, control signal 1624 can be used to adjust the
ratio of transmission
1610 so that sufficient rotational speed can be provided from the engine drive
shaft 1612 to
the compressor 1608 during idling, especially during a cold start, to compress
enough air to
flow through the cold start valve 1620 and ignite catalyzed diesel particulate
filter 1616 with
a sufficient amount of oxygen.
[001031 The requirement for the additional oxygen is typically limited in a
cold start event,
and often lasts only for 30 to 40 seconds. Many vehicles currently include a
separate air
pump to supply this oxygen during the cold start event, at significant cost
and weight
compared to the limited amount of time that such an air pump is required to
operate. By
replacing the separate air pump with the simple cold start control valve 1620,
significant
costs, weight and complexity savings are realized. Because the super-
turbocharger 1604 can
control the speed of the compressor 1608 via the transmission 1610, the cold
start control
valve 1620 may comprise a simple on/off valve. The amount of air supplied
during the cold
start event can then be controlled by controlling the speed of the compressor
1608 via
transmission 1610 under operation of the control signal 1624.
1001041 The cold start control valve 1620 may also be used during periods of
extremely
high temperature operation if fuel is used as a coolant within the engine
and/or for the
catalyzed diesel particulate filter 1616, despite the negative effect on fuel
efficiency. In such
44
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
situations, the cold start control valve 1620 will be able to supply the extra
oxygen necessary
to bring the rich exhaust back down to stoichiometric levels to allow the
catalyzed diesel
particulate filter 1616 to properly reduce the unburned hydrocarbon emissions
in the exhaust.
This provides a significant benefit to the environment over prior systems.
[00105] In embodiments where the cold start control valve 1620 is an on/off
valve, the
system can modulate cold start control valve 1620 to vary the amount of
compressed air
supplied so as to bring the exhaust down to stoichiometric levels. Other types
of variable
flow control valves may also be used to accomplish this same function.
[00106] Figure 16 also discloses a controller 1640. Controller 1640 controls
the operation
of the feedback valve 1618 and the cold start valve 1620. Controller 1640
operates to
optimize the amount of air flow through feedback valve 1618 for different
conditions. The
amount of air that flows through the feedback valve 1618 is the minimal amount
of air flow
that is necessary to obtain a specific desired condition, as described above.
There are two
specific conditions in which controller 1640 operates feedback valve 1618,
which are: 1)
surge limit of the compressor for a given boost requirement is proximate at
low rpm, high
load of the engine; and, 2) temperature of the gas mixture is proximate
entering the turbine
1606 at high rpm, high load conditions.
[00107] As shown in Figure 16, controller 1640 receives the gas mixture
temperature
signal 1630 from a temperature sensor 1638 that detects the temperature of the
gas mixture of
the cooling air supplied from the compressor 1608 that is mixed with the hot
exhaust gases
produced by the catalyzed diesel particulate filter 1616. In addition, the
controller 1640
detects the compressed air intake pressure signal 1632 that is generated by
the pressure sensor
1636 that is disposed in the conduit of compressed air that is supplied from
the compressor
1608. Further, an engine speed signal 1626 and an engine load signal 1628 that
are supplied
from the engine 1602 or a throttle are fed to the controller 1640 .
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
[001081 With respect to control of the temperature of the gas mixture that is
supplied to the
turbine 1606 at high speed, high load conditions, controller 1640 limits the
temperature of the
gas mixture to a temperature that maximizes the operation of the turbine 1606,
without being
so high as to damage the mechanisms of the turbine 1606. In one embodiment, a
temperature
of approximately 925 C is an optimal temperature for the gas mixture to
operate the turbine
1606. Once the temperature of the gas mixture that is fed into the turbine
1606 begins to
exceed 900 C, the feedback valve 1618 is opened, to allow compressed air from
the
compressor 1608 to cool the hot exhaust gases from the catalyzed diesel
particulate filter
1616 prior to passing into the turbine 1606. The controller 1640 can be
designed to target a
temperature of approximately 925 C, with an upper bound of 950 C and a lower
bound of
900 C. The limit of 950 C is one at which damage to the turbine 1606 may occur
using
conventional materials. Of course, the controller can be designed for other
temperatures,
depending upon the particular types of components and materials used in the
turbine 1606. A
conventional proportional integral derivative (PID) control logic device can
be used in the
controller 1640 to produce these controlled results.
[001091 The benefit of controlling the temperature of the gas mixture that
enters the
turbine 1606 is that the use of fuel in the exhaust to limit the turbine inlet
temperatures of the
gas mixture is eliminated. Using the flow of the cooler compressed air to cool
the hot
exhaust gases from the catalyzed diesel particulate filter 1616 requires a
large amount of air,
which contains a large mass to achieve the desired cooler temperatures of the
gas mixture.
The amount of air that is required to cool the hot exhaust gases from the
catalyzed diesel
particulate filter 1616 is large because the cooler compressed air from the
compressor 1608 is
not a good coolant, especially when compared to liquid fuel that is inserted
in the exhaust
gas. The hot exhaust gases from the output of the catalyzed diesel particulate
filter 1616
cause the cooler compressed gas from the compressor 1608 to expand to create
the gas
46
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
mixture. Since a large mass of the cooler compressed air from the compressor
1608 is
required to lower the temperature of the hot exhaust gases from the catalyzed
diesel
particulate filter 1616, a large mass flow of the gas mixture flows across the
turbine 1606,
which greatly increases the output of the turbine 1606. The turbine power
increases by the
difference of the power created by the differential of the mass flow minus the
work required
to compress the compressed air flowing through the feedback valve 1618. By
obtaining the
gas mixture temperature signal 1630 from temperature sensor 1638 and
controlling the
addition of compressed air by feedback valve 1618, the maximum temperature is
not
exceeded.
[001101 Controller 1640 also controls the feedback valve 1618 to limit surge
in the
compressor 1608. The surge limit is a boundary that varies as a function of
the boost
pressure, the flow of air through the compressor and the design of the
compressor 1608.
Compressors, such as compressor 1608, that are typically used in
turbochargers, exceed a
surge limit when the flow of intake air 1622 is low and the pressure ratio
between the intake
air 1622 and the compressed air is high. In conventional super-turbochargers,
the flow of
intake air 1622 is low when the engine speed (rpm) 1626 is low. At low rpms,
when the
compressed air is not used in large volumes by the engine 1602, the mass flow
of intake air
1622 is low and surge occurs because the rotating compressor 1608 cannot push
air into a
high pressure conduit without a reasonable flow of intake air 1622. The
feedback valve 1618
allows flow through the compressed air conduit 1609 and prevents or reduces
surge in the
compressor 1608. Once surge in the compressor 1608 occurs, the pressure in the
compressed
air conduit 1609 cannot be maintained. Hence, at low rpm, high load operating
conditions of
the engine 1602, the pressure of the compressed air in the compressed air
conduit 1609 may
drop below desired levels. By opening the feedback valve 1618, the flow of
intake air 1622
through the compressor 1608 is increased, especially at low rpm, high load
operating
47
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
conditions of the engine, which allows the desired level of boost to be
achieved in the
compressed air conduit 1609. Feedback valve 1618 can simply be opened until
the desired
pressure in the compressed air conduit 1609 is reached. However, by simply
detecting boost
pressure in the compressed air conduit 1609, surge will occur prior to the
feedback valve
1618 being opened to bring the compressor 1608 out of a surge condition.
[001111 It is preferable, however, to determine a surge limit and open the
feedback valve
1618 in advance, prior to the occurrence of a surge condition. For a given rpm
and desired
boost level a surge limit can be determined. The feedback valve 1618 can begin
to open prior
to the compressor 1608 reaching a calculated surge limit. Opening the valve
early allows the
compressor to spool up to a higher boost pressure more quickly because the
compressor stays
closer to the higher efficiency points of the compressor operational
parameters. Rapid boost
pressure rise at low rpm can then be achieved. By opening the valve before
surge occurs, a
more stable control system can also be achieved.
[001121 Opening the feedback valve 1618 in such a way as to improve the
responsiveness
of the engine 1602, is achieved by allowing the engine 1602 to get to a higher
boost pressure
more quickly when the engine 1602 is at a lower rpm. Compressor 1608 is also
more
efficient, which results in less work for the transmission 1610 to achieve
supercharging.
Surge limit control can be modeled within standard model based control
simulation code,
such as MATLAB. Modeling in this manner will allow simulation of the
controller 1640 and
auto-coding of algorithms for controller 1640.
[001131 A model based control system, such as described above, is unique, in
that the
utilization of the transmission 1610 to control the rotation of the turbine
1606 and compressor
1608 generates boost pressure without turbo lag. In other words, the
transmission 1610 can
extract rotational energy from the crank shaft 1612 to drive the compressor
1608 to achieve a
desired boost in compressed air conduit 1609 very quickly and prior to the
turbine 1606
48
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
generating sufficient mechanical energy to drive the compressor 108 at such a
desired level.
In this manner, controls in a conventional turbocharger to reduce lag are
reduced or
eliminated. The model based control of the controller 1640 should be designed
to maintain
the optimum efficiency of the compressor 1608 within the operational
parameters of the
compressor 108.
[00114] The control model of controller 1640 should also be carefully modeled
on the
pressure operational parameters, as mapped against the mass flow allowed by
the engine for a
given target speed and load in which target speed and load may be defined
relative to the
position of the throttle of the vehicle. As shown in Figure 16, the engine
speed signal 1626
can be obtained from engine 1602 and is applied to the controller 1640.
Similarly, the engine
load signal 1628 can be obtained from the engine 1602 and applied to
controller 1640.
Alternatively, these parameters can be obtained from a sensors located on the
engine throttle
(not shown). The feedback valve 1618 can then be operated in response to a
control signal
1642 generated by controller 1640. Pressure sensor 1636 generates the
compressed air intake
pressure signal 1632 that is applied to the controller 1640, which calculates
the control signal
1642 in response to engine speed signal 1626, engine load signal 1628 and
compressed air
intake pressure signal 1632.
[00115] During operational conditions of the engine 1602, in which a surge
limit is not
being approached by the compressor 1608 and the temperature of the gas
mixture, as detected
by the temperature sensor 1638, is not reached, the feedback valve 1618 is
closed so that the
system works as a conventional super-turbocharged system. This occurs over a
majority of
the operating parameters of the engine 1602. When high load and low rpm
conditions of the
engine 1602 occur, the feedback valve 1618 is opened to prevent surge.
Similarly, at high
rpm, high load operating conditions of engine 1602, high temperatures are
produced in the
exhaust gases at the output of the catalyzed diesel particulate filter 1616,
so that the feedback
49
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
valve 1618 must be opened to reduce the temperature of the fuel mixture
applied to the
turbine 1606 below a temperature which would cause damage to the turbine 1606.
[00116] Figure 17 is a detailed diagram of the embodiment of the high
efficiency super-
turbocharged engine system 1600 illustrated in Figure 16. As shown in Figure
17, engine
1602 includes a super-turbocharger that has been modified, as described above
with respect
to Figure 16, to provide overall higher efficiency than conventional super-
turbocharged
engines, as well as providing high, optimal efficiency in low rpm, high load
operating
conditions, and high, optimal efficiency at high rpm, high load conditions.
The super-
turbocharger includes a turbine 1606 that is mechanically connected by a shaft
to compressor
1608. Compressor 1608 compresses intake air 1622 and supplies the compressed
intake air
to conduit 1704. Conduit 1704 is connected to feedback valve 1618 and
intercooler 1614.
As disclosed above, intercooler 1614 functions to cool the compressed air,
which becomes
heated during the compression process. The intercooler 1614 is connected to
the compressed
air conduit 1726 which, in turn, is connected to the intake manifold (not
shown) of the engine
1602. Pressure sensor 1636 is connected to the compressed air conduit 1704 to
detect the
pressure and supply a pressure reading via the compressed intake air pressure
signal 1632,
which is applied to controller 1640. The feedback valve 1618 is controlled by
a controller
feedback valve control signal 1642 generated by the controller 1640, as
disclosed above.
Under certain operating conditions, feedback valve 1618 opens to supply
compressed air
from compressed air conduit 1704 to a mixing chamber 1706.
1001171 As shown in the embodiment of Figure 17, the mixing chamber 1706
simply
comprises a series of openings 1702 in the catalyzed diesel particulate filter
output conduit
1708, which is surrounded by the compressed air conduit 1704 so that
compressed air
supplied from the compressed air conduit 1704 passes through the openings 1702
to mix with
the exhaust gases in the catalyzed diesel particulate filter output conduit
1708. Any desired
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
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
catalyzed diesel particulate filter output conduit 1708 to measure the
temperature of the
exhaust gases in the catalyzed diesel particulate filter output conduit 1708.
Temperature
sensor 1638 supplies a gas mixture temperature signal 1630 to controller 1640,
which
controls the feedback valve 1618 to ensure that the temperature of the exhaust
gases in the
catalyzed diesel particulate filter output conduit 208 do not exceed a maximum
temperature
that would damage to the turbine 1606. Catalyzed diesel particulate filter
1616 is connected
to the exhaust manifold 1710 by way of catalyzed diesel particulate filter
inlet conduit 1714.
By locating the catalyzed diesel particulate filter 1616 proximate to the
exhaust manifold
1710, the hot exhaust gases from the engine flow directly into the catalyzed
diesel particulate
filter 1616, which assists in activating the catalyzed diesel particulate
filter 1616. In other
words, the proximate location of the catalyzed diesel particulate filter 1616
near the outlet of
the engine exhaust gases does not allow the exhaust gases to cool
substantially prior to
entering the catalyzed diesel particulate filter 1616, which increases the
performance of the
catalyzed diesel particulate filter 1616. As the exhaust gases pass through
the catalyzed
diesel particulate filter 1616, the catalyzed diesel particulate filter 1616
adds additional heat
to the exhaust gases. These very hot exhaust gases at the output of the
catalyzed diesel
particulate filter 1616 are supplied to the catalyzed diesel particulate
filter output conduit 208
and are cooled in the mixing chamber 1706 with the compressed intake air from
the
compressed air conduit 1704. Depending upon the temperature of the very hot
exhaust gases
that are produced at the output of the catalyzed diesel particulate filter
1616, which varies
depending upon the operating conditions of the engine 1602, a different amount
of
compressed intake air will be added to the exhaust gas during high speed, high
load
conditions. During low engine speed, high engine load conditions, the feedback
valve 1618
51
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
also functions to allow intake air to flow through the compressor to avoid
surge. Surge is
similar to aerodynamic stall of the compressor blades, which occurs as a
result of the low
flow conditions through the compressor during low engine speed conditions.
When surge
occurs, the pressure in the intake manifold (not shown) falls because the
compressor 1608 is
unable to compress the intake air. By allowing air to flow through the
compressor 1608 as a
result of the feedback valve 1618 being opened, pressure can be maintained in
the intake
manifold so that, when high torque is required at low engine speeds, the high
torque can be
achieved because of the high intake manifold pressure.
[00118] As disclosed above, when the engine is operating under high speed,
high load
conditions, the catalyzed diesel particulate filter 1616 causes a large amount
of heat to be
generated in the exhaust gases that are supplied to the catalyzed diesel
particulate filter output
conduit 1708. By supplying compressed, cooler intake air to the catalyzed
diesel particulate
filter output conduit 1708, the hot exhaust gases under high speed, high load
conditions are
cooled. As the load and speed of the engine increases, hotter gases are
produced and more of
the compressed air from conduit 1704 is required. If the turbine 1606 does not
provide
sufficient rotational energy to drive the compressor, such as under low speed,
high load
conditions, the engine crank shaft 1612 can supply rotational energy to the
compressor 1608
via drive belt 1722, drive pulley 1718, shaft 1724, continuously variable
transmission 1716
and transmission 1728. Again, any portion of the propulsion train can be used
to supply
rotational energy to the compressor 1608, and Figure 17 discloses one
implementation in
accordance with one disclosed embodiment.
[00119] As also illustrated in Figure 17, a cold start valve 1620 is also
connected to the
compressed air conduit 1704, which in turn is connected to the cold start
conduit 1712. Cold
start conduit 1712 is connected to the catalyzed diesel particulate filter
inlet conduit 1714,
which is upstream from the catalyzed diesel particulate filter 1616. The
purpose of the cold
52
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
start valve is to provide compressed intake air to the input of the catalyzed
diesel particulate
filter 1616 during startup conditions, as disclosed above. Under startup
conditions, prior to
the catalyzed diesel particulate filter 1616 reaching full operational
temperatures, additional
oxygen is provided via the cold start conduit 1712 to initiate the catalytic
process. The
additional oxygen that is provided via the cold start conduit 1712 assists in
the initiation of
the catalytic process. Controller 1640 controls cold start valve 1620 via
controller cold start
valve control signal 1644 in response to the engine speed signal 1626, engine
load signal
1628, and the gas mixture temperature signal 1630.
[00120] Hence, the high efficiency, super-turbocharged engine 1600 operates in
a manner
similar to a super-turbocharger, with the exception that feedback valve 1618
supplies a
portion of the compressed air from the compressor to the input of the turbine
for two reasons.
One reason is to cool the exhaust gases prior to entering the turbine so that
the full energy of
the exhaust gases can be utilized and a waste gate is not needed under high
speed, high load
conditions. The other reason is to provide a flow of air through the
compressor to prevent
surge at low rpm, high load conditions. In addition, the catalyzed diesel
particulate filter can
be connected in the exhaust stream before the exhaust gases reach the turbine
so that the heat
generated by the catalyzed diesel particulate filter 1616 can be used in
driving the turbine
1606, and expanding the compressed intake air that is mixed with the hot gases
from the
catalyzed diesel particulate filter 1616, which greatly increases efficiency
of the system.
Further, the cold start valve 1620 can be used to initiate the catalytic
process in the catalyzed
diesel particulate filter 1616 by providing oxygen to the exhaust gases during
startup
conditions.
100121] Hence, a unique super-turbocharger is disclosed that uses a high speed
traction
drive having a fixed ratio that reduces the rotational mechanical speed of the
turbine/compressor shaft to an rpm level that can be used by a continuously
variable
53
CA 02786580 2012-07-06
WO 2011/096936 PCT/US2010/023398
transmission that couples energy between a propulsion train and the
turbine/compressor shaft.
A uniqueness of the super-turbocharger design is that the transmission is
disposed within the
system. The continuously variable transmission is disposed within a lower
portion of the
super-turbocharger housing. The continuously variable transmission 1116
provides the
infinitely variable speed ratios that are needed to transfer rotational
mechanical energy
between the super-turbocharger and the engine. Either a geared continuously
variable
transmission can be used as continuously variable transmission 1116 or a
traction drive
continuously variable transmission can be used. Hence, traction drives can be
used for both
the high speed traction drive 114 and the continuously variable transmission
1116.
[00122] The foregoing description of the invention has been presented for
purposes of
illustration and description. It is not intended to be exhaustive or to limit
the invention to the
precise form 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
principles of the invention and its practical application to thereby enable
others skilled in the
art to best utilize the invention in various embodiments and various
modifications as are
suited to the particular use contemplated. It is intended that the appended
claims be
construed to include other alternative embodiments of the invention except
insofar as limited
by the prior art.
54
