U.S. patent application number 13/881920 was filed with the patent office on 2013-08-29 for simplified variable geometry turbocharger with increased flow range.
This patent application is currently assigned to BORGWARNER INC.. The applicant listed for this patent is David G. Grabowska, John P. Watson. Invention is credited to David G. Grabowska, John P. Watson.
Application Number | 20130219885 13/881920 |
Document ID | / |
Family ID | 46025097 |
Filed Date | 2013-08-29 |
United States Patent
Application |
20130219885 |
Kind Code |
A1 |
Watson; John P. ; et
al. |
August 29, 2013 |
SIMPLIFIED VARIABLE GEOMETRY TURBOCHARGER WITH INCREASED FLOW
RANGE
Abstract
A variable geometry turbocharger is simplified yet able to
maintain pulse energy. In a first embodiment, a turbine housing is
provided with a pivoting flow control valve which pivots about a
point near the entry to the turbine housing. By moving the valve
about the pivot point, the effective volume of the turbine housing
volute is varied, thus effectively reducing the volume of exhaust
gas in the volute, allowing control of exhaust gas flowing to the
turbine wheel. In the second embodiment of the invention, a
rotating wedge segment within the volute is rotated from a first
position to a second position, changing the effective volume of the
volute and allowing control of exhaust gas flowing to the turbine
wheel.
Inventors: |
Watson; John P.; (Arden,
NC) ; Grabowska; David G.; (Asheville, NC) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Watson; John P.
Grabowska; David G. |
Arden
Asheville |
NC
NC |
US
US |
|
|
Assignee: |
BORGWARNER INC.
Auburn Hills
MI
|
Family ID: |
46025097 |
Appl. No.: |
13/881920 |
Filed: |
November 3, 2011 |
PCT Filed: |
November 3, 2011 |
PCT NO: |
PCT/US2011/059043 |
371 Date: |
April 26, 2013 |
Related U.S. Patent Documents
|
|
|
|
|
|
Application
Number |
Filing Date |
Patent Number |
|
|
61410519 |
Nov 5, 2010 |
|
|
|
Current U.S.
Class: |
60/605.1 ;
415/205 |
Current CPC
Class: |
F01D 17/148 20130101;
Y02T 50/671 20130101; Y02T 10/12 20130101; Y02T 50/60 20130101;
Y02T 10/144 20130101; F02B 37/025 20130101; F02B 37/22 20130101;
F01D 9/026 20130101; F04D 1/00 20130101; F02C 6/12 20130101 |
Class at
Publication: |
60/605.1 ;
415/205 |
International
Class: |
F04D 1/00 20060101
F04D001/00 |
Claims
1. A variable flow turbocharger comprising: (a) a turbine housing
(2) having a turbine inlet portion at a turbine housing foot (51)
for coupling to an exhaust manifold for receiving exhaust gas flow,
a turbine exducer portion (52) and a volute chamber between the
turbine housing foot and the exducer portion; (b) a turbine
impeller (70) having a multiplicity of blades located within said
turbine housing for receiving exhaust gas flow from the volute
chamber; (c) at least one volute divider wall to divide the volute
chamber into a larger volute chamber (48) and a smaller volute
chamber (49); and (d) exhaust flow control valve means adapted to
control the degree of blockage of exhaust gases flowing into the
larger volute chamber.
2. The variable flow turbocharger according to claim 1, wherein
said exhaust flow control valve means is variably adjustable.
3. The variable flow turbocharger according to claim 1, wherein
said volute divider wall includes openings through which exhaust
flow can pass between the larger and smaller volute chambers.
4. The variable flow turbocharger according to claim 1, wherein
said divider wall is an axial divider wall.
5. The variable flow turbocharger according to claim 1, wherein
said divider wall is a radial divider wall.
6. The variable flow turbocharger according to claim 1, wherein the
volume of the larger volute comprises at least about 55% of the
divided volute space and the volume of the smaller volute comprises
at most about 45% of the divided volute space.
7. The variable flow turbocharger according to claim 1, wherein the
volume of the larger volute comprises about 55-65% and the volume
of the smaller volute comprises about 45-35% of the divided
volute.
8. The variable flow turbocharger according to claim 1, comprising
at least two axial or radial divider walls.
9. The variable flow turbocharger according to claim 1, wherein the
divider wall is radial, and wherein the smaller volute (49) is on
the exducer side and the larger volute (48) is opposite the exducer
side.
10. The variable flow turbocharger according to claim 1, wherein
exhaust flow control valve means comprises a pivotable valve member
(72) adapted for pivoting between an open position wherein flow to
the larger volute (48) is not restricted and a blocking position
wherein flow to the larger volute (48) is effectively blocked.
11. An internal combustion engine including an exhaust manifold and
having a variable flow turbocharger fluidly coupled to the exhaust
manifold, the variable-capacity turbocharger comprising: (a) a
turbine housing (2) having a turbine inlet portion, a turbine
exducer portion (52) and a volute chamber; (b) a turbine impeller
located in said turbine housing and having a multiplicity of
blades; (c) at least one volute divider wall extending to the
vicinity of said turbine inlet portion to divide said volute
chamber into a larger volute chamber (48) and a smaller volute
chamber (49); (d) a divided exhaust manifold (36) including a first
exhaust introducing passageway for introducing exhaust gases into
said larger volute chamber and a second exhaust introducing
passageway for introducing exhaust gases into said smaller volute
chamber; and (e) exhaust flow control valve means located at least
in the first exhaust introducing passageway to control the degree
of blockage of exhaust gases flowing into said first exhaust
introducing passageway.
12. The variable flow turbocharger according to claim 11, wherein
said exhaust flow control valve means is located in the turbine
housing.
13. The variable flow turbocharger according to claim 11, wherein
said exhaust flow control valve means is located in the exhaust
manifold coupled to said turbine housing.
14. The variable flow turbocharger according to claim 11, wherein
said exhaust flow control valve means is located between the
exhaust manifold and the turbine housing.
Description
FIELD OF THE INVENTION
[0001] This invention addresses the need for a reduced cost,
increased range, turbine flow control device, and accomplishes this
by designing a simplified variable geometry turbocharger housing
with controlled assymetric flow to the turbine wheel.
BACKGROUND OF THE INVENTION
[0002] Turbochargers are a type of forced induction system. They
deliver air, at greater density than would be possible in the
normally aspirated configuration, to the engine intake, allowing
more fuel to be combusted, thus boosting the engine's horsepower
without significantly increasing engine weight. A smaller
turbocharged engine can replace a normally aspirated engine of a
larger physical size, thus reducing the mass and aerodynamic
frontal area of the vehicle.
[0003] Turbochargers (FIG. 1) use the exhaust flow (100) from the
engine exhaust manifold, which enters the turbine housing at the
turbine inlet (51) of a turbine housing (2), to drive a turbine
wheel (70), which is located in the turbine housing. The turbine
wheel is solidly affixed to one end of a shaft. A compressor wheel
(20) is mounted to the other end of the shaft and held in position
by the clamp load from a compressor nut. The primary function of
the turbine wheel is providing rotational power to drive the
compressor. Once the exhaust gas has passed through the turbine
wheel (70) and the turbine wheel has extracted energy from the
exhaust gas, the spent exhaust gas (101) exits the turbine housing
(2) through the exducer (52) and is ducted to the vehicle downpipe
and usually to after-treatment devices such as catalytic
converters, particulate traps and NO.sub.x traps.
[0004] The power developed by the turbine stage is a function of
the expansion ratio across the turbine stage, i.e., the expansion
ratio from the turbine inlet (51) to the turbine exducer (52). The
range of the turbine power is a function of, among other
parameters, the mass flow through the turbine stage.
[0005] The compressor stage consists of a wheel and its housing.
Filtered air is drawn axially into the inlet (11) of a compressor
cover (10) by the rotation of the compressor wheel (20). The power
generated by the turbine stage to the shaft and wheel drives the
compressor wheel (20) to produce a combination of static pressure
with some residual kinetic energy and heat. The pressurized gas
exits the compressor cover (10) through the compressor discharge
(12) and is delivered, usually via an intercooler, to the engine
intake.
[0006] The design of the turbine stage is a compromise among: the
power required to drive the compressor at different flow regimes in
the engine operating envelope; the aerodynamic design of the stage;
the inertia of the rotating assembly, of which the turbine is a
large part, since the turbine wheel is manufactured typically in
Inconel, which has a density 3 times that of the aluminum of the
compressor wheel; the turbocharger operating cycle, which affects
the structural and material aspects of the design; and the near
field (exhaust flow) both upstream and downstream of the turbine
wheel with respect to blade excitation.
[0007] Part of the physical design of the turbine housing is a
volute (47), or pair of volutes, the function of which is to
control the inlet conditions to the turbine wheel such that the
inlet flow conditions provide the most efficient transfer of power
from the energy in the exhaust gas to the power developed by the
turbine wheel, combined with the best transient response
characteristics. Theoretically the incoming exhaust flow from the
engine is delivered in a uniform manner from the volute to a vortex
centered on the turbine wheel axis. To do this, ideally, the cross
sectional area of the volute is at a maximum perpendicular to the
direction of flow, gradually and continuously decreasing until it
becomes zero. The inner boundary of the volute can be a perfect
circle, defined as the base circle (71); or, in certain cases, such
as a twin volute (48,49) as seen in FIG. 2A, it can describe a
spiral of minimum diameter not less than 106% of the turbine wheel
diameter.
[0008] The volute is defined by the decreasing radius of the outer
boundary of the volute (53) and by the inner boundary, as described
above, in one plane defined in the "X-Y" axis as depicted in FIG.
4, and the cross sectional areas, at each station, in the plane
passing through the "Z" axis, as depicted in FIG. 8A. The "Z" axis
is perpendicular to the plane defined by the "X-Y" axis and is also
the axis of the turbine wheel.
[0009] Multiple entry volutes can also be created by dividing the
volute area circumferentially. The volute is divided by axial walls
(103, 104) which follow the decreasing outer boundary of the
volute, as shown in FIG. 15A.
[0010] For consistency of product design, a system is used in which
the development of the volute initiates at slice "A" (FIG. 4),
which is defined as the datum for the remainder of the volute. The
datum, slice "A", is defined as the slice at an angle of "P"
degrees above the "X-axis of the turbine housing containing the
"X"-axis, "Y"-axis and "Z"-axis specifications of the volute
shape.
[0011] The size and shape of the volute is defined in the following
manner: The widely used term A/R represents the ratio of the
partial area at slice "A" divided by the distance from the centroid
(161) of the shaded flow area to the turbo centerline. In FIG. 8A,
the position of the centroid (161) determines the distance R.sub.A
to the turbo centerline. For different members of a family of
turbine housings, the general shape remains the same, but the area
at slice "A" is different, as is the distance R.sub.A. The A/R
ratio is generally used as the "name" for a specific turbine
housing to differentiate that turbine housing from others in the
same family (with different A/R ratios). In FIG. 8A, the volute is
that of a typical divided turbine housing which forces the shapes
of the volutes to be reasonably triangular and equal in area. In
the case of a twin flow design (as depicted in FIG. 8A) the areas
at slice "A" for both volutes are the same. The centroids (160,
161), of the areas are at the same radius R.sub.A. The average
centroid, (163), is on the turbine housing centerline at the same
radius R.sub.A since the individual volutes at this section are
symmetrical about the divider wall.
[0012] Slice "A" is offset by angle "P" from the "X"-axis. The
turbine housing is then geometrically split into equal radial
slices (often 30.degree., thus at (30x+P.degree.)), and the areas
(A.sub.A-M) and the radii (R.sub.A-M), along with other geometric
definitions, such as corner radii are defined. From this
definition, splines of points along the volute walls are generated,
thus defining the full shape of the volute. The wall thickness is
added to the internal volute shape, and, through this method, a
turbine housing is defined.
[0013] The theoretically optimized volute shape for a given area is
that of a circular cross-section since it has the minimum surface
area which minimizes the fluid frictional losses. The volute,
however, does not act on its own, but is part of a system; so the
requirements of flow in the planes from slice "A", shown in FIG. 4,
to the plane at slice "M", and from "M" to the tongue, influence
the performance of the turbine stage. These requirements are often
compromised due to other demands such as architectural requirements
(space availability) outside the turbine housing, method of
location and mounting of the turbine housing to the bearing
housing, and the transition from slice "A" to the turbine foot
(51), which combine to force turbine housing volutes to be of
rectangular or triangular section, as well as in circular, or
combinations of all shapes. The rectangular shape of the volute in
FIG. 1, showing a section "D-K", is a result of the requirement to
not only to fit VTG (31) vanes into the space such that the flow is
optimized through the vanes and that the vanes can be moved and
controlled by devices external to the turbine housing, but also to
minimize the outline of the turbine housing so the turbocharger
fits on an engine.
[0014] The turbine housing foot is usually of a standard design as
it mates to exhaust manifolds of many engines. The foot can be
located at any angle to, or position relative to, the "volute". The
transition from the foot gas passages to the volute is executed in
a manner which provides the best aerodynamic and mechanical
compromise.
[0015] The roughly triangular shape of the volutes in FIG. 2, taken
at the same sections as those above, is the more typical volute
geometry for fixed and wastegated turbine housings. The addition of
the divider wall (25) is to reduce aerodynamic "cross-flow" between
the volutes in an effort to maintain pulse flow, from a divided
manifold (36), to harvest the pulse energy in the work extracted by
the turbine wheel. The pressure pulses in the exhaust manifold are
a function of the firing order of the engine.
[0016] In commercial practice, turbine housings are typically
designed in families (typically 5 to 7 in a family) which, in a
given family, use turbine wheels of the same diameter, or a group
of wheels with close to the same diameter. They may use the same
turbine foot size, although this feature is sometimes customer
driven. For example, a family of turbine housings for a 63mm
turbine wheel may cover a range of A/Rs from 1.8 to 2.2. FIG. 5
depicts the area schedule for three volutes of a family. The
largest volute is a 1.2 A/R volute, represented by the dotted line
(45). The smallest volute is a 0.8 A/R volute; represented by the
dashed line (46), and the mean volute, in the middle of the family,
represented by the solid line. The X-axis depicts the angle of the
slice from 30.degree. (section "A") to 360.degree. (the tongue);
the Y-axis depicts the area of the section at the respective angle.
Typically there is an 8 to 10% difference in cross-sectional area
(in the given case with 12 areas), at slice "A", from one A/R to
the next A/R in a design family. The volute outer wall of the
largest A/R (45) of FIG. 5 is shown in FIG. 4 as the inner surface
of the volute wall (40), and the smallest A/R (46) of FIG. 5 is
shown in FIG. 4 as surface (41).
[0017] Some turbine wheels are specifically designed to harness
this pulse energy and convert it to rotational velocity. Thus the
conversion of pressure and velocity from the exhaust gas for a
pulse flow turbine wheel in a divided turbine housing is greater
than the conversion of pressure and velocity from a steady state
exhaust flow to the turbine wheel velocity. This pulse energy is
more predominant in commercial Diesel engines, which operate at
around 2200 RPM with peak torque at 1200 to 1400 RPM, than in
gasoline engines, which operate at much higher rotational speed,
often up to 6000 RPM, with peak torque at 4000 RPM, such that the
pulse is not as well defined.
[0018] The basic turbocharger configuration is that of a fixed
turbine housing. In this configuration, the shape and volume of the
turbine housing volute is determined at the design stage and cast
in place. Most Diesel turbine housings are of the divided variety
with a radial divider wall (25) as seen in FIG. 2 separating the
two volutes to maintain the pulse energy to the turbine wheel. The
divider wall length is typically such that the inner bound is
approximately at the base circle. The closer the tip of the divider
wall is to the base circle, the greater the preservation of pulse
energy but the greater propensity for cracking of the casting in
the divider wall. The reasons for this cracking are many but
predominant are the dross which is pushed through the pattern at
the casting process which means that the integrity of the material
near the tip of the divider wall is less than optimal, and the
second is the fact that the temperature distribution around the
volutes causes the casting to want to "unwind". The thermal forces
generating the "unwinding" of the turbine housing are resisted by
the vertical divider wall, the resultant being cracking in the
wall. While a crack does little physical damage, the next step in
cracking is for pieces of cast iron divider wall to separate from
the casting and be ingested by the turbocharger or engine which can
cause terminal damage.
[0019] The next level of sophistication, after that of the fixed
turbine housing, is that of a wastegated turbine housing. In this
configuration, the volute is cast in place, as in the fixed
configuration above. In FIG. 2, the wastegated turbine housing
features a port (54) which fluidly connects the turbine housing
volute (49) to the turbine housing exducer (52). Since the port on
the volute side is upstream of the turbine wheel (70), and the
other side of the port, on the exducer side, is downstream of the
turbine wheel, flow through the duct connecting these ports
bypasses the turbine wheel (70), thus not contributing to the power
delivered to the turbine wheel.
[0020] The wastegate in its most simple form is a valve (55) which
can be a poppet valve, or a swing type valve similar to the valve
in FIG. 2. Typically these valves are operated by a "dumb" actuator
which senses boost pressure or vacuum to activate a diaphragm,
connected to the valve, and operates without specific communication
to the engine ECU. The function of the wastegate valve, in this
manner, is to cut the top off the full load boost curve, thus
limiting the boost level to the engine. This, in effect reduces the
effective flow to the turbine, when desired (e.g. to prevent
overdriving of the turbine), while allowing the full range of the
turbine housing flow to the turbine wheel when full flow is
desired. The wastegate configuration has no effect on the
characteristics of the boost curve until the valve opens. More
sophisticated wastegate valves may sense barometric pressure or
have electronic over-ride or control, but they also have no effect
on the boost curve until they actuate to open or close the
valve.
[0021] FIG. 6A depicts the boost curve (65) for a fixed geometry
turbine housing or a wastegated turbine housing in which the
wastegate valve did not open. The X axis depicts mass flow; the Y
axis depicts the pressure ratio. FIG. 6B depicts the boost curve
(67) for a wastegated turbine housing of the same A/R as that for
FIG. 6A in which the wastegate valve opened. In FIG. 6B, it can be
seen that the lower shape (62) of the boost curve (67) is exactly
the same as the boost curve (65) in FIG. 6A to the point (66) at
which the valve opens. After this point, the boost curve is flat.
While a wastegate can be used to limit boost levels, its turbine
power control characteristics are rudimentary and coarse.
[0022] A beneficial byproduct of wastegated turbine housings is the
opportunity to reduce the A/R of the turbine housings. Since the
upper limit of the boost is controlled by the wastegate, a
reduction in A/R can provide better transient response
characteristics, while still controlling the upper limit. However,
if the wastegated turbocharger has a "dumb" actuator, which
operates on a pressure or vacuum signal only and is operated at
altitude, then the critical pressure ratio at which the valve opens
is detrimentally affected. Since the diaphragm in the actuator
senses boost pressure on one side and barometric pressure on the
other, the tendency is for the actuator to open later (since the
barometric pressure at altitude is lower than that at sea level)
resulting in over-boost of the engine. By introducing a smaller A/R
turbine housing to take advantage of the wastegate, this A/R
reduction also reduces the flow range of the turbine stage.
[0023] Engine boost requirements are the predominant drivers of
compressor stage selection. The selection and design of the
compressor is a compromise between: the boost pressure requirement
of the engine; the mass flow required by the engine; the efficiency
required by the application; the map width required by the engine
and application; the altitude and duty cycle to which the engine is
to be subjected; the cylinder pressure limits of the engine;
etc.
[0024] The reason this is important to turbocharger operation is
that the addition of a wastegate to the turbine stage allows
matching to the low speed range with a smaller turbine wheel and
housing. Thus, the addition of a wastegate brings with it the
option for a reduction in inertia. Since a reduction in inertia of
the rotating assembly typically results in a reduction of
particulate matter (PM), wastegates have become common in
on-highway vehicles. The problem is that most wastegates are
somewhat binary in their operation, which does not fit well with
the linear relationship between engine output and engine speed.
[0025] The next level of sophistication in boost control of
turbochargers is the VTG (the general term for variable turbine
geometry). Some of these turbochargers have rotating vanes and some
have sliding sections or rings. Some titles for these devices
include: variable turbine geometry (VTG); variable geometry turbine
(VGT); variable nozzle turbine (VNT); or simply variable geometry
(VG).
[0026] VTG turbochargers utilize adjustable guide vanes (FIGS. 3A
and 3B) rotatably connected to a pair of vane rings and/or the
nozzle wall. These vanes are adjusted to control the exhaust gas
backpressure and the turbocharger speed by modulating the exhaust
gas flow to the turbine wheel. In FIG. 3A the vanes (31) are in the
minimum open position. In FIG. 3B the vanes (31) are in the maximum
open position. The vanes can be rotatably driven by arms engaged in
a unison ring, which can be located above the upper vane ring. For
the sake of clarity, these details have been omitted from the
drawings. VTG turbochargers have a large number of very expensive
alloy components, which must be assembled and positioned in the
turbine housing so that the guide vanes remain properly positioned
with respect to the exhaust supply flow channel and the turbine
wheel over the range of thermal operating conditions to which they
are exposed. The temperature and corrosive conditions force the use
of exotic alloys in all internal components. These materials are
very expensive to procure, machine, and weld (where required).
Since the VTG design can change turbocharger speed very quickly,
extensive software and controls are a necessity to prevent unwanted
speed excursions. This translates to expensive actuators. While
VTGs of various types and configurations have been adopted widely
to control both turbocharger boost levels and turbine backpressure
levels, the costs of the hardware and implementation are high.
[0027] The cost of a typical VTG, in the same production volume, is
from 270% to 300% the cost of the same size, fixed geometry
turbocharger. This disparity is due to a number of pertinent
factors from the number of components, the materials of the
components, the accuracy required in the manufacture and machining
of the components, to the speed, accuracy, and repeatability of the
actuator. The chart in FIG. 7 shows the comparative cost for the
range of turbochargers from fixed to VTGs. Column "A" represents
the benchmark cost of a fixed turbocharger for a given application.
Column "B" represents the cost of a wastegated turbocharger for the
same application, and column "C" represents the cost of a
conventional VTG for the same application.
[0028] Thus it can be seen that, for both technical reasons and
cost drivers there needs to be a relatively low cost turbine flow
control device which fits between wastegates and existing VTGs in
terms of cost.
SUMMARY OF THE INVENTION
[0029] The present invention relates to a simplified, low cost,
variable geometry turbocharger, and more particularly, a turbine
flow controlling device, which uses a divided turbine housing with
assymetric volute A/Rs coupled with a flow modulation device to
change the effective exhaust mass flow to the turbine wheel, while
increasing the turbine stage flow range. By controlling the mass
flow of exhaust, which the turbine housing directs to the turbine
wheel, with a set of asymmetrically configured volute cross
sectional areas, and controlling the flow through the two volutes
with a relatively simple flow controlling device, the flow range
can be both broadened and controlled in a manner exceeding the
range available with a symmetrically configured volute cross
sectional areas without the flow controlling device.
[0030] The variable geometry turbocharger is simplified yet able to
maintain pulse energy. In a first embodiment, a turbine housing is
provided with a pivoting flow control valve which pivots about a
point near the entry to the turbine housing. By moving the valve
about the pivot point, the flow through the turbine housing is
increasingly blocked at the large volute whereby flow is biased to
the small volute and continues from there to the turbine wheel,
thus causing the turbine housing, through effective loss of the
larger volute, to act as a smaller A/R turbine housing. In the
second embodiment of the invention, a rotating butterfly-configured
flow control valve within the volute, which pivots about the center
of the valve blade, pivots about said centerline to vary the flow
from the large volute to the small volute and on to the turbine
wheel, thus causing the turbine housing to act as a smaller A/R
turbine housing.
[0031] Testing by the inventors determined that a 60/40 split of
"A"-section areas with the hub side at 60% and the shroud side at
40% produced a desirable mass flow split with the restrictor valve
fully open. The asymmetric turbine housing has a larger left or hub
side volute (48) and a smaller or shroud side volute (49) situated
axially about a divider wall (25).
BRIEF DESCRIPTION OF THE DRAWINGS
[0032] The present invention is illustrated by way of example and
not limitation in the accompanying drawings in which like reference
numbers indicate similar parts, and in which:
[0033] FIG. 1 depicts the section for a typical VTG
turbocharger;
[0034] FIG. 2 depicts a pair of sections of a typical wastegated
turbocharger;
[0035] FIGS. 3A,B depict a pair of sections of a typical VTG
turbocharger;
[0036] FIG. 4 depicts a section of a typical fixed turbine housing
showing construction radial lines;
[0037] FIG. 5 is a chart of cross-sectional area development;
[0038] FIGS. 6A,B depict the compressor maps for a typical fixed,
and a wastegated turbocharger;
[0039] FIG. 7 is a chart showing turbocharger relative costs;
[0040] FIGS. 8A,B depict the sections of two volute types at slice
"A";
[0041] FIG. 9 depicts a view of an assymetric turbine housing on a
manifold;
[0042] FIGS. 10A,B depict two section views of the restrictor
device on a circumferentially divided housing;
[0043] FIGS. 11A,B depict two views of a variation of a restrictor
device on a circumferentially divided housing;
[0044] FIG. 12 depicts a view of the cracking in a turbine
housing;
[0045] FIG. 13 depicts a section view of closed slots in a turbine
housing divider wall;
[0046] FIG. 14 depicts a section view of open slots in a turbine
housing divider wall;
[0047] FIGS. 15A,B depict two views of the third embodiment on a
radially divided housing;
[0048] FIG. 16 is a chart depicting mass flow;
[0049] FIG. 17 depicts the curtain areas for a sample of production
turbine wheel diameters; and
[0050] FIG. 18 depicts the relationship between the crossflow area
and D.sub.3 for different turbine stages.
DETAILED DESCRIPTION OF THE INVENTION
[0051] As discussed above, variable geometry mechanisms tend to
double and more the cost of the basic turbocharger. The inventors
sought the ability to modulate the exhaust flow to the turbine
wheel in a more cost-effective manner. Therefore the inventors
experimented with designs with divided volute areas, combined with
a flow resistance device to provide both a cost and technically
effective alternative for controlling the required wide range of
exhaust gas flow to the turbine. In addition to the above gains,
the inventors sought to provide a turbocharger matched to low flow
regimes that would provide optimized turbo (and thus engine)
transient response for low flow while still capable of delivering
the high flows demanded by the engine in other than low flow
conditions in the same, cost-effective turbocharger. This target
keeps the gas velocities in the sweet spot which maximizes the
stage efficiencies.
[0052] When a turbocharger is matched to the maximum flow
requirement of an engine, the flow requirements across the entire
engine operating regime are met. The problem is that matching the
turbocharger to the maximum flow requirement means that the size of
the turbine housing volute (and thus flow) is way too large for low
engine flow regimes. The turbocharger's transient response
characteristics are sluggish because the entire volute has to be
filled in order to deliver flow to the turbine wheel. Since
reducing the A/R of a turbocharger turbine housing to match the low
flow requirement would mean that the turbocharger, operating within
typical speed constraints, is not capable of providing sufficient
flow for the high flow requirement of the upper end of the engine
operating regimes, the inventors recognized the need to provide a
novel variable geometry turbocharger. Furthermore with today's EGR
(Exhaust Gas Recirculation) requirements, OEMs are running large
amounts of EGR at part load (say 40% load) and no EGR at high speed
yet they still desire, from a market standpoint, to deliver
best-in-class power at full load. High EGR at low speed or part
load requires low mass flow. Best-in-class power at the rated point
with no EGR requires high mass flow so it can be seen that the
turbine mass flow range needs to be capable of matching the flow
requirements at these two extremes.
[0053] Turbine housing volute shapes and dimensions are defined by
the area of section "A", and all features and dimensions downstream
of section "A" are controlled by the features and dimensions at
section "A". This system is used for consistency of design within
the turbochargers designed and produced by a turbocharger
manufacturer.
[0054] In accordance with the present invention, the inventors
provide a novel turbine design able to produce a wider turbine flow
range than would be available with volutes of equal area.
[0055] By controlling the mass flow through the turbine housing the
inventors sought to control the mass flow of gas passing through
the turbine housing to the turbine wheel. When the engine is
operating in the low speed, low load condition, the boost level
required to supply the required combustion gas (air) is relatively
low. When the engine is in the high speed, high load condition, the
boost level required to supply the engine under these load
conditions is high. When the engine is transitioning from low load
conditions to high load conditions, the turbocharger is required to
supply an increasing volume of air at an increasing pressure ratio.
Since the compressor stage is driven by the turbine stage, the mass
flow of exhaust required to meet the engine (and thus the
compressor) requirements has to change. That is, at the low load,
low speed engine condition, the engine exhaust output, in terms of
mass flow is low. At the high load, high engine speed condition,
the engine output, in terms of mass flow is high. In the transition
stage the exhaust mass flow has to change from low to high.
[0056] The problem is that the turbine stage must be matched to
both of the above-described basic engine conditions, in addition to
the requirements for EGR to allow the turbocharger to supply the
requested flow and pressure ratio at any of these conditions. In
order to force the turbocharger to change speed quickly, one
experienced in the art would select a turbocharger with a small A/R
turbine housing. In order to supply the required flow and pressure
ratio at the high load, high speed condition one would select a
turbocharger with a larger A/R turbine housing. The former small
A/R turbine housing will provide good transient response
characteristics, but insufficient mass flow to the turbine stage to
generate the high speed, high load compressor requirement. The
latter, large A/R turbine housing will provide the mass flow
requirement to the turbine stage for the high speed, high load
boost requirement but will not provide acceleration to the turbine
wheel sufficiently quickly to produce acceptable transient
response.
[0057] Obviously, it would be nice to have a system with two
turbochargers, one larger and one smaller, and to be able to switch
between the two. However, such a system would be expensive, would
represent a large "heat sink", would take much space in the engine
compartment, and would add to the mass of the vehicle.
[0058] A properly matched small A/R turbine stage acting alone will
provide acceptable transient response albeit at the expense of
higher backpressure, compared to that of a turbine stage matched to
the high load, high speed condition. In a non-EGR engine having
high back pressure is a negative to the pressure differential
across the engine and thus the efficiency of the engine. In a high
pressure loop EGR engine configuration (as against a low pressure
loop EGR engine configuration) the high back pressure in the
exhaust system is part of the solution to drive the exhaust gas
from the exhaust side of the engine into the inlet side of the
engine which is seeing boost pressure. A large turbine housing A/R
for a given set of engine parameters will develop lower exhaust
back pressure than would a smaller A/R turbine housing under the
same set of engine parameters. So being able to change the
effective A/R of the turbine housing allows a single turbocharger
to meet both the flow and back pressure requirements of a low
speed, low load condition, and a high speed, high load
condition.
[0059] By controlling the mass flow of exhaust, which the turbine
housing directs to the turbine wheel, with a set of asymmetrically
configured volute cross sectional areas, and controlling the flow
through the two volutes with a relatively simple flow controlling
device the flow range can be both broadened and controlled in a
manner exceeding the range available with a symmetrically
configured volute cross sectional areas without the flow
controlling device.
[0060] After initially experimenting with a symmetrically divided
volute turbine housing, the inventors next experimented with
asymmetric divided turbine housings, and determined that by
substituting one of the volutes with another volute of a smaller
A/R that the flow range would drop and the maximum flow range
through that volute would also drop. Similarly by replacing one
volute with another volute of a larger A/R, the maximum flow range
of that volute would rise. By putting together a combination of a
larger and a smaller volute, and controlling the degree of blockage
of the larger volute, the flow range of the inventive turbine
housing exceeds that of the original prototype turbine housing with
symmetrical divided volute. In FIG. 16 the bars (22) with the
horizontal hatches represent the mass flow of a turbine housing
with equal (50-50) area volutes and the bars (23) with the vertical
hatches represent the mass flow of a turbine housing with
assymetric (60-40) area volutes. While the mass flows with the
restrictor valve fully open are equal to one another, when the
restrictor valve is in the fully closed position (i.e., effectively
blocking the flow into the larger volute) the mass flow of the
assymetric A/R configuration is less than the mass flow of the
equal A/R configuration. The sum of the areas at section "A" for
both configurations is within 0% to 3%, while the change in mass
flow is in the range of 10% to 13%.
[0061] FIG. 8A depicts a typical symmetrical turbine housing volute
configuration in which the centroids (160, 161) of the two volutes
are at the same radius, R.sub.A, from the centerline. Since the
turbine housing is symmetrical, the effective centroid (163) of
both volutes lies in the divider wall between the volutes. FIG. 8B
depicts an example in which the left volute is an A/R size larger
than that of the symmetrical turbine housing of FIG. 8A, and the
right volute is an A/R size smaller in area, thus 2 A/R sizes
smaller than that of the left volute. In this case the centroid of
the right volute is at a radius R.sub.C from the centerline and
axially closer to the centerline of the turbine housing. The
centroid of the left volute is at a radius R.sub.B from the
centerline and axially further from the centerline of the turbine
housing. The effective centroid (164) of both volutes in the
turbine housing is now at a radius R.sub.D offset to the left of
the divider wall centerline.
[0062] To produce an optimal asymmetric turbine housing the
inventors looked at several options of volute sizes from one volute
A/R up, or one volute down from equally sized volutes, to going
from equal sized volutes and making the hub-side one volute A/R up
and the shroud-side one volute A/R down. Testing by the inventors
determined that the latter solution, which was a 60/40 split of
"A"-section areas with the hub side at 60% and the shroud side at
40% produced the desired mass flow split with the restrictor valve
fully open.
[0063] In all divided turbine housings there exists a cross-flow
"curtain" between the tip of the divider wall, at its minimum
diameter, and the tips of the turbine wheel. To minimize turbine
wheel excitation caused by the action of the rotating turbine wheel
blades passing the static tongue, (26) FIG. 4, at the start of the
divider wall, as a rule of thumb, the depth of the divider wall
typically does not extend closer to the turbine wheel blade tips
than a ratio of from 120% to 150% of the turbine wheel tip diameter
D.sub.3. This ratio of D.sub.bc/D3 is usually determined by company
design rules and technical goals. The diameter D.sub.bc is known as
the base circle. Because turbine housing divider wall tips are
prone to cracking, due to the dross of the molten cast iron being
forced into the tip, the innermost, or minimum diameter of the
divider wall is typically not less than 120% to 150% of the turbine
wheel tip diameter. This "curtain" between the divider wall and
turbine wheel allows for crossflow of exhaust gas, between the two
volutes as well as cross-talk between the pulses in the exhaust
flow, the latter being the reason for having the divider wall in
the first place.
[0064] For a turbine stage with a base circle, the diameter of
which is 120% of the turbine wheel diameter, there exists a cross
flow "curtain", with an area which is from 70% to 105% of the area
of both volutes, in a symmetric configuration, at Section "A", for
a turbine housing family of 5 A/Rs. For a turbine stage with a base
circle, the diameter of which is 150% of the turbine wheel
diameter, the cross flow "curtain", has an area which is from 199%
to 299% of the area of both volutes, in a symmetric configuration,
at Section "A", for a turbine housing family of 5 A/Rs. From this
analysis it can be seen that the curtain area can provide a very
large cross sectional area for crossflow from one volute to the
other.
[0065] Since the curtain area is a function of both the turbine
wheel diameter D.sub.3 and the minimum position of the divider wall
D.sub.bc, the curtain area varies for different values of D.sub.3.
FIG. 17 depicts the curtain areas for a sample of production
turbine wheel diameters from 64 mm to 96 mm. The curtain areas
(133) are bracketed between or limited by an upper bound line (131)
and a lower bound line (132). As would be expected the range of
curtain areas increases as D.sub.3 increases. In FIG. 17 the
turbine wheel diameters D.sub.3 are shown as (123) and a line (124)
depicts the trend of turbine wheel diameters, D.sub.3 for the
analyzed turbos. This chart also contains ratios of
D.sub.bc/D.sub.3 ranging from 1.25 to 1.35.
[0066] The inventors determined through testing with a 64 mm
turbine wheel, that for an asymmetrically configured 60/40 volute
combination with a restrictor valve, the optimum cross flow area,
which includes ports in the divider wall plus the area of the base
circle "curtain" (determined by the difference between the area
under D.sub.bc minus the area under D.sub.3), was an area with a
ratio of 289.6% that of a single, symmetrical, volute cross
sectional area at Section "A" (i.e., half the area at Section "A").
This compares to the typical cross flow area of the same sized
turbine housing with no slots or ports, with the same ratio of
D.sub.bc/D.sub.3 which has a crossflow area of only 182.6% of half
the area at Section "A".
[0067] As in the case of the relationship between the curtain areas
and D.sub.3 and D.sub.bc, the total cross flow area (122) is
affected by not only D.sub.3 and D.sub.bc but also the variation in
the area of a single volute at Section "A". The crossflow areas
(122) are bound by an upper bound line (126) and a lower bound line
(127) The chart in FIG. 18 depicts the relationship between the
crossflow area and D.sub.3 (123) for different turbine stages
analyzed by the inventors.
[0068] To select a crossflow area, determine the value of D.sub.3,
the diameter of the turbine wheel in inches. The example is that of
a 76 mm (2.992'') turbine wheel, shown as a horizontal line (128).
From the turbine wheel diameter the vertical line (129) which
intersects the turbine wheel diameter (123) cuts the lower bound
line (127 and the upper bound line (128). The crossflow area is
depicted as the vertical segment (130) of the vertical line (129)
between the lower and upper bound lines (127 and 126).
[0069] The formula generating the data points, which are plotted on
the charts shown in FIGS. 17 and 18, could be as follows:
? = .pi. / 4 ( D bc 2 - D 3 2 ) ##EQU00001## Ratio = A CF ? = A
curtain + A slot A 1 / 2 Asect = ? + ? A 1 / 2 Asect ##EQU00001.2##
? indicates text missing or illegible when filed ##EQU00001.3##
[0070] As depicted in FIG. 9, in the low flow condition, a
pivotable valve member (72) is actuated to generate a flow
restriction to the hub, or bearing housing side, larger volute (48)
which forces the flow from the manifold through the shroud, or
exducer side, smaller volute (49) to the turbine wheel (70).
[0071] The asymmetric turbine housing has a larger left or hub side
volute (48) and a smaller right or shroud side volute (49) situated
axially about a divider wall (25). A flow restrictor, in this case
a pivotable valve member (72) is constrained within the joining
faces of the manifold center section foot (37) and the turbine
housing foot (51). While the inventors chose this configuration for
cost and technical reasons the restrictor could be located in the
hub side exhaust manifold passage (34).
[0072] As depicted in FIG. 10B, in the high flow condition the
pivotable valve member (72) is in the central position, which
biases neither the larger, hub side volute (48) nor the smaller,
shroud side volute (49), to allow the maximum flow to the turbine
wheel. In this high flow condition the pivotable valve member (72)
of the flow restrictor device is aligned with the divider wall (25)
of the turbine housing downstream the turbine foot (51). In the
minimum flow position (depicted by the dashed line in FIG. 10B),
the pivotable valve member (72) is rotated, preferably by a force
exerted on the actuating arm (73), about the axis (74, 78) of the
device, towards the closed position, such that it restricts the
exhaust flow to the large volute (48) and causes the exhaust flow
to flow through the small volute (49). The flow restrictor can be
modulated to any position between fully open and fully closed.
[0073] A sectioned view of this version of the flow restrictor
device is shown in FIG. 10A. In this view one can see that in the
preferred embodiment of the invention, the flow restrictor blade is
fabricated with two cylindrical bearing surfaces for pivoting and
an actuating arm (73) for position control. One side of the cavity
formed in the joint of the turbine housing foot (51) and exhaust
manifold foot (37), which houses a bearing surface is a blind bore
(77) while the other (75) is an open bore. On the open bore side a
piston ring (76) provides not only an axial alignment for the flow
restricting device, but also a gas seal. The placement of the
actuating arm (73) can be optimized to meet architectural
constraints.
[0074] The inventors realized that the ratio of
boost-to-backpressure as well as the backpressure alone increased
as a function of engine speed and load, at both sea level and at
altitude, which made the flow restrictor device in the exhaust
system an ideal controlling parameter. When the pivotable flow
restrictor is rotated towards the closed position, the turbine
housing acts as if it were a smaller A/R turbine housing than would
exist with the flow restrictor in the open position. This causes
the exhaust backpressure to rise which is necessary for EGR flow
from the exhaust side of the engine to the inlet side of the
engine. Thus the rotation of the flow restrictor can be used to
develop a pressure differential (from the exhaust side of the
engine, to the inlet side of the engine) to aid EGR flow from the
exhaust side of the engine to the inlet side of the engine.
[0075] In the first embodiment of the invention, the effective mass
flow to the turbine wheel is controlled by a flow restrictor which
pivots about a point in the turbine housing inlet or foot such that
in the open position the pivotable valve member (72) of the flow
restrictor is in line with the divider wall (25) of the turbine
housing minimizing the restriction to the exhaust flow. As more
restriction, or less mass flow to the turbine wheel, is required
the pivot arm (73) is actuated to rotate about its axis (74, 78)
causing the pivotable valve member (72) to impede the flow of
exhaust gas to the large volute (48), which causes a modulatable
reduction in mass flow to the turbine wheel.
[0076] In a variation to the first embodiment of the invention, as
depicted in FIGS. 11A and 11B, the flow restrictor takes the form
of a butterfly valve (80) which reduces the moment on the pivot arm
(81) enabling the potential use of a lower force, and thus lower
cost actuator. In FIG. 11A, which is a section view of the first
variation of the first embodiment of the invention the
configuration of the bearing surfaces and piston ring are the same
as in the first embodiment. In FIG. 11B, the pivot location, in the
case of the butterfly configuration, is approximately in the center
of the flow path to the large, or hub side volute (48) so that
rotation of the butterfly (77), about its axis (74, 84) provides an
adjustable flow restriction to the hub side volute (48) biasing the
flow to the shroud side volute (49). In the case of this variation
to the first embodiment of the invention, in the minimum flow
restriction position, as depicted in FIG. 11B, the butterfly valve
is aligned with the flow through the volute, such that the tips of
the butterfly valve close or shadow the hub side volute (48). A
butterfly valve solution has the advantage of low actuation loads,
since the moments on the two sides of the pivot cancel each
other.
[0077] When the flow restrictor is in the partially open position,
flow from the shroud side (smaller) volute (49), to the hub side
(larger) volute (48) can be further facilitated by either
shortening the length of the divider wall (25), or by fabricating
slots into the divider wall.
[0078] Typically, in the commercial Diesel world, where the product
can be expected to run for a million miles, turbine housing divider
walls are prone to cracking. The inventors realized an opportunity
to mechanically minimize this propensity for cracking in the
divider wall by introducing pre-cast stress-relieving features in
the divider wall. FIG. 12B depicts a turbine housing viewed along
section A-A of FIG. 12A. This sectioning is typically performed to
evaluate the condition of the turbine housing following a thermal
cycling qualification test in which the turbocharger is subjected
to extreme temperature cycling in an effort to determine its
resistance to cracking. In FIG. 12B the cracks (87) depicted are
typical of a commercial Diesel type turbine housing in the divider
wall area.
[0079] The inventors surmised that if "stress relievers" in the
form of slots or ports were cast into the divider wall then these
ports would not only minimize the propensity for cracking but also
provide a flow path from the un-modulated shroud side volute to the
modulated hub side volute under conditions of partial to full
restrictor valve closure. This additional flow path provides flow
to the turbine wheel over a greater circumferential distance or
area than would be possible without the slots or ports.
[0080] In the second embodiment of the invention, as depicted in
FIG. 13, the effective mass flow to the turbine wheel is controlled
by a flow restrictor in an assymetric turbine housing with
crossflow ports (88) fabricated in the divider wall. In the
preferred second embodiment of the invention the area of said ports
is bound by the leading edge radial (89), the trailing edge radial
(90), the inner edge circular segment (92) and the outer edge
spiral (91) for each port. The sum of the areas of the crossflow
ports in the turbine housing is approximately equal to the area of
the modulated volute at section "A". What is important is the sum
of the areas of the ports not the geometry of the ports.
[0081] In a variation to the second embodiment of the invention, as
depicted in FIG. 14, the effective mass flow to the turbine wheel
is controlled by a flow restrictor in an assymetric turbine housing
with crossflow slots (95) fabricated in the divider wall. In the
preferred second embodiment of the invention, the area of said
slots is bound by the leading edge radial (98), the trailing edge
radial (99), the outer edge spiral (91) for each port and the inner
boundary by the base circle (71). The sum of the areas of the
crossflow slots in the turbine housing is approximately equal to
the area of the modulated volute at section "A". What is important
is the sum of the areas of the slots and the crossflow area inside
the divider wall tip, not the geometry of the slots. In the
preferred second embodiment of the invention the outer bound (97)
of the slot can be characterized by a keyhole configuration as the
outer termination of the slot to minimize the propensity of the
slot to be a stress raiser and initiate cracking.
[0082] Multiple flow turbine housings with the volute divider wall
parallel to the turbocharger axis, i.e., axial surfaces rather than
radial surfaces as in the basic twin flow turbine housing are not
uncommon. The inventors saw the opportunity to use similar logic
for multiple flow turbine housings with assymetric volute areas
accompanied by a flow restrictor to further cost effectively widen
the flow range of a turbine stage with this type of turbine
housing.
[0083] In the third embodiment of the invention, a triple flow
turbine housing as depicted in FIG. 15A is preferably used. Two
axial volute divider walls (103, 104) are fabricated into the
turbine housing such that the ratio of flows through the
unrestricted adjacent volutes, from outer to inner are
approximately 70% to 20% to 10%. These proportions can be varied
depending upon requirements. The ratio of flows is only important
in that the sum of the open areas of the modulated volutes is equal
to the area of the modulating restrictor valve. A flow restrictor
valve is provided. The flow restrictor valve pivots about a point
in the turbine housing inlet or foot such that in the open
position, the blade (89) of the flow restrictor is flush with the
turbine housing outer volute wall minimizing the restriction to the
exhaust flow. As more restriction, or less mass flow, to the
turbine wheel is required, the pivot arm (73) is actuated to rotate
about its axis causing the blade (89) to impede the flow of exhaust
gas to first the outer volute(s) followed by the center volute. In
this manner the effective mass flow to the turbine wheel is
controlled by a flow restrictor which enables a modulatable
reduction in mass flow to the turbine wheel.
[0084] In a variation to the third embodiment of the invention, the
dividing walls (106, 107) are slotted (108) to allow flow from the
outer volutes to reach the inner volutes and then the turbine wheel
(70). The slots (108) also allow for mass flow modulation but with
a more consistent and favorable flow distribution to the turbine
wheel.
[0085] Now that the invention has been described,
* * * * *