Note: Descriptions are shown in the official language in which they were submitted.
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ROTARY ENGINE AND METHOD FOR DETERMINING
ENGAGEMENT SURFACE CONTOURS THEREFOR
a. Field of the Invention
l0
The present invention relates to rotary positive
displacement engines and to methods for determining
engagement surface contours for use in the making of
rotary positive displacement engines.
b. Background
This invention concerns an advanced rotary positive
displacement engine having high power to mass ratio and
low production cost. The term "engine" as used in this
patent document is taken to be a device that converts one
form of energy into another. Hence, the term includes
both devices which impart energy to the fluid flow (e. g.
a pump) and those which employ the fluid flow to generate
an energy output (e.g. an external combustion engine for
providing a power source).
In the case of prior art combustion engines, the
reciprocating piston type is most widely used for its low
cost of production and efficient sealing, while the
turbine has shown that an external combustion engine may
offer greater power, partially from high speed. Rotary
engines such as the Wankel engine have shown higher
power-to-weight ratios than reciprocating engines but at
the expense of increased fuel consumption. The present
invention is a rotary device that offers many of the
advantages of these prior art devices without many of
their shortcomings.
In the case of pumps, there are many general types
of pump designs known, such as positive displacement,
centrifugal and impeller. Pumps of the positive
displacement type are typically reciprocating or rotary.
Many previous rotary combustion engine designs in turn,
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have been of the single plane type in which rotary motion
occurs about axes that are parallel to each other.
Prior forms of rotary pumps and combustion engines
have been limited in their efficiency, in part by
inherent limitations in their operating principles, and
also in many instances by their inability to establish a
seal between operating surfaces which is sufficient to
achieve a high degree of efficiency, yet which also
accommodates the physical characteristics of the fluid
which much pass therethrough.
Many of the deficiencies of prior types of rotary
pumps and engines have been negated by a positive
displacement engine which has been developed by Applicant
(referred to from time to time herein as a "CvR Engine")
For example, as the demand for higher performance
and higher efficiency are increased, machining techniques
have also been improving. Modern manufacturing
techniques such as EDM and wire EDM machining allow
dimensional accuracies on complex surfaces of within
.001". Even higher accuracy is expected to be possible
as new manufacturing equipment and techniques are
developed.
Different applications may require various
clearances and interferences between the surfaces not
always simply the closest possible fit . For example, the
movement of fluids with suspended particles may require
large enough sealing surface clearances to allow these
solids to pass through in the fluid film. In some
applications, such as irrigation pumping, these particles
may be in excess of .1". In other applications, such as
in the semi-conductor or medical industries, the particle
size can be as small as several microns. Hence, there
are many applications where it is essential to establish
a finite, precisely controlled gap between the two
sealing surfaces to provide a positive sealing surface
geometry (SSG) .
Similarly, where comparatively low tolerance
manufacturing techniques are used to produce lower
performance or less expensive designs, a sealing surface
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geometry (SSG) which allows for the inconsistencies of
the final surface. Higher tolerance machining techniques
will also benefit from a predetermined SSG to maintain a
minimum gap clearance or to prevent contact or binding of
the mating rotors. Hard coating of a suitable base
material also requires a pre coated surface geometry
which prevents the coated SSG from binding or
interfering.
Some applications may even benefit from an
interfering or "negative" SSG. Compressible or
deformable materials and coatings can provide increased
seal performance if they are designed to interfere with
the mating surface on the opposite rotor. This can be
accomplished by applying such a coating over a harder
base material having a negative SSG to bring the surface
back to a reduced negative SSG, so as or a positive SSG.
Another advantage made possible by an extremely
precise SSG is the establishment of a fluid film bearing
between the sealing surfaces. Fluid film bearings have
been used successfully in industry to replace ball
bearings or plain bearings in many applications. Fluid
films for bearings range from several ten thousandths of
an inch to several thousandths of an inch. Having a
fluid film between the sealing surfaces of the engine
rotors will decrease friction and wear, however,
establishing this fluid film requires a correctly
designed surface interface. If the surface interface has
a gap space which does not account for the other
variables which affect the fluid film; however, extra
friction and wear, as well as volumetric efficiency
compromises, may result.
An excessive clearance or gap between the sealing
surface, may lead to excessive leak-by, thereby
significantly impairing the overall efficiency of the
engine. For example, if excessive " backlash " develops
between the sealing surfaces of the CvRTM-type engine
described above, this can result in undesirable amounts
of leak-by.
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An additional concern is that for many applications
it is desirable for the engine to be highly efficient in
both forward and reverse directions of operation.
Consequently, if the sealing surfaces of the engine are
able to move apart and create an excessive backlash, due
to deficiencies in the desired SSG or for other reasons
the engine will be unsatisfactory for reverse operation.
Accordingly, there exists a need for a method for
determining the contours of the sealing surfaces of a
rotary engine (as defined herein) so that these will have
a precise, controlled gap during operation of the pump.
Furthermore, for manufacturing purposes, there exists a
need for a method for verifying that the correct contours
have been imparted to such surfaces. Still further,
there exists a need for an engine having such surfaces
arranged so that the proper gap will be maintained during
both forward and reverse operation.
SUMMARY OF THE INVENTION
The present invention is of the rotary positive
displacement type, but is in a class by itself. This
rotary positive displacement device is believed to be the
first rotary engine in which the axes of the moving parts
are offset from each other and the moving parts rotate at
a constant velocity relative to each other when they are
rotating at a constant velocity relative to the casing.
The engine is formed by a pair of facing rotors that are
axially offset from one another and whose faces define
chambers that change volume with rotation of the rotors.
An engine of this type defines a new class of
engines, and includes a minimum number of moving parts,
namely as few as two in total.
In one aspect of the invention, a pump includes a
pair of rotors, both housed on and preferably within the
same housing. The housing has an interior cavity having
a center. Each rotor is mounted on an axis that passes
through the center of the cavity, the respective axes of
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the rotors being at an angle to each other, with the
center of each rotor being at the center of the cavity.
The rotors interlock with each other to define chambers.
Vanes defined by a contoured face on one side of the vane
and a side face on the other side of the vane protrude
from the rotors. The contoured faces of the rotors are
defined so that there is constant linear contact between
opposing vanes on the two rotors as they rotate. The
side faces are preferably concave and extend from an
l0 inner end of one contoured face to the outer end of an
adjacent contoured face, equivalent to the tip of a vane.
The side faces and contoured faces define walls of
chambers that change volume as the rotors rotate. Ports
for intake and exhaust are preferably configured to have
shapes complementary to the intersecting vanes of the
rotors.
Also in accordance with the present invention, a
method is provided for determining a precise,
controllable gap between the sealing surfaces on the
rotors. These methods include both mathematical and
geometric processes, a well as methods for verifying that
the correct contours have been imparted to the surfaces.
Still further, in accordance with a preferred
embodiment of the invention, the vanes on the rotors are
provided with mirror-image contoured sealing surfaces
which both maintain the desired gap during operation by
reducing back-lash, and which also permit efficient
reverse operation of the engine.
These and other aspects of the invention will be
described in more detail in what follows and claimed in
the claims appearing at the end of this document.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is an isometric view of a master rotor and
slave rotor housed within a ported housing according to
one aspect of the invention;
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FIG. la is a top view of the master rotor of FIG. 6A
showing the result of removing material from the master
rotor between four vanes one face of each vane being
formed as shown in the preceding Figure;
FIG. lb and lc are a side view and isometric view
respectively of the master rotor of FIG. lA;
FIG. 2 is a schematic view showing the interior of
the housing of FIG. 1;
FIG. 3 is an end view, partially in section, of the
housing of FIG. 1;
FIG. 4 is a schematic view partially in section, of
the housing of FIG. 1 showing a cantilevered slave rotor
shaft;
FIG. 4a shows a further embodiment of an engine
according to the invention, in section, with vanes of
each rotor extending into the shaft of the other rotor;
FIG. 4b is a section showing the embodiment of FIG.
4a with part of the shaft of the slave rotor extending
around the master rotor;
FIG. 5 is a perspective view of an engine in
accordance with a further embodiment of the invention,
with the casing and the pump being shown separated to
expose the internal components thereof, this embodiment
of the invention having vanes with mirror image contoured
surfaces which maintain closer operating tolerances
between the vanes and also permit the engine to operate
in a reverse direction;
FIG. 6 is an elevational view of a first half of the
engine casing of FIG. 5, showing the port, seal and
bushing structure thereof in greater detail;
FIG. 7A is a side elevational view of the slave and
master rotor of the engine of FIG. 5, showing the
engagement of the contoured surfaces and the incidence
angle between the two rotors;
FIG. 7B is a top plan view of the master and slave
rotors of FIG. 7A, showing one of the chambers at its
point of maximum volume;
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FIG. 7C is a bottom plan view of the master and
slave rotors of FIG. 7A, showing the chamber at its point
of minimum volume;
FIGS. 8A-8E are a series of geometric figures
showing axes, distances, angles, vectors, and other
values associated with the mathematical determination of
the contoured surface contours in accordance with the
present invention;
FIGS. 9A-9D are a series of views of a visual model
l0 illustrating the method by which the contours of the
contoured surfaces are determined in the present
invention, by conceptual rotation of predetermined system
axes based on a predetermined mathematical relationship;
FIGS. l0A-lOD are a series of computer-generated
graphical images, illustrating the manner in which the
contours of the contoured surfaces are determined using
the mathematical relationship in accordance with the
present invention;
FIG. l0E is a perspective view of one of the rotors
2o in accordance with the present invention, where the
dashed line image showing the area of the contoured
surface having the contour which is generated as a result
of the steps shown in FIGS. l0A-lOD.
FIG. 11A is a geometric figure, similar to FIG. 8C,
showing a revised calculation of the contoured surface
contours to provide a modified tip-radius form having a
slightly flattened shape for enhanced wear
characteristics;
FIG. 11B is a partial, cross-sectional view of the
tip portion of a contoured surface formed in accordance
with the relationship shown in FIG. 11A;
FIG. 12 is a schematic view showing the relationship
of a series of mirror image contoured surfaces, somewhat
similar to those shown in FIGS. 7A-7C, with these being
configured to maintain a predetermined fluid film
thickness during operation and also to permit reverse
operation of the engine;
FIG. 13A is a partial, enlarged view of adjacent tip
portions of the mirror image contoured surfaces of FIG.
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12, showing the spacing between the tip surfaces in
greater detail;
FIG. 13B is a geometric diagram, similar to FIGS. 8C
and 11A, illustrating the mathematical determination of
the contoured surfaces having the clearances which are
shown in FIG. 13A;
FIG. 14 is an elevational, somewhat diagrammatic
view illustrating the determination of the engagement
surfaces in accordance with a geometric method which
corresponds to the mathematical processes illustrated in
FIGS. 8A-13B, in which the gap between the sealing
surfaces is controlled by the amount of offset between
the apex of a hypothetical cone and the intersection of
the axes of the rotors upon which the surfaces are
formed;
FIGS. 15A-15F are a series of perspective, somewhat
schematic views illustrating the manner in which the
contoured surfaces on the rotor are formed in accordance
with the method-of FIG. 14, with the movements of the
hypothetical cone corresponding somewhat to those of a
tool for machining the surfaces;
FIGS. 16A-16B are perspective, somewhat schematic
views showing a first rotor, formed as shown in FIGS.
15A-15F, in predetermined angular engagement with a
second rotor having corresponding engagement surfaces,
showing the sealing surface gap which is formed by the
offset between the two sets of surfaces;
FIG. 17 is a schematic, end view of adjacent sealing
surfaces such as those which are shown in FIGS. 16A-16B,
illustrating the manner in which the gap between the
sealing surfaces is increased or decreased by rotation of
the rotor relative to the hypothetical cone which is
shown in FIGS. 14-15F;
FIG. 18 shows a series of schematic views similar to
FIG. 17, showing the different forms of parallel and
angular interfacial gaps which can be formed between the
sealing surfaces by adjusting variable factors in the
methods which are illustrated in FIGS. 14-15F;
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FIGS. 19A-19C are a series of perspective, somewhat
schematic views of a rotor assembly in accordance with an
embodiment of the present invention in which relief areas
are formed in the sides of the sealing surfaces between
the upper and lower ends thereof so as to reduce wear and
provide enhanced characteristics for certain
applications; and
FIG. 20 is a chart demonstrating the relationship
between the relative sliding velocity of the sealing
surfaces of an engine in accordance with the present
invention, as a function of shaft velocity.
DETAILED DESCRIPTION
a. Overview
In discussing the rotors used in the engines
described herein, reference will be made to "top" and
"bottom". Points on a line bisecting the larger angle
formed between offset intersecting axes A and B in the
plane defined by axes A and B will be referred to as
being at the "top", while points on the extension of that
line bisecting the acute angle between axes A and B will
be referred to as being at the "bottom".
In FIG. 1 there is shown an engine 10 in accordance
with one embodiment of the invention, formed by a
housing 12 having an interior surface 14 defining at
least a partially spherical cavity, with a central point
at the center of bearing 16. A master rotor 20 is
mounted for rotation on and within the housing 12 about a
first axis A. The master rotor 20 includes a shaft 22
extending along the axis A and has contoured faces 24, 26
forming plural vanes 25 on the other side of the master
rotor 20 from the shaft 22. A slave rotor 30 is mounted
for rotation on and within the housing 12 about a second
axis B. The slave rotor 30 includes a shaft 32 and has
contoured faces 34, and a side face 36 forming a
plurality of vanes 35a on the other side of the slave
rotor 30 from the shaft 32. Each of the rotors 20, 30
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defines at least part of a sphere, and share a common
center coinciding with the center of the cavity. The
vanes 25, 35 of~the opposed faces of the rotors 20, 30
interlock with each other to define chambers. Axis A and
axis B are non-collinear, being at an angle to each
other, and intersect at the center of the cavity defined
by the housing. The shaft 32 is journalled on an axle 33
(FIG. 9) in this example (configuration as a pump,
turbine or hydraulic engine) since the slave rotor 30
l0 need not be driven. The shaft 32 may also be
cantilevered in the same manner as the shaft 22. The
master rotor 20 and slave rotor 30 face each other within
the housing in an axial direction, each being
predominantly on one side of the common center of the
rotors .
The portion of the interior surface 14 that is
spherical is the portion in which both the vanes of the
master rotor 20 and slave rotor 30 rotate. In an extreme
position, where the vanes of one rotor extend into the
shaft of the other rotor the vanes of both rotors extend
into the shafts 22, 32. The shafts 22, 32 are not
spherical, but rotationally symmetric. In addition, the
master rotor 20 and slave rotor 30 should be generally
spherical in the portions in which they overlap during
operation. The remainder of the rotors 20, 30 and the
interior surface 14 need only have rotational symmetry to
the extent required to have the rotors 20, 30 rotate in
the housing 12.
As will be seen, the contoured faces 24 and 34, and
the side faces 26 and 36 of the master rotor 20 and slave
rotor 30 cooperate with each other and the interior
surface 14 of the housing 12 to form chambers 40 (the
space between the faces of the rotors) that change volume
with rotation of the rotors 20, 30 about the axes A and B
respectively. Ports 42 are provided in the housing 12 to
allow fluid flow in and out of the chambers.
Each contoured face is formed from a contoured face
24, 34 and side faces 26 and 36 defining vanes (blades)
25, 35 between them. The contoured faces 24 and 34 form
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areas of contact between the two rotors 20 and 30.
Sealing of the chambers 40 is accomplished by close
tolerance fit of the rotors 20 and 30 against the housing
12 and bearing 16, as well as the relationship of the
S vanes 25, 35 with respective contoured faces 24 and 34.
As is described in the above-referenced US 5,755,196, the
contours of the surfaces in a CvRTM engine of this type
can be determined by defining the contact faces of the
rotors by a locus which is formed as the rotors rotate
about their respective axes by points on the other rotor,
the points of each rotor that define the locus lying
along an outer edge of a cone whose central axis is
essentially a radius extending outward from the common
centers of the rotors at an angle a/2 from a normal to
the axis of the other rotor. For purposes of the
advantages of the present invention, however, the
contours of the contact surfaces are preferably
determined using the methods which are described below.
Side faces 26 connect inner ends 27 of one contoured
face 24 with the outer ends 29 of adjacent contact faces.
The side faces 26, unlike the contoured faces 24, have a
somewhat arbitrary shape. Clearly, they should not stick
out beyond the tips 28 of the vanes 25, else they will
crash into the side faces 36 of the slave rotor 30. The
shape of the side faces 26 can be adjusted for different
volumetric ratio changes of the chamber 40 defined
between the rotors 20, 30. The chambers 40 may compress
to one seventh their maximum size (compression ratio 7:1)
in a three vane case. For the embodiment shown by the
dashed line in FIG. 1 the ratio will be less. For any
single chamber, the point of maximum compression occurs
when the vanes 25a, 35a are equidistant from the bottom
of their rotation, that is from the line bisecting the
acute angle between axes A and B. Enlargement of the
chambers 40 may be accomplished by removing material from
the side faces 26 and 36 to render them concave. Dotted
lines F in FIG. 1 show preferred cutting lines. The
resulting chambers have considerable volume for the
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efficient pumping of fluid due to reduction in fluid
velocity at the intake and exhaust chambers.
The master rotor 20 and slave rotor 30 could
conceivably rotate cantilevered on their shafts 22, 32
respectively without additional bearings. However,
contact problems and fluid loss at the center of the
cavity poses considerable difficulties. It is preferred
that a spherical bearing housing be formed by removal of
a partial sphere of material from the center of each of
l0 the master rotor 20 and slave rotor; the spherical
bearing housing houses bearing 16.
The laterally extending material of the rotors
housing the bearing 16 is concave over greater than 180°,
creating difficulties in construction. The bearing may
be made integral with or otherwise fixed to either rotor,
preferably the master rotor 20. For the other rotor, the
bearing 16 can be loosely fitted in a less than 180°
bearing housing, resulting in a greater leakage path, or
the bearing may be press fitted into the housing,
thermally contracted and inserted into the bearing
housing, or slotted for insertion and rotated once inside
the bearing housing to present a round bearing surface to
the slave rotor.
As is shown in FIG. l, the master rotor 20 is driven
by a power source (not shown) through shaft 22. Vanes 25
of rotor 20 push on contoured faces 34 of rotor 30 on the
side shown on the other side (not shown) contoured face
24 of rotor 20 push on vanes 35 of rotor 30.
The internal and external configuration of the
housing is shown in FIGS. 2, 3 and 4. In particular, the
location of the ports 42 can be clearly seen, along with
flanges 50 and 50a for connection of the housing 12 to
input and output pipes (not shown). An alternative
threaded coupling 51 is also shown in FIG. 1. The
housing 12 is preferably formed of two halves 12a and 12b
bolted together with bolts 54. The ports 42 and 42a are
located at opposite sides of the housing, with an intake
port 42a and outlet port 42b. As seen in Fig. 4a, areas
55 show contact areas of a vane on contoured faces
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between the master and slave rotors 20, 30. Referring to
Fig. 4c, fluid enters the intake port 42a in expanding
chamber 40a. Chamber 40c is at maximum expansion in this
rotational position. Chamber 40b is contracting and
therefore forces fluid out of port 42b. Chamber 40d is
at maximum compression in this rotational position.
Preferably, the ports 42 have peripheries that match the
chamber configurations at the point the chambers cross
the boundaries of the ports so that as many points as
possible of the chamber edge, defined by a pair of vanes
25 cross the port edges at the same time. The trailing
edge of the set of vanes beginning to cross the exhaust
port or intake port defines the preferred shape of the
port at that position. The leading edge of the vanes
exiting the intake port or exhaust port defines the
preferred shape of the port at that position.
For operation as a pump, the master rotor is driven
by a power source. Rotation of the master and slave
rotors with each other causes the chambers 40 to contract
while moving from the point of maximum separation of the
rotors at the top to the point of minimum separation of
the rotors at the bottom. On the other side, the
chambers expand. While expanding, the chambers intake
fluid, and while contracting the chambers expel fluid,
increasing the velocity and/or pressure of the fluid, and
energy of the fluid. Thus, energy of the motor driving
the pump is converted to energy imparted to the fluid.
The parts described here may be made of any suitable
materials including plastics and metal, depending on the
intended use. Steel may be used for the master rotor 20,
while brass may be used for the slave rotor 30. At
10,000 rpm., a steel and bronze pump is believed to be
able to produce 10 hp per lb. weight of pump, and 20 hp
per lb. weight of pump for titanium rotors. As will be
described below, care must be taken to provide close
tolerance fits of the vanes so that little fluid can
escape past the vane contacts and between the rotor and
the casing. Material may also be added to the vanes to
allow wear.
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This invention provides a positive displacement
rotary pump with high efficiency, believed to be over 900
overall efficiency, and for a pump with eight inches
outside diameter, with seven inch diameter rotors, is
believed to be able to pump one liter per revolution.
100% rotary motion provides low stress on parts and low
vibration. Applications include irrigation, fire
fighting, down-hole water and oil pumping, hydraulics,
product transfer pumps and high rise building water
pump s .
b. Mirror Image Contoured Surfaces.
A preferred embodiment of the invention is shown in
FIGS. 5-7C, the engine in this exemplary embodiment being
configured for use as a pump, although again it will be
understood that the engine can be configured as an
external combustion engine or other power source. As
will be described in greater detail below, a particular
enhancement featured in the embodiment which is shown in
FIGS. 5-7c lies in the mirror image contoured surfaces
which are provided on the leading and trailing sides of
the "lobes".
As with the embodiment of FIG. 1, the engine 100 as
shown in FIG. 5 includes a master or power rotor 112
which rotates about a first axis A and slave or passive
rotor 114 which rotates about a second axis B which is
offset from the first axis A by an angle 8 (see FIG. 7A).
The rotors are housed between the two halves 116a, 116b
of an external casing which seals and supports the
assembly and also has inlet and outlet ports for the flow
of fluid through the engine.
Each rotor 112, 114 is partially spherical with a
common center, and the casing includes a corresponding
spherical cavity 118 which receives and holds the rotors
in engagement. The end shafts 120, 122 of the master and
slave rotors are supported by the casing. The terminal
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end 124 of the slave rotor 114 terminates and is fully
enclosed within the casing 116, which provides the
advantages of simplified sealing and reduced cost of
manufacture, although it will be understood that in some
embodiments the slave rotor shaft may extend through the
exterior of the casing. The master rotor end shaft 120,
in turn, extends outwardly from the casing and is
connected to a suitable external power source (not
shown), such as an electric, hydraulic or other motor.
Each end shaft is supported in a pair of
bearings 126 and 128 to maintain shaft stability and
eliminate end play. The inner bearings 126 include
conical bearing faces (not shown) which engage
corresponding conical tapers 129a, 129b on the backs of
the rotors, so as to react against thrust loads and
maintain the rotors in proper engagement. The bearings
are received in corresponding cavities 130, 132, with
lubricant being supplied to the cavities through a series
of ports 134. The bearings are preferably high speed
fluid film bushings, i.e., bushings which run on a thin
film of air, oil, water, etc., although it will be
understood that other forms of high speed bearings may be
employed in some embodiments. A continuous elastomeric
seal 136 is retained in a channel 138 which extends
completely around the rotor chamber and shafts, and
includes a ring seal 140 which surrounds the master rotor
and shaft where this exits the housing; the seal 136 may
suitably be formed of a moldable polyurethane material.
The clamping force of the two casing halves against the
elastomeric member provides the low pressure seal for the
assembly, while the fluid pressure acting outwardly
against the elastomeric material creates the high
pressure seal.
As was noted above, the casing also includes an
inlet port 142 and an outlet port 144, which communicate
with the rotor chamber 118 and via which the fluid enters
and leaves the engine; the inner edges 146, 148 of the
ports, where these meet the spherical rotor chamber, have
a shape which matches the corresponding edges of the
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contoured surfaces which define the sealed chamber
between the rotors (which shape will be described in
greater detail below), while the outer edges 150, 152 of
the ports are round for connection to conventional
circular cross-section tubing or other conduits.
FIG. 7A shows the engagement of the first and second
mirror image contoured surfaces 160, 162 on each vane
164, and the contoured surfaces 166, 168 on the
corresponding cavity 170. This engagement forms a
l0 substantially sealed chamber which changes in volume with
rotation of the rotors. In contrast to the embodiment
described above with reference to FIGS. 1-4, however,
each vane or lobe is provided with two mirror image
contoured surfaces, i.e., a leading contoured surface and
a mirror image trailing contoured surface.
The relationship between the leading and trailing
contoured surfaces is perhaps best seen in FIG. 7B, which
is the top or "overhead" view of the master and slave
rotors 112, 114. As can be seen, the lobes 164a, 164b,
etc. of the master rotor 112 are angularly spaced so as
to define a plurality of angularly spaced cavities 172,
and the lobes 174 on the slave rotor define corresponding
cavities 176. As can also be seen, each lobe is received
in the corresponding cavity in the opposite rotor i.e.,
the master rotor lobes 164 are received in the cavities
176 in the slave rotor, and the slave rotor lobes 174 are
received in the cavities 172 in the master rotor. The
area in the center of the rotors, between the lobes on
either side, is sealed by a ball 175 or other generally
spherical body.
As can also be seen in FIG. 7B, the leading and
trailing contoured surfaces on each lobe engage the
corresponding contoured surfaces on each socket (these
being the contoured surfaces of the lobes on either side
of the socket), as indicated at the areas 178.
Consequently, a series of sealed chambers 180a, 180b,
180c are formed about the end of the master rotor,
between the ends or "heads" of the lobes in the bottom of
the cavities, and a corresponding series of sealed
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chambers 182a, 182b, 182c are formed around the end of
the slave rotor.
The chambers change in volume with rotation of the
rotor assembly, in the direction indicated by arrow 184.
As can be seen by comparison of chambers 180a and 182a in
FIG. 7A, the volume of the chamber increases as these
rotate past the inlet port 142 (see FIG. 5), thus drawing
fluid into the pump. The ports are shaped so that each
chamber moves out of register with the inlet port just as
the chamber reaches its maximum volume (see chamber 180b
in FIB. 7B), and just before the chamber begins to rotate
into register with the outlet port 144. The chambers
then decrease in volume as they rotate past the outlet
port, forcing the fluid outwardly, and reach a minimum
volume at the bottom of the cycle (see chamber 182c in
FIG. 7C) just after rotating out of register with the
outlet and prior to opening into the inlet port . As a
result, the fluid enters the pump through the inlet port
at a first pressure indicated at P1 and is discharged
through the outlet port at a second, higher pressure
indicated at P2, as shown generally by arrows 186a, 186b
in FIG. 7B.
The embodiment having the lobed vane structure with
mirror image leading and mirror image contoured surfaces
has several advantages over the device which is shown in
FIGS. 1-4. Firstly, the use of mirror image contoured
surfaces enables the engine to run and develop pressure
in either direction of rotation. This is because the
mirror image contoured surface lobes do not require the
force of the faster (power) rotor vanes pushing against
the slave rotor vanes in order to maintain a contact
seal.
Moreover, the mirror image contoured surfaces on the
lobes enable an acceptable fluid film between the
surfaces at a wide range of operating speeds and fluid
viscosities. Maintaining a thin fluid film between the
contoured surfaces is advantageous for reducing wear and
friction. However, when operating at high speeds and low
back pressures the fluid tends to force the contoured
CA 02300420 2004-06-25
18_
surfaces apart, creating an excessively thick fluid film.
This results in a large amount of leakage, or back flow
and reduced operating efficiency. The mirror image
contoured surfaces control the amount of "backlash"
between the slave and power rotors, so that only the
predetermined amount of rotation is allowed between the
two, which in turn defines the maximum clearance/fluid
thickness there can be between the leading and trailing
contoured surfaces of the lobes. Due to the force of the
power rotor, the fluid film at the leading contoured
surface of each lobe will tend to be slightly less than
that of the trailing contoured surface; however,
depending on operating speed, back pressure, fluid
viscosity and other factors, an equilibrium level is
achieved in which a fluid film exists between both
leading and trailing surfaces.
Additional advantages include increased strength of
the rotor lobes, since the area between the mirror image
contoured surface (i.e., the backs of the contoured
surfaces) can be filled in, so that the back side of each
of the faces is reinforcing the other, giving the lobes
strength comparable to that of a gear tooth. Also,
because of the higher strength, it is possible to operate
the pump at higher pressures, which is advantageous in
increasing the power ratio, or power density, of the
pump.
c. Mathematical Calculation of Contoured Surface
Contours
The manner in which the contours of the contoured
surfaces are determined mathematically will now be
described with reference to FIGS 8A-lOD.
FIGS. 8A-8D provide a series of graphical
representations of axes, vectors, angles, and other
values associated with the mathematical computation of
the contoured surfaces of the vanes/lobes, as follows:
CA 02300420 2004-06-25
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FIG. 8A shows the orientation of the two rotor axes,
Axis 1 and Axis 2, intersecting at O and placed at an
angle A° apart. The line 0-0 is initially in the plane
of the two axes and bisects the direction of each, so
that it makes an angle of (90+A/2)° with each axis
direction. Point Q is a radial line on the surface of a
sphere of radius R, which is a point locating the working
surface of the rotor attached to the shaft having Axis 2.
The plane P formed by the line O-O and OQ will be a plane
that changes orientation in space.
When the pump turns about each axis by the same
angle, A - 81 - A2 as shown. To construct the rotor
surface on Axis 2, we need to consider the relative
motion of Axis 1 with respect to Axis 2. Taking a vector
difference of rotations as shown in FIG. 8B, makes the
axis of rotation lie along the direction of vector 81-B2,
which is the direction of the line O-O in the starting
position shown when B=0. A superposition of two
rotations can be used to get the new direction of O-O for
other angles. First, the line 0-O is rotated about
Axis 1 an angle of 8. Thus, with Axis 1 fixed, O-O is
rotated 8 in the opposite direction, or the -82
direction. By analyzing the displacement components
involved in this operation, the xyz coordinates of the
point Q' can be determined as
x = - sinz A cos A cos (2 cosA - cos2A) - cos' A
2 2 2
~_ - sine A cos "- (2 sin8 - sin29)
Rp 2 2
_z = - 2 sin A cost "- (2 cosh - cos29) - cos' ~
2 2 2
where Ro -Rcosa
As A increases, point Q describes a curved path in
space and plane P must remain perpendicular to the
tangent to that curve. This can be insured by making the
plane be perpendicular to a tangent vector, which can be
found from the direction of the velocity of either point
CA 02300420 2004-06-25
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Q or point Q', which remain a fixed distance apart,
00' - Rsino.
A set of unit vectors can be used to describe the
orientation of plane P in space. As shown in FIG. SC,
let ul be the first vector, directed along O-O, and is
defined in terms of the vector Ro:
Since Ro -Roul
1 o a 1 = _Ro = _1 [x, y, z]
Ro I Ro
where Ro - (x2 + y2 + ZZ ~o.s
Vector u2 is tangent to the path of Q' and is obtained
from the vector cross product to get the velocity of Q.
VQ=uOxRou1
VQ = [cos "- cos6, cos "- sin6, -sin "- ] x [x, y, z]
2 2 2
where the component of the two vectors are given. The
vector u0 is a unit vector with direction along O-O,
which changes with rotation, as does ul.
The unit vector along the tangent of the path of Q'
or Q will be
2s VQ
u2 =
VQ
and will be a function of both the shaft angle A and the
rotation 8.
A third unit vector, perpendicular to both ul and u2
will be
u3 - ul x u2
and will be in a transverse direction, along QQ'.
The coordinates of the point Q can now be determined
from the vector equation,
OQ = R = RcosB ul + RsinB u3
in which
OQ' - Ro - RcosB ul
The outer edge of the surface determined by Q is
shown in FIG. 8D. Also shown is the rotation of plane P
for different rotations of the shafts.
CA 02300420 2004-06-25
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The total angular twist S, along the axis O-0 in any
general position can be most easily obtained by
determining the angular change in the normal to the plane
P, which is the unit vector u2. This vector is always
directed along the tangent to the path of Q or Q', and
has already been defined.
As is shown in FIG. 8E, the untwisted position of
the plane P can be obtained by rotating the plane and its
initial normal direction vector u2 about the z axis in
the xy plane through the angle 0, to a new position u2'o
and again about the plane OQ' Q" through the angle yr with
the Z axis in the x-y plane through the angle 0, to a new
position u2'o and in the plane OQ'Q" until it makes an
angle yl with the Z axis to a final position u2"o. These
angles are related to the xyz coordinates by
tan6 = ~ = 2 sin0 - sin20
x cot _A + 2 cos0 - cos E3
2
cosyr = -_Z = sin _A ( 2 cos z _A (1 - cos9) -1)
Ro 2 2
This rotation of u2o to u2o" is done best by defining
another unit vector perpendicular to the plane OQ'Q",
which is:
u4 = [sin0, - cos0, 0]
and then generating yet another unit vector from
u5 = u4 x ul
This vector u5=u2o" represents the untwisted
position of plane P, and u2 represents the twisted
position. The angle between them in space will be the
total rotation about the axis O-O. This is found from
the dot or cross product of the two vectors. Using the
dot product,
cos S = u2 . u5
cos S = u2x u5x + u2y u5y + u2z u5z
from which the angle S is determined from the xyz
components of the two vectors.
CA 02300420 2004-06-25
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FIGS. 9A-9D are a series of views of a model which
provides a visual representation of the relationships
between axes and points in the system described by the
mathematical process above.
FIG. 9A shows the "start" position, in which the
axes 1 and 2 correspond in angular relationship to the
axes of the master and slave rotors, the length O-Q
represents the radius of the rotor, and the point Q, on a
line normal to axis R, represents one point along the
contoured surface of the lobe. The offset between O-O
and Q, in turn, represents the surface depth of the lobe.
By conceptually rotating the Axis 1 through about 90-180°
and following the mathematical process set forth above,
the point Q sequentially plots out a line having a
contour of a line on the contoured surface of the lobe.
For purposes of illustration, FIGS. 9A-9D show
rotation of Axis 1 90° from the start position to the
final position; it will be understood, however, that
determination of the line is ordinarily carried out in
small degree increments, so as to define a smooth,
continuous contour.
Accordingly, FIG. 9B shows the model 190 with the
Axes 1 and 2 having been rotated together by an angle A
of 90°, so that axis R swings from the vertical alignment
(for purposes of illustration) is shown in FIG. 9A to the
horizontal alignment in FIG. 9B. Then, with Axis 1 held
stationary, Axis 2 is rotated back by an angle -8, which
is equal to A but in the reverse direction, rotating
axis R to the position which is shown in FIG. 9C.
Finally, the axis R is rotated by the amount 6S which is
calculated in accordance with the mathematical system
described above, bringing point Q to its final position
Q", as shown in FIG. 9D. For purposes of illustration,
FIG. 9D also includes a broken-line image 192 which shows
the original position of point Q at the start point shown
in FIG. 9A.
CA 02300420 2004-06-25
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FIGS. l0A-lOD are a series of views similar
conceptually to FIGS. 9A-9D, but showing the manner in
which the above process is used to generate or determine
a contoured line 194 in a computer plotting program. As
can be seen, by following the process described above,
the point Q is moved sequentially from position to
position line 194, with each rotation of the Axes 1
and 2. By in essence "connecting the dots", i.e., the
position of point Q at each position of Axis 1, a
l0 continuous contour line is created which corresponds to
the contour line along one of the contoured surfaces,
such as the contoured surface 160 on lobe 164, as shown
in FIG. 10E.
The offset establishes sufficient clearance between
the contoured surfaces to establish the fluid film and
avoid the parts rubbing directly on one another. The
amount of the offset is determined on a basis of fluid
type and viscoscity, operating speeds and pressures, and
materials characteristics, along with other factors.
Also, in some embodiments where the rotors are formed of
resilient material such as urethane, a "negative" offset
may be used, so as to cause some interference between the
contoured surfaces which forms an enhanced seal; this may
be particularly desirable for high-pressure, low-speed
applications.
Having established the contour line at the outer
edge of the lobe (i.e., at the full radius of the rotor),
the three-dimensional surface is generated by one of two
methods. Firstly, the contour line can simply be scaled
down towards the center of the rotor, in which case the
clearances and thickness of the fluid film will also
decrease towards the center accordingly. In other
embodiments, the contour line can be recalculated at the
smallest radius at the lobes/vanes, with the intermediate
contour lines defined accordingly, so as to give a
constant gap/fluid film thickness across the entire
contoured surface; this approach may be particularly
advantageous where the fluid contains particulates of a
CA 02300420 2004-06-25
-24-
known size, and it is therefore important to maintain a
fluid film which is thick enough to hold the particulates
without these being forced into the contoured surfaces.
Whichever approach is used, one contour can be calculated
for the leaving contoured surface of the lobe and then
reversed for the mirror image trailing contoured surface,
or vice versa.
FIGS. l0A-lOD, in turn, illustrate the manner in
which these calculations are employed to produce a
computer generated plot of the contour lines, in FIG. l0E
is a partial perspective view of one of the rotors,
showing the position of the contour line which has been
produced in FIGS. l0A-lOD.
The offset distance from the axis O-O out to the
working surface is (4 sin8 + t), where s = R8. If the tip
were to be reshaped to provide a larger radius of
curvature at the beginning of contact (for the purpose of
reducing wear), the profile of the working surface can
still be calculated readily from the existing computer
program.
An approach is as follows, as shown in FIGS. 11A and
11B. Modify the tip radius to make a slightly flattened
shape, in the vicinity of where first contact occurs.
This shape can be identified as s=s(8), which means the
radius is a function of (depends on) the shaft
rotation 8. Once this is selected, the radius can be
input as a function of small angular increments, and the
profile of the mating working surface calculated for the
same fluid thickness t. Actually, fluid thickness may
not be constant everywhere. It will probably depend on
the relative sliding velocity of the vanes, which
increases from zero at the point of contact and increases
to a maximum near 90° rotation. The initial flattening
of the tip may affect this also.
The working surface would normally follow a radial
line towards the center O, resulting in a film thickness
that tapers towards the center. The relative sliding
velocity between adjacent lobes will be highest at the
CA 02300420 2004-06-25
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outside, so a larger thickness of film there seems
reasonable. However, for applications where small
particulates are contained in the fluid, it may be better
to machine the rotor so that a parallel gap is produced.
This may prevent material from sticking in the small end
of the tapered gap, even though it would tend to be
flushed away during the next rotation.
In FIGS. 12, 13A, and 13B, c is the distance between
centers of adjacent tips (measured along the arc of the
surface of a sphere of radius R). The arc length C is
the distance between like lobe shapes (circular pitch
length) .
2~ R
C=2(c=2s+t) _ -
n
where s is the arc length taken up by the tip, t is the
film thickness (or net interference) and n is an integer
number of pitch lengths to make up a full circle.
If s and t are chosen,
~R-2s-t
C=
n
Therefore, the spacing of both c and C are known in
terms of the film thickness (negative for interference)
and the arc length s. Note that since d=C -c, the center
points of the tips of all lobes are not equally spaced
and
~R
d=+2s+t
n
For input to the computer program, the angle 8=s/R,
and the "offset" distance is
yrQQ = R sing
Consideration should be given to providing variable
spacing, as this would help to alleviate the production
of a pure tone noise (having a single frequency
component) emanating from the running pump. Variable
spacing would produce other frequency components, grouped
around the running speed frequency and its harmonics
CA 02300420 2004-06-25
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(sidebands) . The effect should reduce the overall noise
level slightly, but more importantly, be less annoying
for personnel in the vicinity.
However, rotor unbalance could be produced for
random spacing. If the spacing were arranged
symmetrically in pairs, unbalance can be prevented, but
the beneficial effect of staggered spacing would be
reduced. If the unbalance were the result of a
particular arrangement, each rotor could be balanced
individually before final assembly. For uniform spacing,
whether the number of rotors n is an even or an odd
number, balance would be maintained.
d. Geometric Determination of Surface Contours
FIGS. 13-19 illustrate a method for geometric
determination of the contoured surface contours
consistent with the mathematical calculations described
above, but which corresponds more directly to an actual
manufacturing process for forming the surfaces, as by
bobbing material from a blank so as to form the lobes and
surfaces.
Two of the main considerations when determining the
correct sealing surface gap (SSG) are the "lift off
clearance" and the contact characteristic. The "lift off
clearance" is the thickness of the fluid film between the
sealing surfaces of the two rotors when the engine is
operating in its intended mode. "Lift off clearance" is
affected by the speed of the engine, the viscosity of the
fluid medium, and the differential pressure between the
inlet port and the discharge port.
Contact happens when the one or two or all of these
factors is insufficient to maintain a fluid film between
the mating surfaces. The contact characteristic describes
how the sealing surfaces mate when the fluid film is not
sufficient to achieve "lift off". The three basic types
of contact are (1) Full radial contact, (2) Inner radial
contact, and (3) Outer radial contact. These
CA 02300420 2004-06-25
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characteristics can be different at different angles of
rotor rotation.
Maintaining a fluid film is desirable to reduce
wear, as well as to allow entrained particles to pass
between the sealing surfaces without damaging the
particles or causing excessive abrasion to the sealing
surface .
U.S. 5,755,196 describes a CvRT'" engine configuration
with a "contact" or "close tolerance" seal design which
t0 does not optimize or account for the "lift off
situation". This type of surface geometry relies on a
line to line seal between the rotors and is intended to
operate with each rotor sliding on the other rotor
without consideration of the fluid film between the
rotors. An engine which is designed with this "zero lift
off" seal surface will not achieve a consistent fluid
film thickness during "lift off" because "lift off" of
any type of seal surface in a CvRT'" engine does not occur
"normal" to the contact surface. "Lift off" happens as
the two mating rotors rotate relative to each other
around each of their respective axes. As this happens,
the gap between the rotors increases more at points which
are further from the axis than it does at points which
are closer to the center of the rotors.
The radial difference of the surface speed in this
contact zone may make up for the variation in gap
thickness when the engine is operating at very low
pressures. (relative surface speed is greater at points
further from the rotational center) But the fluid film
"rigidity" is not linear with the thickness of the film
which the surface speed is a linear relationship with the
distance from center. Ideally then, if surface speed was
the only consideration, then the SSG should increase at
points further from center, but only enough to establish
a consistent fluid film pressure.
As the pressure increases, however, the fluid film
is influenced increasingly by the pressurized fluid which
is moving past this area. The fluid film resulting from
the surface speed is affected greatly by the distance
CA 02300420 2004-06-25
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from center and requires an increasing surface gap
towards the outside of the engine. The fluid film
resulting from the differential pressure between the
output port and the input port is independent of the
distance from center and requires a more consistent gap
clearance. The more the fluid film is affected by the
pressure differential of the fluid, the more consistent
the radial gap clearance must be to achieve maximum
efficiency and wear characteristics.
l0 Consequently the present invention provides methods
for determining, defining, and/or constructing this more
consistent gap clearance, as well as a method for
determining, defining, and/or constructing an engine with
a gap clearance that also takes into account the surface
speed of each rotor on the other to maximize the "liftoff
effect" of the fluid film between the rotors. The
methods can also be combined to account for other
variables including the change in relative surface speed
which occurs at different angular rotor positions.
Optimally, in a CvRTM engine, contact between the
sealing surfaces may occur during start-up under high
pressure, but should not continue when the engine is
operating in its intended mode. In order to achieve
"lift off" as soon as possible after start up, and under
as high a pressure, and as low a viscosity, and as low a
speed as possible, it is desirable to determine and
construct a sealing surface with seal surfaces which are
more parallel rather than angular interface surfaces
which radiate from the spherical center of the pump.
U.S. 5,755,196 describes a surface which is defined
by the movement of a cone being rotated around the
opposite rotor axis, in which the apex of the cone should
be as close to the center of the spherical center of the
rotors as possible. The sealing surface of the present
invention can also be described with the movement of a
cone around the opposite rotor axis, but the cone of this
present invention is positioned intentionally above or
below the spherical center of the rotors.
CA 02300420 2004-06-25
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By using an off-center cone position, a more
parallel seal surface interface can be achieved. This
more parallel surface shape will provide a more stable
and consistent fluid film between the rotors for reduced
wear, and more efficient sealing. In applications where
interfering rotor seal interface is desirable, it may
even be desirable to produce rotor seal surfaces which
interfere more towards the center of the rotors than they
do toward the outside of the rotors. The advantage of
l0 this design would include better sealing near the center
of the pump, and lower friction and less resistance
further from center where any resistance will have a
greater effect on the operating efficiency of the engine.
To achieve this "angular interface" effect, as well
as the "parallel interface" effect, it may also be
necessary to introduce a second surface shape variable,
which is the angular position of each contoured face
about the center axis of the pump rotor. By rotating
each seal surface relative to the rest of the pump, a
predetermined surface interface with specific
characteristics can be achieved.
The angular position effect and the off-center cone
apex effect will be covered in the following description
of how to achieve the desired sealing surface geometry:
Referring now to FIG. 14, a spherical rotor RA is
positioned for rotation about its center axis AA. A
second axis AB is positioned at an angle X to axis AA. A
cone C is positioned with its center axis collinear with
a line Y that bisects the obtuse angle between axis AA
and axis AB.
If a positive parallel SSG is desired, the cone C is
positioned on line Y with its apex X below the point P
where the two rotor axes intersect.
If a negative parallel SSG is desired, the apex of
the cone must be positioned above the point P. (The
smaller the angle of the cone, the more its apex must be
positioned off center to achieve a given gap clearance or
interference.)
CA 02300420 2004-06-25
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As is shown in FIGS . 15A-E, the spherical rotor and
the cone are then rotated around their respective axes
(i.e., cone C rotates on axis AB at a fixed angle
thereto) and the path of the cone is removed from the
spherical rotor. This will define the "seal surface" S
of one side of one vane Vlon the rotor RA.
The rotor is then rotated toward the f first cone and
another cone shape C is positioned with its axis
collinear with the line Y. This cone has the same angle
as the first cone and it is positioned with its apex the
same distance from center but on the opposite side of
point P (see FIG. 15) . This cone is added to the rotor
RA and becomes the "seal tip" T of this seal face, as is
shown in FIG. 15E. The sequence is then repeated for the
second rotor RB (See FIGS.16A-16B) with a cone which is
positioned along the center axis of the adjacent "seal
tip" T cone of the rotor RA.
Once this sequence is repeated for each side of each
vane, the engine will have a predetermined parallel
interface gap IG between mating surfaces as is shown most
clearly in FIG. 16B.
Another gap configuration which can be used on its
own or in combination with the "offset cone" gap
configuration, is the "angular interfacial gap".
This type of gap (or interference) is achieved by
rotating each seal surface around the center of its
rotor's axis relative to the seal surface on the opposite
side of the vane it is on as is shown in FIG. 17.
Comparative examples of positive and negative "angular "
and "parallel" interfacial gaps are shown in FIG. 18.
An angular interfacial gap may offer performance
benefits for certain applications. For example, the
centrifugal force of the rotation of the engine could be
used to force particulate matter entrained in the fluid
to the periphery of the engine chamber. In this case an
angular interfacial gap with a larger gap at the
periphery of the rotors would allow the particles to pass
through the thicker fluid film, while a more efficient
CA 02300420 2004-06-25
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seal could be maintained closer to the center of the
rotors where the fluid film is thinner.
A characteristic of the "parallel interfacial gap"
compared to the "angular interfacial gap" is that the
"parallel interfacial gap" method creates a consistent
SSG for the entire seal surface. The "angular IG" method
(of rotating the seal surface relative to the rest of its
rotor), only changes the gap clearance in a plane that is
perpendicular to the rotational axis of the rotor.
This is desirable for applications where a reduced
gap clearance is beneficial during specific areas of the
seal surface interface. Shear sensitive or highly
viscous fluids, for example, might be damaged or cause
excessive friction if a minimal gap were maintained for
the entire rotation of the rotors. In this case, a
smaller gap can be achieved during the sealed portion of
the rotation at the bottom of the rotation while a larger
gap will be more desirable during the unsealed portion of
the rotation.
Further benefit can be realized in this regard if
the relative speed of the rotors is taken into account
(see FIG. 20). The sealed part of the rotation at the
top and bottom of the casing corresponds with the lowest
relative speed of the interjacent rotation of the rotors.
As the seal tip of each vane nears BDC the surface
speed reduces. A reduced gap clearance can be achieved
in this area using the Angular IG method or a combination
of the Angular IG method and the Parallel IG method of
changing the gap at the higher relative speed areas of
the seal surface.
At TDC the surface speed also reduces, but the
angular IG method will increase this gap. To increase
the gap clearance at some places but not at TDC, it is
necessary to use the Parallel IG method of achieving the
desired gap, but the cone must be moved dynamically along
its axis as the rotors are rotated during the shaping
process.
CA 02300420 2004-06-25
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In essence, as one rotor rotates from each contact
extreme, relative to the other rotor, the transitional
gap between the rotors changes from an angular
interfacial gap to a parallel interfacial gap and on to
an angular interfacial gap at an angle in opposition to
the initial angular interfacial gap. Some transitional
gaps will be a variation of the above description in that
they will incorporate only one or two of these
descriptions.
to Although the cone shape described above is the ideal
shape, and the simplest to calculate and design, it will
be understood that other similar shapes (such as a
portion of a much larger cone or simply a sharp edge)
could be used, however, as the mating surface is designed
to maintain the desired SSG as both rotors spin at the
same speed.
Furthermore, it will be understood that, while the
description of the method of the present invention has
been described herein with regard to externally contoured
vanes/lobes, the method is equally applicable to CvR
engines having pistons or corresponding structures which
are house or retained within the lobes, such as the
piston-engine structure which is shown in FIG. 16 of the
above-referenced U.S. patent.
e. Verification of Contours
Many methods for verifying the surface shape are
available. A contact CMM machine, for example, could be
3o used to determine a number of points on the surface of a
completed rotor, and establish what the seal surface
characteristic is. The most basic way of determining if
a rotor design has been manufactured according to the
present invention is to create a plane which is
perpendicular to a point on the seal face (or seal tip)
which passes through the spherical center of the sphere.
Two points on the seal face or seal tip surface which are
CA 02300420 2004-06-25
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also on this plane will be connected and extended toward
the spherical center of the engine.
A rotor face with a parallel interfacial gap will
result in the extended line passing consistently to the
contact surface lobe side of the spherical center.
A rotor face with an angular interfacial gap may
result in the extended line passing through the spherical
center of the rotors or on either side, depending on the
angle, and on the magnitude of the gap. For most
applications, however, the extended line of an angular
interfacial gap will pass through the spherical center or
to the side of the spherical center which is away from
the mass of the seal surface lobe.
A rotor face with a reverse angular interfacial gap
will result in the extended line passing consistently on
the side of the spherical center which is away from the
mass of the seal surface lobe.
A rotor face with an interfering parallel
interfacial gap will result in the extended line passing
consistently on the side of the spherical center which is
away from the mass of the seal surface lobe.
A rotor face with an interfering angular interfacial
gap will also result in the extended line passing
consistently on the side of the spherical center which is
away from the mass of the seal surface lobe.
A rotor face with an interfering reverse angular
interfacial gap will result in the extended line passing
consistently on the side of the spherical center which is
toward the mass of the seal surface lobe.
By checking the surfaces in this way, it is possible
to verify the sealing and fluid film characteristics of a
particular engine design.
f. Interrupted Seal
FIGS. 19A-19C illustrate in the embodiment of the
present invention in which the sealing surfaces are
shaped so as to provide actual fluid sealing during only
CA 02300420 2004-06-25
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selected portions of the rotation of the assembly, i.e.,
at those points during the rotation where the seal is
required in order to maintain efficiency. This
configuration is advantageous in a number of
applications, including for use with pumping sheer
sensitive or abrasive fluids, and for enhanced wear
characteristics.
Accordingly, as can be seen, in the embodiment which
is illustrated in FIGS. 19A-19C, the sealing surfaces 200
on the vanes 202 of the two rotors 204, 206 are each
formed with a recess or channel area 208 which extends
radially across the rotor base and separates the sealing
surface segments 210, 212 which lie proximate the tip and
at base portions of the contoured face.
The sealing surface segments 210, 212 are formed in
accordance with the methods described above, i.e., these
are configured to form the requisite seal with the
corresponding segments on the adjoining contoured face,
with a predetermined gap as desired. Since the sealing
segments are formed at the top and bottom of each
surface, the rotors form an effective seal only when the
chambers defined thereby are approximately at top and
bottom dead center, as is shown in FIGS. 19A and 19B.
At points in the cycle between top and bottom dead
center, however, the channels 208 eliminate direct
contact between the two sealing surfaces so as to form a
relief gap 220, as is shown in FIG. 19C. The relief gap
reduces sheer stresses on fluid in this area, and also
allows particulate or abrasive material to pass
therethrough without causing wear against the sealing
surfaces. Furthermore, the relief gap reduces wear by
eliminating a potential content between the sealing
surfaces during the intermediate phases of the engine
cycle, even in applications not being used with abrasive
fluids. Since sealing is only critical when the chambers
are at top and bottom dead center, these advantages are
achieved without significant cost to the overall
efficiency of the engine.
CA 02300420 2004-06-25
-35-
It is to be recognized that these and various other
alterations, modifications, and/or additions may be
introduced into the constructions and arrangements of
parts described above without departing from the spirit
or ambit of the present invention as defined by the
appended claims.
