Note: Descriptions are shown in the official language in which they were submitted.
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DOUBLE GAS SEAL WII'FI COPLANAR PAD FACES
Field of the Invention
The present invention relates to gas-lubricated non-contacting seals and,
more particularly, to a gas lubricated seal with radial coplanar pad faces.
The gas
seal of the present invention has various applications, and is particularly
well
suited for sealing between a rotating drive shaft sleeve and a pump housing,
so
that the double gas seal reliably seals process fluid within the pump housing.
Background of the Invention
Gas lubricated seals have been used for many years in compressors and,
in some applications, have largely replaced more conventional seals, including
liquid lubricated seals. Since the sealing faces of gas lubricated seals are
not in
dynamic contact, properly designed gas lubricated seals offer significant
benefits
of reduced frictional torque and reduced heat generation compared to
conventional seals. Moreover, since the high pressure gas supplied to a gas
lubricated seal may be selected for its inert qualities in view of the
application, and
since a properly designed gas lubricated seal offers a long life, these seals
are ideal
for applications requiring complete emission control and process purity. In
more
recent years, gas lubricated seals have been applied to pump technology to
seal
between the rotating shaft sleeve and the pump housing. Accordingly, pump
manufacturers have desired improved gas seals for various pump sealing
applications.
One type of gas seal uses circumferentially spaced grooves in one of the
sealing faces. The spiral grooves each extend radially inward from an outer
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periphery of the respective sealing face. Pressurized gas is supplied to these
grooves to block the escape of the fluid being sealed. One such gas lubricated
seal
which is embodied in a cartridge design is the Type 2800 seal manufactured by
John Crane, Inc. Other types of gas lubricated seals with spiraling grooves
are
disclosed in U. S. Patent Nos. 4,423, 879; 5,246,295; 5,385,409; S, 498,007
and
5,713,576. Other variations of gas lubricated seals are disclosed in an
article
entitled "Analysis of Spiral-Groove Face Seal for Liquid Oxygen" by Shapiro et
al., published in ASLE Transactions, Volume 27, 3, pp. 177-188. Another type
of non-contact gas seal marketed by A.W. Chesterton Co. as the 4400
TwinHybrid Gas Seal utilizes radially inward and outward sealing faces with
the
pressurized gas being supplied through the stationary ring and to
circumferentially
spaced elongate grooves spaced radially between the sealing faces.
While various types of double gas seals have been devised, the prior art
has failed to effectively benefit from double gas seal technology. Prior art
1 S coplanar double gas seals do not provide effective lift off of both the
radially outer
sealing face and the radially inner sealing face under various conditions.
Also,
much of the prior art relating to double gas seals provide seal designs which
are
too large for many applications since the seals have a long axial length or
require
a considerable diametral space.
The disadvantages of the prior art are overcome by the present invention.
An improved double gas seal is hereinafter disclosed which provides effective
lift
off of the radially spaced sealing faces and reliably seals pressurized fluid
while
minimizing seal wear.
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Summary of the Invention
A double gas pressure seal is provided for sealing between a stationary
housing and a rotary housing within the stationary housing. The gas pressure
seal
includes a stationary ring and a rotary ring which cooperate to seal fluid
within the
stationary housing. In an exemplary application of the invention, the
stationary
housing may be a pump housing and the rotary housing may be a sleeve which is
rotatable with the pump shaft. The double gas seal is supplied with
pressurized
gas from an external source and at a pressure higher than the pressure of the
process fluid in the pump housing to reliably seal the process fluid.
Each of the stationary ring and the rotary ring has an annular inner sealing
face and an annular outer sealing face, such that the inner and outer sealing
faces
are in sealing engagement. The phrase "sealing engagement" as used herein with
respect to the sealing faces does not mean that the pad sealing faces are
touching,
and instead the faces are separated by a stiff gas film, as discussed more
fully
below, to achieve long seal life. A spring or other biasing member axially
biases
one of the stationary ring and rotary ring toward the other ring, and in a
preferred
embodiment biases the stationary ring toward the rotary ring. An annular
groove
is provided in one of the stationary ring and the rotary ring, with the
annular
groove being radially spaced between the inner sealing faces in sealing
engagement
and the outer sealing faces in sealing engagement. A supply port in one of the
rings supplies the pressurized gas from an external source to the annular
groove.
A plurality of circumferentially spaced inner recesses and a plurality of
circumferentially spaced outer recesses are each provided in one of the
stationary
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ring and in the rotary ring, and preferably in the rotary ring. Each inner
recess is
spaced between the annular groove and a radially innermost portion of the
inner
sealing faces in sealing engagement, and each outer recess is spaced between
the
annular groove and a radially outermost portion of the outer sealing face is
in
sealing engagement. The plurality of inner recesses and the plurality of outer
recesses exert a gas lift-off force on the axial movable one of the stationary
ring
and the rotary ring while providing a stiff pressurized gas film between the
sealing
faces. Each of the plurality of inner recesses and each of the plurality of
outer
recesses has a rotary leading portion and a circumferentially spaced rotary
trailing
portion. A plurality of inner feed channels and a plurality of outer feed
channels
fluidly connect the annular groove with the rotary leading portion of each
inner
recess and outer recess, so that the rotary trailing portion of each of the
inner
recesses and outer recesses is circumferentially spaced from the respective
feed
channel. In a preferred embodiment, each of the sealing faces lies within a
single
plane which is perpendicular to an axis of the rotating shaft.
It is an object of the invention to provide an improved gas seal with
coplanar pad faces for sealing between a stationary housing and a rotary
housing.
A supply port in one of the stationary ring and rotary ring supplies
pressurized gas
to an annular groove. The pressurized gas then passes radially inward from the
annular groove through a plurality of inner feed channels to a plurality of
inner
recesses, and similarly passes radially outward from the annular groove
through
a plurality of outer feed channels to a plurality of outer recesses. The
circumferentially spaced inner and outer recesses provide the desired
pressurized
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gas lift-off force which separates the stationary ring from the rotary ring
while
allowing a relatively small quantity of pressurized gas to leak from the gas
pressure seal.
It is a feature of the invention that the double gas pressure seal may occupy
a small axial and radial space and thereby be used in a large number of
applications. It is a related feature of this invention that the gas pressure
seal may
be used over a wide range of external gas pressures and a wide range of fluid
pressures within the stationary housing.
A significant advantage of the present invention is that the double gas
pressure seal is relatively inexpensive and has a long life. The double gas
pressure
seal may be reliably used at elevated temperatures, and may be used to seal
various
types of fluids in the stationary housing, including abrasive, sticky, and
corrosive
fluids. The gas pressure seal also has the ability to relatively contain
fluids within
the stationary housing in the event that pressurized gas from the external
source
is temporarily lost, and the seal has the ability to return to normal
operation after
pressurized gas from the external source is restored.
These and further objects, features, and advantages of the present
invention will become apparent from the following detailed description,
wherein
reference is made to the figures in the accompanying drawings.
Brief Description of the Drawings
Figure 1 is a cross-sectional isometric view of a portion of a pump housing
and a seal housing according to the present invention, with a portion of the
rotating pump shaft cut off before passing through the seal housing.
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Figure 2 is an enlarged cross-sectional isometric view of a portion of the
seal housing as shown in Figure 1, and illustrating one embodiment of a double
gas seal in accordance with the present invention.
Figure 3 is a cross-sectional isometric view of the stationary ring shown
in Figure 2.
Figure 4 is a cross-sectional isometric view of the rotary ring shown in
Figure 2.
Figure 5 is an end view of the rotary ring shown in Figure 2, illustrating the
plurality of inner recesses, the plurality of outer recesses, and the
plurality of feed
channels.
Figure 6 is an end view of an alternate embodiment of a rotary ring
according to the present invention.
Detailed Description of Preferred Embodiments
Figure 1 illustrates one embodiment of the double gas pressure seal
assembly 10 according to the present invention for sealing the fluid within a
pump
housing. In the depicted application, a stationary seal housing 12 includes
.an
outer housing 14 and an inner housing 16. These housings may be structurally
connected by a plurality of circumferentially spaced bolts 15, and may be
disconnected to repair or replace the seal assembly 10. The pump includes a
shaft
18 which rotates about axis 20 and extends through the housing 8 of the pump,
with only a portion of the housing 8 shown in Figure 1. Pressurized process
fluid
in the pump is sealed between pump housing 8 and the inner housing 16 of the
seal
assembly 16 by gasket 22. Gasket 22, which is secured to the inner housing 16,
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thus seals against the face 7 of the pump housing 8. Process fluid in the pump
thus occupies the annular gap 24 (see Fig. 2) between an outer portion of the
gas
seal assembly 10 and both the inner housing 16 and the outer housing 14.
For the embodiment shown in Figure 1, the rotary housing is a shaft sleeve
26, which rotates with the shaft 18. A conventional pump drive collar 28
includes
circumferentially spaced ports 30 each for receiving a respective set screw
(not
shown) to secure the sleeve 26 to the shaft 18. A retaining ring 36 limits
axial
movement of the drive collar 28 with respect to the sleeve 26. A plurality of
circumferentially spaced centering disks 32 each secured to the outer housing
14
by a respective bolt 34 may be adjusted to properly center the axis 20 of the
shaft
18 within the housing 12, and may apply a desired prealignment to the drive
collar
28 and thus to the shaft sleeve 26. A static seal, such as o-ring 38, may
provide
the seal between the shaft 18 and the sleeve 26. It should be understood that
the
terms "stationary housing" and "rotary housing" as used herein are broadly
intended to refer to any stationary component and rotary component which
include a seal assembly as described subsequently for sealing between these
components during normal operation of the equipment. While the seal assembly
10 of the present invention is particularly well suited for sealing process
fluid
within a pump housing, the double gas seal of the present invention may be
used
in various applications for sealing between a stationary component and a
rotary
component. Exemplary equipment which may benefit from the seal assembly of
this invention includes pumps and blower fans used in various operations,
including chemical processing, hydrocarbon processing, and pulp and paper
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processing. The double gas seal of the present invention may be reliably used
to
seal various fluids within the equipment, including toxic fluids, with no
appreciable
discharge of the fluids from the equipment. All such fluids to be sealed
within the
equipment by the seal assembly of this invention are generally referred to
herein
as "process fluids."
Referring now to Figures 1 and 2, the gas pressure seal 10 includes a
rotary ring 40 and a stationary ring 42. The rotary ring 40 rotates with the
sleeve
26 and is interconnected thereto by a suitable retainer, such as pins or a
retaining
ring 9. If there is a loss of pressurized gas to the gas pressure seal
assembly 10,
item 9 maintains the rotary ring 40 in position and, in the Figure 1
illustration,
stops movement of the ring 40 to the right. The item 9 is thus important to
prevent overload between the stationary ring 42 and the rotary ring 40,
thereby
preventing excessive heat buildup in the seal assembly. The stationary ring 42
is
similarly rotatably secured by pins (not shown) extending through drive ring
58
to the outer housing 14. A static seal, such as o-ring seal 44, seals between
the
rotary ring 40 and the sleeve 26. A pair of pressure responsive cup-shaped
seals
46 and 48 each seal between the stationary ring 42 and the outer cylindrical
surface 50 and the inner cylindrical surface 52, respectively, on the outer
stationary housing 14. Another static o-ring seal 54 seals between the outer
housing 14 and the inner housing 16.
A coil spring or other suitable biasing member 56 is provided in the outer
housing 14 and acts on the drive ring 58, which in cross-section may have a
substantially E-shaped configuration. The drive ring 58 in turn acts on the
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stationary ring 42, which then presses the end surface 60 of the rotary ring
40 into
contact with the surface 62 on the shaft sleeve 26. As the ring 40 rotates
with
respect to the stationary ring 42, pressurized fluid in the pump housing is
sealed
between the planar face 64 on the rotary ring 40 and the planar face 66 on the
stationary ring 42. During normal operation, the faces 64 and 66 are not
actually
in physical contact, but instead are separated by a slight gap, typically less
than
0.0001 inches, which is the gap caused by the gas film which separates these
faces.
The stationary ring 42 includes an annular groove 68 which is discussed in
further
detail below. For the present, however, it should be understood that the faces
b4
and 66 thus form annular inner sealing faces radially inward of the annular
groove
68, and similarly form annular outer sealing faces radially outward of the
annular
groove 68. Each of these sealing faces lies within a plane perpendicular to
the axis
20, and preferably both the radially inner and radially outer portions of both
faces
64 and 66 lie within a single plane perpendicular to the axis 20.
Refernng still to Figures 1 and 2, pressurized gas, and preferably an inert
gas such as nitrogen, may be supplied from an external source, such as supply
cylinder 70. The pressurized gas is supplied to the input port 72 in the outer
housing 14 and is transmitted through drilled passageway 74 to the annular
cavity
76 in the stationary housing 14. Pressurized gas thus flows by the drive ring
58
and into the annular cavity 78 in the stationary ring 42 which is spaced
between
the outer annular leg 80 and the inner annular leg 82 of the stationary ring.
A
plurality of circumferentially spaced drilled supply ports 84 as shown in
Figure 3
provide fluid communication between the cavity 78 and the annular groove 68.
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The cross-section through Figure 3 illustrates two such circumferentially
spaced
ports 84, while the radial position of one of the ports is shown in dashed
lines in
Figure 2. Pressurized gas is thus continually provided to the annular groove
68,
and this gas pressure is maintained at a pressure higher than the anticipated
process fluid pressure within the pump housing 8. This inert gas pressure thus
also energizes the cup shaped seals 46 and 48 to provide a reliable seal
between
the stationary ring 42 and the outer housing 14.
The rotary ring 40 may be manufactured from the relatively hard material,
such as silicon carbide, while the stationary ring 42 may be manufactured from
a
more flexible material, such as carbon. Referring to Figure 3, the cross-
sectional
configuration of the stationary ring 42 is controlled such that the radially
thickest
portion 86 of this component is closely adjacent the face 66. The axial
spacing
between the face 66 and the portion 86 is thus less than 0.050 inches, and
preferably less than 0.040 inches. Moving axially further away from the face
66,
the radial thickness of the stationary ring 42 thereafter is reduced to form
the
reduced width portion 88. The axial length of the portion 86 is from 0.100
inches
to 0.150 inches, and preferably about 0.120 inches. The radial thickness of
the
portion 86 is preferably from 100% to 115% of the radial thickness of the
sealing
face 66, and preferably is about 105% to 115% of the radial thickness of the
sealing face 66. The portion 88 preferably has a radial thickness of about 70%
to
90% of the radial thickness of the sealing face 66, and preferably has a
radial
thickness of about 80% of the sealing face. The annular cavity 78 extends into
the
portion 88 and thus defines the outer leg 80 and inner leg 82 discussed above.
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Moving further away from the face 66, the thickness of each of the legs 80 and
82
is further restricted to form the relatively thin annular extensions 90 and 92
which
are configured to receive the seals 46 and 48. The axial depth of the annular
groove 78 may be controlled to provide a desired radial deflection capability
for
S the legs 80 and 82 so that these legs may deflect in response to a radial
pressure
differential and thereby maintain a desired angular alignment between the
stationary ring face 64 and the stationary housing face 66. The annular groove
78
extends axially into the portion 88, and preferably extends into at least 20%,
and
preferably from about 25% to 3 S%, into the axial length of the portion 88.
The
circumferentially spaced ports 84 have no appreciable effect on the
flexibility of
the portion 86. The recesses or pad faces themselves preferably should stay
relatively flat and parallel. The tabs 93 are provided to fix the rotational
position
of the ring 42 in place.
Figures 4 and 5 illustrate the plurality of circumferentially spaced inner
recesses 94 and the plurality of circumferentially spaced outer recesses 96
each
provided in the end face 64 of the rotary ring 40. Each of these recesses is
in fluid
communication with the annular groove 68 by a respective one of the plurality
of
the inner feed channels 98 and the outer feed channels 100 which supply
pressurized fluid to each respective recess. More particularly, each inner
recess
94 includes a rotary leading portion 102 and a rotary trailing portion 104,
and each
outer recess 96 similarly includes a rotary leading portion 106 and a rotary
trailing
portion 108. As shown in Figure 5, the rotary ring 40 is thus intended for
rotation
in the clockwise direction. A plurality of circumferentially spaced holes 110
as
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shown in Figure 4 may be provided in the rotary ring 40, with each hole 110
being
sized to receive a suitable pin or other securing member to rotatably lock the
rotary ring 40 to the sleeve 26.
By supplying the pressurized gas to the annular groove 68 and then
through the feed channels to the plurality of recesses or pads 94 and 96, it
should
be understood that the pressurized gas, which is sometimes referred to as a
buffer
gas, is supplied between the two sealing faces 64 and 66. This gas may thus
slowly leak radially outward to the annular gap 24 which is fluidly in
communication with the process side of the equipment, while the pressurized
gas
also slowly leaks radially inward to the annular gap 25 between the stationary
ring
42 and the sleeve 26, with the gap 25 being vented to atmosphere. Since the
process fluid is maintained radially outward of the sealing faces 64 and 66,
process
fluid contamination of the sealing faces is minimized.
The double gas coplanar sealing faces 64 and 66 as shown in Figure 2
provide two distinct coplanar sealing faces, with each of these faces
preferably
being within a single plane which is perpendicular to the axis 20 of the
rotary
sleeve 26. Both the stationary sealing face 66 and the rotary sealing face 64
thus
have a radially inward and a radially outward component, i.e., one sealing
face
radially inward of the groove 68 and another sealing face radially outward of
the
groove 68. Since each ring has both a radially inward and a radially outward
sealing face, interaction between the inner and outer sealing faces thus
occurs
whenever there is deflection in one of the faces. This configuration desirably
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provides a reduced space for the coplanar gas seal, and the parts of the seal
are
configured for easy replacement.
By providing the pads or recesses on both an inner portion and an outer
portion of the sealing faces, two different gas seals are effectively provided
with
the inner seal being an outside pressurized seal and the outer seal being an
inside
pressurized seal. The recesses or pad faces are configured to cause the
external
gas pressure to distribute itself across the seal faces so that a film of the
desired
"stii~ness' is achieved to keep the faces from touching while allowing the
film
thickness to reliably form the desired seal. The desired gas seal thus
includes a gas
film with a thickness between these faces so that, as the faces move closer
together, the net force produced by the pressure on each face increases
significantly.
The preferred design of the sealing faces and the circumferentially spaced
recesses will depend upon the application. For many applications, however, the
radial thickness between the annular groove and the radially innermost edge of
the
inner recess 94 will be about 80% of the radial thickness between the annular
groove and the radially innermost portion of the inner sealing face.
Similarly, the
radial thickness between the annular groove and the radially outermost edge of
the
outer recess 96 will be about 80% of the radial thickness between the annular
groove and the radially outermost portion of the outer sealing face.
In the event that the buffer gas pressure becomes less than the process
pressure, i.e., the pressure in the chamber 78 is less than the pressure in
the
annular gap 24, a u-cup seal 46 intentionally will leak so that the process
pressure
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increases the sealing effectiveness of the seal 48. At this time, the seal 46
effectively is performing no sealing function, but the desired sealing
function is still
maintained by the seal 48.
The inner and outer recesses 94 and 96 may either use step pads or tapered
S pads. In a step pad arrangement, the connecting channels 98 and 100 may each
have a depth of approximately 400 microinches, and typically from about 350 to
450 microinches. Each of the inner and outer recesses 94 and 96 have a uniform
depth of approximately 40% to 60% of the channel depth, i.e., typically from
about 150 to 250 microinches, and preferably about 200 microinches. In the
tapered pad arrangement, the radially extending channels 98 and 100 may each
have a depth of approximately 400 microinches, but the recess depth is
uniformly
tapered from the feed channel to the end of the recess, so that the rotary
leading
portion 102 and 106 of each recess has a depth of approximately 400
microinches,
while the rotary trailing portion 104 and 108 of each recess tapers to the pad
face.
In order to obtain the desired lift by the circumferentially spaced inner and
outer pads, each outer recess may have a radial width 140 of approximately 50%
of the spacing between the annular groove 68 and the outermost portion 144 of
the sealing face 64. The feed groove 100 may occupy approximately 15% of the
tangential space between recesses, and the recess 96 itself may use
approximately
60% of the tangential space between the recesses. As shown in Figure 5, the
tangential length 146 of the recess 96 is thus approximately 60% of the
tangential
length 148. Accordingly, the spacing 150 between the end of one recess 96 and
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the beginning of the next feed groove may be approximately 25% of the
tangential
spacing 148. The inner pad 94 is similarly proportioned. The inner pad radial
width 152 is thus approximately 50% of the spacing between the annular groove
68 and the innermost portion 155 of the sealing face 64. The plurality of
inner
recesses 94 and the plurality of outer recesses 96 define an area which is a
significant portion of the total area between the annular groove 68 and the
radially
innermost portion 155 and the radially outermost portion 144 of the sealing
face
64, respectively. Figure S illustrates an outer sealing face area 154 radially
outward of the groove 68 and an inner sealing face area 156 radially inward of
the
groove 68. The stiff gas film layer between the faces 64 and 66 thus fills
these
spaces 154 and 156 for each tangential length 148. The area of pad or recess
96
is at least 25% ofthe area 154, and preferably is at least 30% ofthe area 154.
The
area of pad or recess 94 is similarly at least 25% and preferably at least 30%
of the
area 156.
Figure 6 illustrates another embodiment of the invention, wherein the
rotary pad 120 includes a plurality of circumferentially spaced inner recesses
122
and a plurality of circumferentially spaced outer recesses 124. Inner feed
channels
126 extend between the annular groove and each respective inner recess, while
similar outer feed channels 128 provide fluid communication between the
annular
groove and each of the respective outer recesses.
In both of the embodiments shown in Figures 5 and 6, it is important that
both the inner recesses and outer recesses are configured with respect to the
feed
channels so that the feed channels supply pressurized fluid to the leading
portion
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of the recess and thus at a location circumferentially spaced from the
trailing
portion of the recess. When the recesses are provided on the rotary ring, the
leading portion of each recess is obviously the first portion of the recess
which
intersects an imaginary line, while the trailing portion of each recess
subsequently
intersects this imaginary line. Since the ring as shown in Figure 5 rotates in
the
clockwise direction, the leading portion 106 of the recess 96 thus first
intersects
the imaginary line 158, and the trailing portion 108 then passes by the line
158.
Similarly, a rotary leading portion 102 of a recess 94 first passes by the
line 158,
followed by the portion 104 of that same recess 94. The rotation of the ring
40
thus compresses the gas in each of the recesses so that gas pressure in the
trailing
portion of each recess is higher than gas pressure in the leading portion of
each
recess. The plurality of inner and outer recesses could be provided on the
stationary component rather than the rotary component, and in that case the
leading portion of each recess is the portion of the recess which first
becomes
tangentially aligned with a point on the rotary ring, while the trailing
portion of the
recess thereafter becomes tangentially aligned with the same point on the
rotary
ring. The movement of the rotary ring with respect to the recess thus still
pulls the
pressurized gas from the rotary leading portion of the recess to the rotary
trailing
portion of the recess, thus increasing gas pressure in the rotary trailing
portion of
the recess. Regardless of whether the recesses are provided on the rotary ring
or
the stationary ring, supplying the pressurized gas from the feed channels to
the
rotary leading portion of each recess ensures that pressure in the trailing
portion
of that recess will desirably increase to maintain the desired lifting effect
and thus
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maintain the desired stiff film to achieve reliable sealing with a minimum
loss of
buffer gas.
By utilizing cup shaped seals 46 and 48 rather than o-ring seals, reliable
gas leakage past the faces 64 and 66 may be more accurately maintained,
thereby
maintaining the desired film thickness between these faces 64 and 66. It has
been
determined that using the cup shaped seals 46 and 48 provides low friction to
maintain a desired gas seal between the faces 64 and 66, and this desired
reduced
friction is not easily obtained with an o-ring seal. In part, the increased fi-
iction
attributable to the use of o-ring seals rather than cup shaped seals may be
due to
expansion of the components as the pump heats up during continuous use.
In an alternate embodiment of the invention, the carbon ring has a reduced
axial length from the previously described embodiment, thereby making the
carbon
ring more flexible in bending. An elongated rubber sleeve may then be provided
between the metal drive ring and the stationary ring, so that the rubber
sleeve
axially separates the stationary ring and the drive ring and thereby provides
a high
degree of flexing between these components. The rubber sleeve may be
compressed by the drive ring to transmit mechanical force to the stationary
ring
and then to the rotary ring, but does not affect the stiffness of the
stationary ring.
One of the problems with this embodiment is that the slightest tangential
variation
in Youngs modulus of the rubber may produce waves and thus dragging on the
seal faces. Even small variations in the thickness of the rubber or the
flatness of
the mating surfaces may accordingly cause waves that may adversely affect the
desired film thickness.
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In yet another embodiment of the invention, an additional o-ring may be
provided on the stationary ring and upstream from the cup shaped seals. This
additional o-ring (not shown) would normally be provided on the exterior of
the
stationary ring 42 and out of contact and thus out of sealing engagement with
the
S cylindrical surface 50. In the event that pressurized gas to the chamber 76
is lost,
however, the location of the static o-ring 44 between the rotary ring 40 and
the
sleeve 26 would cause both the rotary ring 40 and the stationary ring 42 to
move
to the right, as shown in Figure 2, thereby bringing the additional o-ring
seal into
sealing engagement with the surface 50. This additional o-ring would thus keep
process fluid from leaking past the cup seal 46 and through the ports 84 and
the
annular groove 68 and then radially inward between the faces 64 and 66.
Initial
tests have indicated, however, that this design does not offer the same high
reliability as the design shown in the figures, and the robustness of the gas
seal
assembly to reliably operate under various conditions is adversely affected.
If gas
pressure is lost, however, this latter design should experience lower leakage
of
process fluid from the gas seal to the environment, although this lower
leakage
also may result in higher heat generation in the area of the sealing faces 64
and 66.
In other embodiments, the biasing spring 56 may be eliminated, and the
biasing force desired to press the axially movable one of the rings against
the other
ring may be provided by the external pressurized gas.
The feed channels that supply the pressurized gas to the recesses and the
recesses may be provided on either the stationary ring or the rotary ring. In
a
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preferred embodiment, both the feed channels and the recesses are provided on
the
rotary ring. Both the feed channels and the recesses are also preferably
provided
in the ring which is formed from the hardest material for the rotary ring and
the
stationary ring. If the faces briefly touch during operation of the equipment,
some
wear the sealing faces will occur. Also, if the external pressurized gas is
briefly
interrupted, some wear between the sealing faces will accur. Preferably the
wear
will thus be concentrated on the ring which does not include the specially
formed
recesses and feed channels. When the rotary ring 40 as disclosed herein is
fabricated from silicon carbide material and contains both the recesses and
the feed
channels, wear between the faces will primarily occur on the softer carbon
stationary ring, which may then be replaced while the more expensive silicon
carbide rotary ring is reused.
The annular groove 68 may be provided on either the stationary ring or the
rotary ring. The cost of forming the annular groove is nominal, however, and
the
groove 68 may be easily provided on the softer material ring. The ports which
supply the pressurized gas to the annular groove are provided on the
stationary
nng.
Various other modifications to the gas seal and method of forming an
improved seal will be apparent from the above description of the preferred
embodiments. Although the invention has thus been described in detail for
various
embodiments, it should be understood that this is for illustration and the
invention
is not limited to the described embodiments. Alternate components and
operating
techniques will be apparent to those skilled in the art in view of this
disclosure.
CA 02325475 2000-09-22
WO 99/49244 PCT/US99/05319
-20-
Additional modifications are thus contemplated and may be made without
departing from the spirit of the invention, which is defined by the following
claims.
