U.S. patent number 3,770,179 [Application Number 05/132,485] was granted by the patent office on 1973-11-06 for self-pressurizing seal for rotary shafts.
This patent grant is currently assigned to General Electric Company. Invention is credited to James Dennis McHugh.
United States Patent |
3,770,179 |
McHugh |
November 6, 1973 |
SELF-PRESSURIZING SEAL FOR ROTARY SHAFTS
Abstract
A self-pressurizing shaft seal for an oil filled submersible
motor is characterized by an inboard spiral grooved face seal and
an outboard conventional face seal disposed in series relationship
along the motor shaft. During operation, the inboard spiral grooved
face seal pumps oil from the motor interior into a substantially
confined zone between the seals to increase the oil pressure at the
outboard face seal without the necessity for structurally
strengthening the entire motor housing. Also disclosed is the
disposition of a shoulder on the spiral grooved seal runner face
remote from the pumping interface to permit the oil pressure within
the confined zone to hydraulicly increase the axial force upon the
runner thereby increasing the pumping pressure of the spiral
grooved seal in boot strap fashion. Other disclosed seals contain
means for measuring the pressure within the confined zone to
actuate remote signaling devices upon a failure of the outboard
seal as well as spiral grooved face seals having valving means to
alter the pumping rate of the inboard seal upon a loss of pressure
in the confined zone between seals.
Inventors: |
McHugh; James Dennis (Santa
Clara, CA) |
Assignee: |
General Electric Company
(Schenectady, NY)
|
Family
ID: |
22454272 |
Appl.
No.: |
05/132,485 |
Filed: |
April 8, 1971 |
Current U.S.
Class: |
277/318; 277/365;
277/400; 277/402 |
Current CPC
Class: |
F16J
15/3412 (20130101); F16J 15/40 (20130101) |
Current International
Class: |
F16J
15/40 (20060101); F16J 15/34 (20060101); F16j
009/00 (); F16j 015/38 () |
Field of
Search: |
;277/2,3,61,65 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Rothberg; Samuel B.
Claims
What I claim as new and desire to secure by Letters Patent of the
United States of America is:
1. A shaft seal for a rotatable machine to inhibit ingress of
ambient fluid into said machine, said seal comprising an inboard
pumping seal characterized by an annular running member mounted
upon said rotatable shaft in juxtaposition with an annular
stationary member disposed in a confronting attitude relative to
said running member, at least one of said juxtaposed members having
viscosity grooves therein extending from a peripheral edge of said
member and terminating in a land along a planar face of the member
to pump sealing fluid contained within said machine into a
relatively confined zone to substantially increase the pressure of
said sealing fluid within said zone relative to the sealing fluid
pressure within said machine, and face seal means axially mounted
upon said shaft at an outboard location relative to said inboard
seal to restrict the out flow of sealing fluid from said confined
zone, said face seal means including a rotorary member axially
mounted upon said rotatable shaft, a stationary member juxtaposed
in a co-planar attitude relative to said rotary member, and means
including a pressure actuated member for driving said rotary and
stationary members toward mutual contact.
2. A seal having a rotatable shaft according to claim 1 wherein
said inboard pumping seal and said face seal means are disposed in
tandem along said shaft, the pressure of said sealing fluid in said
confined zone tending to reduce the span between the stationary and
rotary members of said face seal means.
3. A seal having a rotatable shaft according to claim 1 wherein
said inboard pumping seal and said face seal means are disposed in
back-to-back configuration along said shaft, the pressure of said
sealing fluid in said confined zone tending to increase the span
between the stationary and rotary members of said outboard face
seal means.
4. A seal for a rotatable shaft according to claim 1 wherein one
member of said inboard pumping seal is axially slidable along said
shaft and further including a shoulder notched within the face of
said axially slidable member situated remote from the stationary
member forming said seal, said shoulder being in communication with
said sealing fluid within said high pressure zone to provide a
hydraulic force tending to bias said annular running and annular
stationary members into mutual contact.
5. A seal for a rotatable shaft according to claim 4 further
including an O-Ring seal between said rotary member and said shaft
to seal said rotary member upon said shaft.
6. A seal having a rotatable shaft according to claim 1 further
including means for measuring the pressure within said confined
zone and remote signal means responsive to said pressure measuring
means for indicating a reduction in the pressure of said confined
zone below a predetermined minimum.
7. A shaft seal for a rotatable machine to inhibit ingress of
ambient fluid into said machine, said seal comprising a spiral
grooved face seal disposed at an inboard location along said shaft
for pumping sealing fluid from said machine into a substantially
confined zone to increase the sealing fluid pressure within said
machine, said spiral grooved face seal comprising coaxial rotary
and stationary members juxtaposed in a co-planar attitude, at least
one of said members being axially slidable along said shaft to vary
the span between said members, mechanical means biasing said
axially slidable member toward said stationary member, a shoulder
notched within said axially slidable member face remote from said
stationary member, said shoulder being in communication with the
sealing fluid of said confined zone to bias said axially slidable
member towards said stationary member of said spiral grooved face
seal with increased sealing fluid pressure in said confined zone,
and pressure responsive face seal means disposed at an axially
outboard location upon said rotatable shaft and operable in
response to increased fluid pressure in said confined zone to
restrict the flow of sealing fluid from said confined zone.
8. A seal for a rotatable shaft according to claim 7 wherein the
grooves in said spiral grooved face seal extend from the periphery
of said seal to terminate in an annular land.
Description
This invention relates to a shaft seal for a rotatable machine and,
more particularly, to a shaft seal wherein a high pressure zone of
sealing fluid is formed between a spiral grooved face seal and a
conventional face seal to inhibit ingress of contaminating fluid
into the machine.
One of the major factors limiting the life of submersible motors is
water in-pumping at the shaft seal produced by a slight
eccentricity in the face seal customarily employed to assure
minimum leakage at the shaft. Although water in-pumping can be
overcome by substantially increasing the pressure differential
between the sealing fluid, e.g., oil, typically contained within
the motor and the ambient water, higher pressure differentials
necessarily require reinforcement of the motor housing as well as
substantial alterations in the spring biased diaphragm customarily
utilized to produce the oil/water pressure differential.
Because of the difficulties associated with increaisng the oil
pressure within the motor, a number of different seal
configurations have been proposed to inhibit in-pumping
notwithstanding a low oil/water pressure differential. For example,
rotor shafts have been sealed utilizing an external pump to produce
high and low pressures within sealing chambers situated at axially
displaced locations along the shaft. Similarly, it has heretofore
been proposed that the shaft of a centrifugal pump be sealed
utilizing the rotary speed of the shaft to pump oil from an axially
outboard location to an inboard seal to restrict gas leakage from
the pump. I also have proposed in my co-pending U.S. Pat. No.
3,704,019, issued Nov. 28, 1972, and assigned to the assignee of
the present invention utilization of a spiral grooved face seal
having deep helical grooves to increase the pressure at the seal
interface without increasing the outpumping rate of oil from the
motor. While all these designs have certain advantages, there still
remains a need for seals of different designs with differing
capabilities.
It is therefore an object of this invention to provide a novel
self-pressurizing seal characterized by low leakage.
It is also an object of this invention to provide a seal adaptable
to monitoring at an external location to assure proper seal
functioning.
It is a further object of this invention to provide a
self-pressurizing seal wherein the pressure of the sealing liquid
within the seal is employed to augment the mechanical bias of the
seal thereby maximizing the obtainable pressure from the seal
without extensive wear of the seal during start-up.
It is a still further object of this invention to provide a
self-pressurizing seal wherein automatic closure of the seal is
effected upon a reduction in seal pressure.
These and other objects of this invention generally are achieved by
a self-pressurizing seal for a rotatable machine characterized by
an inboard pumping seal having an annular running member mounted
upon a rotatable shaft in juxtaposition with an annular co-planar
stationary member. At least one of the juxtaposed members is
provided with spiral grooves extending from the perimeter of the
member to a land along the planar face of the member to pump
sealing fluid from the rotatable machine into a substantially
confined zone. The pumping action of the inboard seal increases the
pressure of the sealing fluid within the zone relative to the
sealing fluid pressure within the machine and conventional face
seal means are disposed along the shaft at an axially outboard
location (relative to the inboard pumping seal) to restrict the
flow of sealing fluid from the high pressure zone into the ambient
water. Because sealing fluid at relatively high pressure is
situated only within a zone intermediate the axially displaced face
seals, ingress of water into the motor is inhibited without
structural reinforcement of the entire motor housing and without
subjecting the necessary flexible oil expansion system to large
pressure differences.
Although the features of this invention are defined with
particularity in the appended claims, a more complete understanding
of the invention may be obtained from the following detailed
description of various specific embodiments when taken in
conjunction with the appended drawings therein:
FIG. 1 is an enlarged sectional view of a self-pressurizing seal in
accordance with this invention,
FIG. 2 is a plan view of one member of the inboard face seal
illustrating the disposition of spiral grooves therein,
FIG. 3 is a sectional view of a self-pressurizing seal wherein the
hydraulic pressure of the oil within the seal is employed to
increase the obtainable seal pressure,
FIG. 4 is an enlarged sectional view illustrating the force
distribution along the spiral grooved seal of FIG. 3,
FIG. 5 is an alternate seal configuration illustrating a seal
monitoring device in accordance with this invention,
FIG. 6 is a sectional view of a spiral grooved seal wherein the
outpumping rate of the seal is reduced upon a reduction in the
outboard seal pressure, and
FIG. 7 is a plan view of the stationary member forming the seal of
FIG. 6.
A self-pressurizing seal 10 in accordance with this invention is
illustrated in FIG. 1 and generally includes an inboard spiral
grooved face seal 12 and an outboard face seal 14 disposed in
tandem upon shaft 16 of a dynamoelectric machine, e.g., the pump
motor such as is described in U.S. Pat. No. 2,790,916, issued Apr.
30, 1957 to M.B. Hinman (the entire disclosure of which patent is
incorporated herein by reference). Typically, the pump motor
contains a sealing fluid, e.g., transformer oil 20, biased by a
flexible diaphragm to increase the pressure of the oil
approximately 5 psi relative to the water 24 which forms the
ambient environment for the motor during operation. The oil within
the motor is in communication with the radially outer surface of
inboard seal 12 and is pumped by the spiral grooves of the inboard
face seal into substantially closed annular oil chamber 26 thereby
increasing the oil pressure of the chamber relative to the oil
pressure within the pump motor.
Spiral grooved face seal 12 generally is characterized by an
annular carbon runner 28 mounted upon shaft 16 with planar face 30
of the runner being disposed in a confronting attitude with planar
face 32 of ceramic stationary member 34 fixedly secured to pump
motor housing 36. One of the planar faces of face seal 12,
illustrated in FIG. 1 as face 32 of stationary member 34, has
spiral grooves 22 therein to pump oil from the motor upon rotation
of runner 28 relative to stationary member 34. The grooves, shown
more clearly in FIG. 2, have a geometric configuration and density
dependent upon the quantity of pumping desired by the face seal (as
will be more fully explained hereinafter) and desirably extend
radially from the outer circumferential edge of annular stationary
member 34 to an annular land 38 separating the grooves from central
aperture 40 extending axially through the member. In the event a
failure of outboard face seal 14 should necessitate a shutdown of
the motor, land 38 advantageously functions to block back flow of
water through spiral grooved face seal 12.
Returning again to FIG. 1, the face of carbon runner 28 remote from
planar face 30 is notched to form a lower shoulder 42 which, in
association with backing plate 44, serves to house O-Ring 46
sealing the carbon runner to shaft 16. A second shoulder 48 also is
formed along the radially outer face of carbon runner 28 to seat a
generally L-shaped brass ferrule 50 biased against the runner by
spring 52. To permit axial movement of the ferrule along shaft 16
while restricting movement of the ferrule in a plane perpendicular
to the shaft, elongated body 54 of the ferrule is slidably engaged
within a guide 56 fixedly secured to the motor shaft.
Outboard face seal 14 is conventional in design and is mounted in
tandem with spiral grooved face seal 12 so that the pressure of the
oil within chamber 26 tends to close carbon runner 56 upon
confronting ceramic stationary member 58. A biasing spring 60
augments the oil pressure tending to close the face seal by
providing an axial force against upper extension 62 of ferrule 64
to drive inwardly extending backing plate 66 toward carbon runner
56. The edge of ferrule 64 proximate spiral grooved face seal 12
extends through guide 68 to limit the axial movement of the ferrule
while sealing of the runner to the shaft is accomplished by a
flexible bellows 70 fixedly secured between the shaft and the
overlying ferrule.
To inhibit ingress of solid contamination into the motor, a sand
slinger 71 is secured to motor shaft 16 at an axial location to
shroud the radially outer edge of outboard seal carrier 73. The
seal carrier is fixedly mounted to the motor housing 36 by bolts 74
passing through suitable apertures in the outer flange of the seal
carrier while a radially inner notch in the seal carrier serves as
a seat for stationary member 58 of face seal 14.
During operation of the motor, the rotary motion of carbon runner
28 relative to spiral grooved stationary member 34 pumps oil from
the motor housing into annular oil chamber 26 to increase the oil
pressure within the chamber to a predetermined level dependent
primarily upon the anticipated water inpumping force at outboard
face seal 14 resulting from eccentricity in the outboard seal. This
predetermined pressure level can be calculated (in accordance with
the teachings of an article entitled "Inward Pumping in Mechanical
Face Seals, by J.A. Findlay, presented as paper No. 68 at the Lub 2
ASME-ASIE Lubrication Conference, Atlantic City, N.J., Oct. 1-10,
1968,) from the formula:
.DELTA.p/e = 3.epsilon. cos.alpha..sup.. .mu..omega. (R.sub.o
-R.sub.i)/h.sup.2 (1+1.sup.. 5.epsilon..sup.2) (1)
wherein
.DELTA.p/e is the required oil pressure in lbs./in..sup.2 for each
inch eccentricity (e) of outboard face seal 14,
.epsilon. is the maximum tilt contemplated for face seal 14,
.omega. is the shaft speed in radians per second,
R.sub.o -R.sub.i is the radial span of the juxtaposed faces forming
seal 14 in inches,
cos.alpha. is the maximum misalignment contemplated for face seal
14,
.mu. is the viscosity of the water presumed to penetrate the seal
interface, in lb-sec./in..sup.2, and
h is the average oil film thickness between faces of the seal in
inches. Typically, an oil pressure increase of approximately 4,000
lbs./sq. in. is required to compensate for each inch of shaft
eccentricity to assure zero inpumping at the outboard face
seal.
Although the seal eccentricity can vary dependent upon such factors
as the amount of shaft runout under load and speed, the
out-of-roundness of the carbon washer, etc., the total eccentricity
generally can be estimated with a high degree of reliability for
any given manufacturing procedure. Thus, if manufacturing
experience has indicated that an eccentricity of approximately
0.010 inch normally is not exceeded on fabricated face seals, the
pressure required for chamber 26 to prevent water inpumping is
calculated by multiplying the maximum observed eccentricity by the
pressure per inch of face seal eccentricity as calculated by the
foregoing Findlay equation, e.g., for an empirically determined
maximum eccentricity of approximatley 0.010 inch and a calculated
oil pressure of 4,000 psi per inch eccentricity, a total pressure
of 40 psi is required in oil chamber 26 to inhibit inpumping.
The outpumping rate at outboard face seal 14 also must be
considered to assure that the oil supply within the motor is not
exhausted within an unduly short time in an attempt to inhibit
water ingress through the face seal. The outpumping rate for the
outboard seal therefore is calculated, e.g., from the formula:
q = .DELTA. p .pi. R.sub.i h.sup.3 /6.mu. .DELTA. R (2)
wherein
q is the outpumping rate,
.DELTA.p is the difference in pressure across face seal 14 in
psi,
h is the average film thickness between juxtaposed faces of the
seal in inches,
R.sub.i is the radius to the inner edge of the sealing land,
.mu. is the viscosity of oil in the seal interface in
lb.sec./in..sup.2, and
.DELTA.R is equal to the radial span of the juxtaposed faces
forming the seal in inches. The optimum pressure for oil chamber 26
then is chosen as a compromise between the high oil pressure
desired to overcome inpumping of water into the motor and the low
oil pressure desired to limit the oil outpumping rate at the
outboard face seal.
Once the pressure desired for annular oil chamber 26 has been
chosen, the geometric configuration of inboard spiral groove face
seal 12 required to produce this pressure can be determined in
accordance with the teachings of E.A. Muijderman in an article
entitled SPIRAL GROOVE BEARINGS published 1966 by Philips Technical
Library. One spiral grooved face seal configuration found suitable
for a 12 inch submersible motor having a 2 inch rotatable shaft was
characterized by 10 equally spaced grooves notched to a depth of
0.0013 inch and extending at a spiral angle of 15.degree. with a
groove land to width ratio of 1. The inner and outer diameters of
the seal measured 1.87 inches and 2.37 inches, respectively, while
the groove inner diameter measured 1.95 inches. With the foregoing
seal rotating at a speed of 30 revolutions per second, a maximum
pressure of 67.5 psi was observed in annular oil chamber 26.
FIG. 3 illustrates an improved embodiment of this invention whereby
the force of the spiral grooved face seal biasing spring can be
reduced without a reduction in the pressure obtainable from the
face seal. To achieve this result, a shoulder 72 is notched in
carbon runner 28A at an outboard location relative to O-Ring 46A
thereby permitting pressurized oil within annular oil chamber 26A
to communicate with face 75 and hydraulically drive runner 28A
axially towards mutual contact with stationary member 34A as the
pressure within the oil chamber increases. Although a shoulder 76
is provided in shaft 16A to seal notched runner 28A and the
position of the back support for spring 60A has been changed
slightly, the self-pressurizing face seal of FIG. 3 otherwise is
substantially identical to the face seal illustrated in FIG. 1. The
increased axial force upon runner 28A, however, resulting from
hydraulic pressure on face 75 reduces the gap of the spiral grooved
face seal tending to increase the obtainable pressure from the
seal. This increased pressure, in turn, results in an increased
hydraulic force upon face 75 and the pressure within annular oil
chamber 26A is increased in bootstrap fashion until an equilibrium
pressure is reached.
Assuming zero net flow at outboard face seal 14A, the pressure
generated by spiral grooved face seal 12A (illustrated by pressure
diagram P in FIG. 4) increases approximately linearly from the
outer periphery of stationary member 34A to the inner extent of the
grooves in the stationary member, i.e., from d.sub.3 to d.sub.2
with the pressure along the ungrooved portion of the seal
interface, i.e., from d.sub.2 to d.sub.1, remaining constant at
P.sub.MAX. For simplicity, the average pressure acting over the
area between d.sub.2 and d.sub.3 may be assumed equal to 1/2
P.sub.MAX. The maximum pressure at equilibrium therefore can be
estimated from the approximate formula: ##SPC1##
wherein
F.sub.s is the axial load upon the seal produced by spring 52A in
pounds,
d.sub.2 is the internal diameter of the spiral grooved annular
portion of the face seal,
d.sub.3 is the external diameter of the spiral grooved annular
portion of the face seal, and
d.sub.4 is the diameter of hydraulic shoulder 72 formed in carbon
runner 28A. One bootstrap seal having a seal inner diameter (i.e.,
d.sub.1) of 1.87 inches, a groove inner diameter (i.e., d.sub.2) of
1.95 inches, a seal outside diameter (i.e., d.sub.3) of 2.374
inches and a seal balance diameter (i.e., d.sub.4) of 2.0 inches
produced a hydraulic load of 28.7 lbs. upon the face seal in
addition to a bias of 42 lbs. provided by spring 52A for a total
face seal axial load of approximately 70.7 lbs. The spiral grooved
runner of the face seal contained 15 equally spaced grooves
disposed at a spiral angle of 15.degree. and the runner was rotated
at a speed of approximately 30 revolutions per second.
When the required pressure for the intermediate oil chamber is low,
e.g., approximately 20 psi, the outboard face seal can be disposed
in a back-to-back configuration with the inboard spiral grooved
face seal as illustrated in FIG. 5. The pressure within oil chamber
26B then applies an axial force upon outboard face seal 14B tending
to separate the confronting faces of the seal requiring a biasing
spring 60B having an axial force sufficient to overcome the
hydraulic pressure within chamber 26B to maintain the desired
outboard face seal opening during operation. If a failure of
pressure should occur within chamber 26B, the hydraulic force
tending to maintain the outboard seal open would be removed and
biasing spring 60B would tend to close the faces of the outboard
seal inhibiting ingress of water into the motor. When the
back-to-back seal arrangement is utilized with relatively high seal
pressures, e.g., pressures of approximately 60 psi, care must be
taken to choose a biasing spring 60B having sufficient force to
inhibit excessive outpumping of oil through the outboard face
seal.
A major feature of this invention is the ability to monitor seal
operation at an external location by the disposition of a pressure
transducer 77 within oil chamber 26B as illustrated in FIG. 5. The
pressure transducer is connected in series with an alarm 78 and a
voltage source, e.g., a transformer 80 having a primary winding 80A
connected across the motor energization leads (not shown), and
functions to close the series circuit upon a reduction in pressure
within oil chamber 26B below a predetermined minimum. Alarm 78 then
is sounded permitting shutdown and removal of the motor from a
submerged location prior to permanent damage of the motor interior
by water seepage therein. Should the pressure drop in chamber 26B
be produced by a failure of outboard face seal 14B, seepage of
water through spiral grooved face seal 12B during shutdown is
inhibited by annular land 38 of the face seal. To effectively
function as a flow restricter during motor shutdown resulting from
failure of outboard seal 14B, the annular land desirably should
have a radial span of at least 0.04 inches.
A self-contained motor protective device is illustrated in FIGS. 6
and 7 wherein a spring loaded pressure relief valve 82 is employed
to alter the operation of inboard spiral grooved seal 12C from a
full film to a solid-solid contacting mode in the event of failure
of the outboard seal. Relief valve 82 functions to restrict the
flow of oil from an annular groove 84 situated at the radially
inner terminus of the spiral grooves 22C to a bypass port 86 during
normal operation of the self-pressurizing seal. If the outboard
face seal should fail during motor operation reducing the pressure
within oil chamber 26C confined between the face seals, the
hydraulic pressure on piston 87 of valve 82 communicated to the
valve through axial aperture 88 also drops and the relatively
higher pressure of the oil within annular groove 84 overcomes the
bias of spring 90 to relieve the pressure at the seal interface
through bypass port 86.
With valve 82 open, the operation of spiral grooved face seal 12C
shifts from a conventional thick film operation, i.e., a film in
excess of approximatley 100 microinches typically produced by a
conventional groove depth of 1,000 to 1,500 microinches, to a
solid-solid contacting mode, i.e., a film width below approximately
50 microinches, substantially limiting the outpumping rate of the
spiral grooved face seal. Thus, a portion of the oil pumped by the
spiral grooved face seal is valved back to the suction side of the
face seal thereby reducing both the maximum pressure generated
between faces of the spiral grooved face seal and the quantity of
oil pumped into oil chamber 26C.
It will be appreciated that the spiral groove face seal will tend
to close, even without operation of relief valve 82, upon failure
of the outboard seal because of increased maximum pressure at the
spiral groove seal interface resulting in a changed oil
distribution at the seal interface. If the maximum pressure
required by the seal under conditions of leakage exceeds the
maximum generating capacity of the seal, the seal will inherently
change from a full film mode to a solid-solid contact mode to
reduce the outpumping rate. Thus, by careful choice of spiral
groove design, e.g., spiral groove width, depth and length, a seal
can be fabricated wherein the desired pressure will be produced
with outboard seal 14C functioning properly in a full film mode
while a substantially reduced outpumping rate is produced upon
failure of the outboard seal.
The previously cited formula (3) for estimating the maximum
pressure rise clearly illustrates the effect of the shoulder
d.sub.4 of FIG. 4 upon the pressure created. Formula (3) assumed an
average pressure P.sub.MAX /2 over that portion of the seal
interface where the pressure changes. An alternate, theoretically
exact formula for calculating the maximum pressure rise may be
obtained by integrating the assumed linear pressure rise over the
area between diameters d.sub.2 and d.sub.3 of FIG. 4. If the seal
diameter d.sub.4 is equal to the diameter d.sub.1, no shaft
shoulder exists and the formula for calculating pressure rise
becomes: ##SPC2##
wherein
P.sub.MAX is the maximum pressure generated by the seal in psi in a
full film mode with zero leakage,
F.sub.s is the total force applied to the seal by the biasing
spring in pounds,
d.sub.1 is the span from the radially inner face of the face seal
to the shaft axis,
d.sub.2 is the span from the radially inward end of the spiral
grooves to the shaft axis, and
d.sub.3 is the span from the radially outer periphery of the spiral
grooves to the shaft axis,
The ratio of the maximum pressure developed at the pumping seal
with leakage at outboard seal 14 interface relative to the maximum
pressure capable of being developed by the seal with zero leakage
then can be calculated from the formula: ##SPC3##
From this ratio, the maximum pressure capable of being developed by
seal 12 with no restriction in leakage at the outboard seal can be
calculated to provide an indication of film thickness arising from
the pressure increase. When the calculated maximum film pressure
under leakage conditions exceeds the maximum generating capability
of the seal (as can be calculated from the heretofore cited
Muijderman publication), the seal operation changes from a full
film mode to a solid-solid contact mode upon failure of the
outboard seal.
It should be appreciated that very shallow (e.g., 50 microinches)
or very deep (e.g., 20,000 microinches as described in my
heretofore cited patent application, Ser. No. 47,824) grooves can
be utilized for the inboard face seal to reduce outpumping upon
failure of the outboard seal. However, because the pressures
produced by these face seals during normal operation is difficult
to predict due to variations in fluid viscosity at the seal
interface, such seals generally are not recommended for the inboard
face seal.
* * * * *