CA2147739A1 - Face seal with double groove arrangement - Google Patents
Face seal with double groove arrangementInfo
- Publication number
- CA2147739A1 CA2147739A1 CA 2147739 CA2147739A CA2147739A1 CA 2147739 A1 CA2147739 A1 CA 2147739A1 CA 2147739 CA2147739 CA 2147739 CA 2147739 A CA2147739 A CA 2147739A CA 2147739 A1 CA2147739 A1 CA 2147739A1
- Authority
- CA
- Canada
- Prior art keywords
- groove
- grooves
- seal
- depth
- adjacent
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Abandoned
Links
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
- F16J—PISTONS; CYLINDERS; SEALINGS
- F16J15/00—Sealings
- F16J15/16—Sealings between relatively-moving surfaces
- F16J15/34—Sealings between relatively-moving surfaces with slip-ring pressed against a more or less radial face on one member
- F16J15/3404—Sealings between relatively-moving surfaces with slip-ring pressed against a more or less radial face on one member and characterised by parts or details relating to lubrication, cooling or venting of the seal
- F16J15/3408—Sealings between relatively-moving surfaces with slip-ring pressed against a more or less radial face on one member and characterised by parts or details relating to lubrication, cooling or venting of the seal at least one ring having an uneven slipping surface
- F16J15/3412—Sealings between relatively-moving surfaces with slip-ring pressed against a more or less radial face on one member and characterised by parts or details relating to lubrication, cooling or venting of the seal at least one ring having an uneven slipping surface with cavities
Abstract
2147739 9506212 PCTABS00036 A non-contacting grooved face seal for a shaft rotating at high speed with a combination of two groove patterns (34, 44) on one of the two sealing faces of mating seal rings; one pattern being relatively deep angled grooves (34), and the other pattern being relatively shallow grooves (44). The relatively deep groove pattern (34) is optimized for hydrodynamic operation and on shaft rotation pumps the sealed fluid between the sealing faces to set the running clearance. The relatively shallow pattern (44) is designed to hydrostatically prevent a friction lock of the sealing faces when the shaft is at or near to a stationary condition.
Description
; `'.: WO 95/062122 1 4 7 7 ~ 9 ~CT/US93tO828g FACE SEAL W:l:TH DOUBI.E
GROOVE ARRANG~ENT
, FIELD OF THE INVENTION ..
This invention relates to sealing dèvices for -~
rotating shafts, wherein a sealed fluid is employed to .
generate hydrostatic-hydrodynamic or aerostatic-aerodynami.c forces bètween opposed interacting face~
type sealing elements, one stationary and the other -rotating. These forces provide far slight separation and non-c~ntacting operation of the sealing elements, :~
thereby minimizing face wear and friction power losses 10 ~ while maintaining low fluid leakaye.
BACKGROUND OF THE INVENTION
: Rotary fluid :film face seals, also called gap or non-contacting face ~eals, are usually appIied to high~
speed andlor high-pressure rotating:equipment wherein ~ ;
the use of ~rdinary~mechanical face seals with ~ace j .
' contact woul~ result in excessive heat generation and wear. Non-contacting operation a~oids this undesirable ~ ~
face c~ntact at times when the~ shaft~is~rotating above a ~`` .
certain minimum speed, which is cal:led a li~t off spe There are various ways of accompIishing~the above ?., : non-contacting operati~n~ One Qf the more c~mmonly usad ~ ~' ways inc~udes the formation of a shallow spiral ~roove ~ .
pattern in one of the: ealing faces. The sealing face . .~
W0~/06212 PCl~S93/082g~
GROOVE ARRANG~ENT
, FIELD OF THE INVENTION ..
This invention relates to sealing dèvices for -~
rotating shafts, wherein a sealed fluid is employed to .
generate hydrostatic-hydrodynamic or aerostatic-aerodynami.c forces bètween opposed interacting face~
type sealing elements, one stationary and the other -rotating. These forces provide far slight separation and non-c~ntacting operation of the sealing elements, :~
thereby minimizing face wear and friction power losses 10 ~ while maintaining low fluid leakaye.
BACKGROUND OF THE INVENTION
: Rotary fluid :film face seals, also called gap or non-contacting face ~eals, are usually appIied to high~
speed andlor high-pressure rotating:equipment wherein ~ ;
the use of ~rdinary~mechanical face seals with ~ace j .
' contact woul~ result in excessive heat generation and wear. Non-contacting operation a~oids this undesirable ~ ~
face c~ntact at times when the~ shaft~is~rotating above a ~`` .
certain minimum speed, which is cal:led a li~t off spe There are various ways of accompIishing~the above ?., : non-contacting operati~n~ One Qf the more c~mmonly usad ~ ~' ways inc~udes the formation of a shallow spiral ~roove ~ .
pattern in one of the: ealing faces. The sealing face . .~
W0~/06212 PCl~S93/082g~
2~477~ - 2 -opposite the grooved face is relatively flat and smooth.
The face area where these two sealing faces defin~ a sealing clearance is called the sealing interface.
The abover-mentioned spiral groove pattern on one of . ~
the sealing faces normally extends inward from the outer ~`-circumference and ends at a partisular face dia~eter called the groove diameter, which is larger then the inner diameter of the seal interface. The non-grooved , area between the groove diameter and the inner interface diameter serves as a restriction to fluid outflow.
Fluid deli~ered by the spiral pa~tern must pass through `-this restriction and it can do so only if the sealing faces separate. The way this works is through pressure build-up. Should the faces remain in contact/ fluid will be compressed just ahead of the restriction, thus building up pressure. The pressure causes separation ';
force which eventually becomes larger than the forces that hold the faces together. In that moment the sealing faces separate and allow the fiuid to escape.
During operation of the,sea}, an equilibrium establishes itself between fluid inflow through spiral pumping and fluid outflow through face sepa-ation. Face separation is therefore pre~ent as long as the seal is operating, which means as long as one face i5 rotating in relation to the opposite face. ~-However, spiral pumping is not the only f actor that determines the amount of the separation ~etween the ~ealin~ faces~ Just as the spirals are able to dri~
the fluid into the non-groove portion of the sealing ; 30 'inter~ace pas~ the groove diameter, so can the pressure diff~rential. If enough of a pressure difference exists betwean the groo~ed end of the interface and the non-grooved end, fluid will also be forced into the non- , grooved portion of the interface, thereby separating the faces and ~orming the clearance.
Both ways in which clearan e can be formed between the sealing faces, one with speed of rotation~ the other f - wo 95/OG212 2 1 4 7 7 3 9 PCT~593/08289 with pressure differential, are distinct and separate, even though the effects of both combine on the operating seal. If there is no pressure di~ference and the seal ~ace separation occurs strictly due to face rotation, forces due to fluid flow are known as hydrodynamic forces if the fluid sealed is a liquid, and a~rodynamic ' ~orces if the fluid sealed is a gas.
On the other hand, if there is no mutual rotation between the two sealing faces and face separation is strictly the consequence of pressure differential between both ends~of the sealing interface, forces due to fluid flow are called hydrostatic forces if the fluid sealed is a liquid, and aerostatics forces if the fluid sealed is a gas. In the following, the terms hydrostatic a~d hydrodynamic are used for both liquid and gas effects since these latter terms are more c~nventionally used when describing both liquid and gas seals.
A typical spiral qroove seal needs to provide acceptable performance in terms of leakag~ and the absence of ~ace contact during all regimes of seal op~ration. It must do so not only at top speed and p~ssure, but also at standstill, at start-up, acceleration, at periods of equipment warm-up or at shutdown. At normal operating conditions, pressure and speed vary constantly, which results in continuous adjustments to the running clearance. These adjustments ~-are automatic; one of the key properties of spiral :~
gro~ve seals is their self-adjustment capability. On change in speed or pressure, the face clearance adjusts automatically to a new set of conditions. Hydrostatic and hydrodynamic forces c~use this adjustm~nt.
The operating envelope:of speeds and pressures i~
usually very wide and a seal design of necessity must be a compromise. For its psrformance to be acceptable at : near-zero speed or pressure, it is less than optimum at ~ operating speed and pressure. ~This is simply due to the ~`
WO~5106212 pcT~ss3lo8289~
2 ~ 3 9 fact that, both in terms of pressure and speed, the seal has to be brought up to operating conditions from zero ~ `
speed and ~ero pressure differential.
Especially critical to seal operation is the start- .
up. If the seal is applied to a centrifugal gas compressor, the full suction pressure diff2rential is often imposed onto the seal before the shaft starts turning. This presents a danger in that the sealing faces will lock together with friction~ Face lock results when the hydrostatic force is insufficient-to counter pressure forces that maintain the seal faces in contact. Face lock can lead to seal destruction, in which excessive break-away friction between c~ntacting seal faces can cause heavy wear or breakage of internal seal components.
First t~1en, spiral grooYes must be able to separate the sealing faces hydrodynamically ~or full speed non-contacting operation. This normally requires fairly short and relatively deep spiral grooves. Second, the spiral groo~es must be able to unload the sealing faces hydrostatically for start/stops to prevent face lock.
For this, the grooves have to be extended in length.
The extended grooves in turn cause more separation and leakage during full speed operation. The full speed leakage o~ a typical 3.75 inch shaft seal with short and relatively deep spirals may be about .9 SCFM (i.e.
Standard Cubic Feet per Minute) at l,000 psig and lOrO00 rpm. However, full speed leakage for such a seal with extended grooves may reach 2.4 SCFM un~er the same ~30 `c~nditions, almost triple the previous vaiue. The constant burden of larger-than-necessary leakage represent significant operating costs and is highly . D~""~
undesirable.
Spiral groove design prac~ice goes back to US Patent No. 3,lOg,658 wherein two opposing spiral grooves pump oil against each other to develop a liquid baxrier capable of sealing a gas. Such an arrangement is ~ WO~5/06212 21 4 7 7 3 9 PCT1593/08289 limited in pressure as well as speed capability, as is inherent in the use of liquid forces to seal gas.
Another known arrangement is shown in US Patent No. 3,499,653. This interface design with partial spiral grooves relies heavily on hydrostatic effects.
The interface gap is designed with a tapered shape which is narrower at the non-grooved end and wider at the spiral grooYes. The effect of the spiral grooves and therefore the hydrodynamic forces are suppressed since spiral groove pumping becomes less effecti~e across ~he wider gaps. This likewise affects the stability of the seal and limits its top pressure and speed capability.
A further known arrangement is shown by US Patent ~o. 4,212,475. Here the spiral groove itself attempts to act both as a hydrostatic as well as a hydrodynami~
pattern and is used to eliminate ths need for the tap~red shape o~ the gap so that a considerable degree of spiral groove hydrodynamic force can be applied to impart a self-aligning property to the sealing ~;~
interface. The self-aligning property forces the sealing interface back towards a parallel po~ition, regardless of whether deviations from parallel position dur~ng seal operation occur in radial or tangential directions. This resulted in improvement stability and ;~
increased performance limits in terms o~ pressure and speed~
While the known fluid seals as briefly summarized above have attempted to provide both hydrodynamic~and , i:
hydrostatic sealing properties, nevertheless the known 3iO seals have;bjeen deficient with respect to their ability to optimize the combination of these hydrostatic and hydrodynamic properties so as to provide desirable hydrostatic properties which facilitate starting and stopping of seals while effectively minimizing or a~oiding direct face contact and minimizing face loading between the seals so that the assembly can be start2d up -~' with minimal friction to avoid severe frictional power ~, ',~
W0~5/Ofi~l~ PCT/US~3/08289 ~ ~ ~
~ 39 - 6 - l`
!
re~uirements and direct frictional wear between the faces, and at the same time provide desirable h~drodynamic properties between the relatively-rotatable seal faces under a wide range of operating co~ditions~ ;
particularly those involving high speed and high 1-~
pressure.
Accordingly, it is an object of this invention to provide an improved fluid seal of the type employing a grooved pattern on one of the opposed seal faces, which improved seal provides a more optimi~ed combination of hydrodynamic and hydrostatic sealing characteristics SQ ..
as to permit improved seal performance under a significantly greater range of operating conditions, including operating conditions ranging from start up to conditions involving high speed and high pressure. :
In the improved seal arrangement Qf the present invention, the groove pattern (which is typically defined on only one of the seal ~aces) includes first and ~econd groove arrangements which co~municate with one another, one being significantly deeper than the .;
other, whereby the deeper arrangement is particulaxly effective for providing the desired hydrodyna~ic characteristics, whereas the shallower groove arrangement is more effective for providing the desired ~:
hydrostatic charazteristics. At the same time, these arrangements are positloned such that the shallower .
arr~ngement is interposed generall~ between the deeper groove arrangement and a non-grooYed annular land or dam which effectively separates the groove pattern from the ! 30 lo~ pressure side of the seal, whereby desira~le hydrostatic and hydrodynamic seal properties can both be ~ ~-obtained but at the sam~ time leakage of sealing fluid (for example, a gas~ across the dam to the low pres~ure side is minimized so as to impro~e the performance efficiency of the ~eal.
In addition, this optimization o~ the seal propertîes and performance characteristics is further 21477~9 WO9~/06212 PCT~JS~3/08289 improved by optimizing the groove pattern or configuration relative to the surrounding lands defined ¦ :
on the seal face so that the fluid film which is created between the opposed seal faces provides a more uniform `
pressure distribution and sealing characteristics while minimizing distortion of the seal faoe, ~hich in turn assists in optimizing`the seal performance with minimum width of gap between the opposed seal faces while still ~:
avoiding or minimizing direct contact and frictional ` ;
wear between the opposed seal faces. ;.
In the improved seal of this invention,~ as briefly :~
discussed above, the groove pattern includes the deep ;~`
groove arrangement which i5 defined:by:a circumfer- .
entially arranged series of grooves which angle circum~
ferentially and radially inwardly from the surroundin;g :
hlgh-pressure side of the seal, which angled~grooves may be o~ spiral, circular or straight configuration. These angled grooves are relatively deep and project only ..
partway across the seal:face. The angled deep grooves, at their radially inner:ends, communicate with the ii shallow groove arrangemen~ which is positioned radially inwardly of the deeper groove arrangement,:but which is separate~ from the low pressure side of the seal by:the intermediate non-grooved annular land;or dam.~ This ~
:shallow groove~arrangement has a depth which is a small` :;.
fraotion of the deeper grooYe arrangement and is ~ e~f~ctive for: creati~g:a hydrostatic force:between the :: :opposed sealing~faces substantially in the central : : :~
region thereof as defined between the ra~ially outer and ~ :;
;l3 . inner~boundarieslof the seal interface~ a! preferred ~! ' `'.. ~`-embodiment, all grooves associated with the groove pattern are form~d: ;such ~hat the sides of a~jacent~
~:~ grooves ex~end generally in parallelism~with one another so that the intermediate land area:between~adjacent grooves~maintains a su~stantially~constant width, even adjacent thé radially;~:~inner ends~cf~the~grooves, to :
~àximize squeeze film effects~;~in the~;fluid whi h f1Ows W095/06212 PC~S~3/0828~
2~ 39 - 8 - 1 - over these lands and thus enhance the thrust bearing support these lands provide for avoidance of seal face contact at or near the full speed rotation. ¦ ;
The improved seal arrangement, as aforesaid, also . , preferably forms the shallow groove arrange~ent by a circumfere.ntially-spaced s2ries cf shall~w groove~ which are contiguous with and project radially inwardly from the inner ends of the angled deep grooves, which shallow grooves terminate at the dam. These shallow grooves pro~ide improved hydrostati~ seal characteristics in the central seal face region, and angle radially inwardly at a smaller angle (which angle is zero in a preferred ~mbodiment) re~ative to the radial direction than do the deep grooves so as to increase the land area between the adjacent shallow grooves, particularly adjacent the radially inner ends of the shallow grooves, to provide a ~etter fluid sgueeze film effect between the opposed seal faces during high speed rotation. -~
Further improvement to the hydrostatically effective relatively shallow inner groove pattern is aimed at reduction and elimination o~ any seal face distortions that might occur as a result of circum~erential non- ;
uniformity of hydrostatic pressure fields as these form above groove and land regions of the shallow groove patt~rn at conditions at or near to the zero speed of rotation. This improvement is a narrow and shallow circumferential groove interconnecting inner ends of the shallow inner gxoove pattern. Such a shallow circumferential groove acts to equali2e pressure field ' -nonl-uniformities circumferentially, as a result suppressing any ~ace distortions and producing a uniform face separation with no or only minimal face-to face ~;-contact e~en at extremely low ma~itudes of separation between the f~ces.
Other objects and purposes of the invention will be apparent to persons familiar with seals of this general :
'" '', `
,~ W095/062l2 21 4 7 7 3 9 PCT~S93/08289 _ g _ i .
txpe upon reading the following specific~tion and inspecting the accompanying drawings. i~
BRIEF DESCRIPTION OF THE DRAWINGS 1~
Figure 1 is a f~agmentary central sectional view illustrating a generally conventional fluid face seal arrangement, such as a grooved face ~e~l, associated with a rotating shaft.
Figure 2 is a view taken generally along line 2-2 in Figure 1 and illustrating the groove pattern associated with a face of the rotating seal ring according to an e~odiment of this invention.
Figure 3 is a fragmentary enlargement of a part of Figure 2 so as to illustrate the groove~pattern in ~-greater detail.
Figure 4 i5 a fragmentary sectional vi~w taken ~ubstantially along line 4~4 in Figure 3.
Figures 5 and 6 are views which correspond r ,', respectively to Figures 3 and 4 but illustrate a variation thereof. i~
~0 Figure 7 is a view similar to Figure 2 but showing a further variation of the inner groove pattern.
Figure 8 is a ~iew similar to Figure 2 but showing still a further and prsferred variation of the inner groove pattern.
Figure 9 is a fragmentary enlargement of a part of Figure 8 so as to illustrate the groove pattern in greatex detail. ~-;
Figure lO is a fragmentary sectional view taken ~-substantially along line lO-lO in Figure 9. ~ ;;
~;30 l ! ~igure~ 11 and 12 are views which correspond j ! j respectively to Figures 9 and lO but illustrate a variation thereof.
; Cert~in terminoIogy will be used in the following ~`
description for convènienca in reference onlyj and will ~ ~not be limiting. For exampla, the words~'lupwardly", j ~
; ~ "downwardly", "rightwardly"~and "leftwardly" will refer ''.. .' ' :~ ~ '.;i' W~g~/062l2 PCT~593/08289 ~` ~
~4,~39 -lo- ~`:
to direction~ in the drawings to which reference is made. The words "inwardly" and "outwardly" will re~er to directions toward and away from, respectively, the geometric center of the assembly and designated parts thereof. Said terminolo~y will include the words ~~ -specifically mentianed, derivatives ~hereof, and words of similar import.
DETAILED DESCRIPTION
Referring to Figure 1, therè is shown a typical grooved face seal assembly 10 and its environment. This environmenl comprisas a housing ll and a rotatable shaft 12 extending through said housing. The seial as~embly lO
is applied to seal a fluid (such as a pressurized ga~) within the annular s~ace 13 and to restrict its escape into the environment at 14. Basic components of the seal assembly incl~des an annular, axially movable but non-rotatable sealing ring 16 having a radially extending flat face 17 in opposed sealing relationship with a radially extending ~lat face 18 of an annular ~`
rotatable sealing ring 19 which is non-rotatably mounted on the shaft 12. ~ing 1~ normally rotates in the direction of the arrow ~Figure 2). The sealing ring 16 is located wi~hin cavity 21 of housing 11 and held sub-qtantially concentric to rotatable sealing ring 19.
Between housing 11 and the sealing ring 16 is a conventional anti rotation de~ice ~not shown) for ;
preventing rotation of ring 1~, as well a9 a plurality of springs 22 spaced equidistantly around the cavity 21. ! `~
Springs 22 urge the sealing ring 16 toward engagement with the sPalin~ ring 19~ An 0-rin~ 23 seals tha space between the sealing ring 16 and the housing 11. The -sealing ring 23 is retain~id in the axial position by a ~-sleeve 24 which is c~ncentric with and locked on the shaft 12, such as by locknut 25 threaded on shaft 12 as shown. 0-ring seal 26 precludis leakage between the ~ ;
sealing ring lg and the sha~t 12.
~ woss/o62l2 21 4 7 7 3 9 PCT~593l08289 ¦ ~
The radially extending face 18 of the sealing ring 19 and radiall~ extending ~ace 17 of sealing ring 16 are in sealing relationship~ and define an annular contact area 27 therebetween, this being the seal interface.
This seal interface 27 is defined by a surrounding outer ~`
diameter 28 of ring 19 and an inner di^..me~er 29 of ring 16, these being the diameters exposed to the high and -low pressure fluid respectively in the illustrated ~:
embodiment. In opèration, a very narrow clearance is ' maintained between the seal faces 17-18, due ~o a fluid ;.
~ilm as generated by a groove pattern (as described --below) formed in the sealing face 18 of the sealing ring `~
19. Alternately, the groove pattern San be formed in the sealing fae 17 of the sealing ring 16 and stiIl be ef~ective. Said narrow clearance is maintained by the ~luid between the seal faces which prevents generation of friction heat and wear, but the narrow clearance limits out~low of the sealed fluid from the space 13 j`
into the region 14.
Referring now to Figure 2, there is illustrated the -^~
sealing face 18 of the sealing ring 19, which face has a groove arrangement 31 ~ormed therein. This groove arrangement 31 includes a first groove pattern 3~ which is positioned primarily on the radially outer portion of the ~ace 18. This groove pattern 32 normally provides both hydrodynamic and hydrostatic ~orce in the seal~
interface 27, although it is the pr.mary source for `
generatin~ hydrodynamic force and hence will herein often be referred to as the hydrodynamic region. The ~ ~
1 30 groove arrangement 31 also includes,a second groove i ~ i } ,i.. `
patt~r~ 33 whis_h is disposed generally radially inwardly of the groo~e pattern 32 and is positioned genierally within the center radia} re~ion of ~he face 18, that is ;i' ;~
the region which is spaced radial~y from both of the ~- ;
interface diameters 28 an~ 290 This latter groove pattern or region 33 functions primarily to provide a hydrostatic force betwesen the opposed seal faces 17-18 ,, .
,. .
WO9~/06212 PCT~$9310828g ,~
21~39 - 12 - ' ~
at conditions of near zero rotational speeds. The groove patterns 32 and 33 may be formed in the face 18 using conventional fabrications techniques~ ¦
Considering first the hydrodynamic groove pattern 32, it is de~ined by a plurality o~ angled grooves 34 which are formed in the face 18 in substantially uniformly angularly spaced relationship therearound.
These grooves 34 are all angled such that they open radially inwardly from the outer diameter 28 in such ~ashion that the grooves simultaneously project circumferentially and radially inwardly, and have an angl~.d relationship with respect to both the circumferential and radial directions of the seal face.
The angled groove 3~, as reprasented by the centerline 36 ~hereof where the groove intersects the outer diameter 28, normally opens inwardly of the outer diameter 28 at an acute angle relative to a tangent to the outer diameter, which acute angle may be in the neighborhood of 15 degrees.
Each angled groove 34 is defined by a pair of side or edge walls 37 and 38. The inner ends of grooves 34 -terminate generally at shoulders or abutments 39 which are generall~ rather abrupt and are defined about a radius designated R4 as generated about the c~nter point O of the face ring, this radius R4 defining the groove diameter for the groo~es 34 of the outer groove pattern , 32. The opposed sîde walls 37-38 defining each of the grooves 34 generally and preferabiy slightly converge 3 relati~e to one ano~her as the groove angl~s radially ! 30 inwardly. These side walls 37-38 may assume differen~ i configurations including straight lines, circular arcs or spiral profiles. When the sides 37-38 are de~ined as ~ -circular arcs or spirals, then the side wall 37 is of a convex configuration/ and the opposed wall 38 is of a concave configuration.
In the illustrated and preferred embodiment~ the opposed sides 37 38 are of circular configuration, but :`~
1,````~WO~S/06~12 2 1 4 7 7 3 9 PC~ iS~3/0828~ ~
13 - :
~re preferably generated about different radii having :~
different centerpoints. ~:
For example, and referring to Figure 3, the concave side 38' of groove 34' is generated about a radius :
designated R5 having a first centerpoint Cl, and the j-convex side 37 of the adjacent groove 34 is gerAerated about a radius R6 which is swung about the same centerpoint Cl, whereby the radius R6 exceeds the radius R5 by the perpendicular distance which separates the edges 37 and 38' of the adjacent pair of grooves 34 and 34'. This results in the flat or land 41 as defined ~.`
between the edges 37 and 38' being of constant transverse width as it angles radially inwardly toward :.:
the center of the ring.
In similar fashion, the concave edge 38 of groove'34 is also generated about the radius R5, which radius is . .`~
now generated about a second centerpoint C2 spaced from `,-the first centerpoint, and similarly the convex edge 37" 1-:
of the next groove 34" is generated about the radius R6 which is also swung about the second centerpoint C2, whereby the land 4l ~etween the edges 38 and 371~ again has a constant transverse dimension there~etween as this '````~!' land angles inwardly toward the center of the ring. The `:
two cer1terpoints themselves are located on a circle which is concentric about the center 0, and all of the grooves 34 are generated in a similar fashion.
Each of the grooves 34 is of substantial depth relative to the groove pattern 33, which depth is illustrated by the generally flat k?ottom wall 42 o~ the groove 34 as illustrated by Figure 4. The groove dep~h in a preferred embodiment as illustrated by solid line 42 is substantially uniform throughout the length of the groove 3 4 . i . :
However, the groove 34 can be of a tapered t' ' configuration throughout iti~ length so that the depth varies throughout the length, such being? ;:.
diagrammatically illustrated by the variations indicated '' ~VO~s/062l2 9 PCT~59Y08289 !~
by dotted lines designated at 42a and 42b in Figure 4.
As to the groove bottom wall designated at 42a, this groove has its maximum depth at the radially outer end, and its minimum depth at the radially inner end, although the depth at the radially inner end is still - sufficient so as to result in a significant shoulder or step 39 at the radially inner end thereof. Further, with this variation designated at 42a, the average depth of the groove substantially midway throughout the length thereof preferably substantially corresponds to the uniform depth o~ the groove as indicated by the bottom wall 42. In this tapered variation designated at 42a, the groove depth at the radially outer end is s~fficiently deep as to minimize the hydrodynamic force ef~ect. This latter effect is more pronounced adjacent the ~adial inner end of the groove 34 in the region of the face ring which is more centrally located, and is ~;
believed more effective for applying greater pressure against the central portion of the face ring so as to resist the typical thermal distortion (i~e. crowning) which occurs in operation. , As to the other tapered variation of the groove 34 as illustrated by the bottom 42b in Figure 4, in this variation the groove 34 is shallowest at its radially outer end and deepest at its radially inner end adjacent the shoulde~ or step 3Q. The shallowness vf the groove at the radially outer end i5 such as to effectively starve this region o~ the groove of fluid, and again minimizes the hydrodynamic effe~t in this region so that greater pressuxe is developed closer to the center of the face ring so as to tend to provide increased pressure resistance against the distortion of the ring which normally occurs during operation.
Considering now ~he hydrostatic groove region or pattern 33, this groove pattern is disposed generally radially inwardly of the hydrodynamic groove pattern 32 and is generally~of significantly shallower depth so as ~-.
~',; ""
WO9~106212 21 4 7 7 3 9 PCT~S93/08289 to prevent it from having any signi~icant ~ydrodynamic ¦
effect. This hydrostatic groove pattern 33 also I ;:
includes a plurality of angled grooves 44 which are formed in the oentral radial region of the seal face 18, with these grooves 44 being uniformly angularly disposed around the seal face. ~he grooves ~4 are contiguous with and project radially inwardly from the radially inner ends of the angled grooves 34, with grooves 44 being angled such that they simultaneously project circumferentially and radially inwardly from the diameter which defines the steps 39. The grooves 44 !' thus have an angled relationship with respect to both `.
the circumferential and radial directions of the seal face 18. .
The angled grooves 44, in the embodiment illustrated by Figures 2 and 3, are angled in the reverse circumferential direction from the outer grooves 34, i~
whereby a centerline 46 of the groove 44 intersects a radial line 45 at an acute angle ~ which, in the embodiment illustrated by Figure 3, is about 45.
Each groove 44 is defined between opposed edge or ~ `:
side walls 47-48, with the radially inner ends of grooves 44 terminat.ing at abrupt shoulders or abutments 49, the latter being defined generally on a radius R3 generated about the centerpoint 0, this latter radius defining the inner groove diameter.
The side walls of adjacent grooves, such as the -adjacent side walls 48 and 47", de~ine therebet~een a flat land Sl which is an extension of the flat land 41 defi~ed between the adjacent grooves 34 and 34~O This land 51 projects radially inwardly and connects to a further annular flat land 53, the latter being defined between the inner Xac~ diameter 29 (i.e., radiu~ R2) and , .
the radius R3. This land 53 is free of grooves and ~.
functions as a dam to significantly restrict flow of ~ ~
sealing fluid thereacross into the low pressure region ` ! .
d~fined at the diameter 29.
;'.'`
WO95/062l2 PCT/U593/0828~
2~ 4r1 rl 3 g - 1~ - ' ~` '";`':'' The inclined orientation ~i.e. angle ~) of the ¦ ' grooves 44 relative to the radial direction 45 is selected so that the grooves have a significant radially-directed flow component and hence these grooves 4~ have a less steep angle relative to radial direction ,~ ' 45 than do the grooves 34. More spQcifically, the inclination angle ~ is prefe~ably selected so as to be within the range of about + 45 relative to the radial direction 45. This maximizes the area of the land 51 as meas,ured transversely between the side walls (for -example the side walls 48 and 47") of adjacent grooves, thereby permitting creation of a more effectivP land 51 for trapping pressure fluid therebetween so as to create :~
a thrust bearing effect at times of operation at relati~ely high speeds of rotation. That is, a squeeze film effect is created at the lands 51 which is effective for resisting changes in gap width due to high ''.
speed vibrations or oscillations. In fact, in the preferred embodiment, the directly adjacent sides of adjacent grooves 44 and 44", such as the sides 48 and ~'~
47", preferably extend in parallel relationship to one another. Similarly, the adjacent sides 47 and 48' of the next adjacent pair of grooves 44 and 44' also '`' p~eferably extend in parallel relationship with one `:
another. This necessarily results in the opposed sides -:.
47-48 of each groove being of a slightly converging relationship as they project radially inwardly, and .
results in the txansverse width of the land 51 between , .:
each adjacent pair of grooves 44 being substantially 3iO ¢onstant anq hence of maximum width as the la~d project '.
radially inwardly, and maximizes the width of land 51 at the mouth therec)~ where the land meets the groove diameter def ined by the radius R3 .
The pair of side walls 47-48 which cooperate to de~ine each groove 44 ~ay be straight for manufa~turing ~, con~enience, or may ~e generated with spiral or circular j `.:
profiles, which circular profiles will preferably be ~: WO~5/06212 21 ~ 7 7 3 9 PCT~593/08289 - 17 - [
generated in a manner similar to the circular profiles of the side walls 37-38 for the grooves 34 as explained above.
As to the depth of the grooves 34 and 44, the l :.
grooves 34 have a depth which is several times greater - than th~ depth of grooves 44 and which is pref~rably in ~ `~
the range of about five to about ten times the depth of the grooves 44. More specifically, the deep grooves 34 will normally have an average depth of from about .Oool .~`
inch to about .001 inch althou~h a more practical maximum depth is believed to be about .0005 inch with a depth of from about .OUOl inch to about .0003 inch being preferred, and the shallow grooves 44 will norm~lly have ~`
a depth of from about .00001 inch to about .00008 inch ~`
with a depth of about .00002 inch to about .00005 inch -~
being preferred.
As to the radial positional relationships between the deep grooves 34, the shallow grooves 44 and the land 53, these relationships are determined relative to the radial width ~R of the seal interface 27 as measured batween the high pressure radius ~8 tradius Rl) and the low pressure radius 29 (radius R2). The hydrodynamic ;:
groove pattern 32 will normally occupy about the ..
radially outer one-third of the radial dimension oR, the hydrostatic groove pattern 33 will normally occupy about the middle one-~hird of the radial distance ~R, and the `~:
dam 53 will normally occupy about the radially inner `~
one-third of the distance ~R. However, the shallow groove pattern 33 can be either radially narrowed or ` .
widened as desired so that it will occupy anywhere from the middle one-quarter to a~out the middle one-half of the width ~R so as to maximize the fluid pressure~ in ~;:
this central region of the face ring so as to provide , incr~ased resistance against the conventional distortion and crowning which normally occurs in operation, such as due to thermal~effects.
',~
wo~739 rc~/uss3los2s~
In operation, the high pressure fluid surrounding the outer diameter 28 enters into the deep grooves 34 and the shallow grooves 44, but is then restricted from further radial inward flow by the land or dam 53. T~is . i pressure fluid within the grooves creates sufficient hydrostatic pressure to effect.. significant unloading of '.
force or a small separation between the opposed seal faces 17-18 throughout the interfaGe area 27, there thus :~
being created a hydrostatic force between the opposed seal faces. A small but controlled amount of the sealing fluid will pass over the dam 53 to the low pressure side 29 of the seal. The presence of this ~, hydrostatic force, however, greatly minimizes frictional contact between the opposed sealing faces, and greatly facilitates start-up of the seal both by reducing the stresses imposed on seal structural elements that transmit the seal face friction to the seal hausing 11 :
or shaft 12, and by significantly reducing or eliminating direct frictional contact between the opposed relatively rotatable seal faces 17-18 as ~:
rotation is initiat~d.
As the seal arrangement operates at higher rota~ional speed, the high pressure fluid enters the deep grooves 34, and is effectively pumped out over the shallow groove region 33 and the lands 41 to create and increase the d~m~nsion of the gap or clearance between the opposed faces 17-18 so as to permit relative hiqh sp~ed rotation between the faces while ef~ectively . ~.
avoiding or greatly minimizing any direct friction~
~30 contact therebetween. The fluid pressure profile (i.,e.
hydrodynamic force) cxeated between these opposed f~ces under this later condition, however, is subject to its ¦
highe~ pressure in the vicinity of the steps 39 disposed circu~ferentially between the adjacent groove ~ :
regions 32 and 33. For this pressure fluid to escape to the lower pressure side 29 of the seal, it must first .-flow over the shallow groove region 33 which creates r~ Wo ~S/06:112 2 1 ~ 7 7 3 9 PCTlUS!)3iO8289 - 19 ~
significant flow resistance, and in addition must also flow across -the relatively wide dam or land 53. This .
significant radial extent as defined by the land 53 and the shallow groove region 33 severely impedes the escape of the sealing fluid to the low pressure side of the ~ ;
system, ~nd permits the development of a desir~ble ' ~`
hydrodynamic force while at the same time providing for controlled and accepta~le rates of sealing fluid leakage to the low pressure side. :-,Referencing IlOW Figure~ 5 and 6, there is illustrated a variation of the invention. In this variationt the seal face is again provided with a groove pattern which incorporates a radially outer series of -:~
angled deep grooves 34 contiguous with a radially intermediat~ series of angled shallow grooves 44 constructed and positioned in a manner subs~antially .
identical to that illustrated by Figures 3 and 4. In this variation of Figures 5-6, however, the hydrostatic groove pattern 33' additionally includes a shallow annular groove 61 formed in the seal face 18 in `;
concentric re}ationship to the centerpoint 0, which annu}ar groove 61 is formed at and continuously connects the radially inner ends of the shallow grooves 44. This annular groove 61 has an inner annular wall 62 which effectively defines the radius ~3 which i5 the radially inner groove diameter, whereby the non-grooYed land 53 projects`radially lnwardly from this boundary wall 62O
The groove 61 i5 generally of uniform depth ~.
circumferentially throughout, which depth preferably ~i :
! 30 substantially identically corresponds to the depth of the shallow grooves 44.
The groo~e 61 is prefera~ly of rather narrow radial ~n.
width, which radial width as defined between the radially inner boundary wall 62 and the radially outer boundary wall 63 will typically ~e in the neighborhood of about 1/16 inch or less.
1```
WO95/06212 PCT~S93/08289 i ~ ` .
4When the hydrostatic groove pattern 33' includes therein the shallow annular groove 61 as shown in Figures 5-6, this effectively equalizes pressures ~ !-circumferentially in the vicinity of annular groove 61.
Thus, the fluid film created between adjacent grooves in the presence of the lands 4~1`and ~1 can be mai~tain2d at a substantially uniform magnitude circumferentially.
Since the pressure fluid occupies not only the grooves 34, 44 but also the annular groove 61, this minimizes distortions of both sealing faces in circumferential directions and permits therefore smaIler hydrostatic face separation with smaller leakage while avoiding or minimizing face contact when at or near ~ero rotational speed. ~ :
` Since the high pressure fluid exists continuously throughout the annular groove 61 in a hydrostatic condition, the pressure drop o~ the fluid a~ it escape~
radially across the land 53 to the low pressure side 2 creates uniform pressure gradients which extend ~0 circumferentially of the s~al ring, thereby also minimizing distortion circumferentially of the seal ring in the area of the land 53, and hence minimizing the tendency of the seal ring to deform into a wavy ..
circumferentially-extending configuration. However, under a hydrodynamic condition, the entire shallow groove region 33 effectively acts as an extension of the land 53 to provide for ~ontrolled and minimal leakage of `, sealing fluid thereacross during operation near to or at ~;
full speed. A.' ' ' I While the grooves 44 illustrated by Figure 3 and,as described above are reversely angled relative to the grooves 34, the grooves 44 can also be angled in the same circumferential direction as the groove~ 34 a5 ' il}ustrated by Figure 7. In this latter variation, the ~ -inner grooves 44 stiIl pre~erably have the c~nterlines 46 thereof intersecting the radial direction 45 at an angle ~ which is preerably no greater than about 45, ,, ,; ; WO95/06212 21 4 7 7 3 9 PCT~S93/08289 ~
- 21 - ~ ;
with the inclination of the centerlines of grooves 44 ¦ -preferably being positioned so as to lie within the j ~
extremes illustrated ~y the positions of Figures 3 and I :
7. In the positional relationship wherein the grooves i `~
44 angle in the .came circumferential direction as the grooves ~4, such as illustrated by Fî~1re 7, the inne~
shallow grooves 44 will be angled radially inwardly more sharply than the outer deep grooves 34, where~y the side walls 37-38 where they ioin to the side walls 47-48 `~
effectively define a discontinuity in curvature. That is, the abutting side walls 37, 47 and 38, 48 do not .
define a continuous curvature or straight line, although any discontinuity can obviously be rounded to facilitate the merger of the side walls.
By providing the shallow grooves 44 with a greater radially-inwardly directed inclination, the grooves 44 pe~mit the formation of more effective ~and areas 51 therebetween so as to provide for an improved squeeze film effect during high speed rotation, and at the same ~:
time the retained circumferential angularity of the grooves 44 is believed to permit at least some minimal ~
hydrodynamic force generation in the gap between the .`~
opposed seal faces 17-18 when seal rotation occurs at `;:
low speed, such as during start-up, thereby improving the fluid seal in the central radial region between the seal faces l7-18 prior to the gap being widened due to `~
the full efectiveness of the hydrodynamic force generated by the outer grooves 34 r which latter grooves ~
become fully effective at higher speeds. ¦
~ I More specifica~ly, by providing the shallow srooves 44 and deep grooves 34 with diffexent angularities as discussed above, this enables the wid~h ~ o~ the land f 51, as measured perpendicularly between the side edges of adjacent grooves 44, to be maximized, and made i~
greater than the width B of the land 4l as measured perpendicularly between the side~ edges of the adjacent ' -grooves 34. The angularities are preferably selected so ~ .
WO9~/06212 PCT~S93/08289 ~
~ 4rl ~ 3 9 -- 22 - that the land width A is equal to or greater than about l.3 times the land width B.
F'ur~her, by inclining the shallow grooves 44 in the reverse circumferential direction as indicated ~y the emhodiment of Figure 3, such is believed to provide some }
hydrodynamic force generation in the small g2p ~etween the opposed seal faces when;low speed reverse rotation occurs, such as when accidental back pressure upon shut down causes reversal of rotation in a compressor llpon shut down. Since such reversal of rotation in most use applications occurs only for a relatively short time and normally involves only lower rotational speeds, the re.verse angled arientation of the shallow inner groo~es 44 is believed to provide generation of at least mi.nimal hydrodynamic force to prevent or at least minimize any signi~icant direct contact between the opposed seal ~aces.
This Figure 7 variation is also preferably provided with the annular groove 6l in the same manner as shown in Figures 5~6.
While Figures 2 and 7 illustrate the shallow grooves 44 respectively reversely and forwardly angled relati~e to the deep grooves 34, reference is now made to Figures ::
8-lO wherein there is illustrated the preferred variation o~ the shallow grooves. In this variation (see Figures 8 and 9), the shallow grooves 44 project directly radially inwardly from the radially inner ends o~ the an~led grooves 34, with each groove 44 being position~d such that it has a substantially straight ! 30 centerline~46 which extends lengthwise of~the groove and Which projects inwardly in intersecting relationship to `the c~nterpoint O so as to consti~ute a radial lina.
E~ h groove 44 is also defined between opposed edge or side walls 47-48, both of which ar~ preferably straight.
The side walls of the adjacent grooves ~4, such as the . .~
adjacent side walls 4~ and 47" (Figure 9), define ' :::
therebetweeIl the flat land 51 which is an extension of :~
~.,.
,~ WO95/06212 21 4 7 7 3 9 PCT~S93/08~89 , the flat land 41 defined between the adjacent grooves 34 and 34". This land 51 projects radially inwardly and connects to the further annular flat land 53. The -~
radial orientation of the grooves 44 is highly desirable since this maximizes the area of the land 51 as measured transversely between the side walls (f~r ~xample the side walls 48 and 47") of adjacent grooves, thereby permitting creation of a more effective land (since the minimum transverse dimension across these lands is relatively large~ for trapping pressure fluid therebetween so as to create a thrust bearing effec~ at times of operation at relatively high speeds of rotation. That is, a squeeze film effect is created at ;~
the lands 51 which is effective for resisting changes in gap width due to high speed induced oscillations and vibrations. In fact, in this preferred variation of Fi~ures 8-10, th~ directly adjacent sides of adjacent ~ `-grooves 44 and 44", such as sides 48 and 47", preferably extend in parallel relationship to one anothex.
Similarly the adjacent sides 47 and 48' of the next adjacent pair of grooves 44 and 44' also preferably extend in paràllel relationship with one another. This necessarily results in the opposed sides 47-48 of` each groove b~ing of a slightly con~erging relationship as they project radially inwardly, and results in the transverse width of the land 51 between each adjacent pair of grooves 44 being substantially cons~ant as the ,-land 51 projects radially inwardly, and maxi~izes the width o~ land 51 at the mouth thereof, that is where the ~ ;
land meets the groove~diameter defined by the radius R2 This parallel relationship between the sides of adjacent grooves 4 4, and the radial direction o~ these grooves s 44, hen~e maximizes the area of land 51 to pro~ide signi~icantly improved squeeze film characteristics '-which ar~ particularly important at or near high speed ~ ;
rotation.
., WO95/~G2l2 PCl~593/08289 ~
21 4~ ~ 39 - 24 - . I , Referencing now Figures 11 and 12, there is illustrated a preferred variation of Fi.gures 8-lo 1 :
modified to include the shallow annular groove 61 for continuously connecting the radially inner ends of ~he shallow grooves 44. This annular groove 61 ~unctions iIl ,.
the same manner as described above relative to Figures 5 ~ ~ `
and 6. ~
While the invention illus~rated and described herein has the high pressure region located at the outer diameter, which is the most commonly encountered.use condition, it will be appreciated that the groove pattern can extend radially from an inner diameter if the latter is the high pressure region.
Although a particular preferred embodiment of the invention has been disclosed in detail for illustrative purposes, it will be recognized that variations or modifications of the disclosed apparatus, including the rearrangement of parts, lie within the scope of the present invention.
';' ;:
r,:';' ~ ' ' '~
., .:.' ~.,''''' ', "
~',``~',''.'.
The face area where these two sealing faces defin~ a sealing clearance is called the sealing interface.
The abover-mentioned spiral groove pattern on one of . ~
the sealing faces normally extends inward from the outer ~`-circumference and ends at a partisular face dia~eter called the groove diameter, which is larger then the inner diameter of the seal interface. The non-grooved , area between the groove diameter and the inner interface diameter serves as a restriction to fluid outflow.
Fluid deli~ered by the spiral pa~tern must pass through `-this restriction and it can do so only if the sealing faces separate. The way this works is through pressure build-up. Should the faces remain in contact/ fluid will be compressed just ahead of the restriction, thus building up pressure. The pressure causes separation ';
force which eventually becomes larger than the forces that hold the faces together. In that moment the sealing faces separate and allow the fiuid to escape.
During operation of the,sea}, an equilibrium establishes itself between fluid inflow through spiral pumping and fluid outflow through face sepa-ation. Face separation is therefore pre~ent as long as the seal is operating, which means as long as one face i5 rotating in relation to the opposite face. ~-However, spiral pumping is not the only f actor that determines the amount of the separation ~etween the ~ealin~ faces~ Just as the spirals are able to dri~
the fluid into the non-groove portion of the sealing ; 30 'inter~ace pas~ the groove diameter, so can the pressure diff~rential. If enough of a pressure difference exists betwean the groo~ed end of the interface and the non-grooved end, fluid will also be forced into the non- , grooved portion of the interface, thereby separating the faces and ~orming the clearance.
Both ways in which clearan e can be formed between the sealing faces, one with speed of rotation~ the other f - wo 95/OG212 2 1 4 7 7 3 9 PCT~593/08289 with pressure differential, are distinct and separate, even though the effects of both combine on the operating seal. If there is no pressure di~ference and the seal ~ace separation occurs strictly due to face rotation, forces due to fluid flow are known as hydrodynamic forces if the fluid sealed is a liquid, and a~rodynamic ' ~orces if the fluid sealed is a gas.
On the other hand, if there is no mutual rotation between the two sealing faces and face separation is strictly the consequence of pressure differential between both ends~of the sealing interface, forces due to fluid flow are called hydrostatic forces if the fluid sealed is a liquid, and aerostatics forces if the fluid sealed is a gas. In the following, the terms hydrostatic a~d hydrodynamic are used for both liquid and gas effects since these latter terms are more c~nventionally used when describing both liquid and gas seals.
A typical spiral qroove seal needs to provide acceptable performance in terms of leakag~ and the absence of ~ace contact during all regimes of seal op~ration. It must do so not only at top speed and p~ssure, but also at standstill, at start-up, acceleration, at periods of equipment warm-up or at shutdown. At normal operating conditions, pressure and speed vary constantly, which results in continuous adjustments to the running clearance. These adjustments ~-are automatic; one of the key properties of spiral :~
gro~ve seals is their self-adjustment capability. On change in speed or pressure, the face clearance adjusts automatically to a new set of conditions. Hydrostatic and hydrodynamic forces c~use this adjustm~nt.
The operating envelope:of speeds and pressures i~
usually very wide and a seal design of necessity must be a compromise. For its psrformance to be acceptable at : near-zero speed or pressure, it is less than optimum at ~ operating speed and pressure. ~This is simply due to the ~`
WO~5106212 pcT~ss3lo8289~
2 ~ 3 9 fact that, both in terms of pressure and speed, the seal has to be brought up to operating conditions from zero ~ `
speed and ~ero pressure differential.
Especially critical to seal operation is the start- .
up. If the seal is applied to a centrifugal gas compressor, the full suction pressure diff2rential is often imposed onto the seal before the shaft starts turning. This presents a danger in that the sealing faces will lock together with friction~ Face lock results when the hydrostatic force is insufficient-to counter pressure forces that maintain the seal faces in contact. Face lock can lead to seal destruction, in which excessive break-away friction between c~ntacting seal faces can cause heavy wear or breakage of internal seal components.
First t~1en, spiral grooYes must be able to separate the sealing faces hydrodynamically ~or full speed non-contacting operation. This normally requires fairly short and relatively deep spiral grooves. Second, the spiral groo~es must be able to unload the sealing faces hydrostatically for start/stops to prevent face lock.
For this, the grooves have to be extended in length.
The extended grooves in turn cause more separation and leakage during full speed operation. The full speed leakage o~ a typical 3.75 inch shaft seal with short and relatively deep spirals may be about .9 SCFM (i.e.
Standard Cubic Feet per Minute) at l,000 psig and lOrO00 rpm. However, full speed leakage for such a seal with extended grooves may reach 2.4 SCFM un~er the same ~30 `c~nditions, almost triple the previous vaiue. The constant burden of larger-than-necessary leakage represent significant operating costs and is highly . D~""~
undesirable.
Spiral groove design prac~ice goes back to US Patent No. 3,lOg,658 wherein two opposing spiral grooves pump oil against each other to develop a liquid baxrier capable of sealing a gas. Such an arrangement is ~ WO~5/06212 21 4 7 7 3 9 PCT1593/08289 limited in pressure as well as speed capability, as is inherent in the use of liquid forces to seal gas.
Another known arrangement is shown in US Patent No. 3,499,653. This interface design with partial spiral grooves relies heavily on hydrostatic effects.
The interface gap is designed with a tapered shape which is narrower at the non-grooved end and wider at the spiral grooYes. The effect of the spiral grooves and therefore the hydrodynamic forces are suppressed since spiral groove pumping becomes less effecti~e across ~he wider gaps. This likewise affects the stability of the seal and limits its top pressure and speed capability.
A further known arrangement is shown by US Patent ~o. 4,212,475. Here the spiral groove itself attempts to act both as a hydrostatic as well as a hydrodynami~
pattern and is used to eliminate ths need for the tap~red shape o~ the gap so that a considerable degree of spiral groove hydrodynamic force can be applied to impart a self-aligning property to the sealing ~;~
interface. The self-aligning property forces the sealing interface back towards a parallel po~ition, regardless of whether deviations from parallel position dur~ng seal operation occur in radial or tangential directions. This resulted in improvement stability and ;~
increased performance limits in terms o~ pressure and speed~
While the known fluid seals as briefly summarized above have attempted to provide both hydrodynamic~and , i:
hydrostatic sealing properties, nevertheless the known 3iO seals have;bjeen deficient with respect to their ability to optimize the combination of these hydrostatic and hydrodynamic properties so as to provide desirable hydrostatic properties which facilitate starting and stopping of seals while effectively minimizing or a~oiding direct face contact and minimizing face loading between the seals so that the assembly can be start2d up -~' with minimal friction to avoid severe frictional power ~, ',~
W0~5/Ofi~l~ PCT/US~3/08289 ~ ~ ~
~ 39 - 6 - l`
!
re~uirements and direct frictional wear between the faces, and at the same time provide desirable h~drodynamic properties between the relatively-rotatable seal faces under a wide range of operating co~ditions~ ;
particularly those involving high speed and high 1-~
pressure.
Accordingly, it is an object of this invention to provide an improved fluid seal of the type employing a grooved pattern on one of the opposed seal faces, which improved seal provides a more optimi~ed combination of hydrodynamic and hydrostatic sealing characteristics SQ ..
as to permit improved seal performance under a significantly greater range of operating conditions, including operating conditions ranging from start up to conditions involving high speed and high pressure. :
In the improved seal arrangement Qf the present invention, the groove pattern (which is typically defined on only one of the seal ~aces) includes first and ~econd groove arrangements which co~municate with one another, one being significantly deeper than the .;
other, whereby the deeper arrangement is particulaxly effective for providing the desired hydrodyna~ic characteristics, whereas the shallower groove arrangement is more effective for providing the desired ~:
hydrostatic charazteristics. At the same time, these arrangements are positloned such that the shallower .
arr~ngement is interposed generall~ between the deeper groove arrangement and a non-grooYed annular land or dam which effectively separates the groove pattern from the ! 30 lo~ pressure side of the seal, whereby desira~le hydrostatic and hydrodynamic seal properties can both be ~ ~-obtained but at the sam~ time leakage of sealing fluid (for example, a gas~ across the dam to the low pres~ure side is minimized so as to impro~e the performance efficiency of the ~eal.
In addition, this optimization o~ the seal propertîes and performance characteristics is further 21477~9 WO9~/06212 PCT~JS~3/08289 improved by optimizing the groove pattern or configuration relative to the surrounding lands defined ¦ :
on the seal face so that the fluid film which is created between the opposed seal faces provides a more uniform `
pressure distribution and sealing characteristics while minimizing distortion of the seal faoe, ~hich in turn assists in optimizing`the seal performance with minimum width of gap between the opposed seal faces while still ~:
avoiding or minimizing direct contact and frictional ` ;
wear between the opposed seal faces. ;.
In the improved seal of this invention,~ as briefly :~
discussed above, the groove pattern includes the deep ;~`
groove arrangement which i5 defined:by:a circumfer- .
entially arranged series of grooves which angle circum~
ferentially and radially inwardly from the surroundin;g :
hlgh-pressure side of the seal, which angled~grooves may be o~ spiral, circular or straight configuration. These angled grooves are relatively deep and project only ..
partway across the seal:face. The angled deep grooves, at their radially inner:ends, communicate with the ii shallow groove arrangemen~ which is positioned radially inwardly of the deeper groove arrangement,:but which is separate~ from the low pressure side of the seal by:the intermediate non-grooved annular land;or dam.~ This ~
:shallow groove~arrangement has a depth which is a small` :;.
fraotion of the deeper grooYe arrangement and is ~ e~f~ctive for: creati~g:a hydrostatic force:between the :: :opposed sealing~faces substantially in the central : : :~
region thereof as defined between the ra~ially outer and ~ :;
;l3 . inner~boundarieslof the seal interface~ a! preferred ~! ' `'.. ~`-embodiment, all grooves associated with the groove pattern are form~d: ;such ~hat the sides of a~jacent~
~:~ grooves ex~end generally in parallelism~with one another so that the intermediate land area:between~adjacent grooves~maintains a su~stantially~constant width, even adjacent thé radially;~:~inner ends~cf~the~grooves, to :
~àximize squeeze film effects~;~in the~;fluid whi h f1Ows W095/06212 PC~S~3/0828~
2~ 39 - 8 - 1 - over these lands and thus enhance the thrust bearing support these lands provide for avoidance of seal face contact at or near the full speed rotation. ¦ ;
The improved seal arrangement, as aforesaid, also . , preferably forms the shallow groove arrange~ent by a circumfere.ntially-spaced s2ries cf shall~w groove~ which are contiguous with and project radially inwardly from the inner ends of the angled deep grooves, which shallow grooves terminate at the dam. These shallow grooves pro~ide improved hydrostati~ seal characteristics in the central seal face region, and angle radially inwardly at a smaller angle (which angle is zero in a preferred ~mbodiment) re~ative to the radial direction than do the deep grooves so as to increase the land area between the adjacent shallow grooves, particularly adjacent the radially inner ends of the shallow grooves, to provide a ~etter fluid sgueeze film effect between the opposed seal faces during high speed rotation. -~
Further improvement to the hydrostatically effective relatively shallow inner groove pattern is aimed at reduction and elimination o~ any seal face distortions that might occur as a result of circum~erential non- ;
uniformity of hydrostatic pressure fields as these form above groove and land regions of the shallow groove patt~rn at conditions at or near to the zero speed of rotation. This improvement is a narrow and shallow circumferential groove interconnecting inner ends of the shallow inner gxoove pattern. Such a shallow circumferential groove acts to equali2e pressure field ' -nonl-uniformities circumferentially, as a result suppressing any ~ace distortions and producing a uniform face separation with no or only minimal face-to face ~;-contact e~en at extremely low ma~itudes of separation between the f~ces.
Other objects and purposes of the invention will be apparent to persons familiar with seals of this general :
'" '', `
,~ W095/062l2 21 4 7 7 3 9 PCT~S93/08289 _ g _ i .
txpe upon reading the following specific~tion and inspecting the accompanying drawings. i~
BRIEF DESCRIPTION OF THE DRAWINGS 1~
Figure 1 is a f~agmentary central sectional view illustrating a generally conventional fluid face seal arrangement, such as a grooved face ~e~l, associated with a rotating shaft.
Figure 2 is a view taken generally along line 2-2 in Figure 1 and illustrating the groove pattern associated with a face of the rotating seal ring according to an e~odiment of this invention.
Figure 3 is a fragmentary enlargement of a part of Figure 2 so as to illustrate the groove~pattern in ~-greater detail.
Figure 4 i5 a fragmentary sectional vi~w taken ~ubstantially along line 4~4 in Figure 3.
Figures 5 and 6 are views which correspond r ,', respectively to Figures 3 and 4 but illustrate a variation thereof. i~
~0 Figure 7 is a view similar to Figure 2 but showing a further variation of the inner groove pattern.
Figure 8 is a ~iew similar to Figure 2 but showing still a further and prsferred variation of the inner groove pattern.
Figure 9 is a fragmentary enlargement of a part of Figure 8 so as to illustrate the groove pattern in greatex detail. ~-;
Figure lO is a fragmentary sectional view taken ~-substantially along line lO-lO in Figure 9. ~ ;;
~;30 l ! ~igure~ 11 and 12 are views which correspond j ! j respectively to Figures 9 and lO but illustrate a variation thereof.
; Cert~in terminoIogy will be used in the following ~`
description for convènienca in reference onlyj and will ~ ~not be limiting. For exampla, the words~'lupwardly", j ~
; ~ "downwardly", "rightwardly"~and "leftwardly" will refer ''.. .' ' :~ ~ '.;i' W~g~/062l2 PCT~593/08289 ~` ~
~4,~39 -lo- ~`:
to direction~ in the drawings to which reference is made. The words "inwardly" and "outwardly" will re~er to directions toward and away from, respectively, the geometric center of the assembly and designated parts thereof. Said terminolo~y will include the words ~~ -specifically mentianed, derivatives ~hereof, and words of similar import.
DETAILED DESCRIPTION
Referring to Figure 1, therè is shown a typical grooved face seal assembly 10 and its environment. This environmenl comprisas a housing ll and a rotatable shaft 12 extending through said housing. The seial as~embly lO
is applied to seal a fluid (such as a pressurized ga~) within the annular s~ace 13 and to restrict its escape into the environment at 14. Basic components of the seal assembly incl~des an annular, axially movable but non-rotatable sealing ring 16 having a radially extending flat face 17 in opposed sealing relationship with a radially extending ~lat face 18 of an annular ~`
rotatable sealing ring 19 which is non-rotatably mounted on the shaft 12. ~ing 1~ normally rotates in the direction of the arrow ~Figure 2). The sealing ring 16 is located wi~hin cavity 21 of housing 11 and held sub-qtantially concentric to rotatable sealing ring 19.
Between housing 11 and the sealing ring 16 is a conventional anti rotation de~ice ~not shown) for ;
preventing rotation of ring 1~, as well a9 a plurality of springs 22 spaced equidistantly around the cavity 21. ! `~
Springs 22 urge the sealing ring 16 toward engagement with the sPalin~ ring 19~ An 0-rin~ 23 seals tha space between the sealing ring 16 and the housing 11. The -sealing ring 23 is retain~id in the axial position by a ~-sleeve 24 which is c~ncentric with and locked on the shaft 12, such as by locknut 25 threaded on shaft 12 as shown. 0-ring seal 26 precludis leakage between the ~ ;
sealing ring lg and the sha~t 12.
~ woss/o62l2 21 4 7 7 3 9 PCT~593l08289 ¦ ~
The radially extending face 18 of the sealing ring 19 and radiall~ extending ~ace 17 of sealing ring 16 are in sealing relationship~ and define an annular contact area 27 therebetween, this being the seal interface.
This seal interface 27 is defined by a surrounding outer ~`
diameter 28 of ring 19 and an inner di^..me~er 29 of ring 16, these being the diameters exposed to the high and -low pressure fluid respectively in the illustrated ~:
embodiment. In opèration, a very narrow clearance is ' maintained between the seal faces 17-18, due ~o a fluid ;.
~ilm as generated by a groove pattern (as described --below) formed in the sealing face 18 of the sealing ring `~
19. Alternately, the groove pattern San be formed in the sealing fae 17 of the sealing ring 16 and stiIl be ef~ective. Said narrow clearance is maintained by the ~luid between the seal faces which prevents generation of friction heat and wear, but the narrow clearance limits out~low of the sealed fluid from the space 13 j`
into the region 14.
Referring now to Figure 2, there is illustrated the -^~
sealing face 18 of the sealing ring 19, which face has a groove arrangement 31 ~ormed therein. This groove arrangement 31 includes a first groove pattern 3~ which is positioned primarily on the radially outer portion of the ~ace 18. This groove pattern 32 normally provides both hydrodynamic and hydrostatic ~orce in the seal~
interface 27, although it is the pr.mary source for `
generatin~ hydrodynamic force and hence will herein often be referred to as the hydrodynamic region. The ~ ~
1 30 groove arrangement 31 also includes,a second groove i ~ i } ,i.. `
patt~r~ 33 whis_h is disposed generally radially inwardly of the groo~e pattern 32 and is positioned genierally within the center radia} re~ion of ~he face 18, that is ;i' ;~
the region which is spaced radial~y from both of the ~- ;
interface diameters 28 an~ 290 This latter groove pattern or region 33 functions primarily to provide a hydrostatic force betwesen the opposed seal faces 17-18 ,, .
,. .
WO9~/06212 PCT~$9310828g ,~
21~39 - 12 - ' ~
at conditions of near zero rotational speeds. The groove patterns 32 and 33 may be formed in the face 18 using conventional fabrications techniques~ ¦
Considering first the hydrodynamic groove pattern 32, it is de~ined by a plurality o~ angled grooves 34 which are formed in the face 18 in substantially uniformly angularly spaced relationship therearound.
These grooves 34 are all angled such that they open radially inwardly from the outer diameter 28 in such ~ashion that the grooves simultaneously project circumferentially and radially inwardly, and have an angl~.d relationship with respect to both the circumferential and radial directions of the seal face.
The angled groove 3~, as reprasented by the centerline 36 ~hereof where the groove intersects the outer diameter 28, normally opens inwardly of the outer diameter 28 at an acute angle relative to a tangent to the outer diameter, which acute angle may be in the neighborhood of 15 degrees.
Each angled groove 34 is defined by a pair of side or edge walls 37 and 38. The inner ends of grooves 34 -terminate generally at shoulders or abutments 39 which are generall~ rather abrupt and are defined about a radius designated R4 as generated about the c~nter point O of the face ring, this radius R4 defining the groove diameter for the groo~es 34 of the outer groove pattern , 32. The opposed sîde walls 37-38 defining each of the grooves 34 generally and preferabiy slightly converge 3 relati~e to one ano~her as the groove angl~s radially ! 30 inwardly. These side walls 37-38 may assume differen~ i configurations including straight lines, circular arcs or spiral profiles. When the sides 37-38 are de~ined as ~ -circular arcs or spirals, then the side wall 37 is of a convex configuration/ and the opposed wall 38 is of a concave configuration.
In the illustrated and preferred embodiment~ the opposed sides 37 38 are of circular configuration, but :`~
1,````~WO~S/06~12 2 1 4 7 7 3 9 PC~ iS~3/0828~ ~
13 - :
~re preferably generated about different radii having :~
different centerpoints. ~:
For example, and referring to Figure 3, the concave side 38' of groove 34' is generated about a radius :
designated R5 having a first centerpoint Cl, and the j-convex side 37 of the adjacent groove 34 is gerAerated about a radius R6 which is swung about the same centerpoint Cl, whereby the radius R6 exceeds the radius R5 by the perpendicular distance which separates the edges 37 and 38' of the adjacent pair of grooves 34 and 34'. This results in the flat or land 41 as defined ~.`
between the edges 37 and 38' being of constant transverse width as it angles radially inwardly toward :.:
the center of the ring.
In similar fashion, the concave edge 38 of groove'34 is also generated about the radius R5, which radius is . .`~
now generated about a second centerpoint C2 spaced from `,-the first centerpoint, and similarly the convex edge 37" 1-:
of the next groove 34" is generated about the radius R6 which is also swung about the second centerpoint C2, whereby the land 4l ~etween the edges 38 and 371~ again has a constant transverse dimension there~etween as this '````~!' land angles inwardly toward the center of the ring. The `:
two cer1terpoints themselves are located on a circle which is concentric about the center 0, and all of the grooves 34 are generated in a similar fashion.
Each of the grooves 34 is of substantial depth relative to the groove pattern 33, which depth is illustrated by the generally flat k?ottom wall 42 o~ the groove 34 as illustrated by Figure 4. The groove dep~h in a preferred embodiment as illustrated by solid line 42 is substantially uniform throughout the length of the groove 3 4 . i . :
However, the groove 34 can be of a tapered t' ' configuration throughout iti~ length so that the depth varies throughout the length, such being? ;:.
diagrammatically illustrated by the variations indicated '' ~VO~s/062l2 9 PCT~59Y08289 !~
by dotted lines designated at 42a and 42b in Figure 4.
As to the groove bottom wall designated at 42a, this groove has its maximum depth at the radially outer end, and its minimum depth at the radially inner end, although the depth at the radially inner end is still - sufficient so as to result in a significant shoulder or step 39 at the radially inner end thereof. Further, with this variation designated at 42a, the average depth of the groove substantially midway throughout the length thereof preferably substantially corresponds to the uniform depth o~ the groove as indicated by the bottom wall 42. In this tapered variation designated at 42a, the groove depth at the radially outer end is s~fficiently deep as to minimize the hydrodynamic force ef~ect. This latter effect is more pronounced adjacent the ~adial inner end of the groove 34 in the region of the face ring which is more centrally located, and is ~;
believed more effective for applying greater pressure against the central portion of the face ring so as to resist the typical thermal distortion (i~e. crowning) which occurs in operation. , As to the other tapered variation of the groove 34 as illustrated by the bottom 42b in Figure 4, in this variation the groove 34 is shallowest at its radially outer end and deepest at its radially inner end adjacent the shoulde~ or step 3Q. The shallowness vf the groove at the radially outer end i5 such as to effectively starve this region o~ the groove of fluid, and again minimizes the hydrodynamic effe~t in this region so that greater pressuxe is developed closer to the center of the face ring so as to tend to provide increased pressure resistance against the distortion of the ring which normally occurs during operation.
Considering now ~he hydrostatic groove region or pattern 33, this groove pattern is disposed generally radially inwardly of the hydrodynamic groove pattern 32 and is generally~of significantly shallower depth so as ~-.
~',; ""
WO9~106212 21 4 7 7 3 9 PCT~S93/08289 to prevent it from having any signi~icant ~ydrodynamic ¦
effect. This hydrostatic groove pattern 33 also I ;:
includes a plurality of angled grooves 44 which are formed in the oentral radial region of the seal face 18, with these grooves 44 being uniformly angularly disposed around the seal face. ~he grooves ~4 are contiguous with and project radially inwardly from the radially inner ends of the angled grooves 34, with grooves 44 being angled such that they simultaneously project circumferentially and radially inwardly from the diameter which defines the steps 39. The grooves 44 !' thus have an angled relationship with respect to both `.
the circumferential and radial directions of the seal face 18. .
The angled grooves 44, in the embodiment illustrated by Figures 2 and 3, are angled in the reverse circumferential direction from the outer grooves 34, i~
whereby a centerline 46 of the groove 44 intersects a radial line 45 at an acute angle ~ which, in the embodiment illustrated by Figure 3, is about 45.
Each groove 44 is defined between opposed edge or ~ `:
side walls 47-48, with the radially inner ends of grooves 44 terminat.ing at abrupt shoulders or abutments 49, the latter being defined generally on a radius R3 generated about the centerpoint 0, this latter radius defining the inner groove diameter.
The side walls of adjacent grooves, such as the -adjacent side walls 48 and 47", de~ine therebet~een a flat land Sl which is an extension of the flat land 41 defi~ed between the adjacent grooves 34 and 34~O This land 51 projects radially inwardly and connects to a further annular flat land 53, the latter being defined between the inner Xac~ diameter 29 (i.e., radiu~ R2) and , .
the radius R3. This land 53 is free of grooves and ~.
functions as a dam to significantly restrict flow of ~ ~
sealing fluid thereacross into the low pressure region ` ! .
d~fined at the diameter 29.
;'.'`
WO95/062l2 PCT/U593/0828~
2~ 4r1 rl 3 g - 1~ - ' ~` '";`':'' The inclined orientation ~i.e. angle ~) of the ¦ ' grooves 44 relative to the radial direction 45 is selected so that the grooves have a significant radially-directed flow component and hence these grooves 4~ have a less steep angle relative to radial direction ,~ ' 45 than do the grooves 34. More spQcifically, the inclination angle ~ is prefe~ably selected so as to be within the range of about + 45 relative to the radial direction 45. This maximizes the area of the land 51 as meas,ured transversely between the side walls (for -example the side walls 48 and 47") of adjacent grooves, thereby permitting creation of a more effectivP land 51 for trapping pressure fluid therebetween so as to create :~
a thrust bearing effect at times of operation at relati~ely high speeds of rotation. That is, a squeeze film effect is created at the lands 51 which is effective for resisting changes in gap width due to high ''.
speed vibrations or oscillations. In fact, in the preferred embodiment, the directly adjacent sides of adjacent grooves 44 and 44", such as the sides 48 and ~'~
47", preferably extend in parallel relationship to one another. Similarly, the adjacent sides 47 and 48' of the next adjacent pair of grooves 44 and 44' also '`' p~eferably extend in parallel relationship with one `:
another. This necessarily results in the opposed sides -:.
47-48 of each groove being of a slightly converging relationship as they project radially inwardly, and .
results in the txansverse width of the land 51 between , .:
each adjacent pair of grooves 44 being substantially 3iO ¢onstant anq hence of maximum width as the la~d project '.
radially inwardly, and maximizes the width of land 51 at the mouth therec)~ where the land meets the groove diameter def ined by the radius R3 .
The pair of side walls 47-48 which cooperate to de~ine each groove 44 ~ay be straight for manufa~turing ~, con~enience, or may ~e generated with spiral or circular j `.:
profiles, which circular profiles will preferably be ~: WO~5/06212 21 ~ 7 7 3 9 PCT~593/08289 - 17 - [
generated in a manner similar to the circular profiles of the side walls 37-38 for the grooves 34 as explained above.
As to the depth of the grooves 34 and 44, the l :.
grooves 34 have a depth which is several times greater - than th~ depth of grooves 44 and which is pref~rably in ~ `~
the range of about five to about ten times the depth of the grooves 44. More specifically, the deep grooves 34 will normally have an average depth of from about .Oool .~`
inch to about .001 inch althou~h a more practical maximum depth is believed to be about .0005 inch with a depth of from about .OUOl inch to about .0003 inch being preferred, and the shallow grooves 44 will norm~lly have ~`
a depth of from about .00001 inch to about .00008 inch ~`
with a depth of about .00002 inch to about .00005 inch -~
being preferred.
As to the radial positional relationships between the deep grooves 34, the shallow grooves 44 and the land 53, these relationships are determined relative to the radial width ~R of the seal interface 27 as measured batween the high pressure radius ~8 tradius Rl) and the low pressure radius 29 (radius R2). The hydrodynamic ;:
groove pattern 32 will normally occupy about the ..
radially outer one-third of the radial dimension oR, the hydrostatic groove pattern 33 will normally occupy about the middle one-~hird of the radial distance ~R, and the `~:
dam 53 will normally occupy about the radially inner `~
one-third of the distance ~R. However, the shallow groove pattern 33 can be either radially narrowed or ` .
widened as desired so that it will occupy anywhere from the middle one-quarter to a~out the middle one-half of the width ~R so as to maximize the fluid pressure~ in ~;:
this central region of the face ring so as to provide , incr~ased resistance against the conventional distortion and crowning which normally occurs in operation, such as due to thermal~effects.
',~
wo~739 rc~/uss3los2s~
In operation, the high pressure fluid surrounding the outer diameter 28 enters into the deep grooves 34 and the shallow grooves 44, but is then restricted from further radial inward flow by the land or dam 53. T~is . i pressure fluid within the grooves creates sufficient hydrostatic pressure to effect.. significant unloading of '.
force or a small separation between the opposed seal faces 17-18 throughout the interfaGe area 27, there thus :~
being created a hydrostatic force between the opposed seal faces. A small but controlled amount of the sealing fluid will pass over the dam 53 to the low pressure side 29 of the seal. The presence of this ~, hydrostatic force, however, greatly minimizes frictional contact between the opposed sealing faces, and greatly facilitates start-up of the seal both by reducing the stresses imposed on seal structural elements that transmit the seal face friction to the seal hausing 11 :
or shaft 12, and by significantly reducing or eliminating direct frictional contact between the opposed relatively rotatable seal faces 17-18 as ~:
rotation is initiat~d.
As the seal arrangement operates at higher rota~ional speed, the high pressure fluid enters the deep grooves 34, and is effectively pumped out over the shallow groove region 33 and the lands 41 to create and increase the d~m~nsion of the gap or clearance between the opposed faces 17-18 so as to permit relative hiqh sp~ed rotation between the faces while ef~ectively . ~.
avoiding or greatly minimizing any direct friction~
~30 contact therebetween. The fluid pressure profile (i.,e.
hydrodynamic force) cxeated between these opposed f~ces under this later condition, however, is subject to its ¦
highe~ pressure in the vicinity of the steps 39 disposed circu~ferentially between the adjacent groove ~ :
regions 32 and 33. For this pressure fluid to escape to the lower pressure side 29 of the seal, it must first .-flow over the shallow groove region 33 which creates r~ Wo ~S/06:112 2 1 ~ 7 7 3 9 PCTlUS!)3iO8289 - 19 ~
significant flow resistance, and in addition must also flow across -the relatively wide dam or land 53. This .
significant radial extent as defined by the land 53 and the shallow groove region 33 severely impedes the escape of the sealing fluid to the low pressure side of the ~ ;
system, ~nd permits the development of a desir~ble ' ~`
hydrodynamic force while at the same time providing for controlled and accepta~le rates of sealing fluid leakage to the low pressure side. :-,Referencing IlOW Figure~ 5 and 6, there is illustrated a variation of the invention. In this variationt the seal face is again provided with a groove pattern which incorporates a radially outer series of -:~
angled deep grooves 34 contiguous with a radially intermediat~ series of angled shallow grooves 44 constructed and positioned in a manner subs~antially .
identical to that illustrated by Figures 3 and 4. In this variation of Figures 5-6, however, the hydrostatic groove pattern 33' additionally includes a shallow annular groove 61 formed in the seal face 18 in `;
concentric re}ationship to the centerpoint 0, which annu}ar groove 61 is formed at and continuously connects the radially inner ends of the shallow grooves 44. This annular groove 61 has an inner annular wall 62 which effectively defines the radius ~3 which i5 the radially inner groove diameter, whereby the non-grooYed land 53 projects`radially lnwardly from this boundary wall 62O
The groove 61 i5 generally of uniform depth ~.
circumferentially throughout, which depth preferably ~i :
! 30 substantially identically corresponds to the depth of the shallow grooves 44.
The groo~e 61 is prefera~ly of rather narrow radial ~n.
width, which radial width as defined between the radially inner boundary wall 62 and the radially outer boundary wall 63 will typically ~e in the neighborhood of about 1/16 inch or less.
1```
WO95/06212 PCT~S93/08289 i ~ ` .
4When the hydrostatic groove pattern 33' includes therein the shallow annular groove 61 as shown in Figures 5-6, this effectively equalizes pressures ~ !-circumferentially in the vicinity of annular groove 61.
Thus, the fluid film created between adjacent grooves in the presence of the lands 4~1`and ~1 can be mai~tain2d at a substantially uniform magnitude circumferentially.
Since the pressure fluid occupies not only the grooves 34, 44 but also the annular groove 61, this minimizes distortions of both sealing faces in circumferential directions and permits therefore smaIler hydrostatic face separation with smaller leakage while avoiding or minimizing face contact when at or near ~ero rotational speed. ~ :
` Since the high pressure fluid exists continuously throughout the annular groove 61 in a hydrostatic condition, the pressure drop o~ the fluid a~ it escape~
radially across the land 53 to the low pressure side 2 creates uniform pressure gradients which extend ~0 circumferentially of the s~al ring, thereby also minimizing distortion circumferentially of the seal ring in the area of the land 53, and hence minimizing the tendency of the seal ring to deform into a wavy ..
circumferentially-extending configuration. However, under a hydrodynamic condition, the entire shallow groove region 33 effectively acts as an extension of the land 53 to provide for ~ontrolled and minimal leakage of `, sealing fluid thereacross during operation near to or at ~;
full speed. A.' ' ' I While the grooves 44 illustrated by Figure 3 and,as described above are reversely angled relative to the grooves 34, the grooves 44 can also be angled in the same circumferential direction as the groove~ 34 a5 ' il}ustrated by Figure 7. In this latter variation, the ~ -inner grooves 44 stiIl pre~erably have the c~nterlines 46 thereof intersecting the radial direction 45 at an angle ~ which is preerably no greater than about 45, ,, ,; ; WO95/06212 21 4 7 7 3 9 PCT~S93/08289 ~
- 21 - ~ ;
with the inclination of the centerlines of grooves 44 ¦ -preferably being positioned so as to lie within the j ~
extremes illustrated ~y the positions of Figures 3 and I :
7. In the positional relationship wherein the grooves i `~
44 angle in the .came circumferential direction as the grooves ~4, such as illustrated by Fî~1re 7, the inne~
shallow grooves 44 will be angled radially inwardly more sharply than the outer deep grooves 34, where~y the side walls 37-38 where they ioin to the side walls 47-48 `~
effectively define a discontinuity in curvature. That is, the abutting side walls 37, 47 and 38, 48 do not .
define a continuous curvature or straight line, although any discontinuity can obviously be rounded to facilitate the merger of the side walls.
By providing the shallow grooves 44 with a greater radially-inwardly directed inclination, the grooves 44 pe~mit the formation of more effective ~and areas 51 therebetween so as to provide for an improved squeeze film effect during high speed rotation, and at the same ~:
time the retained circumferential angularity of the grooves 44 is believed to permit at least some minimal ~
hydrodynamic force generation in the gap between the .`~
opposed seal faces 17-18 when seal rotation occurs at `;:
low speed, such as during start-up, thereby improving the fluid seal in the central radial region between the seal faces l7-18 prior to the gap being widened due to `~
the full efectiveness of the hydrodynamic force generated by the outer grooves 34 r which latter grooves ~
become fully effective at higher speeds. ¦
~ I More specifica~ly, by providing the shallow srooves 44 and deep grooves 34 with diffexent angularities as discussed above, this enables the wid~h ~ o~ the land f 51, as measured perpendicularly between the side edges of adjacent grooves 44, to be maximized, and made i~
greater than the width B of the land 4l as measured perpendicularly between the side~ edges of the adjacent ' -grooves 34. The angularities are preferably selected so ~ .
WO9~/06212 PCT~S93/08289 ~
~ 4rl ~ 3 9 -- 22 - that the land width A is equal to or greater than about l.3 times the land width B.
F'ur~her, by inclining the shallow grooves 44 in the reverse circumferential direction as indicated ~y the emhodiment of Figure 3, such is believed to provide some }
hydrodynamic force generation in the small g2p ~etween the opposed seal faces when;low speed reverse rotation occurs, such as when accidental back pressure upon shut down causes reversal of rotation in a compressor llpon shut down. Since such reversal of rotation in most use applications occurs only for a relatively short time and normally involves only lower rotational speeds, the re.verse angled arientation of the shallow inner groo~es 44 is believed to provide generation of at least mi.nimal hydrodynamic force to prevent or at least minimize any signi~icant direct contact between the opposed seal ~aces.
This Figure 7 variation is also preferably provided with the annular groove 6l in the same manner as shown in Figures 5~6.
While Figures 2 and 7 illustrate the shallow grooves 44 respectively reversely and forwardly angled relati~e to the deep grooves 34, reference is now made to Figures ::
8-lO wherein there is illustrated the preferred variation o~ the shallow grooves. In this variation (see Figures 8 and 9), the shallow grooves 44 project directly radially inwardly from the radially inner ends o~ the an~led grooves 34, with each groove 44 being position~d such that it has a substantially straight ! 30 centerline~46 which extends lengthwise of~the groove and Which projects inwardly in intersecting relationship to `the c~nterpoint O so as to consti~ute a radial lina.
E~ h groove 44 is also defined between opposed edge or side walls 47-48, both of which ar~ preferably straight.
The side walls of the adjacent grooves ~4, such as the . .~
adjacent side walls 4~ and 47" (Figure 9), define ' :::
therebetweeIl the flat land 51 which is an extension of :~
~.,.
,~ WO95/06212 21 4 7 7 3 9 PCT~S93/08~89 , the flat land 41 defined between the adjacent grooves 34 and 34". This land 51 projects radially inwardly and connects to the further annular flat land 53. The -~
radial orientation of the grooves 44 is highly desirable since this maximizes the area of the land 51 as measured transversely between the side walls (f~r ~xample the side walls 48 and 47") of adjacent grooves, thereby permitting creation of a more effective land (since the minimum transverse dimension across these lands is relatively large~ for trapping pressure fluid therebetween so as to create a thrust bearing effec~ at times of operation at relatively high speeds of rotation. That is, a squeeze film effect is created at ;~
the lands 51 which is effective for resisting changes in gap width due to high speed induced oscillations and vibrations. In fact, in this preferred variation of Fi~ures 8-10, th~ directly adjacent sides of adjacent ~ `-grooves 44 and 44", such as sides 48 and 47", preferably extend in parallel relationship to one anothex.
Similarly the adjacent sides 47 and 48' of the next adjacent pair of grooves 44 and 44' also preferably extend in paràllel relationship with one another. This necessarily results in the opposed sides 47-48 of` each groove b~ing of a slightly con~erging relationship as they project radially inwardly, and results in the transverse width of the land 51 between each adjacent pair of grooves 44 being substantially cons~ant as the ,-land 51 projects radially inwardly, and maxi~izes the width o~ land 51 at the mouth thereof, that is where the ~ ;
land meets the groove~diameter defined by the radius R2 This parallel relationship between the sides of adjacent grooves 4 4, and the radial direction o~ these grooves s 44, hen~e maximizes the area of land 51 to pro~ide signi~icantly improved squeeze film characteristics '-which ar~ particularly important at or near high speed ~ ;
rotation.
., WO95/~G2l2 PCl~593/08289 ~
21 4~ ~ 39 - 24 - . I , Referencing now Figures 11 and 12, there is illustrated a preferred variation of Fi.gures 8-lo 1 :
modified to include the shallow annular groove 61 for continuously connecting the radially inner ends of ~he shallow grooves 44. This annular groove 61 ~unctions iIl ,.
the same manner as described above relative to Figures 5 ~ ~ `
and 6. ~
While the invention illus~rated and described herein has the high pressure region located at the outer diameter, which is the most commonly encountered.use condition, it will be appreciated that the groove pattern can extend radially from an inner diameter if the latter is the high pressure region.
Although a particular preferred embodiment of the invention has been disclosed in detail for illustrative purposes, it will be recognized that variations or modifications of the disclosed apparatus, including the rearrangement of parts, lie within the scope of the present invention.
';' ;:
r,:';' ~ ' ' '~
., .:.' ~.,''''' ', "
~',``~',''.'.
Claims (19)
The embodiments of the invention in which an exclusive property or privilege is claimed are defined as follows:
1. A fluid seal device cooperating between a housing and a rotatable shaft for creating a fluid seal between high and low pressure regions, said device comprising:
a first seal ring mounted on the shaft for rotation therewith and a second seal ring disposed adjacent the first seal ring and being non-rotatably mounted relative to the housing;
said first and second seal rings respectively defining thereon opposed first and second flat annular seal faces adapted to substantially axially abut to define an annular seal interface which extends radially between and is defined by radially outer and inner diameters which respectively communicate with said high and low pressure regions, one of said seal rings being axially movable and normally urged axially toward the other seal ring;
a groove pattern formed in one of said seal faces for causing a thin film of pressurized fluid to be interposed between said seal faces to create a small clearance therebetween;
said groove pattern including first groove means formed in said one seal face for creating a hydrodynamic fluid seal between the opposed seal faces when the first and second seal rings relatively rotate at high speed;
1 l said groove pattern including second groove means formed in said one seal face for creating a hydrostatic fluid seal between said opposed seal faces when said first and second seal rings are substantially stationary relative to one another;
said first groove means including a plurality of first grooves disposed in generally uniformly angularly spaced relationship around said one seal face, said first grooves being angled so as to project circumferentially and radially inwardly from said high pressure diameter;
said second groove means including a plurality of second grooves which are disposed in substantially uniformly angularly spaced relationship around said one seal face, each said second groove being contiguous with and projecting radially inwardly from a radially inner end of a respective one of said first grooves, said second grooves projecting radially inwardly from the radially inner ends of said first grooves, said second grooves being radially inwardly angled at a smaller angle relative to a radial direction than said first grooves, and said second grooves projecting radially inwardly so as to terminate at a groove diameter which is spaced radially outwardly of said low pressure diameter;
said one seal face defining thereon an annular non-grooved flat land extending radially between said low pressure diameter and said groove diameter; and said first grooves having an average longitudinally-extending depth which is several times greater than the depth of said second grooves, and the inner end of said first grooves defining abrupt damlike steps where said first grooves connect to said second grooves.
a first seal ring mounted on the shaft for rotation therewith and a second seal ring disposed adjacent the first seal ring and being non-rotatably mounted relative to the housing;
said first and second seal rings respectively defining thereon opposed first and second flat annular seal faces adapted to substantially axially abut to define an annular seal interface which extends radially between and is defined by radially outer and inner diameters which respectively communicate with said high and low pressure regions, one of said seal rings being axially movable and normally urged axially toward the other seal ring;
a groove pattern formed in one of said seal faces for causing a thin film of pressurized fluid to be interposed between said seal faces to create a small clearance therebetween;
said groove pattern including first groove means formed in said one seal face for creating a hydrodynamic fluid seal between the opposed seal faces when the first and second seal rings relatively rotate at high speed;
1 l said groove pattern including second groove means formed in said one seal face for creating a hydrostatic fluid seal between said opposed seal faces when said first and second seal rings are substantially stationary relative to one another;
said first groove means including a plurality of first grooves disposed in generally uniformly angularly spaced relationship around said one seal face, said first grooves being angled so as to project circumferentially and radially inwardly from said high pressure diameter;
said second groove means including a plurality of second grooves which are disposed in substantially uniformly angularly spaced relationship around said one seal face, each said second groove being contiguous with and projecting radially inwardly from a radially inner end of a respective one of said first grooves, said second grooves projecting radially inwardly from the radially inner ends of said first grooves, said second grooves being radially inwardly angled at a smaller angle relative to a radial direction than said first grooves, and said second grooves projecting radially inwardly so as to terminate at a groove diameter which is spaced radially outwardly of said low pressure diameter;
said one seal face defining thereon an annular non-grooved flat land extending radially between said low pressure diameter and said groove diameter; and said first grooves having an average longitudinally-extending depth which is several times greater than the depth of said second grooves, and the inner end of said first grooves defining abrupt damlike steps where said first grooves connect to said second grooves.
2. A seal device according to Claim 1, wherein said second grooves are reversely angled in the circumferential direction relative to said first grooves.
3. A seal device according to Claim 2, wherein said second groove has a longitudinally extending centerline which intersects a radially projecting line of said one seal face at an angle of no greater than about 45°.
4. A seal device according to Claim 3, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves.
5. A seal device according to Claim 1, said second groove having a longitudinally extending centerline which intersects a radially projecting line of said one seal face at an angle of no greater than about 45°.
6. A seal device according to Claim 5, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves.
7. A seal device according to Claim 1, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves.
8. A seal device according to Claim 7, wherein each said second groove is defined between first and second side edges which converge toward one another as said second groove projects radially inwardly, and wherein the first side edge of one said second groove and an adjacent second side edge of an adjacent said second groove extend in parallel relationship to define therebetween a flat land which is of constant width as it projects radially inwardly between adjacent said second grooves.
9. A seal device according to Claim 8, wherein each of said first grooves is defined between first and second side edges which converge relative to one another as they project circumferentially and radially inwardly from said high pressure diameter, and wherein the first side edge of one said first groove extends in parallelism with an adjacent said second side edge of an adjacent said first groove to define therebetween a flat land which is of constant width throughout the longitudinal extent of the first grooves.
10. A seal device according to Claim 1, wherein each said second groove is defined between first and second side edges which converge toward one another as said second groove projects radially inwardly, and wherein the first side edge of one said second groove and an adjacent second side edge of an adjacent said second groove extend in parallel relationship to define therebetween a flat land which is of constant width as it projects radially inwardly between adjacent said second grooves.
11. A seal device according to Claim 1, wherein each of said first grooves is defined between first and second side edges which converge relative to one another as they project circumferentially and radially inwardly from said high pressure diameter, and wherein the first side edge of one said first groove extends in parallelism with an adjacent said second side edge of an adjacent said first groove to define therebetween a flat land which is of constant width throughout the longitudinal extent of the first grooves r
12. A seal device according to Claim 1, wherein the sealing fluid is a gas.
13. A seal device according to Claim 1, wherein the first groove has an average depth in the range of about .0001 inch to about .0003 inch, and wherein the depth of said first groove is in the range of at least about five to about ten times the depth of the second groove.
14. A seal device according to Claim 13, wherein each of said first and second grooves are of uniform depth throughout substantially the respective longitudinal extent thereof.
15. A seal device according to Claim 13, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of:
said second grooves.
said second grooves.
16. A seal device according to Claim 1, wherein said angle is zero and said second grooves project solely radially inwardly from the radially inner ends of said first grooves.
17. A seal device according to Claim 16, wherein the first groove has an average depth in the range of about .0001 inch to about .0003 inch, and wherein the depth of said first groove is in the range of at least about five to about ten times the depth of the second groove.
18. A seal device according to Claim 16, wherein each said second groove is defined between first and second straight side edges which converge toward one another as said second groove projects radially inwardly, and wherein the first side edge of one said second groove and an adjacent second side edge of an adjacent said second groove extend in parallel relationship to define therebetween a flat land which is of constant width as the land projects radially inwardly between adjacent said second grooves.
19. A seal device according to Claim 16, wherein said second groove means includes an annular groove concentrically formed in said one seal face, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| US11224093A | 1993-08-26 | 1993-08-26 | |
| US11218093A | 1993-08-26 | 1993-08-26 |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA2147739A1 true CA2147739A1 (en) | 1995-03-02 |
Family
ID=26809666
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA 2147739 Abandoned CA2147739A1 (en) | 1993-08-26 | 1993-09-01 | Face seal with double groove arrangement |
Country Status (4)
| Country | Link |
|---|---|
| EP (1) | EP0670977A4 (en) |
| JP (1) | JPH08502809A (en) |
| CA (1) | CA2147739A1 (en) |
| WO (1) | WO1995006212A1 (en) |
Families Citing this family (9)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US6189896B1 (en) * | 1999-04-08 | 2001-02-20 | Caterpillar Inc. | Controlled leakage rotating seal ring with elements for receiving and holding a lubricant on a face thereof |
| CA2924554C (en) | 2008-07-02 | 2018-04-03 | Air Products And Chemicals, Inc. | Rotary face seal with anti-crowning features |
| WO2014103631A1 (en) * | 2012-12-25 | 2014-07-03 | イーグル工業株式会社 | Sliding component |
| CN106574725B (en) * | 2014-09-04 | 2019-04-16 | 伊格尔工业股份有限公司 | Mechanical sealing member |
| CN106439037B (en) * | 2016-11-18 | 2018-06-29 | 西华大学 | Sealing ring and mechanically-sealing apparatus with combination slot end face |
| CN107218395B (en) * | 2017-07-06 | 2018-11-13 | 浙江工业大学 | Ternary distorts type groove end surface mechanical sealing structure |
| EP3680519A4 (en) | 2017-09-05 | 2021-05-12 | Eagle Industry Co., Ltd. | Sliding component |
| JP7366945B2 (en) | 2019-02-14 | 2023-10-23 | イーグル工業株式会社 | sliding parts |
| EP3926188A4 (en) | 2019-02-15 | 2022-11-09 | Eagle Industry Co., Ltd. | Sliding components |
Family Cites Families (20)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| GB920892A (en) * | 1961-01-17 | 1963-03-13 | Gen Electric Co Ltd | Improvements in or relating to seals for gas filled machines |
| FR1505487A (en) * | 1966-10-28 | 1967-12-15 | Guinard Pompes | Improvement in leak-controlled rotary joints |
| FR1599308A (en) * | 1968-06-08 | 1970-07-15 | ||
| CA919210A (en) * | 1968-12-30 | 1973-01-16 | Westinghouse Electric Corporation | Controlled leakage face type shaft seal |
| US3640541A (en) * | 1970-06-12 | 1972-02-08 | Koppers Co Inc | Hydrodynamic lift-type face seal |
| US3782737A (en) * | 1970-07-13 | 1974-01-01 | Nasa | Spiral groove seal |
| US3675935A (en) * | 1970-07-13 | 1972-07-11 | Nasa | Spiral groove seal |
| US3804424A (en) * | 1972-04-24 | 1974-04-16 | Crane Packing Co | Gap seal with thermal and pressure distortion compensation |
| US4212475A (en) * | 1979-01-15 | 1980-07-15 | Crane Packing Co. | Self aligning spiral groove face seal |
| FR2511109A1 (en) * | 1981-08-06 | 1983-02-11 | Jeumont Schneider | HYDROSTATIC DEVICE FOR SEPARATING TWO CHAMBERS EMPLOYED BY FLUID UNDER DIFFERENT PRESSURES |
| JPS59231269A (en) * | 1983-06-14 | 1984-12-25 | Arai Pump Mfg Co Ltd | Mechanical seal |
| JPS59231268A (en) * | 1983-06-14 | 1984-12-25 | Arai Pump Mfg Co Ltd | Mechanical seal |
| CH677266A5 (en) * | 1986-10-28 | 1991-04-30 | Pacific Wietz Gmbh & Co Kg | |
| US4972986A (en) * | 1988-11-01 | 1990-11-27 | Eg&G Sealol, Inc. | Circumferential inter-seal for sealing between relatively rotatable concentric shafts |
| IT1227360B (en) * | 1988-11-18 | 1991-04-08 | Honeywell Bull Spa | MULTIPROCESSOR DATA PROCESSING SYSTEM WITH GLOBAL DATA REPLICATION. |
| US5039113A (en) * | 1990-01-17 | 1991-08-13 | Eg&G Sealol, Inc. | Spiral groove gas lubricated seal |
| US5066026A (en) * | 1990-06-11 | 1991-11-19 | Kaydon Corporation | Gas face seal |
| US5174584A (en) * | 1991-07-15 | 1992-12-29 | General Electric Company | Fluid bearing face seal for gas turbine engines |
| US5169159A (en) * | 1991-09-30 | 1992-12-08 | General Electric Company | Effective sealing device for engine flowpath |
| US5201531A (en) * | 1992-04-02 | 1993-04-13 | John Crane Inc. | Face seal with double spiral grooves |
-
1993
- 1993-09-01 WO PCT/US1993/008289 patent/WO1995006212A1/en not_active Application Discontinuation
- 1993-09-01 EP EP93921316A patent/EP0670977A4/en not_active Withdrawn
- 1993-09-01 JP JP7507534A patent/JPH08502809A/en active Pending
- 1993-09-01 CA CA 2147739 patent/CA2147739A1/en not_active Abandoned
Also Published As
| Publication number | Publication date |
|---|---|
| JPH08502809A (en) | 1996-03-26 |
| EP0670977A1 (en) | 1995-09-13 |
| WO1995006212A1 (en) | 1995-03-02 |
| EP0670977A4 (en) | 1995-12-20 |
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Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| FZDE | Dead |