Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
~ 21~3~694
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~ GAS LUBRICATED CONTACT FREE SEALING ARR~NGEMENT FOR A SHAFT
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The invention relates to gas lubricated, contact free sealing
arrange~ents for a shaft having a seal housing, a stationary sealing
ring located in the seal housing and a rotatable sealing ring mounted
on and rigidly affixed to a rotatable shaft. The sealing rings work
against each other at their sealing end surfaces while being separated
by a seal gap, which is created by the lifti.ng force of a lubricating
gas flowing between the sealing surfaces. The rotatable sealing ring
is made of a material which has a high heat conductivity as well as a
high modulus of elasticity and is of high hardness. The stationary
sealing ring is separated by a functional gap from a cylindrical
section of the seal housing which is located towards the shaft, during
operation and at a predetermined operating pressure differential. The
functional ~nn~llar gap is sealed through an O-ring, which is made of
rubber or plastic and is positioned in an annular groove. The
stationary sealing ring is pressed against the lifting force of the
lubricating gas with a predetermined force provided by at least one
pressure spring. In such a sealing arrAn~e -t one must differentiate
between the functional annular gap described above and a construction
dependent ~nn~ r asse~bly gap. The f~mctional annular gap develops
out of the ~nmll~r assembly gap through a settling, in the assembled
conditionj under the effect of a pressure difference which is sealed
during operatio~. It is readily apparent t~at the re~in~er of the
' 25 construction oE such a gas lubricated contact free sealing
ar~rangement, taking into consideration the special sealing functions,
may be realized in accordance with constructions taught in the art
and, especially, by using the corresponding methods, which have been
developed on the sub~ect iTI engineerlng science since 1925.
Appropriate hard sealing materials are described, for example, in
VDI-Zeitschrift, volume 102 (1960) number.l8, pages 728 to 732.
In a sealing arrangement known in the art as described, for
example, in European Paeent EP 00 13 678, the sealing end surfaces are
provided with recesses, which are constructed as spiral grooves that
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commence at at least one circumferential edge of the respective
sealing ring end surface. Only ths rotatable sealing ring is made of
a material which has a high heat conductivity as well as a great
modulus of elasticity and hi~h hardness. Consequently, the stationary
sealing ring is ~ade of a material, namely carbon, which has a
comparatively small ~odulus of elasticity and a low hardness and which
heat conductivity is not outstanding. No syecial importance is
attached to the surface roughness and the pore volume of the sealing
end surfaces. In this prior art sealing arrangement a distortion, a
partial turning inside out of the stationary sealing ring, which is
caused by the working temperature of the sealing arrangement, results
from the relatively small modulus of elasticity and the low heat
conductivity of the material of the stationary sealing ring. In fact,
the temperature drop in axial direction is 25~C or more. Such a
distortion of the stationary sealing ring would negatively influence
the sealing conditions and the life of both sealing rings and, thus,
the sealing arrangement, because of an unavoidable contact of the
sealing surfaces during operation. Therefore, in the context of the
generic measures known in the art, the positioning and construction of
the sealing arrangement is selected so that the pressure distribution
in the sealing gap (i.e. the lifting pressure of the gas in the gap)
produces forces which counteract the distortion. In order to achieve
this, it is absolutely required that the recesses a~e actively
conveying spiral grooves, which produce a pumping effect and that the
spiral grooves, which are included in at least the rotatable sealin~
ring, commence at one circumference of the sealing end surface only,
and terminate at a dam or web of the sealing end surface, whereby
certain numerical parameters must be met, with respect to the depth of
the spiral grooves, the so-called width ratio of the webs and the
desired equilibrium. However, even if these absolute requirements are
met, the achieved effect is unsatisfactory. The desired equilibrium
does not exist during all operating conditions. The plano-parallel
positioning of the sealing end surfaces may only be restored to
ro~; ~ol 1 y 70% through this resetting. All of this is based on the
fact that the prior art measures do not take into consideration the
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tribological characteristics of the seal (such a~ ~; supportable
load, hardness, moment of friction), while the deficiencies which
result therefrom cannot be suffici~ntly compensated through the
attempt to reset the distortions of the st,qtionary seallng ring.
Furthermore, an 11nA~ceptably high leakage rate results in prior art
constructions from the proble~s described above, which leakage rate
increases on a large scale with increasing rotation speed of th~ shaft
and, thus, the rotatable sealing ring, because of the pumping action
of the spiral grooves, and is even more increased through the
incomplete resetting.
The temperature and pressure difference dependent distortion of
the stationary sealing ring which is accepted in the above described
prior art construction and even provoked through the provision of a
surface inertia of the stationary sealing ring, which deformation is
supposed to be reset by the pressure distribution in the lubricating
gas, has a further disturbing disadvantage: the above described
distortion requires, in accordance with the laws of mechanics, that
the Annl11Ar assembly gap must be substantially larger than the
functional Ann1llAr gap, which i~ achie~ed when the stationary sealing
~- - ring is distorted, when the resetting of the distortion through the
pressure distribution in the lubricating gas is not taken into
consideration. This resetting is not important for the dimensioning
of the Ann1l1Ar assembly gap and the ratio of the annular assembly gap
to the ~unctional ~nn~llAr gap at different operating conditions of the
sealing arrangement, since the temperature dependent deformation and
resetting of the stationary sealing ring, which is dependent on the
o~erating conditions of the sealing arrangèment, are not achieved
simultaneously, while, on the other hand, a contact between the
sealing end surfaces of the rotatablc sealing ring and the stationary
sealing ring must be reliably prevented. The width of the
construction dependent Anmll~r assembly gap is practically in the
region of 0.4 or 0.5 mm for all common dia~eter ratios of sealing
arrangements of the construction-described above. The width of the
~nn~ r gap must be covered by the 0-ring made of rubber or plastic
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even at the operating pressure difference, and especially when the
final temperature distribution has not been achieved at the start-up
of the operation, as well as when the desired resetting has been more
or less completely achieved after the final adjustment of the
temperature distribution. In prior art constructions, the O-ring ~ust
therefore be constructed in such a way, with respect to the hardness
of its material in Shore, that it may not be pressed into an annular
gap, which corresponds in si~e to the ann~ r assembly gap of about
0.4 or 0.5 mm, by the generated operating pressure difference and is
not prematurely destroyed by this, so-called, extrusion. Thus, in
prior art constructions, the O-ring is selected to have a
corresponding hardness. Its material hardness is generally at least
90 Shore A in accordance with DIN 53 505. Nevertheless, operating
pressure differences of over 80 or r~ir~1 ly 100 bar ~re practically
uncontrollable with prior art constructions, if the application
requires the common lifetime of several thousand hours for the overall
sealing arrangement. On the other hand, practice more frequently
requires sealing arrangements in accordance with the principles of
construction described above, which may be used for operating pressure
differences w~ich are much above 100 ~ar. In addition, the function
of the known overall sealing arr~ng~- t is adversely affected by an
O-ring of high hardness in the annular gap between the stationary
sealing ring and the cylindrical section of the seal housing. In
- fact, in prior art sealing arrangements, the shaft with the rotatable
sealing ring never rotates completely round and completely co-axial to
the stationary sealing ring and ad~acent parts of the housing. This
may result in vibrations which are transmitted into the stationary
sealing ring. Therefore, a certain play must be allowed in the
dimension of the functional Ann111Ar gap (independent of the size it
reaches during operation), while a negative effect on the sealing
action must be prevented. If an O-ring of high material hardness is
used, it does not follow the induced vibrations of the stationary
sealing ring and reduces the given play, which has a negative effect
on the sealing action in the An~ r gap between the cylindrical
section and the stationary sealing ring as well as in the gas
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lubrication. Furthermore, this may lead to unwanted contacts between
the sealing end surfaces of the rotatable sealing ring and the
stationary sealing ring.
It i9 an ob~,ect oE this disclosure to provide a gas lubricated
contact free sealing arrangement of the principle construction
described above, which may be used for substantially higher pressures
than prior art constructions, for example, for operating pressure
differences of up to 300 bar or even up to 500 bar, while a good
sealing action and a long lifetime is achieved.
i Accordingly, here described is a gas lubricated, contact free
sealing arrangement for a shaft including a seal housing, a stationary
sealing ring located in the seal housing, and a rotatable sealing ring
mounted on and rigidly affixed to the shaft. The sealing rings act
against each other with their sealing end surfaces at a sealing gap
filled with a lubricating gas. The rotatable sealing ring is made of
a material having a high heat conductivity, a great modulus of
elasticity and a high hardness. The stationary sealing ring is
separated under operating conditions and at a predetermined operat;ng
pressure difference by a functional ~nn~ r g,ap from a cylindrical
section of the seal housing located towards the shaft. The functional
gap is sealed by an O~ring made of rubber or plastic. The stationary
sealing ring is pressed against the lifting force of the lubricating
gas at a force determined by at least one pressure spring. The
stationary sealing ring is also made of a hard sealing material of
high heat conductivity, great modulus of elasticity and high
hardness. Specifically, the stationary sealing ring, as well as the
rotatable sealing ring are made of a hard sealing material which has a
heat conductivity of over 70 U/~K (~kJ/mhK), and a modulus of
elasticity of over 25Q 000 N/mm2 at a corresponding hardness and a
pore volume of less than 1% as well as a surface roughness of under
0.3 ~m (Ra), preferably 0.03 ym (Ra). The stationary sealing ring has
such a large surface inertia that, in operation and at the operating
pressure difference, the width of the functional annular gap between
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the stationary sealing ring and the cylindrical section of the seal
housing located towards the shaft is practically equal to the
construction dependent width of the annular assembly gap between the
cylindrical section and the stationary sealing ring and smaller than
0.4 mm, preEerably smaller than 0.3 mm. ~le sealing 0-ring is further
used as a compensating and centering ring for the stationary sealing
ring and, to this end, has a material hardness, which is higher than
the extrusion threshold of the material hardness at the given gap
width and the given operating pressure difference, and is smaller than
a material hardness of 90 Shore A in accordance with DIN 53 505.
Preferably, the material hardness is smaller than 80 Shore A in
accordance with the citsd standard. In other words, the positioning
and construction of the sealing arrangement is so that the orientation
of the sealing end surfaces of the rotatable sealing ring and the
stationary sealing ring remains constant and preferably remains
parallel at all times.
It is not necessary to take the production of any resetting
forces into consideration for the construction of recesses whic~ are
provided in at least one of the sealing end surfaces. Therefore, the
recesses and the sealing end surfaces of the rotatable sealing ring
and the stationary sealing ring may be constructed so that an optimal
sealing action is achieved through the lubricating gas. To this end,
;; the recesse~ may be constructed as actively conveying spiral grooves.
However, in ~his context, it is also possible to construct the
recesses as pressure effective recesses having a pressure edge.
With respect to gas dynamics, spiral grooves are elements,
which effect a defined transport of gas. The transport may be used to
counteract the leakage stream, which determines the leakage rate
~iscussed in the beg~ nni ne. In contrast, recesses which have a
pressure edge are elements which counteract a defined transport of the
gas and effect a pressure build-up. In both cases, the sealing action
may be optimized and is not negatively affected by the requirement for
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t~e production of rssetting forces within speci.al e.quilibrium
conditions.
In a preferred embodiment, the stationary sealing ring has a
surface inertia which prevents a temperature dependent distortion of
its sealing end surface. Such a surface inertia may be easily
determined by the modern computer aided calculation methods of
technlcal mechanics. "Ra'i defines the maan roughness value in
accordance with DIN 4768. In a further preferred embodiment of the
invention, the sealing end surfaces have an evenness of 0.4
micrometers per lOO mm diameter at room temperature and at a
temperatura gradient of 0.
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With respect to the described requiremants in relation to heat
conductivity, modulus of elasticity and hardness, the saaling rings
may be made of diffarent materials. The sealing rings are prefarably
made of a material selected from the group of tungsten carbide,
silicon carbide, silicon/silicon carbide-compound, titanium carbide.
The sealing rings are manufactured, for example, through sintering or
pressure sintering, whereby the pore volums may be ad~usted. Both
sealing rings ~ay be made of the same material. However, the sealing
rings can be made, with respect to the stationary sealing ring on one
hand and the rotatable sealing ring on the other hand, of combinations
of two of the above mentioned materials. In order to optimi~e the
sealing arrangement, the sealing rings preferably have a pore volume
of less than 0.5~. Tha stationary sealing ring preferably has a
cross-section, which axial extent is at least twice its radial extent.
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In a preferred embodiment of a gas seal arrangement in
accordance with the invention, which is preferred with respect to the
construction of tha recesseC~ the recesses commence at one
circumference of the sealing end surface and terminate at a dam of the
sea~in~ end surface, which dam consists of a recess-free region of the
sealing end surface. However, tha recesses may also commence at both
the inner circumference and the outer circumference of the sealing end
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surfaces and terminate at a recess free median dam. In this case, the
embodiment may be provided with spiral grooves in such a way that the
pumping act;ons are counter directed. It is preferred, especially in
an embodiment with pressure edges, that the. recesses terminate at a
meandrical dam. The sealing end surfaces may be provided with an
emergency glide finish. This may be, for example, a layer or coating
of several micrometers made of graphite, polytetrafluoroethylene, or
similar materials. The emergency glide finish may also be provided by
carbon which is incorporated into the seal;ng ring material.
The embodiment which includes recesses with pressure edges is
of special importance. In this contex~, a preferred embodiment of the
~ invention is characterized in that the pressure edges of the recesses
; e~tend in radial direction. However, the pressure edges may also be
arcuate segments of recesses which are circular in plan view. In
another embodiment o~ the invention, the pressure edges are
constructed as lateral edges of recesses which are triangular in plan
view and have one apex located at a circumference of the sealing end
surface, which apex is cut off. It is pre~erred to construct the
recesses in such a way that they are symmetrical in relation to a
radially extending line. If this symmetry is achieved, a gas sealing
arrangement in accordance with the invention is rotation direction
independent. If this is not required or desired, the recesses may be
constructed with unsymmetrical, for example, L-shaped, pressure
edges. The depth Qf the recesses is in the micrometer range.
, :
The advantages achieved with the gas seal arrangement here
described may be seen in that, the tribological characteristics are
combined and the recesses constructed in such a way that the
production of a momentum from the pressure distribution in the sealing
gap for resetting the distortion, is no longer required. Recesses
which produee a distinct pumping action may be omitted and are, in the
embodiment with pressure edges, practically completely omitted, which
substantially reduces the leakage rate. In comparison to a prior art
generic gas seal arran~ement, the novel gas seal arrangement,
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constructed for the same operating conditions, provides for a 50%
reduced leakage rate. This is partially due to the effect that
although the sealing rings are heated during operation, they have such
a small temperature gradient after the start up period and the
subsequently resulting temperature equilibri.um that, even for this
reason alon~, unwanted distortions will practically not appear. The
temperature gradient in axial direction i9 below 1~C, while it is
almost 20~C in the above described prior art constructions. This is
true for all common parameters of the gas s0aling arrangement, and,
for example, for a shaft diameter of 50 to 250 mm and a gliding speed
of up to 150 m per second. For the rer~;n~er, the surface inertia of
the stationary sealing ring prevents distort;ons thereof. In
addition, the small pore volume and the small surface roughness in the
region of the recesses provides for a further reduction of the leakage
rate. Surprisingly, no start-up and shut-down problems have been
observed. The latter is due to the fact that the annular assembly gap
and the functional ~nn~ r gap between the stationary sealing ring and
the associated cylindrical section of the housing have substantially
the same width during all operating conditions. This width is so
20 ~;nir~l, that an O-ring can be used which has a very low material
hardness, without the danger of extrusion of the O-ring into the
:~ ~nn~ r gap, even at high operating pressure differences. Therefore,
~; the sealing arrangements described are especially suited as high
pressure seals and have a longer service life.
Embodiments of the invention will now be further described by
way of example only and with reference to the accompanying drawings,
wherein
Figure 1 shows a schematic illustration of half of an axial
cross-section taken through an assembled sealing arrangement embodying
with the invention;
Figure 2 is an end view of the rotatable sealing ring of the sealing
arrangement shown in Figure l;
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Figures 3 to 7 are preferred embodiments of the rotatable sealing ring
shown in Figure 2;
Figure 8 shows an enlarged section of Figure 1 lllustrating the mutual
orientation of the sealing end surfaces of the stationary and the
rotatable sealing ring as well as the sealing gap; and
Figure 9 is an enlargement of the region o~ the functional sealing gap
shown in Figuxe 1.
A sealing arrangement for a shaft as shown in Figures 1 to 9
includes, a sha~t 1, a seal housing 2, a stationary sealing ring 3
which is located in seal housing 2 and a rotatable sealing ring 4,
which is mounted on and rigidly affixed to shaft 1. Sealing rings 3
and 4 move against each other with their respective sealing end
surfaces 3a and 4a at a sealing gap 6, which is not visible in Figure
1 for reasons of scale. The sealing end surfacs 4a of rotatable
sealing ring 4 is, in the preferred embodiment of the invention,
provided with recesses 5 which are open to one circumference of the
sealing end surface ~see Figs~ 2, 4, 5 and 6). The stationary sealing
ring 3 acts in a direction towards the rotatable sealing ring 4 with a
predetermined force which results from springs 7 distributed along the
circumference of the stationary sealing ring 3, in t~is embodiment.
The stationary sealing ring 3 is axially movably supported and has a
ring height 8 which is larger than the width of its sealing end
surface 9.
The sealing rings 3 and 4 are made of a materlal which has a
high heat conductivity, as well as a great modulus of elasticity, and
a high hardness. Both sealing rlngs 3 and 4 have a small pore volume
and a min;r~l surface roughness. The stationary sealing ring 3 is
further provided with an axial surface inertia, whîch prevents heat
distortions of its sealing end surface 3a. This is apparent from the
cross-section shown in Figure 1. Turning now to Figures 2 and 3,
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recesses 5 are constructed and positioned in such a way that the
leakage rate is mini~i7ed, without taking into consideration the
momentum resulting from the pressure distribution in thc sealing gap
and counteracting the creation of distortions. The reduction of the
leakage rate is achieved through pressure edges 5a, which counteract
the pumping effect. The recesses 5, shown in Figures 2 and 3, are of
a T-shape, and have radial pressure edges 5a. Recesses 5 are
circular in Figure 4, and triangular in Figure 5, and have a similarly
cut apex in both Figures. Sealing rings 3 and 4 are made of one
material selected from the group of tungsten carbide, silicon carbide,
silicon/silicon carbide-compound, titanium carbide, o~ combinations of
two of these.
It is apparent, from Figures 2, 4 and 5, that recesses 5
commence at an outer circumference of sealing end surEace 4a, and
terminate at a dam lO, which is formed by a recess-free region of
sealing end surface 4a. In Figure 3, recesses 5 commence at the inner
as well as the outer circumference of sealing end surface 4a and
terminate at a median, recess-free dam lO. In Figure 3, dam lO has a
substAntiAlly meandrical shape. Sealing end surfaces 3a and 4a may be
provided with an emergency glide finish of small thickness made of
graphite, polytetrafluoroethylene or similar materials, which is not
illustrated, for reasons of scale. In the embodiment shown in Figures
6 and 7, recesses 5 are constructed as spiral grooves. The remainder
of the construction is the same as in the embodiments shown in Figures
2 and 3.
Figure 8, which is an enlarged section of Figure l, illustrates
that the two sealing end surfaces 3 and 4 are positioned very exactly
plano-parallel to each other. In operation, they define a sealing gap
D, which is a fixed gap. It is maintained by the lubricating gas. A
functional slnnt1l~r gap F is apparent in Figures l and 9. Functional
stnn~1lstr gap F, which self-adjusts during operation, is equal to the
construction-dependent annt1lStr assembly gap. Both are practically of
the same gap width, The radial gap width is preferably equal to or
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smaller than 0.3 mm. The stationary sealing ring 3 is constructed to
be distortion free, under all operating conditions. The enlarged
section of Figure 1 shown in Figure 9, illustrates that the V-ring ll,
even at high operating pressure differences, may not be pressed into
the functional annular gap F, which has the aforesaid very small gap
width, so that the O-ring may not be extruded into this functional
~nm~l~r gap F. Therefore, O-rlng ll may have a relatively low
material hardness, so that the above described compensation and
centering of tbe rotatable shaft and sealing ring com~ination is made
possible. The elasticity of O-rings ll is advantageous for the mutual
orientation of the sealing end surfaces. The pressure spring 7, shown
in Figure l, may, for example, be constructed as a closed spring
bellows, and may further be used as a centering element.
lS While the sealing end surfaces 4a and 3a of rotatable sealing
ring 4, and stationary sealing ring 3 respectively, are positioned
plano-parallel to each other in the embodiments shown in Figures 1 to
7, Figure 8, which is an enlarged section of Figure l, shows in broken
lines an embodiment wherein the sealing end surface 3a of the
statlonary sealing rlng 3 includes two AnnlllAr surfaces, in radial
direction, which merge at an edge 12. In this embodiment, the annular
surfaces are obliquely positioned to each other to form a ridge. They
may also form a step at the location of t~e ridge.
The stationary sealing ring 3 may be constructed in such a way
that it is not sub~ect to unwanted temperature-dependent distor-tions
or other operation-dependent deformations. This provides for the
; realization of a very small annular assem-~bly gap, and a practically
corresponding functional AnnlllAr gap F, which is small enough to allow
the use of 0-rings which may not be extruded into the annular gap,
which would prematurely destroy the O-ring, even at an operating
pressure difference of 300 or even 500 bar. As a result, the new gas
lubricated sealing arrangement is characterized in that it may be used
for high and even very high pressures, at which it has a mini~
leakage rate, and a very long service life.
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