Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
The present invention relates to rotating ~luid
machines and in particular to ~luid motors æuitable for
controlling the axial loads imposed on the rotor of the
machine during operation.
Rotating ~luid machines are used in a variety of
applications to transfer energy between a fluid and a
rotatin~ mechanioal system. Such machines include
compressors ~hich compress a gas in a continuous manner,
pumps for pumping liquids and turbines for deriving use~ul
work from a fluid flow The machines usually have a housing
with a fluid duct extending through the housing ~d one or
more rotors rotating within the duct. The rotors rotate at
a speed sufficiPnt to aause a pressure dif~erential betw~en
the inlet and outlet of the duct.
The rotors include an impeller mounted on a shaft
which is, in turn, supported in the housing on bearing
assemblies. Because of the high rotational speeds and close
tolerances encountered within certain classes of machines,
typically compressors, high demands are placed upon the
bearing assemblies. Such assemblies tend to be expensive
and of course must be designed to withstand the maximum load
that may be applied ~or extended periods. This in turn
increases the cost of the bearings.
Conventional hydrodynamic and antifriction
bearings incur significant parasitic losses and during
start-up the static friction in the bearings may be
sufficlent to prevent rotation of the rotor assembly
sub~ecting it to adverse conditions.
Magnetic bearings are utilized in some
applications to support the shaft for rotation and also
to oppose axial loads on the shaft. Magnetic bearings avoid
the limitations encountered in hydrodynamic and antifriction
bearings, particularly at high speed, and, through control
systems, per~lit dynamic adjustment of the bearings to
maintain the shaft aentered. However, the specific load
capacity of a magnetic bearing is less than that of a
mechanical bearing and so a physically larg~r bearing is
required to withstand the loads typically encountered in a
gas compressor. Moreover, whexe magnetic bearings are used,
the typical loads imposed on the bearings result in a
relatively large bearing assembly.
The loads imposed on the rotor of the compressor
are caused in part by the pressure differential across the
machine and also by the mass flow through the machine.
Attempts have been made to reduce the axial loads caused by
the pressure differential by utilizing a balance ~ ston
having one surface exposed to the high discharge pressure
and the other surface exposed to the inlet or suction
pressure. However, leakage occurs across the balance piston
which may represent a substantial loss in machine
efficiency. Moreover, the pressure differential and the
momentum forces vary with different operating conditions of
the machine so that a considerable axial force can still be
generated during operation of the machine which must be
accomodated by the bearings.
In these prior machines, gas seals were used
between the rotox and housing to prevent the egress of gas
from the fluid duct. For safety reasons, the seals w~re
typically used in axially spaced pairs so that if one seal
failed, the oth~r seal would provide safe operation until
the machines could be stopped and a repair could be
effe~ed. The seals used usually included components
rotating with the rotor and these were kept as small as
possible so as not to affect the rotor dynamics.
In USP 4,993,917, assigned to the assignee of the
present application, there is disclosed an arrangement in
which advantage is taken of the seal asssmblies to eliminate
the balance piston. A fluid motor is provided by a pair of
sea~s of differing effective diameter~ so that axial forces
can b~ compensated by the action of fluid in the motor.
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While this arrangement is satisfactory in many applications,
in practice the rotating fluid machine experiences a wide
range of ~luid pressure and ~low combinations. To ensure
that the axial loads opposed by the bearings are maintained
within a given range for all machine operating conditions,
the fluid motor must be capable of exerting a wide range of
counter thru~t on the rotor assembly. The magnitude of the
counter thrust applied to the fluid motor equals the
difference in pressure across the seal assembly times the
effective area of the seal assembly. To accomplish this,
the effective diameters of the two seal assembliè~ must be
correspondingly sized, which leads to an increased size of
at least one of the seal assemblies. As a result, the rotor
dynamics may be adversely e~fected due to the large mass
1~ supported on the relatively long shaft to accomodate the
fluid motor.
It is therefore an object of the present invention
to provide a ~luid motor for use in such machines which
obviates or mitigates the above disadvantages.
According to the present invention, there is
provided a rotating fluid machine having a housing, a
fluid duct extending through the housing, a rotor assembly
rotatably supported in the housing to be impinged by fluid
flowing through the duct, a fluid motor acting between said
rotor assembly and said housing, said motor comprising a
pair of radially extending end walls axially spaced along
said~ct with one wall connected to said rotor and the
other connected to said housing, and a pair o~ seal
assemblies extending between said end walls a~ radially
spaced locations to define an annular cylinder therebetween,
each of said seal assemblies including a pair o~ sealing
rings, a first o~ which is secured to said one wall for
rotation therewith and a second of which is secured to said
other wall.
By providing a pair of radially spaced seal
assemblies, extending between the end walls, the seal
assemblies may be nested permitting a reduction of the axial
length o~ the motor. The annulus formed between the seal
assemblies provides an area differential to permit fluid
within the cylinder to exert an axial force between the
rotor assembly and housing.
According also to the present invention, there is
provided a fluid motor for positioning between a pair of
relatively rotating members comprising a pair of axially
spaced end walls, secured to a respective one of ~aid
components for movement therewith, a pair of seal assemblies
extending between said end walls at radially spaced
locations to define an annulax cylinder therebetween, each
of said seal assemblies including a pair of sealing rings
secured to respective ones of said end walls for movement
therewith, whereby upon relative rotation between said
components, one o~ said rings of each seal rotates relative
to the other ring of its respective seal.
Embodiments o the inv~ntion will now be described
by way of example only w~th reference to the accompanying
drawings in which
Figure 1 is a sectional view through an overhung
compressor;
Figure 2 is a view of a pcrtion of Figure 1 on an
enlarged scale;
~ Figure 3 is a schematic representation o~ the
control circuit utilized to control the loads imposed on the
rotor of the Gompressor of Figure 1;
Figure 4 is a sectional view similar to Yigure 1
of a beam compressor; and
Figure 5 is a view of a portion of the compressor
shown in Figure 4 on an enlarged scale.
Referring therefore to Figure 1, a rotary fluid
machine, in this ~mbodiment, a compressor has a housing 10
with a fluid duct indicated generally at 11 extending
between an inlet volute 12 and an outlet ~olute 14. ~he
forward 2nd o~ inlet volute 12 is defined by a wall 13
(commonly referred to as 'a scoop') secured to a door 16
that closes the ~orward end of the housing 10.
A rotor assembly 17 including a shaft 18 is
rotatably supported within the housing 10 by a bearing
assembly 20. The bearing assembly 20 includes a bearing
housing 22 having a magnetic thrust bearing 24 an~^a pair of
magnetic radial bearings 26,28 spaced apart on either side
of the thrllst bearing 24. The magnetic bearings 24,26,28
are conventional in nature as describ~d for example in
USP 3,702,208 and will not be described herein in further
detail. Conventional antifriction bearings 29 are also
provided at spaced locations on the shaft 18 to provide
emergency support for the sha~t 18 in the event that the
magnetic bearings fail.
The rotor assembly 17 further includes a pair of
impellers 30,32 located at the forward end of sha~t 18 that
rotate with the shaft 18. Flow from the inlet volute 12
past the impellers 30,32 to the outlet volute 14 is
controlled by a diaphragm assembly 34 comprising an inlet
diaphragm 35, an interstage diaphragm 37 and a rear
diaphragm 39. The assembly 34 is secured within the casing
10 a~d~inlet vanes 36 direct gas to the first of the rotors
30. An internal passageway 38 directs gas ~rom the
discharge of the first impeller 30 to the inlet of the
second impaller 32 with labyrinth seals 40 positioned
between the rotor assembly 17 to seal between the impellor
30 and the diaphragm assembly 34. A fluid motor 42 is
located betwe2n the impeller 32 and the bearing housing 22
to seal between the discharge and the shaft 18 and to enable
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adjustment of the axial loads imposed on the bearing 24 in a
manner described below.
Fluid motor 42 i5 shown in further detail in
Figure 2 and includes a rotatable carrier 60 secured to
5 shaft 18 and a stationary carrier 62 secured to the bearing
housing 22. Carrier 60 includes a radially extending flange
64 with an axial nose 66 to define a pair of annular
recesses 68,70.
The carrier 62 also includes a radial flange 72
~0 that is axially spaced from the flange 64 and has a pair of
annular recesses 74,76 directed toward the recess~s 58,70.
An annular disc 78 i5 secured to the forward edge of the
carrier 62 and carries a labyrinth seal 80 at its inner edge
which engages the outer surface of the carrier 60. A
labyrinth seal 81 is also located at the opposite end of
carrier 62 to engage the outer surface of carrier 60.
A pair o~ dry gas seal assemblies 82,84
respectively are located between the flanges 64 and 72.
Each seal assembly 82,84 is similar in construction and
includas a pair of sealing rings 88,90. The seal assembly
82 is o~ greater diameter than seal assembly 84 so that an
area dif~erential is provided between the two seal
assemblies. Moreover, it will be noted that the
circumferential wall at the inner diameter of the seal
asse~bly 82 is greater than the circum~erential wall at the
outer diameter of seal assembly 84 so that the assemblies
may ~ nested and define an annular cylinder 86 between
them.
Each of the sealing rings 88 is l~cated in a
respective one of the recesses 68,70 and secured by a dowel
92. An 0-ring 94 is located between the rings 88 a~d the
carrier 62 to prevent gas passing behind the ring 88.
Each o~ the rings 90 is located in a respective
one of the recesses 74,76 and is axially moveable relative
to the carrier 62 by virtue o splines 960
A spring 98 acts through a thrust washer 100 to
bias each of the rings 90 toward the ring 88 50 that radial
faces 102,104 of the ring~ 88,90 abut. The faces 102,104
are configured to provide a pumping action for gas from a
high pressure zone to a lower pre~sure zone and so provide a
controlled leakage of gas across the seal upon relative
rotation between the rings. The con~iguration o~ the faces
is well known in the dry gas seal art and so will not be
described in further detail. It will be noted that th~
abutting faces 104,106 of each seal assembly 82,84 lie in a
common radial plane, perpendicular to the rotatio~al axis of
the shaft.
A vent line 106 i5 provided in the carrier 62 and
is used in a manner described below to control the pressure
within the cylinder 86. A vent line 108 is also provided to
evacuate gas passing through the seal assembly 84 and
contained by labyrinth seal 81.
To control the axial ~orces imposed on the rotor,
the pressure in the fluid motor 42 is regulated by the
control scheme shown in Figure 3.
As can best be seen in Figure 3, the control line
106 is connected to a pressure control valve 110 that vents
gas flowing through the control line 106 to a suitable vent.
Tha pressure control valve 110 is controlled by a pilot
pressure line 112 so that the pressure maintained in line
106 o~ the valve 110 is set by the pressure in the line 112.
The p~r~ssure in line 112 is derived from a signal fed to a
current-to-pressure converter 114 thxough signal line 116
that is itself connected to a ratio bias module 118. The
ratio bias module receives a control signal from a
tachometer 120 that senses the rotational speed of the shaft
18 in conventional manner. The tachometer 120 is also used
to operate the speed control system indicated at 12~.
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As more fully explained in co-pending application
495,920, it has now been recognized that the net ax~al ~orce
imposed on the shaft 18 varies with output speed.
To reduce the net axial forces, the control
arrangement shown in Figure 3 iB used to vary the pressure
in chamber 86 as the rotational speed of the shaft 18
varies. The difference in effective diameter of seal
assemblies 82,84 provides an annular area that may be used
to generate an axial force along the shaft 18. By varying
the pressure of gas in the cylinder 86, the axial force
exerted on the shaft 18 may also be varied. By c~rrelating
the pressure in the cylinder 86 to the rotational speed of
the shaft 18, an appropriate axial force may be imposed on
the shaft 18 to counteract the inherent axial forces
generated by operation of the machine. This maintains the
net axial force on the shaft 18 within a predetermined range
over the range of normal operating speeds.
In operation therefore, the rotor assembly 18 is
rotated by a suitable drive means and gas supplied to the
inlet 12 is compressed and discharged through the outlet
volute 14. A small flow of the high pressure discharge gas
passes by labyrinth seal 80.
The dry gas seal assembly 82 functions by
permitting a controlled but very small amount of gas to flow
between the relatively moving surfaces of the rings 88,90.
Thus a small amount of gas flows into the cylinder 86 where
its ~ressure is applied between the end walls defined by
flanges 64,72 of carriers 60,62 respectively. The pressure
in cylinder 86 is controlled by the valve 110 to be
maintained at the required level as determined by the
tachometer signal.
Rotation o~ the rotor assembly 17 also generates a
signal ~rom the tachom~ter 120 which is applied to the ratio
bias module 118. The ratio bias module may provide varying
gains and varying offsets so that the desired output
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relationship to the input may be obtained. The input signal
to the module 118 therefore produces the desired output
signal in line 11~ and sets the converter 114 at the
required control pressure in line 112 to produce the desired
pressure in control line 106.
As the speed of the compressor increases, the
discharga pressure in volute 14 and the mass flow acting on
the impellers 30,32 increase. The mass flow may also vary
depending upon the inlet and outlet conditions. The net
efect typically is an increase in the axial thrust in the
direction of the inlet volute due to increas2d p ~ssure at
the discharge volute 1~. This may be offset in part by an
increase in momentum forces toward the discharge volute 14.
The pressure in the cylinder 86 is modulated so that the
force acting through flange 64 of the carrier 60 away from
the discharge volute is decreased. In this way, the net
axial forces imposed on the thrust bearing assembly 24 are
reduced, allowing for a smaller bearing assembly.
It will be seen, therefore, that by monitoring the
speed of the compressor shaft 18 and utilizing that signal
as an indication of end thrust, it is possible to reduce the
variations in thrust forces imposed on the shaft lS in a
progressive and controlled manner. Moreover, the compact
configuration of fluid motor 42 by virtue of the nested seal
assemblies 82,84 permits installation of the motor within
the constraints of existlng machines while providing the
nece8sa~ry control function.
An alternative ~orm of compressor known as a beam
type is shown in Figures 4 and 5 in which the sha~t 18a is
supported at laterally spaced locations. The operation of
the compressor shown in Figures 4 and 5 is substantially
similar in many respects to that of the overhung compressor
shown in Figures 1 and 2 and therefore like reference
numerals will be utillzed to describe like components with a
suffix 'a' added for clarity. In the compressor shown in
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Figures 4 and 5, gas from the inlet volute 12a passes
through rotor assembly 17a and into the discharge duct 14a.
Of course, additional impQllers 30a may be mounted upon the
shaft 18a to provide multiple stagPs o~ compression if
desired.
The shaft 18a is supported at spaced locations by
radial magnetic bearings 26a and 28a respectively and axial
forces are accomodated by a magnetic thrust bearing 24a at
the forward end of the compressor. The bearing~ 24a and 26a
are mounted outboard of an end 16a that closes the inlet
volute 12a and utilizes a dry gas seal assembly I~0 to
prevent the flow of gas between the door 16a and the shaft
18a.
Control over axial loading of the shaft 18a is
provided by a fluid motor 42a shown in more detail in
Figura 5. Fluid motor 42a is substantially the same as
motor 42, having a pair of radially spaced seal assemblies
82a,84a defining an annular cylinder 86a between them.
However, an alternative embodiment of the motor 42 is shown
in Figure 5 where the cylinder 42a, he rings 88a of both
seal assemblies are fsrmed as an integral unit identified as
130 that extends radially along the flange 64a and is
secured by dowel pin 92a. An 0 ring 94a seals between the
unit 130 and flange 64a. This arrangement simplifies the
construction of motor 42a by reducing the individual
components. It will be appreciated that this arrangement
could~e used in the overhung compressor shown in Figures 1
to 3 and that separate sealing rings could be used in the
beam type compressor of Figures 4 and 5.
Operation of the motor 42a to control the axial
~orces is similar to that described above with respect to
Figures 1 to 3. It will be appreciated that alternative
control strategies may be used, such as monitoring pressure
di~ferential. In each case, howev~r, the compact
con~iguration of the motor 42a permits incorporation in an
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installation that would otherwise be extremely dlfficult due
to mechanical constraints.
The embodiments of Figures 4 and 5 illustrate the
fluid motor at the discharge end of the compressor lOa. It
will be appreciated, however, that the motor 42a could be
incorporated in place of seal assembly 100 at the inlet end
to control the net axial forces on the rotor assembly 17a
with a conventional dry seal assembly utilized at the
discharge end of the compressor.
As will be appreciated from considering the
arrangement of the seals, the range of thrust for~es that
can be generated is a function of the area differential
between the seals and the maximum pressure that an be
imposed in the cylinder 86. By selecting the seals so that
their physical dimensions permit nesting, the effective area
differential will be determined and thereafter the pressure
range can be determined. Moreover, the nesting of the
assemblies reduces the overall shaft length and enhances the
rotor dynamics.
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