Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
l ~r~a ~ ` 21 66294
W095/02767 PCT~L93/00150
Rotary screw compressor.
The present inventionrelates to a rotaryscrew compressor
comprising a casing, a male rotor and a female rotor cooperating
therewith enclosed in a working space defined by the casing,
the casing having a discharge outlet connected to an outlet
port at the high pressure end of the working space and a suction
inlet at the low pressure end of the wo~king space, at least
one rotor being rotatably supported at an end thereof through
a bearing arrangement comprising a bearing bracket being fixed
to an end cover and having a substantiaLly cylindrical outer
circumferential surface, the bearing bracket projecting into
an axial cavity provided in the rotor forming a first chamber
between the bracket and the rotor, the bracket being provided
with lan oil feed channel to feed oil into the first chamber.
A rotary screw compressor of this kind for the compression
of gas is known from JP-A-59-168290. During operation of a screw
compressor the rotors are subjected to radial loads arising
from the compression of the gas. At the high pressure end of
the working space of the known compressor a cylindrical bearing
bracket is provided for each rotor, leach bearing bracket
projecting from the end cover into an :internal axial cavity
provided in the high pressure end of the corresponding rotor.
Pressurized oil is fed through an oil feed channel into the
chamber between the bearing bracket and the rotor. The oil then
leaves the chamber and enters the working space of the
compressor. Finally the oil is seperated from the compressed
gas and fed into the chamber again. The rotors will also be
exposed to a higher pressure at their high pressure end than
at th(eir low pressure end, resulting in an axial force acting
on each rotor towards the low pressure end~. Therefore each rotor
of the known compressor isprovided with a rolling contact thrust
bearing at the low pressure end.
The bearing arrangement of the known compressor has the
disadvantage that it has a limited load bearing capacity,
particularly in the radial direction of the rotors. Therefore
the known compressor is not capable of producing a high discharge
pressure or large differential pressure between the discharge
outlet and suction inlet.
WO9S/02767 ~ 2 1 6 6 2 9 4 PCT~93/00150 ~
- 2 -
It is an object of the present invention to provide a
rotary screw compressor according to the preamble which has
an improved bearing arrangement with a high load bearing capacity
in order to handle a high discharge pressure or a high
differential pressure.
This has according to the invention been achieved in
that at least one rotor is rotatably supported at the low
pressure end thereof through the bearing arrangement, the
correspondingbearing bracket beingmountedatthe lowpressure
end of the working space and the outer circumferential surface
thereof being provided with at least a groove connected to the
oil feed channel and a recess connected to an oil drainage
channel provided in the bearing bracket, and in that sealing
means are provided between the first chamber and the working
space of the compressor. According to the invention an
uncomplicated bearing arrangement is obtained capable of
supporting high radial loads. The load bearing capacities of
this bearing arrangement not only arise from the hydrostatic
pressure of the pressurized oil fed into the first chamber but
also from hydrodynamic load bearing effects between each
stationary bearing bracket and the corresponding rotor, which
will rotate at a high speed. As the pressurized oil can also
be present in the space of the first chamber between the end
face of a bearing bracket and the bottom of the internal cavity
of the rotor axial loads on the rotor can also be supported.
In a preferred embodiment the end face at the low pressure
end of a rotor, the end cover, the casing, and the corresponding
bearing bracket define a second chamber, the second chamber
being connected to an oil feed channel. In this manner the
pressure of the oil fed into this second chamber acts as a
hydrostatic thrust bearing capable of supporting at least a
part of the axial load on that rotor.
In another preferred embodimentthe outer circumferential
surface of at least one of the bearing brackets is provided
with two longit~ n~l grooves and one recess, the recess being
located on the side of the bearing bracket radially opposite
the outlet port and being connected to the oil drainage channel,
~ WO9S/02767 ~ 2 1 6 6 2 9 4 PCT~L93/00~5~
the longitudinal grooves being located at either side of the
recess and being connected to the oil feeclchannel. The presence
of two longitudinal grooves, each grool~e being connected to
the oil feed channel, provides a zone in the first chamber
wherein a high oil pressure is maintained for counteracting
the radial load on the rotor. The location of the recess, which
is connected to an oil drain channel, on the bearing bracket
radially opposite the outlet port of the working space is
preferred as an optimal counterbalancing of the radial load
on the rotor can be obtained in this manner.
In a particularly advantageous em~bodiment the edges of
the longitudinal grooves adjacent the recess are situated in
a common plane through the axis of the bearing bracket at an
ec~al distance from the recess, and the edqes of the longitn~in~l
grooves most distant from the recess are each situated in a
plane inclined at an angle cr to the common plane.
Preferably each recesshas an approximate maximum length
of 0.7 times the length of the bearing bracket. As each recess
is located at the portion of the bearing bracket adjacent the
end face thereof, a portion of the bearing bracket having a
cylin~drical cross section at the low pressure side of that recess
forms a restriction between the recess and the second chamber
provided at the low pressure end of the rotor. The restriction
thus obtained prevents pressurized oil from flowing from the
second chamber towards the recess and therefore prevents a drop
in oil pressure in the second chamber.
Since the radial load on the male rotor arising from
the compression of the gas is less than the radial load on the
female rotor, due to the geometry of the rotors, the length
of the bearing bracket of the male rotor and/or the length of
the recess thereof is preferably less than the length of the
bearing bracket of the female rotor and/or the recess thereof.
In another preferred embodiment ~a groove connected to
the oil feed ~h~nn~l on the bearing bracket of the male rotor
and a recess on the bearing bracket of the female rotor terminate
at the end face of the corresponding bearing, and each recess
on the bearing bracket of the male rotor and each groove on
WO9s/02767 ~ ~1 6 6 2 9 4 PCT~Lg3/~olSo~
the bearing bracket of the female rotor are located spaced from
the end face of the corresponding bearing bracket. Due to the
geometry of the rotors the axial load on the male rotor arising
from the compression of the gas is as a rule greater than the
axial load on the femalerotor. To compensate for this difference
an additional axial force is exerted on the male rotor as the
pressurized oil supplied to a longitll~;nA~ groove on the bearing
bracket of the male rotor enters the space between the end face
of that bearing bracket and the bottom of the internal cavity
of the male rotor. The return flow of oil to the recess is
obstructed and the oil pressure in this space is maintained.
For a high-speed screw compressor capable of a high
pressure difference between the discharge outlet and the suction
inlet it is advantageous that at least one of the rotors is
provided with a ring shoulder protruding from its low pressure
end, the sealing means being provided between the ring shoulder
and the casing. This provides a further increase of the axial
thrust load bearing capacity of the bearing arrangement according
to the invention.
For a low-speed screw compressor with a relatively low
pressure difference and wherein cooling is obtained by feeding
oil into the working space of the compressor it is advantageous
that at least one of the rotors is provided with sealing means
between the rotor and the corresponding bearing bracket. The
low-speed screw compressor is also preferably provided with
a rolling contact bearing between at least one of the rotors
and the corresponding bearing bracket.
Further advantageous embodiments of the rotary screw
compressor according to the invention are specified in the claims
11-13.
The rotary screw compressor according to the present
invention is capable of achieving considerably higher
differential pressures between the discharge outlet and the
suction inlet and considerably higher~;schArge pressures than
the known compressors of this kind. Traditional screw compressors
having bearings located outside the helical screw part of the
rotors are known to achieve a differential pressure of up to
~ W095/027~ ? ` ~ 2 ~ 6 ~) 2 9 4 PCT~L93/00150
-- 5
15-2~ bar. The rotary screw compressor according to the invention
can achieve high differential pressures and ~isc-h~rge pressures
as much as 3 to 4 times higher. Therefore the inventive
compressor can compete with centrifugal and piston compressors,
and can be used, for example, for compression of natural gas
in gas and oil fields, in gas delivery, gas filling and gas
lift stations for gas and oil production, transportation,
refinery and power Le~overy and chemical plants as well. Further
advantages of the rotary screw compressor according to the
invention are its simple design, reliability and long service
life, in particular regarding the d,esign of the bearing
arrangements at the low pressure end, :its limited weight and
small dimensions.
The invention will now be explained in greater detail
through the following description of pr~eferred embodiments of
the screw compressor according to the invention, wherein
reference is made to the accompanying drawings, in which:
fig. 1 is a longitll~;nAl section through the male rotor
of a first embodiment of the screw compressor according to the
invention,
fig. 2 is a section taken along :Line II-II of fig. l,
fig. 3 is a section taken along ].ine III-III of fig.2,
fig. 4 is cross section of the bearing bracket of the
male rotor of fig. 1,
fig. 5 is a view corresponding l_o fig. 2 of a second
embodiment of the screw compressor according to the invention,
fig. 6 is a diagrammatic view, partly sectional, of a
thircl embodiment of the screw compressor according to the
invention,
fig. 7 is a view corresponding 1;o fig. 6 of a fourth
embodiment of the screw compressor according to the invention,
and
fig. 8 is a view corresponding to fig. 6 of a fifth
embo~iment of the screw compressor according to the invention.
In figs. 1, 2 and 3 a rotary screw compressor is shown
comprising a casing l, a male rotor 6 and a female rotor 18
W095/02767 ~ l i i 2 1 ~ 6 2 9 4 PCT~L93/00150
-- 6
cooperating therewith enclosed in a working space defined by
the casing. The casing has a outlet port 2 and a discharge pipe
4 at the high pressure end of the working space and a suction
pipe 3 at the low pressure end of the working space. Arrow A
indicates the direction of the gas to be compressed. Arrow B
indicates the direction of the discharge of the compressed gas.
Arrow ~ indicates the rotation of the male rotor 6 which can
be driven through drive means not shown in the drawings.
The malerotor 6 isrotatably supported through a bearing
10 at its high pressure end and a bearing bracket 11 at its
low pressure end. The bearing bracket 11 is fixed on a detachable
end cover 5 of the casing 1 and projects into an internal cavity
in the low pressure end of the male rotor 6, thereby forming
a first chamber 9 therebetween.
As can be seen in fig. 1 the cavity and the bearing
bracket 11 inside the cavity extend over a significant part
of the length of the male rotor 6. Therefore the distance between
the bearings 10, 11 at opposite ends of the rotor 6 is
comparatively small, as a result of which the radial forces
on the rotor can be better supported through the bearings and
only a small radial deflection of the rotor will occur.
The low pressure end face of the male rotor 6 is provided
with a protruding ring shoulder 15 having a cylindrical outer
surface 16. A sealing means 7 between the male rotor 6 and the
casing is provided at the high pressure end and a sealing means
8 is provided between the shoulder 15 and the casing 1 at the
low pressure end.
The bearing bracket 11 has a substantially cylindrical
circumferential outer surface, the surface being provided with
two longitll~;nAl grooves 25, exten~;ng parallel to the
longitll~; n~ l axis of the bearing bracket, and with a recess
13. The recess 13 is an essentially rectangular cutout formed
at a distance from the substantially circular end face of the
bearing bracket 11 and is connected to an oil drainage channel
12 through an opening 14. As can been seen in fig. 2 the recess
13 is located on the side of the bearing bracket 11 radially
opposite the outlet port2 for reasons explained further below.
~ W095/02767 ~ f`j~ t ~ 2 1 6 6 2 9 4 PCT~L93/00150
-- 7
The longitudinal grooves 25 are located at either side of the
recess 13 seen in circumferential direct.ion. Each longitudinal
groove 25 is connected to an oil feed channel 27 provided in
the bearing bracket 11 through a number of openings 29 uniformly
distributed along the length of each g:roove. As can be seen
in fig. 3 the longitudinal grooves 25 ter-minate at the end face
of the bearing bracket 11 to provide communication between each
groove 25 and the space formed between the end face of the
beari.ng bracket and the bottom of the cavity in the male rotor
6.
At the low pressure end of the male rotor 6 a second
chamb~er 17 is formed by the annular end face of the ring shoulder
15, sealing means 8, the bearing bracket 11 and the end cover
5. The chamber 17 is connected to oil feed channels 27 through
openi.ngs 35.
The female rotor 18 is at its low pressure end rotatably
supported in a manner similar to the ma.le rotor 6. A bearing
brack.et 20 projects into an internal cavity provided in the
rotor 18 forming a first chamber 19 therebetween. The bearing
brack.et 20 is mounted on the side cover 5. The substantially
cylindrical outer surface of the bearing bracket 20 is provided
with a recess 22 and two longit~l~;n~l grooves 24 located at
eithe.r side of the recess 22. The recess 22 is connected to
an oil drainage channel 21 through an opening 23. The recess
22 is an essentially rectangular cutout and terminates at the
end face of the bearing bracket 22. The longittl~;n~l grooves
24 are located at a distance from the end face of the bearing
brack.et 20 and extend towardsthe low pressure end. Each longitu-
dinal groove 24 is connected to an oil feed channel 26 through
a nunlber of openings 28 uniformly disposed along the length
of the groove.
The low pressure end of the female rotor 18 is provided
with a protruding ring shoulder 31 having a cylindrical outer
surfalce 32. A sealing means 30 is provide.d between the shoulder
31 and the end cover 5 at the low pressure end of the female
rotor 18.
At the low pressure end of the fe~male rotor 18 a second
WO95l027C7 ; ~ 2 i 6 6 2 9 4 PCT~L93/OnlSO~
chamber 33 is formed by the annular end face of the ring shoulder
31 of the rotor, sealing means 30, the bearing bracket 20 and
the end cover 5. The chamber 33 is connected to oil feed channels
26 through openings 34.
The length of the bearing bracket ll of the male rotor
6 projecting into the male rotor is less than the length of
the bearing bracket 20 of the female rotor 18 projecting into
the female rotor. This is indicated by the distance "l" in fig.
3. Also, the length of the recess 13 is less than that of recess
22, both recesses having an approximate maximum length of 0.7
times the length of the corresponding bearing bracket.
Fig. 4 shows a cross section of the bearing bracket 11
of the male rotor 6. As can be seen the recess 13 is essentially
a flat portion formed on the cylindrical outer circumferential
surface of the bearing bracket 11. The recess 13 communicates
with the central oil drainage channel 12 through the opening
14. Each ~Loove 25 is connected to an oil feed chAnnel 27 through
a number of openings 29 to reduce the flowresistance of the
oil feed. The longitl~inAl grooves 25 at either side of the
recess 13 are formed such that their side edges adjacent the
recess 13 are located in a common first plane passing through
the longit~ l axis of the bearing bracket 11 and at an equal
distance from the recess 13. The other longitudinal edges of
the grooves 25 are each located in a second and third plane
through the axis of the bearing bracket respectively. The second
and third plane each being inclined at an angle ~, preferably
equal or less than 45, to the first plane. This embodiment
of the bearing bracket provides optimal conditions for a
combination of hydLod~namic and hydrostatic radial load bearing
capabilities and an excellent radial stiffness of the bearing
arrangement. The bearing bracket 20 of the female rotor 18 has
a cross section substantially similar to that of the bearing
bracket 11 of the male rotor. In an alternative embodiment not
shown in the drawings the location of the oil feed grooves at
either side of the recess on the bearing bracket can be adapted
e.g. for supporting a lower radial load on the corresponding
rotor. In this case the grooves could be located closer to each
2 1 6 6 2 9 4
_ W095/02767 PCT~93/OOlS0
-
_ g
other, therefore a smaller zone in the first having a high oil
pressure is obtained.
A second embodiment of the compr,_ssor according to the
invention is shown in fig. 5. The compressor is provided with
S bearing brackets 11, 20 for the male rot:or 6' and female rotor
18' respectively, the bearing brackets being similar to the
bearing brackets described hereinbefore. A sealing means 56
is provided between the bearing bracket:11 and the male rotor
6'. Towards the low pressure end of the compressor a rolling
contact bearing 57, such as a ball bearing, is mounted between
the male rotor 6' and the bearing bracket 11. A sealing means
58 is provided between the bearing bracket 20 and the female
rotor 18'. Towards the low pressure end of the compressor a
rolling contact bearing 59, such as a ball bearing, is mounted
between the female rotor 18' and the bearing bracket 20. This
embodiment is particularly advantageous for screw compressors
operating with cooling oil injected intol:he gas to be compressed
in the working space of the compressor.I'hese screw compressors
operate at low speed comparedwithoil-free ("dry") compressors
and have small clearances between the rotor teeth, and between
the rotors and the casing. Therefore rolling contact bearings
in general having smaller clearances than bearing brackets are
preferred. The sealing means 56, 58 can]be provided in the form
of a flow obstruction having a smaller clearance than the
clearance between the rotor and the bearing bracket. As can
be seen in fig. 4 no sealing means are provided between the
second chambers 60, 61 and the working~ space.
In the embodiment shown in fig. 6 the bearing brackets
11 and 20 of the male and female rotor respectively have their
oil feed channels 26, 27 connected to a common source 38, e.g.
an oil pump, for supplying pressurized oil as indicated by arrow
k. The oil drainage channels 12, 21 of the respective bearing
brackets 11, 20 are connected to an oil collector 39. The
collector 39 is vented to the atmosphere as indicated by the
arrow M. In this embodiment the source 3~ is designed to supply
the oil at a pressure approximately equal to the pressure of
the gas to be compressed. This embodiment is preferred for screw
WO 95/02767 ~ f~ ~ 1` C 2 ~ 6 62 9 4 PCT/NL93/00150~
-- 10 --
compressors wherein the compressed gas has to be free of oil.
Since the pressure in the chambers 17, 33 (fig. 3) approximates
the pressure in the suction pipe 3 the loads on the sealing
means 8, 30 are limited. As the oil drainage channels 12, 21
are in open communication with the atmosphere the oil collector
39 can be of a simple design.
In the embodiment shown in fig. 7 the bearing brackets
11 and 20 of the male and female rotor respectively have their
oil feed channels 26, 27 connected to a source 38 for supplying
pressurized oil as indicated by arrow k. The oil drainage
channels 12, 21 of the respective bearing brackets 11, 20 are
connected to an oil collector 40. The collector 40 is connected
to the suction pipe 3 to maintain a pressure in the collector
40 equal to the pressure of the gas to be compressed.
In the embodiment shown in fig. 8 the bearing brackets
11 and 20 of the male and female rotor respectively have their
oil feed channels 26, 27 connected to an oil separator 41 for
supplying pressurized oil as indicated by arrow m. The oil
drainage channels 12, 21 of the respective bearing brackets
11, 20 are connected to the suction pipe 3 of the compressor
as indicated by arrow n. The oil will then pass through the
compressor along with the gas to be compressed resulting in
a cooling of the gas during compression. The discharge pipe
4 of the compressor is connected to the oil seperator 41 where
the oil and the compressed gas are separated. This embodiment
of the compressor is preferred if the presence of oil in the
compressed gas is allowed.
The rotary screw compressor according to the invention
operates as follows.
The gas to be compressed enters the suction pipe 3 (fig.
1). The male rotor 6 is rotated at a speed ~ by means of an
external drive acting on the male rotor 6. The gas to be
compressed is entrained and compressed in chambers limited by
the rotor teeth and the casing. During the compression of the
gas a force F, resulting from the differential pressure between
the discharge pipe 4 and the suction pipe 3, acts on the rotors
as is indicated in fig. 2. This force F is composed of radial
~ W095/02767 ~ 2 1 6 6 2 9 4 PCT~L93/00150
-- 11 --
forces Fl, F2 and axial forces F3, F4 acting on the rotors 6 and
18. These forces must be supported by the bearing arrangements
of the rotors.
To counteract these forces F~-F4 pressurized oil is fed
5 through the oil feed channels 26, 27 (arrows D and H in fig.
3), t:he openings 28, 29, and the longitll~;n~l grooves 24, 25
of the bearing brackets 11, 20 and enters the chambers 9, 19
between each bearing bracket and the corresponding rotor. The
pressurized oil is drained from chamber 9, 19 through the recess
13, 22 provided on the bearing bracket, each recess being
connected to an oil drainage channel 12, 21 by an opening 14,
23 (arrows K and E in fig. 3).
The maximum length of the recesses 13, 22, which is
approximately 0.7 timesthelengthofthe corresponding bearing
bracket, is preferred in this embodiment as there must be a
cylindrical section of the bearing brac:ket having sufficient
dimensions present inside the cylindrical cavity in each rotor
near the low pressure end thereof to provide a restriction
between the chamber 17, 33 and the recess 13, 22, respectively.
The presence of pressurized oil in the first chambers
between the rotors and the bearing brackets gives rise to radial
lift~;ng forces Fs and F6 (fig. 2) acting on the rotors 6, 18
~ ~Lively. The position of each recess on the bearing bracket,
radially opposite the outlet port 2, as shown in fig. 2,
faci~itates obtaining a balance between the forces F5, F6 and
the forces Fl, F2. As a result of the location of the longit1l~;n~l
grooves 24, 25 a pressure zone is obtained, the pressure
difference in this zone being equal to the pressure difference
between the oil feed channels and the oil drainage channels.
The dimensions of the recesses 13, 22, the location and
dimensions of the longit1~in~1 grooves 24, 25, and the pressure
levels in the oil feed channels as well as in the oil drainage
cham~els depent on the desired characteristics of the rotary
screw compressor. They are chosen such that the forces F5 and
F6 compensate the major part of the forces Fl, F2 respectively.
The remaining part of each of the forces F~and F2 is supported
through the bearing 10 at the high pressure end of each rotor
W095/02767 ~ c 2 ~ 6 6 2 9 4 PCT~L93/00150~
- 12 -
(bearing lO of the female rotor 18 not shown in the drawings).
As a result of the geometry of the rotors defined by
the toothing thereof the radial force Fl is in most cases less
than the radial force F2. Therefore there is a difference in
lenght between the bearing bracket 11 and/or recess 13 of the
male rotor 6 and the length of the bearing bracket 20 and/or
recess 22 of the female rotor 18. This is indicated in fig.
3 by distance "l".
As a result of pressurized oil being fed into the axial
lOchambers 17, 33 at the low pressure end of the rotors 6, 18
respectively, axial forces F7, F8 (fig. 3) are exerted on the
rotors opposing the axial forces F3 and F4 resulting from the
compression of the gas. The axial forces F7, F8 compensate a
part of the forces F3 and F4. The remaining part of the forces
15F3 and Fi is compensated through the bearings lO of the rotors.
Due to the geometry of the rotors the axial force F3 on
the male rotor 6 is as a rule larger than the axial force F4
on the female rotor 18. To compensate this difference an
additional axial force F9 is exerted on the male rotor 6.
20According to the invention the longitll~inAl grooves 25
terminate at the end face of the bearing bracket to provide
an open communication betweenthegrooves25 andthe space formed
between the end face of the bearing bracket 11 and the bottom
of the chamber 9 of the male rotor 6. As can be seen in figs.
251-3 the passage of oil from this space towards the recess 13
is obstructed, whereby the oil pressure is maintained in this
part of the chamber 9. This results in the axial force F9, which
is exerted on the rotor 6. At the same time the axial force
F4 on the female rotor 18 will be smaller than the force F3 and
30since the grooves 24 on the bearing bracket lO are not in open
communication with that part of the chamber 19 no additional
axial force is exerted on the female rotor. As the recess 22
terminates at the end face of the bearing bracket, the recess
22 is in open communication with the bottom part of the chamber
3519, so that a built-up of oil pressure therein that is prevented.
The provision of bearing brackets at the low pressure
ends of the rotors, which brackets project into internal
WO 95l02767 ~ ~- 2 1 6 6 2 ~ 4 PCT/NL93/00150
-- 13 --
essentially cylindrical cavities provided in the rotors and
extend over a significant part of the lenght of rotors, results
in a bearing arrangement having an excellent stiffness and
capahle of supporting high radial loads on the rotors. In
S combination with the comparatively smal;l distance between the
- bearings at opposite ends of each rotor the deflection of the
rotors resulting from the gas pressure is even further reduced.
The bearing arrangement according to t:he invention is also
capable of counteracting the axial forces on the rotors without
10 having to provide complex additional tlhrust bearings.
The bearing arrangement of the rotary screw compressor
according to the invention permits a co~siderable increase of
the radial and axial forces over existing bearing arrangements,
resulting in an increase of the allowable differential pressure
15 and discharge pressure of the screw compressor.
