lubricants
Article
Specific Features of Aerodynamic Journal Bearings
with Elastically Supported Pads
Jiří Šimek
Centre of Industrial Research Techlab Ltd., 190 00 Prague, Czech Republic; [email protected];
Tel.: +420-607-933-682
Academic Editors: Ron A.J. van Ostayen and Daniel J. Rixen
Received: 7 March 2017; Accepted: 6 April 2017; Published: 11 April 2017
Abstract: Aerodynamic bearings with elastically supported tilting pads have operational properties
comparable with widely-used foil journal bearings. They combine the excellent stability of tilting pad
bearings, as a result of very small cross-coupling stiffness terms, with the positive properties of foil
bearings, namely their ability to adapt to changing operating conditions and presence of additional
damping due to friction between elastic members and bearing casings. Air cycle machines (ACMs)
are used in the environmental control systems of aircrafts to manage the pressurization of the cabin.
An ACM with the abovementioned type of bearings and an operational speed of 60,000 rpm was
designed and successfully tested, even under conditions of strong external excitation. Some problems
with rotor stability in certain operation regimes were encountered. Rotor relative vibrations measured
at both bearing locations increased substantially when excitation frequency was close to the lowest
rotor eigenvalues. In spite of that and the 1000 start/stop cycles passed by the end of the test, any
traces of wear on the bearing sliding surfaces were negligible. When the bearing distance had to be
shortened in order to insert the machine into the defined space, the rotor quickly became unstable at
relatively low speeds. Although rotor stability reserve was reduced only slightly, the rotor had to
be redesigned in order to achieve stability. Operation characteristics of aerodynamic bearings with
elastically supported tilting pads are presented together with rotor dynamic analysis and validated
with measured results.
Keywords: aerodynamic bearings; elastically supported tilting pads; air cycle machine (ACM); rotor
dynamics; rotor stability; measured relative rotor vibration; external excitation
1. Introduction
Air cycle machines (ACMs) are used in the environmental control systems of aircrafts to manage
the pressurization of the cabin. To achieve small dimensions and high efficiency, the rotor should
operate at a very high speed. ACM design is similar to that of a turbocharger, i.e., the rotor is driven
by a turbine and pressurized air is provided by a blower impeller. The majority of contemporary
high-speed rotating machines use aerodynamic foil bearings, which have very good operational
properties [1–6]. However, even better functional properties can be achieved with tilting pad journal
bearings (TPJBs), which have excellent dynamic characteristics. Cross-coupling stiffness terms, which
are known to promote rotor instability, are found in TBJBs two orders lower than the principle stiffness
terms. Rotors in TPJBs are therefore principally stable, unless there is a strong external source of
excitation, such as a high-pressure labyrinth seal. With foil bearing working gap geometry, it is not
possible to achieve such small values of cross-coupling stiffness terms (e.g., [7]), and this is why TPJBs
are a better solution from the standpoint of rotor stability. A new type of aerodynamic bearing with
elastically supported pads, which combines the advantages of foil bearings and tilting pad journal
bearings, was designed and successfully used for ACM rotor support.
Lubricants 2017, 5, 10; doi:10.3390/lubricants5020010
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2. Materials
and5,Methods
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Aerodynamic
pad journal bearing characteristics were designed with help of the computer
2.
Materials andtilting
Methods
program code “DATPJB.” Standard bearing configuration covers three pads with static load acting
Aerodynamic tilting pad journal bearing characteristics were designed with help of the
symmetrically between the two lower pads. The program code “DATPJB” provides the solution of
computer program code “DATPJB.” Standard bearing configuration covers three pads with static
the Reynolds
equation
for compressible
according
toprogram
[8]. In the
process
of calculation,
it was
load acting
symmetrically
between themedium
two lower
pads. The
code
“DATPJB”
provides the
necessary
to
find
the
equilibrium
of
the
pad
position
and
static
position
of
the
journal
in
the
bearing.
solution of the Reynolds equation for compressible medium according to [8]. In the process of
Dynamic
stiffness
and necessary
damping to
coefficients
were calculated
harmonic
journal
movement
calculation,
it was
find the equilibrium
of the for
padsmall
position
and static
position
of the in
the vicinity
position,
whichand
enabled
us tocoefficients
also determine
stability
the pad
journal of
in the
the equilibrium
bearing. Dynamic
stiffness
damping
were the
calculated
forofsmall
harmonic
journal
movement
in the
vicinity of the
equilibrium
position,
which
to also
movement.
The
computer
program
“AXSPDR4,”
based
on principles
stated
in enabled
[9], wasus
used
for the
determine
the stability spiral
of thegroove
pad movement.
The computer
“AXSPDR4,”
onloss
design
of the aerodynamic
thrust bearings,
providingprogram
bearing load
capacity,based
friction
principles
stated
[9], was of
used
for the
of the aerodynamic spiral groove thrust bearings,
and film
stiffness
as ainfunction
speed
anddesign
film thickness.
providing
load capacity,
friction
loss
andthe
film
stiffness“DYNROT-R,”
as a function ofwhich
speedenable
and film
Dynamic bearing
rotor analysis
was carried
out
with
program
us to
thickness.
calculate critical speeds, response to unbalance and pertinent stability reserve. The program involved
Dynamic rotor analysis was carried out with the program “DYNROT-R,” which enable us to
graphic superstructure, enabling easy generation of the rotor model. The program also provided
calculate critical speeds, response to unbalance and pertinent stability reserve. The program involved
graphic outputs, such as Campbell diagrams or graphic representations of rotor response to unbalance
graphic superstructure, enabling easy generation of the rotor model. The program also provided
in selected
computer
programs
are available
according
to standardofcommercial
conditions.
graphicsections.
outputs,All
such
as Campbell
diagrams
or graphic
representations
rotor response
to
Relative
rotor
vibrations
during
tests
were
measured
by
means
of
eddy
current
sensors
unbalance in selected sections. All computer programs are available according to standard
Micro-epsilon
S04
(Micro-Epsilon Messtechnik, Ortenburg, Germany) connected to units Micro-epsilon
commercial
conditions.
eddyNCDT
3300
(Micro-Epsilon
Messtechnik,
Ortenburg,
Germany)
in nanometers
Relative rotor vibrations during
tests were
measured by
means ofwith
eddyresolution
current sensors
Microand bandwidth
up to 100 kHz.Messtechnik,
Vibration signals
were Germany)
recorded by
National
measuring
epsilon S04 (Micro-Epsilon
Ortenburg,
connected
toInstrument
units Micro-epsilon
Messtechnik,
Ortenburg,
Germany)
with
incard
nanometers
USB eddyNCDT
DAQ card 3300
with(Micro-Epsilon
16 analog inputs
and sampling
frequency
up to
1.2 resolution
MHz. The
operation
and bandwidth
up program
to 100 kHz.
VibrationNI,
signals
were recorded
by National of
Instrument
measuring
was controlled
by the
ScopeWin
enabling
graphic processing
measured
signals and
USBelaboration
DAQ card with
16 analog data,
inputse.g.,
and evaluating
sampling frequency
up toof1.2measured
MHz. Thedata
cardor
operation
statistic
of measured
RMS values
providing
was
controlled
by
the
program
ScopeWin
NI,
enabling
graphic
processing
of
measured
signals
and
frequency spectra.
statistic elaboration of measured data, e.g., evaluating RMS values of measured data or providing
External excitation was realized by mounting the ACM on a vibration table, which could provide
frequency spectra.
excitation either in horizontal or in vertical directions. Maximum acceleration of the table was 2g
External excitation was realized by mounting the ACM on a vibration table, which could provide
(~20 excitation
m·s−2 ). either in horizontal or in vertical directions. Maximum acceleration of the table was 2g (~20
m·s−2).
3. Results
3. Results
3.1. Aerodynamic Bearings with Elastically Supported Tilting Pads
3.1.
Aerodynamic
Bearings
Elastically
Supported
Tilting
Padsthe simplest of which is illustrated in the
There
are several
patentswith
covering
bearing
design
[10,11],
left-hand There
section
Figurepatents
1. Tilting
pads 2bearing
are supported
by elastic
with the
shape-enabling
areofseveral
covering
design [10,11],
the members
simplest of4 which
is illustrated
in
free pad
tilting in section
a peripheral
direction.
The
detail
of supported
the pad elastic
support
is in the
centre
Figure 1.
the left-hand
of Figure
1. Tilting
pads
2 are
by elastic
members
4 with
theofshapeThe pad
is held
position
pin 3, which
alsoThe
enables
adjustment
basic bearing
enabling
freeinpad
tilting by
in athe
peripheral
direction.
detailthe
of the
pad elasticofsupport
is in the clearance
centre
of Figure
1. The
is heldbearing
in position
by the pinenabling
3, which easy
also enables
adjustment
of basicwear,
bearing
by means
of the
nutpad
5. Basic
clearance,
run–upthe
without
excessive
is the
clearance
by
means
of
the
nut
5.
Basic
bearing
clearance,
enabling
easy
run–up
without
excessive
most important advantage of TPJBs in comparison with foil bearings.
wear, is the most important advantage of TPJBs in comparison with foil bearings.
Figure
1. Two
types
TPJBsaccording
according to
specifications.
1–bearing
casing,casing,
2—pad,2—pad,
3—
Figure
1. Two
types
ofofTPJBs
toCzech
Czechpatent
patent
specifications.
1–bearing
pin,
4—elastic
member,
5—nut,
6—insert,
7—screw.
3—pin, 4—elastic member, 5—nut, 6—insert, 7—screw.
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2017, 5, 10 with TPJBs of the simplest design may occur due to uneven deformation
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Some problems
elastic member, which can result in two-point contact restraining free pad movement. That is why
Some problems with TPJBs of the simplest design may occur due to uneven deformation of the
a more sophisticated design with defined geometry of contact between the pad and its supporting
elastic member, which can result in two-point contact restraining free pad movement. That is why a
member was devised. Its design is shown in the right-hand section of Figure 1. The pad 2 is supported
more
sophisticated design with defined geometry of contact between the pad and its supporting
Lubricants 2017, 5, 10
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by the
insert
6 with
a curved
surface,
on in
which
the pad section
can freely
tilt in
circumferential
and axial
member
was
devised.
Its design
is shown
the right-hand
of Figure
1. The
pad 2 is supported
directions.
The
supported
the
elastic
4; basic
clearanceand
can
be set by
by the insert
6problems
with6aiscurved
surface,
which
themember
pad
canoccur
freely
tilttobearing
in
circumferential
axial
Someinsert
with
TPJBs
of by
theon
simplest
design
may
due
uneven
deformation
of the
the screw
7.
Dynamic
bearing
properties
are
very
good
due
to
the
very
small
crosscoupling
stiffness
directions.
The insert
6 is can
supported
the elastic
member
4; basicfree
bearing
clearance can
beisset
by athe
elastic member,
which
result inby
two-point
contact
restraining
pad movement.
That
why
screw
7. Dynamic
bearing
properties
are very
good
due
to the
very small
crosscoupling
stiffness
terms,
which
is a property
inherent
to tilting
pad
bearings.
Thanks
totheelastic
members
the bearings
more
sophisticated
design
with defined
geometry
of
contact
between
pad
and
its supporting
member
was
Its design
shown
in the
right-hand
section
of Figure
1. The
pad
2 isthe
supported
terms,
which
is devised.
a property
inherent
to tilting
pad
bearings.
Thanks
to
members
bearings
also acquire
properties
found
in foilisbearings,
i.e.,
the
adaptability
to elastic
changing
operating
parameters
the insert
6 withdue
afound
curved
surface,
which
the
can
tilt
in circumferential
and
axial body.
alsoby
acquire
properties
foil bearings,
i.e.,members
the pad
adaptability
changing
operating
and additional
damping
to in
friction
ofonelastic
onfreely
thetoinner
surface
of theparameters
bearing
directions.
The
insert 6 due
is supported
byof
theelastic
elasticmembers
member 4;
bearing
clearance
canbearing
be set bybody.
the
and
additional
damping
friction
onbasic
the
inner surface
of the
Dynamic
characteristics
oftothe
bearings
used in ACM,
calculated
by the
method
described in
screw
7.
Dynamic
bearing
properties
are
very
good
due
to
the
very
small
crosscoupling
stiffness
Dynamic characteristics of the bearings used in ACM, calculated by the method described in
Section terms,
4, are presented
in Figure
2. Bearings
mm
in diameter
with
l/Dmembers
ratio ofthe
0.72
were loaded
is a property
inherent
to tilting25
pad
bearings.
Thanks
to elastic
bearings
Section 4, which
are presented
in Figure
2. Bearings
25
mm
in diameter
with
l/D ratio
of 0.72 were
loaded
with 4.4also
N at
the
turbine
(TB)
and
4.9
N
at
the
blower
(DB)
side,
respectively.
acquire properties found in foil bearings, i.e., the adaptability to changing operating parameters
with 4.4 N at the turbine (TB) and 4.9 N at the blower (DB) side, respectively.
and additional damping due to friction of elastic members on the inner surface of the bearing body.
Dynamic characteristics of the bearings used in ACM, calculated by the method described in
Section 4, are presented in Figure 2. Bearings 25 mm in diameter with l/D ratio of 0.72 were loaded
with 4.4 N at the turbine (TB) and 4.9 N at the blower (DB) side, respectively.
Figure 2. Stiffness and damping coefficients of air cycle machine (ACM) bearings.
Figure 2. Stiffness and damping coefficients of air cycle machine (ACM) bearings.
The diagrams in Figure 2 show principle bearing stiffness and damping coefficients for the
Figure
2. Stiffness
and 1.
damping
coefficients
of airstiffness
cycle
machine
coordinate
system
in Figure
Cross-coupling
stiffness
terms
Kand
xy,(ACM)
Kyxdamping
as bearings.
well as coefficients
damping terms
The diagrams
inshown
Figure
2 show
principle
bearing
for the
B
xy, Bxy are zero.
coordinate system shown in Figure 1. Cross-coupling stiffness terms K , K as well as damping
xy coefficients
yx
The diagrams in Figure 2 show principle bearing stiffness and damping
for the
terms Bxy
, Bxy aresystem
zero. shown in Figure 1. Cross-coupling stiffness terms Kxy, Kyx as well as damping terms
coordinate
3.2. Rotor Dynamic Analysis
Bxy, Bxy are zero.
ACM with
aerodynamic TPJBs, designed in 2012 [12], is shown in Figure 3. The rotor 1 is
3.2. RotorAn
Dynamic
Analysis
3.2. Rotor
Analysis
supported
byDynamic
two TPJBs
with elastically supported pads 3. Axial forces are taken up by aerodynamic
An ACM
with
aerodynamic
TPJBs,
designed
in 2012 [12],
is impellers
shown inare
Figure
3. overhang
The rotor 1 is
spiral
groove
thrust
bearings
4, 5.
Turbine
(left) and
(right)
An ACM with aerodynamic
TPJBs, designed
in blower
2012 [12],
is shown
in Figure at
3. rotor
The rotor
1 is
supported
by two
TPJBsvibrations
with elastically
supported
pads
3. Axial
forces aresensors
taken up
by aerodynamic
ends.
Rotor
relative
are observed
by two
pairs
of eddy
9, which
have
supported
by two TPJBs
with elastically
supported
pads
3. Axial
forcescurrent
are taken up by aerodynamic
spiralmeasuring
groove
thrust
bearings
4,
5.
Turbine
(left)
and
blower
(right)
impellers
are
at
rotor
overhang
range
of
0.5
mm
and
typical
sensitivity
of
28
mV/µ
m.
spiral groove thrust bearings 4, 5. Turbine (left) and blower (right) impellers are at rotor overhang ends.
Rotor relative
vibrations
observed
two pairs
of eddy
current
sensors
9, which
havehave
measuring
ends. Rotor
relativeare
vibrations
are by
observed
by two
pairs of
eddy current
sensors
9, which
range
of 0.5 mm
and typical
of 28 mV/µ m.
range ofmeasuring
0.5 mm and
typical
sensitivity
of sensitivity
28 mV/µm.
Figure 3. Air cycle machine design. 1—rotor, 2—casing, 3—pad, 4,5—thrust bearings, 6—elastic
member, 7—pin, 8—nut, 9—sensor.
Figure
Air cycle
machine
design.1—rotor,
1—rotor, 2—casing,
2—casing, 3—pad,
4,5—thrust
bearings,
6—elastic
Figure 3.
Air 3.cycle
machine
design.
3—pad,
4,5—thrust
bearings,
6—elastic
member, 7—pin, 8—nut, 9—sensor.
member, 7—pin, 8—nut, 9—sensor.
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Dynamic
analysis of the rotor showed that there are four critical speeds associated with the rigid
Dynamic analysis of the rotor showed that there are four critical speeds associated with the
rotorrigid
on bearing
film, which
from 7,000
to 15,500
rpm.rpm.
The The
firstfirst
bending
critical
rotor onaerodynamic
bearing aerodynamic
film, ranged
which ranged
from 7,000
to 15,500
bending
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speed
of co-rotating
precession
is higher
thanthan
200,000
rpm,
beseen
seenfrom
from
potential
critical
speed of co-rotating
precession
is higher
200,000
rpm,asascan
can be
the the
potential
intersection
of the
redred
lineline
with
thethe
speed
line
inin
the
4. The
The branch
branchofof the
intersection
of the
with
speed
line
theCampbell
Campbelldiagram
diagram in
in Figure
Figure 4.
Dynamic analysis of the rotor showed that there are four critical speeds associated with the rigid
the
same
eigenvalue
with
counter-rotating
(backward)
precession
(light
blue
line)
is
not
excited
by by
same eigenvalue with counter-rotating (backward) precession (light blue line) is not excited
rotorunbalance
on bearing
aerodynamic
film,
which
ranged
from
7,000
to
15,500
rpm.
The
first
bending
critical
therefore
it is
notconsidered
considered as aa critical
unbalance andand
therefore
it is
not
criticalspeed.
speed.
speed of co-rotating precession is higher than 200,000 rpm, as can be seen from the potential
intersection of the red line with the speed line in the Campbell diagram in Figure 4. The branch of the
same eigenvalue with counter-rotating (backward) precession (light blue line) is not excited by
unbalance and therefore it is not considered as a critical speed.
Figure
ofcritical
criticalspeeds.
speeds.
Figure4.4.Campbell
Campbell diagram
diagram of
As can
be seen
in in
Figure
5,5,real
fourrelevant
relevanteigenvalues
eigenvalues
negative,
so that
As can
be seen
Figure
realparts
parts of
of the
the four
areare
negative,
so that
the the
rotorrotor
operation
is stable.
Rotor
resistance
isusually
usuallydetermined
determined
means
of logarithmic
operation
is stable.
Rotor
resistanceto
toinstability
instability is
by by
means
of logarithmic
Figure
4.
Campbell
diagram
of
critical
speeds.
decrement,
designated
as as
loglog
δ, δ,defined
However,totomake
make
situation
more
obvious,
decrement,
designated
definede.g.,
e.g., in
in [13].
[13]. However,
thethe
situation
more
obvious,
we can
thethe
stability
reserve
as aapercentage
percentagebyby
following
equation:
we define
can define
stability
reserveofofthe
the rotor
rotor as
thethe
following
equation:
As can be seen in Figure 5, real parts of the four relevant eigenvalues are negative, so that the
rotor operation is stable. Rotorχ resistance
to instability
determined
by means of logarithmic
= log δ × 100/π=
−200 × is
Reusually
(λ)/Im (λ)
(%),
(1)
decrement, designated as log δ, defined e.g., in [13]. However, to make the situation more obvious,
where
Re (λ)
and
Im (λ) represent
the
real
and as
imaginary
part ofby
eigenvalue
respectively.
we can
define
the
stability
reserve of
the
rotor
a percentage
the following
equation:
Figure 5. Damping of the four lowest eigenvalues.
χ = log δ × 100/π= −200 × Re (λ)/Im (λ) (%),
(1)
Figure5.5.Damping
of the
the four
eigenvalues.
Figure
of
fourlowest
lowest
eigenvalues.
where Re (λ) and Im (λ) represent
theDamping
real and
imaginary
part
of eigenvalue respectively.
The stability reserve of the two lowest eigenvalues at operation speed of 60,000 rpm are 10% and
The stability reserve of the two lowest eigenvalues at operation speed of 60,000 rpm are 10%
13.5% respectively,
which correspond
logarithmic
decrement
0.31 and 0.42. Such values(1)of
χ = log δ × to
100/π=
−200 × Re
(λ)/Im (λ)of
(%),
and 13.5% respectively, which correspond to logarithmic decrement of 0.31 and 0.42. Such values of
stability reserve are in most cases quite sufficient. The results of rotor response to unbalance were
stability
are(λ)
in represent
most cases the
quitereal
sufficient.
The results
of rotor
response torespectively.
unbalance were also
where
Re (λ)reserve
and Im
and imaginary
part
of eigenvalue
also quite
quitesatisfactory,
satisfactory,
with
only
a slight
increase
of vibration
amplitude
inaround
the region
around the
with
only
a
slight
increase
of
vibration
amplitude
in
the
region
the “bearing”
The stability reserve of the two lowest eigenvalues at operation speed of 60,000 rpm
are 10% and
“bearing”
critical
speed
from
8000 to
18,000
rpm, i.e.,
outside
standard
operating range.
critical
speed from
8000
tocorrespond
18,000
rpm,
i.e.,
outside
standard
operating
13.5%
respectively,
which
to logarithmic
decrement
ofrange.
0.31 and 0.42. Such values of
stability
reserve
are in most cases quite sufficient. The results of rotor response to unbalance were
3.3. Operation
Tests
also quite satisfactory, with only a slight increase of vibration amplitude in the region around the
Operation
tests
were
carried
in a test
loop
with
the turbine
driven
by pressurized
air, as in
“bearing”
critical
speed
from
8000out
to 18,000
rpm,
i.e.,
outside
standard
operating
range.
real conditions [14,15]. The machine could be operated up to 68,000 rpm with minimum values of
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3.3. Operation Tests
Operation tests were carried out in a test loop with the turbine driven by pressurized air, as in real
conditions [14,15]. The machine could be operated up to 68,000 rpm with minimum values of relative
rotor vibrations. As shown by the amplitude-frequency characteristics depicted in Figure 6, the5RMS
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values of vibration amplitudes were measured in two different tests. The RMS values of vibration in
radial direction
indirection
both testsinwere
than 2lower
µm (double
than 5lower
µm). The
vibration
in radial
bothlower
tests were
than 2 µamplitude
m (doublelower
amplitude
thansensor
5 µ m).at
the sensor
bloweratside
the horizontal
direction
wasdirection
damagedwas
during
tests, and
therefore
its signal
was its
not
The
thein
blower
side in the
horizontal
damaged
during
tests, and
therefore
included
into
the
right
diagram.
signal was not included into the right diagram.
Figure
Figure6.6.RMS
RMSvalues
valuesofofrotor
rotorvibration
vibrationfrom
fromthe
the first
first and
and second
second test.
test. TB/DB—turbine/blower
TB/DB—turbine/blowerside,
side,
vert./hor.—vertical/horizontal
direction.
vert./hor.—vertical/horizontal direction.
Typical records of rotor run-up and run-down are illustrated in Figure 7. The vibration signal in
Typical records of rotor run-up and run-down are illustrated in Figure 7. The vibration signal
axial direction is distorted by damaged measured surface, so that only quasi-static shifts can be
in axial direction is distorted by damaged measured surface, so that only quasi-static shifts can be
evaluated. Quasi-static shifts in radial directions are caused by the change of rotor axis due to thrust
evaluated. Quasi-static shifts in radial directions are caused by the change of rotor axis due to thrust
runner alignment according to the position of thrust bearing sliding surface. It is more evident from
runner alignment according to the position of thrust bearing sliding surface. It is more evident from
the rotor run-down in the right section of Figure 7, where the change of axial force direction results
the rotor run-down in the right section of Figure 7, where the change of axial force direction results in
in relatively big deviations of the rotor in a radial direction. This phenomenon is caused by relatively
relatively big deviations of the rotor in a radial direction. This phenomenon is caused by relatively
low stiffness of pad supports in comparison with thrust bearing film stiffness. This by-product of
low stiffness of pad supports in comparison with thrust bearing film stiffness. This by-product of
elastic pad support contributes to a significant increase of thrust bearing load capacity. The average
elastic pad support contributes to a significant increase of thrust bearing load capacity. The average
coefficient p/u (specific load, i.e., load acting on unit of sliding surface area, divided by circumferential
coefficient p/u (specific load, i.e., load acting on unit of sliding surface area, divided by circumferential
speed on bearing pitch diameter) was exceeded by more than three times due to this possibility of
speed on bearing pitch diameter) was exceeded by more than three times due to this possibility of
sliding surface alignment.
sliding surface alignment.
Some tests were carried out with strong external excitation with variable frequency. Some
excitation frequencies caused severe vibrations of the otherwise stable running rotor. The regimes with
vibration amplitudes increased by external excitation are documented in Figure 8 for the speeds of
30,000 and 60,000 rpm. Excitation frequencies inducing severe rotor vibrations ranged from about
60 to about 200 Hz; lower or higher excitation frequencies had no adverse effect on rotor vibration.
The abovementioned frequencies are in the vicinity of the four lowest eigenvalues, i.e., critical speeds
of the rigid rotor on the aerodynamic bearing film.
At the end of test, the machine had endured 1000 start-stop cycles without any traces of excessive
wear or sliding surface damage. As is illustrated by Figure 9, neither the pads nor shaft in the bearing
location shows significant sliding surface damage; at the pad surface there are still visible traces of the
original machining.
Figure 7. Vibration amplitudes at run-up to 23,000 rpm (left) and run-down from 60,800 rpm (right).
in relatively big deviations of the rotor in a radial direction. This phenomenon is caused by relatively
low stiffness of pad supports in comparison with thrust bearing film stiffness. This by-product of
elastic pad support contributes to a significant increase of thrust bearing load capacity. The average
coefficient p/u (specific load, i.e., load acting on unit of sliding surface area, divided by circumferential
speed on bearing pitch diameter) was exceeded by more than three times due to this possibility of
Lubricants 2017, 5, 10
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sliding surface alignment.
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6 of 9
Some tests were carried out with strong external excitation with variable frequency. Some
excitation frequencies caused severe vibrations of the otherwise stable running rotor. The regimes
with vibration amplitudes increased by external excitation are documented in Figure 8 for the speeds
of 30,000 and 60,000 rpm. Excitation frequencies inducing severe rotor vibrations ranged from about
60 toFigure
about 200
Hz; lower or higher
excitation
frequencies
had no
adversefrom
effect
on rotor
vibration.
run-down
60,800
rpm (right).
(right).
Figure 7.
7. Vibration
Vibration amplitudes
amplitudes at
at run-up
run-up to
to 23,000
23,000 rpm
rpm (left)
(left) and
and run-down
from 60,800
rpm
The abovementioned
frequencies
are
in
the vicinity
of theDB
four
lowest
eigenvalues,
i.e.,
critical
speeds
Top
TB vertical
vertical
direction,
TB
horizontal
direction,
DB
horizontal
direction,
axial
direction,
Top down:
down: TB
direction,
TB
horizontal
direction,
horizontal
direction,
axial
direction,
DB
of theDB
rigid
rotor
on
the
aerodynamic
bearing
film.
vertical
direction;
maximum
vibration
double
amplitude
10
µm
at
DB
vertical
direction.
vertical direction; maximum vibration double amplitude 10 µ m at DB vertical direction.
(left)
and
60,000
rpm
(right)
with
external
excitation
2 g. Top
Figure 8. Vibration
Vibrationamplitudes
amplitudesatat30,000
30,000
(left)
and
60,000
rpm
(right)
with
external
excitation
2 g.
down:
TB vertical
direction,
TB horizontal
direction,
DB horizontal
direction,
axialaxial
direction,
DB
Top
down:
TB vertical
direction,
TB horizontal
direction,
DB horizontal
direction,
direction,
vertical
direction;
maximum
vibration
double
amplitude
450450
µ mµm
at DB
vertical
direction
DB
vertical
direction;
maximum
vibration
double
amplitude
at DB
vertical
direction
At the
end of test,
machine
hadappeared
endured when
1000 start-stop
cycles
without
traces of excessive
The
problem
withthe
rotor
stability
the bearing
span
had toany
be reduced
to assure
wear
or
sliding
surface
damage.
As
is
illustrated
by
Figure
9,
neither
the
pads
nor
shaft
in the
bearing
that the machine would fit into the defined space. The rotor with a bearing span reduced
from
80 to
location
shows
significant
sliding
surface
damage;
at
the
pad
surface
there
are
still
visible
traces
of
45 mm exhibited instability already at relatively low speeds, around 30,000 rpm. This is illustrated by
the vibration
original machining.
the
signals around 30,000 and 50,000 rpm in Figure 10. Very distinct subharmonic frequency
components, which varied from 1/4 to 1/7 of rotational frequency, can be observed in vibration signals
in Figure 10. Its frequency could be determined by comparing vibration signals with speed signals.
vertical direction; maximum vibration double amplitude 450 µ m at DB vertical direction
At the end of test, the machine had endured 1000 start-stop cycles without any traces of excessive
wear or sliding surface damage. As is illustrated by Figure 9, neither the pads nor shaft in the bearing
location 2017,
shows
Lubricants
5, 10significant sliding surface damage; at the pad surface there are still visible traces
7 ofof
9
the original machining.
Lubricants 2017, 5, 10
7 of 9
frequency components, which varied from 1/4 to 1/7 of rotational frequency, can be observed in
vibration signals in Figure 10. Its frequency could be determined by comparing vibration signals with
speed signals.
Figure
9. Bearing
sliding surface
surface after
after test.
test.
Figure 9.
Bearing pads
pads and
and rotor
rotor sliding
The problem with rotor stability appeared when the bearing span had to be reduced to assure
that the machine would fit into the defined space. The rotor with a bearing span reduced from 80 to
45 mm exhibited instability already at relatively low speeds, around 30,000 rpm. This is illustrated
by the vibration signals around 30,000 and 50,000 rpm in Figure 10. Very distinct subharmonic
Figure
Figure 10.
10. Vibration
Vibration amplitudes
amplitudes around
around 30,000
30,000 rpm
rpm (left)
(left) and
and 50,000
50,000 rpm
rpm (right).
(right). Top
Top down:
down: speed
speed
signal,
TB
vertical
direction,
TB
horizontal
direction,
DB
horizontal
direction,
axial
direction,
DB vertical
signal, TB vertical direction, TB horizontal direction, DB horizontal direction, axial direction,
DB
direction;
maximum
vibrationvibration
double amplitude
225 µm at
TBµ in
direction.direction.
vertical direction;
maximum
double amplitude
225
m horizontal
at TB in horizontal
Vibration amplitudes in the region of instability were very high, reaching practically the whole
bearing
to to
thethe
elastic
padpad
support
thatthat
immediate
failure
did not
bearing clearance,
clearance,and
andititwas
wasonly
onlythanks
thanks
elastic
support
immediate
failure
didtake
not
place.
RotorRotor
dynamic
analysis
carriedcarried
out disclosed
that, due
to the
span reduction,
the stability
take place.
dynamic
analysis
out disclosed
that,
duebearing
to the bearing
span reduction,
the
reserve
the lowest
at 60,000atrpm
decreased
from 9.8%
to 9.8%
7.3%.toOur
experience
shows
stabilityof
reserve
of theeigenvalue
lowest eigenvalue
60,000
rpm decreased
from
7.3%.
Our experience
that
in most
cases
of cases
a rotorofsupported
in aerodynamic
tilting padtilting
journal
bearings,
5% ofeven
stability
shows
that in
most
a rotor supported
in aerodynamic
pad
journaleven
bearings,
5%
of stability
reservetoisensure
sufficient
to ensure
stability,
this
special
the rotor had
to be
reserve
is sufficient
stability,
but in this
specialbut
caseinthe
rotor
had tocase
be redesigned
to achieve
redesigned
to achieve stable operation.
stable
operation.
4.
4. Discussion
Discussion
The
with
thethe
reduced
bearing
span
waswas
discovered
withwith
the
The reason
reason for
forthe
theinstability
instabilityofofthe
therotor
rotor
with
reduced
bearing
span
discovered
help
of several
experiments,
one of
which
was carried
out with
the blower
impeller
removed.
The rotor
the help
of several
experiments,
one
of which
was carried
out with
the blower
impeller
removed.
The
without
the
blower
impeller
exhibited
stable
operation
in
the
whole
operating
range
up
to
60,000
rpm.
rotor without the blower impeller exhibited stable operation in the whole operating range up to
Dynamic
analysis
of the
rotor with
such
geometry
showed
that stability
was increased
60,000 rpm.
Dynamic
analysis
of the
rotor
with such
geometry
showedreserve
that stability
reservefrom
was
7.3%
to 10%,
i.e.,7.3%
the value
calculated
thecalculated
rotor with for
the the
original
span.
To achieve
at span.
least the
increased
from
to 10%,
i.e., the for
value
rotorbearing
with the
original
bearing
To
achieve at least the same stability reserve as that of the rotor without the blower impeller, the thrust
bearing was relocated between journal bearings so that the bearing span could be increased to 63
mm. The rotor dynamic analysis proved that this new configuration increased stability reserve of the
lowest eigenvalue to 12%, which was even higher than that for original rotor with a bearing span of
80 mm. The modified rotor has therefore all the prerequisites for stable operation.
Lubricants 2017, 5, 10
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same stability reserve as that of the rotor without the blower impeller, the thrust bearing was relocated
between journal bearings so that the bearing span could be increased to 63 mm. The rotor dynamic
analysis proved that this new configuration increased stability reserve of the lowest eigenvalue to 12%,
which was even higher than that for original rotor with a bearing span of 80 mm. The modified rotor
has therefore all the prerequisites for stable operation.
5. Conclusions
A new type of aerodynamic tilting pad bearings was used for the rotor support of an air cycle
machine for aircraft cabin pressurization. Bearings with elastically supported pads were able to secure
stable operation of the rotor with a mass of 0.7 kg up to a speed of 68,000 rpm. Pad elastic support leads
to additional positive bearing properties, such as the possibility of adaptation to changing operating
conditions, and provides additional bearing damping. Preset basic bearing clearance reduces sliding
surface wear and contributes to easy run-up. Elastic pad support enables, among other features,
very rapid rotor run-up to maximum speed without the danger of bearing clearance reduction due to
different temperature dilations of the bearing and shaft. Similarly, as with foil bearings, vibrations
from the rotor to the outside structure and vice versa are substantially damped. As a by-product of the
journal bearing design, the thrust bearing load carrying capacity is increased due to the possibility of
aligning the rotor axis according to the thrust bearing sliding surface position.
The above described example shows that no strict rules can be defined as concerns rotor stability.
While some rotors in aerodynamic bearings exhibit stable operation with a stability reserve lower than
5% (log δ ~0.15), in the present case more than 7% was not sufficient. The reason for this behavior
may be found in the effect of high-gyroscopic moments of impellers at both overhanging ends of the
rotor. Most high-speed machines with heavy impellers overhanging the rotor end exhibit somewhat
unorthodox behavior (e.g., [16]), which should be further investigated.
Acknowledgments: This work was supported by the Technology Agency of the Czech Republic under project
PID TA02011295 “Verification of air bearing technology”. Many thanks to Petr Lindovský and Lukáš Chromek for
performing the experiments and providing data for analysis.
Conflicts of Interest: The authors declare no conflict of interest.
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article distributed under the terms and conditions of the Creative Commons Attribution
(CC BY) license (http://creativecommons.org/licenses/by/4.0/).