eccentricity measurements on a five-pad tilting pad journal bearing · · 2015-10-15eccentricity...

The 14th IFToMM World Congress, Taipei, Taiwan, October 25-30, 2015 DOI Number: 10.6567/IFToMM.14TH.WC.OS14.020

Eccentricity Measurements on a Five-pad Tilting Pad Journal Bearing

Phuoc Vinh Dang1

Politecnico di Milano

Via La Masa 1, I-20156 Milan, Italy

Steven Chatterton2



Paolo Pennacchi3



Andrea Vania4



Filippo Cangioli5



Abstract: The study of static and dynamic effects plays a

key–role in the field of rotor dynamics to evaluate the

behavior of hydrodynamic bearings. This paper describes

an experimental apparatus and the procedure for

measuring the static eccentricities and the bearing

clearance distribution in a five–pads tilting pad journal

bearing (TPJB). The tests have been carried out on a five–

pads TPJB with nominal diameter of 100 mm, length–to–

diameter ratio (L/D) of 0.7 and load–on–pad (LOP)

configuration. The results show that the clearance profile,

with a pentagon–shape, is a function of the operating

temperature and of the pad dimensions. Besides, the

barycenter of the clearance profile is very close to the

equilibrium position of the journal center at low rotational

speeds. A procedure for the evaluation of the bearing

clearance with experimental measurement is also

described. Keywords: Tilting pad journal bearing, eccentricity, clearance

profile, bearing clearance, load-on-pad

I. Introduction

It is well known that dynamics of high–speed rotating

machines strongly depends on journal bearing

characteristics. Currently, the titling pad journal bearings

(TPJBs) are widely mounted on high rotational speeds

machinery, such as steam and gas turbines. This is mainly

due to two special features of the TPJBs, which are i)

stability at high rotational speeds and ii) tolerance to

misalignment.

In order to select and design TPJBs compatible with

the required operating conditions of a high–performance

rotating machine, analytical models are commonly used to

forecast, among others, the white–metal temperature, the

load capacity, the minimum oil–film thickness, the

stiffness and the damping coefficients. However, it is

difficult to exactly predict the performances of a TPJB

because of its complex geometry, model uncertainties (in

particular about pivot compliance), large oil temperature

variations, and possible establishment of turbulent flow in

the oil–film.

[email protected] 2steven.chatterton@ polimi.it [email protected] 4andrea.vania@ polimi.it 5filippo.cangioli@ polimi.it

A theoretical study about the effects of TPJB

geometry, on the static and the dynamic characteristics,

has been presented by Jones and Martin [1], by considering

the load direction, the pad clearance, the bearing clearance,

the length–to–diameter ratio and the number of pads

The effects of eccentricity of the journal have been

studied in [2] and [3] for a circular bearing and in [4] for a

TPJB. Good agreements have been obtained between

measured and theoretical eccentricities, for a range of

Sommerfeld numbers between 0.2 and 3.0.

Walton and San Andrés [5] have provided some

experimental results of the static and dynamic

performances of a TPJB with flexible pivot. The

measurements have been employed to validate the

analytical simulations performed on this kind of bearing.

In addition, the phenomena related to heat transfer were

evaluated experimentally to provide an estimation of the

bearing power losses, during operation. The test conditions

consisted of rotational speeds and static loads up to 4,500

rpm and about 1,400 N (311 lbf), respectively. The cross–

coupling effects resulted relatively negligible for this

bearing type, over the range of loads and speeds tested, due

to the small pad displacements in the orthogonal direction

to the applied load. Moreover, a very little sub–

synchronous oil–whirl, was observed within the bearing

during the tests. Besides, the bearing stiffness increased

with the increasing of either the applied static load for a

given journal rotational speed or the journal rotational

speed for a constant load.

Wilkes and Childs [6] have investigated the effects of

the oil temperature, the pad clearance and the radial

displacement of the loaded pad (i.e. the pad having the

static load vector directed through its pivot) on the bearing

clearance. The measured clearances corresponding to

thermal steady state were approximately 30% smaller than

the measured one at the cold start–up and were inversely

proportional to pad surface temperature.

The aim of this paper is to present the results of the

measurements of eccentricity, clearance profile and

bearing assembled clearance Cb in a 100 mm diameter,

length–to–diameter ratio of 0.7, LOP configuration five–

pads TPJB. It is observed that both the actual machined

geometry and the pad assembling cause the clearance

profile to be far from the ideal design one. This could

affect the actual characteristics of the TPJB, also including

the existence of cross–coupling effects.

mailto:[email protected]

II. Description of the test rig and of the bearing

A detailed description of the test–rig, used for the

experimental tests is given in [7, 8, 9]. The rotor is

supported by two equal TPJBs, with nominal diameter of

100 mm and length-to-diameter ratio of 0.7. The bearing at

the non-driven end (NDE) which is labeled 1 of the shaft is

tested. The load is applied in the middle of the shaft by

means of two hydraulic actuators arranged in an

orthogonal configuration at ±45° with respect to the load

cells. The actuators are connected to the shaft by means of

two deep-groove high-precision ball bearings. Thanks to

this configuration, both the static load and the dynamic

loads can be applied in any direction. The nominal force of

the hydraulic actuator is 25 kN. The actuators are able to

move the shaft with amplitude of 100 µm in the frequency

band of 0-50 Hz and are provided by high-resolution

position and force transducers. The test-rig is controlled by

proprietary PC-based software, while data acquisition is

performed by means of Labview software and PCI DAQ

boards (National Instruments cDAQ-9178). A sketch of the

test–rig is shown in Fig. 1.

Fig. 1. Test-rig used for identification of

bearing characteristics

All the pads are instrumented with a temperature

sensor and a pressure probe, as shown in Fig. 2. The

pressure probe measures the pressure in the center of the

pad by means of a tiny hole in the white–metal and a duct

in the shoe. The geometric characteristics of the bearing

under test and the operating conditions are listed in Table

1. A picture of the complete test–rig is shown in Fig. 3.

Fig. 2. Five-pads tilting pad journal bearing

used for the tests

Item Unit Value/Span

Number of pad - 5

Configuration w.r.t

bearing housing - LOP

Bearing diameter (D) mm 100

Machined clearance (Cp) mm 0.125

Nominal pads thickness mm 16.0

Bearing length (L) mm 70

Angular amplitude of pads degrees (°) 60

Upper pads - 3,4

Lower pads - 1,2,5

Lubricant - ISO VG46

Rotational speed rpm 1200

Static load

(on each bearing) kN 5

Table 1. Bearing geometric characteristics

and operating conditions

Fig. 3. Picture of test-rig used to perform the

tests

Motor Ball

bearing

Hydraulic

actuator

2

1

Hydraulic actuator

Load

cell

Pressureprobe

Temperatureprobe

Nozzle

Pressure

probe

Temperature

probe

III. Experimental procedure

The experimental results discussed here have been

obtained with the load centered over the bottom bearing

pad, i.e. in the LOP configuration.

Since the thermal expansion of the bearing housing,

which supports the proximity probes, makes it difficult to

define the actual center of the bearing, steady–state shaft

eccentricities have been measured with respect to a lightly

loaded equilibrium position, corresponding to the shaft

rotating at 50 Hz (3000 rpm) with the external load equal

to about 350 N on each bearing. The Sommerfeld number

for this lightly loaded equilibrium position, which

corresponds to an eccentricity ratio less than 0.01, is

greater than 12 [10]. We can consider this position as the

bearing center. The Sommerfeld number is calculated

using the following formula:

2

p

NLD RS

W C

(1)

where:

µ: oil dynamic viscosity [Pa s], in which ( )T

N: shaft rotational speed [Hz]

L: bearing length [m]

D: bearing diameter [m]

R: bearing radius [m]

Cp: machined clearance [m]

W: vertical load applied on the bearing [N]

The dynamic viscosity, µ, is interpolated from the

curve of dynamic viscosity vs. temperature.

The average temperature over the five pads, measured

on the centerline position for each pad, is calculated as:

1

1

n

pad i

i

Pad temp Tn

(2)

where:

n: number of pads (n = 5)

Tpad-i: temperature of i-th pad

The average temperature is assumed as the actual

temperature of oil–film. The actual oil–film temperature

and corresponding viscosity are used for the calculation of

the Sommerfeld number.

Before and after each test run, once the rotation had

stopped, the clearance profiles of bearing were measured

by using the proximity probes. The rotor is slowly moved

inside of the housing support by means of a rotating force,

applied in the middle of the shaft through the deep groove

precision ball bearings and the two hydraulic actuators [7].

The rotating resultant force has been chosen in order to:

ensure that each pad is loaded by the exerted thrust

along its span;

be sufficiently high to put the shaft in contact with

all the pads;

to prevent significant pivot deflections during the

tests.

The temperature of the oil is kept at 40°C by means of

a PID temperature controller. However, in correspondence

of bearing loads that are extremely high or low, the

temperature may exceed this limit.

After having obtained the clearance profile of the

bearing, using interpolation, five points at the vertices of

the pentagon–shaped profile can be defined. Besides, the

barycenter of each profile was determined. This point has

turned out to be very close to the equilibrium position of

the journal center.

The co–ordinates of the journal center have been

evaluated during all the tests by means of two proximity

probes. Then, the values of eccentricity B Je O O ,

attitude angle ( , which is the angle between x-axis and

B JO O ) and eccentricity ratio ( ) can be calculated as

(see in Fig. 4):

2 2

equilibrium equilibriume x x y y (3)

1tan [ ]equilibrium

equilibrium

x xrad

y y

(4)

b

e

C (5)

Fig. 4. Test–rig reference system

The method used for the bearing clearance

measurement Cb (see Fig. 5) is described in [11]. The tests

for the clearance measurement are performed after those

used to define the clearance profile.

Fig. 5. Titling pad bearing clearance measurement

Y

Xe

ø (+)

Non-driven view

OB

OJ

X

Pad 1

Pad 3Pad 4

shaft

Test

condition

#1 #2 #3 #4

10Hz 25Hz 40Hz W = 3.535kN

1 Clearance profile

(Test A1)

Clearance profile

(Test A2)

Clearance profile

(Test A3)

Clearance profile

(Test A4)

2 Equilibrium

position

Equilibrium

position

Equilibrium

position

Equilibrium

position

3 5 kN 5 kN 5 kN 5 Hz

4 6 kN 6 kN 6 kN 10 Hz

5 4 kN 4 kN 4 kN 15 Hz

6 7 kN 7 kN 7 kN 20 Hz

7 3 kN 3 kN 3 kN 25 Hz

8 8 kN 8 kN 8 kN 30 Hz

9 2 kN 2 kN 2 kN 35 Hz

10 9 kN 9 kN 9 kN 40 Hz

11 1 kN 1 kN 1 kN 45 Hz

12 Bearing

clearance Cb

Bearing

clearance Cb

Bearing

clearance Cb

Bearing

clearance Cb

13 Clearance profile

(Test A2)

Clearance profile

(Test A3)

Clearance profile

(Test A4)

Clearance profile

(Test A5)

Table 2. Operating conditions for measurements on TPJB

As shown in Fig. 5, the shaft is pushed upwards in

between of the two upper pads (pad 3 and pad 4): this is the

highest position of the shaft. Then, the shaft is moved to

the lower pad (pad 1): this is the lowest position of the

shaft. The value of Cb can be calculated using the formula,

based on geometrical considerations, reported in [11]:

0.894bC X (6)

Each series of test consisted of a combination of shaft

rotational speed and bearing loads, corresponding to

different Sommerfeld numbers and eccentricity values.

Before and after each test run, once the rotation had

stopped the clearance profiles of bearing which were

named as test A1 – A5 were defined by rotating force as

mentioned previously (see Table 2). The aim of this

procedure is to see the effect of temperature before and

after each test run to the clearance profile of TPJBs.

After having obtained the clearance profile of bearing,

using interpolation, five points at peaks of the

shaped-pentagon profile can be identified. Besides, the

barycenter of each profile was determined. This point is

very close to the position of equilibrium of the center of the

bearing (see Fig. 6 - Fig. 9).

Four test series were run and each one consisted of

nine different conditions. The tests series #1, #2, and #3

were run at constant rotational frequency, respectively

10 Hz, 25 Hz, and 40 Hz, with measurements taken using

nine bearing loads, from 1 to 9 kN on each bearing, in steps

of 1 kN. Because the oil temperature strongly depended on

the bearing load, the loads were applied on the bearing as

listed in Table 2, for maintaining the temperature in the

aforementioned range. The test series #4 was performed by

applying a constant bearing load, equal to 3.535k N on

each bearing, and by changing the shaft rotational speed

from 5 Hz to 45 Hz, with steps of 5 Hz.

IV. Results and discussions

The bearing eccentricity measurements, shown in Fig.

6 - Fig. 9, are plotted as a function of the rotational speed

and of the applied load, along with the bearing clearance

profiles, before and after each series of tests. The

temperatures of the oil and of the pads during the five tests

are listed in Table 3.

As shown in Fig. 6, when the rotational speed is low,

the measured eccentricity deviates slightly from the

bearing vertical centerline. The largest eccentricity, which

corresponds to the maximum load of 9 kN, is about 70 µm.

The clearance profiles of two tests, namely A1 and A2,

corresponding to different temperatures, are also plotted in

Fig. 6. The equilibrium position and the barycenters of two

pentagon–shaped clearance profiles are either coincident

or very close each other. It can be seen that, when the

temperature of inlet oil increases, then the pentagon shape

tends to shrink. This trend shows a good agreement with

the clearance profile changes reported in [6].

When the rotational speed increases, the deviation of

the measured eccentricity from the vertical centerline

grows (especially at 40 Hz), though the largest

eccentricities decrease with respect to that of the test at

10 Hz. This indicate that the bearing cross–coupling

stiffness coefficients (kxy and kyx) cannot be neglected and

that they have some influence on the dynamic performance

of the bearing under test, at least in these specific operating

conditions. Besides, the distance between the equilibrium

position and the barycenters increases almost

proportionally with the rotational speed.

Fig. 9 shows the eccentricity values for test series #4.

It is evident that the deviations of the measured

eccentricity from the bearing vertical centerline are quite

large, even in correspondence of low rotational speed. This

effect might be caused by the relatively large difference of

the temperature between the pads and the oil. As shown in

Table 3, the temperature difference is about 12°C before

beginning the test (profile A4) and 10°C after ending the

test (profile A5).

Fig. 6. Eccentricity measurement at a constant speed of 10Hz

Fig. 7. Eccentricity measurement at a

constant speed of 25Hz


constant speed of 40Hz


constant static load of 3.535kN

The measured eccentricities, as a function of the

applied static loads, are shown in Fig. 10 for three different

rotational speeds. It results that the eccentricity is almost

inversely proportional to the rotational speed and

practically linear to the static load.

Fig. 10. Titling pad bearing clearance measurement

Test Test

A1

Test

A2

Test

A3

Test

A4

Test

A5

Pads [°C] 39.70 43.60 46.00 50.80 47.40

Inlet [°C] 36.60 39.40 38.50 38.35 37.56

Outlet [°C] 39.41 39.27 40.27 41.22 39.64

Table 3. Operating temperatures

The measured eccentricity ratios and attitude angle, as

a function of the Sommerfeld number calculated using

nominal radial clearance and pads’ average temperature,

are shown in Fig. 11 and Fig. 12 and a summary of

measured static bearing eccentricities and the

corresponding fitting curve for the five-pads TPJB is

shown in Fig. 13. These Sommerfeld number are based on

the oil viscosity corresponding to the bearing pad

temperature.

It should be noted from Fig. 11 and Fig. 12 that large

Sommerfeld number; i.e. denoting small load, high speed

or large lubricant viscosity, determine small operating

journal eccentricities (small eccentricity ratios) or nearly

1000 2000 3000 4000 5000 6000 7000 8000 90000

10

20

30

40

50

60

70

Static Load [N]

Eccentr

icity [

m

]

= 10Hz

= 10Hz - Fitted curve

= 25Hz


= 40Hz


centered operation, 0, 2e (90°). That is, the

journal eccentricity vector is nearly orthogonal to the

applied load. On the other hand, small Sommerfeld

number; i.e. denoting large load, low speed or low

lubricant viscosity, determine large operating journal

eccentricities (large eccentricity ratios), 1, 0e

(0°). Note that the journal eccentricity vector is nearly

parallel to the applied load

Fig. 11. Measured static bearing eccentricities versus Sommerfeld number

for a five-pad TPJB

Fig. 12. Measured attitude angle versus Sommerfeld number for a five-pad TPJB

It is clearly seen that, in Fig. 13 most of experimental

results either touch or coincide with each other when the

Sommerfeld number falls within the range of 0.1 and 0.5.

Fig. 13. A summary of measured static

bearing eccentricities versus Sommerfeld number for the five-pads TPJB

V. Conclusions

The paper presents a procedure, along with its

experimental validation, for the identification of the

eccentricity and the assembled bearing clearance Cb of a

five–pads TPJB with LOP configuration.

The clearance profile of the bearing strongly depends

on the operating temperature and the actual dimensions of

the pads. The measurements also confirmed that the

equilibrium position of journal center is close to the

barycenter of pentagon–shaped clearance when the

rotational speed of the shaft is low.

It is necessary to take into account the cross–coupling

stiffness coefficients (kxy and kyx) when dynamic

characteristics of the test bearing are evaluated, especially

at high speed.

The measured eccentricity is almost inversely

proportional to the rotational speed and practically linear

to the static load applied to the rotor bearing system.

The effects of the preload factor on the eccentricity

will be investigated in the near future.

References

[1] Jones G.J. and Martin F.A., "Geometry Effects in

Tilting-Pad Journal Bearings," ASLE Transactions, vol. 22,

no. 3, pp. 227-244, 1979.

[2] Orcutt F.K. and Arwas E.B., "The Steady State and Dynamic

Characteristics of a Full Circular Bearing and a Partical Arc

Bearing in the Laminar and Turbulent Flow Regimes,"

ASME Journal of Lubritation Technology, vol. 89, pp.

143-152, April 1967.

[3] Tonnesen J. and Hansen P.K., "Some experiments of the

Steady State Characteristics of a Cylindrical Fluid-Film

Bearing Considering Thermal Effects," ASME Journal of

Tribology, vol. 103, no. 1, pp. 107-114, January 1981.

[4] Tripp H. and Murphy B., "Eccentricity Measurements on a

Tilting-Pad Bearing," ASLE Transactions, vol. 28, no. 2, pp.

217-224, 1984.

[5] Walton N.V. and San Andrés L., "Measurements of Static

Loading Versus Eccentricity in a Flexure-Pivot Tilting Pad

Journal Bearing," ASME Journal of Tribology, vol. 119, pp.

297-304, April 1997.

[6] Wilkes J.C. and Childs D.W., "Tilting Pad Journal Bearings -

A Discussion on Stability Calculation, Frequency

Dependence, and Pad and Pivot," ASME Journal of

Engineering for Gas Turbines and Power, vol. 134, no. 12,

pp. 1-17, December 2012.

[7] Chatterton S., Pennacchi P., Dang P.V. and Vania A., "A Test

Rig for Evaluating Tilting–Pad Journal Bearing

Characteristics," Proceedings of the 9th IFToMM

International Conference on Rotor Dynamics, vol. 21, pp.

921-930, 22-25 September 2014. doi:

10.1007/978-3-319-06590-8_75

[8] Chatterton, S., Pennacchi, P., Vania, A., Tanzi, E., and Ricci,

R., "Characterization of Five-Pad Tilting-Pad Journal

Bearings Using an Original Test-Rig," ASME Paper

DETC2011-48166,Washington, DC, USA, vol. 1, pp.

989-993, 2011. doi:10.1115/DETC2011-48166.

[9] Chatterton S., Pennacchi P., Dang P.V. and Vania A.,

"Identification Dynamic Force Coefficients of a Five-pad

Tilting–Pad Journal Bearing," Proceedings of the 9th

IFToMM International Conference on Rotor Dynamics, vol.

21, pp. 931-941, 2014. doi: 10.1007/978-3-319-06590-8_76

[10] Nicolas J.C., Gunter E.J. Jr. and Allaire P.E., "Stiffness and

Damping Coefficients for the Five-Pad Tilting-Pad

Bearing," ASLE Transactions, vol. 22, no. 2, pp. 113-124,

1979.

10-2

10-1

100

101

0

0.1

0.2

0.3

0.4

0.5

Speed = 10Hz

10-2

10-1

100

101

0

0.1

0.2

0.3

0.4

0.5

Speed = 25Hz

10-2

10-1

100

101

0

0.1

0.2

0.3

0.4

0.5

Sommerfeld number [S]

Speed = 40Hz

10-2

10-1

100

101

0

0.1

0.2

0.3

0.4

0.5


Load = 3.535kN

10-2

10-1

100

101

0

20

40

60

80

Ang

le [°]

Speed = 10Hz

10-2

10-1

100

101

0

20

40

60

80

Ang

le [°]

Speed = 25Hz

10-2

10-1

100

101

0

20

40

60

80


Ang

le [°]

Speed = 40Hz

10-2

10-1

100

101

0

20

40

60

80


Ang

le [°]

Load = 3.535kN

10-2

10-1

100

1010

0.05

0.1

0.15

0.2

0.25

0.3

0.35

0.4

0.45

0.5


Eccentr

icity R

atio [

]

Speed = 10Hz

Speed = 25Hz

Speed = 40Hz

Load = 3.535kN

[11] Nicholas J.C., "Tilting Pad Bearing Design," in Proceedings

of the twenty-third turbomachinery symposium, Texas

A&M University, 1994.

[12] Chatterton, S., and Pennacchi, P., and Ricci, R., 2010,

“Application and Comparison of High Breakdown-Point and

Bounded-Influence Estimators to Rotor Balancing”, J. Vib.

Acoust. 132(6), 064502. Doi:10.1115/1.4001842

[13] Pennacchi, P., Chatterton, S., and Ricci, R., 2010, “Rotor

Balancing Using High Breakdown-Point and

Bounded-Influence Estimators”, Mech. Syst. Signal

Process., 24(3), pp. 860-872.

doi:10.1016/j.ymssp.2009.10.004

[14] Pennacchi, P., Vania, A., Chatterton, S., and Pesatori, E.,

2010, “Case History of Pad Fluttering in A Tilting-Pad

Journal Bearing”, ASME Paper No. GT2010-22946, ASME

Turbo Expo: Power for Land, Sea and Air, 6, pp. 227-233.

Doi:10.1115/GT2010-22946

[15] Vania, A., Pennacchi, P., and Chatterton, S., 2012, “Dynamic

Effects Caused by the Non-Linear Behavior of Oil-Film

Journal Bearings in Rotating Machines,” ASME Paper

GT2012-69457, ASME Turbo Expo: Turbine Technical

Conference and Exposition, 7, pp. 657-664.

Doi:10.1115/GT2012-69457

[16] Vania, A., Pennacchi, P., and Chatterton, S., 2012, “Analysis

of the sensitivity to non-linear effects in the oil-film forces of

journal bearings,” Proc. of 10th IMECHE International

Conference on Vibrations in Rotating Machinery, London,

UK, C1326-037

[17] Vania, A., Pennacchi, P., Chatterton, S., and Tanzi, E., 2013,

“Sensitivity Analysis of Non-Linear Forces in Oil-Film

Journal Bearings”, World Tribology Congress 2013, Torino,

Italy, pp. 1-4.

[18] Bachschmid N., Pennacchi, P., Chatterton, S., and Ricci, R.,

2009, “On Model Updating of Turbo-Generator Sets”,

Journal of Vibroengineering, 3(11), pp. 379-391

[19] P., Borghesani, P., Ricci, R., Chatterton, S., and Pennacchi,

P., 2013, "A new procedure for using envelope analysis for

rolling element bearing diagnostics in variable operating

conditions", Mechanical Systems and Signal Processing,

38(1), pp. 23-35. doi:10.1016/j.ymssp.2012.09.014

[20] Pennacchi, P., Borghesani, P., Chatterton, S., and Vania, A.,

2012, "Hydraulic Instability Onset Detection in Kaplan

Turbines by Monitoring Shaft Vibrations", ASME Paper No.

DETC2012-70963, pp. 715-722,

doi:10.1115/DETC2012-70963

[21] Chatterton, S., and Pennacchi, P., and Ricci, R., 2010,

“Application and Comparison of High Breakdown-Point and

Bounded-Influence Estimators to Rotor Balancing”, J. Vib.

Acoust. 132(6), 064502. Doi:10.1115/1.4001842

[22] Pennacchi, P., Chatterton, S., and Ricci, R., 2010, “Rotor

Balancing Using High Breakdown-Point and

Bounded-Influence Estimators”, Mech. Syst. Signal

Process., 24(3), pp. 860-872.

doi:10.1016/j.ymssp.2009.10.004

[23] Pennacchi, P., Vania, A., Chatterton, S., and Pesatori, E.,

2010, “Case History of Pad Fluttering in A Tilting-Pad

Journal Bearing”, ASME Paper No. GT2010-22946, ASME

Turbo Expo: Power for Land, Sea and Air, 6, pp. 227-233.

Doi:10.1115/GT2010-22946

[24] Vania, A., Pennacchi, P., and Chatterton, S., 2012, “Dynamic

Effects Caused by the Non-Linear Behavior of Oil-Film

Journal Bearings in Rotating Machines,” ASME Paper

GT2012-69457, ASME Turbo Expo: Turbine Technical

Conference and Exposition, 7, pp. 657-664.

Doi:10.1115/GT2012-69457

[25] Vania, A., Pennacchi, P., and Chatterton, S., 2012, “Analysis

of the sensitivity to non-linear effects in the oil-film forces of

journal bearings,” Proc. of 10th IMECHE International

Conference on Vibrations in Rotating Machinery, London,

UK, C1326-037

[26] Vania, A., Pennacchi, P., Chatterton, S., and Tanzi, E., 2013,

“Sensitivity Analysis of Non-Linear Forces in Oil-Film

Journal Bearings”, World Tribology Congress 2013, Torino,

Italy, pp. 1-4.

[27] Bachschmid N., Pennacchi, P., Chatterton, S., and Ricci, R.,

2009, “On Model Updating of Turbo-Generator Sets”,

Journal of Vibroengineering, 3(11), pp. 379-391

[28] P., Borghesani, P., Ricci, R., Chatterton, S., and Pennacchi,

P., 2013, "A new procedure for using envelope analysis for

rolling element bearing diagnostics in variable operating

conditions", Mechanical Systems and Signal Processing,

38(1), pp. 23-35. doi:10.1016/j.ymssp.2012.09.014

[29] Pennacchi, P., Borghesani, P., Chatterton, S., and Vania, A.,

2012, "Hydraulic Instability Onset Detection in Kaplan

Turbines by Monitoring Shaft Vibrations", ASME Paper No.

DETC2012-70963, pp. 715-722,

doi:10.1115/DETC2012-70963

http://dx.doi.org/10.1016/j.ymssp.2009.10.004




eccentricity measurements on a five-pad tilting pad journal bearing · · 2015-10-15eccentricity...

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