560130

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http://slidepdf.com/reader/full/560130 1/86

Design Modifications of Papermachine Support Structures to Control Vibration

Andrew K. Costain, B.Sc.Eng.Dale G. Eyre, Eng., B.Sc., M.Sc., D.Sc.

J Michael Robichaud, P.Eng., CMRP

Bretech Engineering Ltd.

ABSTRACT

The generally accepted methods for vibration control of equipment and structures include Force Reduction, MassAddition, Tuning, Isolation, and Damping. This paper will describe the use of tuning, or changing, the natural

frequency of a system or component to control vibration response of large structures. Several case studies will be

presented, with emphasis on pragmatic solutions to papermachine vibration problems.

1. INTRODUCTION

Notwithstanding the replacement of worn or defective components, such as damaged bearings, 5 basic methods exist

for vibration control of industrial equipment, as detailed below;

• Force Reduction of excitation inputs due to, for example, unbalance or misalignment will decrease the

corresponding vibration response of the system.

• Mass Addition will reduce the effect (system response) of a constant excitation force – not considered practical

for papermachine support structures.

• Tuning (changing) the natural frequency of a system or component will reduce or eliminate amplification due

to resonance.

• Isolation rearranges the excitation forces to achieve some reduction or cancellation

• Damping is the conversion of mechanical energy (vibrations) into heat – not considered practical for

papermachine support structures.

2. BASIC METHODS FOR PAPERMACHINE VIBRATION CONTROL

The typical and preferred methods for vibration control of paper machine structures are force reduction and tuning.These methods are described in detail on the following pages.

2.1 Force Reduction

Force reduction of inputs related to rotating components, such as unbalance, misalignment, looseness, and rubbing,

will result in a corresponding reduction of vibration response. Typically, force input increases in proportion to the

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2

2

f eMF ⎟

⎠

⎞⎜⎝

⎛ =

π

[Equation 1 • unbalance force]

where F = force, lb [N]

M = mass, lb-sec2/in [kg]

e = eccentricity, in [m]

f = frequency, Hz

For force inputs at or near the system natural frequency, f n, (resonance), amplification of the vibration response is

likely to occur. This may cause otherwise acceptable (residual) force inputs to result in excessive vibrations. For

well damped systems, force reduction may sufficiently control the vibration response. For lightly damped systems,force reduction is typically used in conjunction with Tuning.

2.2 Tuning

Tuning is a process used to eliminate amplification due to resonance by changing a system or component natural

frequency, f n, so that it is no longer coincident with the frequency of a specific force input. Resonance of industrial

equipment will amplify vibration response, in theory up to ∞, depending on system damping characteristics.

The Synchronous Amplification Factor (SAF) is a measure of how much 1X vibration is amplified when the system passes through a resonance. Systems with a high effective damping tend to have a low SAF, and systems with low

effective damping have a high SAF. The Bode Plot, shown in Figure 1, below, indicates that systems with a high

SAF (lightly damped) have a narrow range of resonance with high amplification; systems with low SAF (well

damped) have relatively broad range of resonance with low amplification. Note that the range of resonanceindicates the amount of tuning (change in the system natural frequency) required to eliminate resonance. Resonant

frequencies that are nonsynchronous exhibit similar behavior.

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Before attempting to tune a resonant system, the natural frequency (and damping characteristics) should be

determined experimentally, typically by a forced response (bump) test. Once the frequency relationship between

force input and system natural frequency, f n, are evaluated, a decision must be made on whether to raise or lower the

natural frequency. For a simple (sdof) vibrating system, f n is proportional to the stiffness to mass ratio. Equation 2,

below, indicates that adding stiffness will raise the f n, and, conversely, adding mass will lower the f n.

M

K

2

1f n

π

≅ [Equation 2 • natural frequency for simple vibrating system]

where f n = system natural frequency, Hz

K = stiffness, lb/in [N/m]

M = mass, lb-sec2/in [kg]

Since increasing mass and/or decreasing stiffness (to lower the f n) may compromise the strength of a machine or its

supporting structure, adding stiffness (to raise the f n) is the most common practical application of detuning. When

adding stiffeners to a resonant system, the optimal location for end connections is the system antinodes, andinstallations at nodes will be ineffective. A modal or operating deflection shape analysis is useful for determining

the location of antinodes. Care should be taken to ensure that: 1) the added mass of the stiffener does not cancel its

stiffening effect (note that pipe has the best stiffness to mass ratio), 2) stiffeners do not introduce new componentnatural frequencies that are coincident with force input frequencies (i.e. resonance), and 3) stiffener end connections

are as fixed (rigid) as practically possible.

0.0 20.0000Hertz0.0

1.2000

Cross Channel Spectrum

FRF, Pump #1, -10Y, Pump Frame

LBFi: 0.7265Hert: 10.5000

26-Feb-0315:39:31JOB ID: 198

12 Hz

0.0 20.0000Hertz0.0

0.4200

Peak

in/sec

Peak Hold

Coast Down Spectrum

in/s: 0.3485Hert: 11.8752

26-Feb-0312:24:20JOB ID: 202198

11.9 Hz

Figure 2A•

forced response test of vertical pump Figure 2B•

peak hold coastdown of vertical pump

Figures 2A and 2B, above, show the evaluation of system f n for a vertical pump, using both forced response testing

and peak hold coastdown methods. In this case, 1X pump rotating speed, 700 rpm, is coincident with a systemnatural frequency (i.e. resonance). Note that by analyzing phase and coherence data, the 10.5 Hz peak shown in

Figure 2A was found to be an external force input, rather than a system natural frequency.

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3.1 Transient and Speed Trial Data

Speed trial and transient testing methods involve the continuous acquisition of data for a period of time and

operating conditions. These methods will capture any events which are not continuous, such as beats resulting from

amplitude and/or frequency modulation. Vibration beating may be of particular interest, since dryer sections produceclosely spaced forcing frequencies due to the draw or speed difference from the wet end to the dry end required to

keep sheet tension. Beats can be very destructive over time to frame condition and directly affect machine

reliability.

Speed trials involve continuous data acquisition over a machine operating range or range of interest. This data is

acquired to investigate the relationship of machine speed versus vibration amplitude and frequency. As speed

changes, so do associated forcing functions - as machine speed increases, vibration levels should show a gradual

corresponding increase. The speed trial data analysis may reveal system natural frequencies and other unusualvibration amplitude patterns present in the machine operation.

Resonance occurs when forcing frequencies are coincident with system natural frequencies. At resonance, systemresponse (vibration amplitude) is amplified. A sudden sharp peak followed by decreases in amplitude as machine

speed increases is an indication of the presence of a system natural frequency. The amount of amplification depends

on the system damping characteristics.

Unbalance in rolls may also be indicated in speed trial data. A sharp and continual increase in amplitude with

increased speed may indicate an unbalance condition, since calculated force due to unbalance increases by thesquare of the rotational frequency (as shown in Equation 2). A key factor affecting vibration response due to

unbalance is the rotating speed of the roll.

Multiple channels of vibration data acquired simultaneously with increasing machine speed on all papermachine

components, tending and drive sides, will also allow the investigation of component phase relationships.

Figure 3 contains the 0 to 40 Hz waterfall plot of data in acceleration (g’s) of a dryer section measurement. This

data was acquired midframe on the tending side in the machine direction. The multi-axis plot contains data on three

axes, the vibration amplitude, time, and frequency.

Amplitude

Frequency

Time

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High vibration response from resonance is identified in several location of this plot as well and a proportional

vibration increase with speed – typical of unbalance in a rotating component. Using the trend over a given speed

range comparison of vibration representing unbalance is separated from resonant response.

3.2 Forced Response

Forced Response Testing (impact testing) is used to identify the natural frequencies of systems and components. The

raw data acquired during impact testing is evaluated for amplitude, coherence and phase shift to verify the accuracyof the sample. This data is then crosschecked with system forcing functions to identify possible areas where

resonance will occur.

Sample data plots for the impact testing are shown in 4 and 5. The data was acquired on a point located at press felt

roll 202F, in the upper frame of the press section. The indicated peak is 25.9 Hz. Note that other valid peaks arealso present.

0 .0 250.0000Hertz0.0

0.0240

Peakin/sec

Spectrum

096 - -50Z:1X(Ch B)

in/s: 0.019 6Hert:25.9375

30-May-0714:05:00JOB ID:UTOPIAFRF

0 .0 2 50 .0 0 00Hertz0.0

0.0150

ec/LBF

Transfer Function B/A

096 - -50Z:1X

in/s: 0.013 7Hert: 25.937 5

30-May-0714:05:00JOB ID:UTOPIAFR

Figure 4A • forced response test Figure 4B • transfer function

The validity of an impact test is considered with reference to the dominance of the peak shown in the natural

frequency transfer function compared with the coherence and the phase shift data. The coherence plot verifies that

the vibration data acquired by the accelerometer is in fact a result of the impact by the modal hammer. In this test, avalid data point will have a coherence value of 0.95 or higher on a scale of 1.0, the indicated value in Figure 5 is

0.996. Additionally, a phase shift of approximately 180 degrees in the cross channel phase plot is consistent with

resonance conditions (see Figure 5B).

1.1000

Cross Channel Coherence096 - -50Z:1X

30-May-0714:05:00JOB ID:UTOPIAFR 180.0000

Degrees

Cross Phase096 - -50Z:1X

30-May-0714:05:00JOB ID:UTOPIAFR

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All Forced Response data requires evaluation using the criteria presented in prior to acceptance as a valid natural

frequency of the tested component(s)

Forced response testing indicates system natural frequencies, typically the 1st and 2nd bending modes. When forcing

functions (such as 1X dryer turning speed) are coincident with the natural frequency, a significant increase invibration amplitude will occur (i.e. resonance).

Additional components of interest to natural frequency modes, other than peak frequency and amplitude, aredamping and amplification. The amount of damping in a structure affects the resonant response, such that, with no

damping, one would expect to get infinite motion with very small excitation. As the damping increases, the resonant

response from a given force reduces.

In a lightly damped system, the frequency peak is narrow at its base in the spectral plot. This narrow peak implieslarge amplification over a narrow frequency range. Conversely, a highly damped system shows a peak at a natural

frequency with a wide base with lower amplification factor over a typically wider frequency range. Associated with

system damping is how much the system vibration amplitude will be amplified by a coincident forcing function. Thedegree of amplification may be quantified by Equation 3, whereQ is the amplification factor, f c is the center or peak

frequency, and f a and f b are the frequencies at which the amplitude peak is 0.707 of its maximum.

fafb

fcQ

−= [Equation 3 • Amplification Factor using ½ power points]

where f c = peak power point, Hz

f a, f b = rms power points, Hz

Q = amplification factor

In Figure 6 below, the amplification factor for the natural frequency peak of the #212 Felt Roll in the upper frame of

dryer section 2, peaks at 0.707 of the center frequency amplitude, and Hz, are used to calculate Q, the amplification

factor. Therefore, the vibration amplitudes at a frequency of 25.9 Hz will be amplified by a factor of 26.0.

0.0 50.0000Hertz0 .0

0.0240

Peak in/sec

Spectrum

096 - -50Z:1X(Ch B)

in/s: 0.01 96Hert: 25.9375

30-May-0714:05:00JOB ID: UTOPIA FR

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3.3 High resolution vibration measurements

Radial vibration measurements are acquired using a data acquisition route configured for all dryer cylinder bearings

on the tending side and drive side. The data acquired is used as an assessment tool for the frequency and amplitudeof each contributor to the vibration problem. Occasionally, there are multiple frequencies that can be contributing tothe overall problems, thus requiring more than a single correction or single course of action.

Machine speed during data acquisition is held constant at the highest normal production rate for the machine.Vibration amplitudes at 1X the dryer turning speed represent movement induced by the turning speed of the roll and

may indicate unbalance.

Vibration amplitudes of 0.04 in/sec peak (ips pk) have been identified as the maximum acceptable vibration

amplitude at 1X dryer turning speed for dryer cylinders based on historical data and experience. Sample spectraland waveform data is shown in Figure 7.

Figure 7 • sample vibration spectrum and waveform

This sample plot indicates the spectra and associated acceleration waveform for a horizontal measurement on thetending side bearing housing of dryer 33. The amplitude of the 1X dryer vibration in this plot, 0.205 ips pk, exceeds

the acceptable amplitude of 0.04 ips pk, indicating the presence of dryer unbalance or frame movement.

Dryer unbalance is typically a combination of static and couple unbalance, and the severity of it is represented by

the vibration magnitude at 1X the roll turning speed. The phase of the 1X component is measured relative to a phase

marker (reflective paint) painted on a dryer head bolt. Generally, if the unbalance force is a result of pure unbalance

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Vibration amplitude and phase data is associated with specific points on a model to visually display the relative

movements of papermachine components. The model can then be animated at any frequency present in the vibration

data. This is valuable in identifying the frequencies causing the deflection of rolls and frames as well as in

determining the phase relationship between these components.

Animations are created in .avi movie format using cross-channel data, images from the animations are shown below.

Figure 8A • ODS press model Figure 8B • ODS former model

Animations are run at specific forcing frequencies identified during previous data analysis using field data.

The results of ODS animation results is a key indication of problem locations and the limit of the affected areas or

components. As data is acquired during a periodic vibration measurements variables including transient or beat

vibration must be identified and addressed to ensure accurate results. Incorrect data acquisition and analysis is oneof the leading potential sources for incorrect diagnosis.

4. Sample Structural Vibration Case Studies

Three case studies selected from previous analysis and corrections follow. Each correction deals with a problemthat was identified and corrected through structural tuning or force reduction. Strong similarities exist in the dataand careful analysis is necessary to determine the cost effective source to solution.

4.1 Case Study #1

A large newsprint machine was to increase nominal operating speed from 3400 fpm to 4000 fpm. Since existing

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0.0 4.2000Seconds

-1.2000

1.2000

G's

Waveform of Dryer Can #20 - Y Direction

G's: -0.2682

Seconds: 0.1699

09/12/02

11:57:29

JOB ID: 202137-3

0.0 200.0000Hertz

0.0

0.4800

Peak

in/sec

Spectrum of Felt Roll F28 - X Direction

in/sec: 0.3924

Hertz: 13.0000

09/12/02

16:04:45

JOB ID: 202137-4A

0.0 200.0000Hertz

0.0

0.0300

Peakin/sec

Spectrum of Dryer Can #20 - Y Direction

in/sec: 0.0217

Hertz: 3.7500

09/12/02

11:57:29

JOB ID: 202137-3

0.0 4.2000Seconds

-0.6000

0.6000

G's

Waveform of Felt Roll F28 - X Direction

G's: 0.0696

Seconds: 0.0

09/12/02

16:04:45

JOB ID: 202137-4A

Figure 9A • dryer bearing defect Figure 9B • felt roll unbalance

As shown in Figs 10A and 10B, below, excessive dryer cylinder unbalance was detected at dryer section #1. Similar

results were found at section #2 and #3, with 1X dryer (4 Hz) amplitude up to 0.3 in /sec pk in the machine direction(target value is 0.04 in/sec pk).

Conversely, the dryer sections #4 and #5 were found to be well within tolerance. Note that sections #1, #2, and #3

are driven by dryer felts, and sections #4 and #5 are driven by enclosed gear trains.

Figure 10A • 1X dryer vibration amplitude Figure 10B • 1X dryer phase / magnitude vectors

Tending Side 1X Vibration - 3650 FPM

0.000

0.020

0.040

0.060

0.080

0.100

0.120

0.140

0.160

0.180

0.200

1 2 3 4 5 6 7

Dryer Cylinder #'s

V i b r a t i o n A m p l i t u d e ( i n / s e c p k )

Tending Side Horizontal

Tending Side Vertical

Tending Side Axial

Vibration Limit

0.238 0.210 0.202

An operating deflection analysis indicated excessive machine direction motion in the supporting structure. The ods

model is shown in Fig 11, below. Machine direction is +x.

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Figure 11•

operating deflection shape model

Visual examination of the support structure revealed previously installed machine direction braces, as shown in Fig

12A, below. These were found to be undersized, and in generally poor condition, with many of the structural welds

broken.

Figure 12A • failed structural brace Figure 12B • example structural brace

Recommendations were submitted to design and install suitable machine direction bracing, similar to that shown inFig 12B above At the time of writing this work is in progress

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Baseline vibration assessment identified: 1) dryer unbalance, and 2) machine direction frame resonance at current

operating speed. See waterfall plot in Figure 13 below. Frame resonance was affecting felt roll vibration.

Figure 13 • waterfall plot – machine direction sill beam

Data acquired using both constant speed and transient techniques identified speed range where residual vibrationamplitude was amplified in resonance. The dryer section was composed of 6 sections with individual sill beam

separations by section. Data indicated both in phase and out of phase machine direction vibration as measured over

time. Steady state amplitudes reached 1X dryer vibration amplitudes in excess of 0.6 ips-pk (target 0.04 ips-pk).

Typical data plots are shown below in Figure 14 A and B.

ROUTE WAVEFORM08-Feb-06 10:10:42

RMS = .0499

PK(+) = .1550PK(-) = .2205

CRESTF= 4.42

0 0.3 0.6 0.9 1.2 1.5 1.8 2.1 2.4 2.7

-0.3

-0.2

-0.1

0.0

0.1

0.2

Timein Seconds

A c c e l e r a t i o n i n G - s

PM 2 - Dryer #52D52 -TSH Tending Side Horizon tal

ROUTE SPECTRUM

08-Feb-06 10:10:42

OVERALL= .4891 V-DGPK = .5311

LOAD = 100.0RPM = 94.

RPS = 1.57

0 30 60 90 120 150

0

0.1

0.2

0.3

0.4

0.5

0.6

Frequencyin Hz

P K V e l o c i t y i n I n / S e c

Freq:Ordr:Spec:

3.9382.501.411

Figure 14A•

typical spectrum dryer bearing Figure 14B•

1X dryer magnitude

An operating deflection analysis indicated excessive machine direction motion in the supporting structure and in

particular vibration for the sill beam and below. This is the root cause of the upper frame vibration and the locationfor correction.

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Figure 15 • operating deflection shape model

Bracing was installed during an annual outage in 4 of the 6 dryer sections (as time permitted) in conjunction with

trim balancing of residual dryer vibration.

Figure 16A • installed structural brace Figure 16B • example dryer balance weights

Vibration amplitudes were reduced to the target amplitude and breaks in the dryer section stopped. Machine speed

is increasing and reliability improving.

4.3 Case Study #3

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PM1 Speed Trial - Machine Direction Vibration

1X Dryer 37, Amplitud e and PhaseMarch 2003

-120

-100

-80

-60

-40-20

0

20

40

60

80

100

120

2000 2500 3000 3500 4000 4500 5000 5500 6000

Machine Speed (FPM)

C o r r e c t e d P h a s e A n g l e

(

D e g r e e s )

0.00

0.10

0.20

0.30

0.40

0.50

0.60

0.70

0.80

0.90

V i b r a t i o

n A m p l i t u d e

( i n . / s

e c p e a k )

Measured Phase

Projected Phase

Measured Amplitude

Figure 17 • waterfall plot – machine direction sill beam

Speed trials identified that although the machine was operating in resonance at speed ranges between 3900 and 4100

FPM, the frame stiffness was lower than expected and dryer unbalance continued to deflect the frame with resulting

high vibration amplitudes.

Structural corrections identified include both stiffness for a natural frequency change and strength for increased

operating speed loads.

As with previous analysis, 1X dryer peak magnitude and phase is evaluated for peak amplitudes vs. target and their associated vector (phase & magnitude).

Tending Side, 1XDryer Vibration - March 2003

Dryer Section 4

0.00

0.04

0.08

0.12

0.16

0.20

0.24

0.280.32

0.36

0.40

0.44

0.48

0.52

0.56

0.60

D28 D29 D30 D31 D32 D33 D34 D35 D36 D37 D38 D39

Dryer Can #

V i b r a t i o n A m p l i t u d e

( i n . / s e c . p e a k )

Vibration Limit

(0.04 in./sec.)

0.70

(in/sec peak)

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Figure 19 • operating deflection shape model

Using a computer FEA of the frame system along with load calculations and structural design two large braces were

added to each side of the machine section. The braces were installed and a subsequent follow-up testing identified

vibration amplitudes at the sill beam were reduced from 0.6 IPS pk at 1X dryer vibration to 0.02 ips pk.

Figure 20A • installed structural brace Figure 20B • installed brace bench

Vibration amplitudes were reduced to the target amplitude and machine speed increased to 4500 FPM.

5 CONCLUSION

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BIBLIOGRAPHY

[1] Costain, A. and Robichaud, J. M., Practical Methods for Vibration Control of Industrial Equipment,CMVA, (2003)

[2] Costain, A., Case Studies on Paper Machine Vibration Problems, CMVA, (2003)

[3] Eshleman, R. Machinery Vibration Analysis: Diagnostics, Condition Evaluation, and Correction,

Vibration Institute, (2002)

[4] Jackson, C. The Practical Vibration Primer

Gulf Publishing Company - Houston, TX, (1979)

[5] Bently, Donald E. Fundamentals of Rotating Machinery Diagnostics

Bently Pressurized Bearings Press, (2002)

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Design Modificationsof Papermachine Support Structures

to Control Vibration

presented by

Andrew Costain

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Î introduction

Î test methods

Î design strategies

Î examples

Design Modifications

of Papermachine Support Structures


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• established in 1984

•

based in Saint John, NB• 45 Engineers & Technicians

computed

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Î introduction

Î test methods


of Papermachine Support Structures


• vibration assessment & condition evaluation

• forced response testing

• speed trial testing• operating deflection shape analysis

• peak magnitude & phase measurements

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ROUTE WAVEFORM

16-OCT-03 21: 20: 34

RMS = .0373PK(+) = .1844

PK(-) = .1670CRESTF= 4.95

0 0.3 0.6 0.9 1.2 1.5 1.8 2.1 2.4 2.7

-0.20

-0.15

-0.10

-0.05

-0.00

0.05

0.10

0.15

0.20

Time in Sec onds

A c c e l e r

a t i o n i n G - s

PM 5 - Dryer #52

D52 -TSH Tending Si de Hor izontal

ROUTE SPECTRUM16-OCT-03 21: 20: 34

OVRALL= .0924 V-DG

PK = .0986LOAD = 100.0RPM = 232.

RPS = 3.87

0 5 10 15 20 25 30 35 40

0

0.01

0.02

0.03

0.04

0.05

0.06

0.070.08

Fr equency in Hz

P K

V e l o c i t y i n I n / S e c

Fr eq:

Ordr:

Spec:

10.22

2.642

.01355

1X dryer

2X dryer

1X felt rol l

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Time Domain

Frequency Domain

MACHINERY VIBRATION ANALYSIS

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ROUTE WAVEFORM

28-FEB-00 11:45:59

RMS = .5096

PK(+) = 2.09

PK(-) = 1.86

CRESTF= 4.11

0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0

-2.5

-2.0

-1.5

-1.0

-0.5

0

0.5

1.0

1.5

2.0

Time in Seconds

A c c e

l e r a

t i o n

i n G - s

PM#1 - Dryer #24

D24 -TSH Tending Side Horiz.

ROUTE SPECTRUM

28-FEB-00 11:45:59

OVRALL= .3989 V-DGPK = .3987

LOAD = 100.0

RPM = 220.

RPS = 3.67

0 4000 8000 12000

0

0.04

0.08

0.12

0.16

0.20

Frequency in CPM

P K

V e

l o c

i t y i n

I n / S e c

Freq:

Ordr:

Spec:

Dfrq:

9161.3

41.64

.07664

213.28

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0.0 150.0000Hertz0.0

0.0200

Peakin/sec

Spectrum

350 - 283Y/001X(Ch B)

in/s: 0.0179Hert: 51.0000

15-Aug-0613:29:05JOB ID: MODDAO2P

0.0 150.0000Hertz0.0

1.2000

Cross Channel Coherence

350 - 283Y/001X

0.0 150.000Hertz-180.0000

180.0000

Degrees

Cross Phase

350 - 283Y/001X

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0

1

2

3

4

5

67

8

9

10

11

12

13

14

15

16

0 200 400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 2800 3000

Speed (FPM)

F r e q

u e n c y ( H z )

1X 48" RollS

2X 42.5" Rol l

1X 42.5" Rol l

Primary Bel-Bond, X Dir., TS

Secondary Bel -Bond, Y Dir., DS

PM1 Fourdrinier • January, 2004PM1 Fourdrinier • January, 2004

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No 5 Paper Machine Baseline Assessment

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D12(0.05 ips)

D14(0.13 ips)

D16(0.08 ips)

D18(0.07 ips)

D20(0.02 ips)

D22(0.10 ips)

D13(0.10 ips)

D15(0.13 ips)

D17(0.08 ips)

D19(0.13 ips)

D21(0.08 ips)

D23(0.12 ips)

No. 5 Paper Machine Baseline Assessment

Drive and Tending Sides, 1XDryer Vibration

Dryer Section 2, October 2003

0.289

0.00

0.04

0.08

0.12

0.16

0.20

D12 D13 D14 D15 D16 D17 D18 D19 D20 D21 D22 D23

Dryer #

V i b r a t i o n A m p

l i t u d e ( i n . / s e c . p e a k )

Tending Side Drive Side

Vibration Limit (0.04 in./sec.)

No. 5 Paper Machine Baseline Assessment

Drive and Tending Sides 1XDryer Vibration

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Drive and Tending Sides, 1XDryer Vibration

Dryer Section 3, October 2003

0.2870.2390.2040.2220.2060.1990.2

0.00

0.04

0.08

0.12

0.16

0.20

D24 D25 D26 D27 D28 D29 D30 D31 D32 D33 D34 D35 D36 D37 D38 D39

Dryer #

V i b r a t i o n A m

p l i t u d e ( i n . / s e c . p e a k )

Tending Side Drive Side

Vibration Limit (0.04 in./sec.)

D27(0.20 ips)

D25(0.20 ips)

D29(0.21 ips)

D31(0.19 ips)

D33(0.22 ips)

D35(0.20 ips)

D37(0.24 ips)

D39(0.29 ips)

D24(0.08 ips)

D26(0.12 ips)

D28(0.08 ips)

D30(0.05 ips)

D32(0.11 ips)

D34(0.04 ips)

D36(0.06 ips)

D38(0.09 ips)

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Îintroduction

Î test methods

Î design strategies


of Papermachine Support Structuresto Control Vibration

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Î FORCE REDUCTION

Î MASS ADDITION

Î TUNING

Î ISOLATIONÎ DAMPING

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Î Mass Addition will reduce the effect (system

response) of a constant excitation force.

aMF =

K

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Î Tuning (changing) the natural frequency of a

system or component will reduce or eliminate

amplification due to resonance.

M

K

f n ≅

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T d i Sid 1X Vib t i 3650 FPM

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Tend ing Side 1X Vib rat ion - 3650 FPM

0.000

0.020

0.040

0.060

0.080

0.100

0.120

0.140

0.160

0.180

0.200

1 2 3 4 5 6 7

Dryer Cylinder #'s

V i b r a

t i o n

A m p

l i t u d e

( i n / s e c p

k )

Tending Side Ho rizontal

Tending Side Vertical

Tending Side Axial

Vibration Limit

0.238 0.210 0.202

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Typical correction recommendations

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2. IN-SITU ROLL BALANCING to reduce the forcing or

exciting function.

Recommendations

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Baseline Dynamic Assessment

& Structural Modifications

case study 2

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ROUTE WAVEFORM08-Feb-06 10:10:42

RMS = .0499

PK(+) = .1550

PK(-) = .2205CRESTF= 4.42

0 0.3 0.6 0.9 1.2 1.5 1.8 2.1 2.4 2.7

-0.3

-0.2

-0.1

0.0

0.1

0.2

Time in Seconds

A c c e l e r a t i o n i n G - s

PM 2 - Dryer #52

D52 -TSH Tending Side Horizontal

ROUTE SPECTRUM

08-Feb-06 10:10:42

OVERALL= .4891 V-DGPK = .5311

LOAD = 100.0

RPM = 94.

RPS = 1.57

0 30 60 90 120 150

0

0.1

0.2

0.3

0.4

0.5

0.6

Frequency in Hz

P K V e l o c i t y i n I n / S

e c

Freq:Ordr:

Spec:

3.9382.501

.411

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Î Very High Vibration Amplitudes

Î 1X turning speed of dryers is excit ing supportstructure natural frequencies.

Î Vibration increasing sheet breaks

Observations

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Baseline Dynamic Assessment

& Structural Modifications

case study 3

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2. IN-SITU ROLL BALANCING to reduce the forcing or

exciting function.

Recommendations

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560130

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