CRITICAL SPEED VERIFICATION OF 18,000 HP ADJUSTABLE SPEEDDRIVE MOTORS FOR OFFSHORE PRODUCTION APPLICATIONS

Copyright Material IEEEPaper No. PCIC-2006-42

David C. RainsWEG Electric - USA3219 Stephens Creek LaneSugar Land, TX, 77478USAdcrains@weg. net

Carlos A. Bavastri, Dr. Eng.UTFPRAv. Sete de Setembro, 3165Curitiba, PR, 80230-901Brazilbavastri@cefetpr.br

Hideraldo L. V. dos SantosWEG Industrias - MotoresAv. Pref. Waldemar Grubba, 3000Jaragua do Sul, SC, 89256-900Brazilhideraldos@weg.net

Kazuo ShojiWEG Industrias - MaquinasAv. Pref. Waldemar Grubba, 3000Jaragua do Sul, SC, 89256-900Brazilkazuos@weg.net

Abstract - This paper describes the tests and analysisperformed on 3 x 18,000 HP purged and pressurizedsynchronous motors in order to comply with API 546 Standardrequirements. These tests are even more important ifconsidered that these motors will be being supplied power froman adjustable speed drive that will vary the motor speed from1260 up to 1890 rpm on offshore platforms.

Index Terms - Vibration tests, API 546, resonant speed,lateral critical analysis.

I. INTRODUCTION

API standards are used globally by the petroleum industry toimprove performance, reliability and enhance quality of theirequipments and installations. One key issue of API 546,required for the form-wound brushless synchronous motors(Fig. 1), is related to their strict vibration performancerequirements.

These motors, with main characteristics listed at Table 1,will drive gas compressors for a FPSO platform. This platformwill be installed at Roncador field, Campos basin, Brazil.FPSOs (Floating, Production, Storage and Offloading

vessels) (Fig. 2) can process and store crude oil, and offloadthe oil and/or natural gas. A process plant is installed on theship's deck to separate and treat the fluids from the wells. Afterthe crude is separated from the water and gas, it is stored intanks on the actual vessel and then offloaded to a relief shipevery so often.

Frame SizeOutputNumber of poleVoltageCooling

Main characteristics

Inam Ul HaqDresser-RandPaul Clark DriveOlean, NY, 14760-0560USAnam_u_haq@dresser-rand .com

Rogerio da Silva CamposPetrobrasAv. Rep. Do Chile, 65 room 21021Rio de Janeiro, RJ, 20035-900Brazilrscampos@petrobras.com .br

TABLEMOTOR DATA

IEC 90018,000 hp

46,600 V

Water cooledEx-p type of protection

ASD operated (1260 - 1890 rpm)BV & INMETRO certifications

1-4244-0559-9/06/$20.00 ©2006 IEEE 1

The relief vessel is an oil tanker that moors on the FPSOstern to receive the crude stored in its tanks and then transportit onshore. Compressed gas is sent onshore through gaspipelines and/or injected back into the reservoir. FPSOoperating data are according to Table II.

Oil productionWater injectionGas compressionWater depthRisers

TABLE IIFPSO OPERATING DATA

180,000 bpd39,000 m3/d6 million m3/d

1 315 m57

Campos Basin (Fig. 3) is considered the largest oil reserve inthe Brazilian continental shelf. It covers an area of some100,000 square kilometers and stretches from the EspiritoSanto state to the northern coast of Rio de Janeiro State.Today more than 400 oil and gas wells, 30 or so productionplatforms, and 3,900 kilometers of offshore pipelines are inoperation there.Campos Basin exploration started at the end of 1976 at

Garoupa field and commercial production began there in 1977at the Enchova field. In 1996 Roncador was discovered and itscommercial production started in 2000.

This paper will present the tests performed at factory in orderto verify that these motors were in compliance with API 546standard vibration requirements.The tests results will be compared then to theoretical values

calculated from an analytical model applied for rotor lateraldynamic analysis in order to check how accurate it is.

II. PERFORMED TESTS AND ANALYSIS

A. Coast-down test

According to item 2.4.6.1.2 of API 546, lateral naturalfrequencies which can lead to resonance amplification ofvibration amplitudes shall be removed from the operatingspeed frequency and other significant exciting frequencies byat least 15 percent. Machines intended for continuousoperation on adjustable speed drives shall meet thisrequirement over the specified speed range.The coast-down test was performed in order to identify all

critical speeds from zero to Maximum Continuous Speed.In this test, the motor was accelerated to its rated speed and

then, turned off. During the motor run down, vibration levelsrelated to rotor speed are recorded. A Bode plot, according toFig. 5, can be used for this test.

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0 200 400 600 800 1000 1200

240

300

360

50

i:2.67/39 3.9)21r

33 rpm1400 1 600 1uu1 iLI .iI1

.I

10 ;=

20

0 160.10 20 40 60 0 100 120 140 160 I8

SPEED' 50 rpmldiv

Fig. 5 Coast-down test (Bode plot)

B. Rotor dynamic analysis

During the tests described at ILA, a critical speed wasidentified in the equipment operating speed range. In order toconfirm this fact, a rotor dynamic analysis was performed (seeattachment A).

The finite element method (FEM) was applied for thisanalysis tool. In this method, the distributed physical system isdivided into a number of discrete elements linked by pointscalled nodes.

Considering the nodal displacement hypothesis, the kinetic,potential and damping energies are calculated as function ofnodal displacements.

2

4 i7 -4 -4

The Lagrangian of the system can be defined as:

L = T - V (1)

TABLE IlIlCALCULATED DATA COMPARISON

Critical Speeds

Where:L Lagrangian of the systemT Kinetic energy of the systemV Potential energy of the system

The Lagrangian fits into a certain equation called Lagrange'sEquation. In this case this equation is defined as:

dQ2LIdt t06

aL+

0JOD

. =FaJ

Mode

1st

2nd

3rd

Measured Values

1650 rpm

Calculated Values

1620 rpm

1759 rpm

2468 rpm

(2)

Where:DF6

Damping energyGeneralized loadsGeneralized coordinates

After determining all rotor parts element matrices, thedifferential equation of global rotor movement can bedetermined. Considering the gyroscope and damping matrix,the rotor natural frequencies and vibration modes may bedefined from the eigenvalues problem solution.The solution at steady-state condition is achieved

considering the modal parameters and unbalance excitation,including their location and intensity. Then, the dynamic rotorcharacteristic may be represented in the Campbell diagram.Figure 6 shows a typical Campbell diagrams for these systems.

. _.

Fig. 8 Purchaser estimated 1st vibration mode

Fig.6 Campbell diagram

Calculated characteristic frequencies of these motor rotorsare listed in table l1l. In Fig. 7 it is showed the first vibrationmode indicated by this procedure and in Fig. 8, the same fromthe purchaser procedure.

C. Electrical-mechanical run-out measurement

API 546 item 4.3.3.3 requires that machines that areprovided with non-contact probes or provisions for non-contactprobes shall be tested to verify that the shaft sensing areasmeet the total electrical and mechanical runout requirements of2.4.5.1.3 (less or equal to 25 percent of the maximum allowedpeak-to-peak unfiltered vibration value). The probe track runoutshall be measured with the rotor at slow roll (200-300 rpm)speed, where the unbalance forces on the rotor are negligible.A continuous unfiltered trace of the non-contact probe output

3

shall be recorded for a 360 degree shaft rotation at each pilocation with the shaft rotating in the assembled machine.rotor shall be held at its axial magnetic center during recordi

This measurement may be done during the coast-downand it is required in order to avoid undesirable effectsmeasured vibration levels.

D. Rotor thermal stability test

As rotor balance conditions may change at hot conditionin order to assure that vibration levels are going to be alvunder specified vibration limits, the rotor thermal stability teperformed.According to item 4.3.3.10 of API 546, the vibration shal

within filtered and unfiltered specified limits throughouttemperature range from the test ambient temperature tototal design temperature.

In this case, due to the fact that these motors areoperated, they were tested over all their operating speed ra(1260 up to 1890 rpm).

Fig. 9 shows the recorded data. The cold conditioishowed in solid lines and the hot condition, in dashed lines.

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25APR2005 25APR2005 25APR2005 25APR20051 3:00 1 4:00 1 5:O0 1 6:00

(D5¾~L

Lu aJ)Cco

360 F

90 H

100270K360

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26APR2005 26APR2005 26APR2005 26APR2005 26APR200510:50 11:00 11:10 11:20 11:30

0..

w~o-360180360180360

40

30

20

1 0

=40

-30

20

1 0n

1 0:50 11:00 11:1 0 11:20 11:3026APR2005 26APR2005 26APR2005 26APR2005 26APR2005

TIME: 2 Mins /divFig. 10 Unbalance and damped response test plot

(In phase)

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(D >

0)Lu CL)

co -o< C,

n

30 4I I I ~~~~~~~~~~30

20~ ~ ~ c~22i0

1 3:00 1 4:00 1 5:00 1 6:0025APR2005 25APR2005 25APR2005 25APR2005

TIME: 1 0 MinsjidivFig. 9 Rotor stability test plot

F. Unbalance and damped response test

As a critical frequency was identified in the machineoperating speed range, the unbalance and damped responsetest was preformed.

Item 4.3.5.4 of API 546 requires that a deliberate unbalanceof 4X of allowable unbalance weight per plane be applied to therotor. The weights shall be placed at the balance planes inphase and also 180 degrees out of phase and the machineshall be run to 120 percent of its rated speed.

According to the same item, for machines which do notcomply with the specified separation margin, a well-dampedresponse shall be demonstrated.

Fig. 10 shows the measured vibration levels when theunbalance mass were located in-phase and Fig. 1 1, when out-phase. These tests were performed at cold (solid lines) and hotcondition (dashed lines) and at worst vibration condition speed.

240

300I

360

GO

I~~ ~ ~ ~ ~ ~ ~ ~ ~~4

I~ ~~ ~ ~ ~ ~ ~ ~ ~ ~ ~ ~~~~~

30

20.

1 0

1 3:40 1 4:00 1 4:2026APR2005 26APR2005 26APR2005

TIME: 2 Mins /divFig. 1 1 Unbalance and damped response test plot

(Out phase)

The amplification factor was considered in order to be surethat the rotor is well-damped. A value below 2,5 is adequate toconsider the system well-damped and this factor is calculatedaccording to the following equation:

AF=-;.c o (3)360 dN

Where:AkF :Amplification FactorNc Critical speed, rpmdo1/dN Rate of change of phase angle with

respect to speed around critical speed,deg/rpm

The phase angle with respect to speed was obtained fromthe bode plot of coast-down test and the calculatedamplification factor was equal to 1.92.

4

i Ft A .- --,-4

bu L 71b U

in , -,; -,Po

d h 115

U 'l lu

x - f

CONCLUSIONS

Due to the fact that a critical speed was found in theoperating speed range of the motor, all three motors wereextensively tested at worst condition according to thoseconditions required by API 546 standard in order to assure theirperformance and reliability.

Related to the vibration analysis tool, we can consider thecalculated values from software were satisfactory. Oneimportant point that has to be mentioned is that proceduresbased on traditional methods for critical speed calculation arenot able to identify the first (rigid body motion) motion

ACKNOWLEDGEMENTS

The authors would like to thanks to Tobias Milani Dietrich forhis assistance during all the tests performed, and to HiltonPenha Silva and Francisco Jose Doubrawa Filho for theirefforts for Vibration Analysis Procedure development

REFERENCES

[1] API Standard 546 - Blushless Synchronous Machines -500 kVA and Larger, second edition, June 1997.

[2] Michel Lalanne and Guy Ferraris, "RotordynamicsPrediction in Engineering" 2a Edition, England 1998, JohnWiley and Sons

{3} Fredric F. Ehrich, "Handbook of rotordynamics", USA1992,McGraw-Hill

VITA

Hideraldo L. V. dos Santos graduated in MechanicalEngineering at Universidade Federal do Parana in 2003. Heworks as Junior Researcher at WEG Industrias S.A. - Motoressince 2003.

Inam Ul Haq is a staff rotordynamics engineer with Dresser-Rand, Turbo Products Division, Olean, NY, USA. He has 24yrs. of practicing experience with turbomachines of energyrelated industries. He holds a Masters degree in MechanicalEngineering from Carleton University, Ottawa, Canada. He is amember of the ASME and authored 15 refereed technicalpapers published in the esteemed international conferencesand periodicals.

Carlos Alberto Bavastri graduated in Mechanical Engineeringat National University of Comahue, Neuquen, Argentina, andobtained his Doctor degree at Federal University of SantaCatarina, Santa Catarina, Brazil. His field of work is optimumvibration and noise control, in a wide frequency band, usingviscoelastic dynamic neutralizers. He is associated to theGroup of Integrated Research on Vibratory and AcousticalSystems, recognized by CNPq, Brazil. He is currently a regularmember of the academic staff ("Professor Adjuntoconcursado") of the Federal Technological University of Parana(UTFPR), Brazil. With other colleagues, he is nominated as ainventor in a request of a Brazilian patent on viscoelasticneutralizers for overhead lines.

David C. Rains graduated from City University; London. Heis Oil, Gas & Petrochemical industry manager at WEG ElectricMotors and is active on IEEE and API standards related tomotors and their applications.

Kazuo Shoji graduated from Escola Politecnica daUniversidade de Sao Paulo in 1985 as a Mechanical Engineer.He has been a design engineer for the WEG Industrias S.A. -Maquinas of Jaragua do Sul SC since 2003. He is a registeredprofessional engineer in the state of Sao Paulo.

Rogerio da Silva Campos, graduated from Rio de JaneiroFederal University in 1986 as a Mechanical Engineer. He hasbeen a Rotordynamic and Vibration Petrobras specialist since1987, working in Petrobras Downstream Department.

5

ATTACHMENT A

ROTORDYNAMIC ANALYSIS - NUMERICAL RESULTS

Shaft Material Properties:

Density: 7850 kg/m3Young Module: 2.06x101 N/M2

* Poisson Coefficient: 0.3

Shaft Dimensions, bearings location, discs mass:

Operational Speed:

0 1260 to 1890 RPM

Lubricant film stiffness and damping as function of rotation:

1200 1300 1400 1500 1600

ROTATION [RPM]

1700 1800 1900 2000 2100

6

Section Inner Diameter Outer Diameter Section Length Disks' Mass Bearing[mm] [mm] [mm] [kg] Location

1 0 270.0 335.0 -

2 0 280.0 562.0 - DE3 0 345.0 148.0 - -

4 0 340.0 158.0 - -

5 0 417.5 314.2 - -

6 0 417.5 1245.6 4850 -

7 50 425.1 282.2 - -

8 50 340.0 269.0 - -

9 50 315.0 140.0 - -

10 50 280.0 500.0 - NDE11 50 255.0 68.0 -

12 50 235.0 76.5 3013 50 234.0 68.0 -

14 50 212.0 180.0 -

15 50 210.0 291.0 22516 50 170.0 123.5 225

1 *109

8.108

DYN BRG (DE) - LUBRICANT FILM STIFFNESSIf

l ~~~ ~ ~ ~ ~~~~~ -

- - - - - - - - - - - - - _

-_ _- _- _- _- _- _- _- _- _- _- _-_ _-_ _- _- _- _- _- _-

I__

6 108 K

4.108 [

2.108 K

0

-2 108 L10(00 1100

kllk12k21k22

ROTOR'S FINITE ELEMENT MODEL:

* The rotor was approached by #16 discrete beam elements with #3 discs, representing the rotor stack, the fan exciterand the exciter.

* The bearings stiffness and damping properties was calculated by a polynomial function of machine's rotation.

Figure below shows the finite element model built to numerical analysis.

CAMPBELL DIAGRAM:

The Campbell's Diagrams is shown below

7

CRITICAL SPEEDS:

1St: 1620.0 rpm 2nd 1759.0 rpm3rd: 2467.7 rpm 4th 3680.8 rpm5th: 3770.2 rpm 6th 4930.4 rpm

MODE SHAPES:

8

LOAD CASES FOR UNBALANCE RESPONSE:

Rotor core ends unbalance weights and position:

e: 300mm* m: 100g

UNBALANCE RESPONSE PLOTS:

9