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Bearing Technology TopicsVolume 2Various Technical Papers

TR-113059-V2

Final Report, February 2000

DISCLAIMER OF WARRANTIES AND LIMITATION OF LIABILITIES

THIS DOCUMENT WAS PREPARED BY THE ORGANIZATION(S) NAMED BELOW AS ANACCOUNT OF WORK SPONSORED OR COSPONSORED BY THE ELECTRIC POWER RESEARCHINSTITUTE, INC. (EPRI). NEITHER EPRI, ANY MEMBER OF EPRI, ANY COSPONSOR, THEORGANIZATION(S) BELOW, NOR ANY PERSON ACTING ON BEHALF OF ANY OF THEM:

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Electricité de France

ORDERING INFORMATION

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Copyright © 2000 Electric Power Research Institute, Inc. All rights reserved.

iii

CITATIONS

This report was prepared by

Electricité de FranceDirection des Etudes et Recherches1, Avenue du Général-de-GaulleBP 40892141 Clamart Cedex, France

Nuclear Maintenance Applications Center (NMAC)1300 W. T. Harris BoulevardCharlotte, North Carolina 28262

This report describes research sponsored by EPRI.

The report is a corporate document that should be cited in the literature in the following manner:

Bearing Technology Topics, Volume 2: Various Technical Papers, EPRI, Palo Alto, CA: 2000.TR-113059-V2.

v

REPORT SUMMARY

BackgroundBearings of many configurations exist in rotating equipment at nuclear power plants. Reliableoperation and maintenance of these components is vital to sustained plant operation. The Frenchutility Electricité de France (EdF) has developed a series of technical papers on the subject ofbearing and related subjects. These documents provide background knowledge of bearing designand operation as well as an understanding of critical parameters for reliable bearing operation.Topics addressed in these papers should aid engineers and maintenance personnel introubleshooting bearing problems and in the development of effective maintenance programs.

Objectivex To share EDF’s research and knowledge

x To provide a vehicle to share this information with maintenance and engineering personnel inthe United States involved in the application, selection, or maintenance of bearings

ResultsThe EdF technical papers were translated to English and edited for use by NMAC members.Volume 1 was issued in the summer of 1999; and Volume 2, in 2000. Collection of thisinformation provides a valuable reference for future use.

TR-113059-V2

KeywordsDesign engineersPlant support engineeringPlant maintenance

EPRI Licensed Material

vii

PREFACE

This document has been developed based on a number of technical papers published aboutbearings by Electricité de France (EdF), the French utility company, in cooperation with localuniversity faculty members. The papers were published in French. With the permission of EdF,they have been translated and edited for use by NMAC members.

A number of recommendations in these papers have been introduced in the EdF plants.Therefore, the contents of these documents represent a combination of academic research testedthrough application experience in an operating nuclear power plant. This publication is not atypical “Bearing Guide” for use by the uninitiated. For those working with bearing application,selection, or maintenance, these papers provide insight into various problems and possibleremedies.


ix

CONTENTS

1 TECHNOLOGY AND TYPES OF ROTATING MACHINE BEARINGS ................................ 1-1

Executive Summary............................................................................................................ 1-2

1.1 Introduction ............................................................................................................... 1-3

1.2 General Makeup of a Roller Bearing ......................................................................... 1-3

1.3 Roller Bearings.......................................................................................................... 1-5

1.3.1 Stiff Ball Bearings ................................................................................................. 1-5

1.3.2 Swiveled Ball Bearings ......................................................................................... 1-5

1.3.3 Dismountable Ball Bearings.................................................................................. 1-6

1.3.4 Angular Contact Ball Bearings .............................................................................. 1-6

1.3.5 Cylindrical Roller Bearings.................................................................................... 1-6

1.3.6 Tapered Roller Bearings....................................................................................... 1-7

1.3.7 Swiveled Roller Bearings...................................................................................... 1-7

1.4 Thrust Bearings......................................................................................................... 1-7

1.4.1 Single-Effect (Direction) Ball Thrust Bearings ....................................................... 1-7

1.4.2 Double-Effect (Direction) Ball Thrust Bearings...................................................... 1-8

1.4.3 Cylindrical Roller Thrust Bearings......................................................................... 1-8

1.4.4 Tapered Roller Thrust Bearings............................................................................ 1-8

1.4.5 Swiveled Roller Thrust Bearings........................................................................... 1-8

1.5 Cages .......................................................................................................................1-9

1.6 Dimensioning and Interchangeability......................................................................... 1-9

1.7 Conclusion .............................................................................................................. 1-10

1.8 Bibliography ............................................................................................................ 1-11

2 ROTATING MACHINES SUPPORTING COMPONENTS.................................................... 2-1


2.1 History....................................................................................................................... 2-3

2.2 Introduction ............................................................................................................... 2-5

2.3 Purpose of Supporting Components.......................................................................... 2-6


x

2.3.1 Purpose of Bearings and Thrust Bearings ............................................................ 2-6

2.3.1.1 Dry Bearings ................................................................................................. 2-7

2.3.1.2 Grease Lubricated Bearings .......................................................................... 2-8

2.3.1.3 Roller Bearings and Thrust Bearings ............................................................. 2-8

2.3.1.4 Film Lubricated Bearings and Thrust Bearings .............................................. 2-9

2.3.1.5 Magnetic Bearings......................................................................................... 2-9

2.4 Purpose of Lubricants ............................................................................................. 2-10

2.4.1 Roller Bearings Lubrication................................................................................. 2-10

2.4.2 Film Lubrication of Bearings and Thrust Bearings............................................... 2-11

2.5 Dynamic Behavior ................................................................................................... 2-12

2.5.1 Description of a Pump ........................................................................................ 2-12

2.5.2 Interaction Forces............................................................................................... 2-13

2.5.3 Exciter Forces .................................................................................................... 2-13

2.6 Interaction Coefficients............................................................................................ 2-15

2.6.1 General Definition............................................................................................... 2-15

2.6.2 Coefficient for Bearings ...................................................................................... 2-16

2.6.2.1 Roller Bearings............................................................................................ 2-16

2.6.2.2 Film-Lubricated Bearings............................................................................. 2-16

2.6.2.3 Magnetic Bearings....................................................................................... 2-17

2.6.3 Coefficients for Labyrinths and Sealing Rings..................................................... 2-17

2.6.4 Coefficients for Impellers .................................................................................... 2-18

2.6.5 Coefficients for Other Components..................................................................... 2-18

2.7 Main Failures .......................................................................................................... 2-19

2.7.1 Roller Bearings................................................................................................... 2-19

2.7.2 Film Lubricated Bearings .................................................................................... 2-20

2.7.3 Wear of Sealing Joints........................................................................................ 2-22

2.8 Conclusion .............................................................................................................. 2-22

2.9 References.............................................................................................................. 2-24

3 EXPERIENCE FEEDBACK ON THE MAINTENANCE OF NUCLEAR POWERPLANT PUMP ROLLER BEARINGS ..................................................................................... 3-1


3.1 A Good Feedback Experience................................................................................... 3-3

3.2 Maintenance Performed on Roller Bearings .............................................................. 3-4

3.2.1 Monitoring During Operation................................................................................. 3-5


xi

3.2.1.1 Monitoring Bearing Temperatures ................................................................. 3-5

3.2.1.2 Monitoring Bearings: Towards Predictive Maintenance ................................. 3-5

3.2.2 The "Metravib" Defect Factor, a Method Used in Power Plants ............................ 3-5

3.2.2.1 Principle ........................................................................................................ 3-5

3.2.2.2 Monitoring Device.......................................................................................... 3-6

3.2.3 Implementation of the Defect Factor ..................................................................... 3-6

3.2.4 Mixed Conclusions Calling for Caution ................................................................. 3-7

3.2.5 Under-Load Operation: A Design Problem Difficult to Manage on a NuclearBase ............................................................................................................................. 3-7

3.2.6 Greasing: Hard-to-Find Compromises .................................................................. 3-8

3.2.6.1 Unsuited Grease Quality/Quantity/Periodicity ................................................ 3-9

3.2.6.2 Choice of Grease Volume to Inject Is Subject to the Same Type ofDifficulties.................................................................................................................. 3-9

3.2.6.3 Greasing Methodology .................................................................................. 3-9

3.2.6.4 Dysfunctional Operation of Grease Valves .................................................... 3-9

3.3 Conclusion .............................................................................................................. 3-10

4 SPARE PARTS FOR THE BEARINGS AND BALL BEARINGS OF THE ROTATINGMACHINE SWIVELING PROVISIONS ................................................................................... 4-1


4.1 Background and Goals.............................................................................................. 4-3

4.2 Tools ......................................................................................................................... 4-4

4.2.1 Identification ......................................................................................................... 4-4

4.2.2 Management ........................................................................................................ 4-4

4.3 Roller Bearing and Bearings...................................................................................... 4-5

4.3.1 Roller Bearings..................................................................................................... 4-5

4.3.2 Bearings ............................................................................................................... 4-6

4.3.2.1 Elliptical Journal Bearings.............................................................................. 4-6

4.3.2.2 Three-Pad Journal Bearings.......................................................................... 4-6

4.4 Policies ..................................................................................................................... 4-7

5 METHODS OF ANALYSIS AND FOLLOW-UP OF THE OILS IN TERMS OF THEMAINTENANCE OF ROTATING MACHINES ........................................................................ 5-1


5.1 Why Follow up on Oil Charges .................................................................................. 5-3

5.2 How to Follow up on an Oil Charge........................................................................... 5-3


xii

5.3 Sampling ................................................................................................................... 5-5

5.3.1 How to Perform Samplings ................................................................................... 5-6

5.4 What Can Be Expected From the Different Control Parameters ................................ 5-7

5.5 Conclusions ............................................................................................................ 5-14

6 R.R.A. (RHR) PUMP BEARING GREASE QUALIFICATION TESTS.................................. 6-1


6.1 Introduction ............................................................................................................... 6-3

6.2 Loading Specifications .............................................................................................. 6-3

6.3 Description of the Test Benches................................................................................ 6-4

6.4 Operating Mode ........................................................................................................ 6-7

6.4.1 Irradiation Resistance Tests ................................................................................. 6-7

6.4.2 Thermodynamic Tests .......................................................................................... 6-8

6.5 Results From Irradition Resistance Tests .................................................................. 6-8

6.6 Conclusion .............................................................................................................. 6-11

7 ROTATING MACHINE DYNAMIC SEAL TECHNOLOGY ................................................... 7-1


7.1 Design Diversity ........................................................................................................ 7-3

7.1.1 Listing of the Different Classifications ................................................................... 7-3

7.1.1.1 Classification According to the Existence or Lack of Relative MotionBetween the Leak-Proof Elements ............................................................................ 7-3

7.1.1.2 Classification According to the Direction of the Sealing Interstice .................. 7-3

7.1.1.3 Classification According to the Existence or Lack of Sealing Material............ 7-3

7.1.1.4 Classification According to the Existence or Lack of the Interstice................. 7-3

7.1.2 Criteria of Choice.................................................................................................. 7-4

7.1.2.1 Purpose......................................................................................................... 7-4

7.1.2.2 Thermodynamic Conditions........................................................................... 7-4

7.2 Labyrinth-Type Joints ................................................................................................ 7-5

7.2.1 Definition .............................................................................................................. 7-5

7.2.2 Description and Operating Principle...................................................................... 7-5

7.2.3 Parameters of Choice ........................................................................................... 7-6

7.2.3.1 Dimensioning Characteristics ........................................................................ 7-6

7.2.3.2 Influence on Vibrational Behavior .................................................................. 7-7

7.2.3.3 Wear of the Labyrinths .................................................................................. 7-8


xiii

7.3 Conclusion ................................................................................................................ 7-8

7.4 References................................................................................................................ 7-9


xv

LIST OF FIGURES

Figure 1-1 Single-Row Stiff Ball Bearing ............................................................................. 1-12

Figure 1-2 Introduction of the Balls in a Stiff Roller Bearing With No Loading Notch ............ 1-12

Figure 1-3 Stiff Ball Bearing With Two Sealing Joints........................................................... 1-12

Figure 1-4 Swiveled Ball Bearing ......................................................................................... 1-13

Figure 1-5 Dismountable Stiff Ball Bearing........................................................................... 1-13

Figure 1-6 Single-Row Angular Contact Ball Bearing ........................................................... 1-13

Figure 1-7 Double-Row Angular Contact Ball Bearing.......................................................... 1-14

Figure 1-8 Type NU Cylindrical Roller Bearing ..................................................................... 1-14

Figure 1-9 Type NU.E Cylindrical Roller Bearing.................................................................. 1-14

Figure 1-10 Type N Cylindrical Roller Bearing...................................................................... 1-15

Figure 1-11 Type NJ Cylindrical Roller Bearing.................................................................... 1-15

Figure 1-12 Type NUP Cylindrical Roller Bearing................................................................. 1-15

Figure 1-13 Type NJ Cylindrical Roller Bearing With Guiding Shoulder Ring ....................... 1-16

Figure 1-14 Type NN Double-Row Cylindrical Roller Bearing............................................... 1-16

Figure 1-15 Needle Bearing ................................................................................................. 1-16

Figure 1-16 Tapered Roller Bearing ..................................................................................... 1-17

Figure 1-17 Large Contact Angle Tapered Roller Bearing .................................................... 1-17

Figure 1-18 Swiveled Roller Bearing With Fixed Guiding Shoulder Rings ............................ 1-17

Figure 1-19 Swiveled Roller Bearing With Symmetrical Rollers and Mobile Guiding Ring .... 1-18

Figure 1-20 Swiveled Short Roller Bearing With Mobile Guiding Ring .................................. 1-18

Figure 1-21 Barrel Roller Bearing......................................................................................... 1-18

Figure 1-22 Single-Effect Thrust Ball Bearing....................................................................... 1-19

Figure 1-23 Single-Effect Thrust Ball Bearing With Spherical and Aligning Washers............ 1-19

Figure 1-24 Single-Effect Double-Row Thrust Ball Bearing .................................................. 1-19

Figure 1-25 Double-Effect Thrust Ball Bearing ..................................................................... 1-20

Figure 1-26 Cylindrical Roller Thrust Bearing ....................................................................... 1-20

Figure 1-27 Tapered Roller Thrust Bearing .......................................................................... 1-20

Figure 1-28 Swiveled Roller Thrust Bearing ......................................................................... 1-21

Figure 1-29 Comparison of the Different Series of Dimensions on the Most CommonRoller Bearings.............................................................................................................. 1-21


xvi

Figure 1-30 Adjustment, Machining Tolerances for Shafts and Casings............................... 1-22

Figure 2-1 Examples of Roller Bearings .............................................................................. 2-25

Figure 2-2 Examples of Film Lubricated Bearings ................................................................ 2-25

Figure 2-3 Operating Principles of Magnetic Bearings.......................................................... 2-26

Figure 2-4 Cross-Sectional View of a Multi-Stage Centrifugal Pump (Picture Allis –Chalmers)...................................................................................................................... 2-26

Figure 2-5 Forces Acting on the Rotor of a Multi-Stage Pump.............................................. 2-27

Figure 2-6 Spectral Analysis of the Vibration in a Pump (Wheel With 5 Buckets)Rotating at 3560 rpm [3] ................................................................................................ 2-27

Figure 2-7 Campbell Diagram in Standard Operation and Degraded Operation ................... 2-28

Figure 2-8 Example of Computation on the Response of an Imbalance in a Pump .............. 2-29

Figure 2-9 Schematic View of the Dynamic Coefficients of a Film Lubricated Bearing ......... 2-30

Figure 2-10 Schematic View and Location of the Sealing Rings on a Pump......................... 2-30

Figure 2-11 Schematic View of a Wheel and Swirl Break of the Balancing Piston on aPump............................................................................................................................. 2-31

Figure 3-1 PWR 900/1300MW Nuclear Base Number and Distribution of Roller BearingFailures Observed During 82/83...................................................................................... 3-3

Figure 6-1 Thermodynamic Aging Profile According to the K1 Profile of the R.C.C.E.......... 6-12

Figure 6-2 Schematic Sectional Drawing of the Test Bench Shaft With the RadialLoading System............................................................................................................. 6-12

Figure 6-3 Sectional Drawing of the Test Bench .................................................................. 6-13

Figure 6-4 Schematic View of Three Test Benches in the Irradiation ContainmentEnclosure ...................................................................................................................... 6-13

Figure 6-5 Schematic View of Test Benches in the Climatic Enclosure ................................ 6-14

Figure 6-6 Type A Grease, Status of the Numbers 1 and 2 Roller Bearings of theNumber 2 Bench After a Cumulated Radiation Dose of 235 kGy While Operating......... 6-15

Figure 6-7 Type B Grease, Status of the Number 3 Roller Bearing of the Number 2Bench After a Cumulated Radiation Dose of 245 kGy While Operating ......................... 6-16

Figure 6-8 Type B Grease, Status of the Number 4 Roller Bearing of the Number 2Bench After a Cumulated Radiation Dose of 245 kGy While Operating ......................... 6-16

Figure 6-9 Type C Grease, Status of the Number 4 Roller Bearing of the Number 2Bench After a Cumulated Radiation Dose of 300 kGy While Operating ......................... 6-17

Figure 6-10 Type B Grease, Status of the Roller Bearings of the Number 3 Bench MotorAfter a Cumulated Radiation Dose of 180 kGy While Operating .................................... 6-17

Figure 7-1 No Contact Seals ................................................................................................ 7-11

Figure 7-2 Schematic Diagram of an Interstice Joint ............................................................ 7-12

Figure 7-3 Schematic Diagram of a Grooved Labyrinth-Type Joint ...................................... 7-12

Figure 7-4 Types of Interstice Joints .................................................................................... 7-13

Figure 7-5 Types of Interstice and/or Labyrinth-Type Joints ................................................. 7-14

Figure 7-6 Schematic Diagram of a Multi-Cellular Pump ...................................................... 7-15


xvii

Figure 7-7 Table of leakage rates as a function of geometrical shapes and rotationspeeds .......................................................................................................................... 7-16

Figure 7-8 Annex 1 Evolution of Leakage Rate as a Function of the PressureDifferential ..................................................................................................................... 7-17

Figure 7-9 Annex 2 Evolution of the Dynamic Coefficients as a Function of the FlowRate .............................................................................................................................. 7-18

Figure 7-10 Influence of Eccentricity on Dynamic Coefficients ............................................. 7-19


xix

LIST OF TABLES

Table 1-1 Different Types of Rolling Bearings ........................................................................ 1-5

Table 2-1 Classification of the Origin of Vibration on a Multi-Stage Pump ............................ 2-14

Table 2-2 Causes of Roller Bearing Failures......................................................................... 2-20

Table 2-3 Causes of Film Lubricated Bearing Failures .......................................................... 2-21

Table 6-1 Types and Characteristics of the Roller Bearings Mounted on the R.R.A.Pumps at the Bugey PWR Plant and on the N4 Plants .................................................... 6-5

Table 6-2 Characteristics and Loading of Test Bench Roller Bearings .................................... 6-6

Table 6-3 Type of Roller Bearing and Mechanical Loading Values for IrradiationResistance Tests............................................................................................................. 6-9

Table 6-4 Summary of the Results of Irradiation Resistance Tests ....................................... 6-10


1-1

1 TECHNOLOGY AND TYPES OF ROTATING MACHINEBEARINGS

Translated from Electricité de France Document

No. 97NB00107

Author:D. Buchdahl

Translator:Vince Cardon

EPRI Project Manager:Mike Pugh


Technology and Types of Rotating Machine Bearings

1-2

Executive Summary

Rotating machine bearing technology stemmed from the simple need to reduce friction forces. Itsapplication to the practical cases of mechanics led to a wide variety of implementations. A highnumber of bearing types exists, and the choice of these is governed by the conditions ofoperation and use.

Bearing designs have been standardized, thus ensuring complete geometric interchangeabilitybetween one supplier and another. However, in operational terms, any change in supplier for agiven application requires verification of the mechanical dimensioning of the bearing. Therefore,it is essential to refer to the catalogues published by the bearing manufacturers prior to anychanges being made. These catalogues provide all the data required for the bearing dimensioningand assembly.

This document briefly describes the general make-up of a bearing and reviews different types ofbearings, thrust ball bearings, and thrust roller bearings.



1-3

1.1 Introduction

Since the beginning of time, laziness, the spirit of invention, and chance have been the cause forthe great discoveries that have made humanity leap forward.

The wheel is one of these great discoveries.

Man has always wanted to leave tangible marks in his wake. To do this, he had to move heavyloads. Large structures indeed symbolize strength and power.

Human strength can be easily increased through the use of a lever. Nevertheless, the mechanics'foe still remains, and that is friction.

Friction forces can be reduced by applying a lubricant between the two surfaces in motion.Another method consists of reducing the surface of contact.

That is probably how rollers, the ancestors of the wheel, originated.

The Egyptians used this technology widely and left some masterpieces behind.

Rollers changed surface friction into friction along a line. The sliding became rolling. Then, inthe sixteenth century, Leonardo de Vinci made scientific studies on force and motion. Anotherartist, Benvenuto Cellini, made the first thrust bearing in 1534.

Since that time, roller bearings have been perfected, and their use has spread to a large variety ofapplications.

1.2 General Makeup of a Roller Bearing

The roller bearing is a mechanical component that allows the motion of two parts, one relative tothe other, and their guides with a minimal amount of friction.

Many types of roller bearings exist. Their range of application is wide, but roller bearingsgenerally share the following common components:

x Outer ring

x Inner ring

x Rolling elements

x Cage

Sealing joints are sometimes mounted on the roller bearings.



1-4

The rolling elements can be of different shapes:

x Balls

x Needles

x Cylindrical rollers

x Tapered rollers

x Spherical rollers

The purpose of the cage is important: it separates and spreads evenly the rolling elements on therolling track during operation. During assembly, it maintains the rolling elements together.

The following materials are used to produce roller bearings:

x Most often, steel 100 C6 (SAE 52100)

x Stainless steel

x Ceramic material for certain very specific applications

x The cages are most often made with buckled sheet or polyamide material (reinforcedfiberglass) or with some heavy gauge brass or steel.

According to ISO norms, there are two main applications for roller bearings:

x Radial

x Axial

Some authors prefer to separate roller bearings into two groups:

x Radial roller bearings

x Thrust bearings

Most roller bearings can support axial loads. Some thrust bearings can support radial loads.

The angle of contact (angle between the direction of the load on a rolling element and theperpendicular to the axis of the roller bearing) determines the two groups. Thrust bearings havean angle of contact more that 45q.

The loading capacity of a roller bearing depends, among other things, on the contact surfacebetween the rolling elements and the rolling track, thus on the size and number of rollingelements.



1-5

Table 1-1Different Types of Rolling Bear ings

Axial Roller Bearings Radial Rolling Bearings

Ball thrust bearings Roller thrustbearings

Radial ball bearings Cylindrical rollerbearings

Single-row single-effect(direction) ball thrust

bearings

Cylindrical rollerthrust bearings

Single-row ball bearings Cylindrical rollerbearings

Single-row single-effect(direction) thrust bearingswith a spherical washerand an aligning washer

"Magneto" type rollerbearings

Tapered roller bearings

Double-row double-effect(direction) thrust ball

bearings

Tapered rollerthrust bearings

Single-row angularcontact ball bearings

Swiveled double rowspherical roller bearings

Double-row angularcontact ball bearings

Double-row double-effect(direction) thrust ball

bearings with a sphericalwasher and an aligning

washer

Spherical rollerthrust bearings

Swiveled double-rowball bearings

Needle bearings

1.3 Roller Bearings

1.3.1 Stiff Ball Bearings

The outer and inner rings on this type of roller bearings have deep grooves in which the balls roll(see Figures 1-1 and 1-2). During the manufacturing of this roller bearing, the rings are mountedeccentrically with respect to one another so the balls may be put in place. Even distribution of theballs is obtained with the cage, which generally consists of two parts assembled together.

Shields or joints are often mounted on this type of roller bearings (see Figure 1-3). Their purposeis to protect the lubricant from pollutants or foreign particles. They also help the lubricantcirculate evenly inside the roller bearing while in operation.

The number of balls and their size determine the radial and axial loading capacity. For highrotation speeds, stiff ball bearings can have an axial loading capacity higher than that of a thrustbearing.

1.3.2 Swiveled Ball Bearings

These roller bearings have two rows of balls (see Figure 1-4). The inner ring has two rollingtracks, and the rolling track of the outer ring is spherical.



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This design allows for the tilting of the inner ring and the balls with respect to the outer ring.However, the swiveling angle is limited because the balls must not leave their rolling track.

This design is particularly interesting for assemblies with overhangs with a bending shaft ormisalignment.

1.3.3 Dismountable Ball Bearings

The groove of the inner ring on this roller bearing is not as deep as that on a stiff rollerbearing. The outer ring has only one shoulder. Thus, this ball bearing can be disassembled.(see Figure 1-5).

This design implies that two roller bearings be mounted as a pair, one facing the other.

Because of its design, this type of roller bearing has a low axial load capacity. Therefore, itsapplication is limited to small machines.

1.3.4 Angular Contact Ball Bearings

The shape of the grooves on the rings of these roller bearings is designed to obtain an anglebetween the result of the applied load and perpendicular to the bearing axis (see Figure 1-6).This angle (called the contact angle) is generally between 30q and 40q.

In general, these roller bearings cannot be disassembled. Their loading capacity is particularlyhigh.

These roller bearings are generally mounted in pairs, one facing the other. Their rolling tracksform either an O or a X. This assembly can support radial loads from either direction.

In some cases, it is possible to mount two roller bearings one against the other in the same way(T configuration) so they share the axial load. Then, one or two additional roller bearings aremounted facing the first two.

Double row angular contact ball bearings (see Figure 1-7) are designed to support largealternating axial loads.

1.3.5 Cylindrical Roller Bearings

The rolling elements on this type of bearing are cylindrical (see Figures 1-8 through 1-13). Theyare guided by shoulders on one of the rings. The cage maintains the rollers inside the ringwithout a shoulder, called the free ring.

This configuration lets the inner ring move relative to the outer ring. This displacement occurswithout any noticeable resistance during rotation. This type of roller bearing allows axialdisplacement of the shaft with respect to the stator (for example, different thermal expansion).



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The rollers are not perfectly cylindrical. Their shape varies with the manufacturer. They areslightly bulged at the ends. This gives a better distribution of the load on the rolling tracks, thusavoiding large loads between the rollers and the rings. This also makes the roller bearings lesssensitive to the bending of the shaft.

Because of the increase in contact surface, these roller bearings support large radial loads.

These roller bearings can be mounted by mounting the inner and outer rings separately.

Variations of this type of roller bearing have two rows of cylindrical rollers (see Figure 1-14).This increases the loading capacity without making the system bulkier in the radial direction.

When the available space is restricted, another variation called needle bearings can be used (seeFigure 1-15).

1.3.6 Tapered Roller Bearings

In this type of roller bearing, the rolling elements are tapered rollers (see Figure 1-16). Therollers and rolling tracks also have a particular shape to insure that the pressure of contact is welldistributed.

Tapered roller bearings are generally mounted in pairs, one facing the other. Their axial andradial loading capacity is high. The axial loading capacity depends on the angle of contact of therollers, generally between 12q and 16q. When the axial load is high, this angle may rangebetween 28q to 30q (see Figure 1-17).

1.3.7 Swiveled Roller Bearings

These roller bearings have two rows of rollers (see Figures 1-18 through 1-20). The inner ringhas two rolling tracks, and the rolling track of the outer ring is spherical.

This configuration is similar to that of the swiveled ball bearing.

Because of the inclination of the rolling elements, the axial load capacity of these roller bearingsis very high.

The barrel roller bearing aligns automatically (see Figure 1-21). Its axial loading capacity is low.In fact, the axis of the rollers is parallel to the axis of rotation of the roller bearing.

1.4 Thrust Bearings

1.4.1 Single-Effect (Direction) Ball Thrust Bearings

These thrust bearings consist of two washers: one lodging washer (outer ring) and a shaft washer(inner ring). A row of balls separates the two washers (see Figure 1-22).



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The supporting surface of the lodging washer is either flat or spherical. In the latter case, thelodging washer is placed on an aligning washer that is also spherical (see Figure 1-23).

This type of single-effect thrust bearing with a 90q angle of contact can support only axial loadsin one direction. It will not support any radial load.

An increase in the axial loading capacity is obtained by interposing two rows of balls (see Figure1-24). In this case, the shaft washer consists of two concentric parts. An elastic plate The evenlydistributes the load on the two rolling tracks.

1.4.2 Double-Effect (Direction) Ball Thrust Bearings

This type of thrust bearing can be described as the superposition of two single-effect ball thrustbearings (see Figure 1-25). The shaft washer, located in the center, has a rolling track on bothsides.

Double-effect ball thrust bearings support axial loads in either direction. They cannot supportany radial load if the angle of contact is 90q.

1.4.3 Cylindrical Roller Thrust Bearings

Cylindrical roller thrust bearings have the same basic design as ball thrust bearings (see Figure 1-26). The washers can have guiding shoulders. They can also have several rows of rollers.However, the rotation of the rollers cannot possibly be kinematically correct. The longer therollers, the worse the rotation of the thrust bearing becomes. Therefore, the use of this type ofthrust bearing is limited to low rotation speeds.

1.4.4 Tapered Roller Thrust Bearings

Tapered roller thrust bearings have a tapered and a flat rolling track (see Figure 1-27). Therolling track can, for example, be located on the shaft washer.

This type of thrust bearing accepts a displacement of the shaft parallel to its axis, without causingany stress to the thrust.

1.4.5 Swiveled Roller Thrust Bearings

Swiveled roller thrust bearings have the same basic design as the swiveled roller bearings (seeFigure 1-28). The large inclination of the rollers provides a very large axial loading capacity.This thrust bearing can also support radial loads.

However, the shape of the rollers induces heavy load between the long base and the guidingshoulder. Much care goes into designing the guiding faces. The long bases of the rollers and theguiding surface of the shoulder are spherical, with slightly different radii. This allows for the



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creation of an oil film during rotation. The recommended uses for this type of thrust bearing andfor other types of roller bearings are different.

1.5 Cages

The cages do not participate in transferring the load on the roller bearing. They are subjected tocentrifugal forces and shocks.

They are guided radially by the rolling elements or centered on one of the rings. The double-rowroller bearings either have a common cage or two half-cages.

Cages made of sheet metal (cut and buckled) or bulk polyamide equip the majority of the ball orroller bearings. They are inexpensive and perfectly suited for mass production. In addition, theyare light in weight and leave plenty of room for the lubricant.

When roller bearings are big or when the operating conditions are severe (for example, highrotation speed or shocks), it is better to use massive steel, brass, cast iron, or even some lightalloy cages.

The use of polyamides is limited to applications where the operating temperature is low.

1.6 Dimensioning and Interchangeability

Roller bearings are used everywhere by a large number of people. Very early, end users andmanufacturers felt the need to rationalize the production of roller bearings with the followinggoals:

x Improving reliability

x Manufacturing in mass

x Reducing costs

x Interchangeability

To address this common need of the end users and manufacturers, the International StandardOrganization (ISO) established dimensioning standards for roller bearings and thrust bearings(ISO 15, ISO 355, and ISO 104).

For the series of shaft diameters (d), these norms set different series of outer diameters anddifferent widths (see Figure 1-29).

Thus, roller bearings are standardized and interchangeable across manufacturers. However, wewant to point out that only the outer dimensions of the roller bearing are standardized. Therefore,for the same level of bulkiness, the number of rolling elements may change. Each manufacturerhas its own expertise, its own technology. Consequently, it is possible to find roller bearings thatlook the same, but have radically different mechanical characteristics.



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In regards to dimensioning, roller bearings having the same standardized designation areinterchangeable. Nevertheless, it is essential to check that their mechanical as well as mountingand operating characteristics are similar. In all cases, one must verify that the roller bearing iswell suited for the application.

As technology advances, manufacturers improve their products, and the end user benefits. Theseimprovements mainly concern increases in loading capacity. Unfortunately, contrary to what onemight think instinctively, an increase in the loading capacity is not always welcome in allapplications.

For example, when a roller bearing is lightly loaded, its replacement by a geometrically similarroller bearing with a higher loading capacity will increase the unloading. The rolling elementsthen tend to slide, and this sliding quickly destroys the tracks and the rolling elements.

Good operation and the lifetime of a roller bearing are conditioned by the tolerance on theadjustment on the shaft or in its casing. The choice of these tolerances depends more particularlyon:

x Magnitude and direction of the load

x Rotation speed

x Level of machine vibration

x Operating temperature

These tolerances are within a short range in the tolerance range for ISO smooth parts (seeFigure 1-30).

1.7 Conclusion

Roller bearing technology stemmed from a simple idea: to reduce friction forces.

Its application to practical cases in mechanics leads to a large variety of designs. Many types ofrolling bearings exist, and choosing the type of roller bearing depends on operating conditions.

Roller bearings are standardized. The norm guarantees a perfect geometrical interchangeabilityfrom one manufacturer to another. However, from a functional perspective, any switch insupplier, for a given application, requires that the mechanical characteristics of the roller bearingbe checked. Before any modification is made, it is essential to refer to the manufacturers'catalogues where the necessary data on dimensioning and mounting roller bearings can be found.



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1.8 Bibliography

1. Les roulements. Description, théorie, applications, Arvid Palmgren.

2. Mémento ISO roulements 1988.

3. Notice technique sur les centrales thermiques. Fascicule Nq51B.

4. Roulements FAG. Programme standard. Catalogue WL 41 510/2FB.

5. S.K.F. Catalogue général 1989.10.

6. Revue des roulements 231



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Figure 1-1Single-Row Stiff Ball Bearing

Figure 1-2Introduction of the Balls in a Stiff Roller Bearing With No Loading Notch

Figure 1-3Stiff Ball Bearing With Two Sealing Joints



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Figure 1-4Swiveled Ball Bearing

Figure 1-5Dismountable Stiff Ball Bearing

Figure 1-6Single-Row Angular Contact Ball Bearing



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Figure 1-7Double-Row Angular Contact Ball Bearing

Figure 1-8Type NU Cylindrical Roller Bearing

Figure 1-9Type NU.E Cylindrical Roller Bearing



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Figure 1-10Type N Cylindrical Roller Bearing

Figure 1-11Type NJ Cylindrical Roller Bearing

Figure 1-12Type NUP Cylindrical Roller Bearing



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Figure 1-13Type NJ Cylindrical Roller Bearing With Guiding Shoulder Ring

Figure 1-14Type NN Double-Row Cylindrical Roller Bearing

Figure 1-15Needle Bearing



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Figure 1-16Tapered Roller Bearing

Figure 1-17Large Contact Angle Tapered Roller Bearing

Figure 1-18Swiveled Roller Bearing With Fixed Guiding Shoulder Rings



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Figure 1-19Swiveled Roller Bearing With Symmetrical Rollers and Mo bile Guiding Ring

Figure 1-20Swiveled Short Roller Bearing With Mobile Guiding Ring

Figure 1-21Barrel R oller Bearing



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Figure 1-22Single-Effect Thrust Ball Bearing

Figure 1-23Single-Effect Thrust Ball Bearing With Spherical and Aligning Washers

Figure 1-24Single-Effect Double-Row Thrust Ball Bearing



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Figure 1-25Double-Effect Thrust Ball Bearing

Figure 1-26Cylindrical Roller Thrust Bearing

Figure 1-27Tapered Roller Thrust Bearing



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Figure 1-28Swiveled Roller Thrust Bearing

Figure 1-29Comparison of the Different Series of Dimensions on the Most Common Roller Bearings



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Figure 1-30Adjustment, Machining Tol erances for Shafts and Casings


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2 ROTATING MACHINES SUPPORTING COMPONENTS


No. 97NB00106

Author:Chan Hew Wai C.




Rotating Machines Supporting Components

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Executive Summary

This document deals with supporting components and succinctly illustrates the most importantpoints for correct operation of rotating machines:

x Choice of bearing and attendant lubricant type and technology

x Dimensioning and/or assessment of bearings with appropriate tools

x Monitoring of bearings and lubricants

x Compilation and use of a bearing failure knowledge base

x Effect of the supporting components on the dynamic behavior of the machine



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2.1 History

As soon as man wanted to make one piece rotate inside another, the problem of guiding it arose,and the bearing was born. The first needs appeared mainly in transportation and the woodenwheel; the cart was invented around 4000 BC. Then, the Romans and the Greeks developedwheel usage with vegetable and animal oil lubricants.

It quickly became obvious that friction and, thus, wear would represent a major problem to thesenew mechanisms. The use of solid lubricant was the first answer to the problem.

Additional progress came at the end of the Middle Ages with the production of numerousmechanical clocks supported by lubricated iron journal bearings, even brass, to minimize frictionand wear.

After the Middles Ages, Leonardo de Vinci was the first to conduct studies quantifying frictionforces and to introduce the notion of friction coefficient. He studied and proposed devices toreduce friction. His ideas were very much in advance of his time. Indeed, he proposed to replace,in the bearings, the sliding of the shaft inside the bearing by the rolling of the shaft on rollers: theancestor of the bearing was born.

In the sixteenth century, machines became more and more complex, and bearings became moreand more sophisticated. Steel shafts supported by copper rings to minimize wear appeared duringthis period. This technology is still in use today with brass (alloy of copper and tin) instead ofcopper.

In the seventeenth century, the first theoretical studies on friction were made, and the first laws(law of Amontons) were formulated, some of which are still is use today.

Also, numerous experiments on bearing friction were conducted, and Jacob Rowe designedmachines with rollers driven by the shaft, somewhat like today's roller bearings.

At the beginning of the industrial revolution, Coulomb worked on several devices to determinethe main parameters (for example, materials, lubrication, contact surface, and loading conditions)having an effect on sliding and rolling friction. Although limited, that work permitted a definiteprogress in the area and has been used as reference for over 150 years, including today. It is alsoduring that period that the technology greatly progressed.

In the eighteenth century, machines supported by smooth bearings and even roller bearings weremade.

In that era, the true prototypes of roller bearings still in use today appeared. In fact, Vaughampatented a ball bearing used on wheel axes.

The invention of the bicycle in the nineteenth century provided the push to the industrialdevelopment of the roller bearing. In the meantime, research on roller bearing design wasundertaken, both theoretically and experimentally, particularly by Hertz in 1881.



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This period was also marked by the development of steam machines and mineral lubricants. Thisevidently invited many scientists to investigate or explore other alternatives in this domain. Wewill mainly mention Him's work in hydrodynamic lubrication. He showed the direct relationbetween the viscosity of the lubricant and torque due to friction. However, the most importantdiscovery of this time is undeniably attributed to Tower, who showed experimentally theexistence of a hydrodynamic pressure in lubricated smooth bearings. His work is the basis for thetheory on hydrodynamic lubrication.

From that work, several theoreticians have conducted studies to try to explain Tower'sobservations. Navier worked out the general equations for the motion of a fluid as a function ofviscosity, representative of the resistance to shear inside the fluid. Poiseuille derived theequations of the flow of a fluid inside a small tube and gave his name to the unit of dynamicviscosity in the SI system.

Reynolds really laid the theoretical bases of today's hydrodynamic lubrication in an 1886technical paper. He derived the basic equation for hydrodynamic lubrication, known as theReynolds equation.

The start of the twentieth century confirmed scientific interest in the area, and it is virtuallyimpossible to search exhaustively all the articles that have been published since.

Even before computers existed to solve Reynolds' equation, the equation was appliedsuccessfully to real cases when simplifying assumptions were used (for example, finite widththrust bearing and short bearing) and helped increase the knowledge in this field.

Cameron and Wood proposed the first numerical methods in 1949.

Thermal effects also play a major role in lubrication. In 1962, Dowson, based on Kingsbury'swork, presented equations that model thermodynamic lubrication.

Couette, Taylor, and Wilcock showed the existence of vortices inside the fluid when they usedlower viscosity fluids and greater rotating speeds. Models that represent these phenomena havebeen proposed and used only recently.

From a technological perspective, Mitchell in Australia patented a tilting pad bearing for use on athrust bearing in 1905. He showed that this device reaches an inclination to obtain the pad’soptimum load if the swivel is at the right place. Practically at the same time in the United States,Kingsbury designed the tilting pad for the bearing.

Hydrostatic lubrication consists of injecting a pressurized fluid to separate the surfaces incontact. This technology was known before the beginning of this century. In 1917, Rayleigh wasthe first to present the principle of a hydrostatic axial thrust bearing and to calculate the load andthe corresponding friction torque. Hydrostatic systems are mainly used to guarantee very preciseguiding without wear and with friction independent of the load.

The most recent studies involve lubrication where thermal effects and elastic deformations of thecontact area are taken into account simultaneously: elastothermohydrodynamics. In that case, the



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pressure and/or the temperature gradient are high enough to elastically deform the surfaces. Theeffect is to affect the behavior of the fluid. The first studies were conducted by Hertz andinvolved two parts in contact without any fluid. With the development of computers, iterativenumerical methods solved these problems.

As for roller bearings, the development of the automobile and mechanical industry in generalgave manufacturers an incentive to propose high performance products that both support heavyloads with little friction and rotate at high speeds.

The history of tribology (study of friction) shows that the major concern of the tribologistsresides in the knowledge of all the phenomena that exist in a mechanism where frictiondominates.

2.2 Introduction

In a rotating machine, bearings are crucial components that must guide the moving part andtransfer loads throughout the lifetime of the machine.

Therefore, determining the type of bearing to use is an essential step during the design stage. Itdepends on several factors associated to the operating constraints (for example, speed, loading,and environmental conditions).

Then, the type of technology and lubricant best adapted for the application must be chosen.

In any case, the focus must remain on minimizing wear and thus friction.

In fact, friction remains one of the main areas of concern for researchers and machine operators.The laws that regulate friction are well documented in certain areas (for example, dry friction,hydrodynamic lubrication, and hydrostatic lubrication), but can still be perfected in other areas(for example, thermohydrodynamic lubrication and mixed lubrication).

These studies have helped develop several tools now available to designers, operators, andexperts.

In addition, due to the wide number of applications in industry, all the different types of bearingfailures are practically known and have been analyzed and categorized.

In the case of a bearing failure, that knowledge base is very useful for technicians working on-site or in a laboratory.

Moreover, to avoid potential incidents and to better organize maintenance, monitoringtechniques developed for these vital devices are already in operation in a large number ofindustrial sites.

These aspects are briefly mentioned here with the details available in specific articles referencedthroughout this paper.



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All these elements come together to reassure and help the end user concerned with his machine’svibrational behavior to use the machine safely and maintain it easily.

Different studies and experiments conducted over the past few years show that all theinteractions with the rotor must be taken into account to have a picture closest to reality. It isparticularly true for areas of the bearing where the film is relatively thin (for example, bearings,seals, hubs, and balancing pistons) [3].

A multi-stage centrifuge pump illustrates this last point.

It appears that bearings determine the good operation of a rotating machine because of twothings: (1) resistance and lifetime of the bearing as a component and (2) active participation inthe dynamic behavior of the machine.

2.3 Purpose of Supporting Components

In a rotating machine, radial and axial static loads (for example, weight, hydraulic forces, andmagnetic forces) and dynamic loads can be either hydraulic or mechanical (for example,imbalance, re-circulation, and turbulence).

Static load. Forces with a constant direction in magnitude. These forces have different origins,such as weight, constant hydraulic load, and constant magnetic force.

Dynamic load. Forces with a time varying direction and magnitude. These forces can bemechanical and/or hydraulic, such as imbalance, instabilities, and cavitation.

2.3.1 Purpose of Bearings and Thrust Bearings

The primary role of a bearing or thrust bearing is to assure the following functions:

x Rotation: guiding rotation

x Load transfer: withstanding radial and axial loads

x Positioning: precise positioning

x Safety: withstanding loads in an accidental situation

The bearing (or thrust bearing) is a device where contact phenomena appear when twocomponents are in contact and moving with respect to one another. Friction occurs there and canbecome very intense if the device is not lubricated. The laws involving dry friction (Coulomb’slaw) have been well known for a long time and introduce the notion of sliding frictioncoefficient.

High friction coefficient values may be needed in some cases and are used in such devices asbrakes, clutches, and belt transmissions.

However, in most cases, friction should be minimized because it often leads to wear, heating,and eventually, gripping.



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To prevent such problems, several solutions are proposed:

x Treat with an anti-wear coating: dry bearings

x Decrease the friction by using a paste-like lubricant: grease lubricated bearings

x Replace sliding with rolling: roller bearings

x Separate the two parts facing each other using a third component or lubricating film that mustbe able to form, to be maintained, and to present a low resistance to shear: this ishydrodynamic and hydrostatic lubrication.

x Separate the two parts without using a third component: magnetic bearings

The choice of the type of bearing depends on several factors, from the loading conditions tospeed, environmental conditions, and cost.

In any case, the initial determination of the bearing type essentially involves the following threeparameters: loads, shaft rotation speed, and available space. Secondarily, the lifetime and ease ofmaintenance can be considered.

This part has been developed in another article presented during this seminar.

This article mentions the first, second, and fifth points but will not look at the details since thesetypes of bearings are practically never mounted on EdF’s machines.

Lubricant. The term lubricant refers to any interposed compound that opposes a light resistanceto shear and reduces the wear of the moving parts. Thus, it can be liquid, solid, or gas.

Bearing (or thrust bearing). The term bearing (or thrust bearing) is somewhat ambiguousbecause it sometimes refers only to the part subjected to friction (journal bearing) and sometimesto the complete ensemble: journal, shaft, and lubricant.

In either case, the possible performances are obtained by combining these three elements.

2.3.1.1 Dry Bearings

The good operation of a dry bearing mainly requires choosing appropriate materials and thesurface coating best suited for the required operating conditions. In this type of bearing, it seemsthat the heat generated in the areas of contact represents the determining factor for the operatinglifetime. Therefore, high rotation speeds are to be avoided. Generally, the shaft is made ofsurface-coated steel to avoid the penetration of abrasive materials. The part subjected to frictionis made of PTFE (Teflon), which presents a good resistance to temperature and corrosion, orPTFE mixed with various components, such as fiberglass, brass, and graphite to help evacuatethe heat.



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2.3.1.2 Grease Lubricated Bearings

In a dry bearing, the contact occurs mainly at rough spots. If grease is added, it becomes trappedand compressed in the cavities; and the friction coefficient decreases between the rough spotswhen contact occurs. The greasing is then said to be unctuous.

So, grease-lubricated bearings are better than dry bearings, but their use remains limited to low-speed applications.

The part subjected to friction is generally made out of brass.

2.3.1.3 Roller Bearings and Thrust Bearings

In a roller bearing (see Figure 2-1), the fundamental idea is to replace sliding by rotation. Aroller bearing is generally made of two rings between which are located rolling elements (balls,rollers, and needles). These allow the relative rotation of the two rings by positioning one withrespect to the other.

Because the roller bearing does not generate much internal friction, rotation speeds can be high.The ball is the rolling part that generates the least amount of friction because it only has onepoint of theoretical contact with its track.

A large variety of roller bearings exist according to the purpose and conditions to which they aresubjected. These include the following:

x Radial or angular contact ball bearings

x Cylindrical or tapered roller bearings

x Needle bearings

Roller bearings are capable of supporting radial and axial loads, depending on the choice of thebearing. However, thrust bearings are more adaptable for supporting large axial loads. However,thrust ball bearings can support only axial loads. However, swiveled roller thrust bearings canalso support radial loads under certain conditions.

Although the application for roller bearings is widespread in industry, roller bearings havelimitations: the load can cause element fatigue, and the speed can induce various mechanicaleffects (for example, vibrations and centrifugal forces) in addition to the limitations alreadyassociated with temperature and cage resistance.

The choice of the lubricant is essential to the lifetime of the roller bearing. Depending oncircumstances, the lubricant can be oil (for example, oil bath, oil fog, or circulating oil), standardroller bearing grease, or special grease (for example, grease for high or low temperatures, heavyloads, or high speeds).

All the details are developed in specific articles presented throughout this seminar.



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2.3.1.4 Film Lubricated Bearings and Thrust Bearings

With film-lubricated bearings and thrust bearings, we enter the world of hydrodynamic and/orhydrostatic lubrication.

Several designs can be used according to needs:

x Fixed shape (for example, cylindrical or with lobes) (see Figure 2a)

x Tilting bearings (see Figure 2b)

x With different types of plugs for hydrostatic bearings

The main principle is to separate the two surfaces in contact by interposing a third compound, orlubricating fluid. Thus, the friction coefficient is considerably reduced and the fluid evacuatesheat generated in the contact area.

The lubricant can be an incompressible viscous fluid or a gas. In the latter case, the bearings aresaid to be aerodynamic or aerostatic.

In lubrication, viscosity is the most important physico-chemical parameter because it determinesfriction losses, loading capacity, and film thickness.

The viscosity of a fluid is a measure of how easily its molecules slide with one another.

The application of hydrodynamic bearings and thrust bearings covers a large range of operation,and these bearings appear well suited for very heavy loads rotating at high speeds.

Inversely, they do not operate well at low speeds and light loads. There are risks of instabilityand contact.

On the other hand, hydrostatic devices can be used over the entire range of loads and speeds.

The details are developed in specific articles presented throughout this seminar.

Viscosity of a fluid. According to the standard NF T 60-100, the viscosity of a liquid is theproperty of that liquid defined as the resistance that its molecules oppose to a force that tends todisplace them through internal sliding.

Viscous fluid. A viscous fluid is a fluid with a resistance to shear (viscosity) that cannot beignored.

2.3.1.5 Magnetic Bearings

The peculiarity of magnetic bearings and thrust bearings resides in the absence of lubricantbetween the rotor and the stator. The phenomenon is called magnetic levitation.

A distinction must be made between passive magnetic bearings and active magnetic bearings(see Figure 3).



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The passive magnetic bearing is generally made with permanent magnets that use eitherrepulsive forces or aligning forces. These bearings have an acceptable load capacity.

Active magnetic bearings always works with attractive forces induced by electromagnets whosevariable fields can be monitored by position detectors and feedback loop electronics.

Therefore, passive and active magnetic bearings operate without mechanical contact. There isneither lubrication nor wear, and friction due to magnetic losses is minimal.

The application of magnetic bearings is theoretically broad, but remains technologically limitedby manufacturing mechanical constraints (for example, behavior of the fretted sheets andbulkiness). In reality, the use of magnetic bearings is mainly limited by the complexity of thedevice and its cost, a dissuasive factor for the most common applications.

2.4 Purpose of Lubricants

Lubrication is essential for the good operation of bearings. In fact, the lubricant is the thirdcompound interposed between the moving parts (rolling components or rotor) and the fixed parts(rolling tracks and journals) to minimize the friction and wear of the elements in contact.

Therefore, its role is major in a lubricated contact, and the aging behavior of its variousproperties is a permanent concern for the user.

Because lubricants are consumables, replacement must not to be overlooked. Frequentreplacements have cost implications. Regular analyses of oil samples can be of great help to theoperator. This also raises the problem of supplying the lubricant to the sites, stocking it, andgetting rid of the waste.

All parts in this chapter are developed in more detail in specific articles presented throughout thisseminar.

2.4.1 Roller Bearings Lubrication

To operate reliably, roller bearings must be correctly lubricated to avoid contact between themoving parts and the rolling tracks to avoid wear and to protect the surfaces from corrosion.

For roller bearings, the lubricant can be oil or grease depending on operating conditions(temperature, speed, and environmental conditions) of the roller bearing. The lubricant alsoprotects the roller bearing against oxidation and external pollution and helps evacuate the heatgenerated inside the bearing.

The lubricating power of grease or oil decreases with time as a result of mechanical constraints,aging, and contamination by foreign compounds. In fact, sealing devices may be mounted toavoid the penetration of impurities or humidity and also to hold the lubricant inside the rollerbearing. The choice of such devices depends on the nature of the lubricant, the peripheral speedof the shaft, available space, and cost.



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Thus, rolling bearings must be relubricated with a predetermined frequency. This relubricationmust always be done when the bearing is still lubricated in a satisfactory manner. Therelubrication period is not simple to determine because it depends on several complex factors.Therefore, this relubrication frequency is often determined statistically.

Some rolling bearings have integrated joints that are generally greased for life. This simplymeans that the operating lifetime of the grease is greater than the operating lifetime of the rollerbearing.

Oil must be used for lubricating roller bearings that operate at high speeds and hightemperatures. In all cases, it is better to use oils whose viscosity does not vary with temperature.

The lifetime of the roller bearing is directly related to the capacity of the lubricant to keep itshigh speed and temperature properties under the given operating conditions of the roller bearing.

To avoid a bad surprise, roller bearings must be checked regularly. To do this, some simplemethods are used: listening for noise during operation, monitoring of temperature, andmonitoring of the lubricant.

2.4.2 Film Lubrication of Bearings and Thrust Bearings

In hydrodynamic and thrust bearings, the lubricant is important because it determines practicallyall the static (for example, flow rate, friction torque, and load) and dynamic (stiffness anddamping) characteristics of the bearing.

Lubricating oils can be synthetic or mineral with several additives according to the application.Additives are used to improve certain characteristics of the oil, such as index of viscosity,detergents, dispersing agents, anti-oxidants, and anti-wear.

Among these characteristics, viscosity plays an essential role in determining the performance ofthe bearing and the hydrodynamic thrust bearing.

Therefore, to avoid problems during operation, it is good to know not only the chemical andphysicochemical behavior of the lubricant, but also the rheological properties that determine theflow.

Oil replacement is necessary because of the thermal and mechanical constraints to which the oilis subjected and because of the quantity of oil in circulation.

The replacement frequency depends on how frequently the entire oil charge circulates in thecircuit and whether there is an oil cooling system.

Generally, the appropriate replacement period is determined by regularly sampling and analyzingthe oil, mainly to check for pollutants or excessive oxidation.

Like all essential parts in a machine, bearings must be constantly monitored. For this, a minimalfollow-up procedure is established: monitoring of the temperature of the babbit at different



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locations, checking of the oil, etc. These measures can be completed with more sophisticatedvibrational analyses and with the follow-up of the temperature of the babbit on the journalsduring slow downs, etc.

2.5 Dynamic Behavior

To be complete, we also have to mention the effects of the supporting components on thedynamic behavior of a rotating machine.

For example, a pump unit is a rotating machine that is widely used and essential for powerplants. In fact, a nuclear power plan has several hundreds of these units with very differentcharacteristics. The purpose of the pump is to transfer the mechanical energy of the drivingrotating machine (for example, electric motor or turbine) to the fluid as pressure in the bestpossible conditions.

These units must operate in a satisfactory, reliable, sure, and stable manner.

For the end user, this means that the unit must present, during normal operating conditions, a lowand very stable level of vibration.

Moreover, in case of vibrational crises, it must be easy to diagnose and correct the unit. As ageneral rule, the exciter forces originate from the pump itself. So, it appears without doubt thatthe pump is the most sensitive part of the unit.

2.5.1 Description of a Pump

The objective for a pump unit is to transfer energy to the fluid to carry it from one point toanother. For this purpose, several technologies are available, including centrifugal pumps,rotating volumetric pumps, and alternating volumetric pumps.

A multi-stage centrifugal pump (see Figure 2-4) consists of a set of wheels contained within anenclosure called the shell. Energy is transmitted to the fluid by the wheel as a result ofcentrifugal forces.

A centrifugal pump can have a vertical or horizontal axis. It consists of a rotating element (rotorand wheels) and fixed elements (casing of the pump, sealing rings, and bearings).

The pump's operating principle itself will not be discussed here. Rather, we are more interestedin the supporting components, such as bearings, thrust bearings, and sealing joints.

To understand the dynamic behavior of a pump, we must consider in detail all the forces that acton the rotor.

Two categories of forces acting on the machine can be distinguished (see Figure 2-5). The firstcategory involves the forces of interaction that can affect the dynamic behavior of the rotor(modes, critical speeds, and response). These forces originate from the confinement zones. The



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second category involves the hydraulic or mechanical forces that directly apply a dynamic loadto the machine. They are the exciter forces (for example, imbalances and vibration of thestructure).

2.5.2 Interaction Forces

These forces occur in areas where the clearance is narrow: bearings, sealing rings, hubs, andbalancing pistons. They can also be found at the wheels' output. The study of the phenomena thatoccur in these areas can be complex. For small displacements, it is possible to assume a linearapproximation of the relation between the forces and the motion of the rotor around its positionof equilibrium.

The combined action of all these forces defines the intrinsic vibrational state of the system(natural frequency and damping).

2.5.3 Exciter Forces

This category includes the forces that excite the machine. They are always there, even though themachine does not vibrate. These include hydraulic and mechanical imbalances, inter-wheelinteraction forces, hydraulic forces caused by recirculation and turbulence, and the forcestransmitted by the supporting structure and the pipes.

Vibrations in a pump thus originate from either an unfavorable dynamic behavior of the system(presence of natural frequencies close to the speed of rotation), the existence of large exciterforces, or even a combination of the two phenomena.

During operation, these problems are diagnosed from the frequency analysis of the vibrationsignals collected on the machine.

This spectral analysis can characterize the vibrational state of the machine by associating certainspectral lines to known causes (see Figure 2-6 and Table 2-1). These data were taken on a multi-stage centrifugal pump with five wheels per level and rotating at 3560 rpm (signal detected on abearing).



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Table 2-1Classification of the Origin of Vibration on a Multi-Stage Pump [3]

Probable Causes Related toLine(see

Fig. 2-6)

Observations

The System The Exciter Forces

1 Line at the rotationfrequency Z

-Critical speed-Resonance of the bearing or ofthe structure

-Hydraulic imbalance-Mechanical imbalance-Bending of the shaft

2 Lines at the rotationharmonics (2Z, 3Z,etc)

-Non-linearities due togeometrical discontinuities-Friction, cracks in the shaft

Misalignment

3 Lines at the frequencyof the wheels (nZ,2nZ, etc)

Acoustical resonance -Insufficient radialclearance-Interaction between thewheels

4 Lines between 0.5 and0.95 Z,subsynchronousvibration

Instability of the rotor, bearing Periodic hydraulic excitation

5 Large vibration bandat a frequency < 0.2Z

Large band hydraulicexcitation (recirculation,turbulence)

6 Low frequency linebetween 15 and 1 Hz

-Improperly damped acousticalfrequencies of the fluid in thepipes-Instabilities caused by a pumpwith an improper hydraulicpressure curve

Hydraulic excitation:-Axial whipping due to thepositioning of the thrustbearing-Hydraulic forces excitingsome modes in thestructure.

7 Large high frequencyband between 0.5 and10 KHz or moredepending on thesucking pressure

Cavitation causing somepressure and vibrationpulses

8 Large band at a nonsynchronousfrequency

Resonance of the bearingor resonance of thestructure caused byturbulence.



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Vibrations are minimized by reducing the exciter forces, but this aspect will not be elaborated onhere. Instead, we are interested in the system behavior that depends on the various componentshaving low clearances, such as in bearings and sealing joints.

In fact, from a mechanical perspective, these components interact with the rotor throughhydraulic phenomena, and their importance depends on several parameters, including speed,clearance, load, and temperature. Thus, these components largely determine the dynamiccharacteristics of the system, such as natural frequencies, modal damping, and modaldeformations.

Figure 2-7 shows the natural frequencies of a pump plotted as a function of the rotation speed intwo cases: standard clearances and degraded clearances (2 times the normal clearance).

With degraded clearances, a critical speed exists that does not exist with standard clearances.

Nevertheless, this first analysis must be completed with the study of the system behaviorsubjected to a dynamic load. Imbalances are applied at different locations on the rotor and theresulting displacements are analyzed (see Figure 2-8).

These two approaches are totally complementary and help explain at least qualitatively thevibrational behavior of the machine.

Figures 2-7 and 2-8 show how imbalance on the bearings, sealing joints, and hubs affect thebehavior of the system through hydraulic interactions. We will describe them briefly.

Natural frequency is the frequency of the free vibrations (oscillations) in the intrinsic behaviorof a system.

Modal damping characterizes the exponential decay of the amplitude of the free oscillations.

Modal deformation is the shape of a mechanical system at its natural frequency. It ischaracterized, at a given instant, by the position of the masses that comprise it in a discreetsystem and by the position of the axes of the system in a continuous system.

2.6 Interaction Coefficients

2.6.1 General Definition

Unlike turbine-driven machines carrying gas or steam, the fluid in the pumps greatly affects thenatural frequencies, modal damping, and modal deformations of the rotor.

Thus, it is important to know as precisely as possible the coefficients of interaction of the wheels,sealing rings, balancing piston, and bearings. These coefficients, called dynamic coefficients,(see in Figure 2-9) represent the stiffness and damping of the fluid around a given static position.

For simplification, "fluid film" refers to the phenomena that occur at the sealing joints, hubs, etc.



2-16

Stiffness is the ratio between the variation of a force (or torque) and the corresponding variationin the displacement in translation (or rotation) of an elastic component.

Dynamic stiffness is the ratio between the variation of a force (or the moment of the torque) andthe linear displacement (or angular variation) during operating conditions. Dynamic stiffness candepend on the applied force (amplitude and frequency), level of constraint, temperature, andother conditions.

Viscous damping is the energy dissipation that occurs when an element or some part of anelement of a vibrating system is subjected to a force whose value is proportional to the speed ofthe element and whose direction is opposite to that of the speed.

2.6.2 Coefficient for Bearings

The main function of a bearing is to support and guide the rotor for designed operatingconditions.

In addition, bearings affect the dynamic behavior of the entire system (rotor + bearings + mounts+ shell, etc).

2.6.2.1 Roller Bearings

For roller bearings, the stiffness in the horizontal direction and that in the vertical direction aregenerally considered to be equal and constant, whatever the operating conditions of the rollerbearing may be. This stiffness can be evaluated experimentally or analytically. If more preciseresults are required, numerical tools are available on the market that will parameterize thestiffness of the roller bearing in a matrix form.

In all cases, roller bearings are responsible for a negligible amount of the damping. Thus, twocoefficients of stiffness are sufficient to represent the bearings in most instances.

2.6.2.2 Film-Lubricated Bearings

The coefficients of film-lubricated bearings can be derived analytically for simple cases [1] or byusing computer codes, such as EDYOS. EDYOS is a software package for the calculation of filmlubricated bearings and thrust bearings (code developed in a partnership between EdF/DER,Solid Mechanics Laboratory of the University of Poitiers, and Mechanics of Contact Laboratoryat the INSA in Lyon).

Regarding dynamic behavior, a stiffness matrix and a damping matrix can characterize thebehavior of the film around a static equilibrium position.

The main difficulty resides in the knowledge of the load affecting the bearing. As a general rule,only the weight of the rotor is considered, but experience shows that this assumption rarely givescorrect results.



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In fact, tensions produced by the sealing joints and balancing piston largely affect the load on thebearings. These tensions depend on the internal alignment of these different elements withrespect to the bearings.

Non-linear calculation tools that include the bending of the rotor in the bearings computer codecan determine the load applied to the bearings as a function of their real position. However, thesecalculations are complex and can quickly become fastidious.

Moreover, radial hydraulic forces originating from a non-uniform distribution of the pressure atthe output of the wheels also affect the load on the bearings.

2.6.2.3 Magnetic Bearings

Passive magnetic bearings have an acceptable load capacity, a low stiffness coefficient, and apractically null damping.

In an active magnetic bearing, feedback loop electronics define stiffness and damping.

This impedance is thus adjustable according to the needs of the user, creating the originality ofthis technology.

A stiffness and damping matrix like that for film-lubricated bearings shows this impeding factor.

2.6.3 Coefficients for Labyrinths and Sealing Rings

The primary role of the ring is to monitor the leakage rate from one level to another according tospecifications, pressure and speed conditions (see Figure 2-10).

Contrary to that of bearings, the study of the phenomena occurring in these areas is only aboutten years old.

In the United States, EPRI (Electrical Power Research Institute), a research center funded byAmerican electric utilities, launched a vast program of experimental characterization of smoothshort sealing joints in 1984. These results still constitute the best base of experimentalknowledge on this topic [4].

Several theoretical methods are proposed. The most widely known is that derived by Childs [5]on smooth joints. Nordmann [6] completed the method for grooved joints. Nevertheless, thesetheories do not yet give total satisfaction, particularly for grooved joints.

User-friendly computation tools have been developed to determine the characteristics of joints ofdifferent shapes.

In general, these studies show that the inertial effects of the fluid cannot be neglected as they canfor bearings because the fluid is subjected to large accelerations at the input of the joint. Themain consequence is the presence of added mass that affects the system significantly.



2-18

Contrary to bearings, a mass matrix is added to the stiffness and damping matrix to represent thedynamic behavior of rings. This evidently makes a reliable characterization of these componentsmore complex.

This is even more true for long joints whose coefficients are difficult to measure. In these cases,interaction forces become large, clearances are small, and rotor displacement becomes difficultto control. Also, the "tilting," or misalignment of the rotor with the joint, changes significantlythe distribution of pressure along the joint and thus changes the stiffness terms. Misalignmentmay simply be caused by the natural bending of the shaft.

Clearly, the most important effect that the circulating fluid has on modal location and shapeoccurs at labyrinths, hubs, and the balancing piston.

The sealing ring or joint is a device used to keep the leakage rate to a minimum between thehigh-pressure chamber located upstream of the joint and the low-pressure chamber locateddownstream.

2.6.4 Coefficients for Impellers

Until recently, the impellers were not particularly considered (see Figure 2-11). At best, the massof the fluid circulating in the impeller was considered in a system analysis. Black's hypotheses[7] show forces originating from impellers make rotor destabilization possible. Ohashi [8]published the first experimental measurements on a wheel, volute chamber, and shell.

Several subsequent studies showed that forces applied to hubs represent impeller interactionforces in addition to forces generated by the main flow of the impeller.

These different studies show that we are still in fluid films. Consequently, modeling of theseinteractions is done through mass, stiffness, and damping matrices.

From these experiments, it appears that stiffness terms are relatively low whereas mass terms andcrossed damping terms are high. They nearly cancel one another.

In addition, crossed stiffness terms are high, and damping terms rather low.

Thus, experiments show that the impeller interaction forces have a relatively weak effect on theglobal system and tend to reduce not only the natural frequencies, but also the correspondingdamping.

2.6.5 Coefficients for Other Components

Thrust bearings and eventual balancing disks are other areas where interaction forces are present.However, in these areas, radial forces are too low to affect the bending shaft behavior. Therefore,the effect of these elements is neglected.



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2.7 Main Failures

Failures affecting supporting components can lead to inopportune unit outages and to damagingproduction losses. The collected feedback and studies on this topic now provide identificationand classification of the main failures affecting supporting components on rotating machines.These studies constitute an interesting knowledge base that allows experts to quickly determinethe nature of failure and to describe probable causes.

This section summarizes the main failures affecting roller bearings and film-lubricated bearings,the only two types of bearings used in EdF power plants. For other components (for example,sealing joints and hubs), direct contacts between stator and rotor components should be fearedand avoided.

2.7.1 Roller Bearings

The lifetime of a roller bearing is essentially a function of its behavior under mechanical loadsleading to fatigue. Presumably, the roller bearing degrades due to fatigue rather than wear. Table2-2 summarizes the main failures affecting roller bearings [2].



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Table 2-2Causes of Roller Bearing Failures

Failure type Definition Probable causes

Galling Cracking or tearing of the surface ormaterial fragments

x Insufficient clearance

x Lack of lubricant

x Misalignment

Gripping Deformed rolling parts, heating,microfusions, and metal rolling


x Lack or excess of lubricant

x Excessive speed

Dent due to deformedrolling parts

Dents on balls and rollers, pushedmaterial

x Lack of care or shock duringmounting

Dent due to abrasiverolling parts

Material removal caused by wear,false Brinelling effect.

x Vibrations without rotation

Wear Wear of rolling parts, tracks, andcages

x Lack of care during mounting

x Pollution due to dust

Craters or fluting Pitting on and around the rollingtrack (balls) or narrow parallelfluting (rollers)

x Electrical current flow

Cracks or breaks Tool dents or ring ruptures x Lack of care or shock duringmounting

Contact corrosion Oxidation on outer diameter andsupport faces of roller bearing

x Loose adjustment

Coloration Coloration of the rolling tracks androlling parts


x Excessive speed

x Lubrication defect

x High temperature

Degradation of thecages

Cage deformation, wear and/orrupture


x Excessive speed

x Lubrication defect

x Loose adjustment

2.7.2 Film Lubricated Bearings

The main study on film lubricated bearings, initiated by EPRI [9]), shows that contamination ofoil, a failing supporting system, and large levels of vibration are the main causes for turbinebearing failures. These incidents can also occur on other rotating machines, such as pumps, fans,auxiliary turbines, and motors.

The same study identified and classified a dozen means of failure of film lubricated bearings(see Table 2-3).



2-21

Table 2-3Causes of Film Lubricated Bearing Failures

Failure type Definition Probable causes

Abrasion Grooves or serration due to action ofsolid particles, such as chips and sand

x Contamination of lubricant

Babbit let-go Partial or total lack of adherence to thejournal

x Badly prepared surface

x Presence of gas, grease, etc.

x Presence of fragile layers

Cavitationerosion

Presence of cavities or pitting on anarea of the journal (cavitation zone)

x Presence of diverging zones wherepressure is less than the cavitationpressure of the film

Corrosion Chemical etching of the base metal(journal or rotor) by reactive agents orelectrolytes

x Presence of reactive substances inthe lubricant

Pitting Local melting of babbit caused byelectrical arcs

x Electrically charged oil

x Potential difference between thebearing and rotor

Erosion Damages due to impact of high speedparticle-contaminated fluid

x Particles in the lubricant. Damage isconcentrated essentially aroundgeometrical discontinuities, such asgrooves.

Fatigue Cracks in metal x Loads greater than the metal's stresslimit and large number of cycles

Fretting Corrosion in a contact subjected tosmall vibrations

x Oxidation of the metal

x Vibrations in the bearing

Black scale Deep grooves on high chromiumcontent shaft

x Solid particles encrusted in thebearing

Excessiveheating

Damage due to local overheating and/ora high temperature gradient

x Presence of chips obstructing flow

x Thermal fatigue

Griping Hard friction on bearing and brutal stopof rotor

x Clearance too low between rotor andbearing

x Large differential thermal expansion

Wear Gradual decay of the babbit leading todegradation of the bearing'sperformance (excessive clearance orchange in the film shape)

x Particles

x Abrasion

x Excessive vibrations

x Frequent startups and shutdowns

Babbitdisplacement

Rotor carries and redepositssuperficially torn babbit downstream ofthe journal

x Friction between surfaces

x Tearing due to cracks

x Lack of lubricant

x Excessive loads



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2.7.3 Wear of Sealing Joints

Among the parameters defining a dynamic sealing joint, clearance plays an essential role becauseit significantly affects the leakage rate. Therefore, it is necessary to consider surface contact wearduring the machine design phase, particularly regarding the choice of materials.

In fact, clearance degradation leads to a higher leakage rate and significantly affects the dynamiccoefficients of interaction of the joint.

If metrology cannot be used to monitor the status of joints during operation, then a visual checkof the surfaces can yield precious information on the dynamic behavior of the machine (refer tothe article on sealing joints).

2.8 Conclusion

Bearings play a major role in rotating machines because they must support the rotor with a highlevel of reliability. Also, they greatly affect the dynamic behavior of the machine.

Plant operators cannot tolerate the failure of supporting components. Unplanned shutdowns arevery expensive.

Also, good design and monitoring of the operating environment are important, including thelubrication system and monitoring the entire machine for vibrations.

Reliable numerical tools (EDYOS) provide understanding of the parameters affecting thebehavior of bearings in most instances, such as static, dynamic, and thermal.

Monitoring of bearings is also possible, thanks to sophisticated methods set in place by experts.

As for the lubricant, monitoring methods exist to assure the good behavior of the oil charge,which greatly affects the performance of the bearings.

For the bearings themselves, a minimum follow-up during operation (temperature and/orpressure) helps to monitor eventual degradation of the mechanical or vibrational performance ofthe entire machine.

If despite all, failures occur on the bearings, more sophisticated and reliable methods of analysisare available for quickly analyzing the causes and effects of the incident. For the end user, thisprovides precious help for the good operation of his machine.

We have also shown the importance of zones of interaction in defining and understanding thedynamic behavior of rotating machines.

For the entire machine, high performance tools (CADYAC) of today can help provideunderstanding of the dynamic behavior by taking into account all elements affecting the rotor,such as bearings, mounts, shells, and fluid films.



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Let’s note that the interaction forces naturally depend on the shape of the components, speed,pressure, and temperature conditions as well as the nature of the lubricating (or cooling) fluid.

Nevertheless, all these tools remain perfectible because uncertainties still subsist on the bestdesign of all components and environmental conditions (for example, load, temperature, andlubricant quality). In addition, the validity of the hypotheses used to build the mathematical ormonitoring models currently in use still has to be verified for difficult and extreme cases.

All these aspects are developed throughout articles presented during this seminar.



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2.9 References

1. J. Frene, D. Nicolas, B. DegUeurce, D. Berthe, M. Godet. Lubrification hydrodynamique—Paliers et Butées. Collection n° 72 de la Direction des Etudes et Recherches. EditionsEyrolles 1990

2. M. Moret. Roulements et butées à billes et à rouleaux. Les Techniques de I'Ingénieur B 5370a-1.

3. U. Bolleter, A. Frei et al. Rotor Dynamic Modeling and Testing of Boiler Feedpumps. EPRITR-100980, September 1992

4. S. Florjancic, R. Stuerchler A, T. McCloskey. "Annular Seals of High Energy CentrifugalPumps." Presentation of Full Scale Measurement, 6th Workshop in Rotordynamic InstabilityProblems. Texas A&M University, College Station, Texas, May 1990.

5. D. W. Childs. Finite Length Solutions for Rotordynamic Coefficients of Turbulent AnnularSeals. ASME-Paper 82-Lub-42.

6. Nordmann et al. Rotordynamic Coefficients and Leakage Flow for Smooth and GroovedSeals inTurbopumps. Proceedings, IFToMM Meeting -Tokyo 1986.

7. H. F. Black. "Lateral Stability and Vibrations of High Speed Centrifugal Pump Rotors."Proceedings IUTAM Symposium Dynamics of Rotors, Lungby 1974.

8. H. Ohashi, H. Shoji. "Lateral Fluid Forces Acting on a Whirling Centrifugal Impeller inVaneless and Vaned Diffuser" 3rd Workshop in Rotordynamic Instability Problems in HighPerformance Turbomachinery. Texas A&M University, College Station, Texas, May 1984.

9. O. Pinkus. Manual of Bearing Failures and Repair in Power Plant Rotating Equipment,EPRI GS-7352, July 1991



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Figure 2-1Examples of Roller Bearings

Figure 2-2Examples of Film Lubricated Bearings



2-26

Figure 2-3Operating Principles of Magnetic Bearings

Figure 2-4Cross-Sectional View of a Multi-Stage Centrifugal Pump (Picture Allis – Chalmers)



2-27

Figure 2-5Forces Acting on the Rotor of a Multi-Stage Pump [3]

Figure 2-6Spectral Analysis of the Vibration in a Pump (Wheel With 5 Buckets) Rotating at3560 rpm [3]



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Figure 2-7Campbell Diagram in Standard Operation and Degraded Operation [3]



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Figure 2-8Example of Computation on the Response of an Imbalance in a Pump [3]



2-30

Figure 2-9Schematic View of the Dynamic Coefficients of a Film Lubricated Bearing

Figure 2-10Schematic View and Location of the Sealing Rings on a Pump



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Figure 2-11Schematic View of a Wheel and Swirl Break of the Balancing Piston on a Pump


3-1

3 EXPERIENCE FEEDBACK ON THE MAINTENANCE OFNUCLEAR POWER PLANT PUMP ROLLER BEARINGS


No. 97NB00113

Authors:J.P. Cailleaux

L. Randrianarivo




Experience Feedback on the Maintenance of Nuclear Power Plant Pump Roller Bearings

3-2

Executive Summary

Roller bearing maintenance in nuclear power plants is not seen as critical. The maintenanceprograms only call for monitoring techniques once an evaluation of the consequences in relationto the failure risks of the bearings has been carried out and the efficacy of such methods has beenvalidated on a case-by-case basis.

However, in terms of design, we are dissatisfied with "under load" operating problemscharacterized by overheating of the bearings which can lead to bearing destruction. Theseproblems have, on occasion, been underestimated by the designers. They also appear later whenbearings with "improved" load capacities appear on the market and, paradoxically, no longermeet the operational requirements of the machines.

Greasing problems must also be mentioned. These can arise because of bearing design problems,which are difficult to modify on an installed base of machines as EdF's. Another problem is thedifficulty of defining as satisfactory optimum in terms of quality, frequency and quantity ofgreasing on machines operated in very particular conditions.

Finally, the drive to optimize maintenance and the desire to reduce risks attributed to the mixingof different greases entail the standardization of certain practices, notably the choice oflubricants. In this respect, the disparate nature of the nuclear base represents a difficult hurdle tosurmount.



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3.1 A Good Feedback Experience

From 225 rpm to 23,500 rpm, tens of thousands of roller bearings operate each day in the nuclearpower plants without causing any particular problem. From this perspective, the feedbackexperience can be considered rather good.

Well-designed, oil-lubricated roller bearings have an excellent behavior. For grease-lubricatedroller bearings of the same design quality, this remark must be weighted. In fact, the choice ofthe method of lubrication causes problems for the maintenance engineers seeking a compromisedifficult to find.

The result of our investigation on large grease lubricated machines shows that there are alwaysfailures.

High Levels of Vibration

High Temperature

Under Load Operation

Grease Evacuation Problem

Mixing of Greases

Hardened Grease

Lack of Grease

Excess of Grease

“False Brinell Effect”

Water in Rolling Bearing

Other

Destruction

Rolling Bearing Failure

Abnormal Noise

0 5 10 15 20 25

Other Motors Auxiliary Feedwater System Containment Spray System

CRDM Power Supply System Chemical and Volume Control System

Safety Injection System Shutdown Reactor Cooling System

Component Cooling System Essential Service Water System

Figure 3-1PWR 900/1300MW Nuclear BaseNumber and Distribution of Roller Bearing Failures Observed During 82/83



3-4

The most frequent problems shown in the figures affect some very specific machines (forexample, shutdown reactor cooling pumps, essential service water pumps, and chemical andvolume control pumps). Because destroyed roller bearings are difficult to investigate, thedegradation modes attributed to the most spectacular incidents are unfortunately difficult tocharacterize.

In the list of problems encountered, greasing problems top the list: lack of grease, greaseevacuation problem, hardened grease, mixing of greases, high temperature. Less frequently,problems such as design problems, under-load operating conditions, and even more rarelydegradation by "false Brinell effect" affect very specific systems.

3.2 Maintenance Performed on Roller Bearings

For a few years, EDF has been developing a method of analysis (OMF) that can determinewhether the machines should or should not be checked for preventive measures. This method isbased on the evaluation of three essential parameters associated with the risk of machine failure:

x Power production unit safety

x Production tool availability

x Maintenance costs

In practice, choices are extremely limited in terms of roller bearing maintenance:

x The greasing/lubrication function is never overlooked; it is a functional requirement of theroller bearings.

x The systematic preventive replacement of the roller bearings is not considered; the reliabilityis good, and the life expectancy of roller bearings is long.

Thus, the problem boils down to choosing whether to monitor the bearing during operation. Thismonitoring can be preventive, even predictive.

Greasing. Except for roller bearings greased for life that are found sometimes on smallmachines, lubrication always requires a minimum of preventive maintenance. The lubricantmust be replaced periodically.

Roller bearings are not systematically replaced as a preventive measure. Roller bearings are oftendesigned for a long operating lifetime (>150,000 hours) and will last for the expected lifetime ofthe machines. On the other hand, operators do not hesitate to replace bearings when they havebeen dismounted during a maintenance check.

The reasons are as follows:

x TechnicalIt is difficult and probably expensive to correctly assess the status of a roller bearing(particularly the tracks and rolling elements)

The dismounting of the roller bearing can generate degradations



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x ReliabilityNew roller bearings can be considered reliable with a rather high level of confidence.

x EconomicReplacement cost of a standard roller bearing (not custom made) is acceptable.

In conclusion, neither the reliability concern nor the cost concern justifies operating the rollerbearing to the extreme end of its life expectancy.

3.2.1 Monitoring During Operation

In this area, maintenance has been the area of a large number of improvements.

3.2.1.1 Monitoring Bearing Temperatures

First, let's recall that critical machines have some monitoring devices mounted on their bearings.This monitoring is very effective and well adapted to show functional problems, such as:

x Dysfunctional operation of the bearings (under load operation)

x Incorrect sliding of the guiding bearing

x Fluctuations associated with unsuitable or insufficient lubrication

Accessing the temperature of the roller bearing (outer cage) presents an important advantage.The main functional limitations of the roller bearing are tightly linked to its operatingtemperature (for example, behavior of the grease or polyamide cages). Therefore, it is possible todefine shutdown criteria that will prevent unavoidable failure in all circumstances.

Second, one must consider that all the degradation modes of a roller bearing appear sooner orlater as a rise in temperature of the bearing. Temperature monitoring with varied levels of actionregarding maintenance is all that needs to be done.

3.2.1.2 Monitoring Bearings: Towards Predictive Maintenance

Roller bearing deterioration originates mainly from galling, gripping, and corrosion. Whatever itsorigin, a degradation of the tracks or contact surfaces of the rollers (or balls) finally occurs.

3.2.2 The "Metravib" Defect Factor, a Method Used in Power Plants

3.2.2.1 Principle

The technique used is based on measuring the peak factor of the vibrational signal equal to theratio of peak to peak value (J c.c) to the effective value (J eff). The measurement of theacceleration (J) is given more attention to selectively characterize the high frequencies attributedto galling phenomena. Low frequency noises are filtered out with a low pass filter.



3-6

From the original signal, a defect factor F is calculated as follows:

effeff

c.cF EJEJ��JJ

JJDD

3.2.2.2 Monitoring Device

Following a series of tests, a portable device was selected to collect the data: the Metravib fromFramatome Diagnostic.

Simultaneously, a data evaluation software was supplied to the power plants: the Galileesoftware.

Since 1990, these monitoring devices have provided feedback experience leading to severalimproved versions.

3.2.3 Implementation of the Defect Factor

The implementation of this monitoring encountered several difficulties: the estimation of theappropriate defect factor criteria according to the evolution of the data collection equipment(MV03 or MOVILOG of Metravib) and the use of a sensor with glued seating. The developmentof the defect factor had to be adapted in such a way that the results would be identical and thatthere would be no ill-timed warnings.

These variances were corrected in two ways:

x The hardware was modified to corroborate the two modes of development of the F factor,according to the device's origin.

x A new defect factor was defined for the new detectors (MOVILOG). This number, called"test factor" FDT, is calculated for frequencies ranging from 1 to 4 kHz (instead of f > 3kHz).

Despite these efforts in development and adaptation, monitoring the roller bearings involves arange of uncertainty requiring a case-by-case validation. Nine percent of the machines checkedgive some sign of unjustified warnings, whereas other machines may have problems that thedetector does not catch.

For memory, the following criteria are used:

F < 5, normal status5 < F < 7.9, slight degradationF > 8, advanced level of degradation



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3.2.4 Mixed Conclusions Calling for Caution

After deciding to implement a predictive maintenance process, the development andstandardization of a detector and a software application created difficulties. These problems arestill being evaluated. The concept is not in doubt, and we are fortunate to have adopted astandard system in a field where exploiting the feedback experience is fundamental. On the otherhand, it is the application of this tool that is being challenged. The problem must be formulatedin terms of costs and benefits.

It is obvious that the development of these monitoring techniques requires resources that are notnegligible. Four hours are spent for collection and analysis, and this represents only the tip of theiceberg. Also to be considered are the time spent on designing and improving this entire systemand on building the national databases for feedback information. What is the return on thisinvestment?

In practice, a choice must be made between two alternatives:

x A basic means of detection with its share of uncertainty and risks as to the possibility ofproducing errors, such as ill-timed warnings

In these conditions, in case the F criterion goes out of its normal range, more sophisticatedmeans are implemented to characterize the observed conditions. In its actual form, thedetector and its associated criteria meet this type of objective rather well. However, the plantoperator needs to know for certain if a machine is available and will tend to distrustmonitoring that leaves doubt.

x A more refined and reliable diagnostic tool, suitable for the characterization of defects onvery specific rotating machines

In that case, the visual check of roller bearings is a monitoring technique that accompaniesthe analysis of other parameters characterizing the operating status of the machine. Thisapproach based mainly on the use of feedback experience was studied as part of the MACH2program (monitoring of small machines).

3.2.5 Under-Load Operation: A Design Problem Difficult to Manage on a NuclearBase

Under-load operating conditions can affect the behavior of the roller bearing. This condition cancause sliding of the rolling elements. In a lightly loaded roller bearing, the rollers do not rollcorrectly on the tracks, and sooner or later the tearing of material causes the destruction of theroller bearing.

In regard to this, two factors have often penalized the designer’s choice:

x Architecture of vertical axis machines, where the rotor weight does not contribute to theradial loading of the roller bearings.

x The expected lifetime needs to be long, which leads to over-dimensioned roller bearings.



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Since 1985, a new problem arose in addition to these well-defined situations: the developmentand commercialization of "extreme load" roller bearings as a substitute for previous products.For an equal size, these roller bearings support much heavier loads and behave much worseunder light loading conditions. This "technological evolution" initiated by the roller bearingmanufacturers has had insidious effects on machines that encountered under-load condition aftertheir roller bearings were changed.

As a remedy, large-scale dispositions had to be taken in accord with the manufacturers to modifythe bearing every time that a new generation of roller bearings was used. The modificationconsists of selecting from the "extreme loading" series—the only one commercially available atpresent—roller bearings suitable for light-loading conditions. Because these roller bearings aresmaller, it is necessary to make rings in which the roller bearing can be mounted and to modifythe grease circulation channel.

What may look simple for the engineering division to do or relatively benign when only a fewmachines need to be modified, becomes proportionately huge when a machine base the size ofEDF's is considered. In short, it was necessary to do the following:

x Justify, as a first step, situations that were really becoming preoccupying.

x Review all design files for large machines.

3.2.6 Greasing: Hard-to-Find Compromises

Classified unduly to elementary maintenance tasks, greasing is at the heart of in-depth studiesinvolving many people. The problems are numerous and complex.

The complexity is largely due to the diversity in practices among the different plants.Maintenance is optimized at different levels through the following:

x Nationally managing the maintenance of "important machines," through a standard exchangepolicy

x Seeking to reduce and standardize the products used

x Using feedback experience to select greases best suited for the required service andprescribing optimal greasing modalities (quantity and periodicity)

Historically, greasing plans have not been defined on a national level. Specifications ofmanufacturers or regional representatives of large suppliers in oils and greases have variedgreatly, evolving over the past 20 years. Each plant has defined with its suppliers the range ofproducts that seemed best.

When standardization efforts occur, the transition phase is even more difficult because manygreases are not compatible with one another. The behavior of grease mixtures in operation isrelatively unknown. Therefore, changing the grease type cannot occur without dismounting theroller bearing to remove the old grease. This process is a relatively heavy and penalizingintervention.



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Let’s recount problems having the same symptom—a rise in temperature in the bearing—although this one sign rarely determines the precise situation. Systematically, the followingprobable causes are mentioned:

3.2.6.1 Unsuited Grease Quality/Quantity/Periodicity

With the roller bearing manufacturer's recommendations, the analyses of the machinemanufacturer, and the choice that is made by the plant operator, an inflationary approach hasalways prevailed. The result is that the choice of grease ends up being high pressure andunsuitable for the service demanded.

3.2.6.2 Choice of Grease Volume to Inject Is Subject to the Same Type ofDifficulties

What may look simple in traditional applications becomes more delicate when safetymechanisms are concerned, such as those in place in nuclear power plants. These systems,designed for installation safety, have a never-operating quality except for a few hours duringannual periodic checks. In case of emergency, however, these systems can be required to operatecontinuously for up to 8,000 hours, without any grease make-up. Under these conditions, themanufacturers' operating guides that traditionally give the quantity of grease to be injected as afunction of the number of operating hours are totally useless.

3.2.6.3 Greasing Methodology

The grease is injected under pressure while the machine is operating. This operation leadsinevitably to a rise in the temperature of the bearing for the time necessary to replace the grease.The implementation of this method quite often poses problems on the emergency pump unitswhen the machine operating time during the periodic check is insufficient for the bearing'stemperature peak to drop.

3.2.6.4 Dysfunctional Operation of Grease Valves

During greasing of a motor, a quantity of grease equivalent to that injected must be evacuatedthrough the grease valve and recuperated in the tanks designed for that purpose.

This scenario is not always observed on the sites: a good number of tanks are either found emptyor full of oil.

These observations show a dysfunctional operation of the grease valves that may be caused byseveral factors:

x Design of the grease evacuation device (disks/baffles), more or less successful

x Using grease with unsuitable viscosity



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x Finally, a drying and filling phenomenon that can occur on machines not operated for a longtime

3.3 Conclusion

Roller bearing maintenance remains an area of concern for maintenance engineering divisions.True, the operating behavior is rather good globally, but a few problems still remain. Designchoices must be corrected. A clear, optimum greasing schedule is yet to be found. Finally,monitoring techniques must be adopted after their limits and the cost and benefit of theirimplementation are evaluated seriously.


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4 SPARE PARTS FOR THE BEARINGS AND BALLBEARINGS OF THE ROTATING MACHINE SWIVELINGPROVISIONS


No. 97NB00128

Authors:J.M. Vialettes

J.P. Deroo




Spare Parts for the Bearings and Ball Bearings of the Rotating Machine Swiveling Provisions

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Executive Summary

The stock of spare parts must meet the requirements of scheduled inspections and be capable ofdealing with the vagaries of operation.

Two existing computer programs have been developed to limit and reduce the financial volumeof the stock and to foster a "team" spirit among the different parties involved. The sitesthemselves undertake the supply of bearings except in the case of the R.R.A. (shutdown reactorcooling) pumps, which are handled by the U.T.O. (Unité Technique Opérationnelle), thecorporate maintenance department. The latter is responsible for the T.G.U. (Turbine GeneratorUnit) journal bearings and for journal bearing for category 1 equipment for the R.R.I.(Component Cooling System), A.S.G. (Auxiliary Feedwater System) and A.P.P. (Turbine DrivenFeedwater Pump System). The T.G.U. journal bearings fall into two families: elliptical withthree pads or segments, and elliptical in two half bearing shells. Only the second type can berepaired.



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4.1 Background and Goals

The Maintenance Department of the EPN (Nuclear Generation Group) defines the policiesregarding the maintenance of machines through the PBMP (Basic Preventive MaintenanceProgram). These policies affect the type of work or checking required and the frequency, thusdirectly affecting the inventory of spare parts.

Therefore, it is necessary to adjust the spare parts inventory to respond to the needs that ariseduring scheduled checks and needs due to inopportune events that occur during operation.

The goal is that, with a good level of maintenance and an operation within operating range,random breakdowns will move to zero and the inventory of spare parts will move to its optimallevel. We are not there yet.

Feedback experience affects the PBMP and, consequently, the inventory of spare parts.

These different cause-and-effect factors have led to a concept of consumable parts and safetyequipment (parts that have a lifetime similar to that of the plant, but must be kept in inventory forfear of long, unplanned shutdowns) to which the EPN, in consideration of the shutdown time, hasadded the concept of off-the-shelf parts, either by the part or sub assembly.

These different concepts and their interpretation are defined in documents published by the EPNand called the IN30 (National Guidelines nq30). The documents also define the roles of differentplayers, such as the CNPE's (Nuclear Power Production Center) and UTO's (OperationalTechnical Units). See the first page of the IN30 in annex 1.

This global approach to spare parts has naturally led to another project called POS (InventoryOptimization Plan), whose philosophy could be:

Before requesting spare parts, the nuclear production group must make sure that the parts arenot already available in inventory and in sufficient quantity.

This project has two objectives:

x To limit and reduce the financial value of the inventory

x To develop a "team" spirit between the different players

Achieving these objectives requires widespread and effective communication tools.

Two existing computer applications, FTPDR (Spare Part Data File) and A39, have been adaptedfor that purpose.



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4.2 Tools

To trade conveniently, the different parties must always keep two criteria in mind: inventoryidentification and management.

4.2.1 Identification

A unique label attached to each part contains the nationwide part number that can bereconstituted to provide different pieces of information characterizing the part. It is essentially adrawing number and a guiding mark in the global assembly drawing, a number referring to adrawing of the part with or without a label, a reference to a catalogue, and the functional EdFlabel on which the part is assembled.

Each number is like a keyword for searching the spare part database.

The UTO, along with the computer application FTPDR (Data File on Spare Parts), gives thealphanumeric, seven-character national number for the parts under its responsibility. Each plantcan assign part numbers for their local use (FPR software, Spare Part File, from SYGMA).However, the UTO imposes a nationwide label for each part under its authority with the FTPDRsoftware application. The production site can then establish a voluntary link between its localnumber and the nationwide number.

The first three characters on the label refer to the machine and, in most cases, provide thedestination of the part (for example, 414 control rods and mechanisms, 511 main turbine, 611main alternator, and 712 screening drum).

4.2.2 Management

Once identified, the part must be located. Each CNPE manages its inventory warehouse with thelocal A39 software application.

The UTO also has its local A39 to manage its three national warehouses: Creil in the Oiseregion, Dieppedalle near Rouen, and Bugey next to the production site.

The UTO also manages three third-party warehouses (EdF inventory stocked in a supplier'swarehouse, GECALSTHOM, CEFILAC, FRATIA).

These local A39s work with an eight-character part number. A code is added before the FTPDRidentifier for accounting and management purposes. The most often used codes for spare partsare the following:

5 consumable part6 safety equipment9 part under reparation8 downgraded part



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All these local A39s have been consolidated to communicate with each other. The AGNES(National Management and Inventory Control software application) software package forecasts,from the seven-character identifier, the need for parts, which the local A39s do not do.

For each part, AGNES gives:

x National level of inventory and its location

x UTO orders with the shipment time for supplying the part and repairs

x Future site demand and the delivery delay

Consolidation of these three parameters gives off a warning that leads to restocking. UTOwarnings occur at two levels: the minimum UTO inventory level and the minimum nationalinventory level.

For more substance, figures for 1994 follow:

x Number of reference parts managed by UTO 19,530

x Value of the UTO stock 2,500,000,000 FF, or 2,500,000 KFF

x Number of parts delivered to sites 40,000

x Number of deliveries made 6,000

4.3 Roller Bearing and Bearings

Let's concentrate now on the theme of the seminar. From a spare part perspective, a rollerbearing or a journal bearing is a part among others. Here are more details on these two types ofparts.

4.3.1 Roller Bearings

Except for three types of roller bearings mounted on RRA (shutdown reactor cooling) pumps, allroller bearings are directly supplied by the sites. The part numbers are local. They can be labeledwith the AM code of the machine, but they are more generally labeled as standard components(150 + 5 numbers).

SKF no longer keeps the three RRA roller bearings in inventory because the SNCF (NationalRailway Company) has abandoned these models and the demand for these bearings has dropped.

When the inventory drops to a critical level, a special work order with a minimum quantity of 50units is placed.

It is logical that UTO provides management for this type of situation.



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4.3.2 Bearings

I will not review the terminology developed by the Maintenance Department.

Whether a spare part is kept by the UTO or the CNPEs depends on the category, value, andspecificity of the part.

The UTO is responsible for the turbine generator journal bearings as well as the journal bearingsclassified as category 1 mounted on the RRI, ASG, and APP machines. These "small" journalbearings are not repaired.

Let's have a closer look at the "large" journal bearings.

There are two families for the journal bearings mounted on turbine generators:

x Elliptical journal bearings in two half bearing shells, repairable

x Journal bearings with three pads or segments, non-repairable

4.3.2.1 Elliptical Journal Bearings

Elliptical journal bearings are mounted on the rotor of the 900 MW CP0/CP1 units. On the otherCP2 and 1300 MW units, some can be found on the front bearing with the turning gear and theattached greasing pump.

For each journal bearing with a standard diameter, there exists a journal bearing with amachining allowance to custom fit the diameter of a shaft bearing.

Moreover, for known non-standard shaft bearing diameters, a replacement journal bearing is alsokept in a national warehouse.

These journal bearings can be repaired. They are delivered to the sites as a trade-in. UTO takescare of repairing at the expense of the site (30% to 50% of the value of the journal bearing).

The following numbers refer to elliptical journal bearings (excluding small journal bearings onthe front bearing of the 900 CP2 and 1300 MW units):

x Number of journal bearings in operation 240

x Number of different parts 15

x Number of trade-ins in 1993 5


4.3.2.2 Three-Pad Journal Bearings

Three-pad journal bearings are used to support the rotors on the 900 MW CP2 and 1300 MWunits. The table in annex 6 shows the type of bearing as a function of diameter (RETRO).



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Similar to elliptical journal bearings, a diameter with a machining allowance corresponds to eachstandard diameter. No specific diameter is listed. The shaft bearings on the generator rotors arenot rebuilt to the minimum, but to repair specification, which explains the absence of an 800 mmdiameter journal bearing for the generator with a machining allowance.

The CHINON B3 and B4 generators have diameters of 710 mm. Spare parts are managed by thesite. That is the only exception.

The following numbers refer to three-pad journal bearings:

x Number of pads in operation 196

x Number of different parts 14



4.4 Policies

The journal bearings are supplied and rebuilt by the original supplier of the turbine generators,Gecalsthom, in order to conform to the original specifications and calculations and to benefitfrom a feedback experience external to EdF.

The inventory quantities of spare parts correspond to about 10% of the total number of partsinstalled on the machine base. A greater or reduced inventory results from feedback experience,particularly for the three-pad journal bearings having a more sensitive supporting pad.

Originally, pads had to be replaced in sets of three. Now, they are replaced individually.

Supporting pads are delivered in sets of three because they are made from a forged, babbit-coated ring. The upper and side pads come from a ring that can be made into four pads.

The pad coating is not reparable. Often, it is possible to save the upper and side pads when thecoating on the supporting pad has been torn away and redeposited onto the upper and side pads.

The main reason for the replacement of the journal bearings, whether they are elliptical or withthree pads, is the tearing away of the babbit coating beyond an acceptable range.

As an indication of cost, the cost of a set of three 425 mm diameter pads is around FF 230,000.

The UTO international telephone number to call after business hours for spare parts is33+1+46.99.27.10.


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5 METHODS OF ANALYSIS AND FOLLOW-UP OF THEOILS IN TERMS OF THE MAINTENANCE OF ROTATINGMACHINES


No. 97NB00117

Authors:G. Court

C. Chan Hew Wai




Methods of Analysis and Follow-Up of the Oils in Terms of the Maintenance of Rotating Machines

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Executive Summary

This document presents the follow-up procedure for an oil charge through the taking of samples.This follow-up is explained in the following five stages with the role of the operator described ineach one:

x Taking the sample

x The accompanying record sheet

x Physical and chemical tests on the sample

x Log

x Diagnosis

The actual taking of the sample is explained in particular detail because this is vital to a correctdiagnosis.

The various physical and chemical tests are described not by principle, but more in terms of whatcan be expected of them. Examples are given.

The evaluation process is also discussed, and the links between parameters are described.

The follow-up sampling of the oil charges also enables a better management of its replacement.Therefore, this operation makes it possible to limit the production of used oil and, thereby, waste.

The importance of the controls of the relationship between the operator and the log is described.

The follow-up of the oils through physical and chemical testing is an important tool in themaintenance of rotating machines because it can provide a diagnosis of the state of the machinesthemselves.



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5.1 Why Follow up on Oil Charges

Following up on the characteristics of the oil charge in rotating machines involve:

x Monitoring the aging of the oil charge

x Monitoring the wear of the mechanical components

x Limiting waste production, that is, used oil

x Establishing a diagnosis

x Getting feedback experience on similar materials or new products

Together, these points can limit expensive shutdowns due to lubrication problems. Also, they cantranslate into preventive action during scheduled shutdowns.

5.2 How to Follow up on an Oil Charge

The follow-up of an oil charge involves a set of sensitive operations managed by differentoperators. These steps are listed below.

Step 1

The first and foremost step is the sampling of the fluid and the degradation products it carries.

Technicians usually perform that operation. It is important to inform and train these people sothey can participate throughout the process. An improperly taken sample can limit the diagnosisor lead to bad decision-making.

The sampling process constitutes at least half of the testing process.

We will return to this step later to define the most important points of this operation.

Step 2

The second step involves the collection of most information on the history of the oil charge sincethe last check. What do we ask for?

x Date and location where the sample was taken

x Type of machine

x Type of oil

x Who took the sample?

x How many hours has the oil charge worked?

x Has any makeup been performed? If so, how much volume has been added or replaced?

x Has any event affected the oil charge?



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Step 3

The third step includes the physicochemical testing of the sample. What do we look for in thissample?

x Level of stability or degradation of the fluid; that is, its lubricating quality as well as thesecondary properties of the fluid due to additives

x Origin of eventual pollution of the fluid after the following:

– Human errors during makeup (errors regarding the quality of the fluid)

– Introduction of foreign products during operation or check-ups

These products can be liquid and act on the basic characteristics of the oil. When theyare solid, they can act, depending upon their hardness and size, as a microscopicmachine tool that accelerates the wear on mechanical parts. Small solid particlesstimulate deaeration degradation by accelerating oxidation of the oil.

x Type and origin of the products of degradation

– Organic products can be found when the oil base polymerizes after exposure tooxidation, radiation, mixing, and cracking.

– Organic compounds can come from coating materials or sealing joints.

– Additives can precipitate during abnormal storing.

– Metallic chips or metallic oxides can appear in small quantities and may be a sign ofcostly consequences.

x Any nonquantifiable variation in oil characteristics

– Change in odor

– Change of color

– Abnormal wettability of instruments

– Any other abnormal characteristic

These checks can assist in following the evolution of the oil characteristics, but must also revealnon-standard incidents.

Staff responsible for these checks must look at the sample critically and detect anomalies eventhough the machine may seem to work properly.

Step 4

The fourth step involves the historical analysis of the collected data as well as a cross-analyticalcomparison with the data taken from other identical machines. This analysis must take intoaccount eventual makeups or oil changes.



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Step 5

The fifth and last step involves the diagnosis of the fluid from which the sample was taken. Areport comments on each main point regarding the sample:

x Status of the fluid and possible remedy

x Status of the material and determination of the origin of observed degradation products

x Global aspect of the oil-material couple

Technicians with a good knowledge of the machines and constitutive materials shouldperform this step. They should propose their solutions, but they must also be aware of thecorresponding financial factors.

The important point to remember beyond these follow-ups is that machine degradation can beprevented. It is necessary to detect as quickly as possible when degradation starts in order toreact accordingly or correct the phenomenon until the next scheduled shutdown.

The end purpose is not to realize that the machine is degraded, but to realize that the currentmode of operation is going to degrade the machine and to propose possible solutions.

5.3 Sampling

We noted earlier the importance of this operation.

To preserve the integrity of the sample, the best container for transportation is a glass flask, butbreakage, which postal services do not appreciate, often occurs.

Therefore, this type of container is only used for tests requiring counting of particles and veryfine spectroscopy.

The container most often used is a tin can that matches the fluid volume variation caused bytemperature changes. Tin cans can also withstand heavy impacts.

This type of container can slightly affect the sample, but experience shows that in most cases thefinal diagnosis is not affected.

Plastic containers must be avoided because the sample can acquire its infrared signature and alsoretain suspended particles.

In all cases, containers are prepared in a laboratory. They are cleaned with filtered (0.45Pm)gasoline ether, dried, and tightly closed.

Containers delivered to the sites are kept in a clean storage room where their integrity ispreserved until used.



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The most commonly used volumes are:

x For follow up on the set of parameters for the oil charges of turbine generator groups: twotimes two liters in a glass flask closed with a Teflon seal

x For a practically complete follow-up on the oil charges of turbine-driven pumps: one timetwo liters in a glass flask closed with a Teflon seal

x For particle counting, 125 ml glass flask previously cleaned with a detergent and then driedusing filtered (0.45Pm) isopropyl alcohol

The sample must be the most feasible example of the oil charge and the degradation by-productsthat it carries.

As much as possible, the sampling operation must be done on a circulating pipe. In other cases,the sampling is done within the oil storage compartment and, if possible, at different levels.

Unless the sample comes from circulating oil, the testing laboratory should be informed of themethod used to collect the sample so the laboratory can refine its diagnosis.

However, in all cases, samples taken on a machine should be collected using the same method atthe same location. This will help with the historical follow-up.

5.3.1 How to Perform Samplings

With the previous information in mind, operators must consider the following points:

x Do not take the sample in an area where thermal insulation is being handled or in very dustyareas. In the same manner, the samplings should be performed without causing any of thedust deposited on the machines to become airborne.

x Follow these preparation and collection steps:

1. Remove from the storage room the right number of containers for taking the sample andmaintain them at an optimum level of cleanliness.

2. Prepare a bucket to collect the flushing oil. If several samples are to be taken, use a largercontainer.

3. Use lint-free, clean cloths for cleaning the sampling valve.

4. Use labels to identify the sample containers.

5. Wipe the end of the sampling valve with a cloth to remove the maximum dust and oxides.

6. Drain a small volume of oil equal to at least three times the volume of the sampling fillerneck at a quick rate which, if at all possible, is close to the circulating speed inside themain pipe.



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7. Without changing the draining speed, use a small volume of oil to flush the sampling can,

8. Empty the can quickly, and reuse the same container to take the sample, maintaining anair cushion under the neck. Close the container and the sampling circuit.

9. Wipe the can and label it. Maintain it at an optimum level of cleanliness.

10. Quickly send the container to the laboratory along with all the relevant information.

5.4 What Can Be Expected From the Different Control Parameters

Approximately twenty different parameters can affect the diagnosis of an oil sample. Thefollowing section explains what can be expected from each one and its importance.

Aspect

This very simple parameter allows the plant operator to quickly recognize an anomaly thataffects the oil. This check is performed by looking at the transmitted light going through 10 cmof oil held in a clear container.

When conducted just after the sample, this check allows quick detection of an abnormaldeaeration or demulsification. This observation is subject to judgment; nevertheless, it provides avalid warning sign.

After the sample has rested, this check allows observation of abnormal particles that havedeposited on the bottom of the container, water droplets, or a persistent cloudiness. This check isimportant to allow detection of coolant leaks.

The aspect check has the advantage of providing a quick diagnosis and, therefore, increasing thereliability of the machines.

At the laboratory, this parameter often helps determine the quality of the sampling. Also, it canbe used to direct maintenance personnel to look for water, deaeration, and demulsification.

Checking sample odor completes the observation. The odor should be soft, not harsh. A samplethat smells burned or thermally degraded can indicate a hot point in the circuit (for example,cracks in a heating element or degradation of mechanical parts under heavy load). Any unusualodor must, if possible, be identified to determine the origin of an eventual pollution (for example,introduction of cleaning solvents, fuel, regulating fluid, or development of bacteria ormushrooms characterized by a smell of fermentation or rot).

Machine operators can check for unusual odors during their shift. Degradation caused by hotpoints tends to produce odors around the oil tank.



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Color

This parameter is characterized by an accelerated oxidation of the oil, which may also carry anabnormal smell.

At the laboratory, color helps determine whether makeups have been performed (decrease of thelevel of color in the same tint) or whether any the oil is oxidizing (increase of the color of thesame tint). In this last instance, the laboratory must perform acidity measurements and check thewater content in the oil and its infrared signature to determine the origin of the oxidation.

Tint variations detected by oil charge follow-up can indicate that different, yet miscible fluidshave been introduced or that bacteria or mushrooms are developing.

Viscosity

This parameter represents the primordial function of the lubricant. Yet, what the laboratory looksfor is an eventual anomaly during oil changes or makeups. Pollution by other fluids (forexample, solvents, fuel, or regulating fluid) and the formation of “light cuts” by cracking (forexample, degradation of the oil in contact with a heating resistor upstream of centrifuges) can benoticed. On older oil charges, it can be observed that viscosity increases continuously. Variationsmost often acceptable are 10% of the viscosity of a new oil charge.

Level of solid impurity

This parameter is very much influenced by sample quality. On nuclear power plant machines, thelow levels observed make this parameter more difficult to interpret. Only the lubricating oil incrane step-down boxes and diesel engines shows significant levels of impurity.

On some machines, it is preferable to complete this check with a particle count.

In all cases, observing the sample for impurities under a microscope greatly enhances thediagnosis of this parameter.

Although this check is somewhat limited, it is necessary to keep in mind that solid particles cancause important damage.

A large number of small particles can significantly affect the deaeration of the oil by acceleratingoxidation due to micro-diesel effect.

Particles greater that 15 Pm, depending upon their hardness, can behave as sharp micro-tools,damaging bearings and pumps.



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Acid number

This parameter characterizes the consumption of certain oil additives; moreover, it characterizesthe oxidation of this oil with acid radical formation. This measurement does not represent all theintrinsic properties of the oxidation phenomena, but indicates a trend in degradation of the oilcharge.

This measurement must be completed with additional tests:

x Level of catalyzing metals as well as small particles.

x Water content. This element removes a large proportion of the polar additives from the oilcharge.

x Measurement of the oxidation lines in the infrared.

x Check for the formation of “gum” and its evolution during observation under the microscope.

These additional checks can show whether polymers are likely to form that can combine oxidesor chips in the high-pressure areas of the bearings.

Water content

This measurement is essential to determine the integrity of the oil charge.

As a matter of fact, oil is characterized by its lubricating properties, but also by the sideproperties that its additives provide.

Unfortunately, a large majority of these additives are polar and highly soluble in water. Thus,every time the oil charge makes contact with water, some of the additives are removed from thecharge, thereby affecting the characteristics of the oil.

Therefore, it is crucial to determine quickly whether water has been accidentally introduced intothe oil charge. Water can come from a leaky heat exchanger, a poor seal in a bearing, orabnormal breathing of the oil tank.

Foaming characteristics

This measurement determines the tendency of the oil charge to foam, that is, to cause a more orless important volume of foam above the oil.

This foam, depending on its volume, can affect the monitoring instruments of the lubricatingcircuit.

Therefore, depending on the circuit type, this parameter can indicate whether the machine isworking poorly.

Also, foam can increase the oil charge oxidation rate by increasing the contact area betweenoxygen in the air and the oil.



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Residual stability to oxidation (RBOT)

This test measures the tendency of the oil to oxidize within a given amount of time. In fact, theoil charge contains anti-oxidant additives that degrade as the oil ages. Thus, the measurementconsists of determining the amount of anti-oxidant left in the oil. This is a difficult and, thus,expensive test to perform. Therefore, it is often better to use infrareds to measure the quantity ofanti-oxidants if the supplier listed them when selling the oil.

This measurement looks at all oxidation phenomena and the counter-effect of the additives,although there is no need to know the details of the product, often kept secret as an industrialintellectual property.

Deaeration time

This parameter is essential for knowing the lifetime of the oil charge. Depending on how thecircuit was designed, meaning the amount of time that the oil spends in the oil tank, the amountof air that mixes with the oil will have some influence.

Oil may tend to oxidize more quickly if this parameter increases.

The presence of small particles in the oil can affect this parameter. These particles carry micro-bubbles, thus increasing the amount of air inside the oil charge. Depending on how the circuitwas designed and whether pressure spikes are encountered, micro-diesel phenomena can occur,that is, oxidation or localized combustion.

This measurement must be interpreted considering the following parameters:

x Measuring oxidation or cracking lines in the infrared

x Measuring the acidity level

x Testing for odor

x Looking for small particles and gum under a microscope

Demulsification time

This parameter is very significant. Since additives dissolve in the water contained in the oil, it ispreferable that they stay in contact with accidental water leaks for the least time possible. Thisparameter quantifies the separation time between the water and the oil.

A large water leak with a short demulsification time can be less significant for the oil charge thana small water leak that separates slowly with the oil. In fact, the effect on the level of additives isdirectly proportional to the time the water and the oil are in contact.

This parameter can be affected by the formation of oxidation products within the oil charge.Therefore, the previous parameter (deaeration) must also be considered while checking for thiscriterion.

The introduction of surface-active pollutants during maintenance tasks affects this parameter inan irreversible way.



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Copper corrosion

This test determines whether the oil charge can corrode copper alloy materials.

These corrosions can be the result of acid radicals formed by oxidation of the oil.

This measurement is only very relative and hardly predictable. Thus, this parameter cannotalways be interpreted and varies from one type of machine to the next.

Rust-inhibiting characteristics

This parameter monitors the quantity of rust-inhibiting additives contained in the oil chargeduring operation.

The results determine whether the oil charge is being well used (for example, starting the oilcleaner, scheduled cleaning of the oil pit, and water-level reduction).

This parameter can also indicate whether a highly corroding agent has been introduced that couldcreate an electrochemical corrosion.

Level of particle contamination

On machines where operation is critical or clearance tolerances are small, this parameter canquantify the use range of the oil charge.

In that case, the mass of particles is often very small, but their granulometric distribution isimportant. Particle contamination tests are performed to ensure that the oil operating range iswithin the machine manufacturer's specifications. Otherwise, the oil must be filtered andpurified.

It is good to complete this test by using a microscope to determine the type of particles, theirorigin, and their likelihood for damage to the machine.

Infrared spectroscopy

This technique checks for different phenomena within the oil charge:

x Measurement of the level of additives, mainly antioxidants and antifoaming agents

x Measurement of the cracking process upon contact with hot points

x Quantitative measurement of the oxidation products that formed

x Quantitative measurement of soluble pollutants, such as regulating fluid or fuel

x Assessment of accidental introduction of aromatic solvents during maintenance

This technique requires a technician skilled in critically recognizing eventual degradations ofpollutants present in the oil charge. This critical observation is important to assess the severity ofthe issue and its origin.



5-12

Quantitative analysis of wear metals

These analyses are performed using atomic absorption spectrometry or emission spectrometry byplasma torch. They are used to determine the number of mineral particles in solution or smallparticles (<5 Pm) in the oil charge.

They determine the concentration of mineral additives in the oil as well as particles presentbecause of wear or pollution.

These analyses only look at elements smaller than 5Pm. Thus, only the evolution of the numberof particles should be considered.

The necessary complement to these analyses is microscopic observation of the particles retainedby a 0.8 Pm filter.

It is possible to observe copper compounds with a size between 20 and 50 Pm in a microscopewithout using the atomic absorption spectrometer.

For precise information about compounds present in the oil, a small portion of the sample shouldbe mineralized. This operation is possible but expensive and would bring little value to thediagnosis.

These analyses can confirm the microscopic observations, especially when anti-friction coatingmaterials have been detected.

Microscopic observation of the particles in suspension after 0.8 PPm filtration

These observations are essential to evaluate the degradation that could affect machines. Anenlargement of 100 is sufficient to see the potentially damaging particles.

But what can we see during these observations?

x For new oil charges before a makeup or an oil change, the test serves mainly to look forparticles likely to behave as micro cutting tools.

These particles are mainly thermal insulation fibers (fiberglass) or remains of fibers whosehardness may be harmful to the metals. It is also possible to find steel or copper metal chipsoriginating from a degrading emptying pump on the supply trucks during delivery. Theseparticles are usually large (50 to 200 Pm). The best solution is a properly handled filtrationduring the makeup.

For new oil charges, performing precipitation tests for additives or paraffinic products canhelp reveal decantation phenomena that can occur during cold temperature or long standingperiods. These products are shaped like a fluffy and transparent gel. The best solution is tocirculate this oil charge for a period of about 48 hours.



5-13

x Oil charges lubricating anti-friction alloy-coated bearings require particular attention fordetecting this metal. After the coating degrades, it looks like a small, flat chocolate candywrapper. Detecting it is important, yet difficult due to the small ratio between the surface ofthe anti-friction coating material and the volume of oil.

On the same types of machines, detection efforts are mainly concentrated on finding steelchips larger that 20 Pm that may affect the bearings.

The oil charge degradation products can be recognized by the formation of developedpolymers (ranging from diluted to burned and strained, indicating lacquer). These lacquers,also referred to as “gum,” can be found in the exhaust area of the pressure peak of thebearing where they stop the larger chips present in the oil. Then, this area behaves as abrasivepaper.

The presence of these gums is quite normal in oil charges that are old or affected by baddeaeration. Their follow-up can provide an estimate for the lifetime of the oil underacceptable conditions.

On these types of machines, it is possible to observe degradations of retention pit paint,appearing as polymer flakes, often colored. This detail implies the necessity to determinewhether this type of paint has been used on the machines.

x Machines using copper radiators or cryogenic compressors often show copper alloy erosionparticles. It is necessary to recognize the importance of this finding, depending on the size ofthe torn metal or the global quantity observed on the filter. This can indicate the disturbanceof stream configurations.

x On machines with step-down gearboxes, testing concentrates on the forces that generatemicro-welded type particles. The coloration of these particles is changed, and areas where thecrystallography of the metal structure appear are torn. These observations can indicate thatthe gear profile is not well suited, that a non-extreme pressure oil is being used, or that thegearbox is overloaded.

x Dirty machines or machines with limited maintenance cleanliness present populations ofoften large oxides, sandy-looking mineral products or remains of fiber very harmful to thesystem. The proposed solutions are, after having filtered the oil charge, to check theventilation of the circuits or the oil pits, and to clean the bottom of the oil pits at everyshutdown that allows it. Controlling the thermal insulation atmospheric pollution on andaround the installation sites should also be done.

x The shape, color, and dimensions of torn chips can give an estimate of the forces present,thus allowing the distinction between grinding phases and the mechanical influences that cantake place.

x Knowledge of the circuits and identification of the particles can help determine the affectedcomponents and, thus, the criticality of the defect. The quality of this check is directlyinfluenced by the appreciation of a few people with a significant level of feedbackexperience.



5-14

Waste management

One objective involving oil charges follow-up can and must yield a better control over thequality of these charges and an improved human behavior during interventions performed on oraround the oil.

Meeting this objective should decrease the amount of oil replaced, either conventional orcontaminated.

The financial cost associated with replacing the oil charges is relatively important. New productsmust be bought and old ones destroyed.

This behavior also reinforces the image of the company. Although this follow-up has aninstantaneous heavy cost, the feedback is more than satisfactory.

Follow-Up

The synthesis of all the previous checks must be compared with the historical data collected onthe machine or the oil charge. The follow-up can be used to observe the following:

x The behavior of the oil charge, along with the makeups and eventual replacements, the wearon the machines as well as accidental pollutions. In this case, there is really a follow-up ofthe oil-machine combination.

x On some machines, the oil charge is replaced every time the machine is serviced, and the oldoil charge is tested. This type of follow-up serves mainly to evaluate the degradation processof the machine. However, a more careful analysis of oil behavior can increase its replacementperiod to every two or three shutdowns. In this case, this follow-up allows eventual reductionin maintenance and unavailability costs. Also, it can decrease waste production.

x The most often used follow-up tool is a computer system that can be accessed permanentlyby the control operators in order to have the reactivity best adapted to the collected samples.It is still possible to visually observe samples originating from five to ten previous samplesand, sometimes, farther back.

x The only constraint concerns the identification of the machines or the sampling valves. Wereturn to the previous comments on proper sampling and labeling techniques.

5.5 Conclusions

Beyond the data readily available to technicians, the testing of lubricating oils byphysicochemical follow-up in the laboratory can yield a directly exploitable diagnosis onmachine maintenance.

The quality of this diagnosis directly depends on the quality of the samples. Therefore, onlyskilled and motivated operators should perform this operation.


6-1

6 R.R.A. (RHR) PUMP BEARING GREASEQUALIFICATION TESTS


No. 97NB00115

Authors:D. Buchdahl

R. MartinP. Domon

J.M. Girault




R.R.A. (RHR) Pump Bearing Grease Qualification Tests

6-2

Executive Summary

Pressurized water reactor (PWR) power plants are equipped with shutdown reactor coolingpumps (R.R.A.s). (Note to U.S. facilities: These are comparable to residual heat removal pumps.)These are installed inside the containment enclosure (Reactor Building) and convey the primarycircuit water. In case of emergency, they play an important safety role; in a small breach steampipe failure or a primary cooler loss accident, they are used to cool the primary circuit.Therefore, these pumps must be capable of continuing to operate in an irradiated and/or degradedenvironment.

The rotational support of these pumps consist of grease-lubricated bearings. To providecontinued operation of the equipment, the grease used must retain its lubricating power under allconditions to which it may be subjected.

Therefore, a qualification procedure has been defined that consists of two series of independenttests on benches reproducing the operating conditions of the grease in the bearings, namely, thethermodynamic test and the irradiation resistance test.

The thermodynamic test is conducted in a climatic enclosure and reproduces profile K1 of theRCC-E temperature and pressure (RCC-E is the French acronym for Conception andConstruction Rules for electrical equipment of nuclear power plants).

The irradiation resistance test is conducted in a pool and reproduces the cumulative dose ofradiation in the case of a reference accident. Several series of tests in addition to the qualificationtests have made it possible to validate this procedure and to perfect the test benches. Theirradiation resistance tests are conducted on three types of grease. The results are conclusive.



6-3

6.1 Introduction

Pressurized water reactor (PWR) power plants are equipped with shutdown reactor coolingpumps (R.R.A.s), or residual heat removal systems. They are installed inside the containmentenclosure (Reactor Building).

These pumps convey the primary circuit water. In case of an accident, the pumps play animportant safety role. In case of a steam pipe failure or a primary cooler loss accident smallbreach, they are used to cool the primary circuit. Therefore, these pumps must be capable ofcontinuing to operate in an irradiated and/or degraded environment.

The rotation support of the pumps consist of grease-lubricated bearings. To ensure continuedequipment operation, the grease used must retain its lubricating power in all conditions to whichit may be subjected.

Because of the short commercial lifetime and the variability in quality from shipment toshipment, EdF established a qualification procedure for grease to be used in R.R.A. pumpbearings.

This procedure determines whether the products supplied to the plants have the requiredqualities. Also, this qualification procedure would help select new greases if currently usedproducts were to be removed from the market.

This quality control method is representative of the operating conditions of the R.R.A. reactorcooling pumps. It is simple, reliable, and accurate.

6.2 Loading Specifications

As a general rule, the stability of grease to irradiation is determined with a static test.

This static test consists of exposing to radiation a test tube filed with the lubricant to be tested.The stability of the grease to irradiation is then estimated according to its variation from acharacteristic value.

That characteristic value is most often a measure of penetrability or of a dielectric constant.However, the degradation of a grease exposed to radiation induces chemical reactions, such asoxidation.

In this type test, only the surface of the product is in contact with the ambient air. Therefore,oxidation phenomena are limited to a small portion of the lubricant.

In a rotating bearing, the grease is permanently being mixed. It makes contact with the oxygenpresent in the air in its entirety. Moreover, it is subjected to important mechanical loads.

Therefore, a static test is not very representative of grease behavior in a roller bearing.Consequently, this method is not applicable to grease used in R.R.A. pump roller bearings.



6-4

For a more accurate picture of grease behavior, tests are performed on mechanically loaded rollerbearings in rotation that are subjected to environmental conditions present during a design basisaccident.

Environmental loads within the range of the design basis accident (primary cooler loss accident,large breach) are defined by K1 qualification specifications. They define the irradiation levelsand thermodynamic profile to take into account. These irradiation levels are actually of 250 kGyfor aging and 600 kGy for an irradiation accident. The K1 profile is included in Figure 6-1.

Accidental situations requiring R.R.A. pumps to operate are a loss of primary coolant accident(small breach) and a steam pipe failure. The accidental loads due to a loss of primary coolantaccident (small breach) and the loads due to the thermodynamic environment present duringsteam pipe failures (high pressure and temperature, steam) are much less than the loads presentduring a K1 type design basis accident. In addition, they do not occur at the same time.

In addition, the grease used on R.R.A. pumps is replaced frequently. Thus, the aging factor isnegligible and neglected.

For the qualification procedure, the following tests are performed:

x A test of irradiation resistance at the maximum level of 100 kGy (simulating the radiation towhich the grease is subjected after a primary cooler loss accident, small breach),

x A thermodynamic aging test according to the K1 profile defined by the R.C.C.E. (simulatingthe internal environment in the containment building after a steam pipe failure).

These two tests are conducted independently of each other.

To determine the grease resistance tolerance levels, the irradiation resistance test is continued upto a maximum level of 600 kGy, if possible.

The irradiation resistance test is conducted in the “Caline” cell at the C.E.A. (Commissariat àl’Energie Atomique, the French research institute for nuclear matters).

The thermodynamic aging tests occur in one of the ADR enclosures of EdF/DER (Direction desétudes et recherches) at the Renardières laboratories or one of the SOPEMEA enclosures in thecity of Vélizy.

For the qualification test, the grease lubricates the roller bearings of a test bench that models asaccurately as possible the operating characteristics of the R.R.A. pumps.

6.3 Description of the Test Benches

The test bench is designed from loading data collected on the R.R.A. pump roller bearings at theBugey PWR plant (900 MW) and the N4 plants (1400 MW).



6-5

For obvious reasons, it is not possible to conduct tests on a 1:1 scale model. Consequently, onemust define proportion coefficients to design the roller bearings of the test bench. They are:

x Rotation factor: N u dm

x Load coefficient: C / P

x For angular contact ball bearings:The ratio between the axial load and the radial load: Fa / Fr

The rotation support on the R.R.A. pumps is made with:

x A load-bearing cylindrical roller bearing, supplied by S.K.F.

x A thrust bearing made of two angular contact ball bearings. They are set up as an X (seeFigure 6-2). They have a 40q contact angle and are supplied by S.K.F.

The characteristics of the roller bearings of the R.R.A. pumps at the Bugey PWR plant and thoseof the N 4 plants are summarized in Table 6-1.

Table 6-1Types and Characteristics of the Roller Bearings Mounted on the R.R.A.Pumps at the Bugey PWR Plant and on the N4 Plants

Bugey N4

Type of rollerbearing

Angularcontact ball

bearing7317 BM

RollerNU 220prior to

1985

RollerNU 20EC

since1985

Angularcontact ball

bearing7322 BM

RollerNU226prior to

1985

RollerNU226EC

since 1985

Inner diameter(mm) 85 100 100 110 130 130

Outer diameter(mm) 180 180 180 240 230 230

Rotation factorN u dm (mm-rpm)

198,750 210,000 210,000 262,500 270,000 270,000

Loadingcoefficient (C /

P)12.3 45.6 62.6 10 45.7 60.6

Ratio betweenthe axial loadand the radial

load Fa / Fr

8.1 _____ _____ 11.1 _____ _____

For practical reasons, the rotation speed is kept at 2,900 rpm for the test bench.



6-6

To most accurately approximate the proportion condition for the rotation factor (N u dm), a shaftwith a 60 mm diameter is used.

The roller bearings that equip the test bench are:

x Angular contact ball bearing S.K.F. 7212 BEM

x Load-bearing roller bearing S.K.F. NU 212 EC

Their characteristics are shown in Table 6-2.

Table 6-2Characteristics and Loading of Test Bench R oller Bear ings

Type of Rolling Bearing Angular Contact Ball7212 BEM

Roller NU 212 EC

Inner diameter (mm) 60 60

Outer diameter (mm) 110 110

N u dm246,500 246,500

C / P 12 62.3

Fa / Fr 11 -

Equivalent dynamic load P(N)

4767 1500

Axial load Fa (N) 4850 0

Radial load Fr (N) 440 1500

The test bench consists of a shaft, the roller bearings to be tested, one "slave" roller bearing, acoupler, and a drive motor (see Figures 6-2 and 6-3).

The test bench shaft has a diameter of about 60 mm. It is equipped with two angular contact ballbearings at one end, one slave cylindrical roller bearing, and at the other end, a cylindrical rollerbearing.

The radial load on the tested bearings is applied by the slave roller bearing and the adjustableloading system consisting of a screw and elastic washers.

The radial load distribution among the different roller bearings occurs at the assembly stage anddepends on their respective axial positioning. The distances between the roller bearings areminimized in order to keep the bench as compact as possible.

The slave roller bearing selected is a cylindrical roller bearing, type S.K.F. NU 2211 EC.

The axial load on the angular contact ball bearings is applied by helicoidal springs.



6-7

The design of the other parts (for example, struts) is as close as possible to that of the parts onthe R.R.A. pumps.

The materials used on the pumps of the N4 bearing are austenitic or martensitic stainless steels(casing: Z3CND19.10, shaft: Z5CND16.04).

The test bench was manufactured using Z30C13 steel. This is stainless steel (to facilitatefrequent assembling and disassembling) with a high hardness level and a thermal expansioncoefficient close to those of the actual pump.

For tests in an irradiated enclosure, the bench is equipped with a special drive motor resistant toradiation and environmental conditions (temperature: 70qC). It has a power of 600W. It isequipped with roller bearings that are lightly loaded and lubricated with the test grease.

The coupling uses flexible stainless steel blades.

The roller bearings are replaced after each test.

The tests in the A.D.R. enclosure are conducted with a modified version of the test bench. Infact, during the first tests, condensation occurred on the motor, and electrical insulation defectsappeared. Thereafter, we were forced to equip the bench with a magnetic coupler and mount themotor outside the enclosure.

6.4 Operating Mode

New roller bearings are used for each series of tests. They are filled with grease from plantinventory. The filling of the bearings is accomplished according to standard procedures:

x The roller bearings are completely filled.

x The free space inside the bearings is partially filled (between 30 and 50%).

The slave roller bearings mounted on the bench and the bearings mounted on the motor arereplaced before each test and lubricated with the same grease as that being tested. The loadsapplied upon them are light; therefore, these bearings do not influence the test.

The proper mounting of the roller bearings is checked by operating the bearing for 24 to 36 hoursoutside the A.D.R. irradiation enclosure.

6.4.1 Irradiation Resistance Tests

The irradiation resistance tests are performed inside the C.E.A.'s Caline enclosure.

They are conducted using three identical test benches mounted horizontally on a cradle. Theassembly is installed inside a watertight box and submerged in a pool (see Figure 6-4).Atmospheric pressure is maintained in the box, and a warm flow of air keeps the temperature at70qC.



6-8

Two rows of radioactive sources slide on a rail mounted on both sides of the box. The dose rateis around 800 Gy/h and remains constant throughout the test. The maximum dose inflicted to thebench is about 600 kGy. This corresponds to a maximum exposure time of 30 days.

During this test, the only parameter continuously monitored is the electrical power consumed bythe drive motors mounted on the benches. It must stay below 600W. If it is greater than 600W,the bench is stopped. The benches are disassembled only after the maximum irradiation dose isreached. Then, the roller bearings are visually inspected.

6.4.2 Thermodynamic Tests

The thermodynamic tests (see Figure 6-5) are conducted in one of the available enclosures(either at the Renardières laboratory or at SOPEMEA).

The test conditions are the following:

x Temperature: from 50 to 156 qC

x Pressure: 1 to 5.5 bar

x Hygrometry: 100%

x pH: 9.25

x Boric acid concentration (H3Bo3): 1.5%

x Sodium hydroxide concentration (NaOH): 0.6%

The pressure and temperature cycles follow a determined cycle according to the K1 profile of theRCC-E (see Figure 6-1). The test bench is set in rotation four hours after the second thermalshock. Again, the only parameter monitored during this series of tests is the electrical powerconsumed by the drive motors mounted on the benches. It must stay below a predeterminedvalue. If it is greater than this value, the test is cancelled. At the end of the test, the rollerbearings are visually inspected.

6.5 Results From Irradition Resistance Tests

Until now, only irradiation resistance tests have been conducted. We have tested three types ofgrease. They are designated A, B and C. The new grease samples were supplied by the Bugeyproduction plant in February 1993.

The dose rate is about 800 Gy/h. The cumulated dose of radiation that was applied is about 600kGy. When the power consumed by the bench reaches 600W, it is stopped automatically.



6-9

Each of the three test benches is equipped with four new roller bearings greased according tostandard procedures. The roller bearings and their respective loads are shown in Table 6-3.

Table 6-3Type of Roller Bearing and Mechanical Loading Values for Irradiation Resistance Tests

Roller BearingReference

Roller Bearing Type Axial Load daN Radial Load daN

SKF 7212BEMMassive brass

casing

1 and 2 Angularcontact ball

485 44 on each rollerbearing

SKF NU212Massive steel

casing

3 Cylindricalroller

0 150

SKF NU2211ECMassive metal

casing

4 (slave) Cylindricalroller

0 62



6-10

The results of these tests are summarized in Table 6-4.

Table 6-4Summary of the Results of Irradiation Resistance Tests

Greasetype

Bench#

Consumedpower at

the end ofthe break inperiod (w)

Cumulateddose of

radiationwhen thebench is

definitivelystopped(p>600w)

Observations

Roller bearings #1and #2

Roller bearing#3

Roller bearing#4

1 73 120

2 74 235

Grease mildlydegraded:

hardened, blackand sticky. Casingof #1 and #2 roller

bearing broken(see Figure 6-6)A

3 85 78

Grease not verydegraded:

hardened andsticky

Grease totallydegraded. Only

a non-lubricating,sticky black

deposit remains.

Grease slightlydegraded:

hardened andsticky

1 119 300

2 130 245B

3 134 180

Grease mildlydegraded: black

and sticky

Grease totallydegraded. Only

a non-lubricating,sticky black

deposit remains(see Figure 6-7)

Grease slightlydegraded, sticky(see Figure 6-8)

1 126 405

Grease slightlydegraded:

hardened, blackand sticky

2 108 300

Greasedegraded,

hardened, blackand sticky (see

Figure 6-9)

C

3 132 372

Grease degraded:hardened, black

and sticky

Grease totallydegraded, blacksticky and non-

lubricatingdeposit

Broken casing,grease totally

degraded, blackand sticky



6-11

At the end of the test, the benches were disassembled to visually inspect the roller bearings.

In summary, the following were determined:

x The roller bearings mounted on the motors are in a relatively good shape (see Figure 6-10):the grease that filled them retained a portion of its lubricating properties. The oil bleeds whenthe roller bearing is set on a piece of paper. These roller bearings are lightly loaded.

x The grease deteriorates more when heavier mechanical loads are applied (for example,loading and rotation speed).

x It appeared that the grease that filled the #3 roller bearing deteriorated the most. This rollerbearing should operate with a large loading coefficient (C/P). In our case, it operates in anunderload mode. Its rollers tend to slip. Consequently, its lifetime is unpredictable.

6.6 Conclusion

R.R.A. pumps must have a high level of reliability. Because of the particular operatingconditions to which they are subjected in accidental situations and the short commercial lifetimeof grease, it is necessary to test the grease lubricating the R.R.A. pump roller bearings inaccidental environments.

Therefore, a test procedure has been developed. It consists of performing two series ofindependent tests on benches that model the operating conditions of the grease inside the rollerbearings.

x Thermodynamic test

x Irradiation resistance test

The thermodynamic test is performed in a climatic enclosure and follows the K1 temperature andpressure profile of the R.C.C.E.

The irradiation resistance test is performed in a pool and models the cumulated irradiation dosein case of a design basis accident.

Several series of preliminary tests have validated this technique and contributed to the design ofthe test benches.

The irradiation resistance tests have been performed on three types of grease. The results areconclusive.

The thermodynamic aging tests on these three types of grease have yet to be performed.



6-12

Figure 6-1Thermodynamic Aging Profile According to the K1 Profile of the R.C.C.E.

Figure 6-2Schematic Sectional Drawing of the Test Bench Shaft With the Radial Loading System



6-13

Figure 6-3Sectional Drawing of the Test Bench

Figure 6-4Schematic View of Three Test Benches in the Irradiation Containment Enclosure



6-14

Figure 6-5Schematic View of Test Benches in the Climatic Enclosure



6-15

Figure 6-6Type A Grease, Status of the Numbers 1 and 2 Roller Bearings of the Number 2 BenchAfter a Cumulated Radiation Dose of 235 kGy While Operating



6-16

Figure 6-7Type B Grease, Status of the Number 3 Roller Bearing of the Number 2 Bench After aCumulated Radiation Dose of 245 kGy While Operating

Figure 6-8Type B Grease, Status of the Number 4 Roller Bearing of the Number 2 Bench After aCumulated Radiation Dose of 245 kGy While Operating



6-17

Figure 6-9Type C Grease, Status of the Number 4 Roller Bearing of the Number 2 Bench After aCumulated Radiation Dose of 300 kGy While Operating

Figure 6-10Type B Grease, Status of the Roller Bear ings of the Number 3 Bench Motor After aCumulated Radiation Dose of 180 kGy While Operating


7-1

7 ROTATING MACHINE DYNAMIC SEAL TECHNOLOGY


No. 97NB00119

Author:E. Landrieux




Rotating Machine Dynamic Seal Technology

7-2

Executive Summary

Providing a leak-proof seal between two devices or parts of a system or machine is an everydayproblem for people, whether in industry or not. From the humble rubber washer and fittingsmade up of metallic rings to the labyrinth-type geometry dynamic seals used in hydraulicmachines, the problem is always the same: how to best impede the flow of a fluid between twoareas of different pressures to be isolated.

In the ’60s and ’70s, seal technology saw tremendous improvements, thanks largely to spacetravel and nuclear energy production (nuclear power plants). The diversification of applicationswidened the performance range required of sealing devices; and the importance of criteria, suchas reliability and safety, increased the stringency of the requirements in the field.

The choice of one particular configuration over another depends on several parameters linked tothe immediate environment of the device and a full analysis of the vibrational behavior of themachine shaft line. This is why it is difficult to propose standardized methods that can bescientifically justified for designing seals applicable to all machines, operating conditions, andfluids. Each method calls for a series of hypotheses that only the designer, with all his know-how, can choose in order to find an acceptable solution.



7-3

Labyrinth: Legendary edifice attributed toDedale, composed of a large number ofrooms built in such a way it was verydifficult to exit.

7.1 Design Diversity

7.1.1 Listing of the Different Classifications

The different sealing systems can be classified according to some characteristic criteria. [1]

7.1.1.1 Classification According to the Existence or Lack of Relative MotionBetween the Leak-Proof Elements

x Mechanical seal, where sealing is accomplished by a mechanical force (spring compression)

x Static seal, where there is no relative motion between the adjacent parts that make up the seal

x Dynamic seal, where the adjacent parts of the seal are in motion relative to one another

7.1.1.2 Classification According to the Direction of the Sealing Interstice

x Axial seal

x Radial seal

7.1.1.3 Classification According to the Existence or Lack of Sealing Material

x Direct seal, where the sealing structure components are in immediate contact with oneanother

x Indirect seal, where components are separated by a leak-proof material

7.1.1.4 Classification According to the Existence or Lack of the Interstice

x Seals with contact, where the leak-proof devices are in contact (without an interstice), such asseals by crushing, pressure seals, seals made of adhesive or elasto-flexible (with spring orelastic ring) materials, sliding joint seals, hydrodynamic seals, ring seals, lip seals, elasticenvelope seals, viscosity seals, cohesion seals (welding, brazing, gluing), or electromagneticseals,

x No contact seals (see Figure 7-1), where the adjacent surfaces forming the leak-proof jointare not in contact, such as constant thickness interstice seals, hydraulic seals operating on theprinciple of communicating vessels, bottleneck seals, pressure generating seals, labyrinth-type joints (in which the space is partitioned in several chambers on one of the connectingparts), and labyrinth-type and interstice joints.



7-4

7.1.2 Criteria of Choice

As the previous listing demonstrates, sealing devices show a large diversity. Different industrialdesigns have generated specific know-how, and the evolved stringency on performance andreliability has lead to research for new designs.

The main technical criteria to consider in choosing an appropriate sealing device are the specificpurpose for its operation and the conditions to which it will be subjected.

7.1.2.1 Purpose

Several reasons may require the use of a sealing joint:

x To control or prevent the flow of a fluid

– Too large a leak is negative for the productivity of the machine (for example, the jointbetween the pump wheel and the flow nozzle prevents the loss of energy by leakage).

– A leak must be controlled as well as feasible to avoid mixing a contaminated orradioactive fluid with a clean medium.

x To control or prevent gas flows or diffusion

x To raise a thermal barrier (A joint can reduce the flow of primary circuit hot water towardsensitive areas.)

Tolerance to leakage depends on the specific design purpose of the joint. For classicalapplications, a leakage rate of 2% to 5% of the total flow rate circulating in the machine can betolerated.

7.1.2.2 Thermodynamic Conditions

x Pressure conditions between the lower and the upper part of the sealing device

The fluid to seal pressure varies within a wide range. It can be lower than atmosphericpressure (vacuum of a condenser pump) or reach over 150 bar (in the primary circuit pumpsof a PWR reactor) [ 8].

x Temperature conditions

Similarly, the temperature range is very broad, from liquefied gas at -200qC (cryogenics) tofluids at normal temperature, to fluids with temperatures over 200qC (superheated waterstations in power plants) [8].

x Type of fluid

The fluid can be a liquid or a gas (sometimes diphasic: petroleum industry) and haveparticular chemical properties (toxic, flammable, crystallizable, radioactive, or abrasive) [8].

x Rotation speed when the parts are moving

The peripheral speed of a pump shaft can range from a few meters per second to over70 m/s [8].



7-5

7.2 Labyrinth-Type Joints

7.2.1 Definition

This section looks at dynamic joints with no contact of interstice and/or labyrinth type. This typeof joint is generally used to cover machine sealing needs where a part rotates, such as the rotoron a turbomachine.

The fluid can be incompressible, as in pumps, or compressible, as in steam turbines orcompressors.

This sealing device has the benefit of avoiding contact between the rotating parts, thussuppressing direct frictional forces. This is why it is widely used in more recent machine designswhere the rotation speed of the shaft becomes greater and greater and where losses due to frictionmust be limited. The use of a well-designed labyrinth-type joint has additional benefits: makeupor frequent maintenance lubrication is not necessary.

7.2.2 Description and Operating Principle

Of course, with such a no-contact joint, it is not possible to eliminate leakage, but it is possible toreduce it to an acceptable level. Selecting a joint depends on the leakage tolerance criteria inaddition to design constraints, thermodynamic or otherwise.

In this section, the most simple device is the joint previously referred to as "interstice joint."(See Figure 7-2).

The operating principle consists of restricting the fluid to a path over a set length to controlleakage. For this, the machine is designed so only a very narrow clearance exists between theshaft and the flange. This generates an interstitial area preceded by a very large sectionalrestriction. Selection of the restricted path and the length of the interstice depend upon theacceptable level of leakage. The important pressure loss at the interstice entrance, due to thesharp restriction plus the pressure loss due to friction along the interstice, control the leakage ratebetween the upstream side of the device where the pressure is high and the downstream sidewhere the pressure is lower [3].

This principle based on pressure losses can be used more efficiently in joints previously referredto as "labyrinth-type" joints (see Figure 7-3).

The fluid must circulate through particular geometrical shapes, like small chambers that force itto accelerate and decelerate through successive narrowing and widening sections. This conditiondissipates the energy and the pressure drop, thus reducing the leakage rate.

More precisely, the sealing mechanism of the labyrinth-type joint is illustrated in the abovediagram. The upstream pressure that must be dissipated through the successive chambers of thelabyrinth-type joint reduces to the downstream pressure level and eventually down toatmospheric pressure. When the upstream fluid flows through the first chamber, three zones of



7-6

flow can be recognized. In zone 1, the fluid accelerates and reaches its maximum speed. In zone2, the fluid spreads, decelerates and dissipates its energy as turbulences. Then, the fluid fills zone3 with a relatively low speed and a reduced pressure level. The first chamber of the labyrinthmust be large enough to let the energy of zone 3 dissipate [4].

When the interstitial areas alternate with the chambers, different effects cumulate. The totalpressure loss then comes from the losses due to friction in the clearance, losses at the sharpwidenings and narrowings, and energy losses inside the chambers where a core of constant massseparates from the stream [3].

From this principle, a wide range of designs was invented:

x Step interstice joint or joint with curves in series (see Figure 7-4)

x Labyrinth-type joint with grooves on the rotor, the stator, or both; joints with projections;comb-shaped joints (see Figure 7-5).

7.2.3 Parameters of Choice

7.2.3.1 Dimensioning Characteristics

The dimensions characterizing an interstice and/or labyrinth-type joint are geometrical andoperating parameters.

The different configurations previously mentioned highlight several design parameters: lengthand clearance of the interstices; position, number, width, length, and depth of the grooves; andconstant or variable radii configurations in the direction of the grooves.

The surface conditions of the parts facing each other are also important, particularly where thefluid is turbulent. Pressure losses are influenced by the roughness of the walls. In this way, these"honeycomb sleeve" joints constitute a particular geometry midway between a labyrinth-typejoint and a rough-walled joint.

The relative position of the sealing surfaces also influence the behavior of the joint:

x Eccentricity characterized by the eventual gap between the position of the centers of the twocomponents making up the joint

x Misalignment characterized by the eventual angular gap between the axes of the twocomponents making up the joint

The operating parameters are mainly the following:

x Pressure differential between the media separated by the joint

x Residual flow rate

x Fluid characteristics: type, density, viscosity

x Temperature gaps



7-7

x Flow characteristics: laminar or turbulent, affected by flow speeds

x Rotation speed of the moving parts

Finally, the labyrinth-type joint can be characterized by the following:

x Load and inclination angle when eccentricity exists

x Forces resulting from the interaction between the fluid and the structure that can be modeledas a system of stiffness, damping, and added mass.

Material used in manufacturing the joint can be a relatively soft metal like aluminum or brass sothe rotor will not sustain damage if it makes contact with the joint [2]. Alternatively, it can bemade of carbon or a material that sublimates instead of melts [5].

Figure 7-6, a schematic diagram of a multi-cellular pump indicates the different sealing joints:flat louver ring, inter-stage ring, and balancing drum.

Figures 7-7 and 7-8 show the possible evolution of the leakage rates as a function of shape,rotation speed, or pressure differential between the sealed media.

7.2.3.2 Influence on Vibrational Behavior

Even though the primary function of a sealing joint is to limit a leakage of fluid or gas, and not tobear the shaft of a turbomachine, interactions between the fluid and the structure influencesubstantially the dynamic behavior of the rotors.

This is why for several years, theoretical and experimental research work has been done tounderstand and to predict this effect [12].

The models are based mainly on breaking up the radial loads exerted on the labyrinth-type jointinto a stiffness component proportional to the displacement, a damping component proportionalto the speed of motion, and an added mass component proportional to the acceleration.

Taking into account this effect is necessary:

x To better determine and/or modify the critical speed values of the shaft line (particularlysensitive to the stiffness terms)

x To avoid the instabilities caused by crossed forces (that is, forces acting in a directionperpendicular to the displacement of the shaft)

This is why the Machines Division of the Research and Development Department of EdF hasbeen searching for several years for validated means of calculation capable of evaluating themain parameters of a dynamic joint (for example, leakage rate and dynamic coefficients) as afunction of its shape and operating conditions. Each type of geometry—interstice joint orlabyrinth-type joint—uses different flow hypothesis leading to numerical models specific to eachfamily of sealing joints. The calculation modules are integrated in the EDYOS software package,a part of the CADYAC (dynamic calculation of the shaft lines) package.



7-8

Figures 7-9 and 7-10 show graphs indicating some values of the dynamic coefficients to beconsidered.

7.2.3.3 Wear of the Labyrinths

Among the parameters defining the shape of the interstice and/or labyrinth-type joint, clearanceappears essential. At first approximation, clearance, along with the length of the joint,dimensions leakage as a function of the pressure differential between the zones to be sealed.

Although this type of joint has the singularity of operating without any contact between thesurfaces, the wear of the surfaces facing each other must be considered when choosing the shapedue to the operating conditions (liquid with particles). In fact, as the gap widens, the leakage rateincreases, and the parameters that influence the dynamic behavior change.

However, deformations (thermal effects) of the rotor and the machine are possible. They alsochange the operating conditions of the joint by decreasing or increasing the clearances andgenerate misalignment or eccentricity. The contact between the surfaces generates seriousproblems and must be avoided.

Finally, a careful inspection of the condition of the labyrinths on a turbomachine can be veryenlightening about its operation [5]:

x Deep grooves on the wheels or the shaft indicate that the shaft is shifting (largedisplacements) at some point during its operating cycle.

x Worn or corroded labyrinths lead to loss of efficacy; and if this occurs on the balancingpiston, the thrust bearings may be deteriorated.

x Friction marks may indicate bad operating conditions, such as operating close to a criticalspeed or a jerky operation, a problem in the dynamic of shaft line, a bending of the shaft dueto heat, or some other similar difficulty.

7.3 Conclusion

The different points mentioned have highlighted the vast variety in sealing devices, even on themere section on dynamic interstice and/or labyrinth-type joints without contact.

The choice of a particular configuration depends on several parameters closely related to thenearby environment of the device, and on the complete analysis of the vibrational behavior of themachine shaft line. That is why it is difficult to propose standardized design methods that arescientifically justifiable and applicable to all machines, operating conditions, and fluids. Eachmethod requires a series of hypotheses that only the manufacturer, with his know-how, canchoose to find an acceptable solution.

Nevertheless, progress is made in flow theory and experimental techniques, particularly forlabyrinth-type joints. The essential parameters, such as the leakage rate and the dynamiccoefficients (for example, stiffness, damping, and added mass) taken into account whenanalyzing the vibrational behavior of the machines, are being evaluated with more and moreprecision. Altogether, these improved techniques constitute a precious help in behavior analysisand design improvements. Efforts in this direction must not be relaxed.



7-9

7.4 References

1. Dictionary of Hydraulic MachineryAdam Tadeusz TroskolanskiElsevier1985

2. Dynamic Seal Technology: Trends and DevelopmentsPart 1: Seal TypesOtto DeckerMechanical Engineering03/1968

3. Mémento des pertes de chargeI.E. Idel'cikEyrolles1986

4. Industrial Sealing TechnologyH. Hugo ButcherJohn Wiley & Sons1979

5. Machinery Component Maintenance and RepairHeinz P. Bloch - Fred K. GeitnerGulf Publishing Company1990

6. Seals and Sealing HandbookThe Trade & Technical Press Limited1986

7. Pompes centrifuges et pompes hélicesA.J. StépanoffDunod1961

8. Théorie et technology des pompes centrifugesÉcole de thermiqueDelplanque

9. Centrifugal PumpsH. H. AndersonThe trade & technical press limited1980



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10. Standard Handbook of Machine DesignJ. E. Shigley - C. R. MishkeMac Graw Hill1986

11. Centrifugal Pump SourcebookJ. W. Dufour - W. E. NelsonMac Graw Hill1992

12. Comportement dynamique des labyrinthes dans les machines hydrauliquesA. Verry - P. GuitonLa houille blanche1986

13. Hydrodynamics of PumpsChristopher E. BrennenOxford University Press1994

14. Lubrification hydrodynamique. Paliers et butées.J. Frêne - D. Nicolas - B. Degueurce - D. Berthe - M. GodetEyrolles1990

15. Turbomachinery RotordynamicsPhenomena, Modeling, and AnalysisDara ChildsJohn Wiley & Sons1993



7-11

Figure 7-1No Contact Seals



7-12

Figure 7-2Schematic Diagram of an Interstice Joint

Figure 7-3Schematic Diagram of a Grooved Labyrinth-Type Joint



7-13

Figure 7-4Types of Interstice Joints



7-14

Figure 7-5Types of Interstice and/or Labyrinth-Type Joints



7-15

Figure 7-6Schematic Diagram of a Multi-Cellular Pump



7-16

Leak Rate (%)Function of Rotating Speed

amm

bmm

1400 1700 2000 2500

1 0.30 28 1.52 1.80 2.00 2.18

2 0.30 28 2.85 3.32 3.52 3.70

3 0.30 17.5 3.52 4.03 4.33 4.50

4 0.43 17.5 6.06 6.65 6.70 6.70

5 0.51 17.5 7.92 8.62 8.86 8.60

6 0.74 17.5 13.2 13.9 14.0 14.0

7 0.99 17.5 18.7 19.6 19.8 20.0

8 0.4. 17.5 4.83 5.38 5.58 5.52

9 0.74 17.5 12.7 13.5 13.7 13.6

10 0.28 17.5 3.18 3.68 3.94 4.08

11 0.53 17.5 8.53 9.04 9.15 9.19

12 0.28 17.5 2.52 2.88 2.92 2.98

13 Spiral Groove1.5 • 1.5 mm

0.53 17.5 6.24 6.68 6.89 6.82

14 0.25 17.5 2.55 3.03 3.28 3.44

15 0.25 17.5 2.07 2.34 2.45 2.52

Leak rate as a percentage of full flow and for different rotating speeds;pump diameter 76 mm, n3 = 21.2, D2 = 257 mm, annular diameter = 105 mm

Figure 7-7Table of leakage rates as a function of geometrical shapes and rotation speeds [7]



7-17

Figure 7-8Evolution of Leakage Rate as a Function of the Pressure Differential [9]



7-18

Figure 7-9Evolution of the Dynamic Coefficients as a Function of the Flow Rate [13]



7-19

Figure 7-10Influence of Eccentricity on Dynamic Coefficients [15]

bearing technology topics tr-113059-v2[1]

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