Feedwater Pump Maintenance Guide

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ABSTRACT

"...although there have been a num ber of specific problems that can be attributed to a specific supplier's pump design or design feature, the major factors that contribute to reduced feed pump reliability are under the direct control of the utility engineers and operators." EPRI Report NP-1571, Causes of Repetitive Feedwater Pump Failures, October 1980. The data derived for this report confirms that the same statement is applicable today, fifteen years later (Reference 3).

The NMAC Main Feedwater Pump Maintenance Guide is the latest EPRI sponsored effort directed toward boiler feed pumps in the fossil and nuclear power generation application areas. For two decades, EPRI has contributed research and development direction in the attempt to improve the awareness of the industry to the critical importance of the boiler feed pump to the reliability of the power generation facility. The failure of a boiler feed pump has economic ramifications in the cost of a forced outage or a reduction in generated power. Conclusions from available data sources, including NPRDS and NMAC plant surveys, show that a majority of failure symptoms are leaks from lube oil systems, stuffing box shaft sealing systems, water piping and gaskets. In the NM AC evaluation of all maintenance issues, reactor boiler feedpumps rose in their combined ranking from thirteenth place in 1991, to fifth place m 1992, to third place in 1993. In 1995, pumps still rank in fourth place as a maintenance issue. This guide is directed only to the boiler feed pump applications in nuclear power generation facilities. In boiling water reactor facilities, the boiler feed pumps are utilized as reactor feed pumps. In pressurized water reactor facilities, the boiler feed pumps are utilized as steam generator feed pumps. For purposes of simplicity, in this guide, the boiler feed pump m either type of facility will be referred to as a main feedwater pump.

This guide provides basic information on the design, construction and maintenance of the main feedwater pump equipment supplied by the six pump manufacturers to the domestic United States nuclear power generation industry. This guide is intended to provide useful information to all disciplines and skills associated with the maintenance of main feedwater pump equipment, the planning of its maintenance, and the monitoring and evaluation of its performance. Data basically derived from 1988 onward is utilized to determine failure modes and effects analysis. The use of diagnostics is discussed and its practical application in illustrative case histories is included. Predictive and preventive maintenance programs utilizing new technology are available and affordable today where they were previously unattainable. EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

ACKNOWLEDGMENTS

This guide was prepared for the Nuclear Maintenance Applications Center (NMAC), under Electric Power Research Institute (EPRI) project number 3814-19. Many individuals and organizations provided assistance and information to make this publication possible.

The manufacturers of nuclear main feedwater pumps, BW/IP International, Inc., Demag Delaval Turbomachinery Corp., Ingersoll-Dresser Pump Company, and Sulzer Bingham Pump Company, are to be commended for providing the information necessary to make this guide complete and useful, and for granting permission to reproduce illustrations and use their technical data.

Special recognition is given to Warren H. Brown for his significant efforts in analyzing data, his critical review of the text, and his participation in all activities associated with this guide.

NM AC and the authors of this guide would like to particularly recognize the following members of the Technical Advisory Group for their timely and detailed contributions to this project: Bob Joines Ingersoll-Dresser Pump Company Tim W otring Ingersoll-Dresser Pump Company Joe Silvaggio Demag Delaval Turbomachinery Corp. Don Spencer Sulzer Bingham Pump Company Steve Davis Public Service Electric and Gas, Salem Randy Dvorka ComEd, Quad Cities Mike Turner Carolina Power and Light, Shearon H arris Chris Farsaci Niagara Mohawk Power Corporation, Nine Mile Point Brendan Duffy Northeast Utilities Don Duenkel Wisconsin Electric Power Corporation Jim W ilson Wisconsin Electric Power Corporation John Forsman Northern States Power Corporation, M onticello W illiam Gates Southern Nuclear Operating Company Robert Porter Entergy Operations, Inc. Lloyd Sullivan Entergy Operations, Inc. Donald Osborne Duke Power Company Harry Pales ComEd Amy Monroe South Carolina Electric and Gas, V.C. Summer Gene Baker South Carolina Electric and Gas, V.C. Summer Roger Beal Northeast Utilities

We also acknowledge the following for their valued assistance: W ayne M otley Carolina Seals, Inc. Watson Tomlinson Duke Power Company EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

CONTENTS

1 INTRODUCTION 1-1

1.1 Background ...... 1-1 1.2 The Main Feedwater Pump...... 1-1 1.3 Purpose of the Guide...... 1-1 1.4 Organization of the Guide ...... 1-2 1.5 Identification of Pump Suppliers ...... 1-3 1.6 Categories of Main Feedwater Pumps by Construction...... 1-3 1.7 Prior EPRI Documentation...... 1-5 1.8 Factors Influencing Main Feedwater Pump Reliability ...... 1-5

2 BASIC MAIN FEEDWATER PUMP DESIGN, CONSTRUCTION,

AND APPLICATION 2-1

2.0 Feedwater System...... 2-1 2.1 System Features That Affect Reliability...... 2-1 2.2 Feedwater Auxillary Systems...... 2-4 2.2.1 Minimum Flow Recirculation System...... 2-4 2.2.2 Stuffing Box Seal Injection System...... 2-4 2.2.3 Drain Systems ...... 2-4 2.2.4 Pump Warming System...... 2-5 2.3 History of System Requirements...... 2-5 2.4 Pump Performance and Hydraulic Characteristics...... 2-6 2.5 Radial and Axial Force Balancing...... 2-7 2.6 Pump Case Design Considerations...... 2-8 2.7 Impeller Design ...... 2-8 2.8 Gap A and Gap B Modifications ...... 2-9 2.9 Wear Ring Design ...... 2-12 2.10 Pump Shaft ...... 2-13 2.11 Hydraulic Balancing Devices ...... 2-13 2.12 Bearings...... 2-15 2.13 Journal Bearings...... 2-15 2.14 Thrust Bearings...... 2-18 2.15 Lubrication Systems...... 2-22 2.16 Stuffing Box Sealing Mechanisms...... 2-22

vii EPRI Licensed Material Nuclear Maintenance Applications Center 2.17 Injection Throttle Bushing...... 2-27 2.18 Injection Control Systems ...... 2-28 2.19 Pressure-Controlled Injection System ...... 2-29 2.20 Temperature-Controlled Injection System...... 2-30 2.21 Mechanical Seals ...... 2-30 2.21.1 Design Features...... 2-31 2.21.2 Cartridge Design ...... 2-32 2.21.3 Computer Design of Faces...... 2-32 2.21.4 Stationary Springs...... 2-32 2.21.5 Face Material Selection...... 2-32 2.21.6 Cooling of Faces ...... 2-32 2.21.7 Shaft Centering Plates ...... 2-32 2.22 Mechanical Seal Operating Environment...... 2-33 2.23 Pump Couplings...... 2-33

3. FAILURE MODES AND EFFECTS ANALYSIS 3-1 3.1 Introduction ...... 3-1 3.1.1 Purpose...... 3-1 3.1.2 Methodology...... 3-1 3.2 Failure Data...... 3-1 3.2.1 Feedwater Pump Failure Narratives...... 3-1 3.2.2 NMAC Main Feedwater Pump Maintenance Survey ...... 3-6 3.2.3 Previous EPRI Pump Failure Data ...... 3-8 3.3 Conclusions...... 3-12 3.3.1 Reconciliation of EPRI Data...... 3-12 3.3.2 Summary of the Reconciliation...... 3-12 3.3.3 Topics of Main Feedwater Pump Maintenanc Guide...... 3-12

4. PREDICTIVE AND PREVENTIVE MAINTERNANCE 4-1 4.0 Predictive and Preventive Maintenance ...... 4-1 4.1 Introduction ...... 4-1 4.2 Condition Monitoring...... 4-3 4.2.1 Parameters To Be Monitored ...... 4-3 4.2.2 Data Collection...... 4-6 4.2.3 Data Storage ...... 4-8 4.2.4 Data Processing...... 4-8 4.3 Predictive Maintenance ...... 4-11 4.3.1 Parameter Trending ...... 4-12 4.3.2 Limits for Monitored Parameters ...... 4-13 4.3.3 Interpretation of Monitored Parameters...... 4-14 4.3.4 Limitations of Predictive Maintenance...... 4-19 4.4 Preventive Maintenance...... 4-20 4.4.1 Major Maintenance...... 4-21 4.5 Diagnostics and Root-Cause Analysis...... 4-22 4.5.1 Case Histories...... 4-23 viii EPRI Licensed Material Main Feedwater Pump Maintenance Guide

5 MAIN FEEDWATER PUMP MAINTENANCE PRACTICES 5-1 5.1 Introduction ...... 5-1 5.2 Maintenance Programs...... 5-1 5.2.1 Experience and Training...... 5-1 5.2.2 Procedures and Worksheet ...... 5-1 5.2.3 Predictive/Corrective Maintenance Interface...... 5-6 5.2.4 Spare Parts Inventory ...... 5-6 5.2.5 Vendor Technical Manuals ...... 5-7 5.2.6 Feedwater System Cleanliness ...... 5-7 5.2.7 Preventive Maintenance Schedules ...... 5-8 5.2.8 On-Line Maintenance...... 5-8 5.2.9 Lubrication System Contaminants...... 5-8 5.3 Disassembly and Recording of As-Found Data ...... 5-8 5.3.1 Alignment...... 5-9 5.3.2 Coupling Disassembly, Removal, and Inspection...... 5-11 5.3.3 End Play Measurement and Thrust Bearing Inspection and Removal...... 5-13 5.3.4 As-Found Rotor Positions...... 5-15 5.3.5 Auxiliary Piping ...... 5-16 5.3.6 Radial Bearings...... 5-16 5.3.7 Shaft Oil Seals ...... 5-20 5.3.8 Injection-Type Seals ...... 5-21 5.3.9 Mechanical Seals...... 5-22 5.3.10 Hydraulic Balancing Devices ...... 5-23 5.3.11 Rotor Removal and Inspection ...... 5-23 5.3.12 Casing and Casing Ring Inspection ...... 5-29 5.4 Pump Reassembly and As-Built Data...... 5-30 5.4.1 Casing Ring Installation...... 5-30 5.4.2 Rotor Reassembly and Balancing ...... 5-30 5.4.3 Rotor and End Cover Installation...... 5-33 5.4.4 Radial Rotor Position or Center...... 5-34 5.4.5 Running Position (Axial Center) ...... 5-36 5.4.6 Mechanical Seal Installation ...... 5-37 5.5 Alignment...... 5-40 5.5.1 Cold Alignment ...... 5-40 5.5.2 Hot Alignment ...... 5-40 5.6 Lube Oil System Upgrades ...... 5-41 5.6.1 Introduction ...... 5-41 5.6.2 Stainless Steel Piping ...... 5-41 5.6.3 Piping Design...... 5-43 5.6.4 Piping Inspection ...... 5-44 5.6.5 Hydrostatic Test ...... 5-44 5.6.6 Lube Oil Pumps ...... 5-44 5.6.7 Lube Oil Coolers ...... 5-45 5.6.8 Lube Oil Piping Flanges ...... 5-46 5.7 Pressure Retaining Joints...... 5-46 5.7.1 Main Casing Joints ...... 5-46 5.7.2 Flanged Piping Joints ...... 5-49 5.8 Mechanical Face Seals...... 5-51 5.8.1 Mechanical seal Upgrades ...... 5.51 5.9 Nozzle Loads ...... 5-51

ix EPRI Licensed Material Nuclear Maintenance Applications Center

5.10 Pump Startup...... 5-52 5.11 Pump Warming...... 5-53

APPENDIX A TYPICAL SECTIONAL DRAWINGS OF EACH MAIN FEEDWATER PUMP TYPE ...... A-1

APPENDIX B USER INSTALLATION LISTS...... B-1

APPENDIX C TYPICAL STUFFING BOX DRAWINGS...... C-1

APPENDIX D TYPICAL BEARING CONSTRUCTION DRAWINGS ...... D-1

APPENDIX E FEEDWATER SYSTEM FLOW DIAGRAMS ...... E-1

APPENDIX F TYPICAL OIL LUBRICATION PIPING SCHEMATICS ...... F-1

APPENDIX G SYMPTOMS OF FEEDWATER PUMP PROBLEMS AND

RECOMMENDATIONS FOR SOLUTIONS...... G-1

APPENDIX H RECONCILIATION OF NPRDS/EPRI MAIN FEEDWATER

PUMPFAILURE DATA ...... H-1

APPENDIX I NMAC BOLTED JOINT MAINTENANCE AND APPLICATION

GUIDE (CHAPTER 3)...... I-1

APPENDIX J METHODS FOR OBTAINING THE CORRECT RUNNING POSITION

OF THE MAIN FEEDWATER ROTORS...... J-1

REFERENCES/BIBLIOGRAPHY ...... BIB-1

x EPRI Licensed Material Main Feedwater Pump Maintenance Guide

LIST OF FIGURES

SECTION 2

Figure 2-1 Flow-Straightening Valve...... 2-3

Figure 2-2 Gap A, Gap B, and Overlap Modifications...... 2-10

Figure 2-3 Secondary Flow Pattern In and Around a Pump Impeller at Off- 2-11 Design Flow Conditions ......

Figure 2-4 Balance Disk...... 2-14

Figure 2-5 Balance Drum...... 2-15

Figure 2-6 Types of Journal Bearings...... 2-16

Figure 2-7 Typical Radial Bearing Bracket Assembly, Tilting-Pad Radial Bearing .. 2-17

Figure 2-8 Typical Insert-Type Sleeve Bearing ...... 2-18

Figure 2-9 Kingsbury-Type Thrust Bearing ...... 2-19

Figure 2-10 Typical Kingsbury Thrust Bearing Assembly...... 2-20

Figure 2-11 Typical Bearing Construction, Outboard End Thrust and Journal 2-21 Bearing Bracket Assembly......

Figure 2-12 Typical Stuffing Box Configuration...... 2-24

Figure 2-13 Typical Stuffing Box Detail, Floating Ring...... 2-25

Figure 2-14 Types of Labyrinth Injection Throttle Bushing...... 2-26

Figure 2-15 Typical Stuffing Box Detail, Injection Throttle Bushings 2-27

Figure 2-16 Typical Intermediate Leakoff Shaft Seal System ...... 2-29

Figure 2-17 Drain Temperature Control System Regulating Condensate Injection 2-30 to Feed Pump Shaft Seals......

Figure 2-18 Typical Cartridge - Type Single Mechanical Seal...... 2-31

Figure 2-19 Diaphragm Coupling ...... 2-35

Figure 2-20 Flexible Disk (Disk Pack) Coupling ...... 2-36

xi EPRI Licensed Material

Nuclear Maintenance Applications Center

Figure 2-21 Gear Coupling...... 2-37 Figure 2-22 Continuous Lubricated Gear Coupling ...... 2-38 Figure 2-23 Continuous Lube Coupling Cover ...... 2-39 SECTION 4 Figure 4-1 Relationship Among Maintenance Terms...... 4-2 Figure 4-2 Pump Shaft Vibration (Outboard)...... 4-17 Figure 4-3 Pump Shaft Vibration (Inboard)...... 4-18 Figure 4-4 Cascade Plot of Pump Shaft Radial Vibration (Inboard) ...... 4-19 Section 5 Figure 5-1 Reverse Dial Indicator Alignment ...... 5-10 Figure 5-2 Laser Alignment System ...... 5-10 Figure 5-3 Spherical Seat Bearing ...... 5-19 Figure 5-4 Stationary-Type Oil Seal ...... 5-20 Figure 5-5 Bearing Isolator User for Main Feedwater Pump Bearing Housings ...... 5-21 Figure 5-6 Shaft Runout and Dimensions Worksheet ...... 5-25 Figure 5-7 Double-Suction Impeller ...... 5-26 Figure 5-8 Rotor Balancing Worksheet ...... 5-32 Figure 5-9 Typical Cartridge-Type Single Mechanical Seal ...... 5-38 Figure 5-10 Potential Casing Deformation During Warming...... 5-54 APPENDIX A Figure A-1 Bingham Type CD...... A-3 Figure A-2 Bingham Type CPD...... A-4 Figure A-3 Bingham Type HSB...... A-5 Figure A-4 Bingham Type MSB ...... A-6 Figure A-5 Bingham Type MSD ...... A-7 Figure A-6 Byron Jackson Type DVSR...... A-8 Figure A-7 Byron Jackson Type HDR ...... A-9

xii EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Figure A-8 Byron Jackson Type HSB...... A-10 Figure A-9 Byron Jackson Type DVS...... A-11 Figure A-10 Byron Jackson Type DBMX, Single-Suction First Stage...... A-12 Figure A-11 Byron Jackson Type DVMX, Double-Suction First Stage ...... A-13 Figure A-12 Delaval Type 1BSX...... A-14 Figure A-13 Delaval Type 1BSX...... A-15 Figure A-14 Delaval Type 2BSX...... A-16 Figure A-15 Delaval Type 4BSX...... A-17 Figure A-16 Ingersoll-Rand Type 16x17 CN...... A-18 Figure A-17 Ingersoll-Rand Type 18x17 CN...... A-19 Figure A-18 Ingersoll-Rand Type C...... A-20 Figure A-19 Ingersoll-Rand Type CA...... A-21 Figure A-20 Ingersoll-Rand Type JT...... A-22 Figure A-21 Pacific Type HVF, Double Cover ...... A-23 Figure A-22 Pacific Type HVCN, Single Cover...... A-24 Figure A-23 Pacific Type BFI...... A-25 Figure A-24 Pacific Type SFI...... A-26 Figure A-25 Pacific Type RHC ...... A-27 Figur A-26 Worthington Type WGID ...... A-28 Figure A-27 Worthington Type WNC, Single Stage, Single Suction ...... A-29 Figure A-28 Worthington Type WNCD, Single Stage, Double Suction ...... A-30 APPENDIX C Figure C-1 Stuffing Box, Floating Ring Construction ...... C-2 Figure C-2 Stuffing Box, Injection Throttle Bushing Construction, Rotating Member Serrations ...... C-3 Figure C-3 Stuffing Box, Injection Throttle Bushing Construction, Rotating and Stationary Member Serrations...... C-4 Figure C-4 Typical Mechanical Seal ...... C-5

xiii EPRI Licensed Material Nuclear Maintenance Applications Center

APPENDIX D Figure D-1 Typical Bearing Construction Thrust Bearing ...... D-2 Figure D-2 Typical Bearing Construction Fixed Radial Bearing ...... D-3 Figure D-3 Typical Bearing Construction Self-Aligning Radial Bearing...... D-4 APPENDIX E Figure E-1 Baboock & Wilcox Pressurized Water Reactor ...... E-2 Figure E-2 Combustion Engineering Pressurized Water Reactor...... E-3 Figure E-3 General Electric Boiling Water Reactor...... E-4 Figure E-4 Westinghouse Pressurized Water Reactor ...... E-5 APPENDIX F Figure F-1 Motor Gear Pump Lubrication Schematic...... F-2 Figure F-2 Turbine Pump lubrication Schematic...... F-3 APPENDIX J Figure J-1 Thrust Collar Utilizing a Spacer...... J-1 Figure J-2 Plan View of Trust Housing...... J-3 Figure J-3 Plan View with Inner Thrust bearing Installed...... J-4 Figure J-4 Shim Type Rotor Centering...... J-5

xiv EPRI Licensed Material Main Feedwater Pump Maintenance Guide

LIST OF TABLES

SECTION 1

Table 1-1 Category 1: Radially Split Outer Case Construction with Single-Stage Double-Suction Impeller ...... 1-4 Table 1-2 Category 2: Radially Split Double Outer Case Construction Containing One or More Impellers with First Stage as a Single or Double Suction ...... 1-4 Table 1-3 Category 3: Axially Split Outer Case Construction with a Single-Stage Double Suction Impeller., or a Multistage Pump with either a Single- Suction or Double-Suction First-Stage Impeller...... 1-5 SECTION 2

Table 2-1 Functional and Design Parameters for Nuclear Feed Pumps...... 2-6 SECTION 3

Table 3-1 NPRDS Reports Failures Each Year ...... 3-3 Table 3-2 NPRDS Reports Failure Symptoms...... 3-3 Table 3-3 NPRDS Reports Failed Components...... 3-4 Table 3-4 NPRDS Reports Failure Causes...... 3-4 Table 3-5 Component Cross-Reference by Sections ...... 3-13 SECTION 4

Table 4-1 Acceptable Life Expectancy of Pumps ...... 4-21

SECTION 5

Table 5-1 Pump Worksheets...... 5-3

Table 5-2 Coupling Fits...... 5-13

APPENDIX B

Table B-1 Category 1: Main Feedwater Pump Installations Radially Split Single Outer Case with Single Stage Double-Suction Impeller...... B-2

Table B-2 Category 2: Main Feedwater Pump Installations Radially Split Double Outer Case containig One or More Impellers with First Stage as a Single-or Double-Suction Impeller...... B-7

xv EPRI Licensed Material Nuclear Maintenance Applications Center

Table B-3 Category 3: Main Feedwater Pump Installations Axial Split Outer with a Single Double-Suction Impeller or a Multistage Pump with either a Single Suction or Double Suction First Stage B-10 Impeller......

Table B-4 Plant Listing of Main Feedwater Pump Installations ...... B-12

Table B-5 Bingham Supplied Main Feedwater Pump Installation List...... B-18

Table B-6 Byron Jackson Supplied Main Feedwater Pump Installation List...... B-20

Table B-7 Delaval Supplied Main Feedwater Pump Installation List ...... B-24

Table B-8 Ingersoll-Rand Supplied Main Feedwater Pump Installation List...... B-26

Table B-9 Pacific Supplied Main Feedwater Pump Installation List...... B-27

Table B-10 Worthington Supplied Main Feedwater Pump Installation List...... B-28

xvi EPRI Licensed Material Main Feedwater Pump Maintenance Guide 1

INTRODUCTION

1.1 Background

For two decades, the Electric Power Research Institute (EPRI) has been involved in improving the reliability of boiler feed pumps in the power generation industry. This has included basic research and development based on determining the root cause of failures to fossil and nuclear boiler feed pumps. The Main Feedwater Pump Maintenance Guide, developed by the Nuclear Maintenance Applications Center (NM AC), continues by providing inform ation to the industry that can be of help to im prove the reliability of boiler feed pumps in nuclear power utilities.

A boiler feed pump failure m a nuclear utility has serious ramifications. Results can be a forced outage or a reduction in power generation capacity, both of which have adverse financial impact on the utility. NMAC's 1995 maintenance issues survey reveals increasing concern among utilities for the reliability of nuclear boiler feed pumps. In evaluating a nuclear boiler feed pump as a maintenance issue, the composite ranking of nuclear boiler feed pumps rose from 13th place in 1991, to 5th place in 1992, to 3rd place in 1993. This guide provides an analysis of failures that have been experienced before, during, and after those years.

1.2 The Main Feedwater Pump

In a boiling water reactor, the boiler feed pump supplies water to the reactor and is known as the reactor feed pump. In a pressurized water reactor (PW R), the boiler feed pump supplies water to the steam generator and is known as the steam generator feed pump. For this guide, no distinction is made to the service application by nuclear system type. The guide refers to either a reactor feed pump or a steam generator feed pump as a main feedwater pump. The equipment used for each of these services can be of the same generic design. Six manufacturers have supplied all the pump equipment for this application in the U.S.

1.3 Purpose of the Guide

This guide is intended to provide useful information to personnel of all disciplines and skills associated with maintenance of main feedwater pump equipment, planning their maintenance, and monitoring and evaluating their performance. The guide can provide new insight to experienced personnel and basic information, guidance, and instruction to personnel recently assigned responsibility for main feedwater pump maintenance. Existing procedures can be reevaluated. Technology advances in data acquisition and

1-1 EPRI Licensed Material Nuclear Maintenance Applications Center analysis afford the possibility of solving long-term problems that have resisted prior root-cause failure identification and resolution.

1.4 Organization of the Guide

This guide consists of five sections with supporting appendices. Section 1 introduces the purpose and content of the guide by providing basic information on the design, construction, and maintenance of nuclear main feed water pumps. The pumps have been categorized into three basic mechanical and hydraulic configurations that include all existing operational installations.

Section 2 discusses the systems in which the main feedwater pumps are required to operate and the design and construction of the pumps themselves. Design details of constituent components, a comparison of the different hydraulic and mechanical design philosophies used, the choices available in stuffing box sealing, thrust and radial bearing constructions, and coupling considerations are discussed. The prime movers and their proper alignment to the main feedwater pumps are considered.

Section 3 contains available failure data, and analyzes its content to direct the maintenance practices and procedures to follow or implement to eliminate or mitigate the reported failures of today from being repeated tomorrow. Primary data sources were the Nuclear Plant Reliability Data System (NPRDS) database, NMAC Main Feedwater Pump Mainte- nance Surveys, and previous EPRI pump failure surveys and reports. This information was complemented by discussions with responsible utility maintenance and engineering personnel, and with additional input derived from the Operating Plant Experience Code (OPEC). Primary ongoing maintenance problems include lubrication oil systems, shaft sealing systems (especially mechanical seals), water piping, and gaskets.

Section 4 discusses condition monitoring as the systematic collection, storage, and processing of machine operating parameters. The parameters to be monitored and measured are itemized. The frequency of data acquisition measurements is listed by parameter.

Section 5 provides guidance as to what should be expected and what should be acceptable for the maintenance of main feedwater pumps. Practical advice to assist the maintenance effort is offered. Discussion is included concerning stuffing box configurations, nozzle loads, gaskets and sealants, alignment, pump startup, and pump warming.

Appendix A offers typical sectional drawings of each pump type, provided by the pump vendors.

Appendix B provides multiple user installation lists by plant, manufacturer, and pump category.

Appendix C provides typical stuffing box construction drawings for each of the three alternative constructions—floating ring, injection throttle bushing, and mechanical seal.

Appendix D contains typical bearing construction drawings for Kingsbury-type thrust bearings and for fixed and self-aligning radial bearings.

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Appendix E provides basic typical feedwater system flow diagrams for Babcock and Wilcox, Combustion Engineering, General Electric, and Westinghouse nuclear steam supply system s.

Appendix F offeres typical forced-feed lubrication system drawings for turbine driven and motor gear-driven systems.

Appendix G includes symptoms of main feedwater pump problems and recommendations.

Appendix H contains NPRDS/EPRI main feedwater pump failure data.

Appendix I presents a section reproduced from the NMAC Bolted joint Maintenance and Application Guide (Reference 16).

Appendix J presents methods for obtaining the correct running position of main feedw ater pump rotors.

1.5 Identification of Pump Suppliers

Because the last of the domestic nuclear main feedwater pumps was purchased about 20 years ago, there has been an extensive rationalization in the pump manufacturing industry due primarily to manufacturing over capacity. Every one of the original six manufacturers has gone through a reorganization, merger, direct purchase, or lever- aged buyout, sometimes more than once. The pumps are no longer properly identifiable to the original organization that manufactured them. Instead of attempting to trace the historical name to its present-day identity, this report identifies each supplier by its manufacturer's common name when the equipment was supplied, which is likely to be the same name used by the facilities to identify the original pump manufacturer.

In this report the main feedwater pumps are identified as being supplied by Bingham, Byron Jackson, Delaval, Ingersoll-Rand, Pacific, and Worthington m this manner.

1.6 Categories of Main Feedwater Pumps by Construction

Three generic pump construction categories were determined to include all the existing pump designs of all the vendors. Discussion will be presented for each of the three generic categories without respect to the supplier, except as deemed appropriate to properly describe a specific construction feature. The three generic construction categories Area • Radially split single other case construction containing a single-stage double-suction impeller (see Table 1-1) • Radially split double outer case construction containing one or more impellers with the first-stage impeller being double or single suction (see Table 1-2) • Axially split outer case construction as a single-stage pump with a single or double- suction impeller or a multiple-stage pump with either a single- or double-suction first- stage impeller (see Table 1-3)

1-3 EPRI Licensed Material Nuclear Maintenance Applications Center Table 1-1 Category 1: Radially Split Single Outer Case Construction with Single-Stage Double-Suction Impeller

Pump Type Manufacturer Construction Features Reference Figure (Appendix A) CD Bingham Volute Type A-1 HSB Bingham Volute Type A-3 DVSR Byron Jackson Volute Type A-6 1BSX Delaval Diffuser/Volute Type A-12, A-13 CN Ingersoll-Rand Diffuser Type A-16, A-1 7 JT Ingersoll-Rand Diffuser Type A-20 HVF Pacific Volute Type A-21 HVCN Pacific Volute Type A-22 WGID Worthington Volute Type A-26 WNGD Worthington Diffuser Type A-28

Table 1-2 Category 2: Radially Split Double Outer Case Construction Containing One or More Impellers with First Stage as a Single or Double Suction

Pump T ype ManufacturerC onstruction Features Reference Figure (Appendix A) CPD Bingham Multistage Volute A-2 HDR Byron Jackson Single-Stage Volute A-7 HSB Byron Jackson Multistage Volute A-8 2BSX Delaval Multistage Diffuser A-14 4BSX Delaval Multistage Diffuser A-15 C Ingersoll-Rand Multistage Diffuser A-18 CA Ingersoll-Rand Multistage Diffuser A-19 BFI Pacific Multistage Diffuser A-23 WNC Worthington Multistage Diffuser A-27

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Table 1-3 Category 3: Axially Split O uter C ase Construction with a Single-Stage Double- Suction Impeller, or a Multistage Pump with either a Single-Suction or Double-Suction First-Stage Impeller

Pump T ype Manufacturer Construction Features Reference Figure (A ppendix A) MSB Bingham Multistage Volute A-4 MSD Bingham Multistage Volute A-5 DVMX Byron Jackson Multistage Volute A-10, A-11 DVS Byron Jackson Single-Stage Volute A-9 RHCNDS Pacific Multistage Volute A-25

1.7 Prior EPRI Documentation Prior EPRI project documentation has furnished extensive information on the design, construction, and application of boiler feedwater pumps in fossil and nuclear facilities. Specifically, these resources provide information on the feedwater to the boiler, reactor, or steam generator as it applies to the type of steam supply system. The following documents are singled out for attention: • EPRI FP-754, Survey of Feed Pump Outages (Reference 1) • EPRI NP-1571, Evaluation of Basic Causes of Repetitive Failures of Nuclear and Fossil Feedwater Pumps, October 1980 (Reference 3) • EPRI TR-102102, Feedpump Operation and Design Guidelines-Summary Report, June 1993 (Reference 15) • EPRI TR-104292-V1, Boiler Feedpump Operation and Maintenance Guidelines, September 1994 (Reference 17)

1.8 Factors Influencing Main Feedwater Pump Reliability

The reliability of a main feedwater pump is influenced by many factors. The basic design of the nuclear steam supply system, the feedwater system, the feedwater auxiliary systems, and the design of the main feedwater pump itself with its driver are all crucial to a successful installation. The level of expertise of site personnel in operating and maintaining the system and equipment must be sufficient regardless of the quality of the systems and available design. On any installation, the reliability of the system and equipment should improve as operations and maintenance personnel become more task-proficient and procedures are refined.

As personnel are reassigned, there might be an increase in failure rates that necessitates reevaluating the skill levels of responsible personnel and operational practices. Changes to the operating point of the pump on its performance curve and in the warmup proce-

1-5 EPRI Licensed Material Nuclear Maintenance Applications Center dures are examples of changes that can adversely influence the reliability of a pump that has been properly maintained yet degrades or fails. Proper operation and maintenance must always be attained to expect reliability from main feedwater pumps. This guide assists in advancing user knowledge and expertise with regard to these important critical pieces of equipment m the feedwater system.

1-6 EPRI Licensed Material Main Feedwater Pump Maintenance Guide 2 - .. BASIC MAIN FEEDWATER PUMP DESIGN, CONSTRUCTION, AND APPLICATION

2.0 Feedwater System

The feedwater system m a nuclear power generation facility provides feed water to the reactor in a boiling water reactor (BW R) system. The feedwater system also provides feedwater to the steam generator in a PWR system. The feedwater is provided at the required pressure and temperature to handle all anticipated normal and transient conditions including startup and power ascension. The main feedwater pump is the motive force that provides the required flow and pressure. Failure or degradation of the . main feedwater pump results in a direct loss or downgrading of the facility's power generation capability. A feedwater system flow diagram for each of the nuclear steam supply system vendors—Babcock and Wilcox, Combustion Engineering, General Elec- tric, and Westinghouse—is furnished in Appendix E.

2.1 System Features That Affect Reliability

Past EPRI studies have concluded that overall design of the feed water system is an important factor in feed pump reliability. The system features that have a major affect on feed pump reliability have been identified as: • Features that influence the normal and transient environment at the pump suction, i.e., pressure, flow, and temperature • Feed pump protective alarm/trip systems • Availability of an installed spare feed pump • Details of the pump auxiliary systems

The most important feed system feature that affects pump performance and reliability is the amount of net positive suction head (NPSH) available at the pump suction. A pump requires an amount of pressure above the vapor point of the pumped fluid to prevent and preclude flashing and cavitation in the pump suction. Cavitation can result in vibration of the pump and its various components and erosion of impeller surfaces, all of which result in reduced reliability and compromised pump life.

The value of NPSH required by the pump is determined by model and full-scale pump tests. Historically, testing determined an accepted value of required NPSH when the developed head degraded 1-3%, depending on the test method used. Recent research by manufacturers demonstrates that this old criteria underestimates the existence of cavitation and its destructive influences. A pump can exhibit no degradation of developed head

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and still be experiencing extreme cavitation with potential damaging effects. A number of nuclear feed pump installations with cavitation problems used flow visualization tests to develop redesigned impellers to achieve minimum cavitation activity and maximum impeller life. Should a NPSH problem exist, the only viable economic approach to correct the problem might be lowering the NPSH requirement of the impeller through redesign rather than adding a booster pump or implementing an expensive system change.

System piping configurations can distort the flow entry into the pump suction. System conditions that reduce suction pressure or increase suction temperature reduce the amount of NPSH available to the pump. Non-planar elbows in proximity to the pump suction can create distorted flow fields. Any of these conditions can have adverse results on pump performance and life. A recent invention, the Cheng Rotation Vane (CRV), introduces a prerotation upstream of non-planar elbows where an insufficient run of straight pipe exists before an elbow or pump suction (see Figure 2-1). This device has been successful in correcting vibration and performance problems associated with poor suction piping configurations.

2-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Figure 2-1 Flow-straightening Vane (Courtesy of C heng Fluid System s, Inc.)

2-3 EPRI Licensed Material Nuclear Maintenance Applications Center Prior reports indicated that an inordinate number of pump protective alarms and reactor trips resulted in pump degradation. Most incidents were found to be the result of excessively conservative alarm limits. Presently, with reappraisal of procedures and limits, this does not appear to be a problem.

Installed spare main feedwater pump capacity has proven itself economically viable. Without installed spare pump capacity to replace a main feedwater pump that requires shutdown, the power generation unit is either subjected to a forced outage or to a reduced level of power generation, both of which have adverse economic impact. Installing spare pump capacity in a mature installation could prove economically unjustifiable.

2.2 Feedwater Auxiliary Systems Main feedwater pump auxiliary systems have not received the attention required to optim ize pum p reliability. Data in this report identifies an inordinate num ber of auxiliary system problems as being responsible for pump failures.

2.2.1 Minimum Flow Recirculation System

Undersized recirculation systems have been troublesome from initial installation startup at numerous locations. Unfortunately, manufacturers might have been influenced by the desire to offer a more competitive marketing proposal by quoting minimum recirculation flow requirements at low flow values. In doing this, they proposed a less costly, smaller sized recirculation system to be evaluated, which in turn created operational problems with the recirculation systems, especially with extreme vibrations. In retrospect, the architect-engineers should have specified minimum acceptable values of recirculation flows. Today, the pump manufacturers have a more complete understanding of pump performance at low flow conditions to the left of best efficiency on the pump performance curve. The recommended values of recirculation flow would probably be higher today for the same pump performance curve. Complete replacement of recirculation systems has been necessary at several plants due to problems caused by initial undersizing of the system.

2.2.2 Stuffing Box Seal Injection System

The inability of seal injection systems to supply the proper pressure and flow to the pump stuffing boxes under all operating circumstances is still a factor at some sites. This applies whether the stuffing box has floating rings, injection type throttle bushings, or mechanical seals.

2.2.3 Drain Systems

Drainoffs for oil return and injection throttle bushing drain systems that have been undersized create problems. Data in this report indicates that utilities have not sufficiently maintained their oil lubrication systems, resulting m repeated shutdowns due to oil piping and gasket leaks.

2-4 EPRI Licensed Material Main Feedwater Pump Maintenance Guide 2.2.4 Pump Warming System

Utilities have not always correctly and successfully installed warmup systems, resulting in ineffective warming or detrimental thermal distortions being introduced into the pumps.

Problems with pump foundations are no longer reported as problems.

2.3 History of System Requirements

The pressure and flow requirements of all nuclear steam supply systems in the U.S. dictate that horizontal centrifugal pumps are used. During the inception of nuclear power generation in the U.S. plant design, megawatt production of a facility was low. The hydraulic requirements in Lose installations resulted in motor-driven multistage pumps for the main feedwater pump function. Specification requirements concerning pump design, metallurgical requirements in those installations, nondestructive examination, and testing were minimal. As the size of generation units increased with matu- ration of the nuclear power generation industry, the type and configuration of the main feedwater pump also changed. When the purchase of new nuclear units ceased in the mid-1970s, the prevalent pump configuration was a single-stage double-suction high- speed turbine-driven pump with injection throttle bushings. The pump industry had considerable capability and capacity to supply the requirement for all four domestic nuclear steam supply vendors of both BWR and PWR configurations. The hydraulic requirements for the BWR versus PWR systems vary but overlap, such that the same physical design of a pump could service both BWR and PWR systems.

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T able 2-1 Functional and Design Parameters for Nuclear Feed Pum ps

PARAMETER BWR Typical PWR Typical RangeValue (1) Range Value (2) Design total dynamic head (TDH )-(F T) 2,200-2,800 2,800 1,700-2,400 2,070 D esign flow (G PM ) 5,000-12,270 11,200 7,000-19,800 17,400 Design suction pressure (PSIA) 170-600 450 150-709 267 Design discharge pressure (PSIA) 1,205-1,5601,560 990-1,338 1058 Design suction temperature (F) 214-378 300320-390 366 Design speed (RPM) 3,600-7,000 5500 3,000-6,800 4,920 Size (HP) 4,000-10,000 8,450 3,500-11,500 10,000 Notes: (1) Typical values for BWR feed pumps based on modern 1,000 MWe plant with three 33% pumps (2) Typical values for PWR feed pumps based on modern 1,000 MWe plant with two 50% pumps.

Table 2-1 is reprinted from EPRI report NP-1571 (Reference 3), Evaluation of Basic Causes of Repetitive Failures of Nuclear and Fossil Feedwater Pumps, to provide parametric information on comparing BWR and PWR hydraulic requirements. This report states that:

A BWR feed pump tends to have a 50% greater total dynamic head (TDH), a lower section temperature, and a higher suction pressure than a comparable PWR installation. The total dynamic head is higher in the BWR application because the pressure in a reactor (where the steam is generated) is typically several hundred pounds per square inch greater than the pressure in the secondary side of a PWR steam generator. The flow per pump is lower because the BWR example selected has three 33% flow pumps. (ED-. a common situation) The total feedwater flow is the same for BWRs and PWRs of the same power rating. The lower suction temperature, coupled with a higher suction pressure, means that, for the most part, BWR reactor feed pumps have a greater operating margin in regard to NPSH; therefore, if all other circumstances are equal, there should be fewer outages attributable to trips in BWRs than to PWRs.

2.4 Pump Performance and Hydraulic Characteristics

A typical TDH head versus flow curve should exhibit a constantly rising curve from the design point to shutoff. The rise to shutoff should approximate 25% to ensure total stability of performance, especially at low flow. Pump manufacturers tended not to test and report on the actual curve characteristic on the left side of the curve. Unfortunately, this is where most instabilities appear. Recent EPRI-sponsored research on boiler feed design has provided considerable detail on feedwater pump performance. Design considerations resulting in revised performance characteristics have been developed.

Pump manufacturers allowed minimum flows at far less than would be recommended today. There is currently a better understanding of cavitation. Incipient cavitation occurs

2-6 EPRI Licensed Material Main Feedwater Pump Maintenance Guide at much higher flow rates than previously anticipated. These research developments will provide greater reliability to any future pumps in this or other boiler feed services. Mini- mum flow recommendations will be higher, and that will mitigate unstable pump performance and the development of detrimental vibrations from low flow operation.

W hen purchased, main feedwater pumps were quoted with efficiencies in the range of 85-89%. Competition drove quotes of higher efficiencies to gain an advantage in the evaluation of bids. The most recent EPRI boiler feed research concludes that pump design modifications to achieve higher efficiencies worked but affected the reliability of the equipment. The recommendation today is to specify pump design features that would ensure reliability over efficiency.

2.5 Radial and Axial Force Balancing

There are two prevalent design philosophies in the basic construction of the main feedwater pumps: the use of (1) a diffuser design or (2) a volute design. The choice dic- tates how radial and axial forces balance In the overall construction of the pump. Consid- erations involve hydraulic, mechanical, and economic factors for pump manufacturers. Diffuser design casing construction, with in-line impellers and a stacked case inner assembly, and volute design casing construction with opposed impellers and a horizontally split inner volute, have both been successfully used in main feedwater pump service.

In the diffuser design pump, the flow of liquid passes to a ring of diffuser vanes in the diffuser construction, resulting in a balance of radial forces completely around the circumference of the diffuser ring. In the volute construction, the flow from the impellers discharges into two takeoffs, or volute lips, located 180 apart, one in each half of the split volute casing. The claim is for com plete radial balance In this design. It is not difficult to machine the diffuser channels with total duplication from one diffuser vane to the next.

The volute design construction starts with a casting, and even though the volute passages are cast from the same core in each of the two halves, there are always minute differences attributable to the casting process that cannot achieve the accuracies of the diffuser-type construction. There are considerable manufacturing cost savings with diffuser-construction. The diametral size of the diffuser is also less in a multistage construction than in a volute construction, as the volute requires greater physical volume than required by the hydraulic construction of the diffuser. The high velocity fluid is converted to pressure in the diffuser passages and in the volute passages.

If the design requires more than one impeller to develop the required head, the impellers are mounted on opposing directions in the volute design, and in the same direction in the diffuser design. Thus, the axial thrust in the diffuser design has the contribution of the thrust developed by each impeller added in the same direction. In the volute design with opposed impellers, the theory is that the thrusts can be balanced one direction against the other to cancel out and provide minimal net inherent axial thrust. The axial thrust developed in the inline impeller diffuser construction is carried by either a balance drum or a balance disk.

2-7 EPRI Licensed Material Nuclear Maintenance Applications Center 2.6 Pump Case Design Considerations

The pump outer case contains the developed pressure of the pump. As the nuclear industry progressed, the configuration of the outer case, its metallurgy and its nondestructive test requirements, changed. This resulted in a higher quality product. Initially, a stainless steel case was required. This requirement was met with an 11-13 chrome cast metallurgy, the best available for service at the time. However, 11-13 chrome metallurgy is difficult to cast and fabricate and is especially resistive to successful repair welding without inducing subsequent cracks. Eventually, all suppliers used newly developed chromium-content alloys, which resulted in a weldable 13-4 cast metallurgy.

The pump nondestructive test requirements escalated dramatically. Early units were often limited to visual and hydrotest nondestructive examination only. In the 1970s, not only were visual and hydrotest examinations required, but various combinations of magnetic particle examination, ultrasonic examination, and radiographic examination became requirements. Manufacturers had to consider the cost implications of producing to these specifications, especially with cast materials. This consideration caused some manufacturers to convert from cast to forged designs to reduce costs. The hydraulic performance could be, and was, the same with either a cast or forged case design. In some cases the pump design went from a cast outer case design to a double-case forged design. Other manufacturers maintained their cast designs and produced them to higher quality levels.

2.7 Impeller Design

The impeller is crucial to the successful performance of main feed water pumps. Many of the early units were multistage pumps containing a number of impellers. Later, as the industry settled on single-stage high-speed units, the quality of the impeller had to improve. The number of impellers was reduced and the entire produced head was developed by a single impeller using large amounts of energy. Mechanical energy is converted by the impeller into the additional velocity head of the fluid. In a double-suction impeller, the feedwater enters both sides of the impeller axially and equally. The flow is discharged into a diffuser or double volute construction, or a combination of both, where the fluid slows and the developed velocity head converts to the pressure head.

The working parts of an impeller are the vanes and shrouds that support the vanes. Both are supported on a hub mounted and attached to the pump shaft. The shape of the vanes and the number of vanes primarily determine pump performance in complex relationships of angles, curvatures, and areas. The relationship of the discharge of flow from the impeller into the diffusers or volute has been the subject of extensive investigation. Field modification of certain geometric relationships, especially those known as Gap A or Gap B modifications, have provided exceptional problem-solving success m many pumps.

2-8 EPRI Licensed Material Main Feedwater Pump Maintenance Guide 2.8 Gap A and Gap B Modifications The most common modifications are those to Gap A and Gap B. Dr. Elemer Makay has championed these modifications for many years and has published considerable information under EPRI auspices and as an independent consultant (Reference 41-46). A1- though Gap A-Gap B information is available from many sources, its many capabilities to attack and correct many problems make it a valuable inclusion m this report.

In high-horsepower, high-pressure pumps such as those used as main feedwater pumps, a significant number of problems have been solvable with modifications to Gap A and/or Gap B. The problems potentially corrected by these modifications follows: • Breakage of volute and/or diffuser lips • Breakage of discharge area of impeller vanes • Breakage of impeller shrouds • Damage to pump coupling • Overheating of thrust bearing • Breakage of stuffing box injection lines and lube oil piping lines • Failure of mechanical seals • Failure of thrust-balancing device • Feedwater piping vibration at reduced load • Failure of the pump shaft • High noise levels at vane passing frequencies and multiples • Unstable head-capacity characteristics • R eversal of thrust loads

Figure 2-2 illustrates Gap A, Gap B. and overlap. Figure 2-3 shows a secondary flow pattern in and around a pump impeller at off-design flow conditions.

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Figure 2-2 Gap A, Gap B. and Overlap Modifications

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Figure 2-3 Secondary Flow Pattern In and Around a Pum p Impeller at Off-Design Flow Conditions

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Gap A is defined as the radial distance between the outside diameter (OD) of the shroud and the case (not the volute lip). A Gap A modification with the proper overlap at the impeller shroud OD minimizes the quantity of pulsating, pressurized flow that recirculates from the discharge of the impeller to the space between the impeller shrouds and the case side walls. The closing down of Gap A can be thought of as intro- ducing a low-frequency filter between the hydraulic channels. A n alternate w ay to conceptualize this modification is to view the reduced Gap A as an orifice. With the proper Gap A and overlap, this clearance can be considered as analogues to a running clearance similar to an impeller/case wear ring combination.

Gap B is defined as the percentage diametral distance between the OD of the impeller vanes and the volute lip (diffuser vane). A Gap B modification can be described as the optimum gap that allows the fluid discharged from the impeller to enter the volute/ diffuser at an angle that approaches shockless entry. At the pump's best efficiency point, the impeller velocity discharge angle is well m atched with the volute/diffuser entrance angle. At low flow conditions with a relatively small Gap B. the fluid does not have the space or time to adjust and, therefore, enters the diffuser at a high incident angle, which results in high-frequency shocks at the vane passing frequency. The dynamic forces acting on the volute/diffuser originate from a nonuniform pressure distribution of the fluid as it exits from the impeller and impinges on the volute/diffuser. This nonuniform pressure distribution is a direct result of the finite thickness of the impeller vanes.

The benefits of performing a Gap A and Gap B modifications are: • Reduced structural failures • Increased system stability • Increased low flow operation stability • Reduced noise and vibration levels • Increased seal and bearing life • Reduced potential for volute/diffuser cavitation • Increased time between maintenance intervals

2.9 Wear Ring Design

Wear rings are devices whose function is to limit leakage between stationary parts and the impeller by means of close clearances. Typical diametral wear ring clearances are in the order of 14 to 20 mils. These clearances are susceptible to wear due to normal flow over time, particulates in the fluid flow, and rubbing action. There is a differential hardness between opposing rings that minimizes the potential for seizure should contact be made. Wear rings are machined from forgings and undergo heat treatment in their manufacturing process to establish desired hardness levels. On either the impeller hub OD or the inner diameter (ID) of the case wear ring, the designers have used groov- ing or serrations to minimize flow leakage across the rings and to reduce the possibility of seizure should rubbing occur.

2-12 EPRI Licensed Material Main Feedwater Pump Maintenance Guide 2.10 Pump Shaft

The pump shaft is a forging that should undergo stress relief heat treatment while being hung vertically during its manufacturing process. The shaft is a support platform for the impellers and other rotating components. The pump shaft transmits axial thrust from the impellers to the thrust bearing and torque from the driver to the impeller.

The rotating element of a multistage volute pump can be balanced as an assembly and installed after balancing, whereas that of the diffuser cannot, as it is built up sequen- tially from a diffuser case to the impeller to the diffuser case and so on. To achieve the best balance, the rotating element is built up without the diffusers, balanced to the best degree possible, disassembled, and then rebuilt with both the impellers and diffusers together. Typically, diffuser pumps can experience more hydraulic and rotor instability problems than volute pumps, due to the difficulty of precisely locating the axial location of the rotating element within the diffuser stackup. These types of problems can often be corrected with Gap A or B modifications. The fit between the stacked diffusers is usually an interference fit, with the amount of interference determined by the size of the equipment.

2.11 Hydraulic Balancing Devices

The balance disk typically consists of at least two orifices m series with one another. One orifice is vertical with a close running clearance. The pressure distribution and drop in the vertical orifice controls the location and stability of the rotating element and can be considered a self-compensating orifice. This is critical to the healthy operation of the equipment away from the best efficiency point. Frequent starts and stops can shorten service life. (See Figures 2-4 and 2-5 for illustrations of balance devices.)

M any multistage pumps employ some type of device to reduce thrust of the rotor due to hydraulic forces. Typically, a multistage stage pump that employs opposing impellers in an equal amount of stages does not employ such a balancing device. Pumps with impellers in series, all facing in the same direction, must use some sort of balancing device to prohibit thrust bearing failure.

Two types of balancing devices are usually employed on main feedwater water pumps: (1) the balancing disc, and (2) the balancing drum. The specifications and the clearances for these type devices might vary, the original equipment manufacturer (OEM ) should be consulted when verifying if the as-found condition and measurements of these devices is satisfactory for installation. W hile balancing the axial thrust, the balancing device reduces the stuffing box pressure adjacent to the last stage impeller. Some manufacturers call the balancing drum a pressure-reducing sleeve and bushing. This is not to be confused with a throttle sleeve and bushing or a fixed-breakdown sealing system.

Hydraulic balancing devices for multiple stage pumps vary in design among manufacturers. The proper dimensions and clearances of the associated parts are paramount in achieving reliable operation, whether they are of the drum or disk design. The setup of these devices can be complicated, and procedures vary with design and the manufacturer.

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Figure 2-4 Balance Disk

Figure 2-5 Balance Drum

2-14 EPRI Licensed Material Main Feedwater Pump Maintenance Guide 2.12 Bearings Axial and radial loads induced by hydraulic and mechanical forces have been considered by pump designers. The design philosophy and practices of manufacturers account for physical differences in various pumps. Not all of the radial and axial forces can be resolved m the hydraulic and mechanical designs. (There is no requirement for handling these loads with radial and axial thrust bearings.) There are three normal choices available for radial loads and one common device used by all manufacturers for thrust loads. The choice for radial load devices is a plain-type journal bearing, a pressure dam bearing, or a tilting pad bearing, also referred to as a self-aligning bearing.

Prevalent main feedwater pump construction is a single-stage high speed configuration that allows the first critical speed to be above the operating speed. The rotors are light weight and the bearings are lightly loaded, which can create instabilities. Other pumps of m ultiple stages also have light bearing loads. C onsideration of these light loads and their instabilities can influence the designer's choice of bearing type.

2.13 Journal Bearings

There is a journal bearing In a housing at the inboard and outboard ends of the pump. The housing contains a bearing mounting structure and provisions for temperature sensors, lubrication oil entry, lubrication oil drain, and drain sight flow glass. The housing is horizontally split, bolted together, and mounted to the pump covers with either with a 180° or 360° mounting flange. The journal bearing consists of split steel inserts with babbitted inserts mounted in split bearing retainers.

The least complex and costly construction is a plain journal bearing with a plain sleeve or an axial groove. This geometry does not resist subsynchronous excitation problems and, thus, can be involved in problem situations. An upgrade in construction to resist excitations is the common pressure dam or pressure pad bearing, which is a commonly used construction that has greater capability to handle increased loads and to resist subsynchronous rotor vibrations.

The third selection (the one used in most main feedwater pumps) is the tilting pad bearing that contains pivot shoes, also referred to as self-aligning bearings. The biting pad journal bearings contain several babbitted pads that can align themselves to accommodate shaft motion. Dowels are used to hold and position the pads. Although the majority of installations use this configuration on high-speed lightly loaded applications, other configurations might also provide reliable service, and Ring pad bearings are not completely problem-free. A ny installation can experience problems in certain situations. Illustrations of journal bearings follow m Figures 2-6, 2-7, and 2-8.

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Figure 2-6 Types of Journal Bearings

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Bearing Assembly at Coupling End

Figure 2-7 Typical Radial Bearing Bracket Assembly, Tilting-Pad R adial B earing

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Figure 2-8 Typical Insert-Type Sleeve Bearing

2.14 Thrust Bearings

M ost designers select a thrust bearing design for handling the residual axial thrust load of minimized or partially balanced out bearings in the pump's internal construction and hydraulics. All main feedwater pumps use a Kingsbury, or Kingsbury-type, pivot shoe thrust bearing that have been tine-tested. The prevalent choice for the number of bearing pads in the basic construction has changed with time, but the conceptual construction has remained the same.

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The thrust bearing handles the unbalanced hydraulic and mechanical axial forces acting on the rotor. These are minimal with the prevalent single-stage double-suction impeller where the axial forces are essentially in balance in either direction due to its symm etry. A small unbalance is designed into the pump to provide a small load that precludes a shuttling effect on the bearing due to oscillatory axial forces. Transients, including startup and shutdown, can create significant unbalanced hydraulic forces. The design is adequate to contain these forces and preclude axial motion that could allow rotor parts to move and rub.

If the internal pum p design has a failure in its axial balance device, then it is possible for excessive loads to be imposed on the Kingsbury thrust bearing. A more common failure mode would be the entry of contamination particles, steam, or water into the bearing housing, which could lead to failure of the thrust bearings. Should a balancing drum or disk fail, besides the additional axial load imposed on the thrust bearing, there is the possibility of steam emission from the stuffing box that could penetrate the bearing housing and contaminate the oil. Illustrations of thrust bearing types follow m Figures 2-9 and 2-10. Figure 2-11 shows a typical outboard end thrust journal bearing bracket assembly.

Figure 2-9 Kingsbury-Type Thrust Bearing

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Figure 2-10 Typical Kingsbury Thrust Bearing Assembly

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Figure 2-11 Typical Bearing Construction Outboard End Thrust and Journal Bearing Bracket Assembly

2-21 EPRI Licensed Material Nuclear Maintenance Applications Center 2.15 Lubrication Systems

The lubrication system for any of the installations feeds pressurized oil to the bearing brackets. Internal construction features within the bracket, which vary by vendor, distribute the cool oil to the bearings according to the design requirements. The system allows the oil to drain from the bearing brackets, return through properly sized and sloped piping to a reservoir source where the oil is filtered and cooled, and returned under pressure in a closed-loop system to the bearing brackets. The reservoir source and the location and type of pump pressure source varies.

On a smaller motor-driven pump installation, the main feedwater pump can have its own integrated pressurized oil lubrication system. This system usually has a separate auxiliary motor-driven oil pum p that is initialized on the m ain feed water pum p starting sequence to provide initial pressure to the system. When the predetermined pressure level is attained, the m ain feed water pump can initiate its start. Norm al construction provides for a shaft-driven oil pump to assume the task of pressurized oil delivery to the bearings, and the auxiliary pump is cut out. The shaft-driven oil pump can either be mounted directly on the end of the pump shaft, or it can be geared vertically downward to a pump at the foundation level with other lubrication system components.

The lubrication system can be designed to provide the required oil flow to both pump and driver bearings, including a step-up gear (if included). The central lubrication system can also originate with the gear lubrication system, providing flow to the pump and motor bearings.

W hen the pump is turbine-driven, the lubrication requirements of the turbine outweigh those of the pump. In this situation, the turbine lubrication system provides pressurized oil flow to the pump bearings. Typical forced-feed oil lubrication systems for motor- gear and turbine-driven situations are shown in Appendix F.

The lubrication oil used is specified by the vendor. Normally, it is a turbine type oil that is sulfur free and contains no chlorine or acid contamination. An oil filter in the integrated system maintains oil cleanliness (and requires periodic maintenance) and a heat exchanger integral to the system provides temperature control. Instrumentation in a normal system includes sight flow glasses in the bearing drain line descents, a temperature sensor for low system pressure alarm in the oil supply header, and temperature switches on each radial bearing to alarm on high bearing temperature. It is possible that there are some systems with thermocouples m the thrust bearing pads to provide additional temperature data.

2.16 Stuffing Box Sealing Mechanisms

Reliability of the device used to seal the two stuffing boxes of a main feedwater pump determ ines the overall reliability of the pump. Failure data presented in this guide verifies that the most prevalent failure of a main feedwater pump is due to the failure of the sealing device used m the pump stuffing boxes. It is necessary to preclude the pumped water, which can exceed 400°F, from reaching the atmosphere and flashing. An added problem in BWR service is the radioactivity of any escaping fluid. A closure

2-22 EPRI Licensed Material Main Feedwater Pump Maintenance Guide mechanism in the stuffing box prevents flow from between the casing of the pump and the shaft at each end of the pump.

Historically, there have been four generic methods of providing a closure mechanism in the stuffing box between the shaft and the pump case—rings of packing, floating rings, injection throttle bushings, and mechanical face seals (see Figure 2-12). Packing is an obsolete method and was never used in main feedwater pumps. Floating ring construction (Figure 2-13) was used in early installations by some manufacturers, but sensitivity of the device with its close running clearances and susceptibility to particle contamination at low speeds coupled with repeated failures caused this construction to be replaced with a labyrinth-type injection throttle bushing (Figures 2-14 and 2-15).

The injection throttle bushing is a long-lived device in high horsepower, high speed service. However, the mechanical seal in boiler feed applications is gaining greater acceptance today. In earlier years, injection type throttle bushings were the predominant sealing device used. European power generators have been the leading users of mechanical seals in high horsepower, high speed boiler feed pumps. The U.S. power generator industry has long preferred the reliability of the injection throttle bushing over a mechanical seal. From an economic standpoint, the thermal savings with a mechanical seal provide a short-term justification payout, assuming a reasonable degree of reliability can be achieved.

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FIXED BREAKDOWN BUSHING

MECHANICAL FACE SEAL

FLOATING RING SEAL

Figure 2-12 Typical Stuffing Box Configuration

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Figure 2-13 Typical Stuffing Box Detail, Floating Ring

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SINGLE INJECTION SEAL

SINGLE INJECTION WITH A HIGH PRESSURE DRAIN SEAL

DOUBLE INJECTION WITH A HIGH PRESSURE DRAIN SEAL Figure 2-14 Types of Labyrinth Injection Throttle Bushings

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Figure 2-15 Typical Stuffing Box Detail, Injection Throttle Bushing

2.17 Injection Throttle Bushing

The injection throttle bushing as a closure device allows a small volume of liquid to be cooled as it passes through its small clearances. Cold condensate ranging from 90-120°F is introduced into the bushing clearances, mixes with hot water from the pump, and either exits to the atmosphere or has sufficiently high-pressure for a small quantity to enter the pump. To increase the throttling action capabilities and improve the pressure loss capabilities of the bushing, radial serrations or grooves are used. There are different design philosophies among manufacturers as to the best location and design of the serrations. The serrations, or grooves, can be placed on the rotating shaft or shaft sleeve, on the stationary bushing surface, or on both stationary and rotating members of the configuration. Some designs use spiral threads to direct the leakage to the inside of the pump. Besides the benefits derived through pressure breakdown throttling, the serrations provide the additional benefit of allowing passage of particulates through the grooves and out of the close clearances.

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Pneumatic control devices for injected water have not provided the reliability expected. Advances m electronic control devices have reduced control problems by replacing the pneumatic controls with electronic controls. Electronic controls have provided exceptional control capability, and have eliminated some problems.

The diametral clearances between the bushing and the shaft sleeve are typically 10-18 mils, depending on the diameters involved. This clearance normally allows a leakage rate of 15-30 gpm. When the clearances in the throttle bushing open due to rubbing and wear, the control system requires additional flow to maintain the desired pressure differential or temperature. When wear becomes excessive and the flow control valve is at its open limit, the temperature m the outflow drain increases until the water vaporizes. This steaming effect eventually introduces water into the bearing housings, mixing the water with the oil. This causes eventual breakdown of the bearing lubrication process, causing bearing degradation and ultimately bearing failure. If not corrected, this can lead to massive component destruction of the pump and rotor, incurring great cost.

2.18 Injection Control Systems

The system used to control the injection of cool water into the throttle bushing is critical to the reliability of the sealing system. Pressure and temperature m the control system regulate the flow to the throttle bushing. Although the specifics within plants differ, essentially all use condensate pump discharge through a series of heaters to obtain the proper main feedwater pump suction temperature. The main feedwater pump discharge then delivers the flow at increased pressure through high-pressure heaters and on to the reactor or steam generator.

There are five basic control systems for a closed feedwater circuit: • Pressure-controlled injection system • Drain temperature-controlled injection system • Intermediate outlet type bushing system • Reduced temperature intermediate outlet type bushing system • Intermediate outlet for closed feedwater circuits

With the options available, a control system is possible for any system configuration. The trend in later years has been distinctly in favor of a temperature drain-controlled system . These two system s are discussed m the following text.

2.19 Pressure-Controlled Injection System

A pressure-controlled system measures and maintains a predetermined set pressure differential between the seal injection water and the suction of the main feedwater pump (Figure 2-16). The injection water has been sourced from the condensate pump system at about 100°F and is delivered to the throttle bushing at a constant differential value pressure slightly higher than suction pressure. There is a minor flow into the pump that, as the bushing wears, can increase and create mechanical problems. Thermal distortions

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created by the cool inflow can cause distortions of the pump case or barrel and can result in rubbing and wear. Operational requirements and procedures have stretched the limits of the capabilities of pressure-controlled systems. In many instances, the pressure-controlled system has been replaced by a drain temperature controlled system.

2.20 Temperature-Controlled Injection System

The temperature drain-control system has the same source of condensate water inject- ing into the throttle bushing clearances. A typical temperature drain-control system is shown in Figure 2-17. The cool condensate water mixes with the hot feedwater as it passes through the stuffing box to the atmosphere and into the seal drain line. The temperature of the drain water is controlled at about 150°F. If the temperature increases, the injection flow control valve opens, allowing additional cool condensate to mix with the feedwater. Advantages of the temperature-controlled injection system are: its reliability and simplicity of temperature detection versus differential pressure-controlled devices; the increased mixing of the, waters; and the basic elimination of cool water flow into the pump causing thermal distortions with resultant problems on the pump case and rotor.

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Figure 2-17 Drain Temperature Control System Regulating Condensate Injection to Feed Pum p Shaft Seals

2.21 Mechanical Seals Although mechanical seals are second in total usage, there has been an increase in their success in high horsepower, high speed applications. Sophisticated seal face finite element deflection analysis capability exists with all seal vendors. This capability did not exist when most of the main feedwater pumps entered service. Today, seal design engineers can evaluate their designs analytically before incurring the expense of performing actual tests or evaluating field installations. Advances in materials have supported significant increases in m echanical seal reliability. The most significant metallurgical advance has been the development of entire families of silicon carbide rotating face metallurgies. New elastomeric compounds with improved frictional and wear characteristics are now available for gasket materials. Figure 2-18 is one example of a cartridge-type mechanical seal.

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Figure 2-18 Typical Cartridge-Type Single Mechanical Seal

The primary advantage of a mechanical seal over an injection throttle bushing is that the leakage is comparatively small. This is a primary consideration in BW Rs as the feedwater is radioactive. The disadvantage of mechanical seals has been a lack of consistent reliability.

The statistics in this report verify that mechanical seals are not as reliable as injection throttle bushings. However, the economics of heat balance favor a mechanical seal, and the requirement for a condensate injection system can be eliminated with a mechanical seal. The mechanical seal can be a more forgiving device than an injection throttle bushing when experiencing large radial movement.

2.21.1 Design Features Critical Design Features for High Performance Seals (Reference 62), and Retrofit of an Unspared Main Boiler Feed Pump to End Face Mechanical Seals (Reference 63) contain valuable information relating to im portant design features. The following is a summ ary of the mechanical seal design features believed to be beneficial.

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2.21.2 Cartridge Design

All seal parts should be preassembled into a cartridge for quick and easy installation in the pump. Preassembly of the cartridge can be done by the seal manufacturer or by specially trained plant personnel. Cartridges are assembled under ideal conditions, prior to the pump outage. Time required to assemble the cartridge in the pump, and assembly error during the outage, are minimized

Positive drive of the seal shaft sleeve is an important cartridge design feature. Relying on cup point set screws on a smooth shaft surface is generally inadequate.

2.21.3 Computer Design of Faces

Finite element computer programs can predict distortions of the seal faces caused by changes in the pressure and temperature of the sealed fluid. Seal faces should be designed for full fluid film lubrication (a slightly convergent gap between seal faces) over the full range of normal and transient pressures and temperatures. New designs should be factory tested over the full range of normal and transient pressures, temperatures, and shaft speeds (rpm ).

2.21.4 Stationary Springs

The springs should be stationary and not rotate with the shaft.

2.21.5 Face Material Selection

Face materials are selected by the vendor as part of the design process. The history of tungsten carbide m this service indicates that it should be avoided. The recommended rotating (hard) face is a silicon carbide compounded to vendor specifications for microstructure and quality. The carbon-graphite stationary (soft) face is impregnated with either a resin or metal and compounded to the seal vendor specifications.

2.21.6 Cooling of Faces

Reference 62 cites tests showing that a multiport injection, with 4 to 12 ports equally spaced around the seal interface, proved beneficial when sealing a more volatile liquid. This design feature is unproven in feed pumps, but could minimize circumferentially uneven stationary face temperatures and distortion.

2.21.7 Shaft Centering Plates

The seal cartridge described in Reference 63 has shaft-to-seal cartridge centering plates with close tolerances. This feature reduces the time required to align the shaft to the pump casing. The pump is of double-casing construction (Pump Category 2 in Table 1-2 in Section 1.6 of this guide).

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2.22 Mechanical Seal Operating Environment The environment in which the seal operates affects its reliability. The pump condition determines the environment. Rotating imbalance creates vibrations. The severity of the vibration determines the detrim ental effect it has on reducing seal life. M isalignm ent of the pump to driver again creates vibrations that reduce seal life. The seal injection system and its adequacy determines how well heat is removed from the seal face area.

Seal design should include a pumping ring that circulates water from each seal chamber (stuffing box) through a heat exchanger and back to the seal chamber while the pump is running. The cooling system must be designed so that there is no possibility of 212°F steam being formed at the low-pressure edges of the seal faces; therefore, 185°F is considered the absolute maximum allowable seal chamber temperature. Design temperature should be much lower (the operating temperature for the seals in Reference 63 w as 100°F to 105°F). The heat exchangers should be located well above the seal chambers to provide thermal cycling to keep the seals cool while the pump is on hot standby. The piping must be properly vented. Magnetic separators might also be installed in the cooling circuits. Condensate injection cooling of mechanical seals is undesirable because it might introduce particulates or create undesirable thermal transients.

The intent of any seal design is to perform reliably during any normal or emergency transient and to provide long life with stable operation. Some operating problems, such as operation at low flow, are true tests of a seal's ability to survive. When the installed main feedwater pumps were originally procured, pump manufacturers w ere far more liberal in allowing low flows as a percent of best efficiency point flow. Competitive evaluations were made based on the size of the recirculation piping that usually ran back to the condenser. Both manufacturers and users suffered from destruction of recirculation components and often had to resize to a larger system. Seals were the victims of vibration levels from operation at low flows. Obviously, no seal could be expected to operate at no flow through the pump. At times, this mode was attempted, but not intentionally. Seal failure ensued. Another source of operational problems was when the pump operated in a cavitation mode. The severity of cavitation determines the severity of the vibrations, which determines the life of the seal.

For all their sensitivities and past problems, the technology of the seal industry as it applies to boiler feed pump applications will undoubtedly result in a greater share of the stuffing box marketplace. With improved operational and maintenance practices, reliability of boiler feed mechanical seals can be greatly improved.

2.23 Pump Couplings The mechanical device used to transmit the driving torque to the main feedwater pump is critical to the overall reliability of the unit. Three types of couplings have been used for this service. Besides the transmission of torque, a function of the coupling is to provide sufficient flexibility to accommodate misalignments between pump and driver without causing vibrations and failures.

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The couplings are located either directly between the pump and a motor or turbine driver as the prime mover, or on either side of a step-up gear located between the prime mover and the pump. Gear couplings require lubrication, either of a continuous design or one that involves periodic injection of lubrication into the assembly.

Three types of couplings are found on feedwater pumps: (1) the flexible disk coupling, or disk pack, (2) the diaphragm coupling with one steel flex plane at each end of the spacer that performs similarly to the disc coupling, and (3) the gear or marine type coupling. The couplings are located either directly between the pump and a motor or turbine driver as the prime mover, or on either side of a step-up gear located between the prime mover and the pump. Gear couplings require lubrication, either of a continuous design or one that involves periodic injection of lubrication into the assembly.

The construction of gear couplings involves two hubs containing external gear teeth that mesh with the internal gear teeth of a sleeve device that fits over the external teeth on the two hubs. Gear couplings have had mixed success. Over time, many gear couplings have been changed out to other designs. Lubrication of the gear coupling has been the source of some of the problems, but this might be more a problem of human neglect or lack of training. Without proper lubrication, the gear overheats, wears, and fails. The limitations of the gear coupling to accommodate misalignment and remain flexible has probably resulted in its replacement with another type. The coupling can lose its capability to accommodate misalignment and remain flexible. The load and torque magnitudes are such that the coupling fixes into a misaligned configuration, causing high vibration levels. The advantage of a gear coupling is that it can be designed to quickly disconnect and be activated when the turbine is on turning gear. Operation of the main feedwater pump at turning gear speeds produces numerous potential problems for the pump, which are eliminated when it is disconnected from the turbine.

Both flexible diaphragm and disk couplings require constant lubrication maintenance. These couplings can accommodate misalignments which gear couplings sometimes cannot. Contoured diaphragms or sheet metal plates transmit torque through the con- nection to the shaft supports. They can flex under misalignment conditions causing problems for gear couplings. However, contoured diaphragms do not have the disconnect potential of gear couplings. The flexible diaphragm or disk coupling is the predominant choice for main feedwater pumps.

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Figure 2-19 Diaphragm Coupling

The diaphragm coupling (Figure 2-19) has been the coupling of choice when retrofits or change-outs occur. The diaphragm coupling imposes one less bending moment on the shafts than the gear-type coupling. The axial shaft movement (displacement) and separation must be measured and calculated much more closely for these couplings.

Regarding concerns about the retrofit of an existing coupling to that of another type, consideration must be given to the advantages and disadvantages of each type of coupling.

Diaphragm coupling versus gear coupling advantages: • No lubrication required • W ill tolerate greater parallel and angular misalignment • No rubbing or sliding parts to wear • Fewer dynamic balance problems • Less angular moment • Zero backlash • Longer life of radial and thrust bearings • Near elimination of hub-to-shaft fretting

Disadvantages of flexible diaphragm couplings: • Lim ited axial travel • Less forgiving in shaft separation errors 2-35 EPRI Licensed Material Nuclear Maintenance Applications Center

• Possible axial resonance • Failure mode might allow runaway without warning • Heat generation due to windage • Gear coupling smaller based on torque requirements

Figure 2-20 Flexible Disk (Disk Pack) C oupling

The disk pack coupling (Figure 2-20) is gaining favor among rotating machinery specialists. These couplings have more overhung weight than diaphragm couplings. The disk pack of flexible disk couplings uses many laminations of flexible steel sheets to provide for proper torque transmission and adequate flexibility. As with the diaphragm-type coupling, the flexible disk coupling offers the advantage of being non- lubricating. The physical size of this coupling is generally greater than a gear coupling and possibly weighs more. The flexible unit in this coupling should be checked for fatigue and cracks during pump inspections.

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Figure 2-21 G ear Coupling

Gear couplings used on feed pumps vary in design. Gear couplings on feed pumps are usually supplied with crowned teeth on the hubs. The coupling in Figure 2-21 uses a spacer or spool piece on feedwater pump applications.

Lubrication is essential to the life of a gear coupling. Gear couplings are specified with grease lubricant or as continuous lubrication. Continuous lube couplings are enclosed with a large cover to contain the oil. The oil is usually supplied by the same oil pump that supplies the bearing oil (see Figures 2-22 and 2-23).

With many of these couplings operating at speeds in excess of 3,600 rpm, proper grease is a must. As misalignment increases (so does temperature) the lube ability of the coupling grease diminishes. Below is a cross-section of a continuous lube coupling showing the oil nozzles.

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Figure 2-22 Continuous Lubricated Gear Coupling

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Figure 2-23 Continuous Lube Coupling Cover

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FAILURE MODES AND EFFECTS ANALYSIS

3.1 Introduction

3.1.1 Purpose

The purpose of failure modes and effects analysis of the main feedwater pump failure data is to determ ine which major topics are to be covered in this guide.

3.1.2 Methodology

All available failure data was analyzed.

Data sources included: • Nuclear Plant Reliability Data System (NPRDS) • NM A C M ain Feedwater Pum p M aintenance Survey • Previous EPRI pump failure surveys NPRDS provided failure narratives from 96 nuclear units. Data from 684 selected failures was entered into a spreadsheet with narratives reduced to failure symptom, failed component, and reported failure cause. Data was then sorted and tabulated for analysis.

The NMAC M ain Feedwater Pump M aintenance Survey was conducted in April 1995. The responses from 27 units were tabulated and analyzed.

EPRI-sponsored surveys of feedwater pump failures in both fossil fuel and nuclear power plants were published in 1978 and 1982. The EPRI-sponsored, Boiler Feedpump Operation and M aintenance Guidelines was published in 1994 (References 17 and 18).

3.2 Failure Data

3.2.1 Feedwater Pump Failure Narratives

Source of NPRDS Data

M ore than 700 main feed water pum p failure narratives com piled by the N PR DS were tabulated by NMAC in April 1995, providing the data for this analysis.

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Screening of NPRDS Data

The data was screened using the following criteria: • All failures occurring within one year of plant startup were excluded. Although most of the failure data is from 1984 through 1994, earlier failures as far as 1976 were included if they occurred more than a year after plant startup. • Failures of equipment items other than the pump were excluded. The pump includes the pump-to-driver coupling, the entire lube oil system (often a common system with the driver), auxiliary water piping (for vents, drains, seals, lube oil cooling), and related instrumentation.

In this guide, a failure is defined as any event reported as a failure in the NPRDS data tabulated by NMAC in April 1995.

Sorting of NPRDS Data

Tabulated data included unit identification, plant type, pump manufacturer, manufacturer's model, failure start date, symptom, failed component, pump identification, and reported cause of failure.

Data was entered alphabetically by unit identification. The symptom, failed component, pump identification, and failure causes were extracted from each failure narrative. After all data was entered, it was sorted by failure start date, symptom, failed component, and failure cause.

For this guide, symptom is defined as the circumstance, event, or condition observed by unit personnel that led them to conclude there had been a failure.

Tabulation of NPRDS Data

NPRDS-reported failures in each year from 1976 through 1994 are listed in Table 3-1. Few failures were reported prior to 1984, although it is known that failures did occur. The number of failures reported in 1993 and 1994 declined substantially from the 1984- 1992 period.

Changes in the NPRDS failure reporting system might have affected the number of failures reported in these time periods:

August 1983: Main steam, main feedwater, main condensate, and auxiliary/emergency feed water systems added to NERDS requiring additional components.

December 1992: Reporting of failures due to leakage of process and operating fluids from all components was revised to include only failures that degraded operability of the component or required prompt action to mitigate the effects of the leak. This revised guidance substantially reduced the number of leakage failure reports submitted.

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Table 3-1 NPRDS Reported Failures Each Year

Year Number Percent* Year Number Percent* of Failures of Failures

1976 1 0 1986 78 12

1977 1 0 1987 51 8

1978 2 0 1988 34 5

1980 2 0 1989 95 14

1981 2 0 1990 74 11

1982 1 0 1991 92 14 1983 7 1 1992 90 13

1984 56 8 1993 27 4

1985 55 8 1994 16 2

* Percent of total number of failures

Failure symptoms are listed in Table 3-2. A majority of symptoms are leaks from lube oil systems, shaft seals (especially mechanical seals), water piping, and gaskets. Casing flat gaskets refers to the gaskets in the axially split main casing joints of Category 3 pumps. (See Sections 1.6 and 5.7.1.)

Table 3-2 N PRD S Reported Failure Sym ptom s

Location of Symptom Number Percent* of Failures

Lube oil system 200 29

Mechanical shaft seals 182 27

Water piping 56 8

Miscellaneous gaskets 55 8

Shaft or bearing housing vibration 50 7

Other 43 6

Bearings 35 5

Controlled leakoff seals 31 5

Casing flat gaskets 19 3

Seizures 13 2 * Percent of total number of failures

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Failed components are listed in Table 3-3. The components that fail most frequently are lube oil systems, shaft seals (especially mechanical seals), water piping, and gaskets. Significant failure rates were also reported for bearings, couplings, impellers, and other parts of the rotating elements.

Table 3-3 NPRDS Reported Failed Components

Failed Component Number Percent* of Failures

Lube oil system 200 29

Mechanical shaft seals 182 27

Gaskets and O-rings 75 11

Water piping 51 7 6 Bearings 40 Other 38 6

Controlled leakage seals 30 4

Casing flat gaskets 19 3

Couplings 14 2

Impellers 13 2

Rotating elements 10 1

Shafts 8 1 Bearinghousings 6 1 * Percent of total number of failures

Reported failure causes are listed in Table 3-4. Few root-cause analyses were done. The root cause of more than one-half of the failures was unknown, and almost one-third was reported as caused by the end of the normal life of the part or component. The remaining causes were classified as maintenance, manufacturer (often not the pump vendor, but the manufacturer of the mechanical seal, gasket, lube oil pump, or other component), and operation (often beyond the control of the plant operators).

Table 3-4 NPRDS Reported Failure Causes

Failure CauseNumber of Failures Percent*

Unknown 348 51 Normal life 218 32

Maintenance 75 11 Manufactures 24 3 Operations 19 3

* Percent of total number of failures

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Discussion of N PRD S D ata

The NPRDS data provides valuable insight into the operating histories of a wide variety of main feedwater pumps. Forty-nine different models of pumps from six different manufacturers were listed showing 684 documented failures m 96 different units.

Twenty two percent of the units reported 55% of the failures. Some units report relatively minor items such as small lube oil or auxiliary water piping leaks as pump failures, while others do not.

Table 3-1 indicates that only 16 failures were reported from 1976 through 1983 and decreased reporting in 1993 and 1994. Changes in the NPRDS reporting system in mid- 1983 and late 1992 caused these differences.

Table 3-2 shows that a large majority of failure symptoms are easily observed by plant personnel inspecting the pumps. Infrequent failure symptoms, listed as other in Table 3- 2, include casing leaks, noise, coupling misalignment or failure, driver trip or lock out, electrical ground, high torque, unsatisfactory hydraulic performance, broken shaft, bearing wear and seal water alarms, and turbine overspeed.

Table 3-3 lists the failed pump components in the order of failure frequency. It does not show the seriousness or cost of the failure of each component. However, it is of great value to identify pump com ponents that require good m aintenance practices. Infre- quently failed pump components, listed as other in Table 3-3, include pump casings, bolts and cap screws, set screws, covers, controls and alarms, splitters, thrust nuts, and inlet turning vanes.

Table 3-4 emphasizes that it was generally impractical to perform a root-cause analysis of a pump failure for the following reasons: • All data needed for root-cause analysis was not available. • Failure narratives often set forth opinions that could correctly identify the problem but are not based on a systematic analysis of available information. • The author of the narrative often listed a symptom as a cause. Vibration is a good example. It might be the immediate cause of a failure (such as a shaft seal failure), but the root cause is the reason for the vibration (such as rotor instability). To carry the example one step further, the vibration might be caused by dynamic unbalance of the rotor for which the root cause was a broken impeller or a bent shaft.

Conclusions from NPRDS Data

The NPRDS data is a valuable asset for the maintenance practices documented in this guide.

Care must be exercised when interpreting the data, because reporting is subjective and varies from unit to unit. Also, we elected to use the entire time frame, so our compila-

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tions do not account for the fact that more recent failures are more relevant. The causes of earlier failures might have been diagnosed and corrected. A majority of failure symptoms are leaks from the lube oil systems, shaft seals (especially mechanical seals), the water piping, and the gaskets. These symptoms are easily observed by operating personnel. This might lead to an inadequate collection of additional data that could better diagnose the root cause of failure and, thus, avoid its recur- rence. Predictive and preventive maintenance, including monitoring and diagnostics, are covered in detail in Section 4 of this guide.

This guide covers proper maintenance of: • Lube oil systems • Mechanical shaft seals • Water piping • Miscellaneous gaskets • Bearings • Controlled leakage shaft seals • Casing flat gaskets

M echanical shaft seals and casing flat gaskets are under-reported in Table 3-3, because many main feedwater pumps do not have mechanical seals or flat casing gaskets.

Maintenance and upgrading of mechanical shaft seals and the environment in which they operate are obviously areas for major improvement.

3.2.2 NMAC Main Feedwater Pump Maintenance Survey

Source of NMAC Data

The NM AC Main Feedwater Pump M aintenance Survey was conducted in April 1995. Updated and detailed data was solicited from all units owned by EPRI members and from selected units in foreign countries. Responses were received from 22 domestic units, and units in Canada, Taiwan, and M exico.

Summary of NMAC Data

A ll responses from foreign countries indicated that their feedwater pump maintenance problems have been minimal or nonexistent.

Nineteen of the 22 domestic units responding indicated minimal maintenance problems. All have fixed breakdown-type controlled leakage shaft seals.

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The other three have significant problems. All reported that mechanical seals were still a problem, but one of these three has installed upgraded seals that seem to be more reliable. In one unit, radial vibration of 0.5 inches per second was observed. Vibration above 0.24 or 0.3 inches per second is generally considered to be cause for concern. The diffuser in one of the pumps was changed from 8 to 12 vanes, and the vibration dropped to 0.1 inches per second.

The NMAC survey asked: "W hat are your major problems/maintenance issues with the feedwater pumps?" The following were indicated as ongoing problems: Mechanical seal leaks Lube oil leaks Water leaks Pipe stress and alignment Gland water adjustment on startups High vibration Breakdown bushing wear Cavitation damage and impeller cracks Alignm ent w ith drive turbine Couplings lose sealed-oil lubricant The NMAC survey also asked: "W hat things would you like to change if you could ?" The following replies were received: Pump seals Oil filtering Pipe stress Vibration monitoring system More accurate instrumentation for trouble shooting and pump performance Difficulty m getting contaminated pump parts reworked Electronic-controlled seal injection system to replace pneumatic system O-rings with more than 54 months life Larger mechanical seal coolers System for easier and more effective venting Upgrade the installed lube oil purification system Lower bearing oil temperature Change from breakdown bushings to mechanical seals Stretch out preventive maintenance frequency M aintenance guidelines Definition of failures The survey also requested the frequencies of scheduled maintenance, minor inspection, and major inspection. Most units have a minor inspection and/or scheduled maintenance every 18 months, the common refueling outage frequency. Major teardown inspections vary from once each refueling outage to as much as every 10 years.

Several respondents indicated a desire for better monitoring and diagnostic capabilities.

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Conclusions from NM AC Data

Although only 22 of the 96 domestic units represented in the NPRDS data responded to the NMAC requests, much useful information was provided. The case studies cover many of the non-responding units that have significant ongoing problems.

The survey reinforced the conclusions from the NPRDS data that ongoing maintenance problem s include lube oil systems, shaft seals (especially mechanical seals), water piping, and gaskets.

Other important concerns included the following: • Coupling alignm ent, especially with drive turbines. M ost responders have not replaced their gear-type couplings and are not reporting coupling problems. Only one reported unacceptable piping loads on the pump. • Controlled leakage shaft seals. Their failure rate appeared to be much lower than that of mechanical seals. Breakdown bushing wear and problems with the injection system control piping leaks were mentioned. Two units would like to change from breakdown bushings to mechanical seals, and one of these is actually installing mechanical seals on a trial basis. • Monitoring and diagnostic systems. They have been cited as needing improvement. • Contaminated pump parts from BWR units. They are a problem when they require rework.

3.2.3 Previous EPRI Pump Failure Data

Sources of EPRI Data

The earliest EPRI-sponsored research into feedwater pump failures was a survey of feed pump outages in 1978 (Reference 1). Both fossil-fueled and nuclear plants were sur- veyed. The 10 major outage-producing failure causes listed in order of total number of failures were: Shaft seal failure Vibration: pump, piping foundation Axial balancing device failure Journal bearing failure Cavitation damage Impeller breakage or cracking Rapid wear of wear rings Unstable head curve Broken or damaged shaft Thrust-bearing failure

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An EPRI-sponsored survey in 1982 (Reference 5) listed the following feed pump problems, not necessarily in order of severity or frequency of problems: Balancing device failures Hot misalignment First-stage cavitation Seizures Failure of inter-stage partitions Cavitation damage Impeller cracking or breakage Internal leakage and wash out Gear-type couplings Vibration Shaft failures Auxiliary system reliability

The same 1982 study tabulated what it revealed was the most costly problem at the top of the list and the least costly at the bottom: Vibration Impeller breakage or cracking Seal failure Rapid wear of wear rings Cavitation damage Axial balancing device failure Broken or damaged shaft Journal bearing failures Seizures of wear rings, etc. Thrust bearing failure Unstable head curve Auxiliary system reliability Hot misalignment Gear-type couplings In another EPRI-sponsored study in 1982 (Reference 4), pump failures were ranked by total plant hours lost from each pump failure cause. Pump failure rankings were as follows: Cavitation, unstable head curve, impeller breakage, vibration Seals Wear rings Axial balancing device Shaft broken or damaged Journal bearing Thrust bearing The EPRI-sponsored Boiler Feedpump Operation and M aintenance Guidelines was published in 1994 (Reference 17). This document is limited to boiler feed pumps in fossil- fired plants, with emphasis on large plants of 600 or higher MWe. Extensive studies of

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these plants showed that the major problems, classified by failed component, still existing for large boiler feed pumps are: Casing (barrel) damage End head damage Seal housing damage Seal damage, static non-gasketed Seal wear, dynamic-labyrinth Seal damage, dynamic-mechanical Journal bearing, damage-mechanical Thrust-bearing failure Thrust-bearing damage Shaft breakage Shaft damage-cracked Shaft damage-mechanical Impeller breakage Impeller damage-cavitation Diffuser/volute breakage Diffuser/volute damage-cavitation Wear ring-wear Axial balancing device-failure Axial balancing device-damage Axial balancing device-wear Bolt breaking Bolt damage-cracked Key damage-mechanical Coupling-gear tooth breakage Coupling-gear tooth damage Coupling-disk/diaphragm damage In addition to the section dealing with the above list of failed components, Reference 17 is an excellent source that addresses operating problems. In this guide, such operating problems are generally referred to as failure symptoms.

D iscussion of EPRI D ata

It is important to understand that, although the 1978 and 1982 studies included nuclear power plants, all EPRI feed pump investigations to date have focused on the boiler feed pumps in large fossil-fueled plants (600 and higher M We). The pumps studied were all barrel-type pumps, typically five or more stages.

The nuclear main feedwater pumps reflect a variety of designs including barrel-type pumps, typically one or two stages. They generate lower pressures, pump higher flow rates, and require lower horsepower drivers. Typical rotating speeds are about the same.

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As a group, the nuclear main feedwater pumps should be more reliable because they have the following advantages: Smaller shaft diameter Smaller journal bearings Smaller shaft seals Sm aller thrust bearings Shorter and stiffer pump shafts Operation at base load rather than cycling

Nuclear main feedwater pumps also have disadvantages: Pumps in BWR plants pump radioactive water. They are far more likely to have mechanical seals instead of pressure breakdown type shaft seals. They may have horizontally split cases with flat gaskets. Impeller stresses caused by centrifugal and hydraulic forces may be greater.

Conclusions from EPRI Data

The 1978 and 1982 surveys (References 1, 4, and 5) plus additional current survey data were available to the technical team members when they produced the 1994 guidelines (References 17 and 18). Thus, the 1994 guide was the best current reference for large boiler feedwater pump operation and maintenance. Its lists of frequently failed components and of operating problems have been carefully evaluated, and information relevant to nuclear main feedwater pumps have been incorporated into this guide.

The EPR I data is not totally applicable to nuclear main feedw ater pum ps, because the emphasis was on large barrel-type boiler feed pumps. The EPRI definition of pump included the drive coupling but excluded or deemphasized the lube oil system, auxiliary piping and related instrumentation.

The definition of failure was almost exclusively limited to problems causing a pump outage and partial or full loss of ability of the unit to generate power. Seriousness of failure was evaluated in terms of cost of replacement power.

Many of the root-cause problems cited in the EPRI data have been cured in some of the nuclear main feedwater pumps. Cures have resulted from actions such as: Eliminating axial rotor shuttling by correcting Gap A Reducing pressure pulsations and rotor distress by correcting Gap B Upgrading shaft seals Replacing gear-type couplings Increased condition monitoring Upgraded lube oil filtration and monitoring Replacing carbon steel, brass, or copper lube oil piping with stainless steel All EPRI studies emphasize vibration, either as a symptom or as a root cause. It is probable that increased vibration monitoring and diagnosis in nuclear units will prove to be a cost- effective predictive and analytical tool. Section 4 of this guide addresses this topic in detail.

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3.3 Conclusions

3.3.1 Reconciliation of EPRI Data

To reach conclusions regarding failure modes in nuclear main feedwater pumps, it is necessary to reconcile the 1994 EPRI data on boiler feedwater pumps with the NPRDS data and look more closely at the NPRDS reported failures of journal bearings, thrust bearings, controlled leakage shaft seals, couplings, impellers, rotating elements, and shafts. Failures of these components are pump failures by any definition. NPRDS failure data was reported by 96 units for a period of at least eight years. See Appendix H for details of the reconciliation of EPRI data.

3.3.2 Sum m ary of the Reconciliation

Journal bearings. Failures are always serious and seem to occur randomly.

Thrust bearings. Failures are always serious and seem to occur randomly.

Balance disks. Do not seem to be a problem.

Controlled leakage shaft seals. All are fixed breakdown-type. Typically, they fail slowly and are almost never the root cause of a serious failure.

Couplings. Almost all failures are avoidable.

Im pellers. Failures are serious, but rare, and are confined to a few units (so far).

Rotating elements. Failures are always serious and seem to occur randomly.

Shafts. Failures are serious but random and rare.

3.3.3 Topics for Main Feedwater Pump Maintenance Guide

Consideration of all data sources leads to the conclusion that major topics to be covered in this guide must include maintenance, preventive maintenance, and/or upgrading of the following pump components. Sections in this guide containing information relating to each identified pump component are listed in Table 3-5.

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Table 3-5 Component Cross-Reference by Sections

Component Sections Mechanical Shaft Seals 2.16, 2.21, 2.22, 5.3.9, 5.4.5, 5.8 Controlled Leakage Shaft Seals 2.16 through 2.20, 5.3.8 Journal Bearings 2.13, 5.3.6 Thrust Bearings 2.14, 5.3.3

Impe lle rs 2.7, 2.8, 5.3.11 Rotating Elements (Rotors) 5.3.4, 5.3.10, 5.4.2, 5.4.3, 5.4.4 Lube Oil Systems 2.15, 4.2.2, 5.2.9, 5.6, 5.7.2 Flanges and Gaskets 5.7.1, 5.7.2 Casing Flat Gaskets 5.7.1

Water Piping 5.3.5, 5.7.2 Couplings 2.23, 5.3.1, 5.3.2, 5.5 Shafts 2.10, 5.3.11, 5.4.2 Wear Rings 2.9, 5.3.11, 5.4.1

Pump Casings 2.6, 5.3.12, 5.9, 5.10, 5.11 Bolting Appendix I

The failure modes and effects analysis supports this conclusion:

....although there have been a number of specific problems which can be attributed to a specific supplier's pump design or design feature, the major factors which contribute to reduced feed pump reliability are under the direct control of the utility engineers and operators.

From a 1980 EPRI document (Reference 3)

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- 4

PREDICTIVE AND PREVENTIVE MAINTENANCE

4.0 Predictive and Preventive Maintenance

4.1 Introduction R elationships am ong m aintenance term s are illustrated in Figure 4-1.

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Figure 4-1 R elationship A m ong M aintenance Term s

Failure of a main feedwater pump can cause a forced outage or a reduction in the power generation capacity of a unit. Efforts should be expended to avoid a forced outage. The primary objective of a good maintenance program is to avoid corrective maintenance resulting from failure during operation. Predictive and preventive maintenance programs precede the actions performed in planned maintenance. (See Figure 4-1.)

Condition monitoring is the systematic collection, storage, and processing of machinery operating parameters. Predictive maintenance includes trending the parameters and maintenance of critical rotating machinery.

4-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide Preventive maintenance is regularly scheduled periodic maintenance that is generally scheduled for refueling outages. Predictive maintenance can make preventive maintenance schedules more logical. The objectives of each are identical.

Diagnostics and root-cause analysis is a related, but separate, maintenance function. It requires predictive maintenance personnel and a group of diagnostic specialists working as a team .

Although this guide only covers main feedwater pump maintenance, predictive and preventive maintenance activities cover all plant machinery. Maintenance activities for a pump must be coordinated with the pump's driver.

4.2 Condition Monitoring

The periodic collection of data necessary to assess the health of an operating pump for predictive and preventive maintenance is known as condition monitoring.

Data is collected m many ways, including: • Manual logging of observations of abnormal noise, leakage from shaft seals or water piping, lube oil leaks, or discoloration • M anual logging of pump-related pressures, temperatures, flows, rotating speeds, and vibrations that are not otherwise recorded • Recording of pump-related pressures, temperatures, flows, rotating speeds, and vibrations on control room computers or charts • Recording of pump-related pressures, temperatures, flows, rotating speeds, and vibrations into a portable or hard-wired computer data collection system with analytical capabilities • Lube oil analysis

4.2.1 Parameters To Be Monitored Hand-held data collectors with vibration spectrum analyzers and trending capabilities have proven to be essential to cost-effective operation and maintenance of critical rotating machinery.

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M inimal data collection includes the following information:

Data Item Method Frequency Abnormal Noise Logged Manually* Daily* Shaft Seal Leak Logged Manually* Daily* Water Piping Leak Logged Manually* Daily* Lube Oil Leak Logged Manually* Daily* Lube Oil Discoloration Logged Manually* Daily* Lube Oil Reservoir Level Logged Manually** Monthly*** Lube Oil Supply Pressure Logged Manually** Monthly*** Lube Oil Supply Temperature Logged Manually** Monthly*** Lube Oil Filter Differential Pressure (Delta P) Logged Manually** Monthly*** Unit MW Output Logged Manually** Monthly Pump Suction Pressure Logged Manually** Monthly*** Pump Discharge Pressure Logged Manually** Monthly Pump Suction Temperature Logged Manually** Monthly Pump Rpm Logged Manually** Monthly Bearing Housing Vibration-Inboard Horizontal Data Collector Monthly Bearing Housing Vibration-Inboard Vertical Data Collector Monthly Bearing Housing Vibration-Outboard Horizontal Data Collector Monthly Bearing Housing Vibration-Outboard Vertical Data Collector Monthly Bearing Housing Vibration-Axial Data Collector Monthly Bearing Housing Vibration-Outboard Vertical Data Collector Monthly Bearing Housing Vibration-Axial Data Collector Monthly

* If observed. ** If not recorded on control room computer or charts. *** These items should have control room alarm and/or shutdown.

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M any main feedwater pumps are instrumented for the following:

Data Item Method Frequency Journal Bearing Temperature-Inboard Logged Manually** M onthly*** Journal Bearing Temperature-Outboard Logged Manually** M onthly*** Thrust Bearing Temperature Logged Manually** Monthly*** Shaft Vibration-Inboard Logged Manually** Monthly*** Shaft Vibration-Outboard Radial Logged Manually** Monthly*** Shaft Vibration-Axial Logged Manually** Monthly***

Desirable data includes:

Coupling Alignment Logged Manually Monthly Torque Input Logged Manually** Monthly Pump Flow Logged Manually** Monthly*** Turning Gear Motor Current Logged Manually** On standby Suction Strainer Differential Pressure (Delta P) Logged Manually Monthly

Desirable data for controlled leakage seals includes:

Data Item Method Frequency Seal Injection Flow-Inboard Logged Manually Monthly*** Seal Injection Flow-Outboard Logged Manually Monthly*** Seal Drain Temperature-Inboard Logged Manually Monthly Seal Drain Temperature-Outboard Logged Manually Monthly

D esirable D ata for M ultistage Pumps: Balance Device Leakoff Flow Logged Manually Monthly Recirculation Line Flow Logged Manually Monthly

* If observed. ** If not recorded on control room computer or charts. *** These items should have control room alarm and/or shutdown.

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Affordable flow meters that can measure the flow to each pump and the flow to each auxiliary water line (seal injection, balance device leakoff, or recirculation line) are available. The transducers are clamped to the outside of the pipe; they are easy to install and do not restrict flow. They are available for the full range of feedwater temperatures, and for radioactive water. One suitable flow measuring system is called a portable ultrasonic transit-time flow meter. Maximum accuracy is achieved with specified lengths of straight run pipe upstream and downstream of the transducers. These devices are useful, even if the specified straight run of pipe is not available, as long as the readings are reproducible. It is the trends that are important and not necessarily the absolute values.

If a portable thermographic camera is available, it should be used to monitor the temperature of the flex points on each end of the driver-to-pump coupling. Excessive or increasing temperatures can identify a coupling problem, generally misalignment. Reference 30 states that portable thermographic cameras that can transfer data to a personal computer are available but are more expensive than portable vibration detectors. This data should be collected monthly, especially just before a shutdown and just after a startup (allowing enough run time for the coupling to reach equilibrium temperature).

4.2.2 Data Collection

The data collected for predictive maintenance must be consistent as follows: • All instrumentation must be calibrated on a regular basis. Annual calibration of the data collector and quarterly calibration of the portable sensors are generally recommended.

•. For hand-held vibration data collection, bearing housings must be marked m such a way that readings are always taken in the same places and in the same sequence. Quick disconnect or magnetic mounting of sensors is preferred. If hand-held sensors must be used, data should always be collected by the same person. • Observations of abnormal noise, water leaks, lube oil leaks, or lube oil discoloration must be logged into the predictive maintenance data collection system in a consistent m anner.

The frequency of collection listed for each data item m Section 4.2.1 is suggested for the initial phase of a predictive maintenance program. As the program matures, the frequencies should be adjusted to suit the conditions and requirements of the unit. If one or more adverse trends are observed, data collection should be accelerated. Adverse trends are listed in Section 4.3.1.

Frequent data collection is advisable when an idle pump is put on line, especially if maintenance was performed on the pump, its driver, or the connected piping while the pump was idle.

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In addition to full-load operation, the unit will operate at one or more part load conditions such as those caused by equipment outages. For each unit load condition, steady- state data should be collected for: Unit M W Output Pump Suction Pressure Pump Discharge Pressure Pump Suction Temperature Pump Rpm Pump Torque Input (If M easured) Pump Flow (If M easured) Bearing Housing Vibration-Inboard Horizontal Bearing Housing Vibration-Inboard Vertical Bearing Housing Vibration-Outboard Horizontal Bearing Housing Vibration-Outboard Vertical Bearing Housing Vibration-Axial Shaft Vibration-Inboard Radial (If M easured) Shaft Vibration-Outboard Radial (If Measured) Shaft Vibration-A Axial (If Measured) Although listed as one data item, radial shaft vibration is measured (with proximity probes) at two points, 90° apart at the same axial location on the shaft. The axial location is between the radial bearing and the shaft seal.

Bearing housing vibration is measured with accelerometers (converted to velocity) or velocimeters. Accelerometers can measure higher frequency vibrations.

Quarterly analysis of the lube oil for water and wear particles is suggested. This is generally done off site. A written report of results is prepared. Such analyses should also be made one to two months prior to a scheduled outage of the unit and after two to four weeks operation of each pump after an outage for maintenance (whether scheduled or unscheduled). The logic of the suggested additional analysis is: • If an analysis discloses a problem one to two months prior to a scheduled outage, appropriate corrective action can be scheduled for the outage. • Laboratory oil analysis will generally disclose any startup or maintenance-induced bearing or lube oil system degradation after two weeks of operation.

Increasing the lube oil reservoir level almost always indicates leakage of cooling water into the lube oil and should trigger (as a minimum) an immediate oil analysis.

As is true of the suggested frequencies of all data collection, lube oil analysis frequencies should be adjusted to suit the conditions and requirements of the unit as the predictive m aintenance program m atures.

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4.2.3 Data Storage

Collected data such as observations of abnormal noise, leaks, and oil discoloration as well as numerical values for pressure, temperature, rpm, oil level, and vibration readings should be entered into a software program for predictive maintenance.

One to two years of data is ample for normal predictive maintenance. Large-capacity hard disks and backup storage devices are inexpensive, so ample quantities of baseline data (three fuel cycles minimum) should be stored for reference to facilitate diagnostic analysis if serious problems develop. Vibration data storage might require significant disk space. It can be selectively erased from the active hard disk, provided that baseline data recordings no more than six months apart are saved.

4.2.4 Data Processing

As a minimum, the predictive maintenance software package must: • Produce reports of observed abnormal noise, leaks, and oil discoloration. • Plot lube oil levels, pressures and temperatures, bearing temperatures, and other directly observed parameters as a function of time. • Store vibration-spectrum analyses of bearing housing vibration, and plot and tabulate the results. • Perform data reduction calculations on available pump performance parameters, and plot and tabulate the results.

Vibration-Spectrum Analysis

Vibration-based predictive maintenance systems commonly use fast Fourier transform (FFT) ) vibration-spectrum analysis to compute the filtered vibration magnitudes over the full range of frequencies required for predictive maintenance. The x is defined as the rotational frequency. For example, for a 3,600 rpm machine x is 3,600 cycles per minute, 60 cycles per second, or 60 hertz (Hz). The z is defined as the number of vanes in an impeller, therefore, zx is vane passing frequency. For example, a 5 vane impeller at 3,600 rpm has a vane passing frequency of 300 Hz. Vane passing frequency is independent of the number of diffuser vanes or volute lips. Vibration frequencies of particular interest are discussed m Section 4.3.3.

Bearing housing vibration magnitude is generally expressed as velocity m inches per second. Shaft vibration magnitude is generally expressed as amplitude in mils peak to peak (1 mil is 0.001 inches).

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Pump Performance

Ideally, data available for calculating pump performance for predictive maintenance purposes includes: Pump suction pressure Pump discharge pressure Pump suction temperature Pump rpm Pump flow Pump torque input Pump motor driver kilowatt (kw) input This is enough data to calculate pump efficiency:

Q (P - P ) E = 2 2 1 p 1714xHP

E pump efficiency p =

Q = pump flow, gpm

discharge pressure P2 =

= suction pressure P1

HP = horsepower, Txrpm/5250

T = torque, foot pounds

HP = horsepower, motor kw x Em /0.746

Em = efficiency of motor drive train

Efficiency of the motor drive train is the product of the efficiencies of all machines m the drive train (motor, gear box, fluid drive, etc.).

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Q1 x rpm2 Q2 = rpm 1

Q cons tan t (corrected ) flow to be trended 2 =

= measured flow Q1

rpm = constant rim to be trended 2

rpm1 = measured flow

The above is a limited-purpose field test (as opposed to factory test) of the pump. A discussion of pump field tests versus factory tests is in Reference 54.

The above equation is not 100% accurate for all centrifugal pumps but is practical for this purpose.

Pump efficiency varies with pump flow, so it should be trended versus time at constant flow. If the rpm is not constant, the flow must be adjusted:

Example:

April May June July Q 10,000 10,100 9,900 10,050 1 rpm 5,000 5,0504,940 5,020 1 Q 10,000 10,000 1,0020 9,960 2 rpm 5,000 5,000 5,000 5,000 2 E 0.8160.820 0.810 0.800 p In this idealized example, the M ay, June, and July flows were corrected to the April speed. The variations in corrected flow are small. No adverse performance (efficiency) trend is detected until July. There may or may not be a problem; however, it would be prudent to take more frequent data on this pump, perhaps every week or two. A very close look at trends in vibration and in balance device leakoff flow, recirculation line flow, controlled leakage, seal injection flow, and drain temperature (if measured) is now in order.

For a less ideal example, only the following data is available: Unit M W output Pump suction pressure Pump discharge pressure Pump suction temperature Pump rpm We cannot measure pump flow, therefore, unit M W output must be used as a substitute. This means that data useful to evaluate pump performance can be collected only

4-10 EPRI Licensed Material Main Feedwater Pump Maintenance Guide during the (hopefully infrequent) periods when only one pump is running. When two or more pumps are operating in parallel, it is not prudent to assume that they each are pumping an identical flow. Although pump efficiency cannot be calculated, the pump differential pressure can be calculated by subtracting suction pressure from discharge pressure.

Another drawback is that pump suction temperature and, therefore, the density (or specific gravity) of the water must be in a narrow range for the times of data collection.

Given all of the above constraints, there can be useful trending of pump rpm versus time at the constant unit MW output while operating with one feed pump. Increasing rpm with time is an adverse trend.

The two examples illustrate the trending of pump performance with maximum and minimum measured data. Intermediate cases also exist, and other trending parameters can be devised to maximize the usefulness of available information.

4.3 Predictive Maintenance The following predictive maintenance parameters should be available as computer output.

Reports of: Abnormal noise Shaft seal leaks Water piping leaks Lube oil leaks Lube oil discoloration Water in lube oil Particulates in lube oil Plots of: Lube oil pressure versus time Lube oil temperature versus time Lube oil level versus time Lube oil filter delta P versus time Bearing housing vibration at five locations at below lx versus time Bearing housing vibration at five locations at lx versus time Bearing housing vibration at five locations at 2x versus time Bearing housing vibration at five locations at zx versus time Bearing housing vibration at five locations, other versus time Pump output at various unit M w versus time M any pumps are instrumented for and can provide data and plots of: Journal bearing temperature-inboard versus time Journal bearing temperature-outboard versus time Thrust bearing temperature versus time Shaft vibration at five locations at below lx versus time

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Shaft vibration at five locations at lx versus time Shaft vibration at five locations at 2x versus time Shaft vibration at five locations at zx versus time Shaft vibration at five locations, other versus time Other desirable computer outputs include plots of: Coupling flex point temperature versus time Coupling misalignment versus time Pump output at various unit M w using torque data versus time Pump output at various unit Mw using flow data versus time Pump output at various unit M w using torque and flow data versus time Turning gear motor current on standby versus time Suction strainer delta P at various flows (or Mw) versus time Seal injection flows versus time Seal drain temperatures versus time Balance device leakoff flow versus time Recirculation line flow versus time

4.3.1 Parameter Trending

Plots of measured and computed parameters as a function of time constitute trending, which is the primary tool of predictive maintenance. Trending is meant for the early detection of adverse trends.

If these conditions are increasing, they are adverse trends: Abnormal noise Shaft seal leakage Water piping leakage Lube oil leakage Lube oil level Lube oil discoloration Water in lube oil Particulates m lube oil Lube oil temperature Lube oil filter delta P Bearing housing vibration Bearing temperature Shaft vibration Coupling flex point temperature Coupling misalignment Turning gear motor current Suction strainer delta P Seal injection flow Seal drain temperature Balance device leakoff flow Recirculation line flow

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If these conditions are decreasing, they are adverse trends: Lube oil pressure Lube oil level Pump output

4.3.2 Limits for Monitored Parameters

One purpose of predictive maintenance is to minimize alarms as well as unscheduled pump outages.

Control room alarms can monitor: Lube oil reservoir level (high and low) Lube oil supply pressure Lube oil filter delta P Pump suction pressure Journal bearing temperatures Thrust bearing temperatures Unfiltered radial shaft and/or bearing housing vibration Unfiltered axial shaft and/or bearing housing vibration Pump flow Seal injection flow or delta P Predictive maintenance limits for parameters with control room alarms should be set well below alarm settings so that appropriate maintenance can be scheduled to avoid alarms, if possible.

Control room alarm vibration signals are generally unfiltered, but predictive maintenance involves filtered vibration-spectrum analysis.

API Standard 610 (Reference 33) allows 0.4 inches per second (peak) unfiltered, and 0.3 inches per second (peak) filtered, vibration levels measured on the shaft for sleeve-bearing pumps. These values are allowed during factory performance tests at rated speed (rpm) and flow, plus or minus 10%. Allowable values for an installed pump or for operation at off-design flow or speed are not given.

ANSI/Hydraulic Institute (Reference 35) allows 0.31 inches per second (peak) unfiltered, and 0.24 inches per second (peak) filtered, vibration levels measured on the pump bearing housings for sleeve-bearing pumps. These values are allowed when the pump is installed with proper piping, foundation, NPSH, and coupling alignment, then operated at rated speed and flow, plus or minus 10%. Allowable values for operation at off- design flow or speed are not given.

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For reference:

V = 3.14 fD

V = velocity - inches per second, peak

f = frequency-cycles per second

D = displacement - inches, peak to peak

For example, the displacement (amplitude) of 1x bearing housing vibration at a rotating speed of 5,000 rpm and 0.3 inches per second is:

0.3x60 D = 3.14 x 5,000

V D = 3.14 x f

D = 0.0011 inches or 1.1 mils peak to peak

Published information relating to vibration limits for large rotating machinery includes a disclaimer, and this guide is no exception. Each installation is unique. Limits are based on operating experience, using only available published numbers.

4.3.3 Interpretation of Monitored Parameters

Filtered Vibration Correlations

Diagnostics and root-cause analysis are discussed m Section 4.5 of this guide. These are highly technical and beyond the scope of predictive maintenance. However, when an adverse trend is detected and verified, it is helpful if maintenance personnel have some preliminary indication of where the problem might be developing.

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The following filtered vibration correlations are generally accepted:

Location Frequency Suspected Problem Bearing housing-inboard or outboard parts below 1x Journal bearing instability, worn rotor, or loose Bearing housing-axial below 1x Gap A Bearing housing-inboardlx Rotor unbalance, rotor wear, rotor critical speed, bent shaft, cracked shaft, stator resonance, coupling misaligmnent, or unbalance Bearing housing-outboard lx Rotor unbalance, rotor wear, rotor critical speed,bent shaft, cracked shaft, or stator resonance Bearing housing-axial lx Stator resonance Bearing housing-inboard 2x Coupling misalignment or bent or cracked shaft Bearing housing-outboard 2x Bent or cracked shaft Bearing housing-axial 2x Bent or cracked shaft or coupling rnisalignment Bearing housing-inboard zx Piping resonance, crossover resonance, or Gap B Bearing housing-inboard, outboard, or axial Other Stator resonance

Shaft vibrations follow the same general pattern as bearing housing vibrations. They tend to be more sensitive to rotor vibration and less sensitive to stator resonance, piping resonance, and crossover resonance (especially in the axial direction). This is logical.

Shaft rotating frequency is x and the number of vanes in the impellers is z. Rotating frequency is 1x and vane-passing frequency is zx.

Journal bearing instability, also called oil whirl, half-frequency whirl, or oil whip, is generally observed at 0.42x to 0.48x.

Gap A problems are generally at frequencies below 0.25x. Axial motion at these low frequencies (resulting from Gap A problems) is generally not detected on bearing housings but is evident as shaft vibrations measured with proximeters.

Measurement of a broad high-frequency energy band, e.g. 1,000 to 10,000 Hz in G's, provides a sensitive monitor of rub and/or cavitation. Rubs are detected as high-fre-

4-15 EPRI Licensed Material Nuclear Maintenance Applications Center quency bursts that repeat each revolution. Cavitation is also high frequency, but random. M odern accelerometers can detect bearing housing vibration at these frequencies.

The term stator resonance is broad. Harmful resonant frequencies have been found in splitters, bearing brackets, bearing housings, pump pedestals, and foundations. It is possible for a pump to have a stator or piping resonance while an identical pump m the same unit does not.

Filtered Vibration Examples

There are many ways to present filtered vibration data. The examples presented here have been selected because they help illustrate predictive maintenance interpretations.

Figures 4-2 and 4-3 are plots of the filtered shaft vibration of the same pump. Note that Figure 4-2 has a scale of 0 to 1.0 mils and Figure 4-3 a scale of 0 to 2.0 mils. Figure 4-2 shows that the shaft vibration on the outboard (thrust bearing) end is less than 0.3 mils at all frequencies. However, Figure 4-3 shows that the lx shaft vibration on the inboard (coupling) end is almost 0.9 mils, a possible symptom of coupling misalignment or unbalance.

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Figure 4-2 Pump Shaft Vibration (Outboard)

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Nuclear Maintenance Applications Center

Figure 4-3 Pum p Shaft V ibration (Inboard)

Figure 4-4 is a cascade plot of the filtered shaft vibration at the inboard (coupling) end of the coast-down (spin-down) of a pump with a constant-speed, 3,350 rpm driver. Vibration is minimal except at lx and 5x. The pump has 5 vane impellers, therefore, 5x is zx.

If this were the only lx vibration data available, suspected problems could include rotor unbalance, rotor wear, rotor critical speed, bent or cracked shaft, stator resonance, coupling misalignment, or unbalance. If a similar plot on the outboard (thrust bearing) end of the pump showed much lower 1x vibration, then coupling misalignment or unbalance would be more likely. The 1x vibration is only about 0.5 mils, which is not enough to cause immediate concern.

Vibration at 5x (zx) increases rapidly between 3,100 rpm and 3,260 rpm, about 1.2 mils at operating speed vane passing frequency (3,350 x 5/60= 279 Hz ). This is a filtered shaft vibration velocity (see Section 4.3.2) of 3.14 x 279x .0012= 1.1 inches per second, more than three times the value allowed by API Standard 610 (Reference 33). Suspected problems include piping resonance, crossover resonance, or a Gap B problem, whereby a root-cause analysis is indicated.

4-18 EPRI Licensed Material Ma in Fee dwa ter Pu mp M ain ten ance G uide

Figure 4-4 Cascade Plot of Pump Shaft Radial Vibration (Inboard)

4.3.4 Limitations of Predictive Maintenance

Data Limitations Limitations on the reliability of data and computed parameters result from instrument inaccuracies and from natural periodic changes in measured parameters.

Predictive maintenance procedures should state the accuracy and calibration frequency of all instrumentation so that instrument inaccuracies will be highlighted for predictive maintenance personnel evaluating trended data.

Seasonal trends, such as the influence of ambient temperature and cooling water temperature, must be accounted for as they might influence such parameters as lube oil temperature, bearing temperature, and seal drain temperature. Pump performance parameters vary with unit load. Their interpretation is discussed in Section 4.2.4 under Pump Performance.

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Limitations on Diagnostics

The purpose of predictive maintenance is to minimize maintenance costs. Major maintenance costs can result from the cost of replacement power for unscheduled outages. Predictive maintenance should focus on cost-effective scheduling and execution of m aintenance activities.

In-depth diagnostics and root-cause analysis should be a separate function performed at the request of maintenance supervisors by specialists who are not involved in maintenance activities. These individuals must be responsive to maintenance personnel, and they must supply analysis and procedures to cure ongoing maintenance problems.

In-depth diagnostics and root-cause analysis are indicated if a problem is ongoing and costly, especially if it involves these suspected conditions: M etallurgical problem Non-metallic materials problem Journal bearing instability Gap A problem Rotor critical speed Stator resonance Piping resonance Crossover resonance Gap B problem The term stator resonance is broad. Harmful resonant frequencies have been found in splitters, bearing brackets, bearing housings, pump pedestals, and foundations. It is possible for a pump to have a stator or piping resonance while an identical pump In the same unit does not have a resonance.

4.4 Preventive Maintenance Preventive maintenance is regularly scheduled periodic maintenance performed during refueling outages or other convenient times.

The following preventive maintenance tasks should be performed at the end of every fuel cycle: • Drain and clean lubrication oil systems • Inspect bearings, journals, and thrust disks • Replace lubrication oil • Inspect gear-type or disk/diaphragm-type couplings • Grease gear-type couplings • Install rebuilt mechanical seals (if applicable) • Install rebuilt floating ring seals (if applicable)

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Floating ring seal construction should be replaced with controlled leakage injection-type bushings or mechanical seals. If operating experience in the first fuel cycle in which the seals were installed indicates no expected problems m the second cycle, mechanical seals should last for two or more fuel cycles (See Sections 2.16 through 2.22 of this guide).

Serious consideration should be given to the replacing the gear-type or disk-type couplings with diaphragm-type-couplings (See Sections 2.23 and 5.3.2 of this guide and Section 2.13 of Reference 17).

4.4.1 Major Maintenance

Specific periodic scheduling of major maintenance is difficult.. The NMAC Survey (Section 3.2.2 of this guide) revealed that major teardown frequencies vary from once each refueling outage to as often as every 10 years. Manufacturers generally recom- mend a complete teardown only when indicated by condition monitoring and predictive maintenance programs. Table 4-1 is reproduced from Reference 17.

Table 4-1 A cceptable Life Expectancy of Pumps

Component Life Expectancy, Years Pump overall lifetim e 35 External components Foundations/piping 35 Hydraulic components Diffuser/volutes-impellers 10 Impellers (with exchangeable wear rings) 35 W ear-rings (stationary) 3 Rotor components Shaft 35 Balance device 6 Seals (labyrinth) 10 Seals (other) 6 Journal bearings 10 Thrust bearings 15 Co up ling (flu id ) 15 Coupling (dry) 35

The reference states that if current technology is applied, the listed life expectancies for components of large boiler feedwater pumps used on cycling service are achievable. Application of these life expectancies to nuclear main feed water pumps is conservative because cycling service for nuclear plants is rare, and nuclear main feedwater pumps typically have shorter and stiffer shafts than fossil boiler feedwater pumps.

Except for the stationary wear rings, Table 4-1 indicates that boiler feedwater pumps should run six years without teardown. If the pumps are not cycled, nuclear main feedwater pump stationary wear rings will last six years because of the stiffer shafts.

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Also, wear ring material can be upgraded. Recommendation 21 m Reference 18 suggests that the wear ring material be changed to reduce impeller hub and wear ring galling and wear. AISI 420F or CA40F have a higher sulfur content that reduces galling and wearing and extends pump life. When ordering replacement parts from the manufacturer, specify the material needed.

It should be added to these recommendations that wear ring hardnesses, and especially the difference m hardness between rotating and stationary parts, must be maintained.

Gaskets and O-rings can have recommended lives of less than six years. The preponder- ance of literature on the subject of bolted joints, including References 16 and 19, presup- pose that if the joint does not leak, it should not be disturbed until such time as it does leak. Pumps with axial (horizontal) split main cases with thin (generally 0.015-inch thick) gaskets between the main casing flanges have unique problems. (See Radially Split Main Casing Joints in Section 5.7.1 of this guide.)

It is recommended that scheduling of a complete teardown for major maintenance be based on a predictive maintenance program rather than on a fixed time period. Six years between teardowns is a realistic goal.

As predictive maintenance becomes more common, the preventive maintenance schedules become more logical and the distinction between predictive and preventive maintenance becomes less clear. The objectives for each are identical.

4.5 Diagnostics and Root-Cause Analysis Many feed water pump problems that existed when the plants were built have been diagnosed and corrected. Failure analysis and evaluation has been an integral part of the nuclear industry and if often required by regulations.

Diagnostics and root-cause analysis are important, but often not required. The infre- quently used computer programs, specialized engineering and metallurgical skills, and diagnostic instrumentation are not required for successful predictive maintenance. Central engineering groups, pump manufacturers, architect/engineering firms, centrifugal pump engineering consultants, and metallurgical failure analysis specialists are all potential sources of talent.

A root-cause analysis involves some or all of the following: M etallographic analysis Scanning electron microscopy Metallurgical physical and/or chemical analysis Hardness testing E xtensive steady-state vibration testing Coast-down or variable speed vibration testing Pressure pulsation measurement Operating deflection shape testing and analysis Field modal analysis Pump and piping acoustic analysis 4-22 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Suction recirculation analysis Cavitation analysis Rotor dynam ic analysis Preventive maintenance data is helpful in planning diagnostics and root-cause analysis, and preventive maintenance personnel should provide guidance to specialists as part of a team effort.

4.5.1 Case Histories

Controlled Leakage Shaft Seal Problem Caused by Rotor Vibration

Early in 1990, a PWR unit replaced floating-ring-type shaft seals with a controlled leakage type with unsatisfactory results. Seal failures typically occurred within two weeks of a transient, usually at plant startup.

Evaluations included: Visual analysis Diametrical measurements Metallographic analysis Scanning electron microscopy Chemical analysis Hardness testing Vibration analysis It was concluded that the root cause of the seal failures was excessive radial pump rotor vibration at reduced rpm. The vibration level was acceptable at normal pump operating speed of 5,000 rpm, but it increased dramatically at 4,000 rpm. The amplitude of shaft vibration at 4,000 rpm was high enough to create excessive seal clearances in a short period of time.

M echanical Seal Problem Caused by Crossover Resonance

A BWR unit experienced high bearing housing vibration levels and frequent mechanical seal failures since initial startup. Several attempts at root-cause analysis failed to diagnose the problem.

Finally, extensive evaluations included: Vibration testing, including coast-down data Pressure pulsation testing, including coast-down data Operating deflection shape analysis Modal analysis for structural natural frequencies and mode shapes Pump acoustic analysis Piping acoustic analysis Suction recirculation analysis C avitation analysis Rotor dynam ic analysis

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It was concluded that there is a resonance (strong acoustic standing wave) in the long crossover of this two-stage pump at the vane passing frequency of the fixed operating speed of 3,355 rpm. This resonance causes pressure pulsations that produce extreme vibrations of both rotating and stationary parts. These vibrations are the suspected cause of frequent seal failures.

Lube Oil Piping and Tilting-Pad Radial Bearing Failures

A PWR unit experienced pump radial bearing failures, high bearing temperatures, and intermittent high levels of drive turbine rotor vibration from initial startup. Replace- ment of carbon steel lube oil piping with stainless steel cured the pump radial bearing failures, but pump bearing temperatures were still high.

Rotor-dynamic analysis of both turbine and pump confirmed oil-whirl instability m the turbine bearings and also resulted m the recommendation to replace both turbine and pump plain journal radial bearings with a tilting-pad type to increase rotor stability and lower pump bearing temperature. Improved pump-to-turbine alignment procedures w ere im plem ented at the sam e tim e.

Subsequent operation has indicated that the modifications have cured the problems.

Team Effort Solves Mechanical Seal Problems

Two units of a BWR plant experienced an unusually high number of mechanical seal failures in 1991 and early 1992. A task force was formed with representatives from site engineering, maintenance, operations, mechanical and structural engineering, and the technical staff. Input was obtained from the pump and seal manufacturers, General Electric, and pump/seal technical consultants.

Evaluations included: Reactor feed pump system and pump and seal design reviews System modification histories reviewed Reactor feed pump testing program Instrumentation of pump and seals Failure analysis of seal components Audit of incoming seal parts from the OEM Review of maintenance and operations procedures Possible contributing factors included: Operation of pumps on minimum flow (with high vibration) for extended periods of time Lack of vibration instrumentation Seal parts furnished by OEM not up to OEM standards Heat checking of tungsten carbide seal faces Inadequate venting of seal piping

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Insufficient seal cooling Non-cartridge design requiring precise field assembly Seal designed for 3,600 rpm, operated at 4,500 rpm Changes implemented included: Increased Gap B reduced minimum flow vibration Silicone carbide replaced tungsten carbide as seal face material Seal piping was properly vented Larger seal coolers and seal piping (actually tubing) size was used Cartridge seal design reduced maintenance installation time and problems Seal was designed for 4,500 rpm Condensate injection to the seal cooling system was eliminated Improvements were made to pump maintenance and operating procedures Since the above changes were implemented in 1992 and 1993, there have been no mechanical seal failures on the six feed pumps in this plant.

In this case, there were many possible contributing factors. It was economically possible and prudent to implement many corrective actions at one time. This was practical because there were probably several contributing factors rather than one root cause of the m ultiple failures.

Loose Coupling Causes Cracked Shaft

Operators of a PWR unit observed excessive vibrations m a feedwater pump and shut it down. Inspection revealed a transverse crack at the end of the coupling hub, extending 330° around the shaft circumference.

An outside metallurgical laboratory performed the following: Visual examination Scanning electron microscopy Quantitative chemical analysis Deep-etch exam ination Metallographic examination H ardness test Tensile test No abnormalities were found m the type 410 shaft material. The root cause of the failure was found to be a high-cycle fatigue fracture resulting from the loose coupling.

Improper Hardness and Impeller Failure

A PWR unit experienced increased vibration m a two-stage feedwater pump approxi- mately two years after a complete overhaul. Unfiltered horizontal bearing housing vibration about 0.1 inch per second following overhaul increased to 0.3 to 0.4 inch per second on restart after a scheduled outage. Filtered vibration spectra showed that most of the increase was at lx. With time, the unfiltered vibration reached as much as 0.9 inch

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per second. Vibration on the outboard (thrust bearing) housing was higher than that on the inboard (coupling end) housing, indicating that coupling misalignment was probably not the problem. The pump was secured and dismantled. Major damage was observed in the second-stage impeller. It was returned to the pump manufacturer, and then sent to an outside metallurgical/failure analysis company.

There was severe damage in the suction eye area of the impeller. Two of the seven vanes in the suction eye had pieces missing from the leading edges. Another vane adjacent to one of the broken vanes had a crack. There was a crack in the front shroud that extended from the outside diameter across the shroud and ring fit and into the suction eye of the impeller. Severe damage on the two suction vanes showed evidence of a projectile that possibly caused this damage. The other five vanes exhibited little or no damage that could be associated with a projectile entering the suction eye.

Analysis included: Scanning electron microscopy of the fracture surfaces M etallographic examination of the following areas: Uncracked vane 180° from broken vanes Both broken vanes Corners of broken vanes Body of casting Crack on shroud Hardness checks Chem ical analysis The cause of the failed vanes was found to be corrosion fatigue, with pitting corrosion initiating the fatigue cracking. The corrosion resistance of the base metal was impaired by less than optimum microstructure. The hardness ratio in the failed regions was 36-38, as opposed 28-32 specified by the vendors. The higher hardness degrades the corrosion resistance and also renders the impeller susceptible to corrosion fatigue. The failure analysis report concluded that there should have been another ultimate cause, because the failures were not in high-stress locations.

The failed impeller was cast from ASTM A743 CA15 material that was standard at the time this pum p was m anufactured. T his material is difficult to cast, weld-repair, and heat treat to specified hardness. The current industry standard is ASTM A743 CA6NM , which is much easier to work with.

As the four pumps in the two units received routine rotating element reconditioning, plant personnel have supervised rebuilding of rotating elements by the vendors. Impel- lers have been reheat treated to the proper hardness (with some slightly lower hardnesses accepted), checked for hardness, and inspected using magnetic particle or liquid penetrant prior to assembly. All new replacement impellers are specified as ASTM A743 CA6NM.

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M echanical Seal Upgrades Prolong Life

Two units of a BWR plant experienced an unacceptable number of mechanical seal failures from 1984 through 1989. A main feedwater pump task force was formed with representatives from engineering, maintenance, training, and the technical staff. The group focused primarily on mechanical seal failures.

Existing seals were of cartridge design and had a closed-loop cooling system (see Sec- tions 2.21 and 2.22).

Changes implemented included: Seal piping was properly vented. Silicon carbide replaced tungsten carbide as the seal rotating face material. Improved pump alignment reduced vibration levels. Quality of parts by the seal vendor was improved. Implementation of the above changes began in 1992. As of December 1995, no upgraded seals had failed. Seals removed and inspected have shown almost no wear of the seal faces.

Although definitive operating data is incomplete, based on experience to date, it is believed that these seals will have a life expectancy of three to five years.

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MAIN FEEDWATER PUMP MAINTENANCE PRACTICES

5.1 Introduction

Good maintenance practices begin with a comprehensive program that includes proper planning, training, data recording, and record keeping. Qualified, detail-oriented personnel are absolutely essential.

This section discusses improving pump reliability through procedural enhancements and proper maintenance techniques. The discussion progresses from pump shutdown through disassembly, inspection, and reassembly of main feedwater pumps.

5.2 Maintenance Programs

Even when engineering and operations departments perform their duties correctly, main feedwater pump failures can continue to occur. A site's maintenance practices directly affect the reliability of the main feedwater pump, and the entire program should be reviewed regularly.

. 5.2.1 Experience and Training

A sound pump maintenance program begins with knowledge of the specific pump. Initial repairs to a main feedwater pump should be made by a field service technical director provided by the manufacturer. In many cases, to save money, subsequent repairs are made by inhouse personnel that have an adequate level of experience to repair the pump on site.

Maintenance training programs can prove beneficial m improving pump reliability. Training programs designed specifically for maintaining main feedwater pumps are recommended, and training supplied by the pump's OEM is essential. Using a mockup or training pump helps achieve proper results. An additional training course should be conducted by an outside expert who has experience with other manufacturers of similar main feedwater pumps. This provides an objective outside viewpoint and might reveal valuable experiences with other types of pumps.

5.2.2 Procedures and Worksheets

Drawings or sketches, and worksheets on a procedure help record an accurate, verified account of the data taken during as-found readings. These tools are useful for comparisons to as-built readings. The drawings and worksheets should be used liberally throughout

5-1 EPRI Licensed Material Nuclear Maintenance Applications Center the procedure. Copies of these records of the pump's general condition should be kept by the pump engineer m addition to archived procedures and work orders.

Detail drawings of pump parts should be used to record measurements and tolerances. If possible, parts should be measured and drawings made based on a survey of several examples of the same part, Ideally, this information should come from the manufacturer mentioned in the operations and maintenance manual provided with the pump. (See Section 5.2.5, Vendor Technical Manuals.)

Worksheets can be customized for a particular pump or procedural format. The generated work order or service request for repair work should contain the name of the responsible engineer or technical director.

Table 5-1 are examples of worksheets for recording as-found or as-built data for a Worthington 24WGID pump.

5-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide Table 5-1 Pum p W orksheets

DIAMETRAL RUNNING CLEARANCES Part Description Recommended As Found / As Built As Found / Clearance As Built Clearance

Wearing rings and .014 - .017 Coupling end wear Coupling end impeller = impeller hubs ring l.D. hub O.D. (rings installed) = = Thrust end wear ring Thrust end impeller = l.D . hub O.D. = = Breakdown bushing .010 - .012 Coupling end bushing Coupling end shaft = and shaft sleeve l.D . sleeve O.D. = = Thrust end bush. l.D. Thrust end shaft = = sleeve O.D. =

Assembled bearing .005 - .007 Coupling end brg l.D. Coupling end shaft = and shaft journal = journal O.D. = Thrust end brg l.D. Thrust end shaft = journal O .D. = =

Thrust bearing seal ring Thrust brg. seal l.D. Shaft O.D. = and shaft or thrust = = collar nut Thrust collar nut seal Thrust collar nut O.D. = ring l.D . = = Bearing seal and shaft .010 - .012 Coupling end outer brg Shaft seal journal O.D. = seal l.D. = =

Coupling end inner brg Shaft seal journal O.D. = seal l.D. = = Thrust end brg seal l.D. Shaft seal journal O.D. = = =

Seal cover and shaft .020 - .024 Coupling end seal Coupling end shaft = sleeve nut cover l.D. sleeve nut O.D. = = Thrust end seal Thrust end shaft = cover l.D. sleeve nut O.D. = =

Casing or end cover .050 - .070 Coupling end casing @ Coupling end pumping = and pumping pumping sleeve l.D. sleeve O.D. = = Thrust end cover @ Thrust end pumping = pumping sleeve l.D. sleeve O.D. = =

5-3 EPRI Licensed Material Nuclear Maintenance Applications Center Table 5-1 (cont.)

ROTOR MOVEMENTS Description Amount As Found As Built To tal end play of fully .250 - .380 = = = assembled pump (thrust shoes removed) F inal rotor position - .060 max. = = deviation from center = of end play Thrust bearing - total .01 0 - .012 = = = end play

Total vertical .006 - .0 1 2 = = = movement (measured Coupling End at either shaft sleeve nut with opposite end support by bearing) .006 - .012 = = = Thrust End DIAMETRAL FITS

Recom m ended Interference Stationary Parts As Found / As Built Fit Clearance

Im peller on Shaft Interference .004 Coupling end imp l.D. Coupling end shaft O.D. = - .003 = = Thrust end im p l.D . Thrust end shaft O.D. = = = Pumping sleeve on Interference .004 Coupling end pumping Coupling end shaft O .D . = shaft - .003 sleeve l.D. = = Thrust end pumping Thrust end shaft O.D. = sleeve l.D. = = Thrust collar on shaft Interference .002 Thrust collar l.D. Thrust collar journal = - .001 = O.D. = Shaft sleeve on shaft Clearance .001 - Coupling end shaft Coupling end shaft O.D. = .0025 sleeve l.D. = = Thrust End Shaft Thrust End Shaft O.D. = Sleeve l.D. = =

Shaft sleeve nut on Clearance .003 - Coupling end shaft Coupling end shaft O.D. = shaft .005 sleeve nut l.D. = = Thrust end shaft sleeve Thrust end shaft O.D. = nut l.D. = = Thrust collar nut on Clearance .001 - Thrust collar nut l.D. Shaft O.D. = shaft .003 = =

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Table 5-1 (cont.)

DIAMETRAL FITS

Stationary Parts Recommended Interference Fit As Found / As Built Clearance

Inboard wearing ring in Interference .009 Inboard wear ring fit Inboard wear ring O.D. = casing - .007 l.D . = = Outboard wearing ring Interference .009 O.B. wear ring fit l.D. O.B.wear ring O.D. = in diaphragm - .007 == Diaphragm in casing Clearance .002 - Casing l.D. Diaphragm O.D. = .005 = = End cover in casing Clearance .002 - Casing l.D. End cover O.D. = .006 = = Breakdown bushing or Clearance .001 - Bushing Fit in Casing Breakdown Bushing O.D = seal housing in Casing .003 l.D . = or end cover = Bushing fit in end cover Breakdown bushing = l.D O.D. = = Seal cover in casing or Clearance .002 - Seal cover casing fit Seal cover O.D. = end cover .006 l.D = = Seal cover end cover fit Seal cover O.D. = l.D . = = Bearing seal in Clearance .001 - Coupling end outer brg Coupling end outer brg. = bearing housing .003 housing seal fit l.l.D. seal O.D. = Coupling end inner brg Coupling end inner brg. housing seal fit l.D. seal O.D. = = = Thrust end brg housing Thrust end brg seal O.D. = seal fit l.D . = =

Bearing retainer in Clearance .001 - Coupling end brg Coupling end brg = bearing housing .003 housing l.D. retainer O.D. = = Thrust bnd brg retainer Thrust end brg housing = l.D . O.D. = =

5-5 EPRI Licensed Material Nuclear Maintenance Applications Center

Signoffs are another useful tool in pump inspections. The practice of having a responsible technical director or engineer with experience evaluating the condition and correctness of dimensions of all parts is recommended. The responsible technical director or engineer should be directly involved in all decisions and activities regarding the main feedwater pump.

Record keeping has changed drastically since the first nuclear stations came on line. Nearly all original manual record keeping is now done by computer, as is all planning and cost analysis. This has made the job of planning and scheduling, as well as failure analysis, much easier. In addition to this information, the pump engineer should keep accurate records of the pump maintenance history.

5.2.3 Predictive/Corrective M aintenance Interface

Predictive and preventive maintenance programs should interface directly with any contemplated or necessary corrective maintenance activities. To ensure that accurate records are available for all work performed, any work done under the predictive or preventive maintenance program should be recorded and correlated to corrective maintenance work. Tasks that might be performed under a predictive or preventive work order might not be revealed under corrective maintenance records.

5.2.4 Spare Parts Inventory

High inventory levels of spare parts are no longer acceptable. Parts inventories are continually being reduced. As a main feedwater pump becomes more reliable, the spare parts inventory requirements to support it decrease. Spare parts for these pumps might have long lead delivery times that need to be considered in maintenance planning. Parts that require frequent replacement or have long lead times should be kept on hand. Throttle bushings and sleeves or mechanical seals are spares that might require frequent replacement.

A fully assembled and dynamically balanced rotating assembly should be stocked. The cost is justified by shortened turnaround times, especially m those situations of unscheduled corrective maintenance.

Recommended spares by the OEM should be weighed against past usage. The following is a suggested spare parts inventory list:

1— R otating assembly, fully assem bled and dynam ically balanced 1—Shaft 1—Impeller 1—Set of case wear rings 1—Set of journal bearings 1— Thrust bearing assembly 1—Set of thrust shoes 1—Thrust collar 1—Set of throttle bushings and sleeves for every two pumps or

5-6 EPRI Licensed Material Main Feedwater Pump Maintenance Glide

2—Mechanical seals with sleeves for each pump 1—Coupling assembly with dynamic balance 1—Set of coupling bolts or hardware This list does not include hardw are such as set screw s, gaskets, or O-rings. O-rings and gaskets should be stored according to shelf-life requirements and ordered as the responsible pump engineer deems necessary.

All new or reconditioned parts should be thoroughly checked, either against drawings or old and new parts, to verify correctness prior to installation. The hardness of casing rings should be verified upon receipt. Sometimes the hardness is unacceptable for use in pumps. Although the manufacturer has tried to ensure the correct hardness, a ring with a low hardness value can still sometimes be found. These rings should not be installed in the pump.

Typically, casing rings are harder than impeller wear hubs. Check the manufacturer's literature for the correct hardness of the casing rings and impeller wear hubs (or impeller wear rings). Improper hardness differentials m this area can lead to outages that are otherwise preventable.

5.2.5 Vendor Technical Manuals

After an overhaul, main feedwater pumps might draw attention, because experience has shown that a unit's startup is the most critical time for main feedwater pump problems. When problems occur, technical and maintenance personnel generally look in the vendor's operation and maintenance manuals and associated drawings for information.

M any vendor manuals lack all of the information necessary to check and verify all parts in the pump. The dimensions, fits, clearances, and interferences furnished are usually limited to the major parts of the pump. D imensions and fits for other parts (for example, a shaft sleeve nut) might not be included. The manufacturer often considers information of this nature as proprietary to m aintain its aftermarket parts business.

In the nuclear industry, many parts are ordered from the manufacturer by purchase or quality-level requirements. If parts are obtained from other sources, consideration should be given to the manufacturer; parts should not be purchased on price alone for this critical service.

5.2.6 Feedwater System Cleanliness

Inspection of a pump might reveal damage to the bearings, impeller and casing rings, bushings and sleeves, or mechanical seals (if so equipped). The condensate and feedwater systems might contain foreign materials detrimental to the pump internals. The cleanliness of these pumps should be monitored closely. Valve maintenance (for example, upstream of the pump) should be performed to preclude contaminants from entering the system and potentially degrading the pump. Even with precautions, the feedwater system must be thoroughly cleaned when such maintenance is performed.

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5.2.7 Preventive Maintenance Schedules

Scheduled maintenance should be done so that only one pump is worked on during an outage. As discussed in Section 4.4 of this guide, six years between major overhauls is a realistic goal.

5.2.8 On-Line Maintenance

Power stations are continually looking for methods to reduce outage times. Performing pump maintenance while the unit is on line is one option being studied.

M ajor pump maintenance on line is very difficult because the pump cannot be totally isolated from the heat of an operating system. Minor pump maintenance may be performed with caution, but major repairs are risky at best. Trying to make dimensional and concentricity checks while the pump is partially hot is poor maintenance practice.

There are other pumps m the station that are more receptive to on-line maintenance than main feedwater pumps.

5.2.9 Lubrication System Contaminants

When performing bearing inspections, it is important not to destroy evidence of dirt or other contaminants in the lubrication system. This information is helpful in acquiring permission to perform corrective lubrication system cleanup.

5.3 Disassembly and Recording of As-Found Data As-found data is valuable in determining the amount of degradation that has occurred during a pump's run period. This data can be used to inform the maintenance depart- ment, the technical support or engineering section, and the pump manufacturer of the location and magnitude of wear or damage. The degradation might not be normal, in that case the information can be used when consulting with other users of this pump type and the manufacturer. The information also assists station personnel in analyzing whether or not improvements are being made in the maintenance practices.

As-found data should be taken as the pump is disassembled. Checks should be made in the following areas: Alignment Coupling End play Thrust bearing Thrust collar runout Rotor positions

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— Journal bearings — Oil seal condition — Shaft seal (throttle bushing, mechanical seal, floating rings) Hydraulic balancing device — Rotor — Case

5.3.1 Alignment

An as-found alignment should be performed prior to any disassembly. If the pump is turbine-driven, the use of the turning gear to rotate the pump is beneficial. The use of a small air motor or ratchet (3/8 or 1/2 drive) can usually be connected to the turning gear to give smooth continuous rotation of the turbine and the pump. If the pump is driven by an electric motor, it might be difficult to rotate the motor. The auxiliary motor-driven oil pump should be operating to ensure that no damage occurs to the bearing, and to ease in rotating the motor/pump or turbine/pump.

If the system is equipped with a hydraulic coupling or speed increaser/decreaser, uncoupling from the motor will aid in rotating the pump and gear. This can be done manually in most cases. Reverse dial indicator, or a laser system, is the most accurate method to acquire alignment data. The coupling spacer may be left in place to rotate both shafts together when using either method.

The reverse dial indicator method is shown In Figure 5-1. Due to the design of many feed pumps and drivers, the reverse-dial method necessitates the need for long indicator rods to span the distance between couplings. This can result in a large amount of sag that must be accounted for in all calculations for optimum alignment. Reverse-dial data worksheets are a must when performing this alignment check.

With laser shaft alignment tools (Figure 5-2), the pump is rotated 90° at a time to ascertain the alignment. Some lasers can take measurements at 90° intervals without stop- ping rotation of the shaft. Unlike reverse-dial indication, there is no sag. Newer laser alignment technology takes data uncoupled and rotates one shaft at a time, enabling the laser to read the opposing shaft as it passes, resulting in multiple-point readings.

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Figure 5-1 Reverse Dial Indicator Alignment

Figure 5-2 Laser Alignm ent System

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When taking as-found alignment data, perform the alignment with all conditions the same at all times. For instance, if the machine should be realigned with the piping full of water, the as-found data should be taken with the piping full of water. If the alignment cannot be performed under these conditions, attempts should be made to install stops on piping supports or springs to simulate a loaded condition. As-built alignment should be performed under the same conditions. (See Section 5.5).

If the as-found data is at variance with the previous as-built alignment data, it might become necessary to monitor the alignment movements as the pump heats up or cools down. This real-time alignment change data can provide insight into achieving proper collinear alignment in the hot condition. Shutting down a pump and trying to acquire data on alignment while hot is not a satisfactory alternative to controlled monitoring. Regardless of how quickly alignment readings can be taken, errors result. (Hot alignment is recommended. See Section 5.5.2 for a discussion of hot alignment.)

Manufacturers of the pump and/or the driver may supply data for cold aligning of the pum p to the driver. T his inform ation m ight not be valid for particular installations. Efforts should be made to determine the exact alignment of each pump at operating and cold conditions.

After an as-found alignment is performed and the coupling spacer is removed, a measurement should be taken between shafts to verify correct distance between the driver and the pump. For motor-driven pumps, care must be taken to ensure that the motor is in the operating position (magnetic center) when measurements between shafts are taken.

After this measurement, the pump should be checked for freedom of rotation, by hand or using a strap wrench. If rotation is difficult or rubs are heard or felt inside he pump, care should be taken during disassembly to preserve the as-found condition of the rubbing parts. The breakaway force to rotate the pump should be reasonable. Failure of the pump to rotate is indicative of problems. The area most often found to cause this problem is the seal area (throttle bushings and sleeves). This might be due to previous improper installation and centering of the rotor radially, improper warming of the pump, excessive nozzle loads, or all of the above.

5.3.2 Coupling Disassembly, Removal, and Inspection

Refer to Section 2.23 for additional coupling information.

Coupling Disassem bly

During removal of the coupling spacer, the condition of the coupling bolts should be visually checked. A policy of not using coupling bolts more than twice is good practice.

The majority of pumps use a tapered shaft at the coupling. The proper fit on a tapered hub coupling is crucial. The as-found hub location should be measured. A measurement

5-11 EPRI Licensed Material Nuclear Maintenance Applications Center from the end of the shaft to the coupling hub face, or from the hub face to a step on the shaft, determines where the coupling hub is positioned. This measurement can be used during reassembly to determine the correct position of the hub on the shaft.

Coupling Hub Removal

Two types of shaft mounting for couplings are used on feed water pumps: a keyed coupling on a tapered fit or a hydraulically dilated coupling hub on a tapered fit.

The keyed hub generally requires heat for removal. Care should be taken not to overheat the hub or place direct flame on the gear teeth of a gear-type coupling. One of the best methods for removal of the keyed hub is to use a strongback with a threaded rod and a small hydraulic ram with a hand-operated pump. Install a stop of some sort to prevent the hub from displacing completely off the shaft. The hub m oves with trem en- dous force and can be dangerous if allowed to travel com pletely off the shaft. W ith a small amount of tension on the hub, heat can then be used to displace the coupling.

Hydraulically dilated couplings are shrunk to an interference fit that precludes the use of a key. An extremely good fit to the shaft is required. A hand-operated hydraulic pump is typically used to swell the coupling for installation and removal. O-rings and backup rings are m the bore of the coupling to contain the hydraulic pressure. Care should be taken as the hub moves quickly once the interference to the shaft is relieved.

The pump and/or the coupling manufacturer should be consulted concerning specific couplings questions.

Coupling Inspection

Visually inspect the flex unit and driving sections of the coupling. A nondestructive examination should be performed on the coupling flex unit (gear, diaphragm, or disk), to eliminate concerns about fatigue cracking of coupling components. Visually inspect the shaft and the bore of the coupling for fretting, fatigue, or other damage. Verify good contact between the coupling and the shaft with a Prussian blue. The coupling hub should blue at least 75 to 80% when advanced up the taper by hand.

Keyways in the shaft and coupling should be checked for proper dimensions and radii. The key should have the proper fit into the keyway, snug on the sides and slightly loose top to bottom.

Table 5-2 depicts the most popular tapers and the method for determining the correct interference for your particular coupling hub.

NOTE: These are "rule-of-thumb" guidelines. The coupling and/or pump manufacturer should be consulted concerning coupling-to-shaft fits.

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Main Feedwater Pump Maintenance Guide

Table 5-2 C oup lin g F its

COUPLING HUB BORE

Bore Shape Straight Taper

Drive Method Keyed Keyed Hydraulic

Interference .5 to 1 M il T otal 1 M il/In. 1.5 to 2.5 M ilsi/ln.

Taper 3/4 In./Ft. 1/2 In/Ft. 1 1/4 In./Ft.

Minimum 70% 80% Contact Area

Vore Out of Roundness (M ax) Upto 4 In. 0.2 Mils TIR Over 4 In. 0.2 Mils TIR

To determine the taper per foot of the shaft and coupling, use the following:

Diameter large end - Diameter small end Taper per foot x 12 = Length of taper

To calculate the advance or pull-up of the coupling on the shaft, the formulas below can be used. Consult the coupling and/or pump manufacturer to determine the correct advance up the taper for the application.

Total required interference Advance = Taper per inch (tpi) taper per foot tp i = 12 tpi for 1/2 inch tapers = 0.0417 tpi for 3/4 inch tapers = 0.0625

5.3.3 End Play Measurement and Thrust Bearing Inspection and Removal

End Play M easurement

The as-found end play should be recorded at this time. The end of the shaft can be accessed for this check through the thrust bearing end cover. If the pump is equipped with a shaft-driven lube oil pump, it should be removed to access the shaft. The end cover and upper half bearing housing should not be loosened or removed prior to this check.

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After the end play has been recorded, the thrust bearing end cover and the upper half of the bearing housing may be removed. This exposes the thrust bearing and radial bearing assemblies. A typical thrust bearing assembly is shown in Figure 2-8.

If the end play is excessive, the previous as-built end play should be checked in the pump records. If there is a continuing problem with end play increasing over pump run periods, a root-cause analysis should be considered. See Section 4.5 of this guide.

Thrust Bearing Inspection and Rem oval

Thrust bearings are also described in Section 2.14 of this guide.

If excessive rotor end play has been found on disassembly, the thrust bearings should be examined closely. Remove the leveling plates by removing the screw on the periphery of the base ring and remove the leveling plates, keeping them in order. Wear might be observed around the pivot pin in the base ring. If you have removed thrust bearings on machines that exhibit thrust shuttling and measure a large increase in end play every time, wear may be occurring. Refer to Section 4.3.3 for interpretation of monitored parameters for diagnostic assistance.

The thrust shoes (pads) should be measured for thickness from the center of the back (button) side to the face of the shoe. The thickness of all the shoes should be within .001 inch of each other.

If the thrust shoes have signs of wear such as scratches, sm earing of the babbitt, or evidence of any damage, the shoes should be replaced. Prior to replacing the shoes and/or thrust bearing assembly, a thorough search for the root cause of the damage should be performed and corrective action implemented.

NOTE: Thrust shoes may be repaired for use under circumstances such as unavailability of a new set. Care should be taken when repairing the bearings by scraping these bearings to ensure that the proper thickness is maintained among the shoes. If the shoes are tapered, ensure that the taper is not disturbed or changed from the original taper.

Before lapping the thrust shoes, the manufacturer should be consulted for proper procedures. Ideally, the thrust bearing assembly should be returned to the manufacturer or a competent bearing shop for repair. The thrust collar should also be checked for damage.

Monitoring rotor shuttling is effectively accomplished by mounting a noncontacting proximity probe on the end cover. This is difficult to do on pumps equipped with a shaft-driven oil pump.

Shaft-driven oil pumps generally align directly to the end of the pump shaft. This precludes installing a probe to read from the end of the shaft. Probe locations must be installed further toward the periphery of the end cover. The probe locations penetrate

5-14 EPRI Licensed Material Main Feedwater Pump Maintenance Guide the end cover housing and pass through the thrust bearing area where they do not hamper the bearing function, yet provide enough area for the probe to read correctly. Small probes mounted internal to the machine, with an exit for the coaxial cable, can be used on pumps with shaft-driven oil pumps.

Some earlier pumps were equipped with three shoes on each side of the thrust collar, and a one-piece pivoting leveling plate. This design is susceptible to fretting between the base ring and the leveling plate, and the load might not to be distributed evenly over the shoes. If a pump is equipped with a single leveling plate and three shoes, consideration should be given to replacing the thrust bearing with the six-shoe design in Figure 2-9, or something similar.

By removing the outer thrust shoes and cage (leveling plates and base ring), the collar can be accessed for a runout check with a dial indicator. By applying a light amount of pressure to the end of the shaft at the thrust end, the rotor can be held in position against the inner thrust shoes. This runout check is valid only if the inner thrust shoes and cage are in good condition.

For a final check; the collar should be placed on a mandrel and checked between centers on a lathe.

Now, the thrust collar nut, thrust collar, and thrust collar spacer (if any) may be removed. Som e thrust collars have an interference fit to the shaft journal. This fit might require a puller to remove the collar. Care must be taken to prevent damage to the collar or the bending of the small section of shaft under the collar. Holes might be present in the thrust collar to facilitate using a puller (if so, these holes are typically inside the outer diameter of the thrust collar nut).

C losely monitor the interference fit between the shaft and the thrust collar. If interference is used in this area, a check should be made every time the collar is removed to ensure presence of the required interference.

If the thrust collar reunite is greater than .001 inch, then the collar must either be machined and ground or replaced. By machining the two faces of the thrust collar on a mandrel and between centers, the faces will be perpendicular to the center line of the shaft. After machining, the collar should be finished on a surface grinder. When a thrust collar is ground, ensure that the thrust collar spacer is also flat. If shims are used behind the base ring, the shims must also be ground flat. Check all new collars for runout prior to installation.

5.3.4 As-Found Rotor Positions

Many pumps, by design, cannot be checked for axial and radial rotor positions with all the parts in place. Seal covers, shaft sleeve nuts, mechanical seals, etc., might inhibit movement of the rotor while trying to perform these checks. Due to the time involved, many maintenance groups would prefer to do these checks upon reassembly rather than by disassembling the parts and reassembling the bearing housings and bearings

5-15 EPRI Licensed Material Nuclear Maintenance Applications Center during the teardown portion of maintenance work. If the pump is experiencing undiag- nosed problems, time should be taken to perform these checks during disassembly.

Measurements of radial and axial rotor positions on reassembly are discussed in Sec- tions 5.4.4 and 5.4.5 of this guide.

Clearances between the bushings and sleeves are typically smaller than the impeller-to- casing ring clearances. The rotor should be centered so that rubs do not occur at these clearances. An out-of-center dimension of .001 inch or less is typical and should be maintained to ensure proper operation of the pump. A typical clearance in the sleeve and bushing area is .010 to .020 inch depending on the manufacturer.

5.3.5 Auxiliary Piping

Certain auxiliary piping will probably need to be removed to access the pump itself. This piping may consist of oil inlet and drain piping, seal injection piping, and seal leakoff piping. Attention to the cold spring of this piping is essential to eliminate external preloads on the pump.

Some auxiliary piping systems may use flexible bellows or braided steel hose, while others may be hard piped. The elimination of cold spring on hard piped applications may require the cutting and welding of the pipe. Many stations are reluctant to perform this operation on oil piping because of the fire hazard. After cutting and welding, the piping requires cleaning to remove grindings and slag.

Seal injection piping is typically not of the size to cause major cold spring on the end covers or pump casing. If this piping is modified, care should be taken to clean the sections thoroughly to eliminate the possibility of small metal particles entering the seals.

5.3.6 Radial Bearings

Bearings survive only with clean oil, proper clearances, correct assembly techniques, and proper loading. By proper loading, inference is made to the correct static and dynamic loading of the bearing, and the proper bearing preloads which are not influenced by other external or internal loads.

The bearing is a primary source of damping and may inhibit vibration in one direction, particularly if loads are excessive in that direction.

Internal loads may be a result of improperly machined or assembled bearing housings, retainers or bearings, incorrect bearing loads due to faulty bearing clearances and configurations. Incorrect radial centering, running position, and imbalance may effect loads. The improper indexing of top half to bottom half of the bearing will affect loads and temperatures.

External loads may be misaligned influencing shaft behavior, or piping forces acting upon the casing and/or the bearing housings.

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The terms above, discussing internal and external loads, should not be confused with bearing preload in the engineering and design context. Bearing preload is defined as any circular journal bearing with pads whose center of curvature (or center of machined surface) does not coincide with the geometric center of the bearing.

Radial (journal) bearings are described in Section 2.13 of this guide.

Inspection of Cylindrical or Elliptical Bearings

When the top half bearing cap or housing is removed, several checks may be performed to verify the condition of the bearings.

The condition of the shaft journals and the bearings should be noted at this tim e. If the bearings display hot spots and/or scoring, this should be noted and exploration into the cause considered.

An as-found check should be made of the clearance or interference between the bearing and the housing. Use plastigage or lead fuse wire and shims at the housing part line.

The use of plastigage is a good method to make the initial check of bearing to shaft clearance to determine whether or not it is acceptable. Prior to returning bearings to service, they should be measured accurately at several places to ensure: the proper clearance; that the bearing inside diameter is concentric with the outside diameter; and that high spots are absent.

Cylindrical or elliptical bearings should be checked for high spots, however minute, by passing a scraper over the bearing. This can be accomplished with an accurately ground piece of tool steel, such as a 5 to 6 inch cutoff tool used on a lathe. Lightly blue the bearing halves with Prussian blue. By holding the cutoff tool parallel to the bearing or along the line of the shaft and passing the tool from one part line of the bearing to the other part line in a steady and continuous motion, the high spots will be exposed from under the blue. This removes very slight high spots, and if care is taken in performing this task, will not open up the overall clearances. This helps eliminate hot spots and slight rubs.

The bearing should be checked for parallelism to the shaft. Prussian blue may be applied to the shaft, with the top half of the bearing off. The blue should be the length of the bearing. The shaft can be rotated to determine if the blue has contacted all the way across the bearing.

If the bearing to shaft parallelism is found to be incorrect, a quick check may be made to the lower half bearing housing after removal of the housing from the pump casing or end cover. This is of particular importance if bearing and housing have a straight bore fit as opposed to a spherical fit.

A precision machinists square may be used on the bearing housing, with the short side of the square contacting the area of the housing (face) that bolts to the pump, and the long side of the square along the horizontal joint (split line). By shining a light from the back side of the square, a determination can be made whether or not the housing is square.

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If the housing is not square, the bearing seat (saddle or fit) is not parallel to the shaft. Machining of the bearing housing(s) may be required to true the housing (bearing seat) parallel to the pump casing and the shaft.

Inspection of Tilting Pad Bearings

Tilting pad bearing clearances may be checked using several different methods. The use of mandrels is preferred. Two different size mandrels may be used to check bearing contour and clearance if the bearing has a designed preload.

The mandrel should be ground to the proper diameter for accurate checks of tilt pad bearings. One method is to use a mandrel that is the shaft diameter plus the clearance. For example, if the shaft is 5.000 inches and the recommended bearing clearance is .0075 then the mandrel should be ground to the finish diameter of 5.0075.

When the tilt pad assembly is placed around the mandrel with Prussian blue on the mandrel, transfer of the blue should be the same on all pads of the bearing. If the bearing is designed with preload, the blue should only contact a portion of the bearing (for example, only one inch of the pad's center).

For correct checking of the contour of the tilt pad bearing with designed preload, another mandrel must be machined and ground to the pad radius (for example, the pad can be turned on a diameter of 5.011 inches). To verify this, the individual pads must be checked to a mandrel of this size. The blue transfer should be across the entire pad.

One method of checking tilt pad bearing clearance, prior to removal, is by using two dial indicators; one reading the shaft close and adjacent to the bearing and the other reading the top half of the bearing. Lift the shaft carefully and read the shaft indicator, while ensuring that the bearing indicator does not move. If it does, subtract to determine the difference between the two readings, which is the diametral bearing clearance.

Bearing Seats

Two types of bearing seats can be found in the main feedwater pumps. One is straight fit and the other is spherical fit.

The straight-seat bearing and housing must be parallel to the shaft, vertically and horizontally. Any skew or cant causes the bearing to misalign with the shaft.

The spherical seat bearing, shown in Figure 5-3, can be checked in the same fashion as the straight-fit bearing using Prussian blue and rotating the shaft. This bearing also has the advantage of being checked and adjusted using a feeler gage at the part line. The top half may be bolted to the bottom half and plastigage or lead wire used to verify parallelism of the gap. A length of plastigage or fuse wire may be placed axially along the full length of the shaft prior to installing and tightening the top half.

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Figure 5-3 Spherical Seat Bearing

Spherical seat bearings should be manually paralleled to the shaft prior to installing the top half of the bearing. With the top half installed and the housing in place, the spherical-fit bearing is held in place by a slight interference. This interference might range (metal-to-metal) from 0 to .002 inch.

Bearing Failure Modes

Failures occur in five general areas: overheating, wiping, corrosion, fatigue, and wear. The ability of maintenance and support personnek- o m ake determ inations as to the failure mode of the bearing is critical. In addition to the five failure modes, bonding of the babbitt to the parent metal should be checked using nondestructive evaluation, such as dye penetrant testing. Overheated bearings reveal discoloration and cracking of the babbitt material.

Wiping is a smearing and displacement of the bearing material circumferentially in the direction of rotation, caused most frequently from bearing overload, loss of lubricant, or lack of sufficient speed to develop a film Wear can be caused by foreign material in the bearing and is obvious by scoring and scratches along the path of rotation.

As material is removed from the bearing by hard or abrasive material passing through the bearing, this material must be relocated, typically within the bearing. This increases clearances in some areas and reductions in others and might progress to the overheating stage and the wiping or smearing stage. Root cause of the problem can be difficult if the problem has progressed from one stage to another. (See Section 4.5 of this guide for a discussion of root-cause analysis.)

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For detailed reading on the subject of bearings, a previously published EPRI report, Manual of Bearing Failures and Repair in Power Plant Rotating Equipment, GS-7352 (Refer- ence 23) is recommended.

5.3.7 Shaft Oil Seals

Shaft oil seals serve two purposes: (1) they prevent oil leaks along the shaft, and (2) they prevent water from entering the bearing housing and contaminating the oil.

Two types of oil seals are generally used on these pumps. One is a rotating seal with labyrinth grooves and a clearance between the seal and the stationary housing. The other is a stationary seal (Figure 5-4) with labyrinths on the shaft side and a small clearance between the seal and the shaft. Both of these are susceptible to leakage in either direction, especially when used with a fixed-breakdown seal injection system.

Figure 5-4 Stationary-Type Oil Seal

Due to a combination of system and pump design, there have been several instances where oil is drawn from the bearing housing over into the condensate system causing fouling water chemistry.

If these conditions exist, first refer to the pump manufacturer. In some cases, utilities have investigated the problem and found solutions on their own. One solution was using a rotating oil seal or bearing isolator originally designed for ball-bearing machines, such as overhung ANSI or API pumps, and for electric motors where lip seals were previously employed. This device is shown in Figure 5-5.

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Figure 5-5 Bearing Isolator Used for M ain Feedwater Pump Bearing Housings

When converting to these types of seals, the bearing housing must be machined to the correct dimensions, allowing for O-ring compression. If not properly installed, the seal might by to carry part of the shaft load.

5.3.8 Injection-Type Seals Injection-type seals and their injection control systems are described in Sections 2.17 through 2.20 of this guide.

Seal Removal The sleeves and bushings may now be removed. If the pump is equipped with floating ring seals, the seal housing may now be removed.

The sleeves and bushings are a labyrinth-type, with either or both being serrated. This contributes to the pressure drop achieved across the length of the sleeve and reduces the surface area to help prevent galling of the sleeve to the bushing. One is typically harder than the other. Some manufacturers do not use sleeves, and the shaft runs against the bushing.

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On radially split or end pull-out pumps (Category 1 or 2, Section 1.6 of this guide), the end cover on some of these models may be removed with the bushing in place. If this method is preferred, the sleeve should be removed first to prevent any damage to the sleeve or bushing that might mask damage that occurred during the previous run.

Inspection

When the bushings and sleeves are removed, they should be measured to determine the size of each and the clearance. The sleeves and bushings should be measured at the OD and ID in a minimum of two directions. The measurements should be taken along the length in at least in two locations.

These measurements determine if each part is out of round, worn tapered, or eccentric. If the bushing or the sleeve is smooth (without any serrations), visual inspection will disclose any rubs. If rubs have occurred and they are minor and within tolerance, the parts may be reused.

Manufacturers vary in the amount of diametrical clearance between the throttle sleeves and bushings. In general, these clearances range from .010 inch to .018 inch. Some manufacturers specify more or less. If a machine has less than .010-inch clearance dia- metrically, the sleeves and bushings are more susceptible to rubbing and galling during warmup and shutdown of the pump.

These different designs have different hardnesses for the sleeve and/or the bushing. Typically, the serrated piece is harder and has been ground to the final dimension. The softer of the two is the one to be machined if the clearances are to be enlarged.

Several stations using fixed-breakdown seals have increased the diametrical clearance. This is not a problem if the clearance is increased from the minimum and rem ains within tolerance. If you are continually seeing damaged bushings or sleeves, then dis- cuss increasing the minimum clearances with the OEM. In one instance, the recommended clearance is .010 to .012 inch (the clearance was increased to .014 to .016 inch). This station desires to sacrifice a small amount of efficiency for reliability and reduction in seal problems during startup. The sleeves on this particular pump have always experienced rubs due to casing flexibility and nozzle loads.

Proper installation and radial centering of the rotor is crucial to prolonged and efficient operation of a feed pump. The radial center uses a hydraulically induced function to aid in maintaining proper loading of the rotor and radial bearings and adds stability to the rotor.

5.3.9 Mechanical Seals

If the pump has mechanical seals, they should be removed at this time.

Refer to Sections 2.21 and 2.22 of this guide for descriptive information on mechanical seals.

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5.3.10 Hydraulic Balancing Devices

The operations and maintenance manual and the manufacturer's field service depart- ment should be consulted when installing and setting the balancing device. When removing the balancing device (disk or drum), the parts should be checked closely for cracks. This can be accomplished with a dye penetrant test.

Axial positioning of the shaft and, therefore, the rotor assem bly is critical at any rotor close-tolerance axial interface (i.e. balancing disk to end head, impeller exit to volute/ diffuser exit and overlap, etc.). It is also important to properly position the thrust bearing collar relative to the balancing disk, as it directly affects the carrying capability of the combined axial thrust load. Dimensions from the disk to the end head are critical.

When using a parallel face balancing disk design, a hardness differential of the interfac- ing rotating disk and stationary end cover of no less than 10 Rockwell C hardness is considered best practice.

Balancing drums or pistons rely heavily on correct clearance for proper operation. Excessive wear of the drum is defined as greater than one and a half times that of the design radial clearance.

The cause of wear in these areas is excessive radial and/or axial vibration. High thrust bearing temperatures can be a result of balance drum wear.

For more detailed reading on the subject of balancing devices, refer to the EPRI report, Boiler Feedpump Operation and Maintenance Guidelines - Volume 1 Troubleshooting, TR- 104292-VI (Reference 17).

5.3.11 Rotor Removal and Inspection

Rotor Removal

Prior to removing the rotor from radially split casing Category 1 and Category 2 pumps, the pump end cover or head must be removed. Guide studs or rods should be used to ensure that the cover does not skew or bind when removing. The use of guide studs often minimizes problems when removing or installing heavy pump parts. On axially (horizontally) split casing Category 3 pumps, the top half of the casing must be removed prior to removing the rotor.

The rotor should be removed with care. Special tooling is sometimes supplied by the pump manufacturer to assist with proper removal of a rotor. These fixtures or special tools might need modifying to perform correctly. Using the manufacturer's design as guidance, special tooling can be custom fabricated.

Caution and safety are of the utmost importance when removing a rotor. Care should be exercised in preventing damage to the rotor. Thin sheeting, such as a rubber gasket material, is ideal for covering the shaft to prevent damage during removal.

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Runouts

Runouts of the shaft and impellers should be kept within the manufacturer's specified limits. When measuring runouts of shafts, ensure that all specified runouts are taken. For best results, runouts should be taken on V-blocks or rollers. If V-blocks are used, the bearings for the pump should be placed in the V-blocks and squared to the shaft. This helps in rotating the shaft and prevents scratching the journals. A small amount of oil should be kept on the bearings. If rollers are used, they should be clean and free of burrs. Rollers in less-than-perfect condition can damage the shaft.

This type of pump typically has a small shaft on the thrust bearing end. The shaft should not be placed between centers in a lathe to check runouts. Therefore, get the runout on the thrust collar of the pump prior to disassembly. Figure 5-6 is an example of a worksheet for recording shaft runouts and dimensions. This worksheet has the impeller removed. A worksheet showing the impeller is also needed for most pumps. Following the shaft worksheet is one for the impeller listing the dimensions that ensure proper fits for installation.

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Figure 5-6 Shaft Runout and Dimensions W orksheet

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Impeller Removal

A complete teardown of most nuclear main feedwater pumps does not require removing the im pellers from the shaft.

NOTE: Remove the impellers only if necessary.

Some impellers have an interference fit all the way through the bore, while others might have interference only on the ends with a relieved area in-between. The impeller in Figure 5-7 shows a relief between the fits.

Figure 5-7 Double-Suction Impeller

A dimensional record should be made of the impeller location on the shaft prior to removal. This dimension should be measured within the tolerance for the running position of the rotor to casing, but should not exceed .010 inch from the original location on double-suction impellers. This allows for continuity of the previous location to the new impeller to casing location and also eliminates major changes to the thrust collar spacer or thrust shim s.

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Several things change with impeller location, including distance between shafts. If a flexible disk or diaphragm coupling is being used, this can be detrimental to the flex units in the coupling. The shaft sleeve location can change by a small amount. Some rotors with shaft sleeves require accuracy of this position to prevent problems with interaction of sleeve nuts later during assembly.

Some manufacturers use a clearance between sleeves and the impeller hubs to allow for growth of all components without causing problems to the rotor. Accuracy of impeller locations are crucial to ensure that sleeves do not contact the impeller. Many shops cannot measure the long distance to the end of the shaft from the face of the impeller hub. If this is the case, a step on the shaft near the hub will suffice By using a straight edge against the step on the shaft and an inside micrometer, the distance can be measured to the existing im peller location.

When heating an impeller for removal from the shaft, the hub areas receive the last heat prior to removal. This allows for easier removal and prevention of galling. Impellers with a fit throughout the bore typically have less interference than ones with fits on each end. When heating an impeller for removal, always consult the manufacturer's literature to determine the amount of heat to be applied in removing the impeller.

The impeller and shaft should never be placed in a hydraulic press for removal of the impeller. The rotor should be suspended from overhead and hung by the coupling end. Cribbing should be placed under the impeller, allowing enough room for the impeller to clear its fit on the shaft. W hen applying heat to the impeller always start heating at the periphery of the impeller and work in toward the hub.

With double-suction impellers, two heating tips should be used. Preferably, use at least N um ber 14 heating tips. W ork the heat in toward the center of the impeller, being careful not to plN the heat on the shaft. Temperature sticks or a contact pyrometer are necessary to control the amount of heat being applied.

When the impeller starts to move off the shaft, tap the impeller on the wear surfaces with a soft hammer, such as a plastic-coated deadblou0Èr rawhide mallet. Never strike the impeller on the shrouds or near the vanes. If heated properly, a large force will not be required. Single-suction impellers typically do not require as much heat on the suction side of the impeller as on the back or hub side.

M ultiple stage pumps might have stepped shafts, with each impeller having a different bore diameter to ease removal of the impellers. Number the impellers to identify the location fromC/hich they were removed.

M any single-suction impellers on multistage pumps have split rings on one side of the impeller, in the direction of removal. In this case, the impeller must be slid away from the direction of removal and the split ring removed. These rings are typically two half- m oon keys that fit into a groove and locate the im peller axially on the shaft.

One must work precisely and quickly when the impeller is ready for removal.

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CAUTION: If an impeller stops moving, STOP immediately. Do not hit the impeller to drive it from the shaft. Galling may have occurred and the impeller will not move further at this time. Allow the impeller and shaft to cool to room temperature. Evaluate the situation before again applying heat to the impeller. If an impeller is seized to the shaft and cannot be removed, an evaluation by the engineering staff or the manufacturer is in order.

When all else fails, the shaft might require cutting on each end of the impeller and the remaining shaft machined from the impeller bore. This is not an ideal situation, but a shaft is typically easier to replace than an impeller. Always try to prevent impeller damage.

Impeller Inspection

Visual inspections of the impellers should be made as soon as possible after removal from the pump. This allows time for a decision about the use of the impellers and/or whether new impellers should be installed. Particular attention should be paid to the areas that show damage from cavitation or erosion. Look closely at the leading edges of the impeller, paying particular attention to both the front and back sides of the vanes throughout. An inspection mirror might be required to see some areas. A magnetic particle or liquid penetrant test is always prudent when the impellers are removed from the pump.

If a new impeller is to be installed, a decision must be made concerning the disposition of the old impeller. Depending on the type and extent of damage, an impeller might be repairable. The manufacturer should be consulted about making weld repairs to impellers. A welding procedure for repair of the specific im peller is necessary, or the im peller may be returned to the manufacturer or an authorized repair agent.

The inspection of the impeller is critical. The following guidelines are typical. The wear surface or wear hub of the impeller should run out no more than .001 inch. Turns and bores should be concentric and be out of round no more than .001 inch. Runouts may be performed on the shaft. If the impeller is off the shaft, the runouts may be measured on a good-fitting mandrel. The impeller bore should be round to within .001 inch. The impeller bore should be true without any taper. The OD of the impeller shrouds and vanes should runout no more than .002 inch and should be within .005 inch of the old or used impeller diameter.

CAUTION: Many impellers are trimmed at the factory after or during the performance test. In certain instances, the drawing for a replacement impeller might not reflect that a trim was performed after a factory test. When receiving an impeller for installation, the OD should be verified to duplicate those of the impeller being replaced. Otherwise, a change in hydraulic performance will occur.

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When receiving a new shaft for use, several dimensional checks should be made. All threaded areas should be checked to ascertain whether or not the threads have been shot peened. Shot peening cleans and smoothes the surfaces of the threads and helps eliminate galling. Shot peening provides a method of cold working the area and inducing compressive stresses, thereby reducing tensile stresses that contribute to stress corrosion cracking. The manufacturer's bill of materials or documentation may or may not reveal whether shot Keening is a requirement and whether it has been performed on the subject shaft.

If shot peening has not been performed on a shaft, an evaluation by the utility or competent outside metallurgist should be done.

The impeller bore and the keyway should be measured to determine if the dimensions are correct. The impeller should be inspected to verify that the keyway is machined true to the bore.

5.3.12 Casing and Casing Ring Inspection

A thorough visual inspection should be performed on the inside of the pump casing. Look for areas damaged by loose objects, particularly volutes and diffuser lips. Signs of washing in the casing should be documented and guidance sought for the necessary repairs, if required.

All as-found measurements should be taken on the casing rings prior to removal. Cas- ing rings may not be round once removed from the casing. New rings may also be out of round. Once the casing rings are removed from the pump casing, the casing fits should be measured to determine concentricity and the interference or clearance between the casing and rings.

Case wear rings (casing rings) of feed pumps vary with the manufacturer as to the way they are installed. Some rings are bolted in with minimum interference or a small clearance. Others depend on large interferences along w ith high-pressure differentials ex- erted m a direction that prevents movement of the rings. Casing rings with large interferences change m size once installed (the IDs are reduced). Measure the ID after installation. The impeller or impeller wear ring might require m achining for correct running clearances.

The difference in ID dimensions between the original rings in place and as removed indicate any change in ID of the new rings expected when installed.

Special tooling from the manufacturer or fabricated on site is crucial to timely and accurate removal of parts (especially casing rings) or subcomponents of the pump. Plant workers have devised some timesaving tools for removing and assembling components of main feedwater pumps.

Manufacturers and specialty vendors supply custom tooling to assist in the removing the rotor. These fixtures might need modification for other facilities. These tools and fixtures can help ensure that parts are installed more accurately without galling or seizing.

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5.4 Pump Reassembly and As-Built Data Prior to assembly of the pump, all parts and components must be thoroughly cleaned and deburred. A close inspection and feel of every part for imperfections is required. The slightest metal upset as a result of shipping and handling can cause disastrous consequences. This is tim e well spent prior to assem bly.

5.4.1 Casing Ring Installation

Casing rings with large interferences might require cooling with a substance such as dry ice or liquid nitrogen. Before using liquid nitrogen, consult with the pump manufacturer to obtain approval. The metallurgical properties of the material might thermally shock the rings.

Ensure that there is more than enough clearance before installing the rings, then pro- ceed quickly to seat the rings effectively. The ring must be seated straight to prevent getting a ring stuck in the casing before reaching the final position. A good precaution is to have another person available on the other side of the ring to drive the ring back out using something such as a piece of wood.

Once the rings are in place, keep pressure against them for a short time to ensure that they do not creep back from their intended location. After the rings have reached ambient temperature, a final measurement should be taken and recorded to verify the clearance between the rings and impellers.

5.4.2 Rotor Reassembly and Balancing

Rotor Reassembly

When measuring the location of the impeller on a new shaft, a split ring with a flange and a worm gear clamp assists in making an accurate measurement. Once this ring is in place at the impeller hub location, the shaft sleeve or impeller nut can be positioned against the locating sleeve and the locating sleeve removed. The impeller can then be installed using the shaft sleeve or impeller nut as a locator.

W hen installing large double-suction impellers, the shaft should be suspended from an overhead crane and the impeller supported on a special stand that allows even heating, if a heating tip is used. If an oven is used to heat the impellers, this oven should be close in proximity to where the work is being performed. The shaft should be suspended from the coupling end on double-suction impellers.

If the pump is not equipped with an impeller (lock) nut and the sleeve (fixed breakdown) comes in close proximity to the impeller, and this sleeve uses a metal O-ring between a step on the sleeve and a corresponding step on the shaft, the ring can be left out to gain enough distance to use the sleeve as a stop. Once the impeller has cooled at its final position, the sleeve can then be removed and repositioned to the correct location.

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Shaft sleeve nuts might be a cause of problems during reassembly. If the shaft sleeve nuts have tight clearances between the nut and shaft, a minor amount of eccentricity can cause binding of the nut on the shaft during installation.

Ensure that when the nut is installed, the nut does not bind on the shaft. The nut might require machining on the outer bore to ensure that clearance is available and the nut does not cant when installing. This causes galling of the nut to the shaft. This also allows for the nut to not seat correctly against the sleeve and to cause shaft bending.

Rotor Balancing Rotor assemblies are balanced at the manufacturer's facility prior to shipment. Impellers might or might not be balanced upon receipt. Impellers may be checked for balance at your facility, if a balancing machine is available. If a new replacement impeller is required for the job, this impeller can receive a preliminary balance on a mandrel. A better alternative is to have a rotor or rotating assem bly readied for installation in the casing the minute the old rotor is removed. The removed rotor can be refurbished at a more convenient time.

When grinding on impellers to remove weight for balancing, the impeller should be checked for identification markings. The markings consist of a part number, serial number, heat number, or other factory-applied identification or combination of markings. All numbers should be recorded. These markings should also remain on the impeller. If identification markings are in the area where metal is to be removed in balancing, the numbers should be reapplied to another location on the part before they are removed In grinding and balancing.

Rotor designs of diffuser multistage pumps have unique balancing problems. The rotors cannot be installed as a unit immediately after having been balanced as a complete assembly. The rotor must be partially disassembled to rebuild the total pump. The assumption is that, when properly balanced as an assembly, the balance of the rotor as rebuilt upon installation m the pump still provides sufficient balancing retention accuracy when properly rebuilt to operate well.

A general rotor balancing worksheet is provided m Figure 5-8.

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Feedwater Pump l.D. # Date Operator Rotor TIR within Limits Prior to Balancing Yes

TIR Allowable Shaft .001 in. Yes TIR Allowable Impeller W ear Hubs .001 in. Yes TIR Allowable Impeller Periphery .0015 in. Yes

Impeller Requirements Two Plane to oz. In. Impeller W eight lbs. Rotating Assembly Two Plane to oz. In. Rotating Assembly W eight Ibs. Rated Pump Speed Balance Machine Speed

FINAL IMPELLER UNBALANCE Plane #1 Grams @ in. radius @ Phase Angle Plane #2 Grams @ in. radius @ Phase Angle

FINAL ROTATING ASSEMBLY UNBALANCE Plane #1 Grams @ in. radius @ Phase Angle Plane #2 Grams @ in. radius @ Phase Angle

Figure 5-8 Rotor Balancing W orksheet

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5.4.3 Rotor and End Cover Installation

Rotor Installation

Care must be taken during installation to cradle or support the rotor properly to prevent bumping it while maneuvering the rotor into position for installation into the casing. Special tooling might aid proper installation of the rotor. Once the rotor is in position for installation in the casing, care again must be taken not to bump it, particularly on the small end of the shaft at the thrust collar and nut location.

If the pump is equipped with injection throttle bushing seals (sleeves and bushings), certain steps should be taken in the assembly sequence to prevent rubbing of internal parts or to keep contact between the rotating and stationary parts at a minimum.

In some cases, based on the particular design of the pump casing, rotor, bushings, and sleeves, the rotor may be installed without the sleeves (See Figure A-1 as an example). After the rotor is in position and the end covers are on the pump, the bushings may be installed. In some cases, the bushings can be installed and then the sleeves installed on the rotor.

End Cover Installation

R adially split m ain casing joints are discussed in detail in Radially Split M ain Casing Joints in Section 5.7.1 of this guide.

Installation of the end covers or end heads on radially or vertically split case pumps can be performed with a minimum of problems using the following technique.

Some end covers on double-suction pumps are nearly as long as the diameter at the outer end of the cover. These types of end covers should be brought against the pump volute or casing by initially tightening the bolts in a circular fashion. Guide studs should be used to ensure that the end cover (head) is correctly piloted to the casing. Snug all nuts evenly and turn each nut slightly until the adjacent nut becomes loose. Tighten this nut slightly and continue this process until the end cover contacts the gasket or m etal to m etal fit, then use the torque sequence specified in the OEM 's instruction manual and torque in a crisscross fashion.

M ultistage pumps with end covers might require the same technique as above. Consid- eration must also be given to the compensator rings and gaskets as well as how to pull the end cover (head) to the proper position and compress all gaskets and rings. The manufacturer's literature should be consulted and followed closely on multistage main feedwater pumps.

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A xially Split Casing Assembly

Axially (horizontally) split main casing joints are discussed in detail under Section 5.7.1 of this guide. Axially split casing pumps use a specified thickness and compound of gasket to seal the pump casing joint. If another thickness or a variance in thickness of this sheet material is used, or a material is used that does not have the approval of the pump manufacturer, substantial harm can result when the pump is assembled or operated. The casing ring fit and clearance, along with any stage bushing fits, depend on the correct thickness and proper torque.

The torque sequence provided by the manufacturer should be closely observed. Thread lubricant must be taken into account when tightening all studs and nuts.

5.4.4 Radial Rotor Position or Center

Several methods are available to verify the radial center of the rotor in the casing.

Lift Check Method

This method is the preferred method of verifying the radial center in the vertical direction. It is quick, easy, accurate, and especially useful when starting to position the rotor and prior to rotation of the shaft after reassembly to ensure that there is no metal contact within the pump.

With the rotor supported by both bearings, use a dial indicator to monitor the shaft travel in a vertical direction just inboard of the bearing. Remove the bearing on one end by rotating the bearing in the housing and resting it on top of the shaft. Zero the dial indicator with the shaft in the bottom position. Lift the shaft to its total travel and note the total lift. If the total lift reading is .022 inch, then this measurement is greater than the clearance between the sleeve and bushing because the reading was taken further out on the shaft. W ith the bearing positioned on top of the shaft, roll the bearing in and let the shaft rest on the bearing in its running position. Read the dial indicator again. If the rotor is in the center, the indicator in this example will read .011 inch. The bearing housing may be positioned vertically as required to achieve the proper running radial position. The rotor must always be positioned with one end of the rotor supported by a bearing.

The horizontal position of the bearing housing, bearing, and the rotor can be checked by rotating the bearing about the shaft and moving the housing until the bearing rolls in from either side of the shaft.

An alternate method that should be used in conjunction with the lift method is measuring the clearances between the bushing and sleeve with a feeler gage. The sleeve nut must typically be removed from the shaft to facilitate this measurement. The design and assembly sequence on some pumps do not allow for this measurement to be taken.

When the lift and feeler gage methods are used in conjunction, they allow checks that ensure proper radial centering of the rotor in the casing.

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Runout or Dial Sweep Method

This method can be used when the shaft can be rotated with a dial indicator mounted to the shaft and swept over a machined surface in the seal chamber that is concentric with the casing. The surface reading with the dial indicator must be smooth and concentric with the bushing bore of the casing.

The rotor position is correct when the rotor is in the center of the casing. When installing the bearing, if the bearing rolls in freely without lifting the shaft, the housing is probably not in the correct vertical position.

The configuration of the bearing housings on these pumps differ among manufacturers, and the method of maintaining the proper position of the housings and, thus, the rotor must be handled differently. Bearing housings on several types of pumps are adjusted by bolts or cap screws with locking nuts to restrict m ovem ent of the bolts once adjusted. These types of housings might not have a taper pin locator to position the housing to the casing. When performing a radial center check or adjustment on pumps with this type of housing, all centering checks and moves must be made at the bearing or around the area of the seal cover (seal flange).

These radial center checks are best performed m the vertical direction by the lift check method. The side-to-side or horizontal centering can usually be performed by checking or measuring the dimension from the shaft to the bearing fit at the horizontal joint or parting face of the bearing housing. The shaft should be sitting in the bottom of the bushing at true bottom center. With the bearing out of the housing and taking care not to disturb the location of the shaft, the bottom half of the bearing can be placed on the shaft and rotated about in either direction to ascertain if the bearing is on the part line on one side and rolling into the housing on the other side.

The housing can then be adjusted in a horizontal direction until the bearing travels into to housing on either side and at the same distance. At this point, the shaft can be lifted slightly and the bearing rolled m to its final position. A vertical lift check should again be performed to determ ine if the rotor is still m the correct position vertically. If not, both the vertical and horizontal checks should be performed again.

Taper Pins

Bearing housings equipped with taper pins return the bearing housing to the centered position. If the centered rotor position has changed from the previous location, the taper pin holes must be reamed. This will bring the housing back to the proper location when the housing is removed to install the seals and/or seal covers. The quality of the taper pin hole is established by proper reaming methods. Great care should be taken in performing reaming jobs.

On pumps equipped with through pins or pins and holes on the back side of the housing or drip pocket, these are fairly typical tasks. On blind tapered pin holes, the task becomes more difficult. When removing metal in a taper pin hole, the reamer and, thus,

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the pin might bottom in the hole. This is not satisfactory. The reamer may have to be shortened by grinding. Care must be taken to keep the reamer cool and lubricated while grinding, and the ends must be finished correctly.

When the hole is reamed to the point where no lines are visible between the housing and its mating surface, the appropriate pin should be installed and blue checked for a correct fit. If these holes are not perfect, the housing will not return to the location from which it was removed.

Pins do not need to be driven into the housing. Only a slight tap to ensure seating is required. These pins should not be placed in shear at the mating surface. The bolts or studs and nuts should provide adequate clamping force to hold the housing in place. Many experienced personnel remove the pins once the housing is in place and insert them again with minimum pressure to ensure they do not fall out and get misplaced.

Taper pins should be purchased from a mill supply. Attempts to make these pins inhouse should be avoided. M any pins have a threaded portion on the visible end of the pin for a nut to assist in removal. More often than not, these pins are damaged.

One method of removing pins in blind pilot pin holes is to drill the pin through its length. A number seven or eight pin will use a 1/16-inch hole. Pin numbers nine and up will use a 1/8-inch through hole. The outer portion of the pin can be drilled and tapped for a grease fitting, which, when pumped with a hand-held grease gun, hydraulically removes the pin.

A grease-fitting plug can be used or a bolt inserted in the hole to assist in removal or kept in place after completion of the job.

CAUTION: This method requires an extremely good pin hole and pin seat for hydraulic removal of the pin.

5.4.5 Running Position (Axial Center)

The axial running position of the rotor in relation to the casing is important. The im peller should be positioned axially in the center of the volutes or diffusers. If a double- suction impeller is used, then the center of the impeller along the axial plane should coincide with the center of the volute or diffuser. (Appendix J details the methods available for achieving the correct rotor running position.) This allows for the proper hydraulic balance of the rotor and impeller and reduces loads on the thrust bearing. This also prevents or reduces side loading of the impeller shrouds, reduces the opportu- nity for rotor shuttling, and also reduces cyclic fatigue on the impeller shrouds. Most importantly, the high thrust bearing temperatures can be reduced on one or the other thrust bearings on double-suction pumps. Thrust bearing thermocouple monitoring is an ideal tool to ascertain whether or not loading is occurring m one direction more than the other, due to improper axial rotor position.

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5.4.6 Mechanical Seal Installation

Achieving successful seal operation requires attention to a number of critical areas besides the design and quality of the seal hardware itself. Other than hardware, attention must be directed to the manner in which the seal is installed, how the seal relates to the pump hardware, the nature and success of any repairs to the pump hardware, and operation of the pump.

Seal hardware failures have been placed in categories of broad ranges. Forty to 70% of seal hardware failures are on one or both of the primary sealing faces. Often, an upgrade in materials suffices. Finite element analysis is used to determine the geometry of the faces, which, in turn, determines whether a full fluid film between faces in a convergent gap configuration can be achieved. This is the desired operational geometry for a successful seal.

The basic construction of a main feedwater pump mechanical seal (as shown in Figure 5- 9) involves primary and secondary sealing paths. The primary sealing path is the interface between the rotating hard face and the stationary face. The secondary seal is the interface area that can experience dynamic axial motion. Loss of axial movement accounts for 10- 20% of hardware failures. Loss of axial movement is usually a physical hang-up of parts due to accumulation of dirt and corrosion products in clearances. Fretting and corrosion each account for a similar number of hardware failures as loss of axial movement. Fretting occurs between parts that rub against each other. Solutions to fretting problems include proper tolerance analysis to avoid rubbing, proper alignment of the seal-to-pump interface, and selection of proper elastomeric materials in both composition and geometry. Corrosion is best precluded with proper selection of materials.

Improper assembly, primarily due to inexperienced personnel and lack of trading, is an avoidable failure cause. A mechanical seal is a precision piece of equipment with necessarily critical tolerances and face flatness measured in light bands. Untrained personnel often contribute to seal failure. Internal alignments of the pump-to-seal interface are critical. Successful seals require these approximate tolerances: • Stuffing box mounting face perpendicular to the shaft within 3 mils • Seal sleeve runout as installed < 2 mils • Axial float < 5 mils and preferably as low as 1 mil • Shaft runout, 2 mils TIR and shaft lift < 3 mils

Seal failures related to pump hardware include excessive shaft deflection, small-diameter stuffing boxes, excessive axial shaft movement capability, and piping loads that distort the pump case. The majority of main feed water pumps have small bearing spans and are more rigid with less deflection than most pumps in other services.

API 610 (Reference 33) specifies a maximum deflection of .002 inches at the seal face location. Few existing pump designs have a better volume within a stuffing box. A larger stuffing box diameter and, therefore, larger volume allows the seal to run cooler.

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Cooler running seals are longer lived seals. Shaft axial movement induces fretting between parts, especially those that function as the secondary seal. Pump manufacturers often allow 3 to 8 mils axial movement. The seal designer would prefer a Emil axial movement allowance to preclude excessive fretting possibilities. When piping loads are excessive, distortion to the pump case can occur. Pump case distortions can cause dimensional relationships beyond those suggested as maximums in this guide.

One problem area is ensuring the seal is concentric to the shaft and stuffing box, and that unnecessary static loads are not applied to the seal during installation. The weight of the seal alone can be a problem when resting on the shaft prior to bolting to the stuffing box.

When installing a seal, it should be placed on the shaft and supported as best as possible under the space constraints for these applications. Guide studs of an appropriate length should be used on the upper half of the seal gland. The lower half bearing housing can be installed with only the bearing in place and without seal chamber covers or oil seals in place. The shaft is statically in a position that is close to its operating position. One guide stud at a time may be removed and replaced with the proper seal gland fastener or cap screw.

These cap screws can be slightly snug to hold the seal in position. In some cases, the bearing housing needs to be removed to facilitate final installation of the lower gland cap screws. The shaft should be supported in this position prior to removing the bearing housing. This is accomplished by means of a jack stand, if on the coupling end of the pump, or from overhead with a nylon or Kevlar sling on the thrust end.

With the bearing housing removed, installation of the other capscrews can be performed and torqued. Care must be taken not to thrust the shaft, if possible. Care of the O-rings is a must.

If there is a step or small radius of groove on the shaft where the shaft makes a transition from the bearing journal to the journal area of the shaft where the seal rides, care must be taken to ensure that the O-ring is not damaged when passing over this area. A strip of electrical tape stretched tightly over this area or thin plastic shim stock can be used in this transition.

Once the seal is in place, the bearing housing with all attachments can be installed. The seal piping might have to be installed before the bearing housing is placed on the pump for the final time.

The drive mechanism in most cases is a drive collar which is set screwed to the shaft, external to the seal.

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NOTE: When replacing the mechanical seal drive collar, new set screws of the cup point type should be used unless otherwise specified. Some seal and pump manufacturers drill the shaft where the set screws are to be set. This prevents the seal from moving axially along the shaft in case the screws do not engage the shaft with sufficient holding power.

If the shaft has an area repaired by a chroming process, extra care must be taken to ensure that the set screws engage the shaft with the proper holding force. The chrome is hard and a cup point screw might flatten at the point before the screw embeds in the shaft.

One of the problems with drilling holes in the shaft to locate the set screws is the accuracy of the seal tension and the accessibility of this area to drill exactly on location and perpendicular to the shaft. A drilled hole offset .015 to .020 inch can be critical to the proper seal setting. Drilling more than one hole adds to the problem because none of the holes will be in the same location axially. This can induce an axial eccentricity or runout to the drive collar.

A multistage pump with impellers mounted in series versus impellers mounted in opposed directions has a definite thrust position. The mechanical seal setting or tensioning should be performed with the shaft thrusted to this position. If this procedure is not performed, the rotor unloads one seal while loading the other.

NOTE: Never set the seals without having the thrust bearings in place.

5.5 Alignment

5.5.1 Cold Alignment

The pump should be aligned with the oil system on and the pump full of water. The alignment should take place in stable conditions with temperatures constant. If the alignment cannot take place with the pump and piping loaded with water, then prepa- rations for stops to be placed in all spring supports must be made prior to draining the pumps. See Section 5.3.1 for cold alignment methods.

5.5.2 Hot Alignment

There is no substitute for a comprehensive hot alignment check to show exactly how the pump is operating under all conditions. Aligning to close tolerances is not the same as aligning to the targets and the hot running position.

Several types of alignment monitoring equipment are available. A laser computer-based system is excellent for all-around monitoring and can capture and update data as frequently as every 10 seconds. Optical equipment can be used to give accurate results, but

5-40 EPRI Licensed Material Main Feedwater Pump Maintenance Guide can sometimes have problems with horizontal data. One misconception is that horizontal misalignment is not a problem. In fact, this might require the most attention.

Benchmark bars (gauges) are another method for accurate results. The dynalign bar system yields hardcopy data when used with a chart recorder. Another method is an instrumented coupling, but these are expensive and can only be used on the pieces of machinery for which the coupling is designed.

On turbine-driven feedwater pumps, the pump is usually the m ovable unit. A few utilities prefer to move the turbine, but this is not recommended. In many instances, when the hold-down bolts are loosened, the pump lifts from the base at one or more of the feet. This is referred to as piping-induced soft foot. If the pump is water solid or the spring stops are in place and the pump lifts from the base, nozzle loads are typically excessive. 5.6 Lube Oil System Upgrades

5.6.1 Introduction

Section 3 of this guide identifies the lube oil system as the m ost frequently failed main feedwater pump component. API Standard 614, Lubrication, Shaft-Sealing, and Con- trol-Oil Systems for Special-Purpose Applications (Reference 34) is based on the accu- mulated knowledge and experience of manufacturers and users of lube oil systems for critical rotating machinery (large turbines, motors, gears, compressors, pumps, etc.) in oil refinery service. The lube oil system requirements for this refinery equipment are essentially the same as those for main feedwater pumps and their drivers. Long life and extended periods of reliable uninterrupted operation are required. Reference 34 specifies stainless steel lube oil piping. Several plants have successfully replaced carbon steel, brass, or copper lube oil piping with stainless steel. Such upgrades appear to be good practice.

The following minimum requirements for lube oil piping are extracted from Reference 34.

5.6.2 Stainless Steel Piping

Pipe ASTM A 312 type 304 or 316 stainless steel, seamless. Schedule 10S and 40S may be electric fusion welded. Schedules 80S and heavier shall be seamless.

Tubing ASTM A 269 type 304 or 316 stainless steel.

Valves Class 150 flanged or wafer class 800 SW, TE carbon steel. Gate and globe valves shall have bolted bonnets and bolted glands. Valves larger than 2 inches shall be flanged.

Pipe fittings ASTM A 182 F304 forged Class 3000 SW , TE. ASTM A 403 WP304 wrought butt weld. Threaded fittings shall be seal welded.

T ube fittings Stainless steel.

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Flanges Carbon steel slip-on or stainless steel weld neck or slip-on, ASME (or ANSI) B 16.5.

Gasket Flat

Flange bolting ASTM A 193 Grade B7 bolts and ASTM A 194 Grade 2H nuts

To minimize the use of flanges and fittings, piping should be fabricated by bending and welding. W elded fittings should be butt welded or socket welded. The root pass of all butt welds on stainless steel pipe should be made by tungsten inert-gas arc welding. Filler passes may be made by tungsten inert-gas arc welding or by the shielded metal arc process

Normally, piping, valves, and fittings should be at least nominal pipe size (NPS) 3/4. However, piping that is at least NPS 1/2 or tubing that has an OD of 1/2 inch is permitted between instrument takeoff valves, valves adjacent to instruments, and the instruments. This size is also permitted for atmospheric oil bleed lines returning to the reservoir from instruments and actuator heads.

Minimum pipe wall thicknesses should be:

NPS (Inches) Minimum Schedule 1 or less 80S 1 1/2-3 40S 4 or more lOS

Where space does not permit the use of 3/4-or 1-inch pipe, seamless steel tubing may be furnished. Minimum tubing wall thicknesses should be:

a Nominal Tubing Sizes M inim um W all T hickness, Inches 1/4 b 0.035 b 3/8 0.035 1/2 0.065 3/4 0.095 1 0.109 a The tubing size is the OD in inches.

bThe sizes 1/4 and 3/8 are permitted for instrument and control air only.

Pipe threads should be taper threads in accordance with ASME B1.20.1. Flanges should be in accordance with ASME (or ANSI) B16.5. For socket-welded construction, a 1/16- inch gap should be left between the pipe end and the bottom of the socket to prevent cracking during welding. Pipe ends should be prepared and the joints assembled to minimize potential sources of contamination in the completed assembly.

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Seal-welded joints should be made in accordance with ASME B31.3.

Carbon steel slip-on flanges (internally and externally welded to the pipe to minimize exposure of the carbon steel to the oil), stainless steel slip-on, or weld-neck flanges should be used. Van Stone or lap joint flanges should not be used.

Pipe bushings, unions, and couplings should not be used.

Nonconsumable backup rings and sleeve-type joints should not be used. Pressure piping downstream of oil filters should be free from internal obstructions that could accumulate dirt. Socket-welded pipe fittings should not be used in pressure piping downstream of oil filters.

Combination stop/check valves should not be used.

Gaskets and packings for flanges, valves, and other components should not contain asbestos.

Butterfly valves should not be used.

Oil drains should be sized to run no more than half full w hen flowing at a velocity of 1 foot per second and should be arranged to ensure good drainage (recognizing the possibility of foaming conditions). Horizontal runs should slope continuously at least 1/2 inch per foot toward the reservoir. If possible, laterals (not more than one in any transverse plane) should enter drain headers at 45° angles in the direction of flow. The minimum size for most oil drains should be NPS 1 1/2. The minimum size for inner seal oil drains should be NPS1.

Flexible sections may be used in oil drain and pump suction piping. The flexible sections should not be more than 18-inches long and consist of Series 300 stainless steel annular corrugated close-pitch bellows overlaid with Series 300 stainless steel braid. The flexible sections should be welded, not brazed, to the end fittings.

5.6.3 Piping Design

The design of piping systems should achieve the following: • Proper support and protection to prevent damage from vibration or from operation and maintenance • Proper flexibility and normal accessibility for operation, maintenance, and thorough cleaning • Installation in a neat and orderly arrangement adapted to the contour of the machine without obstructing access openings or presenting tripping hazards • Elimination of air pockets • Complete drainage through low points without disassembly of piping

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Piping design and joint fabrication, examination, and inspection should be in accordance with ASME B31.3. Welding of piping should be performed by operators who are qualified in accordance with Section IX of the ASME Code and who use procedures qualified in accordance with Section IX of the ASME Code.

5.6.4 Piping Inspection

Piping welds may be inspected by radiography or liquid penetrant inspection.

Radiography

Radiography should be in accordance with ASTM E 94 and ASTM E 142.

The acceptance standard used for welded fabrications is Section VIII, Division 1, UW - 52, of the ASME Code.

Liquid Penetrant Inspection

Liquid penetrant inspection should be in accordance with Section V, Article 6, of the ASME Code.

The acceptance standard used for welded fabrications is Section VIII, Division 1, Ap- pendix 8, of the ASME Code.

5.6.5 Hydrostatic Test

The assembled oil system should be subjected to hydrostatic test at a minimum of 1.5 times the maximum allowable working pressure, using oil compatible with the system oil.

5.6.6 Lube Oil Pumps

The oil system should include a main oil pump and a standby oil pump. The main and standby pumps should both be suitable for continuous operation.

Oil pumps not submerged inside the reservoir should be equipped with: mechanical seals that have carbon rings with mating tungsten or silicon carbide rings; Viton, Neo- prene, or Buna-N (or equivalent) gaskets and O-rings; and end plates with throttle bushings as outlined in API Standard 610 (Reference 33).

Rotary pumps should conform to API Standard 676.

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5.6.7 Lube Oil Coolers Twin coolers should be provided and piped in a parallel arrangement using a continuous-flow transfer valve.

Cooling water systems should be designed for the following conditions:

Design Parameter Design Valve Velocity over heat exchange surfaces 5-8 ft/s Maximum allowable working pressure >75 psig Test Pressure >115 psig Maximum pressure drop 15 psi M aximum inlet tem perature 90°F Maximum outlet temperature 120°F Maximum temperature rise 30°F Minimum temperature rise 20°F Fouling factor on water side 0.002 hr-ft2 °F/Btu

Provisions should be made for complete venting and draining of the system.

NOTE: The criteria for minimum temperature rise and velocity over heat exchange surfaces may result in a conflict. The criterion for velocity over heat exchange surfaces is intended to minimize waterside fouling; the criterion for minimum temperature rise is intended to minimize the use of cooling water.

Each oil cooler shall maintain the lube-oil supply temperature at or below 120°F. Each cooler should be of a water-cooled, shell-and-tube type or of a suitable air-cooled type. A removable-bundle design is required for shell-and-tube coolers with more than 5 square feet of surface. Removable-bundle coolers should be in accordance with TEM A Class C and constructed with a removable channel cover. Tubes should have an OD of at least 5/8 inch, and the tube wall thickness should be at least 18 Birmingham wire gauge (BWG) (0.049 inch). Water should be on the tube side of the coolers. Each cooler should be sized to accommodate the total cooling load. To prevent the oil from being contaminated if a cooler fails, the oil-side operating pressure should be higher than the waterside operating pressure. Oil coolers should not be located inside the reservoir.

U-bend tubes are permitted.

Unless otherwise specified, cooler shells, channels, and covers should be made of steel; tubesheets should be made of naval brass and tubes of inhibited admiralty.

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When fouling or freezing of the water side of a cooler is a factor, an oil bypass line around the cooler should be included to regulate the oil supply temperature (in no case should oil bypass the filter). A flanged, pneumatically operated (air-to-open) temperature-control valve may be provided to bypass oil around the cooler to regulate temperatures, but the oil m ust m ix with the cooled oil ahead of the oil filters. Failure of the control valve causes all of the oil to pass through the cooler. The valve should be provided with a manual override that permits operation independent of temperature conditions. The control valve should be sized to handle all oil flow passing through the cooler with a pressure drop equal to or less than the pressure drop through the-cooler.

The minimum design pressure for coolers should not be less than the maximum operating pressure of the system or less than the relief valve setting for the positive displacement pumps.

Both the water side and the oil side of the cooler should be self-venting and self-draining or completely drainable (provided with valved vent/drain connections).

5.6.8 Lube Oil Piping Flanges

Maintenance of stainless steel lube oil piping flanges is detailed under Lube Oil Piping Flanges in Section 5.7.2 of this guide.

5.7 Pressure Retaining Joints

5.7.1 Main Casing Joints

A xially Split Main Casing Joints

Twelve nuclear plants are known to have pumps with axially (horizontally) split main casings with thin (generally 0.015-inch thick) gaskets between the main casing flanges. These pumps are identified as Category 3 in Section 1.6 of this guide. Section 3 of this guide calls these casing flat gaskets, and identifies leakage of these joints as a significant problem. A recent industry study states that over 60 main feedwater pump leakage events caused a unit derating or outage extension in 1994, and that mating-surface leakage has been identified as a dominant failure mode in main feedwater pumps.

It is generally agreed that a primary root cause of the current problems is the change to asbestos-free gasket materials. There is nothing available today that is a 100% satisfactory replacement for the asbestos-based gasket materials for which these pumps were designed. Nuclear main feedwater pump operating temperatures are near the upper limit for Category 3 pumps.

5-46 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Some helpful maintenance practices are as follows: • If a main casing joint is not leaking and has never leaked (except for short periods on startup or after a temperature transient), leave it alone until the predictive maintenance program indicates that it is necessary to break the joint. • The gasket materials most often recommended by pump manufacturers for this service are Garlock Blue-Guard, 3400 and Flexitallic, Permanite, SF 3500. Durabla, Durlon, 8600 and Anchor Packing, 443A may also be suitable. • The gasket material must be cut with a sharp knife to fit the complex shape of the pump main casing flange sealing surfaces. Do not cut by hammering the gasket material against the edges of the flange. Leave extra gasket material extending beyond the machined surfaces at each end. When the pump is assembled, these surfaces are used as sealing surfaces for static seals. After the two casing halves are bolted together and torqued, dim the gasket flush with these surfaces. • Flange sealing surfaces should have roughness between 125 RMS and 220 RMS (125 RMS to 175 RMS preferred with Blue-Guard and 180 RMS to 220 RMS with Permanite are values furnished by the pump vendors). The pattern of the surface finish is also important. Tool marks should not provide potential leak paths. • Flange sealing surfaces must be flat and not warped, scratched, wire-drawn, or washed out (eroded). Minor imperfections can be repaired with Belzona, 1111, and Belzona, Release Agent 9400 (Molecular Systems of New England, Inc.), or equal. The repaired area generally requires grinding and/or sanding to obtain proper flatness and surface finish. Belzona can provide instructions and/or technical representation. • Flatness should be checked side-to-side, end-to-end, and diagonally with a straight edge. Sealing surfaces must not be concave more than 0.002 inch. There is a pressure-sensitive paper that changes to a predetermined color when compressed, making a useful diagnostic tool. • Some casings are deliberately crowned. Either the top or bottom half is high m the middle and slightly tapered toward each side. Typical taper is 0.0005 inch per inch or less. This provides more gasket squeeze toward the center of the pump where the high-pressure water resides. Crowning is generally believed to be beneficial but not necessary in most cases. • Temporary repairs of flat gasket leaks with sealants are commonly used to maintain pump operability until the next planned outage. The NMAC Static Seals Maintenance Guide, TR-104749 (Reference 19) states:

Use injectable coating to repair defective seals on an interim basis when the joint cannot be readily removed from service. These coatings are injected into the sealing area to replace portions of a gasket that are defective or have been damaged. This injection is performed either through the bolting or by installing injection points in the flange or canister. Injection of coatings such as Furmanite or Leak Repair is performed only wheel other methods of seal repair have failed. • The use of injectable coatings is discussed in more detail in Section 5 of Reference 19.

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• Use vendor-recommended thread lubricant, torque values, and torquing sequences when assembling the joint. If the vendor recommends torquing sequences that do not start from the center of the pump and work outward toward the sides and ends, or if the vendor recommends less than four passes to torque the bolts, the vendor should be questioned. • Pocket milling of the reliefs in the top or bottom half of the pum p case to increase gasket compression has been tried. The reviews are mixed, but most sources believe that it is beneficial. • Adaptation of the principal of live loading to the bolting on these casing joints could minimize leakage during thermal transients, but implementation would be a chal- lenge. The principal of live loading is explained in Section 5.3.1 of Reference 16.

Radially Split Main Casing Joints

Pumps with radially split main casings are identified as Category 1 and 2 (Tables 1-1 and 1-2) in Section 1.6 of this guide. These joints are sealed with fully confined, circular, spiral wound gaskets.

Reference 19 states, in part:

Spiral wound gaskets are used extensively throughout most power plants. This type of gasket is used primarily with raised-face and tongue-and-groove-style flanges. Spiral wound gaskets are constructed as alternate plies (circular layers counted as revolutions) of preformed metal windings and pliant fillers that are spiral wound. The pliant material is essentially gash with, but trot below, the metal windings on both contact surfaces. Tide metal strip in the winding is between 0.006 and 0.009 inch in thickness with the width depending on the gasket thickness. The metal windings might be straight, V or U shaped, W shaped, or other form depending on the manufacturer. Both the first three and the last three windings of the gasket are wound without filler material. The use of these metal windings are to prevent the gasket from unraveling. The first two inner windings are welded with a minimum of three spot welds spaced at a maximum distance of 3 inches. The outer windings are welded with a minimum of three welds, the last being the termi- nal weld. Fabrication from metal windings might be from carbon steel, 304, 304L, 309, 310, 316L, 347, 321, 430 stainless steel, Monel 400, nickel 200, titanium, 20 Cb-3 Alloy, Hastelloy B Grade B2, Hastelloy C Grade C-276, Inconel 600, Inconel 625, Incomes X- 750, Inconel 800, Inconel 825, or Zirconium. The filler material may be Chrysotile Asbestos, PTFE, Mica-graphite,flexible graphite, or ceramic.

Asbestos filler is not available in the U.S.

The horizontal width of the gasket cavity is critical because it controls gasket compression. It must be m aintained w ithin OE M -recom mended tolerances. Spiral wound gaskets should be purchased from the OEM or to OEM specifications.

Gasket sealing surfaces can be smooth, or machined with concentric grooves or phono- graphic grooves. Smooth sealing surfaces are typically finished to 125 to 200 RMS.

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Never polish a blemish out by hand. This leads to a depression in the flange surface and loss of sealing capabilities. If the blemish needs to be removed, refinish the entire flange surface. Special Wachs Plange Facers, or similar special equipment might be available for this job.

Reference 19 states:

Gasket enhancers are typically not recommended by gasket manufacturers. However, wheel time constraints prevent a permanent repair from being performed, they may allow sealing of a joint on a temporary basis. Gasket enhancers may be in the form of a thin film of graphite, such as Graphoil, or a form-in-place gasket compound such as Permatex or Dow Corning sealant. Flange facings may be improved by filling pits and voids with epoxy materials such as Liquid Steel or Belzona. Clean the flange of all corrosion prior to the application of these materials. Failure to completely remove the corrosion products call result in lack of adhesion to the metal surface and the opening of another leak path. Only use these materials on low-temperature, low-pressure systems. Evaluate epoxy material to ensure that it is compatible with both the flange material and the gasket. As noted, gasket enhancers are not typically recommended and, at best, are only a temporary treasure. If tile joint in question has proved to be a chronic problem, consider changing the type of gasket material altogether.

Although the quoted paragraph clearly states that these materials should only be used for low temperatures and low pressures, reports indicate that the use of Belzona and Furmanite m Category 3 pumps is not unusual.

5.7.2 Flanged Piping Joints

The NM AC Bolted Joint Maintenance and Application Guide, TR-104213 (Reference 16) contains detailed information on the proper design, assembly, preload, and inspection of bolted joints, including flange piping joints. Of the 16 sections m Reference 16, Sec- tion 3, Assembly has been selected for reproduction as Appendix I for the convenient use of readers of this guide.

Pump Suction and Discharge Flanges

Many main feedwater pump casing suction and discharge nozzles are welded into the piping, but many others have ANSI B16.5 flanged connections. The vast majority of the information required to maintain these flanged joints is in Appendix I. Additional information is in References 16 and 19. Maintenance of these flanged joints is identical to maintenance of other flanged joints in feedwater piping, except that they are subjected to direct vibration from the pump casing. Effects of vibration on flanged piping joints is covered in this section of the guide.

A uxiliary W ater P iping Flanges

Auxiliary water piping is sm aller, but m aintenance of these flanged joints is basically the same as it is for pump suction and discharge flanges. This piping is subjected to pump case vibration and might have additional vibration from excited resonant frequencies. 5-49 EPRI Licensed Material Nuclear Maintenance Applications Center

Lube Oil Piping Flanges

Stainless steel lube oil piping with ASME (or ANSI) B 16.5 flanges is recommended for main feedwater pumps. This piping is similar to the auxiliary water piping, but is not subjected to high pressures. Appendix I contains the majority of the information required for maintenance, with additional detail m References 16 and 19. This piping is also subjected to direct vibration and to piping resonances.

Effects of Vibration on Flanged Piping Joints

The following is an edited extraction from Reference 16:

Vibration Loosening

Conditions for Loosening

Vibrations can loosen bolts and often causes a complete loss of the nut and bolt. A1- though various explanations are given for bolt loosening, none fully explain all the problem s encountered in the field. The following conditions typically exist when a nut is fully loosened under vibration: • The vibratory force has components at right angles to the axis of the bolts. This force causes the joint members to slip past each other. Vibratory forces parallel to the bolt axis partially loosen the bolt, causing a 10-20% loss of the initial preload. Only right- angled forces loosen the nut. • There is slip clearance between the male and female threads and between the bolts and joint members. This allows a transverse slip to occur.

If the above conditions exist, severe slip cycles (e.g., thermal cycles, flexing of joint members) might fully loosen the nut. More commonly, hundreds or thousands of vibration-induced micro-slip motions can fully loosen a nut.

Minimizing Loosening

A number of techniques minimize self-loosening. The following methods are listed in order of increasing cost or complexity: • Increase thread and joint friction forces. If these forces are high enough, no transverse slip will occur. Friction can be increased by the following: — Increasing preloads. — Compensation for the relaxation of the fasteners; for example, use an extra pass with the wrench. This increases residual preload and results in higher friction forces. • Use anaerobic adhesives to glue the male and female threads together. These are available for operating temperatures up to 450° F.

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• Add collars under the bolt head and nuts and use longer bolts. Some practitioners claim that a bolt with a length-to-diameter ratio of 8-1 or more never loosens. • Use bolts with fine threads instead of coarse threads. The slope of the helix angle for fine threads is smaller than for course threads. This smaller slope help reduce loosening. • Reduce the amount of vibration m the joint. One way to do this is to use shock absorbers and/or shock mounts to dampen vibration. • Have a vibration specialist evaluate the system to determine the causes and possible cures for the vibration. • R edesign the joint so that the axis of the fasteners is parallel to the direction of vibration, rather than perpendicular to it.

Additional notes: • These bolted joints might be subjected to millions of vibration-induced micro-slip motions (not just hundreds or thousands). • Adding collars under the bolt head and nuts and using longer bolts is also known as live loading.

5.8 Mechanical Face Seals

5.8.1 Mechanical Seal Upgrades

Mechanical shaft seals (face seals) were identified in Section 3 of this guide as a major cause of nuclear feed pump failures. Maintenance and upgrading of mechanical seals, and the environment in which they operate, are major areas for improvement. For material descriptions of mechanical seals and their operating environments, see Sections 2.21 and 2.22 of this guide.

5.9 Nozzle Loads Problems resulting from forces and moments induced in the pump casing through the effects of nozzle loads are misalignment and associated problems. These include distortion of the casing affecting critical clearances, stresses to the casing that can result in cracking or fractures of the casing, and stresses applied to the base, hold-down bolts, and piping.

When excessive forces and moments are applied to a pump, the pump can be moved from its intended position. If the pump is being held in position by keys at the bottom of the casing, these forces can move the pump and result in stress and casing deformation. This, in turn, can result in damage to the sleeves, bushings, and bearings. Coupling failure can also result.

To determine if excessive nozzle loads are present, several checks can be performed. First, acquire a set of readings with dial indicators or a laser alignment system (see

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Section 5.3.1 for cold alignment methods). Loosen the hold-down bolts and any jacking bolts that might be against the feet. Take another set of readings and compare to the initial set. This indicates which direction the pump is moving and to what extent. If keys or pins are installed on the underside of the pump, this might limit the displacement of the pump.

Prior to draining the system, all spring-type supports should have the stops installed. If the spring stops for a particular spring do not go into the slot because they are too long or too short, new stop material should be cut and inserted. If the pump is equipped with keys on each end, these keys should be removed. If the pump has a key on one end, and a pin on the other, the key should be removed and the pin bracket cut loose from its base so as not to restrict pump movement. Loosen the hold-down bolts and monitor the side-to-side displacement, for and aft displacement, twist, and rise. Record all displacements. Strain gauges, load cells, and torque wrenches are all good tools to use to determine the amount of force required to move the pump back in position. The problem should be evaluated and corrections made. The pump manufacturer, a repu- table independent firm, and piping engineers should be consulted for resolution.

A hot alignment analysis should be performed during cool down of the pump. A portion of this force might be negated if hot alignment analysis shows a large movement in the direction of the forces. Adjustable piping struts can help if a proper piping analysis proves there is a margin of adjustment.

5.10 Pump Startup After final assembly of the pump and prior to warming, the oil system should be operated and checked for leaks. This helps to filter out any contaminants in the housing or oil supply.

W hen filling the pump for the first tim e on a pum p that has been totally disassem bled and rebuilt, the seal injection system should be used if the pump has fixed-breakdown- type throttle bushings. This assists in forcing sm all particles inw ard to the pum p instead of working their way through tight clearances in the bushings and sleeves.

With the rotor in the center of the casing, rubs that could cause major damage to the internals of the pump and prevent startup are less likely to occur. Galling, scoring, and seizure might occur during the warmup period of the pump. Casing deformation due to thermal differentials and/or nozzle loads on these types of pumps are problems that must be recognized and dealt with during startup.

Due to the short length of the rotors on the majority of these type pumps, the rotor deflection is not a major problem (casing deformation is a concern).

A temperature differential between the upper and lower casing and/or fore and aft m ust be at minim um to ensure that rubs do not occur. In some cases, if turbine-driven, the turning gear motor trips on a high amperage draw if a casing-to-rotor rub occurs.

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This prevents the unit from coming on line when desired and results in a possible pump tear down, again to replace parts in the pumps.

5.11 Pump Warming

Proper pum p warm ing is one of the m ost critical steps in bringing a m ain feedwater pump on line. A thorough analysis of warming procedures assists in determining if this is a problem. M any manufacturers provide users with temperature monitoring capabilities at the pump casing to ensure that temperatures have stabilized within the pump casing prior to rotation of the pump. This is typically provided by thermocouples installed in the upper and lower casings of the pump at strategic locations to monitor temperature.

Many pumps are subject to hot and cold spots due to the design of the casing. Stratifica- tion of the liquid can leave temperature differentials in areas that might cause rubbing between the stationary and rotating parts. An acceptable differential in temperature is given, generally between 20 to 50°F for these readings. Smaller differential temperatures prior to startups are beneficial to the continued operation and reliability of the pump.

Motor-driven machines should not be started until thoroughly warmed. Turbine driven machines should not be placed on turning gear until stable temperatures have been reached.

Some manufacturers specify that the pump casing be within a specified temperature of the pump. This may not be adequate for your particular installation or startup procedures. Consider the installation of temperature monitoring devices if the pumps are not so equipped.

One area of concern is bringing a repaired pump on line while the unit is operating at temperature. Pump warming becomes more complex. Large isolation valves are notori- ous for leaking, and there is heat transfer through the piping. The pump casing can remain at an elevated temperature while work is being done. The dimensional checks and rotation of the shaft during assembly become difficult. Contact between the rotating and stationary parts can occur.

Temperatures must be monitored closely to ensure that the pump is not rotated prema- turely. Deformations such as the ones depicted m Figure 5-10 can be severe enough to lock the rotor for several hours until temperatures stabilize in the casing and rotor.

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Figure 5-10 Potential Casing Deformation During Warming

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APPENDIX A TYPICAL SECTIONAL DRAWINGS OF EACH MAIN FEEDWATER PUMP TYPE Drawings are grouped by original manufacturer as listed below:

Figure Description Page Bingham

Figure A-1 CD ...... A-3 Figure A-2 CPD...... A-4 Figure A-3 HSB ...... A-5 Figure A-4 MSB...... A-6 Figure A-5 MSD...... A-7 Byron Jackson

Figure A-6 DVSR...... A-8 Figure A-7 HDR...... A-9 Figure A-8 HSB ...... A-10 Figure A-9 DBS ...... A-11 Figure A10 DVMX, Single Suction, First Stage ...... A-12 Figure A-11 DVMX, Double Suction, First Stage...... A-13 Delaval

Figure A-12 1BSX ...... A-14 Figure A-13 1BSX ...... A-15 Figure A-14 2BSX ...... A-16 Figure A-15 4BSX ...... A-17 Ingersoll-Rand

Figure A-16 16x17 CN...... A-18 Figure A-17 18x17 CN...... A-19 Figure A-18 C...... A-20 Figure A-19 CA ...... A-21 Figure A-20 JT ...... A-22

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Figure Description Page Pacific

Figure A-21 HVF, Double Cover...... A-23 Figure A-22 HVCN, Single Cover...... A-24 Figure A-23 BFI...... A-25 Figure A-24 SFI...... A-26 Figure A-25 RHC...... A-27

Worthington

Figure A-26 WGID...... A-28 Figure A-27 WNC, Single Stage, Single Suction...... A-29 Figure A-28 WNCD, Single Stage, Double Sunction...... A-30

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APPENDIX B USER INSTALLATION LISTS

Table Description Page Table B-1 Category 1 Main Feedwater Pump Installations, Radially Split Single Outer Case with Single Stage Double-Suction Impeller...... B-2

Table B-2 Category 2 Main Feedwater Pump Installations, Radially Split Double Outer Case Containing One or More Impellers with First Stage as a Single- or Double-Suction Impeller ...... B-7

Table B-3 Category 3 Main Feedwater Pump Installations, Axial Split Outer Case with a Single Stage Double-Suction Impeller or a Multistage Pump with either a Single-Suction or Double-Suction First Stage Impeller ...... B-10

Table B-4 Plant Listing of Main Feedwater Pump Installation ...... B-12

Table B-5 Bingham Supplied Main Feedwater Pump Installation List ...... B-18

Table B-6 Byron Jackson Supplied Main Feedwater Pump Installation List...... B-20

Table B-7 Deleval Supplied Main Feedwater Pump Installation List ...... B-24

Table B-8 Ingersoll-Rand Supplied Main Feedwater Pump Installation List ...... B-26

Table B-9 Pacific Supplied Main Feedwater Pump Installation List...... B-27

Table B-10 Worthington Supplied Main Feedwater Pump Installation List...... B-28

B-1 Table B-1 Category 1 M ain Feedwater Pump Installations Radially Split Single Outer Case with Single Stage Double Suction Impeller

Site Utility NSSS Pump Mfg. Quantity Size Type Gallons Head Horse- RPMTemperature per Power °F Minute

Beaver Vy alle Duquesne Light WE Bingham 16x18x24BCD 15,200 1900 7,400 3560 384 Company 2

Beaver Valley Duquesne Light WE Bingham 2 16x18x24 CD 15,200 1900 7150 3580 384 Company

Braidwood Commonwealth WE Bingham 6 20x20x17 CD 19,450 2110 9,945 5200 32/410 Edison

Byron Commonwealth WE Bingham 6 20x20x17 CD 19,450 2110 9,945 5200 32/410 Edison

Catawba Duke Power WE Bingham 4 16x18x17 CD 19,850/ 2250/ 11500 5510 243/375 Com U nit 1 & 2 pany 18,040 2040

Comnanche PeakTexas Utilities WE Bingham 4 16x18x17DCD 19,800 2322 11,584 5200 362 Electric Copan my

N orth A nna Virginia Power WE Bingham 16x18x25 CD 16,250 1980 9000 3560 390 Com 6 Unit 1 & 2 pany

S urry Unit 1 & 2Virginia Power WE Bingham 4 16x18x24 CD 13,800 1700 6,060 3560 377 Company

Trojan Nuclear Portland GeneralWE Bingham 2 16x18x17 CD 19,800 2020 9,964 5200 373 Plant Electric

Arkansas Entergy BW /CEBingham 2 16x16x16 HSB 14,750 1902 7,100 5700 374 Nuclear One Operations, Inc.

Palisades Consumers PowerCE Bingham 2 14x14x16 HSB 13,500 1920 6,600 5000 376 N uclear Plant Company

Calvert Cliffs Baltim ore Gas CE Byron 2-50% 14x14x17BDVSR 15000 2392 8000-T 5200 375 Unit 1 & Electric Jackson

Calvert Cliffs Baltim ore Gas CE Byron 2-50% 14x14x17BDVSR 15000 2316 8000-T 5130 344 Unit 2 & Electric Jackson Table B-1 (continued) Category 1 M ain Feedwater Pum p Installations Radially Split Single Outer Case with Single Stage Double Suction Impeller

Site Utility NSSS Pump Mfg.Quantity Size Type Gallons Head Horse- RPMTemperature per Power °F Minute

Cooper Nebraska Public GE Byron 2-50%14x14x17ADVSR 12800 2500 9500-T 5500 370 Power District Jackson

Crystal River Baltimore Gas BW Byron 2-50%14x14x17BDVSR 13320 2280 8000-T 5120 390 U nit 3 & Electric Jackson

Diablo Canyon Pacific G as & WE Byro n 2-50%14x14x17CDVSR 17350 2000 9600-T S800 365 Unit 1 Electric Jackson

Diablo Canyon Pacific G as & WE Byro n 2-50%14x14x17CDVSR 17350 2000 9600-T 5800 365 Electric Unit 2 Jackson

Hatch Un it 1 Georgia Power GE Byron 2-50%14x14x17ADVSR 12500 2540 8500-T 5550 321 Com pany Jackson

Hatch Unit 2 Georgia Power GE Byron 2-50% 14x14x17ADVSR 12500 2540 8500-T 5550 321 Com pany Jackson

Maine Yankee Maine Yankee CE Byro n 2-50%14x14x17BDVSR 14000 2000 8000-MG5000 378 Atomic Power Jackson

Peach Bottom PECO Energy GE Byron 3-33%14x14x17ADVSR 12270 2310 8000-T 5300 378 U nit 2 Jackson

Peach Bottom PECO Energy GE Byron 3-33%14x14x17ADVSR 12270 2310 8000-T 5300 378 U nit 3 Jackson

Sequoyah Unit 1 Tennessee Valley WE Byro n 2-50% 14x14x17CDVSR 20000 1680 8500-T 5400 399 A uthority Jackson

Sequoyah Unit Tennessee2 ValleyWE Byro n 2-50% 14x14x17CDVSR 20000 1680 8500-T 5400 399 A uthority Jackson

Three M ile IslandGPU Nuclear B&W Byron 2-50% 14x14x17BDVSR 16100 2480 10200-T5500 375 Corporation U nit 1 Jackson

24000 ANO Unit 1 & 2Entergy BW Delaval 2-80% 24 1BSX 1900 13300-T5050 382 O perations, Inc. Table B-1 (continued) Category 1 M ain Feedwater Pum p Installations Radially Split Single Outer Case with Single Stage Double Suction Impeller

NSSS Head Site Utility Pump Mfg. Quantity Size Type Gallons Horse- RPMTemperature per Power Minute °F

Brunswick Carolina Power &GE Delaval 4-50% 16 1BSX 13100 2730 10000-T5450 289 Light D/L

2-50% Davis Besse Toledo Edison BW Delaval 16 1BSX 15000 2150 11500-T5150 301

Farley Unit 1 Alabama Power WE Delaval 2-50% 16 1BSX 15000 2330 10000-T5250 385 Company

Farley U nit 2 Alabama Power WE Delaval 2-50% 16 1BSX 15000 2330 10000-T5250 385 Company

Oconee Unit 1 &Duke 2 Power Co BW Delaval 4-50% 16 1BSX 13200 2200 7600-T 5000

Oconee Unit 3 Duke Power Co BW Delaval 2-50% 16 1BSX 13200 2200 7600-T 5000

Perry Unit 1 & 2 Cleveland ElectricGE Delaval 4-60% 16 1BSX 19215 2415 14000-T5300 367 Illum inating

W PPS Project Washington Public BW Delaval 4-50% 16 1BSX 19323 2500 12062-T5400 410 1 &4 Power Supply System

Zion Unit 1 & 2Comm onwealth WE Delaval 4-50% 16 1BSX 17400 2070 10000-T4920 366 Edison

Zion Unit 1 & 2Comm onwealth WE Delaval 2-50% 16 1BSX 17400 2120 1000-M G5050 366 Edison

Susquehanna Pennsylv ania GE Ingersoll- 6 16x17 CN 12500 5200 384 U nit 1 & 2 Power & Light Rand

V.C. Summer South Carolina WE Ingersoll- 3 14x17 CN 9970 5100 342 Electric & Gas Rand

Indian Point 3 Consolidated WE Ingersoll- 2 16 JT 15300 4875 390 Edison Rand Table B-1 (continued) Category 1 M ain Feedwater Pump Installations Radially Split Single Outer Case with Single Stage Double Suction Impeller

Site Utility NSSS Pump Mfg. Quantity Size Type Gallons Head Horse- RPMTemperature P per ower °F Minute

lndian Point Unit 2 Consolidated WE Ingersoll- 2 16 JT 15300 4865 390 Edison R and

LimerickU nit PECO GE Ingersoll- 6 16 JT 12900 5160 367 1 &2 R and

M illstone Unit 2Northeast Utilities CE Ingersoll- 2 16 JT 15000 5150 380 R and

Browns Ferry 1 Tennessee Valley GE Pacific 3 16X HVCN 11200 2800 8450-T 5500 300 Authority

Browns Ferry 2 Tennessee Valley GE Pacific 3 16X HVCN11200 2800 8450-T 5500 300 Authority

Browns Ferry 3 Tennessee Valley GE Pacific 3 16X HVCN11200 2800 8450-T 5500 300 Authority

LaSalle Unit 1 Commonwealth GE Pacific 2 20x17 HVF 19400 2450 12000-T5500 365 Edison

LaSalle U nit 2 Commonwealth GE Pacific 20x17 HVF 19400 2450 12000-T5500 365 Edison 2

Vogtle Unit 1 Georgia Power Co.WE Pacific 2 20x17 HVF 20400 3000 14980-T5950 403

Vogtle Unit 2 Georgia Power Co.WE Pacific 2 20x17 HVF 20400 3000 14980-T5950 403 W aterford U nit Enter3 gy CE Pacific 1 20x17 HVF 17940 2150 9780-T 5200 366 Operations, Inc.

DC Cook U nit 1Indiana/ WE Worthington 2 24 WGID 9600-T 5220 346 Michigan Power

8900-T Fitzpatrick New York PowerGE Worthington 3 20 WGID 5150 367 Authority Table B-1 (continued) Category 1 M ain Feedwater Pump Installations Radially Split Single Outer Case with Single Stage Double Suction Impeller

Site Utility NSSS Pump Mfg.Quantity Size Type Gallons Head Horse- RPMTemperature Power per °F Minute

Hope Creek Pu blic S ervice GE Worthington 18 WGID Electric & G as 6

Kewaunee Wisconsin PublicWE Worthington 2 16 WGID 5000-MG5130 360 Service Corporation

McGuire Unit Duke Power WE Worthington 4 24 WGID 1 &2 Company

Prairie Island Northern States WE Worthington 2 16 WGID 5000-MG5130 360 Unit 1 Power

Salem Unit 1 P ub lic Service WE Worthington 2 24 WGID 11500-T5530 368 Electric & Gas

Salem Unit 2 P ub lic Service WE Worthington 2 24 WGID 11500-T5530 368 Electric & Gas Table B-2 Category 2 M ain Feedwater Pum p Installations Radially Split Double Outer Case containing One or More Impellers with First Stage as a Single or Double Suction Impeller

Site U tility NSSSPump Mfg. Quantity Size Type NumberGallons Head Horse- RPMTemperature of per Power °F Stages Minute

North Anna Virgin ia Po wer WE Bingham 6 14x16x19 CP 2 14,600 2734 11,000 3560 320 Company U nit 3 & 4

River Bend Gulf States GE Bingham 3 14x14x17CPD 2 11,058 2582 7380 4000 388 Unit 2 U tilities C om pan y

Caraway Unit Union1 Electric WE Byron 2-50% 20x20x18BHDR 1 17260 2387 12000-T5300 336 Company Jackson

Caraway Unit 2 Union Electric WE Byron 2-50% 20x20x18BHDR 1 17260 2387 12000-T5300 336 Company Jackson

Clinton Unit Illinois1 PowerGE & Byron 2-50% 20x20x18BHDR 16545 2483 11000-T5350 390 Li 1 ght Jackson

Clinton U nit Illinois1 PowerGE & Byron 1 -30% 18x18x18HDR 1000 2483 6000-MG5050 390 Li 1 ght Jackson (standby)

Clinton Un it 2 Illinois Power &GE Byron 2-50% 20x20x18BHDR 1 16545 2483 11000-T5350 390 Light Jackson

Clin ton Unit 2 Illinois Power &GE Byro n 1 -30% 18x1 8x18HDR 1 1000 2483 6000-MG5050 390 Light Jackson (stand by)

Grand Gulf UnitEntergy 1 GE Byron 2-50% 24x24x19CHDR 1 20600 2430 15000-T5400 290 Operations, Inc. Jackson

Grand Gulf UnitEntergy 2 GE Byron 2-50% 24x24x19CHDR 1 20600 2430 15000-T5400 290 Operations, Inc. Jackson

M illstone Unit N3 orth east U tilitiesWE Byron 2-50% 20x20x18BHDR 1 19820 2050 1100-T 5250 364 Jackson

M illstone Unit N3 ortheast UtilitiesWE Byron 1 -50%20x20x18BHDR 1 19820 2050 1100-T 5250 364 Jackson (standby) Table B-2 (continued) Category 2 M ain Feedwater Pump Installations Radially Split Double Outer Case containing One or M ore Impellers with First Stage as a Single or Double Suction Impeller

Site Utility NSSS Pump Mfg. Quantity Size Type NumberGallonsHead Horse RPMTemperature of per Power °F Stages Minute

2-50% 14000-T Palo Verde Arizon a Pub lic CE Byron 24x24x19CHDR 22314 2378 5315 347 Service Co. 1 U nit 1 Jackson

2-50% 24x24x19C Palo Verde Arizon a Pub lic CE Byron HDR 22314 2378 14000-T5315 347 Service Co. 1 Unit 2 Jackson

Palo Verde Arizona Public CE Byron 2-50% 24x24x19CHDR 22314 2378 14000-T5315 347 Service Co. 1 Unit 3 Jackson

Pilgrim Unit 2Boston Edison GE Byron 2-50% 24x24x19CHDR 1 19340 2215 12000-T5135 367 Co. Jackson

Seabrook U nit New1 Hampshire WE Byron 2-50% 20x20x18BHDR 1 17150 2350 11OOO-T5300 364 Yankee Jackson

Seabrook UnitNew 2 Hampshire WE Byron 2-50% 20x20x18BHDR 1 17150 2350 11000-T5300 364 Yankee Jackson

W atts Bar UnitTennessee 1 WE Byron 2-50% 20x20x18BHDR 1 17500 2010 10000-T5020 402 Valley Authority Jackson

Watts Bar UnitTennessee 2 WE Byron 2-50% 20x20x18BHDR 1 17500 2010 1000-T 5020 402 Valley Authority Jackson

Wolf Creek Wolf Creek WE Byron 2-50% 20x20x18BHDR 17260 2387 12000-T5300 336 Nuclear 1 U nit 1 O Jackson perating Co.

Nin e M ile PointNiagara M ohaw kGE Byron 3-50% 20x20x22HSB 1 18700 2444 12000-MG3339 370 Power Jackson Unit 2 Cor poration

Monticello Northern StatesGE Delaval 2-50% 14 2BSX 7820 2750 6000-M 3570 Power 2

Delaval 2-33% WPPSS jProect Washington CE 14 2BSX 2 9568 2174 5375-M 3570 150 3 & 5 Public Power Su pply System -

Table B-2 (continued) Category 2 M ain Feedwater Pum p Installations Radially Split Double O uter Case containing One or M ore Im pellers with First Stage as a Single or Double Suction Impeller

Site Utility NSSS Pump Mfg. Quantity Size Type NumberGallonsHead Horse- RPMTemperature of per Power °F Stages Minute

Millstone 1 N orth east UtilitiesGE Ingerso ll- 3 12 C 3 3580 291 Rand

LaSalle Unit 1 CommonwealthGE Pacific 16RX BFI 10300 2110 12000-T5500 360 Edison 1

Peach BottomPECO GE Pacific 2 4X BFI 11 515 4000 657-M 93550 345

Lasalle U nit 2 CommonwealthGE Pacific 16RX BFI 10300 2110 12000-T5500 360 Edison 1

Fitzpatrick New York PowerGE Worthington 2 18 WNC 1 7300-T 4500 371 Authority

Nine M ile PointNiagara M ohaw kGE Worthington 1 18 WNC 1 Unit 1 Power Corp.

Nine M ile PointNiagara MohawkGE Worthington 2 8 WNC 1 Unit 1 Power Corp. Table B-3 Category 3 M ain Feedwater Pum p Installations Axial Split Outer Case with a Single Stage Double Suction Impeller or a M ultistage Pump with either a Single Suction or Double Suction First Stage Impeller

Site Utility NSSS Pump Quantity Size Type NumberGallonsHead Horse- RPMTemperature Mfg. of per Power °F Stages Minute

Haddem NeckConnecticut WEBingham 2 12x12x17 MSB 2 9,600 2000 4,900 3560 372 Yankee Atomic Power Com pany

Vermont Vermont YankeeGE Bingham 3 12x16x20 MSD 7,630 2570 5,400 3560 320 Yankee Nuclear Power 2 Com pany

Dresden UnitCommonwealth 2 GE Byron 3-50% 10x12x16 DVMX 2 10750 2700 9000-MG4500 298 Edison Jackson

Dresden UnitCommonwealth 3 GE Byron 3-50% 10x12x16 DVMX 2 10750 2700 9000-MG4500 298 Edison Jackson

Duane ArnoldIES U tilities GE Byron 2-57% 12x12x18 DVMX 3 8900 2500 6000-M 3570 365 U nit 1 Jackson

Fort CalhounOmaha Public CE Byron 3-50% 10x10x17DVMX 2 8000 1750 3500-M 3570 396 U nit 1 Power District Jackson

Oyster CreekGPU Nuclear GE Byron 3-33% 8x8x14 DVMX 4 5000 2800 4000-M 3570 214 Unit 1 Corporation Jackson

Pilgrim Unit 1Boston Edison Co.GE Byron 3-33% 12x12x18 DVMX 3 6650 2620 5000-M 3560 290 Jackson

St Lucie Florida Power &CE Byron 2-50% 12x14x18 DVMX 2 14100 1780 7000-M 3570 380 Light Company Jackson

St Lucie 2 Florida Power &CE Byron 2-50% 12x14x18 DVMX 2 14100 1780 7000-M 3570 380 Light Company Jackson

Turkey Point 3 Florida Power &WE Byron 2-50% 12x14x18 DVMX 2 13000 1880 7000-M 3570 362 Light Company Jackson

Turkey Point 4Florida Power &WE Byron 2-50% 12x14x18 DVMX 13000 1880 7000-M 3570 362 Li 2 ght Company Jackson Table B-3 (continued) Category 3 M ain Feedwater Pump Installations Axial Split Outer Case with a Single Stage Double Suction Impeller or a M ultistage Pump with either a Single Suction or Double Suction First Stage Impeller

Site Utility NSSS Pump Quantity Size Type NumberGallonsHead Horse- RPMTemperature per Power Mfg. of °F Stages Minute

Point Beach W isconsin ElectricWE Byron 2-50% 12x12x16 DVS 1 7810 2180 5000-MG6000 348 Power Co. Unit 1 Jackson

Point Beach W isconsin ElectricWE Byron 2-50% 12x12x16 DVS 1 7810 2180 5000-MG6000 348 Power Co. Unit 2 Jackson

Watts Bar Tennessee Valley WE Byron 1-20% 12x12x16 DVS 6100 1890 3000-MG5350 388 (standb 1 Unit 1 Authority Jackson y)

W atts Bar Tennessee Valley WE Byron 1-20% 12x12x16 DVS 6100 1890 3000-M G5350 388 (standb 1 Unit 2 Authority Ja ckson y)

Quad Cities CommonwealthGE Pacific 6 16X RHCNDS 2 11050 2573 8180 4500 298 Unit 1 & 2 Edison EPRI Licensed Material Nuclear Maintenance Applications Center Table B-4 Plant Listing of Main Feedwater Pump Installations

Site Utility Mfg. Quantity Size Type Pump Number Category of Stages

ANO Unit Entergy Delaval 2-80% 24 1BSX 1 1 1 &2 Operations, Inc.

Arkansas Entergy Bingham 2 16x16x16 HSB 1 1 Nuclear One Operations, Inc.

Beaver Valley Duquesne Light Bingham 2 16x18x24 CD 1 1 Company

Beaver Valley Duquesne Light Bingham 2 16x18x24B CD 1 1 Company

Braidw ood Commonwealth Bingham 6 20x20x17 CD 1 1 Edison

Browns Tennessee Pacific 3 16X HVCN 1 1 Ferry 1 Valley A utho rity 16X Browns Tennessee Pacific 3 HVCN 1 1 Ferry 2 Valley Authority

Brow ns Tennessee Pacific 3 16X HVCN 1 1 Ferry 3 Valley Authority

Brunswick Carolina Power & Delaval 4-50% 16 1BSX 1 1 Light

Byron Commonwealth Bingham 6 20x20x17 CD 1 1 Edison

Caraway U nio n E lectric Byron 2-50% 20x20x18B HDR 2 U nit 1 Company Jackson 1 2-50% Caraway Un ion Electric Byron 20x20x18B HDR 2 1 U nit 2 Company Jackson

C alv ert C liffs Baltim ore Gas & Byron 2-50% 14x14x17B DVSR 1 1 U nit 1 Electric Jackson

Calvert Cliffs Baltim ore Gas & Byron 2-50% 14x14x17B DVSR 1 1 U nit 2 Electric Jackson

Catawba Unit Duke Power Bingham 4 16x18x17 CD 1 1 1 and Unit 2 Company

Clinton Un it 1 Illinois Power Byron 2-50% 20x20x18B HDR 2 1 & Light Jackson

Clinton Un it 1 Illinois P ower Byron 1-30% 18x18x18 HDR 2 1 & Light Jackson (stand by) 2-50% Clin to n Un it 2 Illino is P ower Byron 20x20x18B HDR 2 1 & Light Jackson

Clin ton Unit 2 Illinois Po wer Byron 1-30% 18x18x18 HDR 2 1 & Light Jackson (stand by)

Comnanche Texas Utilities Bingham 4 16x18x17D CD 1 1 Peak Electric Company

Cooper Nebraska Public Byron 2-50% 14x14x17A DVSR 1 1 Power District Jackson B-12 EPRI Licensed Material Main Feedwater Pump Maintenance Guide Table B-4 (continued) Plant Listing of M ain Feedwater Pump Installations

Site Utility Mfg. Quantity Size Type Pump Number Category of Stages

Crystal River Florida Power & Worthington 1 10 WC 1 U nit 1 Light Company C rystal R iv er Baltim ore Gas Byron 2-50% 14x14x17B DVSR 1 1 Unit 3 & Electric Jackson Davis Besse Toledo Edison Delaval 2-50% 16 1BSX 1 1

DC Cook Indiana/Michigan Worthington 2 24 WGID 1 1 U nit 1 Power Byron Diablo Canyon Pacific Gas 2-50% 14x14x17C DVSR 1 1 Unit 1 & Electric Jackson Diablo Canyon Pacific Gas Byron 2-50% 14x14x17C DVSR 1 1 Unit 2 & Electric Jackson Dresden Commonwealth Byron 3-50% 10x12x16 DVMX 3 2 Unit 2 Edison Jackson Dresden Commonwealth Byron 3-50% 10x12x16 DVMX 3 2 Unit 3 Edison Jackson Duane A rnold IE S Utilities Byron 2-57% 12x12x18 DVMX 3 3 U nit 1 Jackson Farley U nit 1 Alabama Power Delaval 2-50% 16 1BSX 1 1 Company

Farley Un it 2 Alabama Power Delaval 2-50% 16 1BSX 1 1 Company

Ferm i Un it 2 Detroit Edison Delaval 2-65% 24 1BSX 1 1 2-65% Ferm i Unit 2 D etroit E dison Delaval 24 1BSX 1 1

Fitzp atrick New York Power Worthington 3 20 WGID 1 1 Authority

Fitzpatrick New York Power Worthington 2 18 WNC 2 1 Authority Fort Calhoun Om aha Public Byron 3-50% 10x10x17 DVMX 3 2 Unit 1 Power District Jackson Byron 2-50% Grand Gulf Entergy 24x24x19C HDR 2 1 U nit 1 Operations, Inc. Jackson Grand Gulf Entergy Byron 2-50% 24x24x19C HDR 2 1 Unit 2 Operations, Inc. Jackson

Haddem Neck Connecticut Bingham 2 12x12x17 MSB 3 2 Yankee Atomic Power Company

Hatch Unit 1 Georgia Power Byron 2-50% 14x14x17A DVSR 1 1 Company Jackson Hatch Unit 2 Georgia Power Byron 2-50% 14x14x17A DVSR 1 1 Company Jackson

B-13 EPRI Licensed Material Nuclear Maintenance Applications Center

Table B-4 (continued) Plant Listing of Main Feedwater Pump Installations

Site Utility Mfg. Quantity Size Type Pump Number Category of Stages

Hope Creek Pub lic S ervice Worthingto n 6 18 WGID 1 1 Electric & Gas

Ind ian Po int 3 Consolidated Ingersoll- 2 16 JT 1 1 Edison Rand Indian Point Consolidated Ingerso ll- 2 16 JT 1 1 Unit 2 Edison Rand Wisconsin Public Kewaunee Service Worthington 2 16 WGID 1 1 Corporation LaSalle Unit 1 Commonwealth Pacific 2 20x17 HVF 1 Edison

LaSalle Unit 1 Commonwealth Pacific 1 16RX BFI 2 Edison

LaSalle Unit 2 Commonwealth Pacific 2 20x17 HVF 1 Edison

L asalle Un it 2 C om m onw ealth Pacific 1 16RX BFI 2 Edison

Limerick Unit PECO Ingerso ll- 6 16 JT 1 1 1 &2 Rand Byron Maine Y ankee Maine Yankee 2-50% 14x14x17B DVSR 1 1 Atomic Power Jackson

McGuire Unit Duke Power Worthington 4 24 WGID 1 1 1 &2 Company

M ills tone 1 Northeast Utilities Ingersoll- 3 12 C 2 3 Rand

M illstone Northeast Utilities Ingerso ll- 2 16 JT 1 1 Unit 2 Rand

Millstone Northeast Utilities Byron 2-50% 20x20x18B HDR 2 1 Unit 3 Jackson

M illstone Northeast Utilities Byron 1-50% 20x20x18B HDR 2 1 Unit 3 Jackson (standby) M on ticello N orthern States Delaval 2-50% 14 2BSX 2 2 Power

N ine M ile Niagara M ohaw k Worthington 1 18 WNC 2 1 Point Unit 1 Power Corp.

Nine M ile Niagara M ohaw k Worthingto n 2 8 WNC 2 1 Point Unit 1 Power Corp.

Nine M ile Niagara Mohawk Byron 3-50% 20x20x22 HSB 2 1 Point Un it 2 Power Corp. Jackson

North Anna Virgin ia Pow er Bingham 6 16x18x25 CD 1 1 Unit 1 & 2 Company B-14 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Table B-4 (continued) Plant Listing of M ain Feedw ater Pum p Installations

Site Utility Mfg. Quantity Size Type Pump Number Category of Stages

North A nna Virgin ia Pow er Bingham 6 14xl6x19 CP 2 2 Unit 3 & 4 Company 4-50% 16 Oconee Unit 1 Duke Power Co. Delaval 1B SX 1 1 &2 2-50% Oconee Unit 3 Duke Power Co. Delaval 16 1BSX 1 1

Oyster Creek GPU Nuclear Byron 3-33% 8x8x14 DVMX 3 4 U nit 1 Corporation Jackson

Palisades Consumers Bingham 2 14x14x16 HSB 1 1 Nuclear Plant Power Company Palo Verde Arizona Public Byron 2-50% 24x24x19C HDR 2 1 Unit 1 Service Co. Jackson Palo Verde Arizona Public Byron 2-50% 24x24x19C HDR 2 1 Unit 2 Service Co. Jackson

Palo Verde Arizona Public Byron 2-50% 24x24x19C HDR 2 1 U nit 3 Service Co. Jackson

Peach Bottom PECO Pacific 2 4X BFI 2 11

Peach Bottom PECO Energy Byron 3-33% 14x14x17A DVSR 1 1 Unit 2 Jackson Peach Bottom PECO Energy Byron 3-33% 14x14x17A DVSR 1 1 Unit 3 Jackson

Perry Unit Cleveland Delaval 4-60% 16 1BSX 1 1 1 & 2 Electric Illum inating Perry Unit Cleveland Delaval 2-20% 14 1B SX 1 1 1 &2 Electric Illum inating Pilgrim U nit 1 Boston Edison Byron 3-33% 12x12x18 DVMX 3 3 Co. Jackson Pilgrim U nit 2 Boston Edison Byro n 2-50% 24x24x19C HDR 2 1 Co. Jackson Point Beach W isconsin Byron 2-50% 12x12x16 DVS 3 1 U nit 1 Electric Power Jackson Co. Byron Point Beach W isconsin 2-50% 12x12x16 DVS 3 1 Unit 2 Electric Power Jackson Co.

Prairie Island N orthern States Worthington 2 16 WGID 1 1 U nit 1 Power

Quad Cities Commonwealth Pacific 6 16X RHCNDS 3 2 Unit 1 & 2 Edison

B-15 EPRI Licensed Material Nuclear Maintenance Applications Center Table B-4 (continued) Plant Listing of M ain Feedwater Pump Installations

Site Utility Mfg. Quantity Size Type Pump Number Category of Stages River Bend Gulf States Bingham 3 14x14x17 CPD 2 2 Unit 2 Utilities Company Company

Salem Unit 1 Public Service Worthington 2 24 WGID 1 1 Electric & Gas

Salem Unit 2 Public Service W orthington 2 24 WGID 1 1 Electric & Gas 2-50% Seabrook New Hampshire Byron 20x20x18B HDR 2 1 U nit 1 Yankee Jackson 2-50% Seabrook New Hampshire Byron 20x20x18B HDR 2 1 Unit 2 Yankee Jackson 2-50% Sequoyah Tennessee Byron 14x14x17C DVSR 1 1 U nit 1 Valley A uthority Jackson Byron 2-50% Sequoyah Tennessee 14x14x17C DVSR 1 1 Unit 2 Valley Autho rity Jackson St Lucie Florida Power & Byron 2-50% 12x14x18 DVMX 3 2 Light Company Jackson St Lucie 2 Florida Power & Byron 2-50% 12x14x18 DVMX 3 2 Light Company Jackson

Surry Power Virgin ia P ow er Bingham 4 16x18x24 CD 1 1 Station 1 & 2 Company

Susquehanna Pennsylvania Ingerso ll- 6 16x17 CN 1 1 Un it 1 & 2 Power & Light Rand

Three Mile GPU Nuclear Byron 2-50% 14x14x17B DVSR 1 1 Island Unit 1 Corporation Jackson

Trojan N uclear Portland General Bingham 2 16x18x17 CD 1 1 Plant Electric Turkey Point 3 Florida Power & Byron 2-50% 12x14x18 DVM X 3 2 Light Company Jackson Turkey Point 4 Florida Power & B yron 2-50% 12x14x18 DVMX 3 2 Light Company Jackson V.C. Summer South Carolina Ingersoll- 3 14x17 CN 1 1 E lectric & Gas Rand Vermont Vermont Yankee Bingham 3 12x16x20 MSD 3 2 Yankee Nuclear Power Company

Vo gtle Un it 1 Georgia Power Pacific 2 20x17 HVF 1 1 Co.

V ogtle Unit 2 Georgia Power Pacific 2 20x17 HVF 1 1 Co.

B-16 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Table B-4 (continued) Plant Listing of M ain Feedwater Pump Installations

Site Utility Mfg. Q uantity Size Type Pump Number Category of Stages

Waterford Entergy Pacific 1 20x17 HVF 1 1 Unit 3 O perations, Inc. W atts Bar Tennessee Byron 2-50% 20x20x18B HDR 2 1 U nit 1 Valley Au thority Jackson Watts Bar Tennessee Byron 1 -20% 12x12x16 DVS 3 1 U nit 1 V alley Authority Jackson (standby) W atts Bar Tennessee Byron 2-50% 20x20x18B HDR 2 1 Unit 2 V alley Authority Jackson W atts Bar Tennessee Byron 1-20% 12x12x16 DVS 3 1 Unit 2 V alley Authority Jackson (standby) Byron 2-50% W olf Creek Wolf Creek 20x20x18B HDR 2 1 U nit 1 N uclear Jackson Operating Co. 4-50% WPPS Project Washingto n Delaval 16 1BSX 1 1 1 &4 Public Power Supply System 4-50% WPPSS Washington Delaval 20 1BSX 1 1 Project 3 & 5 Public Power Supply System WPPSS Washington Delaval 2-33% 14 2BSX 2 2 Project 3 & 5 Public Power Supply System 4-50% Zion Unit Commonwealth Delaval 16 1BSX 1 1 1 &2 Edison Zion Unit Commonwealth Delaval 2-50% 16 1BSX 1 1 1 &2 Edison

B-17 Table B-5 Bingham Supplied M ain Feedwater Pump Installation List

Number Gallons Horse- Site & Utility NSSS Pump of per Head Temperature Quantity Size TypeCategory RPM Stages M inute Power °F

Beaver Valley Duquesne Light Company WE 2 16x18x24 CD 1 1 15,200 1 900 7150 3580 384

River Bend Unit 2 Gu lf S tates U tilities C om pan yGE 3 14x14x17 CPD 2 2 11,058 2582 7380 4000 388

Catawba Unit 1 & 2 19,850/- Duke Power Company WE 4 16x18x17 CD 1 1 18,040 2250/2040 11500 5510 243/375

Comnanche Peak WE 16x18x17D CD 19,800 2322 11,584 5200 362 Texas Utilities Electric Company 4 1 1

Haddem Neck Con necticu t Yankee Atom ic WE 12x12x17 MSB 9,600 2000 4,900 3560 372 Power Company 2 3 2

Palisades Nuclear Plant C onsum ers Pow er Com pany CE 2 14x14x16 HSB 1 1 13,500 1920 6,600 5000 376

Surry Unit 1 & 2 1700 Virginia Power Company WE 4 16x18x24 CD 1 1 13,800 6,060 3560 377

Vermont Yankee Vermont Yankee Nuclear Power GE 12x16x20 MSD 7,630 2570 5,400 3560 Company 3 3 2 320

Arkansas Nuclear One Entergy Operations, Inc. BW/CE 2 16x16x16 HSB 1 1 14,750 1902 7,100 5700 374

Beaver Valley D uquesne Light Company WE 2 16x18x24B CD 1 1 15,200 1900 7,400 3560 384

Trojan Nuclear Plant Po rtland G en eral E lectric WE 2 16x18x17 CD 1 1 19,800 2020 9,964 5200 373

North Anna Unit 1 & 2 Virginia Power Company WE 6 16x18x25 CD 1 1 16,250 1980 9000 3560 390 Table B-5 (continued) Bingham Supplied M ain Feedwater Pump Installation List

Number Gallons Pump Horse- Temperature Site & Utility NSSS Quantity Size Type of per Head RPM Category Stages Power °F Minute

North Anna Unit 3 & 4 Virginia Power Company WE 6 14x16x19 CP 2 2 14,600 2734 11,000 3560 320

Byron WE 20x20x17 CD 19,450 5200 Commonwealth Edison 6 1 1 2110 9,945 32/410

Braidwood 32/410 Com monwealth Edison WE 6 20x20x17 CD 1 1 19,450 2110 9,945 5200 Table B-6 Byron Jackson Supplied M ain Feedwater Pum p Installation List

Quantity Number Gallons and % Pump Head Horse- Temperature NSSS Type per RPM Site & Utility Total Size Category of Power Stages Minute (ft.) °F Flow

P ilgrim U nit 1 GE 3-33% 12x12x18 DVMX 6650 5000-m 3560 Boston Edison Co. 3 3 2620 290

Dresden Unit 2 3-50% Commonwealth Edison GE 10x12x16 DVMX 3 2 10750 2700 9000-mg 4500 298

Dresden Unit 3 3-50% Commonwealth Edison GE 10x12x16 DVMX 3 2 10750 2700 9000-mg 4500 298

St Lucie Florida Power & Light CE 2-50% 12x14x18 DVMX 14100 1780 7000-m 3570 380 Company 3 2

St Lucie 2 Florida Pow er & Light CE 2-50% 12x14x18 DVMX 14100 1780 7000-m 3570 380 Company 3 2

Tu rk ey Po in t 3 Florida Power & Light WE 2-50% 12x14x18 DVMX 13000 1880 7000-m 3570 362 Company 3 2

Turkey Point 4 Florida Pow er & Light WE 2-50% 12x14x18 DVMX 13000 1880 7000-m 3570 362 Company 3 2

Duane A rnold U nit 1 GE 2-57% 12x12x18 DVMX 8900 2500 6000-m 3570 365 IE S U tilities 3 3

Oyster Creek Unit 1 3-33% 5000 2800 4000-m GPU Nuclear Corporation GE 8x8x14 DVMX 3 4 3570 214

Fort Calhoun Unit 1 3-50% 3500-m 396 Om aha Public Power DistrictCE 10x10x17 DVMX 3 2 8000 1750 3570

Calvert Cliffs Unit 1 Baltimore Gas & Electric CE 2-50% 14x14x17B DVSR 1 1 15000 2392 8000-T 5200 375 -

Table B-6 (continued) Byron Jackson Supplied M ain Feedwater Pump Installation List

Quantity Number Gallons and % Pump Head Horse- Temperature Site &Utility NSSS Total Size Type of per RPM Category Stages (ft.) Power °F Minute Flow

Calvert Cliffs Unit 2 2-50% Baltimore Gas & Electric CE 14x14x17BDVSR 1 1 15000 2316 8000-T 5130 344

Cry stal Riv er Unit 3 2-50% 8000-T 390 Baltimore Gas & Electric BW 14x14x17B DVSR 1 1 13320 2280 5120

Hatch Unit 1 GE 2-50% 14x14x17ADVSR 12500 2540 8500-T 5550 321 Georgia Power Company 1 1

H atch Un it 2 2-50% 12500 8500-T Georgia Power Company GE 14x14x17ADVSR 1 1 2540 5550 321

Maine Yankee 2-50% 14000 2000 8000-MG 5000 Maine Yankee Atomic PowerCE 14x14x17B DVSR 1 1 378

Three M ile Island Unit 1 2-50% 10200-T 5500 GPU N uclear C orporation B& W 14x14x17B DVSR 1 1 16100 2480 375

Cooper 2-50% 12800 2500 9500-T Nebraska Public Power DistrictGE 14x14x17A DVSR 1 1 5500 370

Diablo Canyon Unit 1 2-50% 2000 9600-T 5800 365 Pacific Gas & Electric WE 14x14x17CDVSR 1 1 17350

D iablo Canyon U nit 2 2-50% 2000 9600-T Pacific Gas & Electric WE 14x14x17CDVSR 1 1 17350 5800 365

Peach Bottom Unit 2 3-33% 8000-T 5300 PECO Energy GE 14x14x17ADVSR 1 1 12270 2310 378

Peach Bottom Unit 3 3-33% 8000-T 5300 PECO Energy GE 14x14x17A DVSR 1 1 12270 2310 378

Sequoyah Unit 1 2-50% 20000 1680 8500-T 5400 T en nessee Valley Au th orityWE 14x14x17CDVSR 1 1 399

Sequoyah U nit 2 2-50% 20000 1680 8500-T 5400 Tenn essee Valley Au th orityWE 14x14x17CDVSR 1 1 399 Table B-6 (continued) Byron Jackson Supplied M ain Feedwater Pum p Installation List

Quantity Number Gallons and % Head Pump of per Horse- Temperature Site & Utility NSSS Total Size Type Power RPM Category Stages Minute (ft.) °F Flow

1-20% W atts Bar Un it 1 1890 Tennessee V alley AuthorityWE (standby) 12x12x16 DVS 3 1 6100 3000-MG 5350 388

1-20% Watts Bar Unit 2 6100 Tennessee V alley A uthorityWE (standby) 12x12x16 DVS 3 1 1890 3000-M G 5350 388

P oin t Beach Un it 1 2-50% W isconsin Electric Pow er Co.WE 12x12x16 DVS 3 1 7810 2180 5000-MG 6000 348

Po in t B each Un it 2 2-50% W isconsin E lectric Power Co.WE 12x12x16 DVS 3 1 7810 2180 5000-M G 6000 348

Palo Verde Unit 1 2-50% 14000-T Arizona Public Service Co. CE 24x24x19C HDR 2 1 22314 2378 5315 347

Palo Verde Unit 2 2-50% 14000-T Arizona Public Service Co. CE 24x24x19C HDR 2 1 22314 2378 5315 347

Palo Verde Unit 3 2-50% 14000-T Arizona Public Service Co. CE 24x24x19C HDR 2 1 22314 2378 5315 347

Pilg rim Unit 2 2-50% Boston Edison Co. GE 24x24x19C HDR 2 1 19340 2215 12000-T 5135 367

C linton U nit 1 GE 2-50% 20x20x18B HDR 16545 2483 11000-T 5350 390 Illinois Power & Light 2 1

Clinton Unit 1 1-30% GE 18x18x18 HDR 1000 2483 6000-M G 5050 390 Illinois Power & Light (standby) 2 1

Clin ton Unit 2 GE 2-50% 20x20x18B HDR 16545 2483 11000-T 5350 390 Illinois Power & Light 2 1

Clin ton Unit 2 1-30% 1000 2483 6000-M G 5050 390 Illinois Power & Light GE (standby) 18x18x18 HDR 2 1 Table B-6 (continued) Byron Jackson Supplied M ain Feedwater Pump Installation List

Quantity Number Gallons and % Pump Head Horse- Temperature Site & Utility NSSS Size Type of per RPM Total Category Stages (ft.) Power °F Flow M inute

W olf Creek Unit 1 W olf C reek N uclear WE 2-50% 20x20x18B HDR 17260 2387 12000-T 5300 336 Operating Co. 2 1

Grand Gulf Unit 1 2-50% 24x24x19C 20600 15000-T 5400 290 Entergy Operations, Inc. GE HDR 2 1 2430

Grand Gulf Unit 2 2-50% 24x24x19C 2430 15000-T 5400 290 Entergy Operations, Inc. GE HDR 2 1 20600

Nine M ile P oint Unit 2 Niagara Mohawk Power GE 3-50% 20x20x22 HSB 18700 2444 12000-MG 3339 370 Corporation 2 1

M illstone Unit 3 1 -50% WE 20x20x18B HDR 19820 2050 1100-T 5250 364 N orth east Utilities (standby) 2 1

Seabrook Unit 1 2-50% 5300 New Hampshire Yankee WE 20x20x18B HDR 2 1 17150 2350 11000-T 364

Seabrook Unit 2 2-50% 5300 New Hampshire Yankee WE 20x20x18B HDR 2 1 17150 2350 11000-T 364

W atts Bar Un it 1 Tennessee Valley AuthorityWE 2-50% 20x20x18B HDR 2 1 17500 2010 10000-T 5020 402

Watts Bar Unit 2 2-50% 17500 1000-T Tennessee Valley AuthorityWE 20x20x18B HDR 2 1 2010 5020 402

Caraway Unit 1 2-50% U nion E lectric Com pany WE 20x20x18B HDR 2 1 17260 2387 12000-T 5300 336

336 Caraway Unit 2 2-50% U nion E lectric Com pany WE 20x20x18B HDR 2 1 17260 2387 12000-T 5300 Table B-7 Delaval Supplied M ain Feedwater Pump Installation List Table B-7 (continued) Delaval Supplied M ain Feedwater Pum p Installation List

Number Gallons Pump Horse- Temperature NSSS of per Head Site & Utility Quantity Size Type Category Power RPM Stages Minute °F

Perry Unit 1 & 2 2-20% 7700 2226 5000-M Cleveland Electric Illum inatingGE 14 1BSX 1 1 3570 345

WPPS Project 1 & 4 Washington Public Power BW 4-50% 16 1BSX 19323 2500 12062-T 5400 410 Supply System 1 1

WPPSS Project 3 & 5 Washington Public Power CE 4-50% 20 1BSX 21575 2369 12563-T 5300 150 Supply System 1 1

WPPSS Project 3 & 5 Washington Public Power 150 CE 2-33% 14 2BSX 2 2 9568 2174 5375-M 3570 Supply System Table B-8 Ingersoll-Rand Supplied M ain Feedwater Pum p Installation List

Number Gallons Pump Horse- Temperature Site & Utility NSSS Q uantity Size Type of per Head Category Power RPM °F Stages Minute

M illstone 1 GE 12 8670 2400 6810 3580 291 Northeast Utilities 3 C 2 3

Indian P oint Unit 2 Consolidated Edison WE 2 16 JT 1 1 15300 1830 7210 4865 390

Indian Poin t 3 Consolidated Edison WE 2 16 JT 1 1 15300 1830 7210 4875 390

Millstone Unit 2 CE JT 15000 2100 7840 5150 N ortheast Utilities 2 16 1 1 380

L im erick Unit 1 & 2 GE JT 12900 2314 7750 5160 367 PECO 6 16 1 1

Susquehanna Unit 1 & 2 Pennsylvania Power & LightGE 6 16x17 CN 1 1 12500 2200 6936 5200 384

V.C. Summer South Carolina Electric & GasWE 3 14x17 CN 1 1 9970 2539 6785 5100 342 Table B-9 Pacific Supplied M ain Feedwater Pump Installation List

Number Gallons Pump Horse- Temperature Site & Utility NSSS Quantity Size Type of per Head Power RPM °F Category Stages Minute

Quad Cities Unit 1 & 2 Co m m on wealth Ed ison GE 6 16X RHCNDS 3 2 11050 2573 8180 4500 298

LaSalle Unit 1 12000-T 5500 Co m m on wealth Ed ison GE 2 20x17 HVF 1 1 19400 2450 365

LaSalle Unit 2 1 2000-T5500 Commonwealth Edison GE 2 20x17 HVF 1 1 19400 2450 365

LaSalle Unit 1 12000-T Commonwealth Edison GE 1 16RX BFI 2 1 10300 2110 5500 360

LaS alle U nit 2 GE 16RX BFI 10300 2110 12000-T 5500 360 Commonwealth Edison 1 2 1

V ogtle U nit 1 WE 20x17 HVF 20400 3000 14980-T 5950 403 Georgia Power Co. 2 1 1

Vogtle Unit 2 14980-T Georgia Power Co. WE 2 20x17 HVF 1 1 20400 3000 5950 403

W aterford Unit 3 9780-T Entergy Operations, Inc. CE 1 20x17 HVF 1 1 17940 2150 5200 366

Peach Bottom PECO GE 2 4X BFI 2 11 515 4000 657-M 3550 345

Browns Ferry 1 2800 8450-T 5500 Tenn essee Valley Au th orityGE 3 16X HVCN 1 1 11200 300

Browns Ferry 2 2800 8450-T 5500 Tennessee Valley A uthorityGE 3 16X HVCN 1 1 11200 300

Browns Ferry 3 8450-T 5500 Tennessee V alley AuthorityGE 3 16X HVCN 1 1 11200 2800 300 Table B-10 W orthington Supplied M ain Feedwater Pump Installation List

Number Gallons Pump Horse- Temperature Site & Utility NSSS Quanity Size Type of Head Category per Power RPM °F Stages Minute

C rystal R iver Unit 1 Florida Power & Light BW 1 10 WC 1 3932 7600 6500 3500 347 Company

N ine M ile P oint U nit 1 Niagara Mohawk Power GE 18 WNC 13000 2650 8500 5000 311 Corporation 1 2 1

N ine M ile Point U nit 1 2700 2600 Niagara Mohawk Pow er GE 2 8 WNC 2150 6772 310 Corporation 2 1

Prairie Island Unit 1 WE 16 WGID 8600 2150 5000-MG5130 360 Northern States Power 2 1 1

Kew aunee W isconsin Public Service Corporation WE 2 16 WGID 1 1 8600 2150 5000-MG5130 360

DC Cook Unit 1 WE 24 WGID 16700 2150 9600-T 5220 346 Indiana/Michigan Power 2 1 1

DC Cook Unit 2 WE 16700 9600 Indiana/Michigan Power 2 24 WGID 1 1 2150 5220 546

Salem Unit 1 WE 24 WGID 18600 2530 11500-T 5530 368 Public Service Electric & Gas 2 1 1

Salem Unit 2 18600 Public Service Electric & GasWE 2 24 WGID 1 1 2530 11500-T 5530 368

F itzpatrick New York Power AuthorityGE 2 18 WNC 2 1 13800 1980 7300-T 4500 371

Fitzpatrick New York Power AuthorityGE 3 20 WGID 1 1 11600 2300 8900-T 5150 367

M cG uire U nit 1 & 2 WE 24 WGID 17965 2045 365 Duke Power Company 4 1 1 9142 5210

Hope Creek Public Service Electric & GGE as 6 18 WGID 1 1 13000 2550 8300 5745 369 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX C TYPICAL STUFFING BOX DRAWINGS

Typical stuffing box drawings for each of the three alternative constructions (floating ring, injection throttle bushing, and mechanical seal) follow:

Figure Description Page Stuffing Box, Floating Ring Construction...... C-2 Figure C-1 Figure C-2 Stuffing Box, Injection Throttle Bushing Construction, Roating Member Serrations ...... C-3

Figure C-3 Stuffing Box, Injection Throttle Bushing Construction, Rotating and Stationary Member Sarrations ...... C-4

Figure C-4 Typical Mechanical Seal ...... C-5

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Figure C-1 Stuffing Box, Floating Ring Construction

C-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Figure C-2 Stuffing B ox, Injection Throttle Bushing C onstruction, Rotating M ember Serrations

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Figure C-3 Stuffing Box, Injection Throttle Bushing Construction, Rotating and Stationary M ember Serrations

C-4 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

C-5 EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX D TYPICAL BEARING CONSTRUCTION DRAWINGS

Typical bearing construction drawings for Kingsbury type thrust bearings and for fixed and self-aligning radial bearings follow:

Figure Description Page Figure D-1 Typical Bearing Construction, Thrust Bearing ...... D-2

Figure D-2 Typical Bearing Construction, Fixed Radial Bearing...... D-3

Figure D-3 Typical Bearing Construction, Self-Aligning Bearing ...... D-4

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Figure D-1 Typical B earing Construction Thrust Bearing

D-2 EPRI Licensed Material Main Feedwafer Pump Maintenance Guide

Figure D-2 Typical Bearing Construction Fixed Radial Bearing

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Figure D-3 Typical Bearing Construction Self-aligning Radial Bearing

D-4 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX E FEEDWATER SYSTEM FLOW DIAGRAMS

Figure Description Page Figure E-1 Babcook & Wilcox Pressurized Water Reactor...... E-2

Figure E-2 Combustion Engineering Pressurized Water Reactor ...... E-3

Figure E-3 General Electric Boiling Water Reactor ...... E-4

Figure E-4 Westinghouse Pressurized Water Reactor...... E-5

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E-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

E-3 EPRI Licensed Material Nuclear Maintenance Applications ions Center

E-4 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

E-5 EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX F TYPICAL OIL LUBRICATION PIPING SCHEMATICS

Figure Description Page Figure F-1 Motor Gear Pump Lubrication Schematic ...... F-2

Figure F-2 Turbine Pump Lubrication Schematic...... F-3

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F-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Figure F-2 Turbine Pump Lubrication Schem atic

F-3 EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX G SYMPTOMS OF FEEDWATER PUMP PROBLEMS AND RECOMMENDATIONS FOR SOLUTIONS

The symptoms and recommendations described in this appendix were extracted from the report RP 188s4.28 Draft Manual for Investigation and Correction of Boiler Feedpump Problems (1990) (Reference 13).

The symptoms, listed below in categories, identify most of the typical symptoms found during prior EPRI investigations of boiler feed pump problems. Symptoms discovered during EPRI visits to the representative power plants in the EPRI Boiler Feedpump program are listed in the reference document (Reference 13).

Each symptom is numbered and described in detail. For each symptom, one or more specific recommendations are listed. Most of the recommendations are directed toward a solution to the problem. However, some recommendations call for additional information or for an analysis to determine the specific cause of the problem.

Symptoms are grouped into three categories: • Operations Symptoms-Symptoms that would typically be seen by the plant operator in the control room or while walking the power plant. • Maintenance Sym ptoms-Symptoms that would typically be noticed by the plant maintenance personnel m the power plant or while performing maintenance on the boiler feed pumps. • Engineering Symptoms-Symptoms that would typically come to the attention of the plant engineer during a review of the boiler feed pumps as part of a problem analysis.

Review of Recomm endations

Typical recommendations for problem solutions are tabulated m this appendix. For extensive detail on any recommendation, please refer to Reference 13.

Symptoms of Feed Pump Problems 1. Increased seal injection flow 2. Increased or high vibration 3. Excessive leakage and steaming at the seals 4. Reduced pump performance

G-1 EPRI Licensed Matenal Nuclear Maintenance Applications Center

5. Pump seizure 6. High radial bearing temperature 7. Shift in axial rotor position 8. Excessive oil leaks 9. High thrust bearing temperature 10. Damaged or broken suction line hangers 11. High suction line vibration, usually at startup or shutdown 12. Loud noise in suction line, usually at start up, low plant load, or during transients 13. Crackling noise in pump, usually at startup, low plant load, or during transients 14. Increased balance drum/disk leakoff flow 15. Leaking/steaming at discharge head and barrel joint

Symptoms of Feed Pump Problems 1. Excessive floating ring seal replacement 2. Lube oil contaminated with water and/or bearing metal 3. Excessive bearing wear or replacement 4. Cracked or broken shaft 5. Pump seizure 6. Excessive internal wear 7. Lack of coupling lubrication 8 Excessive gear coupling wear 9. Cracked or broken impeller vanes or shrouds 10. Cracked oil supply or drain lines. 11. Cracked balance leakoff lines 12. Cracked or broken diffuser or volute vanes 13. Leaking bearing housing at joint 14. Damaged or broken suction line hangers 15. Excessive pump-driver misalignment 16. Deformed dowel pins at feed pump pedestal 17. Inlet vane damage on first stage impeller 18. Increased or high vibration 19. Leaking/steaming at discharge head/barrel joint; gap between barrel and head; and cutting at discharge head/barrel gasket joint 20. C utting washing at internal barrel/elem ent gasket joint.

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Recomm endations for Solving F eed Pump Problems

These first ten recommendations apply to all feed pumps regardless of type and size in all plants regardless of type and size. 1. Proper monitoring of feed pump during operation 2. Prepare and use proper operating procedures 3. Correct element overhaul procedures and policy 4. Install shaft proximity probe vibration monitoring system 5. Institute vibration monitoring program 6. Installation of thermocouples on feed pump barrel 7. Proper rotor balancing and balancing criteria 8. Add flow measurement to balance leakoff line 9. Install or convert to drain temperature control on seal injection 10. Proper alignment and doweling of feed pump and driver 11. Restore or modify barrel insulation 12. Gap A and B modifications to impellers and diffusers or volutes 13. Field vibration and modal testing, and rotor analysis 14. Install swirl break on balance disk bushing 15. Redesign and replace radial bearings 16. Redesign and replace first stage impeller 17. Redesign and modify recirculation system 18. Replace floating ring type seals with fixed-throttle bushing type seals 19. Replace gear coupling with flexible disk type 20. Modify shaft design to eliminate stress concentration areas 21. Change wear ring material to reduce galling and decrease wear 22. Increase or decrease pedestal stiffness 23. Remachine discharge head 24. Suction line transient analysis 25. Inspect and measure barrel and inner element 26. Replace inner element with new design

G-3 EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX H RECONCILIATION OF NPRDS/EPRI MAIN FEEDWATER PUMP FAILURE DATA

T he following inform ation has been obtained from the Nuclear Plant R eliability Data System and is the basis for much of the content and direction to this guide.

NPRDS/EPRI Main Feedwater Pump Failure Data

Journal Bearings

Twenty nine failures of journal bearings (also called radial bearings) were reported. Three of these also involved thrust bearing failures. Two units reported three failures, two units reported two failures, and the remaining 23 units only one failure each. The most recent failure m any of the four units reporting more than one failure occurred m May 1990.

Failure causes were believed to be:

Causes Failures

Unknown 9

M aintenance 9

Normal life 6

Other 5

Symptoms were:

Symptoms Failures

Hot bearing or lube oil 17

Vibration 6

Inspection 3

Other 3

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Journal bearing failures are always serious, often causing outages and/or damage to other pump components. They seem to occur randomly, without any obvious pattern. Detection of incipient failures is discussed in Section 4 of this guide.

Thrust Bearings

Fourteen thrust bearing failures were reported. Three of these also involved journal bearing failures. Two units reported two failures, and the remaining 12 units only one failure each. The most recent failure in one of the units reporting two failures occurred in August 1989.

Failure causes were believed to be:

Causes Failures

Unknown 5

M aintenance 5

W ater in lube oil 2

Other 2

Symptoms were:

Symptoms Failures

Hot bearing or lube oil 5

Vibration 3 Lube oil low pressure or leak 3

Other 2

Inspection 1 Thrust bearing failures are always serious, often causing outages, and damage to other pump components. They seem to occur randomly. None of the pumps involved had balance disks, so none of the reported failures was caused by a balance disk failure. Detection of incipient failures is discussed in Section 4 of the guide.

Controlled Leakage Shaft Seals

Thirty failures of controlled leakage seals were reported by 14 different units, ten failures at one unit, three at one unit, and two each at five other units. The unit reporting ten failures is the subject of a case history in Section 3.2.3 of this guide.

H-2 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

Failure causes were believed to be:

Causes Failures

Unknown 11

Norm al life 10

Operation or maintenance 7

Other 2

Symptoms were:

Symptoms Failures

Detection by plant personnel 27

Seizure 2

Vibration 1

Some feedwater pumps were furnished with floating ring type controlled leakage seals. W e believe that these have been replaced and that all controlled leakage seals are now the fixed breakdown type. Failures of fixed breakdown type shaft seals are typically caused by gradual wear of Me seal parts, causing gradually increased leakage. Such leakage can be monitored by plant personnel, and seal parts can be replaced during scheduled outages.

Seizures of controlled leakage seals almost never are the root cause of a failure. They are the result of excessive debris in the pumped fluid, or of the failure of a journal bearing or an impeller.

C ouplings

Fourteen coupling failures were reported. Four Units reported two failures and the remaining six units reported one each. In many reported failures, the coupling was not damaged but its misalignment caused excessive vibration, a hot bearing, or other cause for concern.

H-3 EPRI Licensed Material Nuclear Maintenance Applications Center

Failure causes were believed to be:

Causes Failures

Maintenance 6

Gear type coupling 3

Unknown 3

Misalignment 2

Symptoms were:

Symptoms Failures

Vibration 9

Other 3

Observed failure 2

Only eight of the reported coupling failures caused the pump to be unavailable. Pump downtime was generally nominal. Only two caused damage to another pump component. One required replacement of a journal bearing, and the other caused major damage to the pump shaft.

Almost all of the reported coupling failures could have been avoided by frequent inspection and lubrication of gear-type couplings, or replacement of gear-type couplings with disk-type couplings, or maintaining pump-to-driver alignment.

Impellers

Thirteen impeller failures were reported. Four units reported two failures and the remaining five units reported one each (although two of these five were identical pumps in different units of the same plant).

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Failure causes were believed to be:

Causes Failures

Unknown 3

Manufacturer 3

M aintenance 3

Normal Life 2

Other 2

Symptoms were:

Symptoms Failures

Vibration 10

Other 3

Six failures were caused by pieces of the impeller breaking off while the pump was running, and involved major damage. In one case inspection of the impeller revealed a crack which soon would have caused major damage if the pump had been operated. Three impeller failures (two in the sam e unit) were caused by foreign objects in the pumped fluid. One involved the impeller locking ring, and one the axial positioning of the im peller on the shaft.

R otating Elements

Ten rotating element failures were reported. One unit had two failures and the other nine units had one each.

Failure causes were believed to be:

Causes Failures

Unknown 3

Normal Life 3

Operation and/or Maintenance 3

Therm al stresses 1

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Symptoms were:

Symptoms Failures

Vibration 5

Seizure 2

Hot bearing 1

Shaft seal leak 1

Turbine overspeed 1

Rotating element failures are always serious. They have many causes and seem to occur randomly, without any obvious pattern. Nine of the ten reported failures caused unscheduled feed pump outages.

Shafts

Eight shaft failures were reported. Three failures were in the same pump (in 1985 and 1986). The other five units had one failure each.

Shaft failures are serious, but random and rare. Detection of incipient failures is discussed in Section 4 of this guide.

H-6 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

APPENDIX I NMAC BOLTED JOINT MAINTENANCE AND APPLICATION GUIDE (SECTION 3)

The following information has been duplicated intact from EPRI document Bolted Joint Maintenance S Applications Guide (Reference 16).

I-1 EPRI Licensed Material EPRI Licensed Material Main Feed water Pump Maintenance Guide

APPENDIX J METHODS FOR OBTAINING THE CORRECT RUNNING POSITION OF MAIN FEEDWATER PUMP ROTORS

The running position of the rotor is accomplished using two different methods on the majority of feedwater pumps in service. One is by the use of a thrust collar spacer. This spacer is located behind the thrust collar next to the step on the shaft (see Figure J-1). The other method is accomplished by two separate shims, one each behind each thrust bearing cage (see Figure J-4). The thickness of these shims must be adjusted as a set. Adding to one requires a reduction to the other, or vice versa, to obtain the correct running position.

Figure J-1 Thrust C ollar U tilizing a Spacer

The thickness of the thrust collar spacer determines the position of the shaft and impeller in relation to the casing. The thrust bearing assembly closest to the pum p casing does not change position relative to the bearing housing . The bearing housing is constructed such that a partition is cast into the bearing housing.The inner thrust bearing is held in position axially by this partition, which also acts as a partition to separate the oil

J-1 EPRI Licensed Material Nuclear Maintenance Applications Center supplies between the journal bearing and thrust bearings. With this type of configuration, the inner thrust bearing position should never change unless machining has occurred to the bearing housing where the inner shoes or cage contacts the housing, or machining of the face of the bearing housings has occurred where it meets the pump casing. The thrust collar spacer thickness allows the shaft to be positioned fore or aft inside of the collar, such that the rotor can be positioned In the casing.

When impellers are changed or installed on a shaft, the position of the impeller may vary from the rotor that was removed from the casing. This changes the running position of the rotor.

The pump casing on a double-suction-type pump should be symmetrical between the casing ring locations. With the rotor moved from one extreme to the other along the axial plane, the impeller wear hubs will contact the casing rings In either direction. Total travel can be measured. W hen the rotor is set to the final running position, half of this travel will be on either side of the impeller, plus or minus the allowable tolerance from the center.

Several methods are used to determine the correct running position and the center offset of the rotor. These methods and allowable tolerances from the center vary among manufacturers and can vary between single and multistage pumps. Different reference points exist between manufacturers. The manufacturer's Operations & Maintenance Manual should be consulted when making changes to the running position.

A manufacturer's literature should contain the tolerances and dimensions from the center that are allowed. These dimensions may range from .015 inch to 1/16 inch. For example, one manufacturer, Worthington, on its model 24 WGID pump, does not supply a thrust collar spacer, but relies on a step on the shaft on the coupling or inboard side of the impeller, for the impeller position. The tolerance used on this model pump is 1/16 inch.

One pump manufacturer recommends that a surface plate or very level table be used to measure the height of the inner thrust bearing and then center the rotor in the casing. At this point, take the difference between the two and adjust the shaft accordingly.

A good method for setting the shaft to the center of the pump casing is described as follows.

The majority of pumps in service are Category 1 pumps, with double-suction single impeller construction and a radially split casing. The following explanation applies to this specific construction for axially centering the rotor or setting the running position. However, with minor deviations, the same principles apply for axially or horizontally split casing pumps.

With the rotor installed in the casing, and the end cover(s), bearing housings and journal bearings in place, the rotor may be positioned from one extreme to the other within the casing. The total axial travel of the rotor is from the impeller wear hub contacting the casing ring on the inboard side of the pump to contact the casing ring in the oppo-

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site direction on the outboard end of the pump. With the spacer type, the thrust collar position is a constant in relation to the housing and thrust bearings, with the exception of any wear to the thrust bearing

Figure J-2 is a plan view of the thrust bearing or outboard housing showing the shaft at the two extremes of the casing as depicted by the shaded and clear areas and the dimensions affixed.

NOTE: All dimensions in all of the following examples are typical and used for illustration only.

Figure J-2 Plan View of Thrust Housing

Figure J-2 reveals the total travel of the rotor to be .460 inch. The correct running position of the rotor from the outer-most measurement would be 3.135 inch.

.460 3.365 - 2.905 + = .230 + 2.905 3.135 2 = The inner thrust bearing assembly can now be installed and the rotor positioned against the thrust bearing. Measuring from the outer casing again yields a measurement of 3.176 as shown in Figure J-3.

3.176 minus 3.135 yields a difference of .041 inch.

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Figure J-3 Plan View with Inner Thrust Bearing Installed

Remembering that what is taken away from one side of the rotor is added to the other side, the .041 must again be divided by 2 to give a move of .020 inch.

In which direction is the rotor to be moved?

The rotor and the center line of the impeller traveled past the center line of the casing (volute or diffuser) toward the inboard end. The rotor must be positioned closer to the outboard end of the pump by .020 inch. The spacer must be thinner by this amount.

The formula below was devised for use to prevent errors in calculations or direction of the required move of the rotor. By using this formula and precision measuring instruments, the running position can be extremely close.

A = rotor to inboard position

D = existing rotor location

B = rotor to outboard position against inner thrust bearing

C = required rotor position

X = change in shim thickness

A -B C -D + B = C X = 2 2

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Using the dimensions found in Figures J-3 and J-4, the following determinations can be made in the example below: 3.365- 2.905 3.135 - 3.176 + 2.905 = 3.135 X = = -.0205 2 2 The thrust collar spacer thickness must be changed by .020 to .021 inch to place the rotor in the center of the diffuser or volute.

A minus (-) result indicates the spacer must be machined. A positive (+) indicates the spacer must be thicker. Manufacturers typically supply thrust collar spacers as spare parts in a thickness of .375 inch to allow for machining. If a thicker spacer is required based on the calculations, a new spacer must be machined to the required thickness. The thrust collar spacer must be finished to size by surface grinding to ensure excessive runout is not induced into the thrust collar.

Figure J-4 Shim Type Rotor Centering

The bearing housing shown m Figure J-4 with the shims behind each thrust bearing is similarly checked for the correct running position. The total travel of the rotor may be measured using either a depth micrometer or a long-range dial indicator to yield the total distance.

The following example will use a long-range indicator in place of a depth micrometer to provide the user with an alternative method of measuring the rotor movement and position.

The dial indicator in this case is placed on the coupling end of the shaft to read the total travel.

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For purposes of this example, the total travel read from the dial indicator is .437 inch. The total travel of the rotor is then divided by 2.

.437 = .2185 2 The inner thrust bearing and the original shim is placed in the housing and the rotor moved to the inboard end against the inner thrust bearing, and another measurement is taken.

The dial indicator travel from the rotor at the outboard position to rest against the inboard thrust bearing is = .197 inch

.2185 - .197 = .0215

The rotor lacks this amount returning to the center of travel and the inboard shim must have this amount removed, while the outboard shim must have this amount added. This amount does not account for the original shim thickness to allow for shaft end play. The total travel of the rotor equals the two shims thicknesses, plus the present amount of end play.

At this time, a prudent measure is to install the outer thrust bearing assembly and original shim along with the housing end cover and perform an end play check.

An example would be the end play, which was found to be .014 inch.

The combined thickness of the two shims is .423 inch.

.423 + .014 = .437

The inner shim has a thickness .232 inches.

The outer shim has a thickness .191.

To make the rotor position change and set the end play at .010 would require the addition of .002 per shim.

In the case of these measurements, the inner shim should have .0195 removed and the outer shim should be .0195 thicker.

The shims may come from the manufacturer packaged as two shims with each shim being one solid circular piece (doughnut) at a maximum thickness. Installing a one- piece shim requires removal of all interferences such as thrust collar, bearing plates, and end covers to install these shims. To reduce time, a proven method is to cut the shims in half or split them so they can be rotated around the shaft such as the two-piece pivot

J-6 EPRI Licensed Material Main Feedwater Pump Maintenance Guide shoe bearing cage. These shims can be place flat on a saw and cut in half, then finished to size and deburred.

Success has been achieved in splitting these shims in half to facilitate removal and installation without removal of the thrust collar. Both halves of the single shim should be surface ground together to achieve the proper thickness and ensure perpendicularity to the shaft. The end cover will probably not contain shims for adjusting end play on this type of configuration.

M ultistage pumps are nearly identical in the method of setting the running positon. Care must be taken and the manufacturer consulted if, after setting the running position, problems continue to persist. Indexing of the rotor or casing may be required if thrust or rotor shuttling during operation is a problem.

Indexing the rotor or casing is the method to determine if all impellers are aligned correctly to the diffusers or volutes. The machining tolerances and dimensions should be such that all im pellers fall within the specified range of tolerance to all volutes or diffusers. If problems are present from the initial startup, then this should be checked. However, if problems are not present but arise after a new rotor or shaft installation, this can also point to a problem with the impeller positions on the shaft.

M ultistage diffuser pumps may require an additional measurement of the last-stage impeller and the location of the diffuser cover to be factored into running position calculations. Your manufacturer's Operations & M aintenance manual should be consulted for details.

If problems occur with running position settings on multistage diffuser pumps after a rotor change out, and doubts arise about the true running position, the pump manufacturer's field service engineer should be consulted. The possibility exists that impeller spacer sleeves may not match the original rotor. This may also require that the diffusers and covers be stacked and measured to get accurate dimensions between the center line of each diffuser over the length of the assembly.

Manufacturers machine diffusers and covers to close tolerances and calculate and machine the impeller spacers or lock ring grooves accordingly to ensure that the impellers are falling within the specified tolerance of alignment to the diffusers.

M ultistage volute or single-stage double suction volute type pumps with horizontally split casings may be checked and verified for the running position while the top half of the casing is removed. This check provides an accurate visualization of the running position.

J-7 EPRI Licensed Material EPRI Licensed Material Main Feedwater Pump Maintenance Guide

REFERENCES/BIBLIOGRAPHY

EPRI Documentation (contract number and title) 1. Survey of Feedpump Outages (1978), FP 754, Energy Research and Consultants, Inc. 2. Recommended Design Guidelines for Feedwater Pumps in Large Power Generating Units (CS-1512), (1980), RP 1512, Energy Research and Consultants, Inc. 3. Evaluation of Basic Causes of Repetitive Failures of Nuclear and Fossil Feedwater Pumps- Project 1327, (1980), NP-1571, MPR Associates, Inc. 4. Feedpump Hydraulic Performance and Design Improvement (CS-2322), (1982), RP 1884- 5, Franklin Research Center Transamerica Delaval, Inc. 5. Feedpump Hydraulic Performance and Design Improvement, Phase 1 Research Program Design (CS-2323), (1982), RP 1884-6, M assachusetts Institute of Technology/Byron Jackson Pum p Division 6. Feedpump Hydraulic Performance and Design Improvements (1983), RP 1884-10, Sulzer Brothers, Limited 7. Feedpump Hydraulic Performance and Design Improvements (1989), RP 1884-18, Pradco 8. Rotor Dynamical Characteristics of Wear Ring Configurations for Feedwater Pumps (1983), RP 1884-22, Case W estern Reserve University, M.L. Adams 9. Operation and Maintenance Guideline for Feedwater Pumps (1986), RP 1884-23, General Physics Corporation 10. Draft Feedwater Pump Technical Specification Guidelines (1986), RP 1884-24, P.R. Stech Corporation 11. Symposium Proceedings: Power Plant Pumps EPRI CS-5857, June 1988, RP 1884-25 12. Wear Resistant Materials and Coatings for Boiler Feedpumps (1989), RP 1884-26, Lehigh University / Ingersoll-Rand 13. Draft Manual for Investigation and Correction of Boiler Feedpump Problems (1990), RP 1884-28, Stone and Webster Engineering Corporation 14. Intervention to Support Maintenance Personnel in Pump and Valve Overhauls (1991), RP 3111-5, Anacapa Sciences, Inc. 15. Feedpump Operation and Design Guidelines- Summary Report (1993), RP 1884-10, TR- 102102, Sulzer Brothers, Limited 16. Bolted Joint Maintenance and Applications Guide (1995), RP 3814-07, TR-104213, Aptech Engineering Services, Inc.

BIB-1 EPRI Licensed Material Nuclear Maintenance Applications Center

17. Boiler Feedpump Operation and Maintenance Guidelines-Volume 1: Troubleshooting (1994), TR-104292-V1, GPS Technologies, Inc. 18. Boiler Feedpump Operation and Maintenance Guidelines-Volume 2: Case Studies (1994), TR-104292-V2, Stone and Webster Engineering Corporation 19. Static Seals Maintenance Guide (1994), RP 3814-06, TR-104749, QES, Inc. 20. Use of Reliability-Centered Maintenance for the McGuire Nuclear Station Feedwater System Project 2508-2 (1986), NP-4795, Los Alamos Technical Associates/Saratoga Engineering Consultants, Inc. 21. Application of Reliability-Centered Maintenance to San Onofre Units 2 and 3 Auxiliary Feedwater Systems- Project 2508-8 (1987), NP-5430, ERIN Engineering and Research Corporation 22. Power Plant Pumps Short Course-1992, EPRI M & D Center 23. Manual of Bearing Failures and Repair in Power Plant Rotating Equipment- Project 1648- 10 (1991), GS-7352, Mechanical Technology, Inc. 24. Power Plant Pumps Short Course-1992, EPRI M & D Center 25. NMAC Targets Plant Leakage Reduction," NMAC MEMO, June 1995 26. Nuclear Unit Operating Experience: 1987-1988, EPRI NP-7191 Project 2940-3 Final Report February 1991, Stoller Power Division of RCG/Hagler, Bailly, Inc.

Books 27. Eugene A. Avallone and Theodore Baumeister III, Mark's Standard Handbook for Mechanical Engineers, McGraw-Hill Book Co, Ninth Edition 28. Val S. Lobanoff and Robert R. Ross, Centrifugal Pumps, Design and Application, Gulf Publishing Company, Second Edition 29. Igor J. Karassik, William C. Krutzsch, W arren H. Fraser, and Joseph P. Messina, Pump Handbook, McGraw-Hill Book Company, Second Edition 30. Michael M. Calistrat, Flexible Couplings, Their Design, Selection and Use, Caroline Publishing 31. Galen Evans and Pedro Casanova, The Optalign Training Book, Ludeca, Inc. 32. Heinz Bloch, Practical Machinery Management for Process Plants-Volume 1- 1992, Gulf Publishing Company

Codes and Standards 33. API 610, Seventh Edition, Centrifugal Pumps for General Refinery Service, American Petroleum Institute 34. API 614, Third Edition, Lubrication, Shaft-Sealing, and Control Oil Systems for Special Service, American Petroleum Institute

BIB-2 EPRI Licensed Material Main Feed water Pump Maintenance Guide

35. API 670, Vibration, Axial Position, and Bearing-Temperature Monitoring Systems, American Petroleum Institute 36. ANSI/HI 1.1-1.5- 1994, American National Standard for Centrifugal Pumps for Nomen- clature, Definitions, Application and Operation, Hydraulic Institute

Articles, Reports, and Publications 37. Aging Management for Commercial Nuclear Power Plants-Contractor Report SAND 93- 7045, 1993, Multiple Dynamics Corporation 38. The Pump Handbook Series, Volume I- Centrifugal Pumps, Pumps, and Systems 39. The Pump Handbook Series, Volume II- Mechanical Seals and Seal Support Systems, Pumps and Systems 40. The Pump Handbook Series, Volume IV- Pump Maintenance and Reliability, Pumps and System s 41. Elemer Makay, Corrective Measures for Utility Pump Low Flow Hydraulic Instability, ASM E/NRC Symposium, Inservice Testing of Pumps and Valves," 1989 42. Elemer Makay and James Barrett, Changes in Hydraulic Component Geometries Greatly Increased Power Plant Availability and Reduced Maintenance Cost Case Histories, 1st International Pump Symposium, May 1987, Texas A & M University 43. Elemer Makay and James Barrett, Field Experience Brings Help to Embattled Pump Users, Power, July 1987 44. Elemer Makay and Douglas Nash, Gap-Narrowing Rings Make Booster Pumps Quiet at Low Flow, Power, September 1982 45. Elemer Makay and James Barrett, Ten Ways to Improve High-Energy Pump Perfor- mance, Power, January 1988 46. Elemer Makay and Isidro Diaz-Tous, Feed Pump Reliability and Efficiency Improve- ments Resulting from Hardware Modifications, EPRI Conference, 1982 47. S. Gopalakrishnan and F.J. Costanzo, A Software-Based Diagnosis of Feed Pump Prob- lems, Third EPRI Conference on Incipient Failure Detection m Power Plants, March 1987 48. S. Gopalakrishnan and John Mieding, Some Causes and Cures for Boiler Feed Pump Failures, Rotating Machinery Repair User's Council Meeting, 1988 49. William O'Keefe, Can State-of-the-Art Research and New Experience Save Your Pumps?, Power, December 1988 50. Igor Karassik, Fact or Fiction? Some Misconceptions About Desirable Construction Features of Centrifugal Pumps, Chemical Processing, Mid-March 1989 51. Robert Ashton, Optimize B-F Pump Throttle Bushings by Close Match to Feedwater Circuit, Power, September 1980

BIB-3 EPRI Licensed Material Nuclear Maintenance Applications Center

52. S. Gopalakrishnan, A New Method for Computing Minimum Flow, Fifth Texas A & M Pump Symposium, June 1988 53. W illiam O'Keefe, Packing and Seals for Pumps and Valves, Power, August 1984 54. Pump Testing- Comparison of Factory vs. Field Test of Centrifugal Pumps, Proceedings of the Second NRC/ASME Symposium on Pump and Valve Testing, NUREG/CP- 01233, EGG-2676, Idaho National Engineering Laboratory, EG & G Idaho, Inc. 55. Richard Colsher, Robert Frank, Robert Matushisky, Benefits of an Integrated Predic- tive M aintenance Program in a Power Plant Environment, EPRI Maintenance and Diagnostic Center, PDM Conference, September 1992 56. W.E. Nelson and J.W. Dufour, Pump Vibrations, Proceedings of the Ninth Interna- tional Pump User's Symposium, 1992, Texas A & M University 57. John A. Gibbs, Predicting Rotating Machinery Failures, Instruments and Control Systems, May 1981 58. S. Goldman, Periodic Machinery Monitoring: Do It Right, Hydrocarbon Processing, August 1984 59. William V. Adams, Consider Mechanical Seals for Large Boiler-Feed Pumps, Power, October 1987 60. Alastair J. C ampbell, Static and Dynamic Alignment of Turbomachinery, Orbit, Bently Nevada Corporation, June 1993 61. Donald E. Bentley and Malcolm Werner, Extending Machinery Life, Orbit, Bently Nevada Corporation, September 1990 62. William Key, Critical Design Features for High Performance Seals, September 1995 63. William V. Adams and Peter Lytwyn, Retrofit of an Unspared Main Boiler Feed Pushup to End Face Mechanical Seals, ASME/IEEE Power Generation Conference, October 19- 23, 1986

Manufacturer's Instruction Manuals 64. The following manufacturer's instruction manuals have been utilized m the prepa- ration of the guide: 65. A-Line M anufacturing Company 66. BW/IP International, Inc.: Byron Jackson Pump Division 67. BW/IP International, Inc.: Seal Division 68. Bently-Nevada Corporation 69. Coupling Corporation of America: Flexxor Couplings 70. Demag Delaval: Pump Division, Centrimarc Bearings 71. Falk Corporation: Gear Couplings 72. Five Star Seal Corporation

BIB-4 EPRI Licensed Material Main Feedwater Pump Maintenance Guide

73. General Electric Company: Turbine Division 74. Koppers Company: Kop-flex Couplings 75. Ingersoll-Dresser Pump Company: Ingersoll-Rand Pump Company, Pacific Pump Company, Worthington Pump Company 76. Pruftechnic-Ludeca Incorporated: Optalign 77. Sulzer Bingham Pump Company 78. Westinghouse Electric Corporation 79. Zurn Corporation: Couplings

BIB-5 EPRI Licensed Material