Ez E.?

NUREG /CR-4597

Volume 1

ORNL-6282/V1

RV'

OAK RIDGE

NATIONAL

LABORATORY Aging and Service Wear of Auxiliary

Feedwater-Pumps for PWR

Nuclear Power Plants A a

V.

I

f '.

I

I r

I'

Volume 1. Operating Experience

and Failure Identification

M. L. Adams

E. Makay

208 LIBRARY

OCT 13 1986

REACTOR SAFElY and

Work Per-forM~tE HSC

U.S. Nuclear Regulat r o mission

Office of Nuclear Regulatory Research

Under DOE Interagency Agreement No. 40-551-75

NRC FIN No. B0828

OPERATED BY

MARTIN MARIETTA ENERGY SYSTEMS; INC.

FOR THE UNITED STATES

DEPARTMENT OF ENERGY
NOTICE

This report was prepared as an account of work sponsored by an

agency of the United States Government. Neither the United

States Government nor any agency thereof, or any of their

employees, makes any warranty, expressed or implied, or

assumes any legal liability or responsibility for any third party's

use, or the results of such use, of any information, apparatus

product or process disclosed in this report, or represents that its

use by such third party would not infringe privately owned

rights.

Available from

Superintendent of Documents

U.S. Government Printing Office

Post Office Box 37082

Washington, D.C. 20013-7982

and

National Technical Information Service

Springfield, VA 22161
NUREG/CR-4597

Volume 1

ORNL-6282/Vl

Dist. Category RV

AGING AND SERVICE WEAR OF AUXILIARY FEEDWATER PUMPS

FOR PWR NUCLEAR POWER PLANTS

Volume 1. Operating Experience and

Failure Identification

M. L. Adams E. Makay

Manuscript Completed - June 24, 1986

Date Published - July 1986

Report Prepared by

Energy Research and Consultants Corporation

900 Overton Avenue

Morrisville, Pennsylvania 19067

under

Purchase Order No. 1IX-28630V

for Oak Ridge National Laboratory

Work Performed for

U.S. Nuclear Regulatory Commission

under

DOE Interagency Agreement No. 40-551-75

NRC FIN No. B0828

OAK RIDGE NATIONAL LABORATORY

Oak Ridge, Tennessee 37831

operated by

MARTIN MARIETTA ENERGY SYSTEMS, INC.

for the

U.S. DEPARTMENT OF ENERGY

under Contract No. DE-AC05-840R21400
iii

CONTENTS

Page

LIST OF FIGURES .......... ... ... ... ........... *.*. * ** **** ....... vii

LIST OF TABLES ...................................... ... ix

ACKNOWLEDGMENTS ...................... xi

SUMMARY ProjectiScopei....i..... .. ....... ... ..... .. ...... .. .. .*.. xiii

ABSTRACT Definitions... ... .... ... *.* * * *** * ** ***** ...... 1

1. INTRODUCTION . ..................... .1... .. ............ ..... . 4

1.1 Background T.of ....... AuFeedwater.Pumps................ .. 1

1.2 Project Scope undaries2...*............*........*.. 2

1.3 Definitions Reur ....... t . .. .. . .. . .. . .. .. . . .. .. . .. . .. .. 3.. 8

2. BACKGROUND INFORMATION ......... ............................. 4

2.1 Principal Types of-,Auxiliary Feedwater Pumps ..............................4

2.2 Equipment Boundaries ................................... 7

2.3 Functional Requirements ... ooo................o..............8

2.3.1 General requirements o*. oo**oe............................. oo#*........ 8

2.3.2 Seismic requirements *.............................. 8

2.3.3 Miscellaneous *.. . .............................. .. 8

2.4 Materials of-Construtction ....* ......................... 8

3. TECHNICAL SPECIFICATION REQUIREMENTS ... ..................... 10

4. SUMMARY OF OPERATING CONDITIONS AND STRESSOR INFLUENCES ... o.. 11

4.1 Description of Typical Operating Regimes ................ . 11

4.1.1 Regimes ... ...................................... 11

4.1.2 Pump operating information for each regime ...... 11

4.1.3 Pump external environment ... ......................... 12

4.2 Stressor Description .... ............. *...........**..... 12

4.2.1 Mechanical .... oo... .... *oo................ .... 12

4.2.2 Hydraulic oo..................................................... 12

.4.2.3 Tribologicalo oo.o..oo.o.o.o..oo.o.o.o..........................13

4.2.4 Chemical ....... 00.. ...... .......................... . 0*.000*... 13

4.2.5 Low-relevance factors .....ooo.....o...ooo......... 13

4.3' Stres-sor-Influeonce o..............o................'ooooooo .o.o.o.o.ooo.. 13

4.3.1 Pump parts and components ... o ... o-ooo.ooo...... 13

4.3.2 Operating 'regimes .. .. .o.o..o.. o...............o.o ... . 18
iv

Page

5. OPERATING EXPERIENCE .. ......... 24

5.1 Failure Modes and Failure Causes .. ..................... 24

5.2 Frequency of Failure ................. .................. 26

5.3 Methods of Detection ............................ ....... 26

5.4 Maintenance Action .. ................................... 26

5.5 Modifications Resulting from Failures .................. 26

6. FAILURE MODES AND FAILURE CAUSES ............... o.............. 27

6.1 Failure Modes .......................................... 27

6.1.1 Failure to operate * ...... .. ... 27

6.1.2 Failure to operate as required .................. 27

6.1.3 External leakage ............ ............ .. .... 28

6.2 Failure Causes . .......... . .o. .. ..... . ................. . 28

6.2.1 Causes of failure to operate .......... ............ 28

6.2.2 Causes of failure to operate as required ........ 30

6.2.3 Causes of external leakage .... .................. 31

6.3 Failure Cause Analysis .................. ............... 32

6.3.1 Large hydraulic forces and vibarations created

by pump hydraulics .......... ........ . . . . . . . . . . . . . . 32

6.3.2 Shaft breakage .*..o.. ......... .. . . ............... 33 £

6.3.3 Impeller and diffuser breakage .................. 34

6.3.4 Thrust bearing and thrust balancer failures ..... 34

6.3.5 Coupling breakage ........ . . . ........ ... .*.O. ... . .. . . .. 34

6.3.6 Seizure ......... ........................... . . . 35

7. RECOMMENDED MAINTENANCE, SURVEILLANCE, AND MONITORING

PRACTICES .... ; .............. so ............................. 36

7.1. Regular Maintenance . ... ..... ............ ... ... . 36

7.2. Surveillance s...................... o.............. * *o........... ... 37

7.2.1 Current surveillance practice and

limitations ........... 000..oo.................................. 37

7.2.2 Interim recommendations for surveillance and

monitoring practices ......... ................... 37

7.2.3 Interim recommendations for detailed

inspection program . .......... .... .......... . . . . . . ... .. 38

8. AGING AND SERVICE WEAR MONITORING ........................... 39

8.1 Monitoring of Present Configurations ................... 39

8.1.1 Rotor binding check ........... . ....... . ...... 39

8.1.2 Shaft seal leak-off flow ........................ 39 £

8.1.3 Disassembly'and detailed inspection ..... so ...... 40

.1
v

8.2 Continuous Parameter Monitoring ......................

8.2.1 Rotor vibration monitoring ...................

8.2.2 Bearing temperatures and noises ................

8.2.3 Rotor axial position ...........................

8.2.4 Pump head-capacity curve ..... . .

8.3 Summary of Aging and Service Wear Factors .............

9. SUMMARY AND RECOMMENDATIONS .............. ...... .s.. ..... ...

9.1 Summary ................................

9.1.1 Bypass flow criteria .........................

9.1.2 Secondary bypass flow test loop .....

9.1.3 Monitoring and establishing trends ............

9.1.4 Scheduled disassembly and detailed t'

inspections so ....... .... . ... . .....

9.1.5 Pump specifications ..... . ...... .......

9.2 Recommendations .... .o. . ........ . .......

REFERENCES ....... ...O ........... I R AP .. .... ............ E C E ....

BIBLIOGRAPHY ...................... *....................

APPENDIX A. SUMMARY OF ASME BOILER AND PRESSURE VESSEL CODE

Page

43

43

44

44

44

45

51

51

51

51

52

52

52

52

54

55

59

63

67

APPENDIX

APPENDIX

APPENDIX

B.

C.

D.

SECT. XI REQUIREMENTS ..............................

OPERATING EXPERIENCE DATA BASES AND REPORTS ........

AUTOMATIC TRIPPING AND FAILURE ..................

ENGINEERING INFORMATION RELEVANT TO AUXFP

RELIABILITY ...................... a........... ....... 69
vii

LIST OF FIGURES

Figure Page

1 Double-suction, opposed-impellers AUXFP ................ 5

2 Single-suction, opposed-impellers AUXFP ......... ** ..... 6

3 Single-suction, in-line-impellers AUXFP ................ 6

4 NPAR program strategy .* ...... . .. .. .. ... ................ e .... 53

D.1 Anticipated useful operating ranges for pumps used

in large nuclear and fossil power generating units 70

D.2 Formation of stall (a) in diffuser and (b) in eye

of impeller .... *..o .. * .*.. e..*.... 71

D.3 Secondary flow pattern in and around pump impeller

stage at off-design flow operation ..................... 72

D.4 Head-capacity characteristics of multistage boiler

feed pumps . . ... . ... ... . . .. .. ..*. * .. ..... 73

D.5 Influence of pump impeller to diffuser/volute radial

gap on pressure pulsation at blade-passing frequency

and rotor deflection caused radial forces 74

D.6 Frequencies of hydraulically induced dynamic forces

acting on the rotor of a centrifugal pump ....... * ... 76

D.7 Vibration frequency vs speed .......... ................. 77

D.8 Allowable rotor vibration levels measured relative

to the bearing cap ... . ... ..... so......... 78

D.9 Customary axial balancing devices for high-pressure

multistage boiler feed pumps ........................... 81

D.10 Parallel and tapered face balance disk designs ......... 82

D.11 Shaft seal types used in boiler feed, nuclear

feed, and feedwater booster pumps ...................... 84

D.12 Typical AUXFP bearing system ...... ..................... 85

D.13 Typical multistage centrifugal pump shaft failure

locations ........................................ 86
ix

LIST OF TABLES

Table Page

1 Examples of typical AUXFP BEP operating parameters ...... 12

2 Stressor influence on rotating elements .................. 19

3 Stressor influence on nonrotating internals ............. 20

4 Stressor influence on pressure-containment casing ....... 21

5 Stressor influence on mechanical subsystems ............. 22

6 Stressor influence on support ... *.................. ... 23

7 Summary of AUXFP pump failure information available from

operating experience and plant documents .....*........... 25

8 Failure modes ..... .... .. .. . ... . ... .. ...... 0 *.... ...-*....*. 28

9 Summary of failure modes and causes .... .... ............. 29

10 Pump failure causes related to aging and service

wear o.....oo.ooooooooooo...oo.ooo..oo.oo.. o.oo............... 45

11 Methods currently used to detect AUXFP failure modes .... 46

12 Methods for differentiating between failure causes ...... 47

13 Measurable parameters ......... o...................... .o.o. 49

14 Summary of important AUXFP part failure assessment ...... 50

B.1 Summary of AUXFP-type failures reported in LERs

(1973-1983) ......... 0....................... 0.. ..0.0.0........ .. 64

B.2 AUXFP-type failures reported in NPRDS data base

(1974-1985) p.lrea .. .... d.a.astudy.. . .... .. o...o......... 65

B.3 In-plant reliability data study ...................... .o.o.o................ 66
xi

ACKNOWLEDGMENTS

The authors gratefully acknowledge the continuing support and counsel of the Nuclear Regulatory Commission Nuclear Plant Aging Research

Program manager, J. P. Vora, in the planning and implementing of this

study. W. L. Greenstreet of Oak Ridge National Laboratory directed the

work on this report, which was substantially configured through his

direction and advice.

The information, insights, conclusions, and recommendations contained in this document are based primarily on the design, troubleshooting, and redesign experience of Energy Research and Consultants

Corporation in power plants for electric utility companies. The authors

gratefully acknowledge the cooperation of several Electric Utility Companies, in particular, operating personnel at the Palo Verde nuclear

plant of Arizona Public Service Company.

In addition, we wish to express our thanks to the following pump

manufacturers for their openness ~in discussing their points of view:-

Ingersoll-Rand, Byron Jackson;--Pacific-Pump Division of Dresser Clark,

and Bingham Pump Companies.
xiii

SUMMARY

The primary focus of this report is on factors that are important to

aging and service wear of auxiliary feedwater pumps (AUXFPs). These

pumps are from the generic family of-multistage high-head centrifugal

pumps and are basically small boiler feed pump designs used in small

capacity, mostly older fossil fuel electric generating plants.

A general description of AUXFPs is provided and includes illustrations, construction and configuration details, defined equipment boundaries, functional requirements, and materials of construction. Technical

specification requirements and operating experience topics are summarized. Operational stressors are categorized, and a detailed stressor

list is provided, component-by-component.

Failure modes are defined as follows: O(1) failure to-operate,

(2) failure to operate as required,--and (3) external leakage. Failure

causes are identified in general terms and subsequently described in more

specific terms in a section on failure cause analysis. The single most

'important-factor-relevant-t6-AUXFP potential failures is the presence 6f

large hydraulic dynamic forces'within such pumps, particularly at flow'

rates substantially different -than the best-efficiency flow. Correspondingly, determining safe minimum AUXFP flow rates (i.e., bypass or recirculation flow) is a very important task. This topic is discussed at

length in Sect. 6 and in Appendix D.

Methods for detecting failure modes and differentiating between

failure causes are described. In addition, measurable parameters (including functional indicators) are identified for potential use in detecting and monitoring degradation and for tracking degradation trends.

Parameters for identifying failure causes of pumps include

Vibration Clearance

Delivered flow Rotor axial position

Rotational speed Leakage rate

Bearing temperature Appearance

Transmitted torque Local shaft temperature

Noise Bolt torque

The appropriateness and utility of these and other parameters will be

addressed in subsequent phases of the AUXFP investigation.

At present, most AUXFP installations contain no monitoring.devices

except for flow and head measurement, in contrast to continuously running

.power plant equipment. Furthermore, present surveillance practice consists primarily of starting each AUXFP at bypass flow once every 1 to 3

months for a short-duration test to verify operational readiness. Thus,

the establishment and correlation of AUXFP operating parameters (e.g.,

vibration and bearing temperatures) to wear and aging criteria are not

addressed in nearly all installations as presently configured and instrumented. Actions are detailed to address these present weaknesses. These

actions include (1) disassembly and inspection and component renewal at

refueling intervals and (2) parameter monitoring and instrumentation

retrofits. These and other important factors are addressed in Sect. 9.
xiv

Appendix D provides a major background section on engineering information and design factors critically important to AUXFP reliability.

The Bibliography contains an extensive list of over 40 publications and

reports, which have been prioritized to aid in selecting those that are

most critical to understanding the potential operating problems and failure modes of AUXFPs.
AGING AND SERVICE WEAR OF AUXILIARY FEEDWATER PUMPS

FOR PWR NUCLEAR POWER PLANTS

Volume 1. Operating Experience and

Failure Identification

M. L. Adams* E. Makay*

ABSTRACT

This rep6rt was produced under the Detection of Defects and

Degradation Monitoring element of the Nuclear Plant Aging Research Program. Typical auxiliary feedwater pump (AUXFP) configurations are described in terms of configuration details,

materials of construction, operating requirements, and modes

of operation. AUXFP failure modes and causes due to aging and

service wear are identified and explained, and measurable

parameters (including functional indicators) for potential use

in assessing operational readiness, establishing degradation

trends, and detecting incipient failures are given.

A series of measures to correct present deficiencies in

surveillance, monitoring, and in-service testing practices is

discussed. The main body of the report is supplemented by a

number of relevant appendixes; in particular, a major appendix

is included on engineering and design information useful to

assess AUXFP operational readiness.

1. INTRODUCTION

1.1 Background

The Office of Nuclear Regulatory Research of the Nuclear Regulatory

Commission (NRC) has instituted studies aimed at understanding the timerelated degradation (aging) of nuclear power plant systems and equipment,

assessing the effectiveness of methods of inspection and surveillance to

monitor such degradation, and establishing guidelines for maintenance.

This study is one in a series intended to provide technical bases to

assess the ongoing operational safety of operating plants. The strategy

followed can be used by others interested in analyses of equipment in

nuclear applications.

This report addresses time-related degradation of pressurized-water

reactor (PWR) power plant auxiliary feedwater pumps (AUXFPs). fBecause

failures of these components can reduce-the amount of feedwatereavailable

*Energy Research and Consultants Corporation, 900 Overton Avenue,

Morrisville, Pa. 19067.
2

for removing heat when the usual feedwater supply is unavailable, such

failures can result in altered safety margins for PWR systems.

1.2 Project Scope

This report is Volume 1 of a three-part report to be prepared on

AUXFPs. The contents of the three volumes are summarized below.

Volume 1 - Operating experience and failure identification (Phase 1)

1. Background information on AUXFPs - boundary of A1JXFPs to be studied,

types, functional requirements, and materials of construction;

2. Reviews of regulatory requirements, guides, and standards;

3. Summary of operational and environmental stressors;

4. Summary of operating experience;

5. Manufacturers' input; and

6. State-of-the-art aging and service wear monitoring and assessment.

Volume 2 - Aging assessments and evaluation of monitoring methods

(Phase 2)

1. Results from completion of comprehensive aging assessment based on

postservice examination and tests of aged components and in-situ

assessments,

2. Identification of monitoring techniques and review of information

produced, and

3. Evaluations of monitoring methods.

Volume 3 - Analysis and recommendations

1. Value impact analysis, and

2. Recommendations for guidelines for monitoring methods and maintenance.

One of the objectives of the Phase 1 research effort is to provide

baseline information for use in subsequent phases of the study. To prepare this report, operating experience, manufacturers' information, and

information derived from design, troubleshooting, and redesign experience

were reviewed to identify (1) failure modes and causes resulting from

aging and service wear of AUXFPs and (2) measurable parameters (including

functional indicators) with potential for quantifying and tracking degradation. These parameters are to be applicable for detecting and establishing time-dependent degradation trends before loss of function.
3

1.3 Definitions

For the purpose of this report, the following definitions apply.

Failure mode - the way a component does not perform a function for

which it was designed (e.g., fails to actuate or leaks to outside).

Failure cause - degradation (the presence of a defect) in a component that is the proximate cause of its failure (e.g., bent shaft, loss

of lubricant, or loosening of a bolt).

Failure mechanisms - the phenomena that are responsible for the

degradation present in a given component at a given time. Frequently,

several failure mechanisms are collectively responsible for degradation

(synergistic influences). One major failure mechanism, where identified,

has been called the "root cause." Generic examples of failure mechanisms

(and of root causes) include aging, human error, or seismic events.

Aging - the combined cumulative effects over time of internal and

external stressors acting on a component, leading to degradation of the

component, which increases with time. Aging degradation may involve

changes in chemical, physical, electrical, or metallurgical properties,

dimensions, and/or relative positions of individual parts.

Normal aging - aging of a component that has been designed, fabricated, installed, operated, and maintained in accordance with specifications, instructions, and good practice and that results from exposure

to normal stressors for the specific application. Normal aging should be

taken into account in component design and specification.

Measurable parameters - physical or chemical characteristics of a

component that can be described or measured directly or indirectly and

that can be correlated with aging. Useful measurable parameters are

those that (1) can be used to establish trends of the magnitude of aging

associated with each failure cause, (2) have well-defined criteria for

quantifying the approach to failure, and (3) are able to discriminate

between the degradation that leads to failure and other observed changes.

Inspection, surveillance, and condition monitoring (ISCM) - the

spectrum of methods and hardware for obtaining qualitative or quantitative values of a measurable parameter of a component. The methods may be

periodic or continuous, may be in-situ, or may require removal and installation in a test stand or disassembly and may involve dynamic or

static measurements.
4

2. BACKGROUND INFORMATION

2.1 Principal Types of Auxiliary Feedwater Pumps

AUXFPs for PWR systems are basically small boiler feed pumps used in

small-capacity mostly older, fossil fuel electric generating plants.

Thus, they retain most of the design features, plus the potential operating and reliability problems inherent in the feedwater pumps. A comprehensive bibliography of publications and reports document extensive field

troubleshooting experience and analyses that are most relevant to AUXFP

machinery. This bibliography also represents the several years of

experience upon which the contents of the report rest.

AUXFP configurations are all multistage (from 6 to 12) with lowspecific-speed, high head-per-stage (HHPS) impellers, having either vanediffuser or volute collection chambers at the discharge section of each

stage. The overall construction is always an axially (horizontal) split

outer casing with internal stage-to-stage connecting flow passages (see

Figs. 1-3).

More specific details of configuration vary considerably from vendor

to vendor. Even within the array of available designs from a single

manufacturer, there can be a considerable variety of construction details. The main factor that determines these more specific configuration

details is the design approach used in handling the very large axial

thrust forces on the rotor, which arise as a natural by-product in HHPS

centrifugal pumps. A second factor is the design approach used to avoid

cavitation in the inlet (suction), that is, single-suction (Figs. 2 and

3) or double-suction (Fig. 1) inlet stage.

Some manufacturers use a configuration with all stages positioned in

the same axial direction (Fig. 3, called "in-line" type). Other manufacturers use a configuration with the stages on one-half of the shaft

axially opposed to those on the other half of the shaft (Figs. 1 and 2,

called "opposed" type) for thrust cancellation. The basic difference

between these two configurations involves the previously referenced different design approaches to handling the very large axial thrust loads

inherent in multistage HHPS centrifugal pumps. Opposed type configurations (i.e., thrust cancellation approach) are often designed with only a

thrust bearing. In-line type configurations are always designed with an

additional thrust balancer of either "balancing drum" type or "balancing

disk" type (see Appendix D).

Plants have traditionally been configured with two AUXFPs per

pressurized-water reactor (PWR). In most plants, one AUXFP is steam turbine driven, and the other is electric motor driven. The steam-turbinedriven unit has the advantage of variable speed, and the maximum design

operating speeds are generally in the 4000- to 5000-rpm range. The

motor-driven unit is usually designed for two-pole induction motor speed

and powered from either one of two standby diesel-driven generators.

Newer plants may also have a third "backup" AUXFP that is electric motor

driven either from main line power or switchable to the diesel-driven

generator bus. Thus, to provide very high standby reliability of the

AUXFP system, the number of pumps employed (now three) and the number of
ORNL-DWG. 86-4223 ETD

DOUBLE SUCTION

FIRST STAGE /p\\ * s/ SVW A.. \E 9 \ ................. SHAFTSEAL

STUFFING BOX) JOURNAL

JOURNAL (STUFFING BOX) RTFNGBO BEARING BEARING ADUIMNG Bo

DRIVE END I 14

| S.S. I > s# # >SHAFT SLEEVES i f S.S. . .. ........ . ............................... t\W~~~~~~~~~~~iiA- - - -- - - .. ....... ...... ---- ---_ =4w

Fig. 1. Double-suction, opposed-impellers AUXFP.
6

ORNL-DWG. 86-4224 ETO

Fig. 2. Single-suction, opposed-impellers AUXFP.

ORNL-DWG. 86-4225 ETD

SUCTION

I

THRUST

BALANCER

Fig. 3. Single-suction, in-line-impellers AUXFP.
7

separate power sources (i.e., steam, diesel-generator, main line) are

triple redundant. The overall-approach that now is emerging in the newest plants is to have-the backup AUXFP be a non-safety-related, nonnuclear class (nonclass). pump.(thus cheaper) and to use it for all normal

plant startup, shutdown, and nonemergency service. However, the Plant

Technical Specification requires readiness testing of nuclear class

(N-class) AUXFPs at least every 3 months.

The rated discharge pressure for nuclear plant AUXFPs ranges from

1100 to 1500 psig. The rated capacity ranges from 200 to 1000 gal/min.

However, rated head and capacity are generally pegged to the bestefficiency-point (BEP) flow at a given speed, and AUXFPs are required to

operate over a wide range of percent BEP flow. They are typically required to operate from 10 to 140%-of--BEP capacity. Thi s-is considered a

rather abusive operating requirement from the point of view of acceptable

practice for main feedwater pumps.

2.2 Equipment Boundaries

For present purposes,-the-primary focus is on-the pump per se, the

drivers being outside the scope of this report. However, it is appropriate to give consideration to defining-a-'system, that is, equipment boundaries.

The pump is mechanically coupled to its driver (steam turbine or

electric motor). The pump-driver assembly is typically mounted on a common fabricated steel base 'that is stiffly connected to some structural

"floor" section within the plant. Furthermore, the pump is hydraulically

and mechanically connected to a fluid flow circuit through its inlet

(suction) and discharge nozzles. For a turbine-driven pump, the turbine

inlet is, of course, piped into a steam flow circuit and the outlet to a

condenser. For a motor-driven pump, the motor is electrically connected

to a 60-Hz, 3-phase power supply. Thus,' the equipment boundaries will be

taken to include the driver and pump on a common base and with the boundary taken to include all inlet and discharge nozzles as well as direct

electrical connections. However, aging, wear, and failures of the driver

and connecting steam or electrical power lines are not addressed in this

report.

Failures outside this defined boundary will also not be considered

in this report. The small steam turbines and the electric motors used to

drive AUXFPs are standard designs used in many other applications as well

and as yet do not appear to exhibit any aging'or wear factors that are

unique to the AUXFP application. Based on operating experience on similar pumping systems for other applications, driver failure can be caused

by a massive failure of the pump. Primarily, this would be a pump failure resulting in loss of thrust balancer-and thrust bearing load-carrying

capacity. In such a situation,-depending upon the specific- pump and

coupling design,-there is a potential for subjecting the driver (i.e.,

its thrust bearing) to dadmaging-load-levels. Excessive pump vibration is

another factor that could accelerate wear and aging of the turbine or

motor driver.
8

2.3 Functional Requirements

AUXFP functional requirements may vary somewhat from plant to plant

and also vary as a function of plant vintage. The following itemization

of functional requirements includes the typical ranges for quantifiable

parameters where appropriate.

2.3.1 General requirements

1. Supply feedwater to the steam generators under plant startup, normal

shutdown, hot standby, and emergency conditions;

2. meet automatic starting requirements;

3. operate under normal and accident conditions of temperature, pressure, humidity, radiation, and available net positive suction head

(UPSH); and

4. withstand seismic loadings without loss of function.

2.3.2 Specific requirements

1. Capable of producing the rated flow against the rated head within

20 s (typical) after actuation,

2. total of 1000 service-life operating cycles [includes all items in

2.3.1 (1)],

3. AUXFP operating cycle including up to 10 h of continuous pumping,

4. operation of pump at shutoff (upset condition) for up to 10 min,

5. a plant life of 40 years,

6. a steadily rising head-capacity characteristic with a shutoff head in

the range of 115 to 130% of best-efficiency head at a given speed,

7. an efficiency characteristic commensurate with state-of-the-art

hydraulic design, and

8. capable of operating at any point on the head-capacity curve, from

minimum flow to full run-out condition (i.e., maximum pump flow

possible against system resistance) during and following safe reactor

shutdown at earthquake condition.

2.3.3 Miscellaneous

1. To be available for operation, except when taken off-line for maintenance or testing, and

2. to be supplied with appropriate protective devices such as overspeed

trip.

2.4 Materials of Construction

Precise material designations vary somewhat from vendor to vendor.

Also, in a number of AUXFP components, the current material or method of

forming may not be the optimum for this application. The primary reason

for this is that purchaser specifications are not always optimally explicit, potentially resulting in competing vendors using lower cost options.
Thus, the following ''arag-hs''l'oit-only summarize-current- practice but

also contain suggested alternatives that should be studied.

The two parts of the split casing, including inlet and discharge

nozzles, are made of a cast carbon or stainless steel. The most common

is the cast equivalent of 416 stainless steel, designated when cast as

CA6NM.

The shaft is currently machined from'400-series'stainless steel bar

stock by all vendors. Typical material designations are 410, 414, and

416 stainless steel. A more durable alternative for consideration is to

make the shaft from a forging.

Impellers are made of high-chromium alloy steels and are sand cast

by all vendors. Commonly used alloys are 13-5 (i.e., 13% Cr, 5% Ni),

15-5, and 17-4 Ph. Here again, current practice may not have been optimally constrained by purchaser specifications. That is, impellers could

be specified as precision cast to provide maximum quality both in terms

of structural integrity and dimensional control of critical hydraulic

passages.

The diffusers and stage-to-stage return channel components are also

sand cast with the material varying from vendor to vendor, for example,

stainless steel 440A, 440B, and 17-4 Ph. As in'the case of the impellers, optimum dimensional control of critical hydraulic passages suggests

that precision cast or milled diffusers and return channels should be

considered.

The coupling (used to connect pump to driver) materials vary considerably with coupling type and manufacturer. The type of coupling used

in nearly all installations is the gear type. However, tooth clearances

in such couplings, always produce some measurable rotor unbalance that, is

not correctable. Furthermore, lubrication is required to prevent wearout

in service. For these reasons, the dry flexible diaphragm type of coupling should be considered.

The main pressure-closure retaining bolts (studs) are made of high

tensile strength Cr-Mo steels, suitable for the high prestressing necessary to ensure a leaktight and structurally sound closure at the casing

joint.

Various shaft sleeves are made of the same steel alloys as the

shaft. Shaft-sleeve retaining bolts are also made of similar steel

alloys as the shaft.

Balancing disks, balancing drums, and wear rings are typically

machined from stainless steel alloy material such as 416 and 420. A

considerable enhancement to reliability can be ensured if the material is

designated as 420F, which is a "free machining nongalling' stainless

steel with a trace of phosphorus (0.17%) or sulfur (0.3%).

Oil-film journal bearings and thrust bearings are typically constructed of carbon steel shells with bearing surfaces lined with babbitt

white metal (usually tin base for best results) that is cast into the

preheated bearing shell and finished machined to the specified final

dimensions. Rolling contact bearings, which are available to take radial

and axial loads, are constructed of various specialty steels.typically

used in such bearings. These bearings are case hardened to rprovide very

high strength outer (surface) properties and a more ductile inner core of

the rolling contact elements. _
10

3. TECHNICAL SPECIFICATION REQUIREMENTS

Testing requirements for AUXFPs in nuclear power plants are described in the plant Technical Specifications that state the overall inservice inspection requirements for safety-related systems and components.

In-service inspection and testing of ASME Class 1, 2, and 3 components

are to be in accordance with Sect. XI of the ASME Boiler and Pressure

Vessel Code.

The Sect. XI testing requirements for pumps involve measuring inlet

pressure, pressure differential across the pump, flow rate, and vibration

amplitude; all are measured while operating at specific design conditions.

The test quantities are compared with reference values.

The ASME Code Sect. XI contains surveillance intervals, frequencies,

and test requirements for Code Class 1, 2, and 3 systems. A review of

Subsection IWP, In-service Testing of Pumps in Nuclear Power Plants, is

given in Appendix A.

The purpose of Technical Specifications is not specifically to monitor degradation of performance but to ensure operability of components

and systems within specified limits required to perform the desired

safety function.

*Work performed by G. A. Murphy, ORNL Nuclear Operations Analysis

Center.
I

11

4. SUMMARY OF OPERATING CONDITIONS AND STRESSOR INFLUENCES

4.1 Description of Typical Operating Regimes

4.1.1 Regimes

AUXFPs are used to supply feedwater to the steam generators under

plant startup, shutdown, and emergency conditions. These pumps have a

BEP flow at any operating speed, and if used in a continuous operating

mode, normal delivered flow is between 50 and 120% of the BEP flow. However, the AUXFP application is by definition a transient operational

mode, that is, startup,* shutdown,* and emergency. Thus, there are basically two operating regimes: (1) standby and (2) normal. Normal includes

any flow from shutoff or bypass flow to full run-out flow. Alternately,

if the term "normal operation" pertains to that regime in which the pump

resides most of the time, then normal operation would be the standby condition, and any condition under which the AUXFP is pumping would be considered off-normal.

It is therefore more rational to categorize operating regimes in a

manner that is most relevant to aging and wear factors. For that reason,

the following are the defined operating regimes: (1) standby (i.e., 0

flow), (2) 0 to 50% BEP flow, (3) 50 to 120% BEP flow, and (4) 120% BEP

flow to full run-out flow (typically 150% BEP flow). The relative importance and relevance of some operational stressors will be different for

each of these defined operating regimes.

4.1.2 Pump operating information for each regime

The following operating parameter ranges are based on a representative number of plants listed in Table 1.

1. BEP flow, 200 to 1000 gal/min;

2. BEP discharge pressure, 1100 to 1500 psig;

3. available suction pressure, -2.6 to 60 psig;

4. speed, 3560 to 4400 rpm;

5. shutoff head, 115 to 130% BEP head; and

6. pumped water temperature, 60 to 1250F.

These BEP parameter values, when keyed with the operating regimes

defined in the previous section, provide the essential quantitative information for each regime.

*In some plants, startup and normal shutdown are handled by a separate (nonclass) motor-driven pump, frequently referred to as the auxiliary feedwater backup pump.
12

Table 1. Examples of typical AUXFP BEP operating parameters

Number Speed Flow Suction Discharge

Pe pn s(rpm) (gal/min) (psig) (psig) stages

TVA Sequoyah 5 3950 920 25 1120

TVA Sequoyah 9 3570 440 25 1252

North Anna 1 and 2 6 4200 735 -2.6 1216

North Anna 1 and 2 8 3560 370 -2.6 1216

TVA Watts Bar I and 2 6 3850 1000 10 1277

TVA Watts Bar 1 and 2 9 3577 500 10 1277

Seabrook 1 and 2 9 3577 500 10 1277

Shearon 9 4400 850 0 1270

Shearon 9 3550 425 0 1270

Ginna 1 10 3560 200 23 1475

Maine Yankee 5 4400 530 0 1095

Maine Yankee 8 3575 500 0 1095

Donald C. Cook Station 6 4350 900 25 1195

Donald C. Cook Station 8 3560 450 25 1195

Beaver Valley Station 8 3560 370 0 1165

Arkansas Power and Light Co. 9 3560 780 60 1172

Three Mile Island 8 3560 470 23 1128

Three Mile Island .6 4250 940 23 1128

4.1.3 Pump external environment

1. Indoor installation;

2. atmospheric pressure;

3. temperature, 60 to 105'F;

4. 40-year cumulative radiation, 200 rads;

5. pumped fluid not expected to be radioactive; and

6. maximum relative humidity, 100%.

4.2 Stressor Description

4.2.1 Mechanical

1. Torque transmitted loads (static and dynamic),

2. assembly (fastener) loads,

3. rotor-dynamic loads (e.g., unbalance),

4. piping forces,

5. seismic loads, and

6. vibration (for all sources).

4.2.2 Hydraulic

1. Hydraulic loads (static and dynamic),

2. fluid impingement,

3. internal pressure, and

4. cavitation.
13

4.2.3 Tribological*

1. Rubbing between rotating and nonrotating members [potentially more

severe in the presence of 4.2.3 (7)],

2. bearing lubricant breakdown (viscosity and various chemical additive

breakdowns),

3. surface fatigue (life limiting for rolling contacts and gears),

4. contamination and degradation of lubricant,

5. starts and stops,

6. fretting, and

7. surface oxide abrasive formation (see 4.2.4).

4.2.4 Chemical

Corrosion (oxidation) of 400-series stainless steels through chemical reaction with stagnant water can produce an oxide surface scale.

Chlorides and other feedwater impurities have occurred from turbine condenser leakage and can increase the rate of this type of corrosion. Furthermore, this chemical process can result in atomic hydrogen diffusing

into the metal surface to form molecular hydrogen that can cause surface

"blisters."

Other chemical-effects that may have 'influence are stress corrosion

cracking, pitting,,-'and crevice-corrosion. Stress corrosion cracking may

occur, depending on the impurities present in the water, but it.-is-very

unlikely to.happen -at the operating temperatures of these pumps. Pitting

is possible, depending on the type of impurities- in the water-and their

concentrations; potential pitting damage is projected to be small for

AUXFPs, however. Crevice corrosion is another-possible''factor,-but because of the low impurity-content-and the-relatively low temperature of

the water it is not likely to-be 'significant.

4.2.5 Low-relevance factors

From the point-of-view of aging and wear considerations, the following stressor categories are not significant to AUXFPs:

1. thermal,

2. radiation, and

3. environmental [except for earthquakes, covered in 4.2.1 (5)].

4.3 Stressor Influence

4.3.1 Pump parts and components

The pump has been subdivided into five major segments: rotating

elements, nonrotating internals, pressure-containment casing, mechanical

*Tribology is now the unifying label for that field that encompasses

friction, wear, lubrication, and machinery components affected by same.
14

subsystems, and support base. Each of these segment categories is further broken down into individual components listed below.

Pump segment Parts

Rotating elements

Nonrotating internals

Pressure-containment

casing

Mechanical subsystems

Shaft

Impellers

Miscellaneous spacers

Diffusers or

volutes

Return channels

Upper casing

Lower casing

Fasteners

Thrust bearing

Radial bearings

Shaft seals

Thrust runners*

Fasteners

Wear surfaces

Fasteners

Suction and

discharge

nozzles

Thrust balancer

Coupling

Fasteners

Support Base frame Fasteners

The particular stressors that are significant to each part are

enumerated in the following subsections, segment-by-segment. The relative importance of some stressors is a significant function of the particular operating mode. The single potentially important chemical stressor previously listed (Sect. 4.2.4) is carried through this document as a

tribological stressor because its relevance to wear, aging, and failure

modes stems from the abrasive surface oxides and blisters that might

potentially be formed while the pump is on standby [see 4.2.3 (7)].

4.3.1.1 Rotating elements

Shaft

Mechanical stressors: Transmitted torque, fastener loads, rotor-dynamic

loads

Hydraulic stressors: Hydraulic loads

Tribological stressors: Rubbing between rotating and nonrotating members, bearing lubricant breakdown, dirt in lubricant, starts and stops,

surface oxide abrasives

*A thrust runner (also called a thrust collar) is the rotating part

captured by a double-acting oil-film thrust bearing. In addition, the

rotating part of a balancing disk assembly is called a thrust runner. In

many AUXFPs, these two components are superceded by the use of a ball

bearing for carrying axial thrust loads.
15

Impellers

Mechanical stressors: Transmitted torque, fastener loads, rotor-dynamic

loads

Hydraulic stressors: Hydraulic loads, cavitation, fluid impingement

Tribological stressors: Rubbing between rotating and nonrotating members, fretting, surface oxide abrasives

Miscellaneous spacers

Mechanical stressors: Fastener loads, rotor-dynamic loads

Hydraulic stressors: Hydraulic loads

Tribological stressors: Rubbing between rotating and nonrotating members, fretting, surface oxide abrasives

Thrust runners

Mechanical stressors: Fastener loads, rotor-dynamic loads

Hydraulic stressors: Hydraulic loads

Tribological stressors: Rubbing between rotating and nonrotating members, bearing lubricant breakdown,* dirt in lubricant,* starts and stops,

fretting, surface oxide abrasives

Fasteners

Mechanical stressors: Transmitted torque, assembly loads, rotor-dynamic

loads

Hydraulic stressors: Hydraulic loads

Tribological stressors: Fretting

4.3.1.2 Nonrotating Internals

Diffusers or volutes

Mechanical stressors: Fastener loads

Hydraulic stressors: Hydraulic loads, fluid impingement, cavitation

Tribological stressors: Fretting

Return channels

Mechanical stressors: Fastener loads

Hydraulic stressors: Hydraulic loads, fluid impingement

Tribological stressors: Fretting

*May not apply since most, if not all, AUXFP thrust bearings are of

the rolling contact type, not the hydrodynamic oil-film type.
16

Wear surfaces

Mechanical stressors: Assembly loads, rotor-dynamic loads

Hydraulic stressors: Fluid impingement

Tribological stressors: Rubbing between rotating and nonrotating menbers, starts and stops, surface oxide abrasives

Fasteners

Mechanical stressors: Assembly loads, vibration

Hydraulic stressors: Hydraulic loads

Tribological stressors: None

4.3.1.3 Pressure-containment casing

Upper casing

Mechanical stressors: Assembly loads, piping forces, seismic loads,

vibration

Hydraulic stressors: Hydraulic loads, fluid impingement, internal pressure

Tribological stressors: None

Lower casing

Mechanical stressors: Assembly loads, piping forces, seismic loads,

vibration

Hydraulic stressors: Hydraulic loads, fluid impingement, internal pressure

Tribological stressors: None

Suction and discharge nozzles

Mechanical stressors: Assembly loads, piping forces, seismic loads,

vibration

Hydraulic stressors: Hydraulic loads, fluid impingement, internal pressure

Tribological stressors: None
17

4.3.1.4 Mechanical subsystems

Thrust bearing

Mechanical stressors: Assembly loads, vibration

Hydraulic stressors: Hydraulic loads

Tribological stressors: Surface fatigue, dirt

Radial bearings

Mechanical stressors: Assembly loads, rotor-dynamic loads

Hydraulic stressors: Hydraulic loads

Tribological stressors: Rubbing between rotating and nonrotating members, bearing lubricant breakdown, surface fatigue, dirt in lubricant

Thrust balancer

Mechanical stressors: Assembly loads, vibration

Hydraulics stressors: Hydraulic loads

Tribological stressors: Rubbing between rotating and nonrotating members, starts and stops, surface oxide abrasives

Shaft seals*

Mechanical stressors: Assembly loads, vibration, improper adjustment

(overtightening) to correct for normal wear of packing

Hydraulic stressors: Internal pressure

Tribological stressors: Rubbing between rotating and nonrotating members, starts and stops, surface oxide abrasives

Coupling

Mechanical stressors: Transmitted torque, assembly loads, rotor-dynamic

loads

Hydraulic stressors: None

Tribological stressors: Lubricant breakdown, surface fatigue, dirt in

lubricant, fretting

*Mechanical or stuffing-box seals.
18

Fasteners

Mechanical stressors: Transmitted torque, assembly loads, rotor-dynamic

loads, vibration

Hydrodynamic stressors: Hydraulic loads

Tribological stressors: Fretting

4.3.1.5 Suppcrt

Base frame

Mechanical stressors: Assembly loads, piping forces, seismic loads,

vibration

Hydraulic stressors: None

Tribological stressors: None

Fasteners

Mechanical stressors: Assembly loads, seismic loads, vibration

4.3.2 Operating regimes

The influence of stressors is described in Tables 2-6. Each of

these tables pertains to a specific pump segment as previously defined in

Sect. 4.3.1. Each segment is broken down into specific parts, and each

part is briefly commented upon in regard to stressor influence for each

operating regime. In Sect. 4.1.1, operating regimes were defined or

categorized in a manner that is most relevant to aging and wear factors:

(1) standby, (2) 0 to 50% BEP flow, (3) 50 to 120% BEP flow, and (4) 120%

BEP flow to full run-out flow (typically 150% BEP flow).

The relevant stressor influence identified for the standby condition

is abrasive surface oxide and blister formation. However, the potential

wear mechanism resulting from this would arise when the pump is operated,

not at standby per se. Also, this is the single significant chemical

stressor identified in Sect. 4.2.4 but is subsequently herein referenced

as a tribological stressor [see 4.2.3 (1) and (7)] due to its relevance

primarily as a potential wear accelerator at close running clearances.

Therefore, Tables 2-6 do not list this as a chemical stressor nor reference the standby condition.

The approach to AUXFP usage adopted in at least one of the newest

plants (Palo Verde) is the use of an electric-motor-driven nonclass pump

(of similar if not same design as AUXFPs) for startup and shutdown service. This approach renders the AUXFPs on standby all the time, except

for occasional short duration testing. This approach, which clearly

tends to preserve the AUXFPs for true emergency service, probably also

reduces operating costs. That is, when comparing two pumps of the same

design where one is purchased as an N-class pump and the other as a nonclass pump, the class pump and its related spare parts are priced considerably higher by vendors than the corresponding nonclass pump.
r V V

4

~~~~~~~~~~~~~~ -

Table 2. Stressor influence on rotating elements

Operating Stressor influence and remarksa

- Part regime

(% BEP flow) Mechanical Hydraulic Tribological

Shaft 0-50 (1) (3) Hydraulic (3) Rubbing at close-running

dynamic loads clearances due to hydraulic-load-induced rotor

vibration

50-120 (1) (1) (1)

120-R.0. (1) (3) Hydraulic (3) Same as 0-50

dynamic loads

Impellers 0-50 (1) (3) Cavitation (3) Rubbing at wear surfaces

50-110 (1) (1) (2) Same as 0-50

120-R.O. (1) (3) Same as 0-50 (3) Same as 0-50

Miscellaneous 0-50 (1) (1) (3) Rubbing at shaft seal

spacers sleeve due to rotor

vibration

Thrust 0-50 (1) (3) Hydraulic (3) Rubbing due to hydraulic

runners dynamic loads loads

50-120 (1) (1) (1)

120-R.O. (1) (3) Same as 0-50 (3) Same as 0-50

Fasteners 0-50 (1) (2) Hydraulic (1)

dynamic loads

50-120 (1) (1) (1)

120-R.0. (1) (2) Same as 0-50 (1)

aTo indicate a relative measure of importance of each stressor category, an approximate

quantifier is shown as either low (1), medium (2), or high (3).

bR.O. = run-out flow.

,D
Table 3. Stressor influence on nonrotating internals

Operating Stressor influence and remarksa

Part regime

(% BE? flow) Mechanical Hydraulic Tribological

Diffusers or 0-50 (1) (3) Unsteady flow (1)

volutes forces

50-120 (1) (2) Same as 0-50 (1)

120-R.0. (1) (3) Same as 0-50 (1)

Return 0-50 (1) (2) Unsteady flow (1)

channels 50-120 (1) (1) (1)

120-R.0. (1) (2) Same as 0-50 (1)

Wear surfaces 0-50 (1) (1) (3) Rubbing due to rotor

vibration

50-120 (1) (1) (2) Same as 0-50

120-R.0. (1) (1) (3) Same as 0-50

Fasteners 0-50 (1) (2) Hydraulic (1)

loads

50-120 (1) (1) (1)

120-R.0. (1) (2) Same as 0-50 (1)

aTo indicate a relative measure of importance of each stressor category, an

approximate quantifier is shown as either low (1), medium (2), or high (3).

bR.O. = run-out flow.

03
0 * 21

Table 4. Stressor influence on pressure-containment casing

Operating Stressor influence and remarksa

Part regime

(% BEP flow) Mechanical Hydraulic Tribological

Upper casing 0-50 (1) (2) Near shutoff (1)

head, overpressure may

cause casing

distortion

sufficient to

allow excessive leakage

at casing split

line

50-120 (1) (1) (1)

120-R.O. (1) (I) (1)

Lower casing 0-50 (1) (2) Near shut-off (1)

head, overpressure may

cause casing

distortion

sufficient to

allow excessive

leakage at casing split line

50-120 (1) (1) (1)

120-R.O. C1) (1) Cl)

Suction and All oper- C1) (1) (1)

discharge ating

nozzles regimes

Fasteners All oper- (1) C1) (1)

ating

regimes

aTo indicate a relative measure of importance of each stressor category, an

approximate quantifier is shown as either low (1), medium (2), or high (3).

bR.O. - run-out flow.

_
22

Table 5. Stressor influence on mechanical subsystems

Operating Stressor influence and remarksa

Part regime

(% BEP flow) Mechanical Hydraulic Tribological

Thrust 0-50 (1) (1) (3) Axial hydraulic load

bearing (static and dynamic)

50-120 (1) (1) (1)

120-R.0. (1) (1) (3) Same as 0-SO

Radial 0-50 (1) (1) (3) Dynamic loads from

bearing hydraulic forces

50-120 (1) (1) (1)

120-R.0. (1) (1) (3) Same as 0-50

Thrust 0-50 (1) (1) (3) Radial and axial dynamic

balancer loads can cause rubbing

50-120 (1) (1) (2) Static loads can increase

gradually with use as

wear-ring clearance

opens up via normal wear

120-R.0. (1) (1) (3) Same as 0-50

Seals 0-50 (3) POvertight packing gland (3) PExcessive rubbing possip - for pack- causes excessive wear rate ble due to dynamic hying gland draulic forces

type seals 50-120 (3) PSame as 0-50 (2) mPotential for oxide

m - for mechani- abrasives to cause wear

cal type under all operating

seals regimes

120-R.0. (3) PSame as 0-50 (1) PSame as 0-50

(2) mExcessive vibration (3) PSame as 0-50

Coupling 0-50 (3) Axial rotor dynamic loads (3) Axial rubbing between

originating at the mating teeth on gear

impellers couplings

50-120 (1) (1)

120-R.0. (3) Same as 0-50 (3) Same as 0-50

All (2) Breakdown or loss of

lubricant in gear couplings

(2) Running time on liferated parts, fretting

aTo indicate a relative measure of importance of each stressor

quantifier is shown as either low (1), medium (2), or high (3).

bR.O. - run-out flow.

category, an approximate
23

Table 6. Stressor influence on support

Part Operating Stressor influence and remarks regime Mechanical Hydraulic Tribological

Base frame All regimes (1) None None

Fasteners All regimes (1) None None

In Tables 2-6, to indicate a relative measure of importance of each

stressor category, an approximate quantifier is shown as either low (1),

medium (2), or high (3). To some extent these quantifiers are somewhat

subjective, but are based-on the insight derived from-extensive field -

-experience on power plant pumps. Where warranted on the basis of field

experience with multistage pumps, specific stressors of high relevance

are referenced on Tables 2-6.
24

5. OPERATING EXPERIENCE*

The primary basis for information, insights, conclusions, and recommendations given in this report was the extensive data collection developed by Energy Research and Consultants Corporation from design, troubleshooting, and redesign experience. This collection was augmented through

examination of nuclear power plant operating experience records as discussed below.

This section summarizes AUXFP aging information obtained from various nuclear power plant operating experience records. Several Licensee

Event Report (LER)-based pump failure studies were examined for relevant

pump operating and failure information. While these documents do not

always contain specific pump aging-related failure data, the operating

experience summaries and failure cause data, along with the overall

analysis results, are helpful in understanding the aging degradation of

pumps.

A number of operating experience data bases for nuclear power plants

were examined for this report:

* LER file

* Nuclear Plant Reliability Data System (NPRDS)

o In-Plant Reliability Data System (IPRDS)

Specific information needed for AUXFP failure characterization includes: (1) failure modes, causes, and mechanisms; (2) frequencies of

failures; (3) methods of failure detection - incipient, degraded, catastrophic; (4) maintenance actions; and (5) modifications resulting from

failures.

Each of these items serves to build a failure "signature" that, when

taken in total, can provide a comprehensive assessment of the component

failure.

Unfortunately, no single data base provides all of the information

desired for each failure. But each data base does possess some useful

data elements that can be extracted for pump failure study. Table 7

lists the information available from various sources of operating experience. A summary of pump failure information available from each identified data source is contained in Appendix B.

5.1 Failure Modes and Failure Causes

The primary reported failure causes of AUXFPs were failure of the

bearings, followed by shaft packing and seal failures. The predominant

reported failure mechanism was improper or insufficient lubrication or

cooling. Maintenance errors (that were not readily apparent) and wear

were other major reported pump failure mechanisms. Wear of bearings,

*Work performed by G. A. Murphy, ORNL Nuclear Operations Analysis

Center.
25

Table 7. Summary of AUXFP pump failure information available

from operating experience and plant documents

Operating experience Plant-specific

Data/sourcea data bases documents

LERb NPRDSb IPRDSb SAR SD TS ISI/IST

Pump type and X X X X

description

Manufacturer and X X

model No.

Operating environment X X X

Failure cause 0 X

Failure mechanism 0 0

Discrete failed part 0 X

Maintanance action 0 0 X

Modification to prevent 0 0 X

recurrence

Failure trend data X X

Incipient failure X X X

detection

Specific application X X X

aAcronyms

IPRDS

ISVI/ST

LER

NPRDS

SAR

SD

TS

X

0

In-Plant Reliability Data Study

In-Service Inspection/In-Service Testing Program

Licensee Event Report

Nuclear Plant Reliability Data System

Safety Analysis Report

(Plant) System Description

Technical Specification/Surveillance Test Program

Generally available

Occasionally included in failure report

bExamined in this study.
26

packing, and seals is anticipated with extended use, but the factors

mentioned above can accelerate the wear.

5.2 Frequency of Failure

No frequency of failure data for the AUXFP (as defined in this report) are available from data sources examined in this study. Numerical

values published by NPRDS (and in IEEE-500, 1984) are derived from failure of pump function, which could include numerous components - valves,

instrumentation, controls, etc. - all associated with the pump function.

For example, an NPRDS data request for failures of AUXFPs yielded 70

records. Of the 70 records, only 14 (20%) reported a failure of the

pump; the balance of the records described personnel and maintenance

errors or failure of peripheral equipment that disabled the pump function. Similarly, the LER data base search for failure of AUXFPs yielded

1139 events. Only 53 events (4.6%) involved pump failures; the balance

again described loss of pump function.

5.3 Methods of Detection

In the 14 NPRDs and 53 LER events, nearly one-half of the failures

were detected during surveillance or other tests. About one-third of the

failures occurred while the pump was in operation.

5.4 Maintenance Action

Listed maintenance action for the failure events, consisted mainly

of replacing a worn or broken subcomponent. In less than 10% of the

failures modifications were indicated as required to return the pump to

service.

5.5 Modifications Resulting from Failures

Operating experience data bases do not always contain detailed descriptions of postfailure modifications. Some of the data included the

following:

o impeller modifications were generally made to meet head-flow requirements, reduce stress in the key-way area, and increase clearance between the hub and the labyrinth rings;

• maintenance and repair procedures were continually revised to include proper alignment of pump subcomponents;

* bearing and packing cooling often required modification to prevent

overheating; and

* special training for maintenance personnel complements improved

repair procedures to gain overall pump reliability.
27

6. FAILURE MODES AND FAILURE CAUSES

In Sect. 5, failure modes and causes were introduced within the

context of "operating experience," relying upon the available data-bases

therein referenced. However, that information provides primarily sparse

statistics for recorded failure experiences. Furthermore, the total number of operating hours on AUXFPs, industry-wide, is probably insufficient

to establish from such data potential cause-and-effect relationships

between failure modes and causes. Fortunately, there is a much larger

body of applicable experience stemming from (1) feedwater pumps specifically and (2) turbomachinery and rotating machinery in general.

It is recognized that AUXFPs are subjected to intermittent operation

which makes the experience differ from that of feedwater pumps, but there

is not an available data base with better correspondence. There are many

similarities in design and function between the two to justify use of

feedwater pump experience. Thus, what has become fairly common practice

with similar machinery in identifying failure modes and causes must be

used to provide viable approaches to maximize operational readiness of

AUXFPs.

6.1 Failure Modes

Failure modes must be defined carefully and uniquely because the

word "failure," as herein used, has a much broader meaning than just

total loss of function due to sudden breakage. The following subsections

define failure modes.

6.1.1 Failure to operate

1. Required driver torque is applied to coupling, but rotor does not

rotate.

2. Power interruption by automatic tripping devices, such as overspeed

trip on turbine-driven units. This includes the possibility of a

true trip or one resulting from malfunction of the trip device.

6.1.2 Failure to operate as required

1. Failure to provide required head-capacity pumping characteristic.

2. Critical parameter measurements (e.g., vibration, bearing temperature, delivered flow) outside acceptable ranges. Although the pump

is still able to perform the design function and is, therefore, not

failed, it is either repaired or replaced because it might fail in

the immediate future. Because the pump has to be taken out of service to repair the potential problem, these events are considered

failures.
28

6.1.3 External leakage

A leak of the pump body (i.e., seals or casing) could allow the

contained medium (water) to escape from the boundary. This failure mode

pertains to leakage rates that either indicate impending loss of function

or significantly reduce the flow delivered by the pump.

6.2 Failure Causes

In this subsection, only the causes of the failure modes due to

aging and service wear are identified in general terms; analyses of the

causes are covered in Sect. 6.3 and substantially discussed in Appendix

D. In the following subsections, potential failure causes for each failure mode of 6.1 (summarized in Table 8) are described and also summarized

in Table 9 and deal primarily with proximate causes of failure in each

case.

Table 8. Failure modes

Failure mode Remarks

Failure to operate Rotor does not rotate

Failure to operate Pump fails to provide required headas required capacity pumping characteristics

Critical parameter measurements

outside acceptable ranges

External leakage Escape of contained medium from

component boundary

6.2.1 Causes of failure to operate

Material lodging in rotor clearance spaces can precipitate failure

to operate, as can broken coupling gear teeth. It is quite possible for

dirt or loose metallic fragments (particularly in new systems) to become

lodged in small rotor clearances. In most instances, such occurrences

will "clear" themselves with some minimal amount of resulting wear. However, under extreme conditions, this can lead to galling and a rapid

seizure of the rotor at some location with the stationary parts of the

machine. Then the rotor will become "locked" against rotation. A far

less likely result would be imposition of excessive torque due to drag

that would limit rotation to slower than normal speeds.

Breaking off gear teeth may cause large side loadings on the shafts,

which can produce a locked rotor condition. It may also cause the shaft

to break and thus sever the transmission of torque. In either case, the

pump rotor will cease to rotate.
O. i

Table 9. Suimnaryof failure ,odes and causes

Failure Segments and parts Relative probability

modea involved Failure causes Failure mechanisms of occurrenceb

I Rotating element at close Binding between rotor and Dirt or metallic debris, M

clearance with stationary stationary parts galling and seizing

parts Loss of drive torque Coupling or pump shaft

Pump shaft and coupling breakage H

2 Impellers and nonrotating Seizure or breakage of shaft, Hydraulic forces at high and H rL

internals, seals, wear breakage of impellers or non- v low flows, material fatigue, I

surfaces, bearings, thrust rotating internals, bearing high vibration

' balancer, shaft, coupling or thrust balancer seizure

2 Impellers and nonrotating Deterioration of pumping ca- Hydraulic forces, high vi- H

internals, interstage pacity resulting from rapid bration, cavitation

sealing clearances (wear wear, structural failure of

surface) internals

2 Rotating element, seals, High vibration, high bearing All of the above under H FAtq,-A

bearings, thrust balancer, temperature, abnormal seal failure mode 2

internals leak-off flow, abnormal pump

performance

2 Speed control governor, Overspeed trip Large speed overshoot H or L, depending

overspeed trip device on start-up due to ex- upon system concessively fast speed-up figuration

True overspeed trip

3 Shaft seals, casing Deterioration or breakage High vibration, improper H

of seal components packing adjustment

Leakage through casing Casing distortion and L

static sealing joint "wire drawing'

Erosion, corrosion, or

structural failure of

casing

aFailure mode legend: (1) Failure to operate

(2) Failure to operate as required

(3) External leakage

bL - Low

H - Medium

H - High

'1
/

30

The one automatic tripping action that is built into all turbinedriven AUXFPs is an overspeed trip. In a number of plants and some quite

recently, overspeed tripping at startup has been a problem. When startup

action is initiated, the steam valve in the line to the drive turbine

(typically a 6-in. line) fully opens, and maximum torque is applied to

the pump, resulting in rapid acceleration to and above operating speed;

this occurs in about I s. Before the speed control governor has time to

pressurize its oil system and react correctively, the overspeed trip

value of speed is exceeded and the automatic trip mechanism closes the

steam valve, shutting down the machine (see Appendix C for additional

discussion).

6.2.2 Causes of failure to operate as required

Causes include journal bearing or thrust bearing failure, galling

and seizure of the thrust balancer, sudden shaft breakage, sudden disintegration of an impeller(s) or detachment of a stationary (diffuser)

vane(s). None of these causes are uncommon in HHPS multistage centrifugal pumps. An additional cause is deterioration by cavitation of the

first-stage impeller; others are unacceptable critical parameter readings

and automatic tripping.

Operation of these pumps "far" from the BEP flow induces very strong

unsteady flow conditions within the pump hydraulic passages. The strong

unsteady flow phenomena result in very large dynamic forces (radial and

axial) on pump internals, both stationary and rotating parts. The immediate by-product is high-amplitude vibration that causes rapid wear at

critical clearances in the pump due to severe vibration-induced rubbing.

This leads to rapid increase in stage-to-stage leakage, resulting in measurable reduction in delivered pump capacity. In addition, these large

dynamic fluid forces can break loose large pieces of diffuser vanes, impeller side plates, and impeller vanes. Such rapid deterioration of pump

internals will naturally result in considerable reduction in delivered

capacity of the pump as well as catastrophic structural failure.

Deterioration of the first-stage impeller due to cavitation erosion

also degrades performance. Not having sufficient Net Positive Suction

Head (NPSH) available at the pump inlet over the entire operating range

is sufficient to produce this type of first-stage impeller damage. The

net effect will be a slow deterioration in the delivered capacity of the

pump.

Monitored parameters such as vibration, bearing temperature, seal

leakage, delivered flow, and rotor speed are typically measured and recorded continuously on main power cycle pumps (e.g., main feedwater, condensate, and reactor coolant pumps). At present, such monitoring capability is not typically installed on AUXFPs. One or more of these

parameters can be reliable indicators of an impending pump failure as

well as being correlated with aging and service wear factors and analyses. Obviously, an upward or downward trend away from normal seal leakoff flow is a fairly sure indication that a seal is wearing out or needs

adjustment. An upward trend in vibration level or a sudden increase in

vibration is a sure indication of impending pump malfunction (if proper

action is not taken). Temperatures of journal bearings and especially
31

oil-film thrust bearings are indicative of impending difficulties when

they climb above established limits for reliable operation. Furthermore,

excessive vibration or elevated bearing temperature has serious deteriorating effects on the pumps and thus requires immediate corrective

action, regardless of the specific cause. A significant deterioration of

delivered pump capacity would also be indicative of a potentially serious

problem with the pump that should be immediately investigated and corrected.

The measured parameters by themselves may not divulge the full

nature of a problem. However, they can at least give warning that the

machine must be shut down as soon as possible for further inspection and

possibly disassembly of the pump for inspection of internal parts for the

purpose of determining the probable cause(s) of out-of-normal-range

parameter readings.

Any one of these monitored parameters or a combination thereof could

be incorporated into various automatic tripping schemes. This is common

practice for large steam turbine-generator units, reactor coolant pumps,

main feedwater pumps, large fans, etc., primarily to detect potentially

catastrophic failures before they occur and thus prevent costly damage.

However, the "philosophy" behind using such tripping schemes on AUXFPs

should very definitely be given special examination. It would make no

sense at all to trip an AUXFP during an emergency to save it from lifeshortening wear due to excessive vibration or somewhat high bearing temperatures. The primary purpose of the AUXFP is to protect the reactor

system, not to protect itself.

6.2.3 Causes of external leakage

Three causes are related to this failure mode. The first and most

common is failure of the shaft seals to properly control leakage to within

acceptable limits. This can occur either from an outright breakage of a

sealing component or from a progressively deteriorating seal that may

need replacing or simply a tightening adjustment as in the case of stuffing-box type shaft seals.

The second cause is malfunction of the flat static sealing joint

formed at the split between the upper and lower casing halves. Though

far less common than the first cause (i.e., shaft sealing malfunction),

it is a possibility and does occur on structurally marginal split-casing

multistage centrifugal pumps. When it occurs, it generally indicates

that the overall casing design is marginal from a rigidity or deflection

standpoint, and the joint simply opens up a little at maximum pressure

conditions; that is, at or near shutoff head, the casing design is not

quite stiff enough. The worrisome aspect of this phenomenon is that

"wire drawing" damage to the flat-joint sealing surfaces is a possible

by-product and thus a continually increasing leakage rate may develop,

even at lower operating pressures, once the "wire drawing" erosion progresses.

The third and least likely cause is associated with erosion, corrosion, and structural failure of the pressure envelope. Based on extensive field experience with all types of multistage high-pressure centrifugal pumps, this type of failure should be categorized as one of extremely

low probability.
32

6.3 Failure Cause Analysis

The purpose of this section is to provide more specific discussion

on the various failure-mode causes, based on troubleshooting experience

and insight on AUXFPs, main feedwater and booster pumps (both nuclear and

fossil plants), and multistage centrifugal pumps for other service such

as in petrochemical applications. The failure modes herein identified

for AUXFPs are inherent as well in these other multistage pump applications.

HHPS multistage centrifugal pumps remain an engineering challenge.

Unfortunately, this type of pump has not been viewed by manufacturers and

users as the high-technology device that it really is. The amount of

engineering research in the United States devoted to HHPS pumps over the

last 25 years has been rather insignificant (see Appendix D).

The fact that AUXFPs spend most of their plant life in the standby

mode is at the same time an advantage and a disadvantage. The advantage

is that they are, on the average, less likely to fail from wear than

their "big brothers" - the boiler feedwater or nuclear feedwater pumps

that are operated continuously. On the other hand, whatever inherent

design weaknesses exist in present AUXFP designs, they are less likely to

fully surface on the short-term basis - a disadvantage that gives considerable importance to the main theme of this report.

Through recent field troubleshooting, specifically on AUXFPs and on

quite similar pumps for nonclass applications, some specific factors have

surfaced that impact significantly on the reliability of AUXFPs. These

factors are discussed in the following subsections.

6.3.1 Large dynamic forces and vibrations created by pump

hydraulics

Further review of stressor influence factors (Tables 2-6) and, correspondingly, the summary of failure modes and causes (Table 9) clearly

shows that the large dynamic forces resulting from pump internal flows

is the major failure-mode factor. Within this context, the most severe

operating mode is at low flow (i.e., close to shutoff conditions). The

hydraulic forces generated are large and the resulting vibration levels

are high. Consequently, AUXFPs are installed with a bypass or recirculation line, which has long been standard practice on nuclear main feedwater pumps and boiler feedwater pumps.

A bypass flow line of 25% best-efficiency capacity ensures that no

matter how low the delivered pumping capacity is at part-load operation,

the pump internally never "sees" less than 25% flow. In main feedwater

and boiler feedwater pumps, the bypass flow line is valved so that, at

flow rates above bypass flow, the full capacity of the pump is available

and the energy loss associated with bypass flow is incurred only under

low-flow operation. This same approach of valving the bypass flow line

has been applied in many AUXFP installations. However, an emerging trend

has been to remove this bypass line valve from already installed AUXFPs

or to delete it in newer plants; that is, to eliminate completely the

possibility of having a bypass valve malfunction in the closed position

or to be mistakenly left closed.
33

The required flow of main boiler and nuclear feedwater pump bypass

flow lines is typically a relevant evaluation factor in pump vendors'

sales proposals. This is because the larger the bypass flow required,

the more pumping energy is wasted at part-load operation. In retrospect,

pump manufacturers now indicate that this same evaluation philosophy was

applied in the purchase of AUXFPs, forcing them to compete by providing

small required (typically 10 to 15%) bypass flow figures. However, the

energy consideration just noted is surely not relevant to safety-related

standby pumps such as AUXFPs. Much more attention and engineering investigation is warranted, and each AUXFP design should be thoroughly

tested to determine the true safe minimum flow rate. Because energy

consumption is not a major consideration in AUXFP service, one simply can

size the pump so that its full capacity is sufficient to always ".waste"

the bypass flow and, thus, protect the pump from the abusive operation

and accelerated wear that it could otherwise encounter.

Obviously, the larger the percentage of BEP flow routed through the

bypass line, the better it is for the pump. In well-engineered pumps, 25

to 35% bypass flow is sufficient to avoid the dangerous range of offdesign operating flows (see Appendix D and Fig. D.1). However, for some

pumps any flow below 50% of best-efficiency capacity may cause severe

vibration. This raises a crucial and not yet widely recognized (and thus

unresolved) potential safety-related problem. Apparently, determination

of an appropriate AUXFP bypass flow rate has been done only on a plantby-plant basis and then only in response to piping vibration problems,

not in appreciation of AUXFP abuse considerations.

6.3.2 Shaft breakage

Several factors, identified through field troubleshooting, result in

shaft breakage as an initiating event in a failure (see Appendix D for

additional discussion). Because of the presence of large hydraulic

forces, any design or quality control (QC) shortcoming affecting the

shaft can readily lead to shaft breakage. A common such design flaw is

an undersized shaft diameter, generally motivated to improve pump hydraulics by providing larger impeller inlet flow areas.

In some failures, it was determined that the shaft failed as a consequence of the retaining bolt (which retains the runner of the thrust

balancer) not being square with the mating shaft shoulder when fully

tightened. This results in high bending prestress of the shaft, leading

to fatigue-initiated breakage of the shaft. This is a QC correctable

problem.

Another source of shaft breakage observed in the field is improper

heat treatment of the shaft. This is also a QC correctable problem.

A further source of shaft failure arises from overheating caused by

overtightening of a shaft seal stuffing-box gland. This is a maintenance

correctable problem. Stuffing-box shaft seals may require frequent

tightening adjustment to control shaft seal leakage, even in the standby

mode. Thus, one can envision an inadvertent overtightening of the packing gland in an effort to reduce the required frequency of adjustment.

Although AUXFP manufacturers do not use forgings for the shaft, this more

expensive option would certainly improve shaft durability and thus increase reliability.
34

6.3.3 Impeller and diffuser breakage

If the bypass flow line is undersized (see Sect. 6.3.1), then impellers and diffusers are among those pump internal components that are subjected to severe dynamic loading at low-flow operation. This can create

fatigue failure resulting from cyclic stresses.

Even if the bypass flow line is adequately sized, a too small radial

gap between impeller-vane outside diameter and diffuser-vane inside diameter frequently results in impeller and diffuser breakage. This breakage

results from high-amplitude shock loads at vane-passing frequency (see

Fig. D.5). The offending factor here is poor hydraulic design. This can

be fixed in the field by experienced personnel and entails machining and

recontouring the diffuser vanes to a larger radial clearance with the

impeller. This fix also generally makes the pump run much quieter, which

by itself generally is indicative of improved internal hydraulic configuration.

Cavitation damage to the first-stage (i.e., suction-stage) impeller

can result in structural failure of the impeller. Insufficient NPSH is

a common problem of feedwater systems in general and stems from a wide

range of mostly optimistic opinions on what constitutes adequate NPSH.

A significantly reduced potential for impeller and diffuser breakage

could be achieved if impellers were precision cast rather than using the

lower quality sand castings as is presently done by all AUXFP vendors.

6.3.4 Thrust bearing and thrust balancer failures

Even for standardized pump configurations, there exist many possible

variations that a vendor can tailor to meet a given application specification. For all the available designs and variations thereof, there is

simply not adequate engineering test data on hydraulic axial thrust loads

over the full flow range of each specific design. Thrust loads change

considerably on a given pump as a function of flow, speed, and state of

wear at wear-ring stage-to-stage leakage control clearances. Essentially,

the axial thrust loads are not accurately known over the full ranges that

feedwater pumps are commonly required to operate. Consequently, thrust

balancers and thrust bearing failures in feedwater pumps are not uncommon

because the inaccurate definition of axial thrust loads (both static and

dynamic) often results in undersized thrust balancer and thrust bearing

designs.

Excessive operating temperatures in thrust bearings are common symptoms indicative of this overall problem. Furthermore, on thrust balancers, improper heat treatment and material property control result in

galling and seizing events, especially when combined with undersized

geometries (see Appendix D for additional discussion).

6.3.5 Coupling breakage

Coupling breakage is not uncommon. The combination of excessively

low-flow operation of the pump and an undersized thrust balancer can

result in excessive axial dynamic loads on the coupling. Additionally,

when the radial gap between impeller and diffuser vanes is too small,
35

excessive coupling vibration at vane-passing frequency can readily

destroy the coupling (gear teeth).

Breakage of seal injection piping due to excessive vane-passing

frequency vibration also can occur when the impeller-to-diffuser radial

gap is too small.

6.3.6 Seizure

Seizure is not uncommon in feedwater pumps and has been traced to a

variety of causes including (1) excessive rotor vibration (both lateral

and axial); (2) inadequate clearance at thrust balancer; (3) seizureprone material hardness and geometry combination at impeller wear rings,

thrust balancer, or any other close-running clearance; (4) overall

hydraulic mismatch between impeller and diffuser; and (5) journal bearing

failure resulting from excessive rotor vibration, poor babbitt bonding to

bearing shell, dirt in the lubrication oil system, or combinations of

these.
36

7. RECOMMENDED MAINTENANCE, SURVEILLANCE, AND

MONITORING PRACTICES

AUXFPs spend most of their plant life in a standby mode. Consequently, present maintenance, surveillance, and monitoring practices are

not elaborate. Pump manufacturers are mainly concerned with other applications involving continuous long-duration running where certain components clearly do wear out and where scheduled major overhaul and refurbishment is common practice. Regarding AUXFPs, the manufacturers make no

definitive recommendations.

At present, most AUXFP installations contain no monitoring devices

except for flow and head measurement, in contrast to continuously running

machinery. Present AUXFP surveillance practice consists primarily of

periodically* starting each pump for a short-duration test to verify its

operational readiness.

7.1 Regular Maintenance

The only present regular maintenance activity pertains to shaft

sealing components. In the normal standby mode, the shaft seals are

still subjected to upstream system pressure, that is, from a condensate

or storage tank. This pressure is typically 2 to 4 atm and thus maintains the shaft seals in a functional sealing mode, even though the shaft

is not rotating.

Stuffing boxes as well as mechanical seals are in common use in

AUXFPs. With stuffing-box shaft seals, the primary maintenance activity

is to keep the packing gland properly adjusted. Based primarily on an

observation of excessive leakage, a tightening adjustment is made. However, stuffing boxes must be allowed to maintain some continuous leakage.

Overtightening of stuffing-box glands to eliminate leakage completely

will likely result (and has on AUXFPs) in severe operating problems when

the AUXFP is put into a pumping mode, that is, at normal rotational

speed. Overtightened stuffing-box glands result in excessive temperature

rise both in the packing and on the shaft, increasing considerably packing wear and the probability of shaft breakage. In many plants, AUXFP

stuffing-box packing glands are now maintained somewhat looser than would

be typical of non-AUXFP applications. Although this results in somewhat

higher continuous leak-off flow, it is a prudent practice, since it gives

a further safety margin against such heating problems and related potential failures during any pumping mode.

With mechanical shaft seals, the leak-off flow in the standby mode

should be zero, if the seal is functioning properly. That is, at zero

shaft rotation, a mechanical seal essentially closes the flow path for

shaft-end leakage. If a mechanical seal is observed to be leaking in the

standby mode, this is a strong indication that something is wrong with

*Plant Technical Specifications require operational readiness testing at least once every 3 months. However, in many plants, this testing

is performed as frequently as once a month.
I

37

the seal, and thus it should be fixed or replaced as required to correct

the problem.

In addition to these seal maintenance items, lube-oil sight glasses

on the AUXFP should be regularly inspected to check oil level. Most

other maintenance items would require some amount of disassembly to perform and thus would be objectionable, unless there were strong indications of pump malfunction, for example, as indicated by excessive vibration.

7.2 Surveillance

7.2.1 Current surveillance practice and limitations

Unfortunately, AUXFPs are rarely installed with the type of monitoring devices typical of equipment provided for continuous operation

service. Installed rotor vibration probes and bearing inside temperature

transducers (e.g., babbitt-contacting thermocouples) have generally not

been required by AUXFP purchaser specifications. Only in a very few

plants, where recurring AUXFP operating problems have otherwise been

detected, has such monitoring equipment been retrofitted. Thus, in most

plants, comprehensive surveillance of AUXFPs during pumping mode service

is not readily done.

As noted in Sect. 4.3.2, some newer plants use a nonclass pump for

startup, shutdown, and any other nonemergency service. This "saves" the

N-class AUXFPs for strictly emergency service with one important exception. The Plant Technical Specification requires that, nominally every

three months, each N-class pump is put into a pumping mode for a short

period of time to demonstrate its continued operational readiness. During this regular test, the loads imposed on the AUXFP depend on the bypass flow line configuration. In most plants, the AUXFP bypass line is a

single line sized to pass 5 to 15%,of BEP flow. Thus, this type of testing could be the main contributor to wear and aging of various AUXFP components. In some newer plants (e.g., Palo Verde) an additional full-flow

bypass line is provided to allow testing of the AUXFPs over the full

operating flow range, even when the plant is operating in a normal generating mode. However, in most plants, this test provides neither the

proper operating range of flow nor sufficient running time to comprehensively trend and assess an AUXFP's "vital life signs," even if a full

complement of state-of-the-art monitoring devices were installed.

7.2.2 Interim recommendations for surveillance

and monitoring practices

Based upon the described impediments to comprehensive surveillance

and performance trending, some obvious interim recommendations can be

manifestly considered.

First, present auxiliary feedwater systems need to be studied for

modifications that would permit ease of testing the AUXFPs over the entire range of flow, independent of the main feedwater system. One possibility is to permanently install bypass flow lines to allow testing the
38

AUXFPs at operating conditions that fully simulate the various emergency

pumping requirements that the AUXFPs have to meet. Built-in full-flow

testing capability could be engineered to avoid adversely affecting the

reliability of the auxiliary feedwater system in normal and standby modes.

This has been demonstrated at the Palo Verde plant.

Second, a full complement of permanently installed monitoring devices

should be considered on every AUXFP. Major candidate parameters should

include at least (1) rotor orbital motion; (2) oil-film bearing temperatures; (3) head, flow, and speed values; (4) rotor axial position; and

(5) metal fragment and/or sound emission for early detection of incipient

failure of rolling contact bearings.

Third, a standardized sequence of full-service tests need to be

developed and performed at appropriate regular intervals. The full complement of parameters monitored could be appropriately analyzed, trends

established, and results compared over the accrued history of the pump.

This data base could be used to indicate need for preventive maintenance,

overhaul, and replacement of worn parts. Such monitoring and trending

systems for rotating machines are now commercially available at a small

fraction of the cost of an N-class AUXFP. Section 8.2 provides a more

detailed discussion of parameter monitoring.

7.2.3 Interim recommendations for detailed inspection program

In addition to the comprehensive testing, surveillance, and monitoring approach recommended in the previous section, a detailed disassembly

and inspection of AUXFP internals could be carried out whenever the reactor is in a scheduled shutdown for a number of weeks, such as at refueling

time.

During this disassembly, a number of procedures and reliability

enhancing actions could be taken: (1) replace rolling contact bearings;

(2) replace renewable components of mechanical shaft seals; (3) replace

any wear rings having clearance-doubling wear; (4) inspect impellers,

diffuser vanes, and other hydraulic-passage internals for damage and

replace any damaged or significantly worn parts based upon conservative

criteria; (5) inspect journals, bearing surfaces, and thrust balancer

components, replacing any distressed components; (6) inspect main sealing

joint surfaces at casing joint to detect any sign of leakage; (7) check

shaft for run-out (can readily occur through creeping) and correct as

required by straightening or replacement of shaft; and (8) inspect and

replace any fastener that shows signs of distress. Section 8.1 provides

a more detailed discussion of the overall topic of detailed inspection.
39

8. AGING AND SERVICE WEAR MONITORING

As described in Sect. 7.2, nearly all AUXFPs presently in service

are installed with little or none of the specific parameter monitoring

devices that are commonly integrated into continuously running main power

cycle equipment. Furthermore, most auxiliary feedwater systems are not

presently configured to facilitate regular periodic testing of AUXFPs

over the complete range of flow rates that would simulate the various

emergency scenarios for which AUXFPs are installed. Consequently, with

present typical installations, aging and service wear monitoring would

have to be based primarily upon a detailed and thorough inspection of the

pump in the disassembled state. Unless specific operating problems arose

with a particular AUXFP, such a detailed inspection would be practical

only during a scheduled plant shutdown, such as at refueling time.

Shaft seal leak-off flow is the one parameter that can be and typically is readily monitored, both in the standby ready mode as well as

running modes. In the following sections, monitoring specifics that are

advisable and can be readily performed with present typical configurations

are described. In addition, monitoring specifics that could further

assist in aging and wear analyses, utilizing state-of-the-art monitoring

packages, are described for consideration as upgrading options.

8.1 Monitoring of Present Configurations

Here the discussion is aimed at existing configurations that typically have available neither a full complement of state-of-the-art continuous monitoring devices nor are set up to routinely run full-flow

range AUXFP tests.

8.1.1 Rotor binding check

The following is a simple procedure that is probably not commonly

practiced in existing plants although it is highly advisable. Just prior

to starting the pump, the rotor should be hand rotated, grabbing it at

the coupling. Generally, a pair of nonslip gloves must be worn to accomplish this. This procedure could also be performed at regular intervals

even if the pump is not about to be started. For safety reasons, the

pump driver must be "tagged out." Naturally, this procedure should be

disregarded in the event of a true emergency startup of the pump.

If it is not possible to hand turn the rotor, then a suitable wrench

or wrenchlike device should be used to attempt a wrench-assisted hand

turning of the rotor. If this action fails to turn the rotor, then the

cause of the binding can be identified and appropriate corrective action

taken.

8.1.2 Shaft seal leak-off flow

8.1.2.1 Mechanical seals. Properly functioning mechanical seals

should not exhibit leak-off flow large enough to be seen or measured. If
40

leakage is clearly visible, then the seal is potentially in need of parts

replacement and at least the reason for the leakage should be determined

and appropriate corrective action taken.

8.1.2.2 Packing gland (stuffing box) seals. This type of seal

should always leak a small amount to ensure that the packing gland has

not been overtightened. Excessive leakage, which cannot be corrected

through normal gland adjustment, could be indicative of worn or damaged

components in the seal subassembly. The defective or worn part(s) should

be replaced.

8.1.3 Disassembly and detailed inspection

The following subsections pertain to those aging and service wear

monitoring actions that require disassembly of the pump and that could be

performed during a scheduled plant outage, that is, at refueling time.

To disassemble this type of split casing pump, the following steps are

taken: (1) the coupling is disengaged, (2) the journal bearing caps are

removed, (3) the shaft seals are removed, (4) the casing stud bolts are

removed and the upper casing half lifted off of the bottom casing half,

and (5) the rotor is lifted out of the lower casing half. At this point,

each wear ring will be loosely surrounding its respective impeller and

can be moved axially to hang on the adjacent shaft section so that the

inside diameter of each wear ring can be measured.

Upon disassembly of the pump as outlined above, the following measurements and visual inspections can be made.

8.1.3.1 Journal bearings, journals, and thrust bearing. Visually

inspect all bearing surfaces for evidence of (1) bearing metal (babbitt)

fatigue cracks, (2) loss of babbitt-substrate bonding, (3) babbitt wiping

or smearing, and (4) babbitt gouging that would occur from entrained

foreign matter in the oil. If any of the above deteriorations are in

evidence, then the bearing should be replaced, or repaired to essentially

as-new condition. The journals should also be inspected for signs of

distress or wear. Specifically, measurable or visual evidence of surface

deterioration, such as deep scratches, heat checking, or scoring should

require refurbishment to return journal surfaces to their as-new condition. If the journal surface distress is quite extreme, then the journals may have to be chrome-plated to a sufficient thickness so that

resurfacing of the journals does not reduce the journal diameter to less

than its original dimensional tolerance zone. If the journals and bearings are repaired during the same disassembly, the original bearing

clearance could be preserved by remachining a rebabbitted bearing to a

slightly reduced bore diameter, sufficient to offset any slight reduction

in journal diameter resulting from redressing its surface.

The AUXFP thrust bearing is usually a rolling contact type, such as

a double-row preloaded angular contact (duplex) ball bearing arrangement.

This bearing can be inspected for signs of wear or damage although subsurface fatigue cracks are not visible. However, such bearings are relatively inexpensive, and thus automatic replacement is strongly advised.

8.1.3.2 Shaft seals and mating rotor sleeves. For mechanical seals,

the mating sealing surfaces (rotating and nonrotating parts) should be

closely inspected visually, with high magnification if necessary, to
41

determine the presence of any measurable signs of wear. Any evidence of

wear is sufficient cause to replace the worn seal component.

For stuffing-box seals, the rotor sleeve should be inspected for

signs of wear and/or surface finish deterioration. It is not uncommon

for the rotor sleeve to have deep grooves worn in from continual running

with an excessively tightened packing gland. The shaft sleeve should be

replaced in any case where this inspection uncovers measurable wear or

surface deterioration. The packing material, though adjustable in service to compensate for wear, will eventually need replacing. It is inexpensive and thus should be automatically replaced at any such disassemblyfor-wear inspection, using the same rationale previously applied to rolling contact bearings.

8.1.3.3 Coupling. The disassembled coupling should be inspected

for any deterioration of the gear teeth. If the coupling shows measurable wear or tooth surface deterioration, then the coupling should be

replaced. Here it is highly recommended that when coupling replacement

is deemed appropriate, it be replaced by the dry diaphragm type of flexible coupling that requires no maintenance and no lubrication and has a

very successful retrofit record on all types of feedwater pumps.

8.1.3.4 Fasteners. Each fastener that is disengaged during the

disassembly process should be closely examined using approved procedures.

8.1.3.5 Wear ring clearances. Each wear ring inside'diameter and

mating impeller-eye outside diameter should be accurately measured using

a micrometer. If the clearance has worn as large as twice the as-new

clearance, the wear ring should be replaced to return the clearance to

its as-new nominal value. For example, if the as-new clearance is nominally 0.015 in. diametral (typical), then a worn clearance of 0.030 in.

or larger is justification for wear ring replacement.

When wear ring clearances are allowed to open excessively (as commonly occurs in poorly maintained pumps) a number of undesirable consequences will occur as the wear worsens:

* First, the delivered head-capacity performance of the pump will

deteriorate because of the increased stage-to-stage leakage.

* Second, the axial thrust load on the thrust balancer and thrust bearing could increase, further taxing all thrust carrying components.

* Third, the rotor-dynamic performance of the pump will deteriorate

significantly because proper (as-new) wear ring clearances assist in

controlling residual rotor vibration levels arising from all sources

of excitation (e.g., unbalance, hydraulic forces, etc.).

In cases of excessive clearances, the deterioration of pump hydraulic performance will significantly exceed that of just the leakage loss

subtracted from the head-capacity characteristic, because the leakage

itself becomes a large flow component and disrupts the main through-flow

as it approaches and enters the suction side of the impeller. Thus,

proper periodic monitoring of wear ring interstage sealing clearance is

vitally important so that wear rings are replaced before clearances, become excessive.

8.1.3.6 Thrust balancer. When a thrust balancing disk is employed,

two areas should be closely inspected. First, the vertical face of the

balancing disk (rotating part) and the mating stationary surface should
42

be closely inspected (visually) to detect signs of heavy rubbing (e.g.,

scoring marks or wear). If any such distress of these surfaces is detected, the component(s) should be replaced or, if possible, repaired by

remachining the surfaces. Furthermore, for thrust balancing disk configurations, the radial clearance of the cylindrical pressure breakdown serrated bushing should be measured using a micrometer on the mating diametral surfaces. The criteria here are exactly the same as those previously

detailed for wear ring clearances (see Sect. 8.1.3.4). Namely, if the

clearance has worn to twice the as-new clearance or larger, parts should

be replaced to return this clearance to the as-new condition.

When a thrust balancing drum is employed, the inspection procedure

and criteria for replacement are the same as for the cylindrical pressure

breakdown clearance section of the balancing disk assembly. More discussion and technical insight on thrust balancers are given in Appendix D.

8.1.3.7 Suction (first stage) impeller. The first-stage impeller

should be inspected to detect any evidence of erosion (i.e., pitting)

indicative of cavitation-caused damage. If detected, the impeller should

be replaced. Furthermore, if such cavitation damage is detected, the

design parameters of the suction side of the auxiliary feedwater system

should be reviewed and analyzed to determine why the pump NPSH is insufficient to suppress cavitation. In a properly designed system, adequate

NPSH should be available to inhibit cavitation damage. The results of

such an analysis could potentially suggest a redesign of the system to

eliminate cavitation damage (see Appendix D).

8.1.3.8 Impellers and diffusers. A complete inspection of all

impeller and diffuser vanes and impeller side plates should be made to

detect any structural damage to these components. The unsteady flow

dynamic forces that are produced, particularly under low-flow operation

and especially on designs employing relatively small impeller-to-diffuser

vane tip clearances, frequently break off pieces of impellers and diffuser vanes. Obviously, any such structurally damaged components must be

replaced. Furthermore, modifications to pump hydraulic internals that

will eliminate such structural damage in the future should be considered.

8.1.3.9 Shaft. The shaft should be closely inspected for any

structural damage or cracks. If such are detected, the shaft must be

replaced. Also, the shaft should be inspected to determine if any outof-tolerance nonconcentricities have developed, such as may occur through

creep bending of the shaft. In this case, the shaft may be salvaged

through a proper straightening process in a qualified pump repair shop.

A bent shaft can produce obvious undesirable results, including accelerated wear at wear ring clearances, seals, thrust balancer, bearings, and

excessive rotor vibration.

All shaft sleeves, spacers, and fasteners should be closely inspected for wear or structural damage and replaced as needed.

8.1.3.10 Casing split surfaces. The flat casing split surfaces

should be inspected to detect any sign of casing leakage, for example, as

would be evidenced by "wire drawing" damage. Any such surface damage can

be repaired, for example, through hand-lapping the damaged portions of

the surface.
43

8.2 Continuous Parameter Monitoring

The aging and service wear monitoring methods here described are

proposed as additional to (not in place of) those that can already be

performed. The methods addressed in this section may not be applicable

to most present AUXFP installations because of the absence of required

monitoring devices and piping configurations that facilitate regular

periodic testing of AUXFPs over the complete flow and head ranges to

simulate the emergency conditions. In spite of these limitations, the

monitoring methods described in the following sections are proposed to

further assist in aging and wear analyses, utilizing state-of-the-art

monitoring packages. These methods are thus recommended for consideration as upgrading options on existing plants as well as original equipment for future new plants.

8.2.1 Rotor vibration monitoring

Rotor vibration is a primary vital sign of rotating machinery just

as body temperature is to animals and humans.

Firstly, excessive rotor vibration accelerates machinery deterioration. The major consequences of excessive vibration are increased potential of fatigue failures, shortened bearing life, breakage of shafts and

shaft seals, and accelerated wear at internal leakage control clearances.

Thus, early detection of excessive rotor vibration is an essential factor

in the timely identification and correction of such problems.

Furthermore, real time analysis of rotor vibration signals can

frequently allow one to detect and isolate the progression of various

deterioration phenomena. For example, Fast Fourier Transforms (FFT) or

spectrum analysis, which transforms time-base signals into the frequency

domain, can reveal sources of deterioration through a study that includes

establishing trends for dominant frequency components. A slow but steadily increasing vibration component with a frequency equal to the rotational speed multiplied by the number of coupling gear teeth can reveal

that the coupling is approaching the end of its usable life. Strong subsynchronous (below rotational speed) vibration frequency components are

generally indicative of either rotor-bearing instability or significant

deterioration of pump hydraulic internals (primarily, wear ring clearances are badly worn). Strong synchronous vibration can be indicative of

excessive unbalance, bent shaft, or both.

The generally accepted practice for rotating machinery is to have

two noncontacting proximity probes mounted near each journal bearing at a

90° angular separation from each other. This allows one to observe the

rotor (i.e., journal) orbit of vibration when the two time-base signals

are spatially superimposed using an oscilloscope. The journal orbits can

clearly reveal the detailed nature (on a continuous basis) of the rotor

vibration. From this information, and a pre-existing data base for the

AUXFPs, correlations and analyses for wear and aging factors can be made

to assess the "state of deterioration" on a continuous basis.
44

8.2.2 Bearing temperatures and noise

8.2.2.1 Oil-film bearings. Bearing liner (typically babbitt) temperatures are commonly measured continuously on major power plant machines.

As in the case of rotor vibration, excessive bearing temperatures by themselves accelerate deterioration, at least of the bearing. Furthermore,

excessive bearing temperatures are generally a consequence of other problems such as lubricant starvation, general deterioration of the bearing,

or excessive bearing loads as may occur in further consequence of deterioration of pump hydraulic internals, pump-driver alignment, and looseness

of various centering fits or fasteners. A continuous time record of bearing temperature, combined with a pre-existing data base for the AUXFPs,

can be used singly or in concert with other monitored information to

develop a running evaluation of wear and aging deterioration of AUXFPs.

8.2.2.2 Rolling contact bearings. Properly measured acoustic emissions of rolling contact bearings can be an effective approach to detect

early signs of bearing deterioration as well as incipient bearing failure. It is unlikely that rolling contact bearings would ever be employed

in continuously running high-pressure multistage centrifugal pumps.

Their frequent use as an axial thrust bearing in AUXFPs is indicative

that these pumps are not intended for long continuous running. In other

applications, such pumps are typically configured with double-acting

pivoted-pad oil-film thrust bearing. Even for AUXFPs, the use of rolling

contact bearings warrants engineering review to assess whether ball bearings are capable of performing reliably in this application.

8.2.3 Rotor axial position

Boiler feedwater pumps are typically supplied with a mechanical

readout device (located near the thrust bearing) that indicates the axial

position of the rotor relative to the stationary structure of the pump.

Significant axial shifting of the rotor is generally indicative of wear

of the thrust balancing disk. If a thrust balancing drum is employed,

then significant axial shifting is generally indicative of thrust bearing

wear. In either case, this type of wear is not consistent with reliable

operation and, when detected, should immediately be investigated and

fixed.

8.2.4 Pump head-capacity curve

A significant deterioration of the pump head-capacity curve is a

certain indication that something in the pump internals has significantly

deteriorated. Badly worn internal clearances or structural damage to

pump internals has probably taken place. When combined with a pre-existing data base for the AUXFPs, periodic rerunning of the head-capacity

curve could become an effective tool in assessing aging and service wear

factors, singularly or in concert with other monitored information.
45

8.3 Summary of Aging and Service Wear Factors

A tabulation of AUXFP failure modes is given in Table 8. As more

fully described in Sect. 6, the three failure modes are: (1) failure to

operate, (2) failure to operate as required, and (3) external leakage.

A more definitive delineation of failure modes and corresponding

failure causes is summarized in Table 9. Each failure mode is associated

with segments and parts involved, causes, failure mechanisms, and relative probability of occurrence.

The relationship of aging and service wear factors and failure

causes is summarized in Table 10. The lead entry of this table is pump

segment with failure causes and failure modes keyed to it.

Table 10. Pump failure causes related to aging and service wear

Failure modesa

Pump segment Failure cause

1 2 3

Rotating elements Binding between rotor and x

stationary parts

Shaft breakage x

Impeller wear, breakage x

Thrust runner wear, breakage x x

Fastener loosening, breakage x x

Nonrotating Structural damage to stationary x

internals vanes (diffuser or volute)

Wear-surface wear, erosion, x x

corrosion, seizing

Fastener loosening, breakage x

Pressure Leak at casing split x

containment Suction nozzle leak, breakage x

casing Discharge nozzle leak, breakage x

Fastener loosening, breakage x

Mechanical Bearing wear, corrosion, breakage x x

subsystems Shaft seal deterioration, breakage x

Thrust balancer galling, seizing x x

Coupling wear, breakage x x

Fastener loosening, breakage x

Support Base frame breakage x

Fastener loosening, breakage x

aFailure mode designation:

1 - Failure to operate

2 - Failure to operate as required

3 - External leakage
46

Currently used methods for AUXFP failure detections are listed in

Table 11. Essentially, this simple table relates failure modes and means

of detection.

Table 11. Methods currently used to detect

AUXFP failure modes

Failure mode Means of identification

Failure to operate Visual observation

Pressure readings at flow-measuring orifice

Pump driver current and voltage or steam

flow measurement

Failure to operate Pressure readings at flow-measuring orifice

as required Measurements of key parameters, that is,

vibration, temperature, rotational speed,

and flow

External leakage Visual

Table 12 summarizes the methods for differentiating between the different failure causes. Here, the failure mode is the main entry parameter with pump segments, failure causes, and differentiation methods keyed

to it.

Table 13 lists the various measurable parameters. Here again, the

main entry parameter is failure mode, with pump segments, failure causes,

and measurable parameters keyed to it.

Table 14 is a major summary that condenses all key information on

AUXFP part failure assessments, including parts, materials, significant

stressors, failure causes, and measurable parameters.
47

Table 12. Methods for differentiating between failure causes

Failure mode Pump segment Failure causes Methods for differentiation

Failure to operate Rotating elements Binding between

rotor and

stationary parts

Shaft breakage

Thrust runner

wear, breakage

Fastener loosening,

breakage

Bearing wear,

breakage

Mechanical

subsystems

Coupling wear,

breakage

Thrust balancer

galling, seizing

Nonrotating

Internals

Wear-surface

seizing

Wear-surface wear

erosion, corrosion

Fastener loosening,

breakage

Impeller wear,

breakage

Pump rotor cannot be manually

rotated

Visual examinationa

Visual examinationa

Visual examination, inspection

during maintenance

Visual examinationa clearance

measurement, rotor axial position measurement

Visual examinationa

Pump rotor cannot be manually

rotated, clearance measurement,

rotor axial position measurement

Pump rotor cannot be manually

rotated, visual examination

(requires disassembly)

Visual examination, clearance

measurement, rotor axial position measurement

Visual examination, inspection

during maintenance

Visual examination (requires

disassembly), large unbalance

rotor vibration, delivered flow

measurement

Transmitted torque measurement,

rotational speed measurement,

visual examination (requires

disassembly)

Visual examination, inspection

during maintenance

Visual examination (requires

disassembly), vibration monitoring, delivered flow measurement

Transmitted torque measurement,

vibration monitoring

Visual examinationa delivered

flow decrease, clearance measurement, vibration monitoring

Visual examination, inspection

during maintenance

Failure to operate

as required

Rotating elements

Nonrotating

Internals

Thrust runner wear

Fastener loosening,

breakage

Structural damage to

stationary vanes

(diffuser or

volute)

Wear-surface

binding

Wear-surface wear,

erosion, corrosion

Fastener loosening,

breakage
48

Table 12 (continued)

Failure mode Pump segment Failure causes Methods for differentiation

Failure to operate Mechanical Bearing wear, Bearing temperature measureas required subsystems corrosion, breakage ment, vibration monitoring,

(continued) transmitted torque measurement,

visual examinationsd rotor

axial position measurement

Coupling wear, Vibration monitoring, transbreakage mitted torque measurement,

visual examination

Thrust balancer Transmitted torque measurement,

galling, seizing rotor axial position measurement, rotational speed measurement, visual examinationa

Fastener loosening, Visual examination, inspection

breakage during maintenance

Support Base frame breakage Bearing temperature measureFastener loosening, ment, transmitted torque

breakage measurement, acoustic monitoring, vibration monitoring,

visual examination, inspection

during maintenance

External leakage Pressure contain- Leakage at casing Visual examination

ment casing split

Fastener loosening, Visual examination, inspection

breakage during maintenance

Mechanical Shaft seal Visual examination, vibration

subsystems deterioration, monitoring, local shaft tembreakage perature measurement

aMay require disassembly.

I
-

49

Table 13. Measurable parameters

Failure mode Pump segment Failure causes Measurable parameters

Failure to operate Rotating elements

Mechanical

subsystems

Nonrotating

internals

Failure to operate

as required

Rotating elements

Nonrotating

internals

Mechanical

subsystems

Support

Binding between

rotor and

stationary parts

Shaft breakage

Thrust runner

wear, breakage

Fastener loosening,

breakage

Bearing wear,

breakage

Coupling wear,

breakage

Thrust balancer

galling, seizing

Wear-surface seizing

Wear-surface wear,

erosion, corrosion

Fastener loosening,

breakage

Impeller wear,

breakage

Thrust runner

wear

Fastener loosening,

breakage

Structural damage

to stationary vanes

(diffuser or volute)

Wear-surface bindIng

Wear-surface wear,

erosion, corrosion,

Fastener loosening

breakage

Bearing wear,

corrosion, breakage

Coupling wear,

breakage

Thrust balancer

galling, seizing

Fastener loosening

breakage

Base frame breakage

Fastener loosening,

breakage

Leakage at casing

split

Fastener loosening,

breakage

Shaft seal

deterioration,

breakage

Appearance (heat checking,

welding)

Appearance

Appearance, clearance

Appearance, bolt torque

Appearance, clearance, rotor

axial position

Appearance

Appearance, clearance, rotor

axial position

Appearance, clearance

Appearance, clearance, rotor

axial position

Appearance, bolt torque

Appearance, rotor vibration,

delivered flow

Transmitted torque, rotational

speed, appearance

Appearance, bolt torque

Delivered flow,

appearance, vibration

Transmitted torque,

vibration

Appearance, delivered

flow, clearance, vibration

Appearance, bolt torque

Temperature, rotor vibration,

transmitted torque, appearance,

rotor axial position

Vibration, transmitted torque,

appearance

Transmitted torque, rotor axial

position, rotational speed,

appearance

Appearance, bolt torque

Bearing temperature, transmitted torque, noise, vibration, appearance, bolt torque

Leakage rate

Appearance, torque

Seal leakage rate,

rotor vibration,

local shaft temperature

External leakage Pressure containment casing

Mechanical

subsystems
50

Table 14. Summary of important AUIXFP part failure assessment

Part Materials Significnt strfirorh/ Measurable parameters

Shaft and 400-series B.S. (a) Mechanical/breakage

fasteners

Impellers CrNi alloy steels,

17-4Ph

(b) Hydraulic/breakage and wear

(c) Tribological/wear and

seizing

(a) Mechanical/breakage

(b) Hydraulic/breakage and wear (c) Tribological/wear-surface

wear

(a) Mechanical/breakage, seizing

(b) Hydraulic/breakage, seizing

(c) Tribological/rubbing, lubricant dirt and breakdown

(a) Hydraulic/breakage

Thrust

runners

Stationary

vanes (diffuser or

volute)

Wear rings

Thrust

balancers

Thrust

bearings

400-series B.S.

400-series S.S.

400-series S.S.

400-series S.S.

Rolling contact

elements

(Specialty steels)

(a)

(b)

(c)

(a)

(b)

(c)

(a)

(b)

(c)

Mechanical/seizing

Hydraulic/seizing

Tribological/wear-surface

wear

Mechanical/breakage, seizing

Hydraulic/breakage, seizing

Tribological/wear

Mechanical/breakage

Hydraulic/breakage

Tribological/wear

(a), (b), and (c) vibration,

bearing temperature, appearance, transmitted torque

(a) and (b) rotor unbalance

vibration, appearance,

delivered flow; (c) rotor

vibration, appearance

(a), (b), and (c) transmitted

torque, rotational speed,

rotor axial position

(a) delivered flow, appearance, vibration

(a), (b), and (c) vibration,

transmitted torque, delivered flow, appearance,

clearance

(a), (b), and (c) rotor

axial position, transmitted

torque, rotational speed.

appearance

(a), (b), and (c) rotor

axial position, transmitted

torque, rotational speed,

appearance, clearance

(a), (b), and (c) rotor

vibration, bearing temperature, transmitted torque,

appearance

(a), (b), and (c) seal leakage rate, rotor vibration,

local shaft, temperature

(a), (b), and (c) rotor

vibration, transmitted

torque, appearance

Radial

bearings

Shaft

seals

Bearing white

metal (typically

tin-base babbitt)

Stuffing-box or

mechanical type

(a)

(b)

(c)

(a)

(b)

(c)

(a)

(b)

(c)

Mechanical/breakage, seizing

Hydraulic/breakage, seizing

Tribological/seizing, wear

Mechanical/breakage, wear

Hydraulic/breakage, wear

Tribological/wear

Mechanical/breakage, wear

Hydraulic/breakage, wear Tribological/breakage wear

Coupling Gear type

(usually)
51

9. SUMMARY AND RECOMMENDATIONS

9.1 Summary

AUXFPs are critically-important safety-related equipment.; Several

design, operation, maintenance, and condition-monitoring factors of these

pumps warrant investigation. For example, as detailed in Sect. 2.4,

tighter specification of certain materials of construction and fabrication methods could potentially provide marked improvements in AUXFP durability and thus higher reliability.

As described in Sect. 6.3, present HHPS multistage centrifugal pumps

suffer from years of dormant engineering research and development efforts

in the United States. For example, bearing and thrust balancer designs

may in some cases be based on marginally reliable definitions of the

static and dynamic rotor loads produced by the complex fluid flow phenomena that occur within such pumps. Engineering test data of specific pump

configurations over the full range of operating conditions and possible

levels of wear are not always available to designers because such testing

is not always performed on every-specific configuration. This has often

resulted in undersized and marginal load carrying components in similar

pumps used in other (non-AUXFP) applications.

As described in Sect. 6.3.1, the question of a prudent bypass flow

capacity (to minimize dynamic loads and vibration), which varies from

design to design, needs to be further studied and clarified.

AUXFPs are now typically installed with virtually no monitoring

devices for establishing data trends nor secondary bypass flow loops that

would facilitate periodic regular testing over the full-flow ranges necessary to simulate the various emergency pumping scenarios. Also, there

is presently no standardized requirement for scheduled disassembly and

intensive pump internal inspection (as proposed in Sect. 7.2.3 and

detailed in Sect. 8.1).

In response to these important considerations, the following observations are made.

9.1.1 Bypass flow criteria

Present practice should be reviewed with the intent of determining

if present bypass flow rates are a significant contributor to wear and

aging factors.

9.1.2 Secondary bypass flow test loop

Present practice should be reviewed with the inteit of determining

the degree of need for in-plant full-flow test loops to simulate the

various emergency pumping scenarios. Such loops would permit the pump to

be regularly "exercised" at typical emergency condition flow rates,

rather than the potentially abusive bypass flow rates of most present installations. This would also permit performance monitoring and trend

establishment at realistic operating conditions.
52

9.1.3 Monitoring and establishing trends

State-of-the-art monitoring capability should be studied in regard

to its value in assessing wear and aging factors in AUXFPs. Monitoring

factors to be studied should include at least the following: (1) orbital

rotor vibration, (2) bearing temperature and noise, (3) rotor axial position, and (4) delivered head-capacity performance. Trends in this monitored information should be studied and correlations with wear and aging

factors sought. Details of these topics are covered in Sect. 8.2.

9.1.4 Scheduled disassembly and detailed inspection

Periodic disassembly and detailed inspection of AUXFPs are necessities. The proposed inspection and component renewal factors are outlined

in Sect. 7.2.3 and discussed in Sect. 8.1. The most likely time for

these functions to be performed would be during each refueling period.

9.1.5 Pump specifications

It is recommended that, when new pumps are purchased, specifications

be written to take advantage of the latest proven advances in pump design.

Results from projects such as those being carried out by the EPRI2 can be

helpful in preparing such specifications.

9.2 Recommendations

It is recommended that the appropriateness and utility of the surveillance and monitoring methods described in this report be examined in

the next phase of the investigation on AUXFPs. The measurable parameters identified also are to be evaluated in terms of effectiveness for

use in detecting degradation and establishing degradation trends.

The relationship of the first-phase study to the NPAR Program

strategy is illustrated by the cross-hatched part of Fig. 4.
0RN'L-DWG. 86-4226 ETD

PN.ctuco. Cost ah inth Code0

Kft.lnIdg P.,fo...ane. St.dy Sto.dawds Kr~~~~owt~~~~~dq. - I In~~~~~~~~~dicaows co~~ 00001 4

to.$~~~~o

5.rv.ce .1

I N.~~~~~~~~~~~~~~~~~

Fig. 4. NPAR program strategy.
54

REFERENCES

1. U.S. Nuclear Regulatory Commission, Nuclear Plant Aging Research

WPAR) Program Plan, NUREG-1144, Oak Ridge Natl. Lab., July 1985.

2. Electric Power Research Institute, "Feed Pump Hydraulic Performance

and Design Improvements," Research Project RP1884-10 (ongoing).
S3I8V88i1 SJ

'IN
55

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56

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01, Electric Power Research Institute, July 1979.

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t2 7. E. Makay, "Enlarge Radial Gap in Feed Pumps," Power, March 1981.
57

*28. M. L. Adams and E. Makay, "How to Apply Pivoted-Pad Journal Bearings," Power, October 1981.

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Power Plant, ERCO-10,756, Energy Research and Consultants Corp.,

March 1985.

34. Hydraulic Institute Standards, 14th Edition, 1983.

35. American Petroleum Institute (API) Standard 610, "Centrifugal Pumps

for General Refinery Services," 6th Ed., January 1981.

36. M. L. Adams and J. Padovan, "Insights into Linearized Rotor Dynamics," J. Sound Vib. 76(1), 1981.

37. M. L. Adams, "Nonlinear Dynamics of Flexible Multi-Bearing Rotors,"

J. Sound Vib. 71(1), 1980.

38. M. L. Adams, E. Makay, and J. P. Dimmer, "Increasing Feed Pump Reliability," Proceedings, 1980 Reliability Conference for the Ezectric Utility Industry.

39. M. L. Adams, E. Makay, and J. P. Dimmer, "Feed Pump Vibration

Research - Interstage Fluid Annuli," ASME Paper No. 80-JPGC/Pwr-20.

40. M. L. Adams, "An Introduction to Rotor Dynamics," Mach. Des.,

August 1972.

41. M. L. Adams, "Keep Rotor Vibration Under Control," Power, August

1978.

42. M. L. Adams, ed., Proceedings of Symposium on Rotor Dynamical Instability, ASME Book No. AND Vol. 55, ASME Applied Mechanics Division, June 1983.
58

43. S. E. Pace, Project Manager, Feed Pump Hydrautic Performance and

Design Improvements, Research Project RP1884-10, Electric Power Research Institute (ongoing).
59

*

Appendix A

SUMMARY OF ASME BOILER AND PRESSURE VESSEL CODE

SECT. XI REQUIREMENTS

The Sect. XI requirements for pumps are given in Subsection IWP,

In-service Testing of Pumps in Nuclear Power Plants. This subsection

discusses in-service testing of Class 1, 2, and 3 pumps required to

achieve the cold shutdown condition for a reactor or to mitigate the

consequences of an accident.

The in-service testing procedure given in Paragraph IWP-3100 is as

follows:

"An in-service test shall be conducted with the pump operating at nominal motor nameplate speed for constant speed drives

and at a speed adjusted to the reference speed for variable

speed drives. The resistance of the system shall be varied

until either the measured differential pressure or the measured

flow rate equals the corresponding reference value. The test

quantities shown in Table IWP-3100-1 shall then be measured or

observed and recorded as directed in this Subsection. Each measured test quantity shall then be compared with the reference

value of the same quantity. Any deviations determined shall be

compared with the limits given in Table IWP-3100-2 and the

specified corrective action taken."

*Work performed by G. A. Murphy, ORNL Nuclear Operations Analysis

Center.

Table IWP-3100-1. In-service test quantities

Quantity Measure Observe

Speed N (if variable speed) V

Inlet pressure Pi la

Differential pressure AP v

Flow rate Q V

Vibration amplitude V V

Proper lubricant level or pressure V

Bearing temperature Tb

aMeasure before pump startup and during test.

Source: ASME Boiler and Pressure Vessel Code, Sect. XI,

IWP, p. 209 (1983).
Table IWP-3100-2. Allowable ranges of test quantities

Alert rangea ~Required action

Test quantity Acceptable Alert rangea range

Low values High values Low values High values

Pib

AP 0.93 to 1.02APr 0.90 to 0.93APr 1.02 to 1.03APr <0.90aPr >1.03 APr

Q 0.94 to 1.02Qr 0.90 to 0 .

94Qr 1.02 to 1 . 0 3Qr <0 9

OQr >1.03 Qr

V, when 0 4 Vr 4 0.5 nil 0. to I nil None I to 1.5 mil None >1.5 mil

V, when 0.5 ail < Vr c 2.0 ail 0 to 2Vr mil None 2Vr mil to None >3Vr mil

3vr nil

V, when 2.0 mil < Vr < 5.0 mil 0 to (2 + Vr) nil None (2 + Vr) mil to None >(4 + Vd)

(4 + Vr) ail mil

V, when Vr > 5.0 mil 0 to 1.4Vr mil None 1.4Vr mil to None >1*8 vr Mil

1.8Vr Mil

TbC

aSee IWP-3230.

bpi shall be within the limits specified by the Owner in the record of tests (IWP-6000).

CTb shall be within the limits specified by the Owner in the record of tests (IWP-6000).

Source: Adapted from ASME Boiler and Pressure Vessel Code, Sect. XI, IWP, p. 210 (1983).

An

0
61

Paragraph IPW-3110 states the following:

"Reference values are defined as one or more fixed sets of

values of the quantities shown in Table IWP-3100-1 as measured

or observed when the equipment is known to be operating acceptably. All subsequent test results shall be compared to these

reference values or with new reference values established in

accordance with IWP-3111 and IWP-3112. Reference values shall

be determined from the results of the first in-service test run

during power operation. Reference values shall be at points of

operation readily duplicated during subsequent in-service testing."

Corrective action is given in Paragraph IWP-3230 as follows:

a. If deviations fall within the Alert Range of Table

IWP-3100-2, the frequency of testing specified in IWP-3400 shall

be doubled until the cause of the deviation is determined and

the condition corrected.

"b. If deviations fall within the Required Action Range of

Table IWP-3100-2, the pump shall be declared inoperative and not

returned to service until the cause of the deviation has been

determined and the condition corrected.

"c. Correction shall be either replacement or repair per

IWP-3111, or shall be an analysis to demonstrate that the condition does not impair pump operability and that the pump will

still fulfill its function. A new set of reference values shall

be established after such analysis.

"d. When tests show deviations greater than allowed (see

Table IWP-3100-2), the instruments involved may be recalibrated

and the test rerun."

Paragraph IWP-3400 specifies the frequency of in-service tests:

"a. An in-service test shall be run on each pump nominally

every three months during normal plant operation. It is recommended that this test frequency be maintained during shutdown

period if this can reasonably be accomplished, although this is

not mandatory. If it is not tested during plant shutdown, the

pump shall be tested within one week after the plant is returned

to normal operation.

"b. Pumps that are operated more frequently than every

three months need not be run or stopped for a special test, provided the plant log shows each such pump was operated at least

once every three months at the reference conditions, and the

quantities specified were measured, observed, recorded, and analyzed."
62

Paragraph IWP-3500 provides the duration of tests:

"a. When measurement of bearing temperature is

quired, each pump shall be run at least five minutes

ditions as stable as the system permits. At the end

time at least one measurement or observation of each

quantities specified shall be made and recorded.

not reunder conof this

of the

"b. When measurement of bearing temperature is required,

each pump shall be run until the bearing temperatures (IWP-4310)

stabilize, and then the quantities specified shall be measured

or observed and recorded. A bearing temperature shall be considered stable when three successive readings taken at ten

minute intervals do not vary by more than three percent."
63

Appendix B

OPERATING EXPERIENCE DATA BASES AND REPORTS

Failure information obtained from LER, NPRDS, and IPRDS data bases

is summarized below.

B.1 Nuclear Operations Analysis Center RECON

Licensee Event Reports Data Base Survey

Abstracts of all LERs and reports issued prior to LERs by U.S.

utilities are stored on the Nuclear Operations Analysis Center data base

that can be accessed through the DOE RECON system'. A search was made of

this data base for all events indexed as auxiliary or emergency feedwater

pumps. Each 100-word abstract was reviewed to determine (1) the mode of

failure, (2) the mode of detection, (3) the maintenance activity, and

(4) the cause of failure.

This review found 53 events out of a total of 1139 events involving

AUXFPs during the time period 1973-1983. Results are summarized in Table

B.1.

During the review of the 1139 events certain types of failures were

noted that were excluded from this study: (1) failure to test; (2) obvious design errors including seismic analysis; (3) obvious operator errors;

(4) failure of the pump driver (motor, turbine, diesel generator, etc.);

(5) failure of a valve in the system; and (6) failure of instruments or

controls. Only failures of the pump itself are included in Table B.1.

Bearing, packing, and seals are the primary failedjparts. Failures

of shafts, impellers, internals, and housings appear to be isolated cases.

Forty-two percent of all failures were detected during testing of the

equipment, while 29% were detected during pump operation. A few problems

(6%) were identified during maintenance action.

Sixty-seven percent of the maintenance action was replacement or

repair of damaged or worn subcomponents. In 6% of the events a modification was made to correct the deficiency. The major failure cause was

insufficient or improper cooling or lubrication (23%). Improper maintenance accounted for 17%, while wear was the cause in 15% of the events.

For 29% of the failures, the cause was either unknown or not stated.

B.2 Nuclear Plant Reliability Data System

The NPRDS data base was searched for AUXFP failures. Table B.2 is a

summary of the results that contains the failure information derived from

codes specified by the utilities. 'Out of a total of 70 events-found, 14

were actual failures of the pump itself. In cases where more than one

*Work performed by G. A. Murphy, ORNL Nuclear Operations Analysis

Center.
64

Table B.1. Summary of AUXFP-type

failures reported in LERs

(1973-1983)

Item Ratea

Failed component

Bearings 48

Packing and seal 30

Casing 4

Internal components 4

Impeller 2

Capacity 2

Shaft 2

Other 8

Methods of detection

Testing 42

Operation 29

Maintenance 6

Not stated 23

Maintenance action

Replacement 67

Repair 25

Modification 6

No repair required 2

Identified cause

Lack of lubrication or 23

cooling

Maintenance error 17

Wear/end of life 15

Design error 6

Crud 4

Operator error 2

Other 4

Not stated 19

Unknown 10

aTotal of 53 events.
65

Table B.2. AUXFP-type failures

reported in NPRDS data base

(1974-1985)

Item Ratea

Item ~~~~~(%)

Failed component

Packing/gasket 50

Bearings 38

Internal components 6

Shaft 6

Methods of detection

Incidental observation 33

Surveillance testing 25

Routine observation 12

Audiovisual alarm 12

Operational abnormality 12

Special inspection 6

Maintenance action

Repair/replace 94

Modify 6

Failure cause

Wear 58

Lubrication 12

Binding 12

Aging 12

Abnormal stress 6

aTotal of 14 failure reports.

cause code was given, only the first or the one supported by the event

narrative was used.

The predominant failed subcomponents were pump packingand bearing

failure, with only isolated failures of other subcomponents. Testing or

observation detected most of the failures (88%), with 12% involving

operational abnormalities.

Repair or replacement of subcomponents was the primary maintenance

action (94%), with only 6% being modified. Wear was listed as the primary failure cause (58%). Lubrication, binding, and aging were 12% each.
66

B.3 In-Plant Reliability Data Study

The IPRDS contains equipment service histories obtained from maintenance records from four nuclear power plant sites covering six units.

The various data are recorded in coded form rather than descriptive.

Table B.3 lists IPRDS data for AUXFPs. A total of 12 age-related failures were found in 79 maintenance reports for these units. The listed

failure mode is based on symptoms, including leaks (50%) followed by low

output (25%). Vibration accounted for 8%, while 17% were unknown or not

given. As reported in other data bases, repair or replacement of the

subcomponent was the predominant maintenance action. Modifications were

made in only 17% of the events. The failure cause (listed as failure mode

in the other data bases) was seals and packing (50%). Bearing failures

caused 17%, while personnel errors caused another 17%. Loose fasteners

and couplings and design errors accounted for 8% each. The severity

levels of the incidents were coded as incipient (75%) and degraded (25%).

Table B.3. In-plant reliability

data study

Item Ratea

Failure mode

Leak 50

Low output 25

Vibration 8

Other/unknown 17

Maintenance action

Repair/replace 83

Modify 17

Failure cause

Seals/packing 50

Bearings 17

Personnel error 17

Loose fastener/coupling 8

Design error 8

Severity

Incipient 75

Degraded 25

aTotal of 12 records.
67

Appendix C

AUTOMATIC TRIPPING AND FAILURE

An overspeed trip on turbine-driven AUXFPs gives the one automatic

tripping action that is built into all turbine-driven AUXFPs. However,

overspeed tripping at startup has been a problem. The cause (and thus

available fixes) of this problem is well understood by operating personnel at some plants.

One recommendation by the major governor supplier is the addition of

a separate oil-pressurizing pump for the governor, which comes on prior

to the steam-valve-open action. Plant engineers are not favorable to

this proposal because it involves an additional pump with additional

potential reliability and maintenance problems. In at least one plant

(Palo Verde), the corrective fix was to install an additional steam line

(1 in.) to the drive turbine that is used for the initial part of startup

and is opened before the main steam inlet valve is opened. This approach

allows the speed control governor time to gain control of the situation

and to achieve the operating speed without such a large overshoot as to

overspeed trip the machine.
69

Appendix D

ENGINEERING INFORMATION RELEVANT TO AUXFP RELIABILITY

D.1 Introduction

Since the March 1979 TMI incident, considerable attention and

scrutiny has been focused on auxiliary feedwater systems by the NRC and

industry groups. This process is still in the phase of fact-finding,

assessment, and determination of what industry-wide measures are required

to ensure the high reliability appropriate to safety-related systems and

equipment. AUXFPs are critically important safety-related components.

Therefore, to provide further insight into the pertinent engineering

areas, the following sections of this appendix present selected technical

background information critical to high-head-per-stage centrifugal pumps.

D.2 Pump Hydraulic Instability

"Hydraulic instability" is now the term most commonly used in the

centrifugal pump field to label the highly active unsteady flow phenomena

that become progressively more pronounced the farther away from bestefficiency flow that a pump is operated. As clearly identified throughout the main body of this document, these unsteady flows are the most

significant contributor to deterioration (i.e., aging and wear) of pump

components because of the high-amplitude dynamic forces that they produce. In main feedwater systems, these unsteady flow phenomena also frequently produce hard-to-control conditions in the feedwater system, leading to transients, trips, and upset conditions.

Figure D.1 shows delineation of stable and unstable flow regimes for

various classes (i.e., different specific speeds) of power plant pumps.

It is based on a composite of field troubleshooting experience, shop

tests, and laboratory tests; it represents what current hydraulic design

technology can typically provide. Current EPRI-sponsored research

addresses this area of feed pump design technology.

Hydraulic instability has also been characterized by the label internal "flow recirculation," which occurs both at the inlet and discharge

regions of a pump stage at off-design operating flows. These flow recirculation cells, as illustrated in Figs. D.2 and D.3, are highly unsteady,

giving rise to large vibration excitation forces and hard-to-control flow

pulsations in the entire pump loop. Figure D.4 illustrates the ramification of hydraulic instability with regard to the head-capacity characteristic of a pump. Figure D.4 shows how different pumps, all satisfying

the same best-efficiency flow condition, can have a significantly different quality of hydraulic performance at off-design operating flows.

Figure D.5 shows how the vane-passing shock intensity increases with

smaller impeller-diffuser vane tip clearance.

The feedwater pump is also an active element in the overall feedwater flow loop. Thus, proper analysis of pump system stability problems

requires the complete feedwater loop to be treated as a connected system.
70

ORNL-DWG. 86-4227 ETO

10

Cn

U)

z

U)

co0

z

C] z

0

aU)4 IL

0~

U)

2

L _.

0 25 50 75 100

PUMP FLOW, % OF RATED

Fig. D.1. Anticipated useful operating ranges for pumps used in

large nuclear and fossil power generating units. (Inner line of design

margin area is preferred; if hydraulic instability occurs at higher flows

various pump and system problems can be expected.) Source: E. Makay and

0. Szamody, Survey of Feed Pump Outages, EPRI FP-754, Electric Power

Research Institute, April 1978.
71

ORNL--DWG. 86-4228 ETD

DYNAMIC FORCES

(AT VANE PASSING FREOU.)

(a) DEVELOPMENT OF STALL, CAVITATION DAMAGE AND FINALLY VANE

DAMAGE IN A FIRST STAGE BOILER FEED PUMP VANED DIFFUSER

BLADE

INLET

ANGLE

FORMATION OF STALL IN AN IMPELLER EYE DUE TO

FLOW INCIDENCE ANGLE

(VISUALIZED ON EXPERIMENTAL TEST RIG)

Fig. D.2. Formation of stall

impeller. Source: E. Makay and 0.

Zines for Feedwater Pumps in Large

Electric Power Research Institute,

(a) in diffuser and (b) in eye of

Szamody, Recommended Design GuidePower Generating Units, EPRI CS-1512,

September 1980.

._ 4_ .
- -

- -

72

ORNL-DWG. 86-4229 ETD

Ae / ZI

~~~<)~~~~~t (S~~~~~

L

U 1-11 70F

f~~~~~~~~

Fig. D.3. Secondary flow pattern in and around pump impeller stage

at off-design flow operation. Source: E. Makay and 0. Szamody, Survey

of Feed Pump Outages, EPRI FP-754, Electric Power Research Institute,

April 1978.
73

ORNL-DWG. 86-4230 ETD

125

0

I

C-

* 50

% BEST EFFICIENCY FLOW

Fig. D.4. Head-capacity characteristics of multistage boiler feed

pumps. Curve "A" is correct and desired for stable system operation.

Curve "B" represents hydraulically unstable impeller-diffuser design.

Parallel- as well as single-pump operation is difficult in the unstable

flow regime. Curve "C" shows design with flat head curve at part load

resulting in control system malfunctioning. Single-pump operation is

possible in unstable regime. Designs "B" and "C" are not acceptable for

utility applications. Source: .E. Makay and 0. Szamody, Survey of Feed

Pump Outages, EPRI FP-754, Electric Power Research Institute, April 1978.
74

ORNL-DWG. 86-4231 ETD

z

0

cc 0

-w

Uo.

:1 < DJ w

- In 00

<a.

1 2 3 4 5 6 7 8

C = % IMPELLER -DIFFUSER/VOLUTE GAP

Fig. D.5. Influence of pump impeller to diffuser/volute radial gap

on pressure pulsation at blade-passing frequency and rotor deflection

caused radial forces. Source: R. Makay and 0. Szamody, Survey of Feed

Pump Outages, EPRI FP-754, Electric Power Research Institute, April 1978.
75

Laboratory tests' are now being conducted to determine the transfer

matrix of pump stages for imbedding into a total system stability

analysis.

D.3 Rotor Vibration

As described earlier, strong hydraulic forces at off-design flows

are one major source of pump vibration. Other significant sources, include rotor unbalance, resonance, rotor-bearing instability (oil whip),

internal rubbing, misalignment, coupling, and combinations of these. The

use of spectrum analyzers has been an invaluable tool in diagnosing and

correcting various pump vibration problems in the field. Through extensive field troubleshooting experience and related redesign work on highpressure pumps, Fig. D.6 has been assembled to aid in isolating the

fundamental cause(s) of excessive pump vibration. Figure D.6 uses pump

flow as the independent parameter and is supplemented by Fig. D.7, which

uses rotational speed as the independent parameter.

Rotor vibration levels, as usually measured at the bearing journals,

can be evaluated as being acceptable or unacceptable, based on long

experience of what various types of machinery can comfortably endure

throughout the equipment's usable lifetime. Figure D.8 is typical of

such guidelines used throughout the industry in assessing, from monitored

vibration, whether the vibration levels are acceptably low to be consistent with high reliability and acceptable plant life.

D.4 Cavitation and NPSH

Boiler feed pump failures caused by cavitation erosion damage are

among the highest-outage-producing pump problems, because cavitation frequently causes severe pump internal damage, requiring new internal components or lengthy factory repair. Cavitation is caused by the production

of very low local pressures adjacent to flow boundaries,.such that vaporfilled pockets form and then collapse violently as they are transported

into higher pressure regions. Damage is most likely to occur in the inlet

of the stage but may be carried through the impeller causing erosion of

the impeller exit or the diffuser (or volute) inlet. Cavitation can be

caused by:

1. inadequate NPSH of the feedwater system (i.e., not enough pressure

at the pump suction);

2. flow recirculation at the impeller eye while operating in the offdesign flow regimes;

3. incorrect hydraulic design of the first-stage impeller (incorrect

blade inlet angle);

4. localized high velocities caused by sharp corners and other flow

disturbances such as misplaced inlet guide vanes;
76

ORNL-DWG. 86-4232 ETD

BLADE PASSING FREQUENCY: SHOULD NOT BE DOMINANT

IF YES, MAY EXCITE LOWER FREQUENCY RESPONSES. Z x N

_ I iI- - -I- --- -

(Z-NO. OF IMPELLER BLADES. IF DOUBLE VOLUTE OR STAGGERED

BLADES OF DOUBLE SUCTION IMPELLER, TWICE BLADE PASSING

FREQUENCY MAY ALSO BE HIGH)

z

0,

0

z

0

z

!2

a.

0~

6

z

uJ

n

LiVIB. RARELY DOMINANT IN THIS FREO. RANGE

1.01

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LU

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u-I-

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00

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0I- W ,;

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ALWAYS PRESENT,

.7

.6

SOMETIMES DOMINANT IN=ROTOR REV.

FUNCTION /

OF% FLOW/

Z22Zi

>0~z

MAY BE STRONG FOR BOTH VOLUTE AND DIFF

TYPES, REDUCED IN SOME CASES WITH TILT-PAD

AND/OR WEAR-RING OR SEAL REDESIGN

UNSTABLE BEARINGS OR _ // 77/

SEALS, INTERNAL RUB _ _ _ _ _ _ //

- _ _____ ..:..7..7_

.5

.4

.3

.2

.1

0

..W I^ .( YI II\ 

iXX - z X A

E Z 10<% ALWAYS PRESENT IN BOTH DIFF AND VOLUTE TYPE PUMPS

<_MAY BE MORE DOMINANT IN VERTICALS AT OFF-DESIGN FLOWS

AXIAL VIB: MAY BE STRONG W. SINGLE STAGE

DOUBLE SUCTION HORIZONTAL PUMPS .............. ~~~ ..... _ ~~~ I

0 25 50 75

% FLOW

100 125 150

Fig. D.6. Frequencies of hydraulically induced dynamic forces

acting on the rotor of a centrifugal pump. Source: E. Makay and 0.

Szamody, Survey of Feed Pump Outages, EPRI FP-754, Electric Power

Research Institute, April 1978.
77

ORNL-DWG. 86-4233 ETD

4.

3,

2

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a

w

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6 -.... ...

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0. - .~~~sJ . . *c~~~~fr. ..... CY - .'tA6E'F . . .

0.

0.0 1. 2.

OPERATING RPM/FIRST CRITICAL RPM

3.

Fig. D.7. Vibration frequency vs speed. Causes and cures of common

vibration problems of centrifugal pumps, steam and gas turbines, fans and

blowers, compressors, and other centrifugal equipment. Source: E. Makay

and 0. Szamody, Survey of Feed Pump Outages, EPRI FP-754, Electric Power

Research Institute, April 1978.
78

ORNL-DWG. 86-4234 ETD

FREQUENCY - (CPM OR RPM)

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N1 I

0.3

0.2

N

2 3 4 6 8 10 20 40 60

FREQUENCY - (CPS)

100 200 400

__I S 0.1

, 6 7 8

Fig. D.8. Allowable rotor vibration levels measured relative to

bearing cap. (Values shown are filtered readings at that particular RPM

or frequency; experimental standards). Source: E. Makay and 0. Szamody,

Survey of Peed Pump Outages, EPRI FP-754, Electric Power Research Institute, April 1978.
i

79

5. vortex formation due to obstacles in the flow path, sharp elbows in

the suction piping, incorrect pump inlet geometry, and blunt inlet

guide vanes; and

6. high-frequency machine vibration can displace water particles perpendicular to a solid surface creating vapor pockets, thus creating

cavitation (least important type with centrifugal pumps).

Operation of pumps at off-design conditions for extended periods of

time can cause cavitation damage independent of available NPSH, due to

high-incidence-angle-caused stall and secondary flows like eye recirculation as shown in Fig. D.2 and D.3. In a multistage boiler feed pump,

impeller cavitation damage usually occurs in the first stage, but it can

occur at other locations where flow conditions satisfy the above requirements. It is important to distinguish between first-stage impeller cavitation and pump internal cavitation; because the'latter is not related to

pump NPSH. Feed pumps producing high-head/stage are more receptive to

cavitation damage because of the higher energy-input densities to the

fluid. Velocities and dynamic forces are high enough to accelerate cavitation and fatigue damage of pump internal components.

If cavitation damage occurs at design capacity due to insufficient

NPSH, usually cavitation damage'can be seen on either or both sides of

the impeller blade inlet portion. The damage starts at the leading edge

of the vane and may cover a large area. Another type of damage can be

observed on the exposed side of the vane located in the corner where the

blade joins the impeller hub. This type of damage indicates a mismatch

between approach flow and impeller inlet angles that can be caused by

extended operation of the pump in the low flow regime, even if NPSH is

adequate to prevent cavitation. If severe impeller erosion appears somewhat downstream from the vane inlet edge at the periphery of the impeller

eye, the damage may be caused by inlet flow recirculation (see Fig. D.3).

The impeller is then operated in the off-design regime or the impeller

eye is too large, causing development of flow recirculation at the impeller eye. If the damage starts from the vane inlet and is on the nonexposed side of the vane, then the pump is undersized for the application,

that is, operated at substantially larger than best efficiency flow for

extended periods of time.

D.5 Pump Internal Clearances

The critical clearances in a high-speed feed pump are those between

rotating and stationary parts where high-pressure differentials exist

(wear-rings of impellers, balancing device cylindrical surfaces, and seal

surfaces) and the gap between impeller periphery and diffuser vanes or

volute tongues. All these clearances affect pump efficiency as well as

pump reliability. The closer the clearances, the higher the efficiency,

but in most components, the lower the reliability (i.e., seizure and

internal breakage are more probable).

The commonly used close clearance internal dimensions, which have

evolved over years of experience, are nearly the same for all manufacturers and are suitable for reliable operation. If unreasonably high
80

pump efficiencies are specified or demanded, the manufacturer is inclined

to reduce these internal clearances below the commonly used values. Such

a reduction of these clearances improves hydraulic efficiency because of

the resulting reduction in interstage leakage, balancing device leak-off

flow and seal flow. However, this efficiency improvement exists only

during the factory acceptance test and for a short period of time in the

field. The clearance surfaces wear-in to approximately the commonly used

values, and the artificially produced higher efficiency vanishes. However, in the process the reliability of the pump is jeopardized by an

increased potential for rotor seizure and rubbing-induced subsynchronous

rotor vibration, either of which can result in destruction of the rotor

and unexpected outage.

Reduction of the normal radial gap between impeller and diffuser (or

volute) improves efficiency to some degree. As an impeller vane passes

by a stationary blade (diffuser tip or volute tongue), a hydraulic shock

occurs that can be observed in the liquid, on the structure, or on rotor

vibration measured at the bearings or any part of the shaft. The influence of the radial gap on pressure pulsation at blade passing frequency

and rotor-deflection-caused radial forces are shown in Fig. D.5. Numerical values are not given on the vertical scales, because they are also

functions of other design parameters. The radial gap is given as a percentage of the impeller diameter. If the gap is too small (e.g., 1%) the

phenomenon can be self-destructing, because the rotor exciting forces

increase exponentially as shown in Fig. D.5. Rubbing at wear surfaces

may also introduce subsynchronous vibration amplitudes that can rapidly

destroy the rotor. If the impeller and stationary components are structurally marginal, the result can be disintegration of these elements. If

these structures are strong, the result may be complete destruction of

the whole rotating element. If such failure occurs, the radial gap is to

be examined and if found too small, it is to be opened up to normal

dimensions. Generally accepted dimensions are

Diffuser type, minimum gap: 3%

Volute type, minimum gap: 6%

D.6 Pump Component Design

A brief review of pump components subject to frequent failures or

malfunctioning is given in this section. The experience was gained

primarily on boiler feed pumps and therefore has equal applicability to

AUXFPs, which are essentially of the same multistage high-head-per-stage

configuration. In fact, AUXFPs are less robust than typical boiler feed

pumps, because AUXFP designs appear to reflect the relatively small

amount of design operating hours.

D.6.1 Axial balancing device

Axial forces in a boiler feed pump are, in the minority of cases,

held in equilibrium by opposing equal number of impellers with a thrust
81

bearing to assist the pump during startup and to take up the residual

unbalance force caused by casting tolerances, minor dimensional differences, and wear. If the impellers all face the same way, either a

balance drums or a balance disk is used to take the high axial thrust.

When the impellers are opposed, although the forces are in equilibrium,

some designs still utilize a balance drum for safety, because at partload operation, large unpredictable hydrodynamic forces are produced that

could damage the thrust bearing.

A balance drum is basically a rotating piston that has the characteristics of an ineffective water-lubricated radial bearing, so it

influences the dynamic behavior of the rotor (Fig. D.9). The balance

disk, on the other hand, is basically a water-lubricated hydrostatic

thrust bearing, Figs. D.9 and D.10. The small gap "e" controls the

pressure and consequently the thrust balancing pressure in cavity "A."

If the gap "e" becomes too small or closes entirely, the faces will touch

and destruction of the mating parts results. Introduction of a small

taper oetween faces is an effective design improvement successfully

recommended in numerous plants. It is the relative taper angle "alpha,"

and not the orientation of the faces, that is important. Figure D.10

shows the parallel and the tapered face designs. Also shown is-a force

balance diagram, clearly indicating the superior behavior of the tapered

disk design. The disk force counterbalances the forces produced by the

impellers, transmitted to the disk through the shaft. This force can

easily have a magnitude over 100,000 lb and is responsible for internal

damage in many pumps. Figure D.10 shows the force To, which is the force

when the disk is closed. It is vital that the forces produced by the

impellers can never be higher than To, otherwise failure results. Many

ORNL-DWG. 86-4235 ETD

BINGHAM DE LAVAL INGERSOLL RAND

BYRON JACKSON WEIR, LTD. KSB

HITACHI WORTHINGTON SULZER

PACIFIC ALLIS CHALMERS

~~~~~~~jP~~~~~~~~~~L

HP~~~L

(a) (b) (c)

Fig. D.9. Customary axial balancing devices for high-pressure

multistage boiler feed pumps. (a) Balance drum (b) and (c) balance

disk. Source: E. Makay and 0. Szamody, Survey of Feed Pump Outages,

EPRI FP-754, Electric Power Research Institute, April 1978.
82

ORNL-DWG. 86-4236 ETD

T12 / p2 7Trdr

R1

T1 2 -

PO _

T1: jP 1

PARALLEL FACE DESIGN

TAPERED FACE DESIGN

O }\ >~ LEAKGE

J

I

To

I

l V 3 d FLOW. GPM

THRUST. LB

VI

0.0005 0.008

GAP e, in

. _

T=j+T 12 -T2

(a) BALANCE DISK WITH PARALLEL FACES.

FORCE EQUILIBRIUM IS SHOWN WHEN

DISK IN NORMAL OPEN POSITION.

FORCE DEVELOPED BY DISK WHEN

FACES JUST CLOSED. FACES ARE

PARALLEL TAPERED

(c) BALANCE DISK THRUST

CARRYING CAPABILITY

AND LEAK-OFF FLOW.

0.0005 GAP e, in

(b) TAPERED FACE BALANCE DISK. (d) DYNAMIC STIFFNESS OF

BALANCE DISK.

Fig. D.10. Parallel and tapered face balance disk designs.

Source: E. Makay and 0. Szamody, Survey of Feed Pump Outages," EPRI

FP-754, Electric Power Research Institute, April 1978.
83

failures are due to the fact that the disk design load capacity is marginal when the pump is new. As the wear-ring surfaces wear with normal

use, the hydraulic forces on the impellers grow, resulting eventually in

a higher force than the marginal thrust-carrying capability of the disk.

Figure D.10 shows the basic principal difference between parallel and

tapered face designs, the tapered design having only advantages over the

parallel. Note that when the balancing disk is approaching closed position, the tapered face-design is able to take much higher thrust loads

than the parallel face design, hence it is much more reliable, particularly under large transient loads that accompany severe hydraulic instability.

Complete disintegration of the rotating balance disk reported from

one large generating station called attention to a typical failure mode.

Customary material for that component is type 416 SS, where the material

specification clearly states not to.heat-treat above 42 Rockwell C hardness, because the possibility of surface cracking is high. When the

failed component was tested for hardness, it was found to be over 50

Rockwell C. Other unused disks were tested and found to be over 42

Rockwell C hardness. Surface cracks were found severe enough to ensure

future failures. This discrepancy was found in at least three manufacturers' products; thus, it was not a localized problem. It is a good

practice to test for hardness every time a new disk is put in-service to

intercept this type of failure cause. The best practice, however, is to

heat-treat the stationary part to somewhat higher hardness, because that

part is not subjected to centrifugal forces as is the rotating disk.

D.6.2 Shaft seals

Shaft seals have the highest failure-rate for any component in

boiler feed pumps. Figure D.11 shows the three seal designs used for

boiler and nuclear feedwater pump services with single injection. Double

injection is used for high-temperature applications to avoid flashing in

the seal area. Injection water can be regulated by temperature or pressure control. Low-speed booster pumps may employ packing; however, highspeed applications (3600 rpm or above) are exempt from that seal application with the exception of AUXFPs. The majority of failures occur with

seal designs other than labyrinth types. Labyrinth seals are the least

demanding "work horses" of utility pumps.

Floating-ring type or mechanical seals have the highest failure rate

among all pump component failures. When in good condition, the mechanical

seals have lower leakage rate than the labyrinth or packing types but

obviously are more prone to wear and failure. eAte: some wear, the seal

leakage increases rapidly, especially with the floating-Tgt ype. This

then requires excessive maintenance that results in lower pump, hence

unit, availability or if not repaired, rapidly decreasing pump delivered

capacity. Considering the high cost of unit down time to replace failed

seals and the high repair cost with the mechanical or floating-ring seal

types, it becomes obvious that the best seal type for large feed pumps is

the labyrinth type. Several utilities report successful conversions to

labyrinth seals after a long history of failures with the other seal

types.
84

ORNL-DWG. 86-4237 ETD

(A) (B) (C)

Fig. D.11. Shaft seal types used in boiler feed, nuclear feed, and

feedwater booster pumps. (a) Labyrinth, (b) floating ring seals with

single injection, and (c) mechanical seal. Source: E. Makay and 0.

Szamody, Survey of Feed Pump Outages, EPRI FP-754, Electric Power

Research Institute, April 1978.

The demand of other seal types results from the higher apparent

leakage rate of the labyrinth type. The leakage of a properly designed

labyrinth seal is not higher than for a 'floating-ring type. Mechanical

face seals, in spite of their high failure rate, are favored many times

because they do not require external injection water, and when performing

well, have the lowest leakage rate among all seal types. The tradeoff

between apparent higher efficiency and pump reliability should be fully

recognized.

D.6.3 Journal bearings and rotor

Practically all boiler feed pumps apply "flexible" shafts, which

means that the pump operating speed is above the rotor first lateral

critical speed. This is also true of most AUXFP designs. In addition to

this, because of the relatively light rotor weight, the journal bearings

are lightly loaded, which makes them prone to instabilities such as oilwhip (subsynchronous vibration component). Single-stage nuclear feed

pumps in general operate below the first critical speed, but the rotor

weight is even lighter than boiler feed pump rotors. The difficulty with

light bearing loads is that the speed at which rotor dynamic instability

starts is lower than normally expected. This speed is called the threshold speed, above which the bearing fluid film looses its ability to damp

out rotor excitation forces at frequencies below approximately half the

rotational frequency. Self-excited subsynchronous rotor whirl instability (oil-whip) may also occur. This loss of low-frequency bearing
85

damping is particularly harmful in feed pumps (even at speeds below the

threshold speed) because of the large low-frequency hydraulic forces that

are produced by hydraulic instabilities, especially at part-load operation. Subsynchronous or low-frequency excitations also originate in the

seals and wear-ring surfaces, induced both by fluid dynamical phenomena

and by rubbing. Feed pump reliability would therefore be improved considerably by the development of advanced bearing and rotor configurations

that introduce large amounts of low-frequency damping into the system.

A typical bearing system configuration for AUXFPs is shown in Fig.

D.12.

D.6.4 Impeller breakage

Impeller breakage usually results in major pump damage, and causes

are difficult to determine. Breakage is frequently the result of vibration, hydraulic instability, and cavitation damage. However, damage may

also result from design deficiencies such as stress risers, inadequate

strength, or faulty casting quality.

D.6.5 Shaft breakage

The existence of stress risers or material flaws can contribute to

shaft breakage due to fatigue, but failure of other pump components will

produce exceptionally high stress conditions. In most cases, the shaft

failure is diagnosed as a secondary failure mode caused either by another

component disintegration just prior to the shaft breakage or as a result

of high-vibration amplitudes at impeller vane-passing frequency for an

ORNL-DWG. 86-4238 ETD

HEAT EXCHANGER

Fig. D.12. Typical AUXFP bearing system.
86

extended period of time. Frequent failures are a result of overloading

the shaft by the high hydraulic forces acting on the axial balancing

device. If the retaining mechanism of the disk (or drum) or the fitting

of it is not proper on the shaft, very high cyclic forces result, accelerating fatigue failure (see Fig. D.13).

ORNL DWG. 86-4239 ETD

SHAFT GEOMETRY AT FAILURE LOCATIONS:

IMPELLER HUB

1. SHAFT GROOVE OR INTERSTAGE

EXTERNAL DAMAGES:

IMPELLER RETAINING RING

-/ - i

/ 4i

FAILURE LOCATION

POSSIBLE FAILURE LOCATION

SHRUNK-ON PART

SUCH AS:

COUPLING.

SLEEVE. ETC. 2. THREADED OR SHRUNKON PARTS: - y

Fig. D.13. Typical multistage centrifugal pump shaft failure locations. The four most frequent locations are shown in order of failure

frequency. Source: E. Makay and 0. Szamody, Recommended Design GuideLines for Feedwater Pumps in Large Power Generating Units, EPRI CS-1512,

Electric Power Research Institute, September 1980.
87

D.6.6 Wear rings

Excessive wear of the impeller wear rings is in most cases the

result of excessive shaft flexibility and operation at conditions where

large shaft vibrations are encountered; it can also follow journal bearing wear. Overly close wear-ring clearances in a new pump or new components will also lead to rapid wear of the rings. In any case, wear of

these rings will make it impossible for pump efficiency to be maintained

at the initial level. Therefore, new pumps should have sufficiently

large clearances to ensure against rubbing or seizure. As with balancing

disks, a proper choice of component materials is important for providing

good accommodation of occasional rubs, thus avoiding consequential pump

failures and premature replacement of wear rings.

Reference

1. Electric Power Research Institute, "Feed Pump Hydraulic Performance

and Design Improvements," Research Project RP1884-10 (ongoing).
89

NUREG/CR-4597

Volume 1

ORNL-6282/V1

Dist. Category RV

Internal Distribution

1. R. J. Borkowski 18-21. G. A. Murphy

2. W. G. Craddick 22. H. E. Trammell

3-6. D. M. Eissenberg 23. ORNL Patent Office

7. J. A. Getsi 24. Central Research Library

8-14. W. L. Greenstreet 25. Document Reference Section

15. H. D. Haynes 26-27. Laboratory Records Department

16. R. C. Kryter 28. Laboratory Records (RC)

17. A. P. Malinauskas

External Distribution

29-32. M. L. Adams, Case Western Reserve University, Room 417, Glennan

Building, Cleveland, OH 44106

33. K. I. Baron, American Society of Mechanical Engineers, United

Engineering Center, 345 East 47th Street, New York, NY 10017

34. J. H. Ferguson, Connecticut Yankee Atomic Power Company, R. R.

No. 1, Box 127E, East Hampton, CT 06424

35. A. B. Johnson, Pacific Northwest Laboratory, P.O. Box 999,

Richland, WA 99352

36-39. E. Makay, Energy Research and Consultants, 900 Overton Avenue,

Morrisville, PA 19067

40. C. Michelson, ACRS, 20 Argonne Plaza, Suite 365, Oak Ridge, TN

37830

41. S. E. Pace, Electric Power Research Institute, P.O. Box 10412,

Palo Alto, CA 94303

42. J. E. Richardson, Division of Engineering Technology, Office of

Nuclear Regulatory Research, U.S. Nuclear Regulatory Commission,

5650 Nicholson Lane, Rockville, MD 20852

43. R. L. Simard, Institute for Nuclear Power Operations, 1100

Circle 75 Parkway, Atlanta, GA 30339

44. M. L. Stanley, EG&G Idaho, Inc., P.O. Box 1625, Idaho Falls, ID

83401

45. D. F. Sullivan, Division of Engineering Technology, Office of

Nuclear Regulatory Research, U.S. Nuclear Regulatory Commission,

5650 Nicholson Lane, Rockville, MD 20852

46. J. H. Taylor, Engineering and Risk Assessment Division, Department of Nuclear Energy, Brookhaven National Laboratory, Upton,

NY 11973

47. J. P. Vora, Division of Engineering Technology, Office of

Nuclear Regulatory Research, U.S. Nuclear Regulatory Commission,

5650 Nicholson Lane, Rockville, MD 20852

48. Office of Assistant Manager for Energy Research and Development,

Department of Energy, Oak Ridge Operations Office, Oak Ridge, TN

37831

49-50. Technical Information Center, Department of Energy, Oak Ridge,

TN 37831

51-425. Given distribution as shown in NRC category RV (10 - NTIS)
NRC FORM 33S US.NUCLEAR RIGULATORYCOMMISSION I REPORTNUMBERIAsUd'b ,TIOC. ddVol No. -lYi

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NRCM 1102. DT HE 3201. 3202 BIBLIOGRAPHIC DATA SHEET NUREG/CR-4597,

SEE INSTRUCTIONS ON THE REVERSE Vol. 1, ORNL-6282/Vl

2 TITLE AND SUBTITLE 3 LEAVE BLANK

Aging and Service Wear of Auxiliary Feedwater Pumps

for PWR Nuclear Power Plants Volume 1. Operating

Experience and Failure Identification 1 DATE REPORT COMPLETED

MONTH YEAR

5 AUTHORISI June 1986

M. L. Adams and E. Makay 6 DATE REPORT ISSUED

Energy Research and Consultants Corporation MONTH YEAR

July 1986

7 PERFORMING ORGANIZATION NAME AND MAILING ADDRESS tOdeNro. Z Cod) B8 PROJECTtTASK.YWORK UNIT NUMBER

Oak Ridge National Laboratory

P. 0. Box Y 9 FIN OR GRANT NUMB ER

Oak Ridge, TN 37831

B0828

10 SPONSORING ORGANIZATION NAME AND MAILING ADDRESS .c1.,d#ZDpCode) Ila TYPE OF REPORT

Division of Engineering Technology T

Office of Nuclear Regulatory Research Topical

U.S. Nuclear Regulatory Commission I. PERIOD COVERED f.lcfs..atesi

Washington, DC 20555

12 SUPPLEMENTARY NOTES

'3 ABSTRACT 1200 .ooiC oc Pi,#

Typical auxiliary feedwater pump (AUXFP) configurations are described in terms of

configuration details, materials of construction, operating requirements, and modes

of operation. AUXFP failure modes and causes due to aging and service wear are

identified and explained, and measurable parameters (including functional indicators)

for potential use in assessing operational readiness, establishing degradation trends,

and detecting incipient failures are given.

A series of measures to correct present deficiencies in surveillance, monitoring,

and in-service testing practices is discussed. The main body of the report is

supplemented by a number of relevant appendixes; in particular, a major appendix

is included on engineering and design information useful to assess AUXFP operational

readiness. t

This report was produced under the U.S. Nuclear Regulatory Commission's Nuclear Plant

Aging Research Program.

IS AVAILABILITY 14 DOCUMENT ANALYSIS - a KEVWORDS.OESCRIPTORS

Pumps, centrifugal pumps, auxiliary feedwater pumps, aging, service wear,

maintenance, degradation, trend, failure mode, failure cause, measurable

parameter, incipient failure, inspection, surveillance, monitoring,

functional indicators, operating experience

a IDENTIFIERS CPEN ENDED TERMS

STATEMENT

Unlimited

16 SECURITYCLASSIFICATION

ITA's ofiel

Unclassified

Unclassified

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Id PRICE