NUREG/CR-5583

KEI No. 1656

I Prediction of Check Valve

Performance and Degradation

in Nuclear Power Plant Systems --

Wear and Impact Tests

Final Report

September 1988 - April 1990

Prepared by M. S. Kalsi, C. L Horst, 3. K. Wang, V. Sharma

Kalsi Engineering, Inc.

Prep re-d for,

t U.S. Nuclear Regulatory Commission

I

I ,
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NUREG/CR-5583

KEI No. 1656

RG, RV

Prediction of Check Valve

Performance and Degradation

in Nuclear Power Plant Systems --

Wear and Impact Tests

Final Report

September 1988 - April 1990

Manuscript Completed: May 1990

Date Published: August 1990

Prepared by

M. S. Kalsi, C. L. Horst, J. K. Wang, V. Sharma

W. S. Farmer, NRC Project Manager

Kalsi Engineering, Inc.

745 Park Two Drive

Sugar Land, TX 77478

Prepared for

Division of Engineering

Office of Nuclear Regulatory Research

U.S. Nuclear Regulatory Commission

Washington, DC 20555

NRC FIN D2510
ABSTRACT

Check valve failures in nuclear power plants have led to safety concerns as well as

'extensive damage and loss of plant availability in recent years. Swing check valve

internals may experience premature deterioration if the disc is not firmly held open

against its stop. At the present time no guidelines exist for the prediction of degradation

trends and the determination of suitable inspection intervals.

This research Is aimed at developing a reliable quantitative wear and fatigue prediction

model for swing check valves which can be used to improve the safety and reliability of

their operation. Additionally, the predictive techniques can be used to augment check

valve preventive maintenance programs by providing the plant engineer with quantitative

and conservative estimates of damage potential for check valve internals.

- iii -
TABLE OF CONFENTS

Page

EXECUTIVE SUMMARY 1

INTRODUCTION 2

Background 2

Overall Research Objectives 2

Phase I Research 2

Phase II Research 3

WEAR PREDICTION 5

Theoretical Wear Equation 5

Wear Coefficient, K 5

Sliding Distance 5

Load 6

ACCELERATED WEAR TESTS 7

General 7

Valve and Hinge Pin Modifications for Wear Testing 9

Selection of Materials to Accelerate Wear 9

Preliminary Wear Tests, Accuracy of Measurements, and Eliminating Effects 12

of Corrosion and Erosion

Matrix of Accelerated Wear Tests 13

WEAR TEST RESULTS AND DISCUSSION 15

Accelerated Wear Test Results 15

Normalized Wear Rates 19

Comparison Between Theoretical Wear Prediction and Tests 2D

WEAR PREDICIION METHODOLOGY - SlUMARYAND CONCLUSIONS 2

IMPACT AND FATIGUE PREDICTIONS 23

Theoretical Prediction Techniques 23

Disc Natural Frequency Approach 23

Pipe Eddy Frequency Approach 24

Mean and 3G Disc Speed Approach 24

IMPACT AND FATIGUE TESTING 25

General Methodology 25

Test Valve 25

Calibration of Strain Gaged Valve Disc 29

Instrumentation 29

-V.
TEST RESULTS 3)

Typical Impact Signature 30

Discussion of Impact and Disc Motion Signals 2

Discussion of Disc Motion Measurements and Modified Disc 33

Use of 3a Disc Speed for Disc Impact Force Estimates 34

Rate of Occurrence of Impacts 35

CORRELATION OF IMPACT TEST DATA TO THEORY 37

Kinetic Energy Conversion Approach 37

Disc Oscillation Amplitude and Frequency Approach 37

IMPACT FORCE AND FATIGUE PREDICTION MEHODOLOGY- 4

SUMMARY AND CONCLUSIONS

CORRELATION AGAINST PLANT DATA 41

Plant Correlation Example 1: 41

18-Inch Swing Check Valves with Excessive Hinge Pin Bushing Wear

Plant Correlation Example 2: 44

18-Inch Swing Check Valves with Low Hinge Pin Bushing Wear

Plant Correlation Example 3: 44

4-Inch Tilting Disc Check Valve Modifications to Reduce Hinge Pin Wear

Plant Correlation Example 4: 45

10" Swing Check Valve Disc Stud Fatigue Failure

SUMBIARY OF CHECK VALVE REVIEW OME ODOLOGY 47

Overview of Check Valve Review and Analysis Procedure 47

Vmin Calculation 47

Calculation of Modified Vmin using Cup 60

Wear and Fatigue Calculations 51

Conclusions 51

REFERENCES

APPENDIX Az Impact and Fatigue Prediction Methods

APPENDIX B.: Calculation of Vmin, Flow Velocity, and Wear Rate for

Plant Correlation Example 1

APPENDIX C: Correlation of Fatigue Analysis with Available Plant Data

for Plant Correlation Example 4

APPENDIXD-. Plots of Mean Disc Speed and 3-Sigma Disc Speeds for

the 3-Inch and 6-Inch Test Swing Check Valves

- vi -
LISr OF FIGURES

Figure No.

1 Typical Swing Check Valve Geometry

2 Instrumented Test Valve

8 Valve Modifications for Accelerated Wear Tests

4 Modified Hinge Pin for Accelerated Wear Tests

6 Typical Aluminum Hinge Pin and Bushings Used in Wear Tests

6 Multihole Orifice Plates Used to Generate Turbulence

7 Typical Hinge Pin Wear Results for the 3-Inch Swing Check Valve

(Upstream Disturbance: 36 x 3/16-Inch Hole Turbulence Plate at 2.5 D)

8 Accelerated Wear Test Results for the 3-Inch Valve

9 Accelerated Wear Test Results for the 6-Inch Valve

10 Comparison of Accelerated Wear Test Results to Theoretical Predictions

11 Instrumented Disc Cross Section

12 Strain Gage Bridge Schematics

13 Strain Gaged Check Valve Disc

14 Detail of Strain Gage Installation

15 Disc Showing Sealed Wire Exit

16 Wire Exit from Disc

17 Completion of Pressure Boundary Using Hermetic Connector

18 Impact Signature, Upstream Turbulence Source

19 Impact Signature, Upstream Elbow

23 Impact Signature, Upstream Turbulence (Modified Disc)

21 Raw and Filtered Impact Force Signatures

22 Disc Displacement Plot Corresponding to Impact Event of Figure 21

23 Impact Measurement Assembly

24 Impact Force Survey (Based on Filtered Signal, Ibs)

25 Impact Force Survey (Based on Filtered Signal, Ibs)

26 18-Inch Swing Check Valve Installation Used in Example 1

27 Hinge Pin Bushing Wear in 18-Inch Valves (Example 1)

28 Severe Hinge Pin Wear in 4-Inch Tilting Disc Check Valve

29 Minimum Velocity Formula Comparison Against Test Results

SD Significance of Disc Projection (Y/D) on Performance

Page

4

7

10

n

n

12

15

17

is

21

26

26

26

27

27

28

26

31

31

31

32

33

34

96

36

42

43

44

48

49

- Vi-
L1ST OF TABLE

Table No. Page

LA Data for 3-Inch Valve 8

1B Data for 6-Inch Valve 8

2 Matrix of Hinge Pin Accelerated Wear Tests 14

3 Summary of Hinge Pin Accelerated Wear Tests 14

4 Summary of Test Parameters and Wear Rates Results of 16

Hinge Pin Accelerated Wear Tests

5 Disc Assembly Resonant Frequencies, hz 30

6 Impact Force Ranges (After Low-Pass Filtering) 35

7 Comparison of Theoretical and Measured Impact Forces 39

6-Inch Modified Disc Assembly; High Turbulence Source

at 1.5 Diameters Upstream

- viii -
ACKNOWLEDGEMENTS

We are greatful for the funding provided by the Small Business Innovation Research

program of the U. S. Nuclear Regulatory Commission which made this research possible.

Special thanks are due William S. Farmer of the NRC for his guidance and

encouragement throughout the project.

The authors would also like to pay special recognition to Daniel Alvarez for the

development of the load cell concept employed in these tests, and to Bobbie Lambert for

patiently working through many revisions and for her painstaking attention to detail in

preparing this manuscript.

- ix -
EXECUlMVE SUMMARY

Check valve failures at nuclear power plants in recent years have led to serious safety

concerns, and caused extensive damage to other plant components which had a significant

impact on plant availability. In order to understand the failure mechanism and improve

the reliability of check valves, a systematic research effort was proposed by Kalsi

Engineering, Inc. to U.S. Nuclear Regulatory Commission (NRC) under their Small

Business Innovation Research (SBIR) program. The overall goal of the research was to

develop models for predicting the performance and degradation of swing check valves in

nuclear power plant systems so that appropriate preventive maintenance or design

modifications can be performed to improve the reliability of check valves.

Under Phase I of this research, a large matrix of tests was run with instrumented swing

check valves to determine the stability of the disc under a variety of upstream flow

disturbances (elbows, reducers, butterfly valves, and multiple hole orifice plates as high

turbulence sources), covering a wide range of disc stop positions (50 to 75 degrees) and flow

velocities (up to 20 ft/sec) in two different valve sizes (3- and 6-inch). Phase I research led

to the development of upstream flow disturbance factors which should be taken into account

to determine the minimum velocity required to achieve a stable, full open disc position.

The matrix of tests also quantified the severity of the disc fluctuations that can be expected

when these minimum velocity requirements are not met. The results of Phase I research

were published in NUREG/CR-5159.

The goals of Phase II research were to develop predictive models which quantify the

anticipated degradation of swing check valves that have flow disturbances closely

upstream of the valve and are operating under flow velocities that do not result in full disc

opening. Two major causes of swing check valve failure are premature degradation due to

hinge pin wear and fatigue of the disc stud connection to the hinge arm. A matrix of

accelerated wear tests were performed using aluminum hinge pins and bushings in the 3-

and 6-inch valves to quantify wear experienced in the hinge pin area. A special disc

instrumented with strain gages was used in the 6-inch valve to measure the impact forces

and their rate of occurrence to quantify the fatigue damage caused by tapping of the disc

against the stop. Based on this theoretical and experimental research, wear and fatigue

prediction models have been developed which show good correlation against laboratory test

results as well as against a limited number of check valve failures at the plants which had

been sufficiently documented in the past.

This research allows the inspection/maintenance activities to be focussed on those check

valves that are more likely to suffer premature degradation. The quantitative wear and

fatigue prediction methodology can be used to develop a sound preventive maintenance

program. The results of the research also show the improvements in check valve

performance/reliability that can be achieyed by certain modifications in the valve design.

-1-
INTRODUCTION

Background

A review of check valve failures in nuclear power plants shows that severe degradation of

the internals during normal plant operation is responsible for most of the failures (Refs. 1

through 8). Ideally, check valves should be sized to provide full disc lift under normal flow

conditions, and should be located sufficiently away from upstream flow disturbances to

avoid premature degradation. Where these ideal conditions are not met, all check valves

can be expected to experience some degradation. However, the level of degradation may

vary from negligible to very severe, depending upon a number of factors relating to the

design, installation, operating conditions, usage, and maintenance practices. Previous

surveys have shown that, even though over 70 percent of the check valve installations do not

meet the "ideal" requirements of minimum flow velocity and distance from upstream

flow disturbances (Ref 9), serious problems do not exist with this large population of

valves. In fact, the actual percentage of check valves that fail in service due to premature

degradation is estimated to be only 1 to 2 percent (Refs. 2, 3). This shows that the previous

classification of "misapplied" check valves (based on failing minimum velocity and

proximity of the upstream disturbance criteria) is too coarse, and better screening

techniques are need to truly identify the valves that are most likely to suffer from

premature degradation.

Overall Research Objectives

In order to accomplish this goal of better identifying swing check valves that are most

likely to suffer premature degradation, Kalsi Engineering, Inc. proposed a systematic

research program to the U.S. Nuclear Regulatory Commission under the Small Business

Innovation Research (SBIR) program. The major objective of the proposed research was to

develop quantitative techniques for predicting swing check valve performance and

degradation in nuclear power plant systems. These predictive techniques can be

integrated into preventive maintenance programs with the aim of identifying the few

valves most likely to be suffering from accelerated degradation and ensuring that

maintenance efforts are expended first on those valves with the greatest need. The

proposed research was selected and funded by NRC in two phases. Phase I was completed

in April 1988, the results of which were published in Reference 1. Since the details of Phase

I research are important for a complete understanding of the overall predictive techniques

developed under Phase I and II, the reader is encouraged to familiarize himself with the

contents of Reference 1 which are briefly summarized here.

Phase I Research

The major objective of Phase I research was to investigate the stability of the swing check

valve disc at different flow velocities in piping systems that have significant flow

disturbances within 10 pipe diameters upstream of the check valve. The effects of elbows,

reducers, and turbulence sources of varying degrees of severity on the disc stability were

investigated. A matrix of over 2,000 flow tests was completed to determine the disc stability

of three-inch and six-inch swing check valves. Under these upstream flow disturbances,

the minimum flow velocity requirements to achieve full disc opening, as well as the

maximum disc fluctuations (when the flow velocities are less than the minimum

requirements), were determined for a wide range of test conditions.

- 2 -
PhaseflResearch

The objective of Phase II research was to develop quantitative techniques for predicting the

degradation of internals of swing check valves operating under velocities insufficient to

fully open the disc. Figure 1 shows a typical swing check valve. There are two major

areas in swing check valves that are subject to premature degradation: the hinge pin, and

the disc stud connection to the hinge arm. Hinge pin failures are caused by excessive

wear, and disc stud connection failures occur because of impact related fatigue.

The overall goals of Phase II research were accomplished by the development of appropriate

theoretical models supported by necessary laboratory testing, and then correlation to actual

plant data that were available. Accelerated wear testing using aluminum hinge pins and

bushings was done on three-inch and six-inch instrumented test valves for a number of

upstream flow disturbance configurations. Impact force measurements were made in the

six-inch test valve using a disc that had been modified to incorporate strain gages.

As a result of this extensive theoretical and experimental research, predictive methodology

for quantifying degradation of the check valve internals caused by hinge pin wear or disc

stud impact has been developed. The predictions were found to be in reasonable agreement

with the actual performance at the operating plants, as shown in the plant correlation

examples. For reliable predictions based on the methodology developed, a detailed

knowledge of the geometry and design of the check valve internals, its installation details,

operating conditions, and usage is required. The predictions should be compared with the

actual performance history, whenever possible, to refine various assumptions used in the

analytical methodology. When properly implemented and integrated with the preventive

maintenance activities, the quantitative degradation prediction techniques developed

under Phase I and II research have the potential of significantly improving the overall

reliability of the check valves in nuclear power plants.

-3 -
HINGE ARM

r DISC/STUD CONNECTION

HINGE PIN

D

Figure 1

Typical Swing Check Valve Geometry

-4 -
WEAR PREDICTION

Theoretical Wear Equation

Excessive hinge pin wear is a common cause of many swing check valve and tilting disc

check valve failures. Adhesive wear is the most commonly encountered and least

preventable form of wear in sliding contacts. Hinge pin wear, whether it is caused by a

two-body adhesive mechanism in which wear particles escape the rubbing interface or a

three-body abrasive wear mechanism in which wear particles remain entrapped at this

interface, can be predicted by a simple equation which has the same form for both types of

wear mechanisms (Ref. 10):

W=KLVt

H

where W = Volume of material worn away

L = Load between the two bodies in sliding contact

V = Relative velocity between sliding surfaces

t = Time over which sliding occurs

H = Penetration hardness of the surface of material being worn away

K = Nondimensional "wear coefficient" for the sliding pair of

materials

Wear Coefficient, K

Ideally, an experimentally determined value of the wear coefficient for the sliding

material combination under the actual environmental conditions should be used to make

accurate wear predictions. The wear coefficient can be determined experimentally by

using any of the standard wear test machines described in Reference 10. In the absence of

an experimentally determined wear coefficient for the exact material combination and

environmental conditions of interest, a reasonable choice can be made by judicious use

and interpolation of available data from various sources. The use of previously published

data for wear coefficients requires a careful consideration of several factors such as

compatibility of the alloying elements used in the metal pair, the degree of lubrication

provided by the fluid, and the operating temperature. Adhesive wear coefficients span a

range from 3 x 10-6 to 1500 x 10-6 for various combinations of metal surfaces sliding under

varying degrees of lubricity provided by water and steam to no lubrication. Extensive

work has been done in this area by Rabinowicz (Refs. 10, 11) to develop a table of wear

coefficients for adhesive wear, which can be used within its statistical bounds of

uncertainty to make wear estimates. However, it is important to note that the individual

behavior of a specific tribological pair of metals may be higher or lower than the average of

the category into which they fall (Ref. 11). Actual data for the specific metal combinations,

therefore, remains the best choice for accurate wear predictions.

Sliding Distance

In the wear equation (1), the term V t can be equated to d, the total distance the two wear

surfaces have slid relative to each other. For check valves, this total sliding distance can

be estimated from the disc oscillation frequency and its amplitude. The amplitude of disc

-5-
fluctuation depends upon the type of upstream disturbance and its proximity. The results of

extensive testing on three-inch and six-inch valves under Phase I in which disc

fluctuations were measured under a variety of common upstream disturbances are

documented in Ref. 1. Evaluation of the disc motion data obtained from the Phase I tests

shows that in general the disc does not oscillate at a well-defined single frequency.

However, an estimate of the dominant frequency at which the disc is likely to oscillate can

be made by calculating the disc natural frequency based on one of three methods: fluid

spring stiffness, pendulum frequency, and eddy frequency approaches as described in

Appendix A. Alternatively, the mean disc speed can be used directly in making wear

calculations, if it is known. Appendix D includes the results of Phase I test data which

were processed to obtain the mean and 3a disc speeds as a function of flow velocity,

upstream disturbances, and full disc opening angles for the three- and six-inch test

valves. These data span a range of flow velocities and upstream disturbances and their

proximity to the check valve. These disc speeds can be used directly or for reference and

comparison against the frequency- and amplitude-based disc speed estimates.

Load

The load, L, between the hinge pin and the bushing is the result of disc weight and fluid

dynamic lift and drag components on the disc. A free body diagram analysis of the

complete force system shows that this resultant load remains within ± 20 percent of the

effective disc weight for swing check valves over a wide range of disc opening angles.

Since the uncertainty in wear coefficients is much larger than this, the effective disc

weight, which includes the hinge arm weight contribution (Weff = WdiSc + 0.5 Warm), can

be used as the total reaction load in making wear estimates for swing check valves.

However, in tilting disc check valves, since the disc remains in the flow stream, fluid

dynamic drag forces at very high fluid velocities and for non-aerodynamically shaped

discs will be much higher than for swing check valves. The total hinge pin reaction load

can therefore be significantly higher than the disc weight alone. Fluid dynamic drag

forces should, therefore, be considered in determining total load estimates for tilting disc

check valves.

-6 -
ACCELERATED WEAR TESTS

General

Accelerated wear tests were performed using aluminum hinge pins and bushings in the

three-inch and six-inch swing check valves under a number of upstream flow disturbance

configurations and flow velocities to validate and support the development of a reliable

wear prediction model. Aluminum wear parts have been primarily used by others in

determining relative wear trends in double disc and swing check valves; however, no

attempt was made in these studies to achieve correlation against absolute wear rate

predictions (Refs. 13, 14). By substituting soft aluminum for the hardened carbon steel or

stainless steels normally used as hinge pin/bushing materials, significant amounts of

wear can be produced within a few hours that normally take several months or years of

actual operation in the plant. The ability to run many tests allowed the important

parameters to be varied in a systematic manner in a matrix of accelerated wear tests, discussed in the following sections.

Figure 2 shows the three-inch test valve with a Plexiglass bonnet and a displacement transducer for continuous disc motion monitoring. Dimensional and weight data for the

three-inch and six-inch test valves used in Phase I and II tests are included in Tables 1A

and 1B.

I t1

Figure 2

Instrumented Test Valve

-7-
Valve Type

Manufacturer

Weight and Dimensional Data

Disc Weight (incl. nut), Wdisc

Hinge Arm Weight, Whinge

Hinge Pin Diameter, dhinge

Disc O.D., Ddise

Seat Bore Diameter, D

Seat Tilt From Vertical, a

Full Open Angle From Vertical, a +

Fluid Impingement Angle, 0

Hinge Pin to Disc Stud Distance, R

Disc Projection, Y

3", 300# Swing Disc Check

MCC Pacific Valve Company

1.94 lbs.

0.72 lbs.

0.375"

3.90"

3.0"

30

730

17°*

2.85"

0.0"

Table 1A

Data for 3-Inch Valve

Valve Type

Manufacturer

Weight and Dimensional Data

Disc Weight (incl. nut), Wdisc

Hinge Arm Weight, Whinge

Hinge Pin Diameter, dhinge

Disc O.D., Ddijs

Seat Bore Diameter, D

Seat Tilt From Vertical, a

Full Open Angle From Vertical, a +,

Fluid Impingement Angle, 0

Hinge Pin to Disc Stud Distance, R

Disc Projection, Y

6", 300# Swing Disc Check

MCC Pacific Valve Company

8.94 lbs.

3.38 lbs.

0.500"

6.90"

6.0"

30

700

200*

5.12"

0.8"

Table 1B

Data for 6-Inch Valve

* See discussion on the effect of valve geometry on fluid impingement angle 0 in section

entitled "Summary of Check Valve Review Methodology."

-8-
Valve and Hinge Pin Modifications for Wear Testing

Figures 3, 4, and 5 show the modifications made in the hinge pin and the hinge arm

bushing area of the test valves to incorporate certain key features. The hinge pin wear area

was reduced to relatively narrow bands, approximately 1116-inch wide, so that reduction in

the wear diameter could be accurately measured to within ± .0001-inch. Even though actual

weight loss measurement (within ± 0.2 milligrams accuracy) was used as the primary

method of wear quantification, the diametrical measurements were used as an

independent check on the location and magnitude of wear. In order to ensure that wear

took place at the same location on the wear bands, an indexing slot and tang arrangement

was used on one end of the pin to prevent rotation of the pin and ensure the correct angular

orientation after disassembly and reassembly to make intermediate weight loss

measurements. Pin rotation was prevented to ensure that all of the wear occurred between

the wear bands on the aluminum hinge pin and the aluminum bushing and that no sliding

took place between the aluminum hinge pin and the carbon steel body.

To allow for easy removal of the hinge pin for frequent weight-loss measurements, a

dummy pin could be inserted which pushed out the test pin and simultaneously supported

the hinge arm. The hinge pin was provided with generous chamfers so it would slide in

and out smoothly.

Selection of Materials to Accelerate Wear

A thorough investigation of the type of aluminum to be used was done to ensure that reliable

wear data could be developed. Heat treatable aluminum alloys such as 6061-T6, 7075-T6,

etc., were ruled out to avoid possible problems caused by continuing change in their

mechanical properties due to their age hardening characteristics. To achieve high wear

rates and maintain reasonably constant material properties, commercially pure

aluminum (Alloy 1100F) was selected because of its low hardness and its freedom from age

hardening phenomenon. After machining, the aluminum hinge pin was fully annealed

to develop the desired mechanical properties. In this condition, the hardness of the

material was tested to be 23 Brinell. The material had a yield strength of 3,600 psi, tensile

strength of 12,000 psi with 51 percent elongation, and 93 percent reduction of area. In order

to guarantee consistency of material behavior from pin to pin, all test pins of the same size

were made from the same bar, even though different bars were used for different sizes.

-9-
Figure 3

Valve Modifications for Accelerated Wear Tests

- 10-
WEAR SURFACES

ANT I-ROTATISON

\ \& I NDEX ING SLOT

LEAD IN CHAMFERS

Figure 4

Modified Hinge Pin for Accelerated Wear Tests

 % ,0  1! 1,1pa I q- .  -.1 - __ 1

F~gur 5-e ; ;;

Typical~~~,, Hig i Aluinu n uhnsUe nWa e

-11-
Preliminary Wear Tests, Accuracy of Measurements,

and Eliminating Effects of Corrosion and Erosion

All of the preliminary wear tests were performed on the three-inch Pacific swing check

valve with a multi-hole orifice plate placed 2.5 diameters upstream. A number of multihole orifice plates were used in Phase I tests (Fig. 6); the specific orifice plate used in Phase

II tests had 36 holes of 3/16-inch diameter, which was determined to be the most severe

turbulence source during the Phase I tests. The flow velocity chosen for these tests was 7

ftlsec, which corresponds to a zone of operation where the mean disc speed of the fluctuating

disc was near its maximum value (around 220/sec), but the disc stud was not tapping

against the backstop. These test conditions were chosen to create maximum wear rates for

the three-inch valve during preliminary testing.

Figure

Multihole Orifice Plates Used

6

to Generate Turbulence

A series of tests were run with the objective of accurately quantifying wear rates and

separating them from other effects such as corrosion or erosion which can also cause a

material loss from the hinge pin. Initial tests were run with the aluminum hinge pin

being run against 17-4 PH stainless steel bushings. In order to separate the contribution of

material loss caused by sliding wear from that caused by corrosion and erosion, a series of

tests was run: one in which the disc was allowed to oscillate, the second in which the disc

was prevented from fluctuating by bringing the backstop down to the appropriate position,

- 12 -
and another with pins immersed in still water. These tests revealed that the material loss

due to corrosion was not an insignificant part of the total hinge pin weight loss for the threeinch valve hinge pin. From these results, it was concluded that provisions will need to be

made to reduce the corrosion material loss so that it is well below the levels encountered in

these tests.

It was decided to anodize the aluminum hinge pin to eliminate corrosion related material

loss, leaving the wear band areas on the hinge pin unanodized by masking them during

the anodizing process. The next series of tests run with the anodized aluminum hinge pin

against 17-4 PH stainless steel bushings confirmed that corrosion wear effects were

virtually eliminated, and practically all of the weight loss encountered was attributable to

sliding wear. However, it was also found that the actual wear rates obtained by using the

aluminum hinge pin against 17-4 PH stainless steel bushings were less than our goal to

accomplish accelerated wear testing within 20 to 30 hours average for any given set of

conditions being tested.

Since it is well documented (Refs. 10, 11) that the wear coefficients using an identical pair

of materials (e.g., Al vs Al) at the sliding contact are on the average three times higher

than those obtained with compatible material ( e.g., Al vs. Fe), the next series of tests were

run with the aluminum hinge pin sliding against aluminum bushings. The specific

aluminum alloy selected for the bushings was 7075-T6, which is harder than the 11OOF

alloy used for the hinge pin. This was done to ensure that wear would occur predominantly

on the hinge pin. As expected, this change produced higher wear rates, thus permitting

typical test durations of 20 to 30 hours during subsequent tests. As described in more detail

later, the actual test duration for any given configuration ranged from 6 hours to 40 hours.

The minimum duration of each test was governed by our goal of reducing the overall error

to less than 10 percent in the weight loss measurements considering the accuracy and

repeatability of the laboratory balance used, and the amount of material that could be lost

while removing and inserting the hinge pin for intermediate measurements.

This concluded the preliminary debugging stage of the wear tests. All of the subsequent

long-term wear tests were done with the aluminum hinge pin (with the non-wearing

surface anodized) sliding against the aluminum bushing.

hMatrix ofAccelerated Wear Tests

Table 2 shows the matrix of accelerated wear tests that were run on the three- and six-inch

valves. Two types of disturbances at varying upstream distances were used in a judicious

manner to cover the anticipated range in the mean disc speed, which, as predicted by the

fundamental wear equation, should directly affect the wear rates in a linear manner. The

turbulence plates with small holes used in Phase I tests that produced the most severe disc

oscillations were used as one of the upstream flow disturbances, with upstream distances

ranging up to 10 diameters. An "elbow up" configuration at different upstream distances

was used as the second source of flow disturbance. Table 2 identifies flow disturbance,

flow velocity, test number, and the total hours of testing for each configuration.

For each valve size, the flow velocities chosen corresponded to the approximate range of

maximum disc fluctuations. In some tests, the flow velocity was intentionally selected to

be different to determine if the wear behavior followed the predictions based on the

fundamental wear equation. In one of the tests, a combination of the two flow disturbances

- 13 -
(elbow up plus turbulence plate) was used to study its combined effect on the wear rates. All

together, a total of eight unique test configurations identified by A-1 through A-8 were tested

for the three-inch valve, and seven configurations identified by B-1 through B-7 were tested

for the six-inch valve. As summarized in Table 3, the average duration of tests was 30.6

hours and 19.7 hours respectively for the three-inch and six-inch valves. Each test was

spanned by several intermediate segments to inspect the wear surfaces and make

intermediate weight measurements. The test segments averaged 5.1 hours and 7.7 hours

respectively for the three-inch and six-inch valves.

11 ~~~~~Ts~ubrTotal ~ Valve | Disturbance | Flow TestNumber Tet

Size f ource Velocity Distance To Disturbance Source Hours

l l QD | I1.~~~0 5D |2.5D I 3D I 4.5D |IOD -

36 x3/16"7fpA8 A1A2 A4

Turbulence 7 A-i A-2 A-4 13

PlateA3

3" Elbow Up 7.5 fps A-5 A-6 76

Elbow Up + 7.5 fps A-7 8

361x 3116"

Turbulence

Plate

361x 3/8" 55fsB1 B6B55

Turbulence 5.5 | B-i B-6 B-S |8

Plate

63 Elbow Up 5.5 fps B-4 |3 B-7 

Elbow Up 9.0 fps B-2 21

Elbow Up, .fp B36

(w/worn pin) ps B-3 l 6 

Table 2

Matrix of Hinge Pin Accelerated Wear Tests

Table 3

Summary of Hinge Pin Accelerated Wear Tests

- 14-
WEAR TEST RESULIS AND DISCUSSION

Aocelerated Wear Test Results

Figure 7 shows a typical plot obtained from the accelerated wear test (Test No. A-1) on the

three-inch test valve, with a severe turbulence source (orifice plate with 36 holes of 3/16-

inch diameter each) located 2.5 diameters upstream of the check valve. Intermediate

measurements after each test segment show that the hinge pin wear follows a linear trend

as predicted by the wear equation. The number of intermediate weight measurements was

different for each test, but in each case no significant deviations from a linear trend was

noted. The average wear rate for this test configuration is found to be 0.95 mg/hr.

30

25

E

I0o

CL

E

20

15

10

5

0

0 5 1 0 1 5 20 25 30 35

Elasped Time, hrs.

Figure 7

Results for the 3-Inch Swing Check Valve

x 3/16 Inch Hole Turbulence Plate at 2.5 D)

Typical Hinge Pin Wear

(Upstream Disturbance: 36

The overall results of all the accelerated wear tests performed on the three-inch and sixinch test valves are summarized in Table 4. The wear rates were found to span a range

from 0.11 mg/hr to 0.95 mg/hr for the three-inch valve and 0.37 mg/hr to 2.8 mg/hr for the

six-inch valve, depending upon the upstream disturbance source and its proximity.

- 15-
Mean 3a Mean 3a

Number Disc Disc Disc Disc Wear

Valve Test of~or Angl Fluct. Seed Speed Plat

Size No. of Testing (0) (° Isec) (

0 lsec) (mgslhr)

A-1 32 59.0 12.8 22.3 50.8 0.95

A-2 30 60.6 7.8 15.1 34.8 0.39

A-3 8 60.5 11.2 20.0 47.1 0.51

3" A-4 31 60.9 3.8 6.3 14.4 0.16

A-5 36 60.0 5.4 10.7 25.0 0.33

A-6 40 60.8 3.4 6.1 14.3 0.11

A-7 8 60.8 14.0 27.9 63.4 0.86

A-8 37 59.6 11.6 20.3 47.5 0.70

B-1 13 58.1 13.8 17.0 42.4 2.80

B-2 21 66.1 6.0 9.9 23.6 2.48

6" B-3 6 60.3 5.2 4.9 13.2 1.04

B-4 29 60.3 5.8 6.4 15.3 0.37

B-5 23 58.2 9.6 11.6 26.3 1.38

B-6 22 53.4 13.0 14.6 32.7 2.52

B-7 24 60.5 4.4 5.5 12.9 0.47

Table 4

Summary of Hinge Pin Wear Test Results

To provide better insight into the overall results from accelerated wear tests, wear rates are

plotted as a function of mean disc speed for the three-inch and six-inch valves in Figures 8

and 9 respectively. The tables in these figures are arranged in order of decreasing mean

disc speed, regardless of the type of upstream disturbance and its proximity.

- 16 -
S.

E

0~

1..

E

Mean Disk Speed, Olsec

Upstream IUpsrm Mean Disk Speed 1 Wear Rate Test

Disturbance Distance degreeslsecond mgslhr_ Nunber

Elbow Up + Elbow Up + ~1.51)27.9 0.86 A-7

Turbulence Plate

Turbulence Plate 1.5D 20.3 0.70 A-8

Turbulence Plate 2.5D 20.0 0.51 A-3

22.3 0.95 A-1

Turbulence Plate 4.5D 15.1 0.39 A-2

Elbow Up OD 10.7 0.33 A-5

Turbulence Plate 10D 6.3 0.16 A-4

ElbowUp 3D 6.1 0.11 A-6

FIGURE 8

Accelerated Wear Test Results for the 3-Inch Valve

- 17 -
3

0~

0

Ia

0

C

C

2

1

0

0 5 10 15

Mean Disk Speed, l/sec

20

Upstream Upstream Mean Disk Speed Wear Rate Test

Disturbance Distance degrees/second mgslhr Number

Turbulence Plate 1.5D 17.0 2.80 B-1

Turbulence Plate 2.5D 14.6 2.52 B-6

Turbulence Plate 4.5D 11.6 1.38 B-5

Elbow Up /

Increased OD 9.9 2.48 B-2

Velocity

Elbow Up OD 6.4 0.37 B-4

Elbow Up 3D 5.5 0.58 B-7

Elbow Up P OD 4.9 1.04 B-3

Figure 9

Accelerated Wear Test Results for the 6-Inch Valve

- 18-
The most important conclusion from these plots is that the wear rate depends upon the mean

disc speed, regardless of the type and proximity of the upstream disturbance responsible for

it. This confirms that the fundamental wear equation is a valid qualitative form of

predictive model for hinge pin wear in swing check valves. The quantitative correlation

between the predicted and the actual wear rates will be discussed next.

Normalized Wear Rates

To facilitate a comprehensive comparison between theoretical predictions and test results,

the fundamental wear Equation (1) can be rearranged to express wear rates in a

normalized form as shown below. The volume of the material worn away is given by

Equation (1) and is repeated here for convenience:

W=KLVt (1)

For a material weight density of p, the above equation can be converted to give the weight,

m, of the material lost:

pmKLVt (2)

H

Therefore, wear rate can be expressed as

m =pLV ( )

t H

The average sliding speed, V, at the hinge pin of diameter dhinge can be related to the

angular mean disc speed, e (degree/sec) as follows:

V = 0.5 dhinge x 3,600 (4)

V = 10Ondhinge in/hr

From Equations (3) and (4),

p K L (10c 1) dhing

H

Rearranging

m| (Ld: = pK(10i |6) (5)

- 19 -
where

= Defined as normalized wear rate in mg/hr/lblin

m = Wear rate in mg/hr

L = Load at the sliding surfaces, lb

dhinge = Diameter of the hinge pin, in

p = Weight density of material, mg/in3

K = Nondimensional wear coefficient for the sliding pair of materials

0 = Mean angular disc speed, degrees/sec

H = Penetration hardness of the surface of the material being worn

Comparison Between Theoretical Wear Prediction and Tests

All of the accelerated wear test data for the three-inch and six-inch test valves were reduced

to a plot of normalized wear rate, M, vs. mean disc speed, 9, using Equation (5). The

effective disc weight was used for the load on the sliding surfaces in this equation. Data for

the weights and hinge pin diameters shown in Tables 1A and 1B were used.

Theoretical wear calculations were done by using a density of 44,452 mg/in3 for

aluminum and a surface hardness of Brinell 23, which corresponds to 32,700 psi. Wear

coefficients for adhesive wear for identical metals range from K = 3 x 10-4 for poor

lubrication to K = 15 x 10-4 for no lubrication. Wear coefficients for abrasive wear are

typically an order of magnitude lower (Refs. 10, 11, 12).

Figure 10 shows an overall comparison between all the wear test data and theoretical

predictions. Theoretical predictions based on the generic table of wear coefficients for

adhesive wear for identical metal sliding pairs with poor lubrication to no lubrication

nicely bound all the test results. This confirms the wear mechanism to be the adhesive one

and the quantitative validity of the predictive approach.

A more precise prediction can be made if the actual wear coefficients for aluminum vs.

aluminum, instead of the generic value for identical metals was used. However, no data

for this specific material combination were found in the available published literature.

Further research into the wear coefficients of aluminum against aluminum revealed that

aluminum has a higher wear coefficient than the average. Unpublished data by the

leading tribologist Rabinowicz shows that K for aluminum vs. aluminum is two to three

times higher than the generic average for identical metal pairs, and the use of a geometric

average (log scale) between the poorly lubricated and unlubricated values was

recommended* . This gives a value of K = 6.7 x 10-4. With this specific wear coefficient,

experimental results and theoretical predictions were found to be in excellent agreement.

From this laboratory comparison, it is concluded that the wear model presented in this

report is a valid approach for predicting hinge pin wear in swing check valves. The

application of the model and its correlation against limited plant data that were available

will be presented in a separate section.

* Private communication between Prof. Ernest Rabinowicz, Massachusetts Institute of

Technology, and M. S. Kalsi (principal investigator, NRC/SBIR project), May 8, 1990.

-20-
20,

b.

A.

I

I

AS

1.04

1.0

-~ 9'~ Ton flO Iuredj~lcthn' based

hb (KaIsX ,O:4

Avg of e0p data (K.9g2 X 104

Therglc'Ped~Cbnbased

0"R~ 1 O-67X10 4)

TertclPrediction based

10 

on~~~~~~~~~f poor Iutricl 0~ fK-3 X 1d4) 0.O.

0o

4 Mean Disk Speed$ deg/" C

30

Comparison

IFlgure 10

Of Accelerated Wear test Results to the

-21.
WEAR PREDICTION MXtHODOLOGY - SUMMARY AND CONCLUSIONS

The overall results and conclusions from the swing check valve hinge pin wear prediction

research are summarized below:

1. The wear mechanism causing hinge pin wear is adhesive wear.

2. The fundamental wear equation predictions were found to provide excellent

correlation with the accelerated wear test results for the three-inch and six-inch test

valves under a total of 15 different test configurations that were tested.

3. Mean disc speed can be used to predict wear rates, regardless of the upstream

disturbance source that is responsible for causing the disc motion.

4. Data from the large matrix of tests performed under Phase I research have been

processed and included in Appendix D to provide mean disc speeds for a large

number of upstream disturbances at varying proximities. Where applicable, this

matrix of wear disc speed results can be used in hinge pin wear analysis.

5. The combined effect of more than one upstream flow disturbance was determined in

only one test (A-7). A high turbulence orifice plate and an elbow were tested at 1.5

diameters upstream. The wear rate for this combined disturbance was found to be

approximately the same as would be expected with the more severe source alone.

6. The use of appropriate adhesive wear coefficient requires a detailed knowledge of the

tribological system to make accurate wear predictions. The actual wear coefficients

for different metal sliding pairs under various environmental conditions span a

very wide range. A number of factors, such as compatibility of the alloy elements in

both of the metals, environmental fluid, and operating temperature, should be

considered in selecting wear coefficients from previously published wear data. For

best prediction, experimentally determined wear coefficient for the specific

combination of sliding metals of interest should be used.

7. Even though hinge pin wear was found to follow linear predictions, excessive wear

can cause the wear rates to increase nonlinearly. This can result from increases in

hinge pin load due to possible impact between the sliding parts due to increased

clearances.

8. Differences in the range of factors of 2 to 3 between the actual wear rates and the

prediction based on a sound application of tribological principles are considered

normal (Refs. 10, 11). If the differences are larger than this range, the assumption

used or the actual conditions warrant a closer scrutiny.

9. Measurements of the actual parts in service obviously provide the most reliable

information regarding actual wear rates. Whenever practically possible, these data

should be used to refine future wear predictions and to determine suitable

maintenance intervals.

This summary concludes our discussion on the accelerated wear test research on the threeinch and six-inch valves in the laboratory. Application of this prediction methodology to

actual plant conditions and their correlation are presented in a separate section.

-22 -
IMPACT AND FATIGUE PREDICTIONS

When the flow velocity through a check valve is insufficient to cause the disc to reach the

fully open position, disc oscillation and tapping can occur. Tapping against the backstop

can generate significant impact forces that may result in fatigue failure of the disc stud.

This section describes an analytical approach developed to predict these impact forces as

well as documents the results of impact force measurements made on a specially modified

and instrumented check valve disc during actual flow operation. Theoretical predictions

are compared to the measurements for model verification.

Theoretical Prediction Tedhniques

The kinetic energy of the oscillating disc assembly is assumed to be completely converted

into strain (potential) energy of the impacting members at the instant contact occurs.

Since the stiffness of the elements involved in the impact (disc, disc stud, valve bonnet) is

either known or can be calculated, this strain energy can be used to determine the peak

force delivered during impact. It can be shown that the stiffness of the disc and bonnet is

considerably higher than that of the disc stud; therefore, the strain energy will be absorbed

principally by the much softer disc stud. This system model assumes that the disc stud

behaves as a linear spring element and that no other losses occur that would decrease the

net amount of energy delivered to the disc stud.

The most important element in this model becomes the determination of the disc velocity

and, therefore, the kinetic energy of the disc at the moment of impact. Several approaches

were developed which can be used to predict the disc velocity at impact. Each method relies

on determining a characteristic frequency of disc motion and corresponding amplitude

from which an estimate of linear velocity can be derived. Briefly, these approaches for

determining velocity before impact are:

1. Disc natural frequency based on fluid spring stiffness,

2. Pipe eddy frequency,

3. Mean and 3a disc speed from Phase I test data.

Each of these techniques will be described more fully in the following sections. For a

complete derivation of the velocity and impact force equations presented below, refer to

Appendix A.

Disc Natural Frequency Approach

This approach is based on the assumption that the disc behaves as a single degree of

freedom vibratory system and will therefore oscillate at its natural frequency when

subjected to the random fluctuating forces present in the turbulent flow. The disc weight at

any opening angle is supported by the fluid impingement force, which has a nonlinear

force vs. opening angle characteristic (Ref. 3.). Due to this nonlinear force deflection

characteristic, the disc natural frequency will vary as a function of disc position defined

by angle, 0, which in turn depends upon the flow velocity (Appendix A). The disc velocity

can be calculated as follows:

VdIac = VR I [ 2KpAdiRc V2ZJ (6a)

-23 -
where

Z = [cos e sin.0 + 2 cos 0 sin 0] (6b)

The impact force can be directly calculated using the method described in Appendix A if the

disc fluctuation angle, A6, is known. The best available data for disc fluctuation angle

under a variety of upstream flow disturbances and their proximities were developed under

Phase I of this program and are reported in Reference 1.

Pipe Eddy Freency Approach

This estimate of disc velocity is based on the eddy frequencies observed in fully developed

turbulent flow in a pipe (Refs. 17, 18, 19). Since disc oscillations in water (or other liquids)

were found to be highly damped in our Phase I tests, the disc has a tendency to respond to the

disturbing frequencies present in the turbulent flow stream. The disc velocity based on its

response at this frequency is given by:

Vdisc = 0.2513 VRA (7) | ~~~~Dpipe}

Again, the impact force can be directly calculated using the method described in Appendix

A.

Mean and 3ca Disc Speed Approach

The disc motion data gathered during the Phase I portion of this program was extensively

reviewed to ascertain the characteristic disc oscillation frequencies for the many tests

performed. However, this analysis showed that only a small percentage of test cases had

an energy peak near a dominant frequency. Typically, the oscillation energy was spread

over a range of frequencies up to approximately 6 hz. Therefore it appears that the disc does

not behave as a simple single degree of freedom system that would respond at its single

natural frequency, but is able to respond over a range of frequencies contained in the input

flow stream. Considering the chaotic nature of the turbulent flow stream, this response

seems reasonable.

In order to characterize the disc motion and velocity in some usable way, a numerical

technique was developed which considered the randomness of the disc response. The result

is a parameter we call the Mean Disc Speed which is, in effect, the average disc angular

velocity taken over a time long enough to average out instantaneous variations which

would otherwise skew the velocity in one direction or another. When fatigue calculations

are to be performed, the maximum velocity attained prior to impact is of particular interest

because the highest velocity impacts, even though relatively infrequent, will account for the

majority of fatigue damage. To this end, a second parameter is calculated by taking the

standard deviation of the disc velocity data. When multiplied by three, we have the 3a disc

speed which will bound the highest 0.3 percent of the disc velocities. A discussion of how

these parameters are calculated is contained in Appendix D. Also contained in this

appendix is a complete set of the mean and 3a disc speed plots derived from the Phase I work

for both the three- and six-inch valves.

-24 -
IMPACT AND FATIGUE TESTING

In order to measure the impact forces generated when disc tapping occurs, a test program

was devised in which a specially modified check valve could be subjected to a variety of

flow conditions known to cause the disc to tap against the full open stop. By instrumenting

the valve disc with strain gages, actual disc stud impact forces could be captured in real

time and compared to theoretical predictions.

General Methodlovy

The six-inch valve was installed in the flow loop and tested with two different upstream

disturbances. An elbow was used to generate low levels of turbulence and a perforated plate

was used for high levels of turbulence. Tests were performed with the disturbances at two

proximities and flow velocities were varied over a range from 2 to 18 feet per second.

During the tests, data from the instruments were recorded on a wide-band FM tape recorder

for subsequent review and signal processing.

Test Valve

Modifications were made to the six-inch swing check valve to permit direct measurement

of disc stud impact loads. This was accomplished by building a special disc with

provisions for mounting strain gages in a machined recess in the disc (Fig. 11). A thin

diaphragm (0.125-inch thick) was created so that mechanical amplification of the strain

caused by impact would produce adequate strain signals. Extensive finite element

analyses were conducted to optimize the load measuring characteristics of the disc and to

provide a relatively constant strain field at the gage locations.

THIN DIAPHRAGM

STRAIN GAUGES

SILICONE RUBBER SEALANT

ALUMINUM COVER PLATE

Figure 11

Instrumented Disc Cross Section

The disc was instrumented with three separate strain gage bridges as shown in Fig. 12.

Eight gages were mounted in the recessed cavity, four of which were used in a full bridge

-25 -
configuration for axial load measurement. Two additional full bridges, each with two

active arms, were built to measure bending loads along two orthogonal axes (Figs. 13, 14).

The gages were protected from moisture penetration by various epoxy and silicone

coatings. The remaining cavity volume was then filled with a special silicone potting

compound, covered with an aluminum plate, and again sealed from the environment with

an adhesive coating (Fig. 15).

3 7

R c

BENDING #2

HALF BRIDGE

BENDING #1

HALF BRIDGE

AXIAL

FULL BRIDGE

Figure 12

Strain Gage Bridge Schematics

Figure 13

Strain Gaged Check Valve Disc

-26 -
Figure 14

Detail of Strain Gage Installation

\SEALED WIRE EXIT

l STRAIN GAUGES

SILICONE RUBBER SEALANT

ALUMINUM COVER PLATE

Figure 15

Disc Showing Sealed Wire Exit

Wires from the gages were routed through a drilled passage to the top of the disc and routed

along the hinge arm with appropriate strain relief. The wire exit was also thoroughly

potted in silicone adhesive and potting compounds to preclude moisture penetration (Fig.

16). Finally, the wires were terminated in a special glass sealed hermetic connector

mounted on the valve bonnet which served as part of the valve body pressure boundary as

well as an exit to the outside world for the strain signals (Fig. 17).

-27 -
-i . 10

Figure 16

Wire Exit from Disc

EXTENSION TUBE

WIRE BUNDLE

FROM STRAIN GAGES

Figure 17

Completion of Pressure Boundary Using Hermetic Connector

-28 -
Calibration of Strain Gaged Valve Disc

Each of the three strain gage bridges was calibrated before starting the impact tests. The

bridges were found to respond in a very linear fashion, and the amplified axial bridge

output was 2,900 lb/volt. During the course of testing, the bending bridges were damaged

and became unusable. A check of the axial bridge found it to be operating satisfactorily

with no detectable change in output scale factor. After completing the test series, the axial

bridge output was found to have decreased somewhat to 2,288 lbs/volt; but the bridge

behavior remained very linear and repeatable. Since the shift in output scale factor could

not be traced to a particular point in time, all data reduction has used an average of these

2,900+ 2,288

two values, or 2 = 2,594 lbs/volt. Therefore, the maximum error in the force

measurements reported here is approximately bounded by ± 12 percent.

Lnstrumentation

Much of the instrumentation used in the course of the impact testing was the same as that

used in the Phase I research (Ref. 1). The additional equipment used consisted of the

following:

1. Strain gage amplifiers and signal conditioners,

2. Seven-channel FM tape deck,

3. Two-channel spectrum analyzer,

4. Two-channel digital storage oscilloscope,

5. Digital signal processing and filtering software.

An important aspect of the force measurement instrumentation was to ensure adequate

instrument frequency response for capturing the impact load signatures. Measurements

of impacts made during the Phase I effort showed that the load event duration was

approximately three milliseconds. Since a stiffer steel bonnet was being used in the Phase

II tests than the one made of polycarbonate used in tHe Phase I tests, even shorter impact

durations could be expected. Therefore, the strain gage amplifiers and tape deck used to

record these events were capable of measuring signals from DC to 10 khz ± 0.5 db.

-29 -
TEST RESULTS

The impact tests were divided into two areas of investigation. First, the impact load

signature was characterized in terms of its amplitude, duration, and shape. This

information, gathered for different upstream disturbances and flow velocities, would be

used for correlation to the impact force prediction techniques. Second, the frequency at

which these impacts happen was determined. The number of impacts of a given magnitude

that occur over the "tapping zone," which is defined as the range of flow velocities over

which tapping takes place, was determined for a very large number of impact events. This

information combined with the data about impact magnitudes provides the basis for

performing disc stud fatigue calculations.

Typical Impact Signature

Shown in Figs. 18 to 20 are impact load signatures generated with the strain-gaged disc

assembly. Figures 18 and 19 are impacts recorded with the high turbulence source and

elbow located 1.5 diameters and 3 diameters upstream of the valve respectively. Figure 20

is also for the high turbulence source, but includes additional modifications to the disc and

hinge arm assembly to be discussed later. The plots show that the characteristic shape of

the impact signal is independent of the type and proximity of the upstream disturbance as

well as the flow velocity. The principal difference among the signals is the amplitude,

which depends on the velocity of the disc at the time of impact. The plots also show that the

impact excites various resonant frequencies in the load measuring disc. These

frequencies are superimposed on the basic impact load signature and, because they are

relatively high in magnitude compared to the impact force itself, result in a low signal-tonoise ratio. In addition, the resonant frequencies are of the same order as the frequency of

the impact event itself, further heightening the difficulty of extracting the signal of

interest.

Before we could attempt to filter out the high frequency information from the impact force

signals, the resonant frequencies had to be identified. A separate series of controlled

impact tests were performed to determine the natural frequencies of the load measuring

disc. The measured natural frequencies are shown in Table 5:

Test Case | T fl I 2

Disc Assembly in Air | I WO _

Disc Assembly in Water | 3875 1875

Table 5

Dlsc Assembly Resonant Frequencles, hz

-30 -
Turbulence Plate QIJ5D (Unmodiiled Disk)

I

I

I Figure 18

Impact Signature, Upstream Turbulence Source

Figure 19

Impact Signature, Upstream Elbow

Figure 20

Impact Signature, Upstream Turbulence

(Modifled Disc)

Time, seconds

Elbow Up 0 3D (Unmodified Disk)

A

I

Tme, seconds

Turbulence Plate 1.50 (Modified Disk)

A

I

Time, seconds

-31 -
By performing the tests in and out of water, we were able to quantify the added mass effect

of the water, which increases the effective weight of the dynamic portion of the strain gaged

assembly by 14 percent. Spectral analysis of the test data also confirmed our qualitative

observation that the impact energy was concentrated below 1000 hz. Since the disc

assembly natural frequencies are higher than the effective frequency of the impact event, a

digital low pass filter was implemented to remove these unwanted frequencies from the

impact force signal. The result is shown in Figure 21 where the original and filtered

signals are shown superimposed. The remaining impulse is now clearly visible and the

amplitude and duration can be easily extracted.

Discussion of Impact and Disc Motion Signals

Examples of an impact signal (filtered and unfiltered) and the corresponding disc motion

signal are shown in figures 21 and 22. The peak force can be measured directly from the

plot as 950 pounds. Since both signals were recorded on tape simultaneously, the disc

excursion that resulted in the impact event can be identified and the average velocity at

impact determined as 10.5 inches per second. This velocity is used to calculate the kinetic

energy of the disc at impact which is then converted to the peak impact force for later

comparison.

Turbuience Plate @1.50 (Modifled Disk)

Time, seconds

Figure 21

Raw and Filtered Impact Force Signatures

-32 -
Turbulence Plto Os1.5D (Modiflid Disk)

__ _

I

9~

If

w.~~A A /l A A

?.,. C-I \ 41 I 

40.1

.-*. *C -I - 0; C 4 C *, 0. VI *

Time, seconds

Figure 22

Disc Displacement Plot Corresponding to Impact Event of Figure 21

Discussion of Disc Motion Measurements and Modified Disc

The technique used to measure disc motion is illustrated schematically in Figure 23. As

can be seen, the physical connection from the displacement transducer is made to the hinge

arm which moves with the disc in response to the fluid flow. While this system has served

very well in previous test programs, certain limitations became apparent in this project.

The important variable in the force calculation is the impact velocity of the disc. Since the

measurement connection is actually made to the hinge arm, the disc velocity is, in effect,

inferred from the motion (hence, velocity) of the hinge arm. The clearance, C1, between

the disc stud and hinge arm allows the disc a small but significant amount of freedom

independent of the hinge arm. To eliminate this problem, the disc was modified by rigidly

pinning it to the hinge arm so that the two behaved as a single entity. While this helped the

situation, impact force calculations, based on instantaneous hinge arm motion, remained

in disagreement with the measurements.

We believe that, although minimized as far as practicable, the remaining clearances

shown in the illustration, namely those between the hinge arm and pin, C2, and clevis and

hinge arm, C3, still prevent the LVDT from following the disc motion to the degree of

accuracy needed for these measurements to provide instantaneous correlation between

velocity and impact.

This can be quantified as follows. The measured clearance, C3, is 0.005-inch. At disc

velocity, V, of 3 in/sec, the time required to traverse clearance C3 is

t=C= 0-005 inch =0.0017 seconds

V 3in1sec

-33 -
Figure 23

Impact Measurement Assembly

This length of time is almost twice the duration of the impact events themselves.

Therefore, the actual disc impact force measured by the disc mounted strain gages could

easily occur out of complete time synchronization with the hinge arm motion/velocity

measurement. The only way to eliminate this problem and get true disc motion

measurement would be to attach an accelerometer directly to the disc stud and use this

signal as the basis for disc motion and velocity. This was not easily achievable; therefore,

emphasis was put on comparing measurements with the upper-bound predictions discussed

earlier.

Use of SaDisc Speed For Disc Impact Force Estimates

The problems associated with determining the instantaneous disc impact velocity from the

disc motion signals caused us to reexamine the 3a disc speed data presented in Appendix D

as an alternate measure of the disc velocity. The 3a disc speed, which has proven a good

measure of the maximum disc velocity, should establish an upper bound on the impact

velocity, and therefore on the calculated impact force. In reviewing the data presented in

Appendix D, one can see that as the flow velocity increases through the tapping zone, the

disc speed begins to decrease. For the purpose of calculating the bounding values, the

maximum 3a disc speed has been used and no credit taken for any decrease. This

measure of disc velocity has been taken from data collected during the Phase I portion of

this work and reduced as part of the Phase II program.

-34 -
Rate of Occurence of Impacts

In order to perform a fatigue analysis of the disc stud, the frequency of occurrence of

impacts of a given magnitude must be known. To quantify this factor, the six-inch valve

flow test data was surveyed and the number of impacts above a given threshold force were

counted and recorded. The test data were divided into the impact force ranges presented in

Table 6:

Range E1b 3d High lurbutence

m EhwpS Source@1.5d

A 130-2601bs 260-5201bs

B 260-3251bs 520-780lbs

C 325 -3901bs 780- 1040lbs

D 390 - 5201bs 1040 - 1170Ibs

E 520 - 650 1bs 1170 - 15756 bs

F 650 - 7801bs 1575 - 19501bs

G 780- 10401lbs 1950-20801bs

Table 6

Impact Force Ranges

(After Low-Pass FllterIng)

At forces below the lowest incremental level (Range A), the impact frequency was so high

(about four impacts per second) that manually counting them was unfeasible. Since the

stresses developed in the stud at these lower force levels are well below the endurance limit

of the stud material, their contribution to cumulative fatigue damage can be neglected. The

method used to identify these impacts used the raw, or un-filtered strain gage signals

containing the high amplitude ringing due to resonant excitation. For comparison to the

impact force data presented earlier, which had been processed with the digital low pass

filter to eliminate the ringing effects, the forces reported in this section have been reduced

by a factor of two, which was the typical reduction observed after filtering the raw impact

force signals.

Data from the high turbulence and elbow up surveys is presented in Figures 24 and 25. In

both sets of data, as the impact force decreases, the number of impacts rises quickly. This

agrees with observed valve behavior where the statistical occurrence of low velocity

excursions that would result in low impact forces is greater than high velocity excursions

that would create high impact forces. Another observation is that the number of impacts

that occur at a given load range is much greater for the high turbulence source than for the

elbow up. For example, impact forces between 130 and 260 pounds occur at a rate of about 0.7

impacts per second with the elbow disturbance, whereas the turbulence source generates

forces between 1,040 to 1,170 pounds at roughly the same rate. This is in keeping with the

data gathered during Phase I where the amplitude of disc motion generated with the

turbulence source was much greater than with the elbow.

-35 -
High Turbulence Source @ 1.5D

>5 Hz

5

0

a 4

a 3-

a

E 2-

8E 1

3

a

0-

A. 260-520 D: 1040-1170 G: 1850-2080

B: 520-780 E: 1170-1580

C: 780-1040 F: 1580-1850

N Unmodified Disc

El Modified Disc

1

.A WINEM _-,, ~~~~~~ -,, --,, ~ 

A B C D E F G

Figure 24

Impact Force Survey

(Based on Filtered Signal, Ibs)

Elbow Up @3D

n 3i -.

la

0

co

I 0.6-

CL0

4 0.2

z

A:

B3:

C:

130-260

260-390

390-520

D: 520-590

E: 590-790

F: 790-930

G: 930-1040

* Unmodified Disc

0.0 - -_ .

I - -- --. T T

A B C D E F G

Figure 25

Impact Force Survey

(Based on Filtered Signal, Ibs)

-36 -
CORRELATION OF IMPACT TEST DATA TO THEORY

A total of twelve impact events have been analyzed for comparison to theoretical

predictions of impact force. The data is presented in Table 7. All tests were performed with

the modified disc assembly previously described. The upstream disturbance used was the

high turbulence source located 1.5 diameters upstream. Flow velocities ranged from 6 fps

(start of tapping) to 18 fps. It is important to note that cavitation begins at approximately

10 fps and increases in intensity to 18 fps. Cavitation will increase the level of turbulence

as well as create higher frequency disturbances in the flow stream to which the disc may

respond.

Kinetic Energy Conversion Approach

Using the 3G disc speed to estimate the maximum disc impact velocity, the conversion to

kinetic energy yields predicted impact forces that are within 80 percent of the measured

loads up to a flow velocity of 9.5 fps. Beyond this flow velocity, where cavitation begins, the

difference grows larger, reaching about 250 percent at 14 fps. Beyond this point, the level of

cavitation is quite high, and the predictive technique falls short of the measured loads by a

large margin.

Disc Oscillation Amplitude and Frequency Approach

A second predictive technique involves use of the disc natural frequency estimation

method along with the appropriate disc oscillation amplitude data gathered during Phase I

of this project. (Refer to Appendices A and C for the methodology used.)

Impact Force Calculations for Six-Inch Instrumented Disc

Flow Condition : Water @ 8 f/s - 709F

Upstream Disturbance Perforated plate at 1.5D upstream

Disc Oscillation Angle 160 before tapping

Disc Equilibrium Angle 630

Disc Natural Frequency:

fn = 1 Ksif

W/g

1 2KApV2 Z

RW

where

K = 2.0

A = 6.9 x40.26ft2

p = 62.4 lb/ 3

-37 -
V2 = 8 ft/sec

R = 5-04=0.42ft

12

W = Wdi. + 2 Whinge +p3 p disc

=10.19+ -x62.4x (619 3 k12)

= 14.14 lb

Z= cosO sin 0+ cossin9

22 c2 21

= cos 270 sin 13.5° + I cos 13.5 sin 270

= 0.429

A fl]1 12x2x0.26x62.4 x82X0.429

*--~~ - In 0. 42 x 14.14

-2.757hz v

Maximum Disc Velocity at

2.757 hz and 16° Amplitude (peak to peak)

Vmn, = R x0x 2 7if

2

=5.04x8x X .x2nx2.757

180

= 12.19 in / sec

Maximum Disc Inpact Force

Fmax = 162.4 Vmax

= 1,980 lb

The highest force measured in the tapping zone was 1,276 lbs. As can be seen by examining

the measured impact loads in Table 7, the theoretical value of 1,980 lbs bounds this value by

a comfortable margin.

-38 -
Flowv Average 3 Sigma Ftheorefical Fieasured Ratio of

Velocity Velocity at cSpeed based on FIaJsured

Inpact 3 Sigma to

Disc Speed Ftheorefical

filsec degisec deglsec lbs lbs

6.0 7.8 78 1,063 916 0.86

7.9 19.5 8) 818 998 L22

8.5 13.2 55 750 920 1.23

9.0 16.4 50 682 1,276 1.87

9.5 12.3 45 613 1,098 1.79

10.0 18.8 40 545 1,407 2.58(2)

1L0 5.9 33 450 1,148 2.55(2)

12.0 12.2 25 560 894 1.60(2)

13.0 14.7 20 341 792 2.32(2).

14.1 5.6 15 204 971 4.75(2)

16.0 4.7 7 96 795 8.28(2)

17.1 3.6 6 32 530 6.50(2)

(1) Modified disc is rigidly connected to the hinge arm to eliminate excess clearance

(2) This data affected by onset of cavitation

Table 7

Comparison of Theoretical and Measured Impact Forces

6-inch Modified Disc Assembly (1)

High Turbulence Source @ 1.5 Diameters Upstream

-39 -
IMPACT FORCE AND FATIGUE PREDICTION METHODOLOGY -

SUMIARYAND CONCLUSIONS

The overall results and conclusions from the six-inch swing check valve impact and

fatigue prediction research are summarized below:

1. Impact forces as high as 1,276 pounds can occur when the valve is operating at flow

velocities in the tapping zone. Testing shows that, as the disc enters the tapping zone,

the impact forces increase in magnitude to a maximum and then diminish as the

flow velocity increases.

2. Measured impact forces are bounded by predicted values based on the disc frequency

and oscillation amplitude method discussed earlier in this section.

3. Impact forces calculated using the disc velocity inferred from hinge arm motion

were within 80 percent of the theoretical values in the flow range of 6.0 ft/sec to 9.5

ft/sec where other factors such as cavitation were not present.

4. To make valid comparisons of predicted versus measured impact forces, an accurate

measurement of the instantaneous disc motion (instead of the hinge arm motion) is

required.

5. Experimental measurement of the number of impacts occurring at a given load is

bounded by predictions made using the disc natural frequency method.

6. Higher amplitude impacts occur less frequently than lower level impacts.

7. As discussed later in Plant Correlation Example 4, the techniques described in this

report adequately predict actual disc-stud fatigue failure.

-40 -
CORRELATION AGAINST PLANT DATA

As discussed in detail in other sections of this report, wear and fatigue/impact predictive

models were found to be in good agreement with the laboratory tests conducted under Phase

II research. This section presents the correlation between model predictions and actual

plant test data for some important, well-documented cases of premature degradation or

failure of check valves. Also included are some examples of correlations where relatively

low degradation was predicted for valves used in long-term service with the disc not fully

open. Separate correlation examples are given for failures of each type: those resulting

from excessive hinge pin wear and one caused by fatigue of the disc stud nut. The fatigue

example is the San Onofre Unit 1 check valve failure event in November, 1985 that caused

the NRC and the nuclear industry to focus their attention on check valve problems (Refs. 5,

6, 7).

It should be pointed out that this part of the project was limited in scope because specific

tasks in our original proposal to obtain detailed performance data from several operating

plants under the NRC (SBIR Phase II) research could not be funded. Despite this, our

literature search, review of previously published plant data, and our involvement with

utilities in root cause analyses of check valve failures provided us with some excellent

examples against which the validity of the predictive methodology could be established. As

additional plant data with sufficient details are obtained in the future, further validation

and refinements of the predictive model will be made.

Plant Correlation Example 1:

18-Inch Swing Check Valves with Excessive Hinge Pin Bushing Wear

This is one of the best cases for wear correlation against the predictive model because all

the pertinent valve design data, installation data, dimensional inspections, and operating

and maintenance history records for a large population of valves for many years of

operation are well documented (Refs. 15, 16).

Sixteen 18-inch, Class 900 swing check valves (CV-3) installed in the primary loop of the

core discharge system of the pressurized water reactor operated by UNC Nuclear

Industries, Inc. for the U.S. Department of Energy near Richland, Washington were

found to have high hinge pin bushing wear. Excessive wear of the hinge pin bushings had

caused some of the bushings to completely wear through, and the debris of the worn-through

bushings for some valves were found downstream. Figures 26 and 27 show installation

details and the typical hinge pin bushing wear observed. These valves were installed

within two feet downstream of a side discharge dead-end line tee. The average wear rate

was found to be 0.022"/year for these valves based on the first detailed inspection in 1973

after approximately eight years of operation. In a subsequent surveillance inspection

program, the wear rates were more precisely monitored and found to be 0.027"Iyear of

service. At this rate, the 3/16-inch thick bushing wears through in about seven years,

necessitating massive change-out of components. It should be noted that both the hinge pin

and the bushing are made of cobalt-based alloys (Haynes Stellite #25 and Stoody #6

respectively). These materials are known to have relatively good wear characteristics.

-41-
~~~~~~~ J~~~4 s

CV-3 Checkt 114/va-

/8 w 0 33*A43

004/A elJA4C d0w6. #/8/

I

-6

- r

0.4 2

E

r,

a

AM O IV

Figure 26

18" Swing Check Valve Installation Used In Example 1
Figure 27

Hlnge Pin Bushing Wear In 18" Valves

(Example 1)

Appendix B contains the pertinent data and analysis details for these valves based on the

predictive models developed under this NRC-sponsored research. The results show that the

minimum velocity required to fully open the valve without any upstream disturbance is 19

ftlsec; however, when the effect of the upstream disturbance is accounted for as described in

the Phase I report, the minimum velocity to fully open goes up to 22.8 ft/sec. The actual flow

velocity during normal operation was 20 ft/sec, thus showing that the valve disc will be

oscillating in the tapping zone. Since the valve disc and hinge arm are of an integral cast

design that eliminates the disc stud connection to the hinge arm, the typically fatigueprone area is not present in this design, and no fatigue problems were found.

The upstream flow disturbance is a tee connection with one end dead-end on the main run.

The intensity of this flow disturbance is estimated to be more severe than an elbow, but less

severe than the high turbulence source. This gives a (3f) disc fluctuation angle of

approximately eight degrees, based on Phase I research. The natural frequency of disc

oscillation is calculated to be 2.0 hz based on Appendix A approach. The overall sliding

motion is shared by the two bushings between the disc and shaft, and two bushings between

the support trunnions and shaft, within which the shaft is free to rotate.

Because the bushing was made of a cast Stoody-6 (Co-CRW alloy) and the shaft was made

of wrought Haynes Stellite-25 (Co-Cr-W-Ni alloy), the sliding pair of metals can be

classified as compatible. The value of the wear coefficient for this combination for water at

high temperature is estimated to fall between 1 x 10-4 (for poor lubrication) and 5 x 104 (for

no lubrication). The average (on a logarithm scale) of these two values is 2.24 x 104, which

is our best estimate of the applicable wear coefficient in lieu of actual tests on the exact

material combination.

The theoretical wear rate prediction based on this value of K is calculated to be 0.0335-inch

per year, compared to the measured wear rate of 0.022 to 0.027 inch/year. Based on our

predictive methodology, the 3116-inch thick bushing will be expected to wear through

-43 -
completely in a little less than six years of operation. This is in excellent agreement with

the actual experience which caused a number of the bushings to wear through in eight years

of operation.

Plant Correlation Example 2:

18-Inch Swing Check Valves with Low Hinge Pin Bushing Wear

This is an extension of Example 1 because the only difference between the two is that these

sixteen 18-inch valves (CV-5) of identical design were installed in the core inlet side at the

end of a long run of approximately 70 feet of pipe. The operating flow conditions and their

durations are the same as for the valves in Example 1.

In the absence of an upstream flow disturbance, the disc fluctuations are predicted to be in

the range of baseline 3a disc fluctuation of 1.5 to 2 degrees (instead of the 8 degrees for

Example 1) as shown in the Phase I report.

These low levels of disc oscillations are predicted to result in bushing wear rates of .006-

inch to .008-inch per year, which corresponds to a bushing wear-out life of 22.8 to 31.2 years.

This correlates well with actual wear rates, which were reported to be "negligible based on

eight years of actual operation for these valves in Ref. 16.

Plant Correlation Example &:

4-Inch Tilting Disc Check Valve Modifications to Reduce Hinge Pin Wear

Severe hinge pin wear was observed in two four-inch, Class 600 tilting disc check valves

used in the auxiliary feedwater turbine steam supply system after one year of operation at

San Onofre Unit 2 in 1986 (Ref. 8). The valves are normally closed, but the system had

been recently modified to permit a small flow of 600 to 800 lb/hr of steam through these

valves to avoid condensation and accumulation of water in the downstream piping. In one

year of operation, the 0.5-inch diameter 410 stainless steel hinge pins had almost

completely worn through due to the disc oscillation caused by the low rate of steam flow

(Fig. 28). In this case, the 410 SS hinge pins were sliding against the bushings made from

Stellite 6 material.

WORN CONTACT SURFACE

ORIGINAL SHAFTl

Figure 28

Severe Hinge Pin Wear In 4-Inch Tilting Disc Check Valve

-44 -
Wear analysis for' this valve was performed using the pendulum frequency approach

because the disc is'oscillating near the closed position due to low flow conditions. Disc

oscillation frequency was calculated to be 2.9 hz, and mean disc oscillation angle of * 10

degrees was used in wear analysis. The upper bound of disc oscillations found in the

Phase I tests with water was around 16 degrees, and somewhat higher values are

anticipated in steam due to lower damping. Wear coefficient for the 410 stainless steel (FeCr-Ni alloy) hinge pin and Stellite-6 (Co-Cr-Ni alloy) bushing, a compatible metal sliding

pair in steam (unlubricated surfaces) is estimated to be 5.0 x 104. The hardness of the pin

was approximately 275 Brinell.

Analysis predicted severe wear of 0.36-inch in one year of operation, which is more than 70

percent of the hinge pin diameter. Under such high wear conditions, wear depth rate

increases rapidly after half the hinge pin diameter is worn due to reduction in the bushing

area. These analytical estimates were found to be in agreement with actual inspection

results from the two valves.

Even though absolute wear predictions based on assumptions made in the analysis should

be considered accurate only within a factor of two or three (Refs. 10, 11), relative

improvements by making material modifications can be predicted more accurately.

Modifications were made in the hinge pin area by changing to Stellite-6 vs. Stellite-6

materials, which was predicted to reduce the wear rate to between 0.007-inch and 0.030-

inch per year. Actual measurements made on the modified arrangement after five

months of operation showed a wear rate of 0.010"/year, which is in good agreement with the

predicted improvement, and about 30 times less than the original design!

Plant Correlation Example 4:

10-Inch Swing Check Valve Disc Stud Fatigue Failure

At San Onofre Nuclear Generating Site Unit 1, a water hammer event occurred in the

pressurized water reactor horizontal feedwater line caused by failed check valves in

November 1985 (Ref. 7). The disc stud/nut connection in these check valves had fractured

due to repeated impact against the open stop, which had allowed the disc to separate from the

hinge prior to this water hammer event. These valves had operated satisfactorily for

several years at full power conditions. Fifteen months prior to this event, the plant was

operated at approximately 85 percent reduced power. This resulted in insufficient flow

velocity to fully open the disc, causing disc tapping and eventually fatigue failure of the

threaded connection. Root cause analysis of this failure is documented in earlier reports

(Refs. 6, 7); however, the root cause analysis did not attempt to quantify the fatigue

failures.

Using the predictive fatigue/impact model developed under this research, this failure was

reexamined quantitatively from a fatigue standpoint. The details of the analysis approach

are contained in Appendix C. The important results from this analysis are presented here.

The minimum flow velocity to fully open the valve with no upstream disturbance is

calculated to be 14.6 ft/sec. When adjusted for the upstream disturbance effect of the severe

turbulence source (control valve) that is present immediately upstream, the minimum flow

velocity requirements increase to approximately 17.5 ft/sec to 20.4 ft/sec. The actual flow

velocity at full power is 18.5 ft/sec which is in the same range as the minimum flow

-45 -
requirements. Therefore, the disc is estimated to be only marginally seated even at full

power. At 85 percent power, the actual flow velocity is 16.5 ft/see, which is just below the

minimum requirements. Therefore, severe disc tapping is predicted to occur in this zone

of operation.

Disc oscillation magnitude for this installation, which has a severe turbulence source

immediately upstream of the check valve, is estimated to be 16 degrees based on Phase I

research (Ref 1). The disc oscillation frequencies, based on free body flow equilibrium

method and the eddy frequency method, are calculated to be 2.65 hz and 1.58 hz. The high

amplitude motion and impacts are most likely to occur near the lower of the two

frequencies, based on comparisons between theory and actual tests on the three- and sixinch valves tested under Phase II. The peak stress at the base of the stud due to this level of

disc motion is estimated to be 24 ksi, which falls in the low cycle fatigue range for the

material used. Fatigue analysis results show that there is a very high propensity to failure

due to disc stud fracture, and the analytical predictions are in the same range as the

observed failure life of a few months when operated in this flow velocity range.

-46 -
SUMMARY OF CHECK VALVE REVIEW METHODOLOGY

As stated in the introduction of this report, one of the major goals of the Phase II research

was to develop a method to predict the behavior of swing and tilt disc check valves.

Specifically, we are interested in predicting several common modes of check valve

degradation so that potentially troublesome installations can be identified before the valve

causes a plant problem. We believe the techniques described here, combined with the large

amount of data presented in Reference 1, go a long way toward meeting this goal. In this

section, we will present a summary of the method to be followed in predicting check valve

degradation. We believe the predictive methodology outlined below is of particular merit

both because of what it does do and, just as importantly, for what it does not do. The ability

to identify and quantify, in advance, a check valve's propensity for accelerated

degradation should be the aim of any review and analysis technique. The methodology

described here does this.

Many check valve reviews depend on determining whether the Vmin requirements are met

for each valve application. It is very easy to create a maintenance nightmare if open-andinspect decisions are based solely on the results of Vmin calculations since 70 percent to 80

percent of all check valves will fail this single criterion and will end up with

recommendations to be opened and inspected. This is what our methodology does not do.

Overview of Check Valve Review and Analysis Procedure

1. Calculate minimum velocity, Vmin, requirements (Ref 3).

2. Determine Cup factor and calculate modified Vmin (Ref. 1).

3. Do not get discouraged when 70 percent to 80 percent of valves fail to meet Vmin

requirements. Most will prove to be in satisfactory condition.

4. Calculate hinge pin wear and, if necessary, disc stud fatigue potentials using the

predictive models described in this report.

5. Disposition the valve based upon analysis results and any available maintenance

records.

These steps will be reviewed in more detail in the following sections.

Vmin Calculation

The important first step in the review process is determination of the valve minimum

velocity requirements. In the absence of actual test data on the specific value in question,

the Vmini formula described in Ref. 3 should be used.

This formula has been proven many times to provide an excellent estimate of Vmin

requirements for many swing and tilt disc check valves if certain constraints are

followed. Figure 29 presents the results of seventeen valve tests compared with their Vmin

predictions calculated using the EPRI guideline formula.

.47 -
MINIMUM VELOCITY FORMULA

COMPARISON AGAINST TEST RESULTS

M.

0

0.

a

cc0L

w

0 5 10 15 20

TEST DATA, FPS

Figure 29

Minimum Velocity Formula Comparison

Against Test Results

As can be seen, there is excellent correlation between the measured and predicted values,

except for two data points. These two points illustrate the first important constraint of the

Vmin formula. As previously noted in Ref. 3, when the disc projection, Y, into the flow

stream is less than about 25 percent of the valve size, d, the predicted results will begin to

deviate from test results. This Y/d ratio is illustrated in Figure 30. Both valves have a disc

angle, 0, with respect to the flow stream of 65 degrees; however, Valve A has a Y/d of 25

percent and Valve B has a Y/d of almost zero. Valve B is typical of what are generally

classified as clearway swing check valves, and the Vmin formula will underestimate the

flow velocity requirements by as much as 50 percent for valves of this type. In Figure 29,

the two data points labeled three-inch and six-inch are, in fact, clearway swing checks

with Y/d ratios of zero and 0.14 respectively. While the fluid impingement angle, 4, is an

important factor in the Vmin calculation, it must be recognized that the Y/d ratio is an

independent factor which can significantly affect the accuracy of the Vmin estimate. If this

important fact is recognized, then judicious adjustments can be made to estimate a more

realistic Vmin for a clearway valve. Refinement of the Vmin formula is possible to take

-48 -
Geometry A, YID = 0.25

Geometry B, YID a 0

Notes:

(1) Vmn and disc stability are seriously affected by independent parameter "Y

(2) When Y/D c .25, actual Vmin may be much higher than predicted by generaglzed

VmIn equation, and disc stability may not be achievable with velocity increase

Figure 30

Significance of Disc Projection (Y/D) on Performance

-49 -
into account the Y/d ratio. Preliminary work has shown the validity of an approach

presently under development. However, systematic testing is needed to properly quantify

the effect of this important factor.

In addition to the Y/d ratio, care must be exercised in the following areas when applying

the Vmin formula:

1. Fluid impingement angle, 0, and seat angle, at

Most valves have an inclined seat with angle, a. Some will have an angle as high as

30 degrees or more. In most cases this will not affect the net fluid impingement

angle, O. A good illustration of this is in Plant Correlation Example 1, discussed

earlier. These valves have seat inclination angles of 20 degrees, yet the inlet of the

valve is straight and horizontal. The seat angle by itself will not alter the flow

direction through the valve to any appreciable extent. Therefore, the fluid

impingement angle will equal f3, the angle the disc makes with the pipe axis when

fully open. Accounting for the full value of a in the net impingement angle can

result in grossly underestimating the flow velocity requirements of a valve. For

those valves which have both an inclined seat and a deliberately inclined inlet

geometry leading into the seat, some contribution to the net impingement angle can

be taken into account. Typically, this additional amount is in the area of 3 to 10

degrees.

2. Reduced Inlet Port Geometry

Related very often to the previous discussion are concerns about valves with reduced

inlet ports. This design feature is usually incorporated in an attempt to increase the

flow velocity through the seat and, in effect, reduce the Vmin requirements based on

the flow velocity in the (larger) piping system. Unfortunately, there are often

detrimental side effects which accompany the velocity increase. For example, the

fluid often undergoes a severe directional change while passing through the valve or,

after passing through the reduced seat area, expands abruptly upon entering the valve

body. In these cases, one has in effect created a self-contained flow disturbance that

may very well eliminate any beneficial effect of the increased seat velocity. In these

situations, a valve which has only straight pipe for 10d upstream and would otherwise

be treated as having no upstream disturbance, might best be modeled for Vmin

purposes as having a low turbulence source at Od.

3. Effective Disc Diameter

As recommended in Ref. 3, the disc diameter used to calculate the disc flow

impingement area, A, should not exceed 110 percent of the valve seat diameter.

Significant errors will often result if the disc diameter is assumed to be equal to the

basic valve size.

Calculation of Modified Vmin using Cup

After determining the generalized minimum velocity, the effects of any upstream

disturbances must be accounted for. These disturbances will always serve to increase the

Vmin requirements above those calculated using the basic Vmin formula. A table of Vmin

modification factors for upstream disturbances is included in Reference 1. These were

determined from the extensive set of tests performed as part of that research. While every

-50-
conceivable disturbance type cannot be covered, engineering judgment can be used to

make sound and conservative judgments based on the test data at hand.

Having completed this Vmin calculation, it is not uncommon to find that 70 percent or more

of all check valves in nuclear power plants systems do not meet VmIn requirements.

Technically these valves are misapplied, yet it is clear from industry records that serious

problems do not exist with this large population of valves. Only a small percentage of the

so-called misapplied valves suffer from accelerated degradation. The problem for the

utility engineer becomes one of determining which of these misapplied check valves will

turn out to be the truly "bad apple" likely to be suffering from accelerated wear and

degradation. Knowledge of the valve Vmin is necessary; but it is insufficient information

to progress further. The following steps must 'be completed before making this

determination.

Wear and Fatigue Calculations

The techniques developed in this Phase II work form the basis for predicting two of the most

important check valve degradation processes: hinge pin wear and disc stud fatigue. Only

after these calculations have been performed can one make an informed decision with

respect to the need to-open and inspect a valve, and with what frequency. Our experience

has shown that by applying these techniques, using conservative estimates where

information may be lacking (valve dimensions, typically), the majority of valves which

do not meet Vmin requirements are found to have very low predicted hinge pin wear rates

and, often, no predicted disc stud fatigue problems. These valves are operating under suboptimal conditions and their behavior must be quantified unless a massive plant-wide

open-and-inspect program is desired. This is where the Phase II work can be used to

predict hinge pin wear and disc stud fatigue. As shown in the plant correlation examples,

considerable engineering judgment must be exercised in certain areas. Nevertheless, the

methodology will greatly reduce the number ofvalves in need of inspection.

Conclusions -

This Phase II research has resulted in quantitative techniques that can be used to predict

the performance of swing check valves and help identify those valves most likely to suffer

premature degradation from hinge pin wear and/or disc stud fatigue. Predictive models

for hinge pin wear and disc stud fatigue have been presented with experimental

verification as well as correlation to actual plant examples.

) 3

- 51-
REFERENCES

1. M. S. Kalsi, Todd Horst, J. K Wang. Prediction of Check Valve Performance and

Degradation in Nuclear Power Plant Systems, NUREG/CR 5159, prepared for the

U.S. Nuclear Regulatory Commission as Phase I SBIR Research, May 1988.

2. Check Valve Failures or Degradation, Significant Operating Experience Report

No. 86-3, Institute of Nuclear Power Operations, Atlanta, GA, October 1986.

3. Application Guidelines for Check Valves in Nuclear Power Plants, EPRI NP-5475,

January 1988.

4. Aging and Service Wear of Check Valves Used in Engineered Safety-Feature

Systems of Nuclear Power Plants, NRC NUREG/CR-4302, Vol. 1, December 1985.

5. Loss of Power and Water Hammer Event at San Onofre Unit I on November 21,

1985, NRC NUREG-1190.

6. Chong Chiu and M. S. Kalsi. Failure Analysis of Swing Check Valves - San

Onofre Nuclear Generation Site - Unit 1, NRC Docket No. 50-206, April 1986.

7. Chong Chiu and M. S. Kalsi. Plant Availability Improvement by Eliminating

Disc Vibrations in Swing Check Values, ASME Paper No. 86-JPGC-NE-6, October

1986.

8. M. S. Kalsi. Analysis of the 4-Inch Tilting Disc Check Valve Problems and

Proposed Modifications, Kalsi Engineering, Inc. report (3.10.2) to Southern

California Edison Company, 1986.

9. Todd Horst and M. S. Kalsi. Integrating the Check Valve Application Review with

Preventive Maintenance Programs, ASME Paper No. 89-JPGCINE-2, October 1989.

10. E. Rabinowicz. Friction and Wear of Materials, John Wiley & Sons, 1964.

11. Wear Control Handbook, edited by Peterson & Winer, ASME Publication, New

York, 1980.

12. J. F. Archard. "Wear Theory and Mechanisms," Wear Control Handbook, ed. by

Peterson and Winer, p. 35, ASME Publications, New York, 1980.

13. J. P. Tullis and W. J. Rahmeyer. Performance Tests on Feedwater Check Valves

for San Onofre Nuclear Power Plant, Phase A and Phase B, Utah State University

Foundation Report, April 1986.

14. S. L. Collier, C. C. Hoerner, and C. E. Davila. Behavior and Wear of Check

Valves, ASME Paper 82-PET-12.

-52 -
15. D. J. Trimble and T. A. Galisto. Selection of a Non-Cobalt Alloy for Sliding Wear

Service in a Nuclear (Check) Valve, Wear of Materials, edited by Rhee, Ruff, and

Ludena, p. 219, ASME Publication, New York, 1981.

16. W. J. Love. Value Bushing Wear in the Atwood-Morrell Co. Check Valves CV-3,

CV-5, PCSV-202, UNI-WJL-1, March 10, 1977, UNC Nuclear Industries, Richland,

Washington.

17. John E. Sununu. Wear in Check Valves Due to Flow Induced Motion, Master's

Thesis, Massachusetts Institute of Technology, May 1987.

18. J. 0. Hinze. Turbulence, McGraw-Hill, New York, 1959.

19. Laufer, The Structure of Turbulence In Fully Developed Pipe Flow, I9f54

20. K. T. Patton. Tables for Hydrodynamic Mass Factors for Translational Motion,

ASME 65-WA/IUNT-2.

21. Criteria of the ASME Boiler and Pressure Vessel Code for Design by Analysis in

Section VII and VIII, Division 2. Published by ASME, 1969.

i A

-53.
APPENDIX A

IMPACT AND FATIGUE PREDICTION METHODS
APPENDIX A

IMPACT AND FATIGUE PREDICTION MH ODS

A. 1. PEAK DISC STUD IMPACT FORCE ESTIMATION

The peak impact force estimates in the disc stud can be based upon the disc velocity before

impact using the energy balance relationships. The disc/disc stop assembly may be

considered as a simple mass-spring system where the disc is the moving mass and the

combined assembly has an equivalent stiffness as in a simple spring. During the impact,

a majority of the kinetic energy of the moving mass is converted to the elastic strain

energy in the spring with a small energy loss due to friction, vibration, or small contact

deformation. Neglecting the energy loss, the upper bound elastic strain energy in the

system at the peak of impact where the disc movement stops momentarily can be expressed

as

Kinetic Energy of Disc = Elastic strain energy in the assembly

For a small amplitude of disc oscillation, the kinetic energy of the disc before impact can

be estimated as:

K.E. = 2 m V2 (in-lb) 2

where

lb -sec2

m = Disc mass,

V = Linear velocity of disc before impact, in/sec

The equivalent stiffness of the assembly can be approximated without resorting to a

detailed stiffness analysis. As shown in most swing check valve designs, the most

flexible element in a disc/disc stop assembly is the disc stud. Assuming that the disc stud

is the spring element which absorbs all the kinetic energy during the impact and the disc

and disc stop are rigid, the total strain energy in the stud is

Stud Strain Energy = Strain energy in compression + strain energy in bending

=_ + 2

2 2

F 2 L M2 L

2AE 2 E1

F2 L e2 L)

2 (AE +El

where

F = uniform axial force on stud, lb

L = Stud length, in

- A.1 -
A = Average stud cross-sectional area, in2

M = Uniform bending moment on stud, in-lb

= Fe

e = Eccentricity of axial force, in

8 = Axial deformation of disc stud

9 = Angular deformation of disc stud

The upper peak impact force can be estimated based upon the kinetic energy to the strain:

M2= F2(L e2L

2 2 (AE EEI)

F= V

L e2L

- A.2 -
A.2. PEAK FORCE ESTIMATION FOR THE SIX-INCH INSTR 1h E DISC

I STRAIN GAUGES

SILICONE RUBBER SEALANT

Figure A.1

Instrumented Disc Cross-Section

Figure A.1 shows the cross-section sketch of the instrumented disc modified from a sixinch disc design to accommodate the strain gauge installation for the axial impact force

measurements. The axial stiffness of the disc was precisely analyzed by the finite

element method and found to be 1.0 K 106 lb/in. The combined disc weight plus one-half of

the hinge weight is 10.19 pounds. The peak impact force under a straight axial impact can

be estimated using the above derived equation as

F=V I

(L E2L

V 1

= 10.19 x LOxl 106 V

38S 6.4

F= 162.4V lb

The tested valve has a hinge length of 5.04 inches. Therefore, the disc angular velocity

versus impact force can be expressed as

F= 162.4 x 5.04 x co 180

= 14.285 am

where ID = Disc angular velocity before impact, deg/ sec

The input force varies linearly with the disc angular velocity. For convenience, the table

below shows the magnitude of angular/linear disc velocity vs. impact force.

- A.3 -
Disc Angular Velocity, Disc Linear Velocity, Impact Frce

degisec | in/sec l l b

2 0.1759 29

4 0.3519 57

6 0.5278 85

8 0.7037 114

10 0.8796 143

12 1.055 171

14 1.232 200

16 1.407 229

18 1.583 257

20 1.759 286

24 2.111 343

28 2.463 400

32 2.815 457

36 3.167 514

40 3.519 571

45 3.958 643

50 4.398 714

55 4.838 786

60 5.278 857

70 6.158 1,000

80 7.037 1,143

90 7.917 1,286

100 8.796 1,429

150 13.195 2,143

200 17.593 2,857

Table A.1

Magnitude of Angular/Linear Disc Velocity vs. Impact Force
A.S. ESITION OF DISC OSCIILATION FREQUENCY

Disc oscillation frequencies may vary with many factors such as the types of upstream

disturbance, fluid media, and disc positions (disc opening angles). The actual disc

frequency for a specific case is based on comparison against available test data from

references (Refs. 1, 6, 7, 17, 18, 19) and engineering judgment. The three methods

discussed below have been used in estimating disc oscillation frequencies: (1) the free body

flow equilibrium method, (2) the eddy frequency method, and (3) the pendulum frequency

method based upon the disc mass and hinge length.

A.B.1. FreeBodyFlowEquilibriunmMethod

The natural frequency of a single degree of freedom spring-mass system is given by:

2 x~Wig

where

fn = Natural frequency, hertz

Kstff = System spring rate, lb/in

W = Effective disc weight, lb

= Wdisc + 0.5 Whin + I3 p (dasc)s

g = Gravitational acceleration, in/sec2

= 386A

dcisc = disc diameter, in

p =Fluid weight density, lb/ns

For the check valve, effective spring rate can be expressed as the force required to open the

disc to some equilibrium angle, 0. An expression for this force, as developed in Reference

3, is:

2kApV2 . sinG F ~sin - g 2

where

F = Fluid force acting on the disc, lb

k =an empirical constant dependent on valve geometry

= 2.0 (average value)

A = Disc area, in2

V = Fluid mean velocity, in/sec

0 = Fluid impingement angle (see Fig. 1 in report), deg

Ksuf = R do where R = radius from hinge pin to disc %, in

R dO

Kif= 2 KApV2 d (sin0 sn

R g do 2)

- A.5 -
Kstiff = 2(KAV2 cosO sin ! + 2 cos 2 sin 0 Rg 1. 2 2 2

Finally

fn= 1 2KApV2 Z

2iRW

where

Z = [cos sin os 0sin 0]

A.3.2. EdyFrequecy Mehod

As shown in Reference 17, the eddy frequency of the fluid can be estimated as

0.08 V feddy =

where

feddy = Eddy frequency, Hz

V = Flow velocity, in/sec

Vs = Nominal valve size, in

A.3.3. PendlumFrequencyMethod

The pendulum frequency is most suitable for disc frequency estimations for low flow

conditions where discs are in almost vertical hanging positions. The pendulum

frequency is:

fpend = 1

where

fpend = Pendulum frequency, Hz

g = Gravitational acceleration, in/sec2

= 386.4

R = Disc hinge length, in

- A.6 -
A.4. ESIMATION OF DISC OSCILLATION ANGLE

Different types of upstream disturbances and distances of disturbance source produce

various degrees of disc fluctuation as shown in the test results in References 1 and 3.

These test results can be utilized for the similar types of upstream disturbances to estimate

oscillation angles for the disc fatigue and wear predictions.

A.5. ESTIMATIONOFDISC STUD FATIGUE STRENGTH

The disc stud fatigue strength under disc tapping conditions may be estimated using the

following steps:

i) Calculate the minimum flow velocity requirement to fully open the disc,

including the velocity margin factor to account for upstream disturbances as

given in References 1 and 3. If this disc is not fully open and possibly tapping,

then proceed with the following steps for stud fatigue calculations.

ii) Calculate the disc oscillation frequency as discussed in Section A.3.

iii) Calculate the disc oscillation angle as discussed in Section A.4.

iv) Calculate the maximum disc velocity based on the above calculated disc frequency

and amplitude as (Appendix D also provides the measured maximum disc velocity

from tests):

VzaX = R x 2 x 2x f 2

where

R = Hinge length, in

0 = Disc oscillating angle (peak to peak), rad

f = Disc oscillating frequency, Hz

v) Calculate the maximum impact force using the method described in Section A.1.

vi) From the above calculated peak impact force, the stud axial and bending stresses

at the weakest cross-section can be calculated as:

F

fa.d = A

M Fe

Obending =- =- z z

where

F = Peak impact force, lb

A = Cross-Section area, in2

e = Load eccentricity, in

Z = Section modulus, in3

- A.7 -
The maximum stud surface alternating stress is:

calt = x SCF x (a.jA + ('bending)

where

SCF = Stress concentration factor at the selected cross-section

vii) Calculate fatigue usage based upon the above calculated alternating stress, hours

of service, oscillation frequency, and the stud mterial fatigue curve (such as the

ASME fatigue curve used in the example in Appendix C).

It should be noted that the disc oscillation angle provided in Reference 1 is the 3u

disc angle on the Gaussian distribution curve (i.e., the actual disc oscillation

angles due to flow turbulence are less than the 3a value 99.7 percent of the time).

Therefore, the disc impact forces are expected to be much lower than the calculated

force the majority of the time. Statistical method such as Gaussian distribution

may be used to provide a more realistic estimation of the stud fatigue usage.

Detailed calculations are given in the example in Appendix C.

-A.8 -
APPENDIX B

CALCULATION OF VMIN, FLOW VELOCITY, AND WEAR RATE

FOR PLANT CORRELATION EXAMPLE 1
APPENDIXB

CALCULATION OF VxM, FLOW VELOCITY, AND WEAR RATE

FOR PLANT CORRELATION EXAMPLE 1

B.1. DATA USED IN CALCULATIONS

Dimensions

Valve ID

Disc Diameter

Disc & Hinge Weight

Hinge Length

Hinge Pin Diameter

Hinge Pin Wear Length

Fluid Impingement Angle

16.76"

15.875"

200 lbs

11.375"

1.625"

4.38"

(This is accounting for both pairs of bushings, namely, the disk-shaft bushings and

the trunnion bushings, that the shaft can

slide in.)

20°

Materials

Bushings

Hinge Pin

Hinge Pin Hardness

:Stellite-6/Stoody6

Haynes Alloy 25

38RC

Installation

Valve Orientation

Upstream Disturbance

Fluid

OperatingConditions

Flow Rate

Operating Temperature

Operating Pressure

Valve Usage per Year

Horizontal

Side discharge dead end line tee at 2 ft

Water

:1SOMgpm

530PF

1,875 psi

70%
B.2. VN AND FLOW VELOCITY CALCULATIONS

[g j WffCos 0 5

Vngeneradized - K p A sin 2 J

g = Acceleration due to gravity = 32.2 ft/sec2

C = Bouyancy factor = 0.9 for water

Weff = Weight of disc-hinge arm assembly = 200 lbs

0 = Impingement angle = 20°

K = Empirical construct = 2.0

p = Density of constant water (operating fluid)

at 530°F (operating temperature)

= 46.9 lb'ft3

A = Disc Area = 4 (x5 172 ft2=L3745ft 2

* Vminlgnerlized = 19.0 ft/sec

Since the upstream disturbance is a medium turbulence source,

Cup= L2 to L 4 - Table 3, pg. 20, NUREG/CR-5159

VMil = Cup X Vminlgenemlized

= L2 x 19.00

Vmn = 22.8 fps

Determination of Flow Velocity Under Operating Conditions

Operating Flow Rate 12,500 gpm

Valve ID : 15.75"

=w Cross-Sectional Area = ' x(15.75)2 = 194.8 in2

4

F low Velocity (fps) = 12,500 gal in3 1min x 1 ift

min gal 60 sec 194.8 in2 12 in

Flow Velocity = 20.59fps

This flow velocity is less than Vmin; accounting for upstream disturbance.
B.S. WEAR CALCUlATIONS

The wear equation is

WEKLd

H

where

K = Wear coefficient

L = Effective weight of disc = (Wdisc + Warm) + bouyancy factor

d = Total distance traveled by the hinge pin

H = Penetration hardness

Determination of Wear Coefficient, K

The geometric average of the wear coefficient of a compatible metal combination under

poorly lubricated conditions (1 x 104) and an unlubricated condition (5 x 104) is:

K = 10"

where [log 0 (1 x 104)+loglo(5x 104e)

X K=2.236x10 4

Determination of L, the Effective Disc Weight

Disc Weight = 200lbs

Bouyancy Factor = 0.9 (for water)

X Effective Disc Weight = 0.9 x 200 lbs = 180 lbs

L = 180lbs

Determination of Sliding Distance, d

d=0 x x2xf xx0.5xtxs 180

where

0 = Oscillating angle

f = Frequency of oscillation

t = Total time over which wear is calculated

s = Statistical factor

- B.3 -
The side discharge dead end tee line is a medium turbulence source. In NUREG/CR-5159,

page 58, Figure 30, the disc fluctuation angle is seen to be approximately 8 degrees for a

medium turbulence source.

9 = 8°

The equation described in Section A.3.1, page 6 of Appendix A, the frequency of oscillation

is given by

1 = 2KApV2 Z

2x X RW

where Z = [cos a sin 2 + 2 cos 0 sin 0]

K = an empirical constant dependent on valve geometry

= 2.0 (average valve)

A = Area of seat port= 2 x (1875)2 = L3745 ft2

4 144

V = Flow velocity = 20.6 fps (computed in Appendix B, Section 2)

9 = Impingement angle = 200

p = Fluid density = 46.9 lW'fM3 at 5300F

R = Hinge arm length = 11.375" = 0.948 ft

W = 200+!PD3 =200+ 1 x 624 x( 1.875) = 248 lbs 3 3isc

= f = 1.97Hz

1 12x2xL3745x46.9x(20.6)2x0.3316 ft2 X lbx ft2

n 2X 0.948x 248 ft ft3 sec2

Z[cos20 sin 10+ 1 cos 10sin20]=0.3316 2

= f=2.Ohz

* = Hinge pin diameter = 1.625"

Total Valve Usage per Year = 70%

0.7x8,760 hrs x3,6005.

X Total Operating Time, t = year hr

=n t = 22,075,200 seconds

- B.4 -
For a Gaussian distribution, the statistical factor, s - 0.309 from the sum of the products of

magnitude multiplied by probability.

Z d = 8oxXL x2x2.0xL625xO.5x22,075,200xO.309in

180

= 3.095 x 106 in

Determination of Penetration Eardness, H

H = RC38 = 38 x 14,232 psi

w H = 540,810psi

Wear Rate =K L d = 2.236 x 10 4 x 180 x 3.095 x 106

H 5.40,810

w W = 0.2303 cu.inlyear

Bearing Area = sd x I = = x L625 x 2.69 =,3.433in2

4 4

Total Bearing Area = Bearing area at disc to shaft bushing

+ bearing area at trunnion bushing = 3.433 in2 x 2

= 6.866 in2

Wear Depth VolumetricWear Rate 0.2303in8/yr

Total Bearing Area 6.866 in2

i Wear Depth = 0.0335 in/yr
APPENDIX C

CORRELATION OF FATIGUE ANALYSIS WITH AVAILABLE PLANT DATA

FOR PLANT CORRELATION EXAMPLE 4
APPENDIXC

CORRELATION OF FATIGUE ANALYSIS WIH AVAILABLE PLANT DATA

FOR PLANT CORRELATION EXAMPLE 4

C.1. CASE EHISORY AND AVAILABLE PLANT DATA

At San Onofre Nuclear Generating Site Unit 1, a water hammer event occurred in the

pressurized water reactor horizontal feedwater line caused by failed check valves in

November 1985 (RefS. 5, 6, 7). The disc stud/nut connection in these check valves had

fractured due to repeated impact against the open stop, which had allowed the disc to

separate from the hinge prior to this water hammer event. These valves had operated

satisfactorily for several years at full power conditions. Fifteen months prior to this event,

the plant was operated at approximately 85 percent reduced power. This resulted in

insufficient flow velocity to fully open the disc causing disc tapping and eventually fatigue

failure of the threaded connection. The specific valve and system data available from the

plant are:

Valve Description : 10" x 600# MCC Pacific swing check

Fluid Condition Water 0 3750F

18.54 ft/sec 0 full power

16.50 Wsec 0 reduced power

Disc Weight : 33 lb

Hinge Weight 12 lb

Total Impingement Angle : 15 deg

Disc Diameter : 11.25 in

Upstream Disturbances

Disc Stud Diameter

Disc Stud Length (approx.)

Hinge Length

: Control valve 0 19"

Expander 0 14"

: 1-1/8 in

: 3in

: 7.5in

Disc Material: A217 Tensile strength 70 ksi

Yield strength 40 ksi
C.2. STUD FATIGUE STRENGTH CALCULATIONS

C.2.1. Minim Required Flow Velocity

Vmj,, n = =e C a ' 0 mlf K A sin 0

where

g = 32.2ft/sec2

p = 54.56lb/ft3

C 489-p = 0 . 89

489

Weff Wdisc+P2 Whjng,=33+6+391b

K = 2.0

A= j (d., )

2 = 0.7854 (1L225) = 0.6903 ft2

0 = 15 deg

Vmin = 14.61 ft/sec

Minimum required flow velocity with upstream disturbance, V'min (see Ref. 1):

V'min = Cup Vmin

The estimated Cup for the upstream disturbance as stated in Ref. A.7 is 1.2 to 1.4.

Therefore,

V'min = 1.2 x 14.61 = 17.68 ft/sec

= 1.4 x 14.61 = 20.61 ft/sec

The calculated minimum required flow velocity is approximately in the same range as the

actual flow velocity under the full power condition. Therefore, the disc is not held firmly

against the backstop even at full power operation. At the reduced power condition, the

actual flow velocity is less than the minimum required flow velocity. Therefore, disc

tapping is expected.

- C.2 -
C.2.2. Disc Oscillating Frequency

Using the three methods given in Section A.3, we have

Wig

where

2x2x0.6903x 144x54.56x(16.5x 12)2

7.5x386.4x 128

x (cos 13.30 sin 6.7° +1cos 6.7° sin 13.3)

= 38.7 lb/in

Note: The disc angle, 0, is calculated from the disc equilibrium position

based on the actual flow velocity of 16.5 ftsec.

W= Wd~c+ 0.6Wbing+ 1lp (ddilc ~~~~~~

=(1I125)3

=54 lb

1 fn I = 38. ~ 7 X386.i265H ~ 2.65 Hz

2xr 54

0. 08 V =. 96 x 16.5 x 12 =L8H feft = ~O 9

==XI 6X H

Vs 10

Fpend= 21 = 2ui 385 L.14 Hz

Among the above calculated frequencies, the eddy frequency of 1.58 hz is closest to the

estimates based on Phase I test data and will be used in later fatigue calculations.

C.2.3. Disc Oscillation Angle

Based on the Reference 1 test data, the estimated disc oscillation angle is 16.2 degrees (3a

disc oscillation angle). The following impact force and stress are calculated using this

oscillation angle.
C.2.4. Maximum Disc Velocity

VmaX=Rx2xf

= 7.5 x 16. x xx 2 x 158

2 x 180

= 10.5 in/sec

C.2.5. MaximumImpactForce

F=V m

Assuming the loading eccentricity is 0.4-inch, then

39

F= 10.5 386.4

_ _ 3 __ _ _ _0.42 X3

b x L 1252 x 30 x 10+ 30 x 10" x x 1. L 1254

6,195e b

C.2.ff. Maximum Alternating Stress

1 (FFe~

Oak= 2 x SCFx (A + -Z)

Assuming SCF at the stud fillet radius is 2, then

(ait= 1 x2x{ 6,195 + 6,195x0.4] 2 X4 xL 1252 3 x L 1259) 4 32 )0

23,960 psi
C.2.7. Fatigue Usage

The above alternating stress calculations show that under 16.2 degrees of disc oscillation

angle and impacting at the maximum disc velocity of 10.6 in/sec, the maximum peak

stress in the stud is 23,960 psi. Using this conservative peak stress as the 3a peak stress, the

Gaussian distributions for peak stress are:

Gaussian Distibution_[ Strea Magnitude, W Id ,% |

3.75 - 4.00 a 30.0 - 3L9 0.02

3.50 -3.75 a 28.0 -30.0 0.02

3.25 -3.60a 26.0-28.0 0.08

3.00 - 3.25 a 24.0 - 26.0 0.14

2.75-3.000 22.0-24.0 0.34

2.50 -2.75 a 20.0 - 22.0 0.60

2.25 - 2.506 18.0 - 20.0 1.20

2.00 -2.25 a 16.0- 18.0 2.12

1.75 - 2.00 a 14.0 - 16.0 3A6

L50 - 1.75 a 12.0-14.0 5.34

L25 - 1.50a F 10.6-12.0 7.76

1.00 - 1.25 a 8.0 - 10.0 10.62

Table A.1

Gausslan DIstributIons for Peak Stress

The fatigue usage of the stud can be estimated based on the above peak stress distributions,

hours of service, oscillating frequency, and a fatigue curve such as Figure C.1 (from

ASME, Section III, Fig. I-9.1, 1989 Edition). The fatigue usage in an hour's tapping can be

estimated as shown in Table C.2:
Gaussian Avg. Peak Design Impact Cycles

Dishribution Stre*s, ksi Cycles 1lr Usge

3.75- 4.00 a 31.0 2.4x 104 0.4557 L9 10-5

3.50-3.75a | 29.0 3x104 0.4557 1.51 10-5

3.25 - 3.50 a 27.0 3.6 x 104 L823 5.1 x 10-

3.00-3.25a 25.0 4.5x104 3.2 7.1x10-5

2.75 - 3.00 a 23.0 6x 104 7.7 L3 x 10-4

2.50 -2.75 a 21.0 1 x b 13.7 1.37 x 104

2.25 - 2.50 a 19.0 1.3x 105 27.3 2.10x 10-4

2.00-2.25a 17.0 2x 105 48.3 2.42 x 10-4

1.75 - 2.00 a. 15.0 3x 105 78.8 2.63 x 10-4

1.50 - 1.75 a 13.0 5x 105 121.7 2.43 x 10-4

1.25- 1.50 a 11.0 lx 107 176.8 L8x 10-

1.00-1.25a 9.0 0S 242.0 0

Total 1.399 x 103

* Number of impact cycles is calculated as N - 3600idja, x Probability

Table C.2

Fatigue Usage Estimations

Therefore, the total fatigue usage in one hour of service is estimated to be 1.399 x 103. It is

equivalent to a stud design life of 715 hours. The actual disc stud failed within 10,000 hours

of service. It should be noted that the ASME fatigue curve used in the above calculations is a

design fatigue curve; i.e., a safety margin of 2 on peak stress or 20 on fatigue cycles is

included in deriving the fatigue curve (Ref. 21). Therefore, the actual stud fatigue life

wouldi be more than one order of magnitude higher than the above estimated value if the

actual mean fatigue curve of the material is used for the calculations.

- C.6 -
104

UI ---- I l

l l l l + l _ .I

la

1*1

14.14. .9.

C

-4

102

10

10 102 IDS

NOTE: E -30 X 104 pu1

- - - UTS C 80.0 ksl

- UTS 116.0 - 130.0 kOi

Interpolet for UTS 80.0- 115.0 hil

Number of cycles. N

Figure CA

ASME, Section iII Design Fatigue Curves for Carbon, Low Alloy, and High Tensile Steels
APPENDIX D

PLOTS OF MEAN DISC SPEED AND 3-SIGMA DISC SPEEDS

FOR THE 3-INCH AND 6-INCH TEST SWING CHECK VALVES
LTST OF FIGURES

Figure No. Page

D.1 3-Inch Valve Mean Disk Speed(Elbow Up, OD) D.4

D.2 3-Inch Valve Mean Disk Speed (Elbow Up, ID) D.4

D.3 3-Inch Valve Mean Disk Speed (Elbow Up, SD) D.5

DA 3-Inch Valve Mean Disk Speed (Elbow Up, 5D) D.5

D.5 6-Inch Valve Mean Disk Speed (Elbow Up, OD) D.6

D.6 6-Inch Valve Mean Disk Speed (Elbow Up, ID) D.6

D.7 6-Inch Valve Mean Disk Speed (Elbow Up, 3D) D.7

D.8 6-Inch Valve Mean Disk Speed (Elbow Up, 5D) D.7

D.9 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Up, OD) D.8

D.10 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Up, ID) D.8

D.11 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Up, 3D) D.9

D.12 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Up, 5D) D.9

D.13 3-Inch Valve Mean Disk Speed (Elbow Down, OD) D.10

D.14 3-Inch Valve Mean Disk Speed (Elbow Down, ID) D.10

D.15 3-Inch Valve Mean Disk Speed (Elbow Down, 3D) D.11

D.16 3-Inch Valve Mean Disk Speed (Elbow Down, 5D) D.11

D.17 6-Inch Valve Mean Disk Speed (Elbow Down, OD) D.12

D.18 6-Inch Valve Mean Disk Speed (Elbow Down, ID) D.12

D.19 6-Inch Valve Mean Disk Speed (Elbow Down, 3D) D.13

D.20 6-Inch Valve Mean Disk Speed (Elbow Down, 5D) D.13

D.21 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Down, OD) D.14

D.22 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Down, ID) D.14

D.23 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Down, 3D) D.15

D.24 3- and 6-Inch Valve 3 Sigma Disk Speed (Elbow Down, 5D) D.15

D.25 3-Inch Valve Mean Disk Speed(3/16" Holes, Orifice Plate at 1.5D) D.16

D.26 3-Inch Valve Mean Disk Speed(7/8" Holes, Orifice Plate at 1.5D) D.16

D.27 6-Inch Valve Mean Disk Speed(3/8" Holes, Orifice Plate at 1.5D) D.17

D.28 6-Inch Valve Mean Disk Speed(1.75" Holes, Orifice Plate at 1.5D) D.17

D.29 6-Inch Valve Mean Disc Speed (3/16" Holes, Orifice Plate at 1.5D) D.18

D.30 6-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 1.5D) D.18

D.31 3-Inch Valve Mean Disc Speed (3/16" Holes, Orifice Plate at 2.5D) D.19

D.32 3-Inch Valve Mean Disc Speed (7/8" Holes, Orifice Plate at 2.5D) D.19

D.33 6-Inch Valve Mean Disc Speed (3/8" Holes, Orifice Plate at 2.5D) D.20

D.34 6-Inch Valve Mean Disc Speed (1.75" Holes, Orifice Plate at 2.51)) D.20

D.35 6-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 2.5D) D.21

D.36 6-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 2.5D) D.21

D.37 3-Inch Valve Mean Disc Speed (3/16" Holes, Orifice Plate at 4.5D) D.22
D.38 3-Inch Valve Mean Disc Speed (7/8" Holes, Orifice Plate at 4.5D) D.22

D.39 6-Inch Valve Mean Disc Speed (3/8" Holes, Orifice Plate at 4.5D) D.23

D.40 6-Inch Valve Mean Disc Speed (1.75" Holes, Orifice Plate at 4.5D) D.23

D.41 3-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 4.5D) D.24

D.42 6-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 4.5D) D.24

D.43 3-Inch Valve Mean Disk Speed(3/16" Holes, Orifice Plate at 10D) D.25

D.44 3-Inch Valve Mean Disk Speed(7/8" Holes, Orifice Plate at 10D) D.25

D.45 6-Inch Valve Mean Disk Speed(3/8" Holes, Orifice Plate at 10D) D.26

D.46 6-Inch Valve Mean Disk Speed(1.75" Holes, Orifice Plate at 10D) D.26

D.47 3-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 10D) D.27

D.48 6-Inch Valve 3-Sigma Disc Speed (Orifice Plate at 10D) D.27

D.49 3-Inch Valve: Relative Severity of Various Upstream Disturbances

Expressed in Terms of 3 Sigma Disc Speed D.28

D.50 6-Inch Valve: Relative Severity of Various Upstream Disturbances

Expressed in Terms of 3 Sigma Disc Speed D.28

LIST OF TABLES

Table No. Page

D. IA 3-Inch Valve: Mean Disc Speed vs. Flow Velocity Plots D.3

D.1B 3-Inch Valve: 3 Sigma Disc Speed vs. Flow Velocity Plots D.3

D.2A 6-Inch Valve: Mean Disc Speed vs. Flow Velocity Plots D.3

D.2B 6-Inch Valve: 3 Sigma Disc Speed vs. Flow Velocity Plots D.3

- D.ii -
APPENDIXID

DATABASE OFMEANDISC SPEED AND S-SIGMADISC SPEEDS

FOR THE 3-INCHE AND G-INCH TEST SWING CH3ECKVALVES

Compiled in this appendix is a series of graphs that plot mean disc speeds and three-sigma

disc speeds vs. flow velocity. The graphs cover four different types of upstream

disturbances located over a range of distances upstream from the valves. Data is presented

for both the three-inch and the six-inch test swing check valves.

Table D.1A provides a cross-reference index linking the upstream disturbance and the

distance it was located from the three-inch valve to the number of the figure in which the

data of mean disc speed vs. flow velocity is plotted.

Table D.1B provides a cross-reference index linking the upstream disturbance and the

distance it was located upstream of the three-inch valve to the number of the figure in which

three-sigma disc speed vs. flow velocity data is plotted.

Table D.2A provides a cross-reference index linking the upstream disturbance and the

distance it was located upstream of the six-inch valve to the number of the figure in which

the data of mean disc speed vs. flow velocity is plotted.

Table D.2B provides a cross-reference index linking the upstream disturbance and the

distance it was located upstream of the six-inch valve to the number of the figure in which

the three-sigma disc speed vs. flow velocity data is plotted.

The objective of determining the mean disc speed was to estimate the average angular

distance traversed by the disc per second. The scaled LVDT signal that measured the

instantaneous disc angle was first differentiated in time. The resultant signal was the

instantaneous velocity of the disc. Since the disc oscillates about a mean position, the

direction of the disc velocity varies constantly. Because it was the magnitude of angular

distance traversed by the disc and not the direction of disc travel that was of significance,

the magnitude of the velocity signal was extracted by determining the absolute value of

each of the data points in the velocity signal. The average of the resulting signal was the

disc speed, a number that represented the angular distance traveled by the disc per second.
This value has been determined for both the three- and six-inch valves for a variety of

disturbances located at several different upstream distances from the test valve and for a

wide range of flow velocities.

The objective of determining the three-sigma disc velocity was to estimate the maximum

possible velocity at which an impact of the disc stud against the disc stop could occur. The

mean standard deviation and the mean of the mean disc speed signal was first computed.

The three-sigma disc velocity was equal to the sum of mean and thrice the standard

deviation of the mean disc speed data. Figures 49 and 50 are used to summarize an overall

trend that is observed from the 3c disc speed data presented in this appendix. The

dependence of 3a disc speed on the level of severity of the turbulence source and its

proximity to the valve is clearly noticeable.
Upstream 1 U pleam Severe Medium

|Distnce ElbowUp ElbwDon| Distance Turbulence ITurbulence

OD mean 1 13 1.6D mean 25 26

1D mean 2 14 2.6D mean 31 32

3D mean 3 15 4.5D mean 37 38

|D mean 4 16 I lOD mean 47 44

Table D.1A

3" Valve: Mean Disc Speed vs. Flow Velocity Plots

iUpstram Elbow U Eo w Upsheam Severe Medium

Distnce IDiace aI Turbulence [Turbulence

OD3Sigma 9 21 j 1.5D mean 29 29

|lD3Sigma 10 22 2.5D mean 35 35

3D3Sigma | 1 23 4.5D mean 41 41

5D3Sigma 12 24 | lODmean | 47 47

Table D. B

3" Valve: 3 SIgma Disc Speed vs. Flow Velocity Plots

IUpsram I fo~ Ilo~w 11ps~tram 1 S&vere IMedium I

Disance V | D wn |Distance Turbulence [Turbulence

OD mean 5 17 1.5D mean_| 27 [ 28

1D mean 6 18 2.5D mean| 33 | 34

3D mean 7 19 4.5D mean | 39 40

5D mean 8 2) lOD mean 45 46

Table D.2A

6" Valve: Mean Disc Speed vs. Flow Velocity Plots

Upsre 1 E Up Ebw F Usa Severe Medium Disance _E p_11 Distance Turbulence Turbulence

OOD3Sigma 9 21 |j 1.5D mean 30 30

1D 3 Sigma 10 22 2.5D mean 36 36

3D3Sigma | 23 4.5D mean 42 42

5D3Sigma 12 24 10 lODmean 48 48

Table D.2B

6" Valve: 3 Sigma Disc Speed vs. Flow Velocity Plots
25-

u1 2 3- Elbow Up at OD| + Legen u+ 73 stop

20 - + + + * 630 stop

+ o53 stop

.- a 15-

10 . .

0 2 4 B 8 10 12 14 16 18 20 22 24 2B 28 30

Flow Velocity, feet/sac

Figure DA.1

3-lnch Valve Mean Disk Speed

12 -

*~ ~~~~~(lo Up OD

oJ 12 - | 3U Elbow U Legend

10 24881011416120222+ + 7328stop

lo -+Flo Velocty, fet/se

Fi+ g 53u stop

8 2- nl+

co

3-Ic Valve Mean Disk Speed

(Elbow Up, 1OD)

12-

0 4 8 0 1214 1_1 ___2 2022 4 6 2 3

(Elbow Up,3 EloDpa)1D Lgn
12 -

10[ 3' Elbow Up at 3D Legend

10 + | El + ~~~~~~~~~~~~~~~+ * 730 5630stop stop

A + + 530 6 stop

o 6- +~~+++ ++

X 4- 0 +

~~~~~~CFO

4 -2 El 13 * + + 2

C 2 s m. ** +

S~~~~~0C m8 gk a ^ E + +

0 I.. A I I I I A

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.3

3" Valve Mean Disk Speed

(Elbow Up, 3D)

12-

0: | 3T Elbow Up at 5D Legend

810 * 30 stp

+ + 0 530stop

X 6 ~ i co

0 ~ + +

40 ~~~~~+

2 +1

S * 0 *p~ +~~~K* 

0 2 4 6 8 10 12 14 16 18 20 2224 2628 30

Flow Velocity, feet/sec

Figure D.4

3" Valve Mean Disk Speed

(Elbow Up, 5D)
12

0

0

a.1

0

C

01

10

8

8

4

2

0

| 6 Elbow Up at OD || Legend

+ + 70° stop + * 600 sto

+ 1 sow stop

+ +

an

13 ~ ~ ~ +

1l mu aa

12 D* * + 4

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.5

6" Valve Mean Disk Speed

(Elbow Up, OD)

Q

a

a

a

a

a.

0

0.

12 -

10 -

8-

6-

4-

2-

1 lbw p atIDI

+ 4~~4

+ + 4 ++ q

* 90 G *+ +

U. 9 * * + 4+

~ct~aU to I4 ++

ALege-nd

I+ 70 stop

* 600 stop

l 500 stop

0 .1

0 2 4 6 8 10 12 14

Flow Velocity,

II . .I II I I 1

16 18 20 22 24 26

feet/sec

Figure D.6

6" Valve Mean Disk Speed

(Elbow Up, I D)
o i~~~~~~~~~~~~~~~~Lege

* 6 * | 6^ Elbow Up at 3| e

8+ + 70° stop

* 5 + * 60 stop

+ 0 504l Stop

- l

8 3

0 E +

C0 ~ 130+

a 1 *ji *w t +

0

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.7

6" Valve Mean Disk Speed

(Elbow Up, 3D)

|n Elbow Legend6^ Legend

' - ' V~~~~+700stopI

E . * 600 stop

X. 50O stop 4~~~

4 -~~~~~~~~~4

+

0,, 2+Ex 2

22

a +

CZC 1In tim a 40.0*0 +m+ o 4.

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.8

6" Valve Mean Disk Speed

(Elbow Up, 50)
a

0

a

a

0

0

*0

-i0

a

0

a

S

cm

a

a

4

0

a

a

01

0.b

a,

a

a,

4

ci

100

90

80

70

60

50

40

30

20-

10

I Elbow Up atOD

S

a aQ1

3~~~~

a U

m * *

Et ~ * ~ ~ ~ * * 'P * 

*. C

.I~~~~~~v

0

50

45

40

35

30

25

20

15

10

5

0

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.9

3" and 6" Valve 3-Sigma DIsc Speed (Elbow Up, OD)

Elbow Up at ID

C0 13

C . 1 1. I a l l I a I I O I . B .

I_ * @ @ -~~~~

Legend

G0 3* Valve

* 60 Valve

Legend

* 3V Valve

3 S Valve

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/see

3"' and S"

Figure D.10

Valve 3-Sigma DIsc Speed (Elbow Up, 1 D)
S

0.

V

-P

a

;S

0

0)

0

0

a)

E

a

B

50

40

30

20

10

0

50

45

40

35

30

25

20

15

10

5

0

IElbow Up at 3DI

B

va

MP *~~ #

*S *O G do 5 l

., ., ., ., ., ., ., .,.,.L

0 2 4 6 8 10 12 14 16 18202224 26283C

Flow Velocity, feet/sec

Figure D.11

3" and 6" Valve 3-Sigma Disc Speed

(Elbow Up, 3D)

| Elbow Up at 5D | Legend

m 3' Valve

* 6" Valve

B

U.M

0coo mea

*4 0 * * so *ma

I

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28

Flow Velocity, feet/sec

Figure D.12

3" and 6" Valve 3-Slgma Disc Speed

(Elbow Up, 5D)

30
0

0

0

a.

g2

VS

06

X0

21.

0

co

0

0.

0

16

14

12

10

8

6

4

2

0

16

14

12

10

8

6

4

2

0

13 low Downat OD

+4+4

3-*+~~~~~44

a 3 +

a

a pU ea E

I ,, , * .I I ,I .*, , I I

0 2 4 8 8 10 12 14 16 18 20 2224 26 28 30

Flow Velocity, feet/sec

Figure D.13

3" Valve Moan Disc Speed (Elbow Down, OD)

1+ 73', stop

I*630stopI

1+ 73"1 stop I*63*0stopI

30Elow onat1IDI

IPO +

**I4t.I + +

m * ~ 4

4.

0 2 4 6 8 10 12 14 18 18 20 22 24 28 28 30

Flow Velocity, feet/sec

Figure D.14

3" Valve Mean Disc Speed

(Elbow Down, 1D)

-D.10-
m

-e

06

f

(0

CO

(09

S

C

Se

W

12'-

12 FElbow Down at 3D

10

+

8-~~+4 + + +.

6- 1+ + , ++

4 - *C z +

2 la ++

+¢n n=*44+ +

0 1 0 2 0 3 0

Flow Velocity, feet/sec

Figure D.15

3" Valve -Mean Disc Speed

(Elbow Down, 3D)

12-

I Elbow Down at 5D I

10 F

,- + + +fi 

U + 4.~~~

4 a 0 +

2 at

0* p * -

O- 10. 20 3-.........

Legend

+ 730stop

* 63- stop

a 530 stop

Legend

+ 730 stop

* 630 stop

o 530 stop

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.16

3" Valve Mean Dlsc Speed (Elbow Down, SD)

- D.1 -
8 0 ~ ~ *Legend * 7 16 Elbow Down atODI +70-stopIl

e+ + * 60- stop

6 + + 50 stop * +~~~~~~~~

.0i 4 e + + +

0.* 4 - l +

3 + + +

o 2 v2

0

0 2 4 6 8 10 12 14 1618 20 22 24 28 28 30

Flow Velocity, ftet/sec

Figure D.17

6" Valve Mean Disc Speed

(Elbow Down, OD)

<,6-

e 1~~6- Elbow Down t 1 D Legen [i~~~o~~~i~~ownat1D ~ ~ ~ ~ +700stop

5 stop

o 4. +

.0 1- 4 4. Z 4 >. Dr

-O 13 4 8 01 41 8202 42 83

0

Flow Velocity, feet/soc

Flgure D.18

6" Valve Mean-Disc Speed

(Elbow Down, ID)

-D.12-
0

0

04

0

0

An

0

CL

0

Ok

C

a

6

5

4

3

6 Ebw Downat 31)ID

m. a * in

a Ma. a a U I

2

1

0

6

i I a I a a * a * * * X A I

0 2 4 6 8 10 12 14 16 18 20

Flow Velocity, feet/sec

Figure D.19

6" Valve Mean Disc Speed

(Elbow Down, 3D)

Legend

+ 700 stop

* 600 stop

l 50° stop

Legend

+ 70° stop.

* 60°stop

n 50° stop

U

0

0

E

(In

CL

W

0

5.

4.

3.

2-

I1-

16 Ebo own at 51)

a4.

+

4.

4.

. 4-

IF

4+

- *a gb 8

C 1 la 14 +Is

0 B - I -

0 2 4 6 8 10

Flow Velocity,

12 14

feet/sec

1 6 18 20

i Figure D.20t

6" Valve Mean Dlscs Speed

(Elbow Down, 5D)

-D.13 -
0

a

S

a

b.

0

0.

a

a

CL

U)

0

I.

0

0.

CO

U)

(a

CO)

60

50

40

30

20

10

0

60

50

40

30

20

10

0

I lbwaowta0 D I

U U3

*3a

U

13 a

1% 0~~~S

113" ValveI

I3 3'Valve

1.1. I A I. aI&IIIII

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 3C

Flow Velocity, feet/sec

Figure D.21

3" and 6" Valve 3-Sigma DIsc Speed

(Elbow Down, OD)

I lbwDown at D ID

U

Ur M

_%Y. 13 *U

I

.50O 2 . I I B f f I. 1 I . I 1 I . . 1 . . .

2 4 6 8 10 12 14 16 IS 20 22

I .I. I I I 1

24 26 28 30

Flow Velocity, feet/sec

Figure D.22

3" and 6" Valve 3-SIgma Disc Speed

(Elbow Down, ID)

-D.14-
a

0

a

0

C

0

0

la

0

I,

AC

2

a

a

E

0

S

a

0

J9~

40

30

20-

10-

Elbow Down at 3D

a

a e FP

m*U

U00

a

U 0

a

a C a aa 0

0 0

40

30

20

10

0

l' I a I a I a I a .a . I I ,., I . I .I li

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 3C

Flow Velocity, feet/sec

Figure D.23

3" and 6" Valve 3-Sigma Disc Speed

(Elbow Down, 3D)

jDowDownaat5D I

a a aa

' *.* .. .* .* a .* .- . .................

J030 Valvej

Legend

fa3' Valve~ * G Vlv

0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

Flow Velocity, feet/sec

Figure D.24

3" and 6" Valve 3-Slgma DIsc Speed

(Elbow Down, SD)

-D.15-
30

0

S

a

*0

a.

S

C*

a

a

a

a0

S

*0

0

0

0.

2

25

20

15

10

5

0

20

15

10

5

0

13/16- Holes Orifice at 1.5(

Fe* + +

Me. .~~~~~~~~4 4.+

El

In la 's*

l

0 2 4 6 8 10 12 14 16 18 20

Flow Velocity, feet/sec

Figure D.25

3" Valve Mean Disc Speed

(3116" Holes, Orifice Plate at 1.5D)

17/8* Holes Orifice at 1.5DI

+° * a a a + * 

+ +

I . I I4I I

I+73.VstopI

IB83 stopI

1+ 73"1 stop I' 631 stop 0 50tsop

0 2 4 6 8 10 12 14 16 18

Flow Velocity, feet/sec

Figure D.26

3"' Valve Mean Disc Speed

(7/8" Holes, Orifice Plate at 1.5D)

20

-D.16 -
30 I

0

0

0

0

E

Ca

8

0

I 13/8' Holes Orifice at 1.51

25-

20 -

15

10

:1 ,

Legend

+ 700 stop

* 60° stop

0 501, stop +

+

++

+

+4

C ++ *4'O:

* . a + d u~~ e~

I-. Ip*,*p*p-i-.

0

I. -- a .0~~~~~~~~~~~~~~~~~~~ .a. I .I . I .I .

2 4 6 8 10 12 14 16 18

Flow Velocity, feet/sec

Figure D.27

6" Valve Mean Disc Speed

(3/8" Holes, Orifice Plate at 1.5D)

20

&' .'

a

0

a

0,

0

0

0.

CD

m

C

01

15-

10 -

5-

11.75' Holes Orifice at 1.5DI

if 4+I ++

%a

% G~ +

I+700stopI

* 600stopI

, . . . . ~ 1

0 2 4 6 8 10 12 14 16 18

Flow Velocity, feet/sec

20

Figure D.28

6" Valve Mean Disc Speed

(1.75" Holes, Orifice Plate at 1.5D)

- D.17 -
0

a

a

-S

ao

2

IS

a

0

0

.9

C.)

120

110

100

8 10 12 14 16

Flow Velocity, feet/sec

Figure 0.29

6" Valve Mean Disc Speed

(3116" Holes, Orifice Plate at 1.50)

0

a

Cs

5-

a

a

S

0.

a)

a

0

100

90 -

80 -

70 -

60 -

50-

40-

30 -

20 -

10-

0 -

I

70 Sto Orifice at 1 .5D

I Orifice at 1.51)

0

oU 

00

9...* ."

+4 03

I Legend

l 318' Holes

* 1.7520Hole

2 0

U

.

a 40

a ~ ~ . . . . . . . o . 2 I . . 8 I . 1 2 1 1 6 1 I8 

0 2 4 6 8 1 0 1 2 1 4 1 6 1 8

Flow Velocity, feet/sec

Figure D.30

6" Valve 3-Sigma Disc Speed (Orifice Plate at 1.50)

-D.18-
25 C0

0

0)

i

F

0

CL

-0

0

C

£0*

S

a

CD

01

1 3/18" Holes Orifice at 2.5D I

+

20 -

15-

10 -

4.

0'

0T +

13

4.

0

l

0 092

0o C s 1 0

as

0

25

20

15

10

5

0

a I I I I .

0 2 4 6 8 10 12 14 16 18 2C

Flow Velocity, feet/sec

Figure D.31

3" Valve Mean Disc Speed

(3/16" Holes, Orifice Plate at 2.5D)

Legend

T 730 stop

* 63e stop

o 53° stop

Legend

+ 730 stop

* 63 stop

13 530 stop

| 7/8' Holes Orifice at 2.5D I

4

T4.

4 + 

1-,

o #* + 4

a 4. +

4. Cim *a £ a U - T U - * a * ... I * a * I * I * I * * I

0 2 4 6 8 10 12 14 16 18 20

Flow Velocity, feet/sec

Figure D.32

3" Valve Mean Disc Speed

(7/8" Holes, Orifice Plate at 2.5D)

-D.19 -
0

S

a

a

0

0

h.

m

0

00

a

C

ai

12

10

8-

6-

4-

2-

3/80 Holes Orifice at 2.5D

U D d

Legend

+ 70- stop

* 600 stop

a Sol, stop

0

0 2 4 6 8 10

Flow Velocity, feet/sec

Figure D.33

6" Valve Mean Disc Speed (3/8" Holes, Orifice Plate at 2.5D)

0

0

i

01

0

-01

0.

0

as

C

ai

0

12

10

8

6

4

2

0

Legend

+ 701 stop

* 601 stop

l 5011 stop

0 2 4 6 8 10 12 14 16

Flow Velocity, feet/sec

Figure D.34

6" Valve Mean Disc Speed

(1.75" Holes, Orlifce Plate at 2.5D)

18 20

- D.20-
0

0

0

b.

0

0,

-i

4)

0

CL

co

0

0

a

S

E1

R

100

90

80

70

60

50

40

30

20

10

0

100

90

80

70

60

50

40

30

20

10

0

I

Orfieat2.51)

0

IF

C0 0

* 0 0

*

9 0 %?

I . -I A I I I a .. . .I .... 

0 2 4 6 8 10 12 -14 16 18 2C

Flow Velocity, feet/sec

Figure D.35

3" Valve 3-Sigma Disc Speed (Orifice Plate at 2.5D)

I . *1

Legend

# 7/81 HolesI

j r 3/8" Holes

L .7" olesj

I riice at 2.51)

0moEmm la

0

10 0.

* * 0

I MO~~ 0 in

U

0 2 4 6 8 10 12 14 16 18 20

Flow Velocity, feet/sec

Figure D.36

6" Valve 3-Sigma Disc Speed

(Orifice Plate at 2.5D)

*D.21 -
0

a

0

0

-01

0)

a

C

.x

aa

0

0

0

a

0

0.1

C

a0

0

20

311 6 Holes Orifice at 4.50

15 -9

El+

5 D +

0 2 4 6 8 10 12 14 18 18 20

Flow Velocity, feet/see

Figure D.37

3" Valve Mean Disc Speed

(3/16" Holes, Orifice Plate at 4.5D)

20 -

7/8B Holes Orifice at 4.50

15

10 - +

* +~~ ~+4+

+ 4 + + +

5 +~~~~~~~ C ~ ~ ~

M13g * 0 lp *

0 2 4 6 8 10 12 14 16 18 2(

Legend

+ 73e stop

* 63e stop

El 530 StOp

Legend

+ 73e stop

* 630 stop

D 53e StOp

)

Flow Velocity, feet/sea

Figure D.38

3" Valve Mean DIsc Speed

(7/8" Holes, Orifice Plate at 4.5D)

- D.2:2-
0

0

-0

a

0

E

0

*0

0

CL

S

x

20

is10 -

5-

| 3180 Holes Orifice at 4.5D I Legend

+ 70O stop

* 600 stop

a 5V0 stop

04. + +

0 +++

4..~~~41

a *0 *

a~5~O

a

.60

C+

4.

4.

+ +

$

is a4. +'m+1

0 'I

0 2 4 6 8 10 12 14 16 18 20

Flow Velocity, feet/sec

Figure D.39

6" Valve Mean Disc Speed

(318" Holes, Orifice Plate at 4.5D)

in C

0

-0

0

*0

0.

10

le

f

0L

01

8-

I

6-

4-

2 -

1 1.75 Holes Orifice at 4.51I

*

B +T

B 1B * + 4

* 4

,'mm&. a 4T + m 0 

n4.~ b. U B I

I+700stopl

*00stop

L 50. stop

0 I.

..... . I. I I a ..... I . I. ..... I . I- . I .A

0 2 4 6 8 10 12 14 16 18 20

Flow Velocity, feet/sec

Figure D.40

6" Valve Mean Disc Speed

(1.75" Holes, Orifice Plate at 2.5D)

-D.23-
0

0

S6

0.

C*

0.

Cl)

0

21

01

0

0.1

Cl)

B;

50s

40

30

20

10

0

60

50

40

30

a a 3 1 rifie at 4.51)

0 * 0U

0~~~~~ *~~~~~1

9 a

I

, I., ., . ., . I . I I

0 2 4 6 8 10 12 14 16 18 2C

Flow Velocity, feet/3sc

Figure D.41

3" Valve 3-Sigma Disc Speed (Orifice Plate at 4.5D)

a 3/1 6"Hates

1U7/8' HolesI

a /WHoles

Orifice at 4.5D)

a am U

mm, V

U aN

13a *g~~U 5 *

20 -

10 -

0 .1--I . - - - - -~ ~~ . . E . ^ .

0

I ' . a * a . I - , . a . .

2 4 6 8 10 12 14

Flow Velocity, feet/sec

16 18 20

Figure D.42

6" Valve 3-Sigma Disc Speed

(Orifice Plate at 4.5D)

- D.24 -
a

U

I

00

CL

S)

SK

6

| 3/16- Holes Orifice at 1 OD

1-

0 + * +.

+ + ld5

* B

* 3

Legend

+ 730 stop

* 630stop

a 530 stop

4

0 '.u.

0 2 4 6 8 10

Flow Velocity, feet/sec

Iigure D.43

3" Valve Mean Disc Speed (3/16" Holes, Orifice Plate at 10D)

S

0

U01

i

SL

S)

8

. 6

4

2

0

I+73-stopI

1* 630 stop

20

Flow Velocity, feet/sec

Figure D.44

3" Valve Mean DIsc Speed

(7/8" Holes, Orifice Plate at 10D)

-D.25 -
6 a

0

0

-01

-o

0.b

0

-a.

0

-01

0

0.

0

i

S

5

4

3

2

1

0

5

. 318' Holes Orifice at 10D

l +

a *0

03 Q a * a* 0 0*

9 *~~ a * & 

0 2 4 6 8 10

Flow VelocIty, feetlsec

Figure D.45

6" Valve Mean Disc Speed

(3/8" Holes, Orifice Plate at 1OD)

. -~~~~~~~~~~~~~~~~~~~

Legenct

+ 70- stop

* 60' stop

o 50 stop

Legend

+ 70' stop

* 60' stop

13 500 stop

3 -

2-

1-

1 1.75S Holes Orifice at 100D

++~~ ~~ U

+~ P a +

U l

q . . . . . . . . . _

0

I . I . - U- --.

2 4 S 8 10

2I- 1- 16 1 1 2 14 16 18 20

Flow velocity, leetlsec

Figure D.46

6" Valve Mean Disc Speed

(1.75" Holes, Orifice Plate at 10D)

- D.26 -
a)0

0

00

S

0

1.9

-0

10

9..

8

7.

6-

5.

4.

3.

2-

1*

n .

9

0~~~~~~a 0

a

1U

A

j0n3/16' Holes

* 78' oles

16 16 2

U . .. . .. . . .U -.............

0 2 4 6 8 10 12 14

Flow Velocity, feet/sec

Figure D.47

3" Valve 3-Sigma Disc .Speed (Orifice Plate at 10D)

S

06

S.1

C.

0

SW

0

E

0

24

22

20

18

16

14

12

10 -

6-

4

2-

0 :

I Orifice at 10D

0 0 d 0 Ed 0

0 * EU

& a * .

Leged

IU3/8SHolesI

[ 175*Hole

-- - -

Ir

0 2 4 6 8 10 12

Flow Velocity, feet/sec

Figure D.48

6" Valve 3-Sigma Disc Speed (Orifice Plate at 10D)

- D27 -
3" 300 lb Swing Check Valve

100 _ * ~~~~~~~~~~~~Strong Turbulence

80- 0 WekTublec

X * X * 1|1 ~~~~~~~~~Elbow Up

~60 ffi | 60 - + | | ES}~~ Elbow Down

j 40-

E20

Lw, 0

OD 10-1.5D 2.5D-3D 4.50-50 IOD

Upstream Distance to Disturbance

Figure D.49

3-inch Valve: Relative Severity of Various Upstream Disturbances

Expressed In Terms of 3 Sigma Disc Speed

6" 300 lb Swing Check Valve

100

_ * ~~~~~~~~~~~~Strong Turbulence

so 8- M eakTuzte

X~~~~~~~~~~~~ Eb EUpw 

~40-

220

OD 1 D-1.5D 2.5D-3D 4.5D-5D 1OD

Upstream Distance to Disturbance

Figure 0.50

6-Inch Valve: Relative Severity of Various Upstream Disturbances

Expressed In Terms of 3 Sigma Disc Speed

- D.28 -
NRC FORM 335 US. NUCLEAR REGULATORY COMMISSION 1. REPORT NUMBER

12-89) = ige b NFC. Add Vol.. e~ e NRCM 110:2 VW C tdm tnbma. Io V.

3201.3M BIBLIOGRAPHIC DATA SHEET

(See iflstfucti -on ~NUREG/CR-5583 nndI 

2. TITLE AND SUBTITLE KEI No. 1656

Prediction of Check Valve Performance and Degradation in 3. DATE REPORTPUBLISHED

Nuclear Power Plant Systems -- Wear and Impact Tests 3. j VYEAR

August 1990

4. FIN OR GRANT NUMBER

X________________________________ _ -D2510

S. AUTHOR (S) S. TYPE OF REPORT

Technical

7. PERIOD COVEREDIm*cAu gs;

M.S. Kalsi, C.L. Horst, J.K. Wang, V. Sharma 9/88 - 4/90

L. MrcnRvING'4 unuRGuAIZAHO - NAME~ PIU MULFwncaz ~L v.WW.U~ Or ff~g4I. I&M W -- 5' VWMSWY~l,, ~ P smea and ffWeflln .d~id.

Kalsi Engineering, Inc.

745 Park Two Drive

Sugar Land, Texas 77478

P. SPONSORING ORGANIZATION-NAME AND ADDRESS ifN. type- .me o.wioN to.pvid RCivbion. Office as, ReoCgion. U.& NchaeRguAtorC.mmsln.

and w eddmm)

Division of Engineering

Office of Nuclear Regulatory Research

U.S. Nuclear Regulatory Commission

Washington, DC 20555

10. SUPPLEMENTARY NOTES

11. ABSTRACT 10 0wo.'o ai

Check valve failures in nuclear power plants have led to safety concerns as well as

extensive damage and loss of plant availability in recent years. Swing check valve

internals may experience premature degradation if the disc is not firmly held open

against its stop and significant flow disturbances are present upstream within 10 pipe

diameters. The objective of the current Phase II research was to develop and experimentally verify a quantitative methodology for predicting swing check valve performance

and the degradation of internals caused by hinge pin wear or disc stud impact. Phase I

research had focussed on investigating the stability of the swing check valve disc at

different flow velocities for a wide variety of upstream flow disturbances located

within 10 pipe diameters of the check valve. Valve performance predictions based on

methodology developed as a result of Phase I and II research correlate well with actual

valve operating history at plants. The conservative guidelines provided by this

methodology, tempered and refined by actual performance history and integrated with

preventive maintenance activities, have the potential for significantly improving the

overall reliability of check valves in nuclear power plants.

12. KEY WORDS/DESCRIPTORS sartwo and ophmm shot wil henw&e*s hi csdnu m tpwr.i t o AVAIABILITY STATEMENT

check valve behavior Unlimited

check valve degradation 14. SECURITY CLA3SIFICATIO

aging research rr ")

wear and impact on check valves Unclassified

Ih X4gf.

Unclassified

5. NUMBER OF PAGES

16. PRICE

kRC FORM 335 42491
UNITED STATES

NUCLEAR REGULATORY COMMISSION

WASHINGTON, D.C. 20555

OFFICIAL BUSINESS

PENALTY FOR PRIVATE USE, $300

FOUIL XRTH-CLASS RATES

POSTAGE & FEES PAID

AMNoC

PERMITNo. G J6