NUREG/CR-5720

EGG-2643

Motor-Operated Valve

Research Update

Prepared by

R. Steele, Jr., J. C. Watkins, K. G. DeWall, M. J. Russell

Idaho National Engineering Laboratory

EG&G Idaho, Inc.

Prepared for

U.S. Nuclear Regulatory Commission
t I

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NUREG/CR-5720

EGG-2643

R1, RM, 1S

Motor-Operated Valve

Research Update

Manuscript Completed: October 1991

Date Published: June 1992

Prepared by

R. Steele, Jr., J. C. Watkins, K. G. DeWall, M. J. Russell

G. H. Weidenhamer, 0. 0. Rothberg, NRC Project Managers

Idaho National Engineering Laboratory

Managed by the U.S. Department of Energy

EG&G Idaho, Inc.

Idaho Falls, ID 83415

Prepared for

Division of Engineering

Office of Nuclear Regulatory Research

U.S. Nuclear Regulatory Commission

Washington, DC 20555

NRC FINs A6857, B5529

Under DOE Contract No. DE-AC07-761D01570
ABSTRACT

This report provides an update on the valve research sponsored by the U.S.

Nuclear Regulatory Commission (NRC) that is being conducted at the Idaho

National Engineering Laboratory. The update focuses on the information applicable

to the following requests from the NRC staff:

* Examine the use of in situ test results to estimate the response of a valve at

design-basis conditions

* Examine the methods used by industry to predict required valve stem forces or

torques

* Identify guidelines for satisfactory performance of motor-operated valve

diagnostics systems

* Participate in writing a performance standard or guidance document for

acceptable design-basis tests.

The authors have reviewed past, current, and ongoing research programs to

provide the information available to address these items.

FIN A6857, B5529-Investigation of information used to estimate

valve response, methods used to predict valve stem forces/torques,

guidelines for MOV diagnostics systems.

iii NUREG/CR-5720
CONTENTS

ABSTRACT..................................................................... iii

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

LIST OF TABLES .................. xi

EXECUTIVE SUMMARY .................. xiii

ACKNOWLEDGMENTS .................. xv

1. INTRODUCTION ............................................................ 1

2. GENERAL OBSERVATIONS .................................................. 3

3. SPECIFIC OBSERVATIONS ................................................... 4

3.1 Use of In Situ Test Results to Bound the Response of a Valve at

Design-Basis Conditions ............. ........................ ............ 4

3.2 Assessment of Butterfly Valves Closing against a Compressible Fluid

(Containment Purge and Vent) ........... .................................. 4

3.2.1 Background .................................................... 4

3.2.2 Flow Phenomena through a Butterfly Valve ........................... 5

3.2.3 Existing Butterfly Valve Data and Extrapolation Techniques .... ......... 7

3.2.4 Butterfly Valve Dynamic Flow Testing and Results ...... ............... 8

3.2.5 Butterfly Valve Test Results and Torque Bounding Methods .... .......... 8

3.2.6 Effect of an Upstream Elbow on the Torque Requirements of a

Butterfly Valve .............................................. 22

3.3 Assessment of Wedge-Gate Valves Closing against Medium to High

Flow Conditions . .............................................. 24

3.3.1 Assessment of the Disc Factor Term in the Industry Equation ..... ........ 30

3.3.2 Development of a Correlation to Bound the Stem Force on a Gate

Valve during Closure .37

3.3.3 A Correlation to Bound the Stem Force on a Gate Valve during

Closure .42

3.3.4 Use of the Correlation to Bound the Stem Force on a Gate Valve

during Closure .47

3.3.5 Identifying the Pressure Dependency Contributing to the Stem Force

on a Gate Valve during Closure .51

3.3.6 Low Differential Pressure Test Verification and Bounding of the

Stem Force on a Gate Valve Closing against Design-Basis Flows .51

3.3.7 Nuclear Maintenance Application Center (NMAC) Gate Valve Stem

Force Equation ......... 54

v NUREG/CR-5720
4. UNDERSTANDING DIAGNOSTICS AND DIAGNOSTIC TESTING OF

MOTOR-OPERATED VALVES ................................................. 55

4.1 Overview . ..................................................... 55

4.2 Understanding MOV Diagnostic Testing ........... .......................... 56

4.2.1 Motor Operator Switch Position .......... .......................... 56

4.2.2 Motor Current ................................................... 56

4.2.3 Motor Voltage .................................................. 59

4.2.4 Motor Operator Torque ................ ........................... 62

4.2.5 Valve Stem Force . .............................................. 62

4.3 Load-Sensitive Motor-Operated Valve Behavior ........ ....................... 67

4.4 Direct Current Powered Motor Operators ......... ........................... 79

5. WRITING A PERFORMANCE STANDARD .. 83

5.1 Overview .83

5.2 Detailed Observations .84

6. REFERENCES .. 85

NUREG/CR-5720 vi
'LIST OF FIGURES

1. Effect of low pressure zones on butterfly valve torque-disc oriented with the flat face

of the disc facing upstream ........................................................ 6

2. Effect of low pressure zones on butterfly valve torque-disc oriented with the curved face

of the disc facing upstream . ............. ........................... 6

3. Cross section of Valve 1, the first 8-in. butterfly valve. 9

4. Cross section of Valve 2, the second 8-in. butterfly valve, and Valve 3, the

24-in. butterfly valve .10

5. Containment butterfly valve disc overlay .11

6. Typical installation-butterfly valve uniform inlet flow test section .11

7. Uniform inlet flow butterfly valve positions ............................... 11.......... I

8. Typical installation-butterfly valve nonuniform inlet flow test section .12

9. Nonuniform inlet flow butterfly valve positions .12

10. Butterfly valve differential pressure to upstream pressure ratio versus valve position .13

11. Static pressure 15 diameters downstream of a butterfly valve versus valve position .... ..... 13

12. Torque versus upstream pressure and angle for Valve 1, the first 8-in. butterfly valve,

FFF orientation ............... :: .- 14

13. Torque versus upstream pressure and angle for Valve 2, the second 8-in. butterfly valve,

FFF orientation .......... 15

14. Torque versus upstream pressure and angle for Valve 3, the 24-in. butterfly valve,

FFF orientation .15

15. Torque versus upstream pressure and angle for Valve 1, the first 8-in. butterfly valve,

CFF orientation . ........... ... i.................. - 16

16. Torque versus upstream pressure and angle for Valve 2, the second 8-in. butterfly

valve, CFF orientation. .............. : o 16

17. Torque versus upstream pressure and angle for Valve 3, the 24-in. butterfly valve,

CFF orientation .......... ,.17

18. Peak torque versus static upstream pressure for Valve 1, the first 8-in. butterfly valve 17

19. Peak torque versus static upstream pressure for Valve 2, the second 8-in. butterfly valve 18

20. Peak torque versus static upstream pressure for Valve 3, the 24-in. butterfly valve .18

21. Extrapolation exponent, butterfly valve oriented with the curved face of the disc

facing upstream ............................. 20

Vil NUREG/CR-5720
22. Extrapolation exponent, butterfly valve oriented with the flat face of the disc

facing upstream . .............................................................. 20

23. Actual to predicted butterfly valve torque (percent) as a function of extrapolation

exponent versus valve diameter ratio ............................................. 21

24. Predictions for a 48-in. butterfly valve based on extrapolating the torques of an 8-in.

and a 24-in. butterfly valve at upstream pressures of 15 and 60 psig .................... 23

25. Torque versus upstream pressure and angle for Valve 3, the 24-in. butterfly valve

in the CFF orientation .............. ........................................ ... 25

26. Peak torque versus static upstream pressure for Valve 3, the 24-in. butterfly valve,

comparing the response of the peak torque with elbow and peak torque without

elbow orientations . ............................................................ 25

27. Comparison of the standard industry gate valve stem force equation with selected

test results .................................................................. 29

28. Comparison of the NMAC gate valve stem force equation with selected test results ........ 30

29. Disc factor for Gate Valve 2 closing on line break flow, effect of subcooling at

1000 psig . .................................................................. 32

30. Disc factor for Gate Valve 2 closing on line break flow, effect of pressure at

100l F subcooling . ............................................................ 33

31. Disc factor for Gate Valve 2 opening on line break flow, effect of pressure at

100WF subcooling . ............................................................ 34

32. Gate valve disc cross section showing pressure forces and measurement locations .... ..... 36

33. Gate valve internal pressure distribution ............................................ 36

34. Gate valve disc cross section showing unbalanced forces just before wedging ...... ....... 37

35. Resolution of horizontal and vertical gate valve disc forces into surface normal

and sliding forces . ............................................................ 39

36. Gate-valve stem-force history, indicating severe damage to the guides ................... 40

37. Gate-valve stem-force history, indicating damage to the guides but not to the seats .... ..... 41

38. Normalized surface sliding force versus normalized surface normal force for

gate valves closing against less than 70'F subcooled fluid . ............................. 45

39. Normalized surface sliding force versus normalized surface normal force for gate.

valves closing against 70'F, or greater, subcooled fluid ............................... 46

40. Stem-force history of a small gate valve during an opening test ........................ 48

41. Stem-force history of a large gate valve during an opening test .......... .. ............. 49

NUREG/CR-5720 viii
42. Gate valve load-dependent friction factor using the INEL correlation .................... 53

43. Limit switch history showing stroke time for a position-controlled valve. Limit

switch actuation points are important when analyzing other data .............. ......... 57

44. Torque switch history showing torque switch trip and the termination of current flow

to the motor controller holding coil .57

45. Current history from a design-basis test showing current increasing as the valve closes;

the rapid increase in current indicates torque switch trip. The test began with the valve

75% open .58

46. Motor torque-speed curve showing the speed-torque-current relationship for the ac

motor test results shown in Figure 45. Beyond the knee of the curve, the torque increase

is small in proportion to the speed loss and the current increase. For some motor

configurations there is no increase; for some there is actually a decrease in torque beyond

the knee of the speed curve.58 ........ '. I

47. Current history from a design-basis test showing the motor going into a stall, saturating

the current transducer. The test began with the valve 30% open ......................... 59

48. Motor torque-speed curve showing the speed-torque-current relationship for a dc motor.

The torque (load) continues to increase as the speed drops and the current increases. .. 60

49. Disc position histories from three dc-powered closing tests showing that the stroke time

is longer with higher loads .60

50. Motor heating under load caused the current demand of this dc motor to decrease at a

rate of 1 amp per second during the 20-second test, resulting in a decrease in the output

torque of approximately 1 ft-lb every 2 seconds . .61

51. Disc position histories from four dc-powered MOV closing tests, showing that the stroke

time is longer with higher loads. Three of the tests did not fully seat the valve, and the test

at the highest load stalled the motor ................................................... 61

52. Ratio of torque spring force to stem torque from MOVLS Test 1, with a small

stem force ........ ;..... 63

53. Ratio of torque spring force to stem torque from MOVLS Test 4 at a higher stem force ..... 63

54. Ratio of torque spring force to stem torque from MOVLS Test 5, at a high stem force, in

which final seating was not achieved, showing that the average torque spring force to

stem torque ratio was not influenced by the load on the operator .64

55. Stem force history from MOVLS Test 4, showing a 3000 lbf margin between the peak

stem force and the stem force at final seating .64

56. Stem force'history'from MOVLS Test 5, showing that a 1000 lbf increase in the initial

stem force eliminated the 3000 lbf final seating stem force margin observed in

MOVLS Test 4 .65

ix NUREG/CR-5720
57. Typical torque spring calibration plot shows the initial force offset at zero deflection.

This figure also confirms that for this spring, the spring force to deflection relationship is

linear. The actual values are omitted from this figure because the data are proprietary

until the MOV diagnostic equipment validation effort is complete ...................... 65

58. MOVLS layout drawing .......................... ............................. 68

59. Traces of thrust versus time for Tests 8 through 10 of the MOVLS test sequence. These

are identical low-thrust tests .................................................... 69

60. Traces of thrust versus time for Tests 10 through 14 of the MOVLS test sequence ......... 69

61. Traces of thrust versus time for the final three tests of the MOVLS test sequence.

Tests 12-14, which exhibit load-sensitive MOV behavior, were followed by a series of

low-thrust tests (Tests 15-17), with resistance to closing much like the first three of

the series .... 70

62. The design of the operator allows two potential output motion paths ......... ........... 71

63. The application of spring pack force to the operator mechanism at an offset distance

from the stem centerline generates the torque input to the mechanism ................... 73

64. Potential causes of torque loss in the operator are identified .......... .. ............... 73

65. Traces of spring pack force versus time for Tests 8 through 10 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 59 ......................... 75

66. Traces of spring pack force versus time for Tests 10 through 14 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 60. Spring pack force does

not change significantly from test to test ..... ...................................... 75

67. Traces of spring pack force versus time for Tests 14 through 17 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 61. Spring pack force does

not change significantly from test to test .......... I ................................. 76

68. The ratio of spring pack force to stem torque for Tests 8 through 10 of the MOVLS

test sequence. These traces correspond to the thrusts of Figure 59 ...................... 76

69. The ratio of spring pack force to stem torque for Tests 10 through 14 of the MOVLS

test sequence. These traces correspond to the thrusts of Figure 60 ....................... 77

70. The ratio of spring pack force to stem torque for Tests 14 through 17 of the MOVLS

test sequence. These traces correspond to the thrusts of Figure 61. Spring pack force

does not change significantly from test to test ................................ 77

71. Traces of coefficient of friction in the stem nut versus time for the first three tests of

the MOVLS test sequence. Very little change in coefficient of friction is observed at

the low thrust levels of these tests ................................ ' : 78

72. Traces of coefficient of friction in the stem nut versus time for Tests 10 through 14 of

the MOVLS test sequence. A steady increase in coefficient of friction is evident in these

tests. Coefficient of friction is highest during Test 14, the test exhibiting load sensitive

MOV behavior ............................................................ 78

NUREG/CR-5720 x
73. Traces of coefficient of friction in the stem nut versus time for the last four tests of

the MOVLS test sequence. A full opening/closing low-thrust cycle (Test 15) is needed

following the test exhibiting load-sensitive MOV behavior (Test 14) before the

coefficient of friction returns to the levels recorded for the initial low-thrust level tests ...... 80

74. Typical stem thrust histories from the dc-powered MOVLS tests ....... ................ 80

75. Typical stem nut coefficient of friction histories from the dc-powered MOVLS test.

Note the increase in friction between low-load Test 1 and high-load Test 12. The load

was doubled between Tests 12 and 15 without a significant increase in stem nut coefficient

of friction .................................................... 82

LIST OF TABLES

1. Comparison of torque prediction methods butterfly versus valve orientation .22

2. Ratio of peak torque to uniform flow peak torque for a butterfly valve in the

CFF orientation .23

3. Phase I gate valve flow interruption test temperatures and pressures ........ ............. 26

4. Phase II gate valve flow interruption test temperatures and pressures ....... ............. 26

5. Phase II gate valve test data supporting extrapolation .................. .............. 43

6. Phase I gate valve test data supporting extrapolation . ................................ 44

7. Comparison of gate valve stem forces, estimates versus actual .......... ............... 52

xi NUREG/CR-5720
EXECUTIVE SUMMARY

The U.S. Nuclear Regulatory Commission

(NRC) is supporting motor-operated valve

(MOV) research at the Idaho National Engineering Laboratory (INEL). The MOV test programs

performed as part of the research provide the

basis for assessing the effects various factors have

on the valves and for evaluating current industry

standards. This report discusses several research

items, including

* Use of in situ test results to estimate the

response of a valve at design basis conditions

* Methods used by industry to predict

required valve stem forces or torques

* Guidelines for satisfactory performance of

MOV diagnostics systems

* Composition of a performance standard or

guidance document for acceptable designbasis tests.

We have reviewed all of our past, current, and

ongoing valve research that included full-scale

valve test programs to address these items. This

review revealed that (a) use of in situ test results

to estimate the response of gate and butterfly

valves at design-basis conditions is possible, but a

long list of caveats exists; (b) the methods used

by industry to predict the required stem force for

a gate valve and the required stem torque for a

butterfly valve are incomplete; (c) satisfactory

performance of MOV diagnostic systems is

possible, but very few of the currently available

systems measure enough parameters to be completely effective; and (d) our participation in

writing a performance standard or guidance document for acceptable design-basis tests requires us

to continue to exchange information with the

American Society of Mechanical Engineers

standards writing committees.

The research documented in this report forms

the basis for the technical presentations given at

the NRC inspector training courses, and has also

enabled the INEL to create a PC-based computer

program to assist NRC personnel during their

MOV inspections.

.x.i. NUREG/CR-5720
ACKNOWLEDGMENTS

Researchers at the Idaho National Engineering Laboratory (INEL) developed the

basis for the INEL gate valve stem force correlation presented in Section 3.3 and

have presented the results to a wide range of industry and utility experts for review.

We wish to thank the following reviewers, who responded with comments and

insights to this effort:

Siemens/Kraftwerk Union Dr. Nabil Schauki, Dr. Norbert Rauffmann,

Dr. U. Simon, Helmut Knoedler

Duke Power Neal Estep, W. Scheffler

Consumers Power

Portland General Electric

Nuclear Management and

Resources Council

Electric Power Research

Institute

Idaho National Engineering

Laboratory

George Smith

M. Lee Kelly

Review Team

John Hosler

Dr. Romney Duffy, Tim Boucher, Vic Berta.

We would also like to thank the following individuals from Limitorque who have

reviewed our research on load-sensitive, motor-operated valve behavior: Bob

Bailey, Pat McQuillan, Mike Bailey, Greg Pence, Charles Cox, and Walt Sykes.

Their comments and insights are appreciated.

xv xv ~~~~NUREG/CR-5720
Motor-Operated Valve Research Update

1. INTRODUCTION

The Idaho National Engineering Laboratory

(INEL) is performing motor-operated valve

(MOV) research in support of the U.S. Nuclear

Regulatory Commission's (NRC's) efforts

regarding Generic Issue 87, "Failure of HPCI

[High-Pressure Coolant Injection] Steam Line

Without Isolation," and Generic Letter 89-10,

"Safety-Related Motor-Operated Valve Testing

and Surveillance."

This report updates the research reported in

NUREG/CR-5558, Generic Issaie 87 Flexible

Wedge Gate Valve Test Program Phase II

Results and Analysis (Steele et al., 1990). This

update also provides research results on program

objectives not covered completely in that report.

These objectives include

* Examine the use of in situ test results to

estimate the required forces of a valve at

design-basis conditions

* Examine the methods used by industry to

predict required valve stem forces or

torques

* Identify guidelines for satisfactory performance of MOV diagnostics systems

* Participate in writing a performance standard or guidance document for acceptable

design-basis tests.

In preparing this report, we reanalyzed selected

results from all four of the full-scale valve test

programs performed for the NRC Equipment

Operability Program, as well as results from the

ongoing separate effects testing currently being

conducted at the INEL. The full-scale valve test

programs include

* The Test Program for Containment Purgeand-Vent Valves is reported in NUREG/

CR-4648, A Study of Typical Nuclear

Containment Purge Valves in an Accident

Environment (Watkins et al., 1986). In this

test program, three butterfly valves, two

8-in. and one 24-in., were tested at line

break flows at closing differential pressures

of 5 to 60 psig.

The testing reported in NUREG/CR-4977,

SHAG Test Series: Seismic Research on an

Aged United States Gate Valve and on a

Piping System in the Decommissioned

Heissdampfreakior (HDR) (Steele and

Arendts, 1989). A 25-year old, 8-in.,'dc

powered, motor-operated Crane gate valve

was refurbished and installed in an experimental reactor flow loop. Test loadings

included flow, pressure, temperature, and

seismic excitations.

* The Phase I Generic Issue 87 Test Program

reported in NUREG/CR-5406, BWR [Boiling Water Reactor] Reactor Water Cleanup

System Flexible Wedge Gate Isolation Valve

Qualification and High Energy Flow Interruption Test Program (DeWall and Steele,

1989). This test program subjected two

6-in., motor-operated, flexwedge, containment isolation gate valves to (a) the applicable hydraulic qualification test outlined in

American National Standards Institute

(ANSI)/American Society of Mechanical

Engineers (ASME) B 16.41, (b) a test at

normal reactor water cleanup (RWCU)

system flow and pressure, and (c) tests at

full line break flows at design-basis and

other parametric fluid conditions.

* The Phase II Generic Issue 87 Test Program, reported in NUREG/CR-5558 (Steele

et al., 1990). This test program subjected

three 6-in. and three 10-in. flexwedge,

motor-operated gate valves to (a) applicable

hydraulic qualification tests outlined in

ANSI/ASME B16.41, (b) tests at normal

system pressure and flow, and (c) full line

break flow tests at design-basis and other

I 1 ~~~~NUREG/CR-5720
Introduction

parametric fluid conditions for the BWR

containment isolation valves. The valves

were representative of those installed in the

BWR HPCI turbine steam supply line, the

RWCU system, and the reactor core

isolation cooling turbine steam supply line.

The ongoing separate effects testing in our

laboratory uses the INEL motor-operated valve

load simulator (MOVLS). This device uses actual

motor operators and valve stems, which are

loaded using a hydraulic cylinder to produce

rising stem valve loadings typical of those we

have observed in field testing. The MOVLS is

currently instrumented to directly measure the

following:

* Force and torque on the stem

* Position of the stem

* Force on the torque spring

* Displacement of the torque spring

* Torque and limit switch actuation

* Root mean squared (rms) current and

voltage

* Peak to peak current

* Power, power factor, and speed of the

electric motor.

Through the Data Acquisition System, the

MOVLS can also display real time calculations

such as stem factor.

The design and calibration of the MOVLS has

been upgraded from strictly a research device to a

standard that can be used by the industry to evaluate their diagnostic equipment. Current research.

includes load-sensitive motor-operator behavior

and comparison testing of ac- and dc-powered

motor operators.

NUREG/CR-5720 2
2. GENERAL OBSERVATIONS

Manufacturing tolerances, piping geometry

effects, and specific installation effects all

influence whether a valve can be grouped with

other like valves and categorically qualified by

similarity. The following examples provide some

insights into why we believe such influences need

to be considered:

* When the same 6-in. gate valve was tested

with flow in one direction, then refurbished

and tested with flow in the other direction,

the failure mechanism changed from

guiding surface failure to seating surface

failures. We believe this to be the result of

the internal valve tolerances stacking up

differently.

* A 6-in. and a 10-in. gate valve from the

same manufacturer were design-basis flow

tested. The 6-in. valve passed the tests with

good results. However, with the 1 0-in. valve

the flow forces plastically deformed the

valve body guide rails, which increased the

required stem force to close. Upon disassembly and inspection of the valve, it was

observed that the guide rails were not

welded far enough down into the body. The

current nuclear valve qualification standard

ANSI/ASME B16.41 would have allowed

similar 3- to 12-in. valves to be qualified,

based on the 6-in. qualification test.

Two like 4-in. gate valves were recently disassembled, inspected, and tested by a utility.

The internal tolerances of both valves were

within the manufacturer's tolerances, but

the utility lapped and polished the second

valve until it was as similar to the first valve

as possible. The first valve was subjected to

both no-flow static tests and design-basis

flow tests. The second valve was also

subjected to no-flow static tests and

required a higher torque switch setting to

produce the same stem force as the first

valve. Both motor operators had been

dynamometer tested by Limitorque,a showing equal output for the same torque switch

settings. The need for the higher torque

switch setting may have been caused by a

difference in the stem factor.

* Testing of containment purge-and-vent

valves revealed that butterfly valves

installed in certain orientations downstream

of an elbow could require up to 133% of the

torque required to close the same valve in a

straight piece of pipe.

* Two like valves at the same utility, but at

two different units, performed quite differently. One valve burned up two dc motors.

The fault was traced to cable sizing. The

cables supplying power to the valve were

undersized, resulting in excessive voltage

drop and motor stall. The cable size at the

other unit was larger,'thus avoiding'the

voltage drop, motor stall, and subsequent.

motor burnout.

The list could go on, but the point is that we do

not know that like valves will behave alike until

sufficient testing can show that they will perform

in the -generic group. Since all valves cannot be

design-basis tested in situ, we believe it will be

necessary to conduct both testing and analysis to

prove that a given valve can be related to a

prototype test or a generic qualification group.

a. Mention of specific products and/or manufacturers in this document implies neither endorsement or

preference nor disapproval by the U.S. Government,

any of its agencies, or EG&G Idaho, Inc., of the use

-of a specific product for any purpose.

3 3 ~~~~NUREG/CR-5720
3. SPECIFIC OBSERVATIONS

The first two research update objectives, (a) to

examine the use of in situ test results to estimate

the response of a valve at design-basis conditions,

and (b) to examine the methods used by industry

to predict required stem force or torques, will be

covered in this section. At this time, the INEL can

provide information only on butterfly valves used

in purge-and-vent-valve applications and on

wedge-gate valves used in a number of mediumand high-flow applications. Testing currently

being performed in Europe and at selected

utilities in the United States may provide

information on other valve designs and flow

applications later this year.

We have found it useful to distinguish between

predictable valves (those whose performance is

repeatable) and nonpredictable valves (those

whose performance is not repeatable, usually

because they experience internal damage when

subjected to high loads during operation). In the

previous section, we discussed the pitfalls

associated with predicting the performance of

various kinds of valves, both predictable and nonpredictable. The remainder of this report will be

limited to discussing predictable valves only.

Valves that do not exhibit predictable behavior

under load are discussed extensively in NUREG/

CR-5558 (Steele et al., 1990).

3.1 Use of In Situ Test Results

to Bound the Response of a

Valve at Design-Basis

Conditions

The results of INEL testing indicate that the

response of a valve can be bounded for specific

valve types and fluid conditions from in situ test

results obtained from either small-scale test

valves or low differential pressure tests. The

results of recent European testing support this

conclusion. The fact that valves of a given type

respond linearly with pressure for a specific fluid

condition leads one to believe that, with a sufficient amount of testing, bounding the response of

a valve is possible. On the other hand, the

equations the industry has used in the past to predict the perfonnance of gate and butterfly valves

are incomplete. The INEL has confidence in

bounding the stem force of predictable wedgegate valves closing against medium to high flows

and in bounding the torque of high aspect ratio

offset disc butterfly valves used in purge-andvent applications closing against compressible

flows.

3.2 Assessment of Butterfly

Valves Closing against a

Compressible Fluid

(Containment Purge and

Vent)

3.2.1 Background. The expression "butterfly

valve" is a generic term for a rotating-disc, in-line

valve. Of interest to this discussion is the

application of butterfly valves in nuclear containment purge-and-vent systems. These systems

penetrate the containment boundary and allow air

to circulate through the containment; however, in

the event of an accident, these systems must close

to isolate the environment inside the containment.

Consequently, the butterfly valve installed in

these systems must be functional both during and

following an accident. Industry operability

assumptions have been based to a large degree on

empirical information obtained from work with

incompressible fluids or from small valves tested

with compressible fluids. Previous experimental

work with incompressible fluids has, for the most

part, been done at very low pressures, with very

low pressure drops with large valves, or with

small valves. The operability issue concerning

containment purge-and-vent valves was raised

after the Three Mile Island Unit 2 incident. The

first question dealt with valve actuator sizing:

Would an actuator stall and fail to close a valve

because of the dynamic loads that might be

produced by a high differential pressure across

the containment boundary resulting from a

design-basis loss-of-coolant accident? The

second question dealt with stress margins to withstand the loads imposed during the closure. The

NUREG/CR-5720 4
Specific Observations

stress margins are totally dependent on the analysis of the predicted loads and are not part of this

discussion.

3.2.2 Flow Phenomena through a Butterfly

Valve. As a part of the program for testing butterfly valves, reported in NUREG/CR-4648

(Watkins et al., 1986), we conducted a study into

the physics of a butterfly valve closing against a

compressible flow to better understand the torque

response of a butterfly disc. We observed that a

butterfly disc responds as an airfoil from the fullopen position to the position where maximum lift

occurs (peak torque). With the exception of a disc

that is perfectly symmetric and oriented so that

flow is evenly split over both faces, the flow that

passes around the disc, and the resulting flow

perturbation and pressure distribution, will

depend not only on the degree of closure of the

valve, but also on which surface of the disc is

closing into the flow. The torque characteristics of

a valve are the result of the pressure that acts over

the surface of the disc, which must be counterbalanced to move or control the motion of the

disc. With the valve partially closed and the disc

oriented so that the flat face of the disc is facing

upstream when the valve is fully closed

(Figure 1), the flow will separate around the disc,

with vortices and a low-pressure region developing at location "L." The result of this low-pressure

region is a torque acting on the disc in the opening

direction. However, if the valve is partly closed,

with the disc oriented so that the curved face of

the disc is facing upstream when the valve is fully

closed (Figure 2), the flow separation and

resultant low-pressure region (location "L") will

result in a torque acting on the disc in the closing

direction. However, the fact that this torque acts in

the closing direction is not necessarily helpful or

even benign, because most butterfly valve designs

do not have positive stops. Without actuator

control, the disc would go beyond the fully closed

position to some partially open position in the

other direction.

Immediately after peak torque occurs,. the

airfoil will stall. The disc then becomes an

increasingly larger flow obstruction, with a more

uniform pressure distribution and decreasing

torque requirements as the valve completes its

closing cycle. Based on this hypothesis, one.

might believe that the pressure acting on a disc

can be directly related to the torque for any degree

of valve closure, any type of fluid, or any flow

velocity through the valve. Unfortunately, test

results indicate that, unlike many other valve

designs, there are no proven equations for

predicting the torque required for a butterfly

valve to close against a compressible fluid flow.

Torque predictions must come from valid testing

and extrapolation. The validity of test results is

influenced by the following:

* A compressible flow medium must be used

during the tests because of the phenomena

of flow separation around a disc and the

creation of a low-pressure region. An

incompressible fluid will usually cavitate,

resulting in a less pronounced low-pressure

region. If the fluid is cavitating, the low-

;pressure region acting on the disc will be

limited to the vapor pressure of the fluid.

* The inlet pressure, rather than the differential pressure, should be used for

extrapolation, because flow rates through

the valve that cause the flow to become

choked can result in supersonic flow downstream of the valve. The sonic plane in the

valve will depend heavily on the degree of

valve closure, the shape of the disc surface,

the relative location of the valve seals to the

disc, and the properties of the fluid. These

effects should be minimal, resulting in a

torque extrapolating ability for a given

valve, provided the smaller valve is

geometrically representative of the larger

valve.

* The flow rates and pressure drops achieved

during testing must be typical of conditions

expected in the actual plant installation. If

the flow rate through a valve is too low, it

will produce a minimal pressure drop,

which in turn eliminates the effect of

compressibility. Consequently,.the flow

field through the valve and around each side

of the disc will be much more uniform, as

will the resultant pressure profile acting on

5s NUREG/CR-5720
Specific Observations

0 -4 Flow

Q

Resultant

torque

6 3210

Figure 1. Effect of low pressure zones on butterfly valve torque-disc oriented with the flat face of the

disc facing upstream.

QD

Resultant torque -4 Flow

6 3202

Figure 2. Effect of low pressure zones on butterfly valve torque-disc oriented with the curved face of

the disc facing upstream.I

NUREG/CR-5720 6
Specific Observations

the disc. Thus, the resultant torque will be

less than expected at high flow.

* For test purposes, the valve should always

be oriented so that the curve face of the disc

would be facing upstream when the valve is

fully closed: This orientation results ina

bounding extrapolation of the closing

torque using the methods proposed in this

report, whereas the other valve orientation

does not.

* For very low flow applications, the last

20 degrees of disc closure should be

included in the test, because the torque reaction from the seals and bearings in the valve

may be much larger than the corresponding

dynamic torque and will always work

against valve closure.

3.2.3 Existing Butterfly Valve Data and

Extrapolation Techniques. An analytical

assessment of the loads on a butterfly valve

resulting from the increasing pressure environment of a design-basis accident is difficult

because of the complex geometries and flows

through such a valve and the lack of empirical

information on the dynamic response of a

butterfly valve in a compressible fluid flow. Also,

nonuniform inlet flow configurations will impact

the dynamic response of a butterfly valve. Consequently, we performed a number of tests to

determine the response characteristics of nuclear

containment isolation (butterfly) valves during

accident conditions, in an attempt to ensure that

the flow dynamics of such a valve were understood. Prior to this work, the only public domain

information was that produced for the Allis

Chalmers Company by the National Aeronautics

and Space Administration Langley Research

Center (Allis-Chalmers Corporation, 1979). The

Langley Research Center conducted this testing

using a 6-in. valve body and three interchangeable butterfly valve discs. The program did not

include testing a larger-sized valve to verify the

extrapolation theory. Also, Langley performed

the testing for a specific vendor, and did not

compare the results to other vendor designs.

Because of the very high volumetric flow rates

of concern, testing a large butterfly valve under a

simulated accident environment is not always

economical or even feasible. Consequently, the

valve manufacturers have developed a method for

testing scale-model test valves and then

extrapolating their performance to predict the

torque requirements of a larger valve. All of the

extrapolation techniques used to predict the

torque requirements of a larger butterfly valve

have a common underlying assumption as to the

nature of the flow. Specifically, the flow is

assumed to be quasi-one-dimensional, and the

response is assumed to be linear. Again,

unfortunately, the flow through a butterfly valve

has very real and complicated three-dimensional

flow perturbations; therefore, an inherent

compromise must be accepted when extrapolating the performance of a scale-model test valve

to a larger valve. To further complicate the flow

field, the effects of compressibility must be

acknowledged. Compressibility effects can cause

the flow through a valve to become choked and

allow the downstream pressure to vary independent of the upstream pressure.

Each of the extrapolation techniques used by

the manufacturers contains a common extrapolation term that relates the size of a large valve to

the size of the scale-model test valve. This term is

the cube of the nominal diameter of the valve

being predicted, divided by the cube of the

nominal diameter of the scale-model test valve.

Later sections of this report will show that this

term will yield a bounding prediction, if data are

used with the scale-model test valve oriented so

that the curved face of the disc is facing upstream

when the valve is fully closed. A nonbounding

prediction could result if the flat face of the disc is

facing upstream when the valve is fully closed.

Most of the extrapolation techniques have a

differential pressure term. The manufacturers

assume that the differential pressure across a

scale-model test valve will be the same as in a

larger valve. Consequently, if the differential

pressure term is eliminated from the equation, the

torque extrapolation technique reduces to the

torque being a direct function of the diameter

ratio cubed.

7 NUREG/CR-5720
Specific Observations

For our testing, we selected three nucleardesigned butterfly valves typical of those used in

commercial nuclear power plant containment

purge-and-vent applications. For a comparison of

response, we tested two 8-in. nominal-pipe-size

butterfly valves with differing internal designs.

For extrapolation insights, we also tested a 24-in.

nominal-pipe-size butterfly valve (made by the

same manufacturer who made one of the 8-in.

valves). Figures 3 and 4 show cross-sectional

views of the valves, which were ASME Code

Class III, ANSI 150 pound class, in-line, off-set

disc, elastomer-sealed, high aspect ratio butterfly

valves (thickness of the disc is relative to the

diameter). These valves are typical of designs up

to, and including, 24-in. nominal diameter. In

Figure 3, the elastomer seal is part of the body; in

Figure 4, the elastomer seal is part of the disc

assembly. Figure 5 is a composite cross-sectional

view of all three discs, with the 24-in. disc

reduced by a factor of 3.

3.2.4 Butterfly Valve Dynamic Flow Testing

and Results. The results of the testing of butterfly valve dynamic flow were analyzed to assess the

butterfly valve closing torque extrapolation

methodology used by the industry and to quantify

the influence of piping geometry on the torque

response of a butterfly valve.

-We performed the experiments with various

piping configurations and disc orientations to

simulate various installation options that could be

encountered in the field. As a standard for

comparing the effects of the various installation

options, we initially performed our testing in a

standard ANSI test section. The nominal or

uniform inlet flow ANSI test section is shown in

Figure 6, and the valve test positions for this configuration are shown in Figure 7. During the flat

face forward (FFF) tests, the disc was oriented so

that the flat face of the disc would be facing

upstream when the valve was fully closed. Likewise, during the curved face forward (CFF) tests,

the disc was oriented so that the curved face of the

disc would be facing'upstream when the valve

was fully closed.

The test section for nonuniform inlet flow is

shown in Figure 8; the valve test positions for

this configuration are shown in Figure 9. As the

flow bends around the elbow immediately

upstream of the test valve, the resultant flow

profile results in higher velocities near the outside

radius of the elbow. Unlike the uniform inlet flow

test section, this nonuniform flow profile will

interact with the disc differently, depending on

the direction in which the disc is rotating toward

the closed position. As such, the clockwise (CW)

and counterclockwise (CCW) notations

associated with the nonuniform inlet flow tests

identify orientations with the disc rotating clockwise or counterclockwise relative to the figure.

Each test was performed while the valve

upstream pressure was controlled at a relatively

constant pressure throughout the valve closure

cycle. Each test cycle consisted of stabilizing the

valve upstream pressure with the valve' in the

fully open (90-degree) position. We then closed

the valve at 18 degrees/second to the fully closed

(0-degree) position and reopened it after a 250-ms

delay. Testing cycles were performed with the

upstream pressure varied up to the design-basis

pressure of 60 psig while monitoring the position

and torque of the valve shaft, the mass flow rate,

and the temperature and pressure at various locations throughout the system (shown in Figures 6

and 8).

3.2.5 Butterfly Valve Test Results and

Torque Bounding Methods. The test results

were first reduced and entered into a computer

data base. This data base was then used as a

common base for presentation, comparison,

analysis, and plotting. The ratio of valve differential pressure to upstream pressure versus valve

position was plotted at pressures of 15, 30, 45,

and 60 psig for one of the 8-in. CFF valve tests.

The results (Figure 10) indicate that the response

for the 15-psig test is very different from the

response for the higher-pressure tests. However,

the response for the higher-pressure tests is very

similar beyond the 40-degree position. This

difference is indicative of choked flow at the

higher test pressures and results in a supersonic

flow region downstream of the valve that fluctuates as the valve closes. This, in turn, results in a

downstream pressure profile (Figure 11) that is

NUREG/CR-5720 8
Specific Observations

Shaft

Shaft seals

- EPT seal ring

. Metal seat

Purge valve INEL 4 0488

Figure 3. Cross section of Valve 1, the first 8-in. butterfly valve.

9 NUREG/CR-5720
Specific Observations

Shaft seals

Metal seat

-EPT seal ring

Figure 4. Cross section of Valve 2, the second 8-in. butterfly valve, and Valve 3, the 24-in. butterfly valve.

NUREG/CR-5720 10
Specific Observations

Valve 1 (full scale)

Valve 2 (full scale)

-- - Valve 3 (1/3 scale)

------ ---

- ...-

f ~~~~~~~~~~--,-4 IN,

Sealing surface

5 2979

Figure 5. Containment butterfly valve disc overlay.

TE-1 02 . PT-1 02

ET6 0 V-102D

PT-106 |TE10

PT

PDT

TE

TV

= pressure

= differential pressure transducer

= temperature

= test valve

PT-C

OQ-ORupture disc

TE-M

TE- K, PT-K 0 0

PT-D 00

PT-J, TE-JPT-H

.. TE-H

PT-G

T

0

TE-G

PT-A

PT-B 0 ~~PT-M

TE-C TE-D

IPDT I

PT-E PT-F _ a 0 a 0 O~

TE-E TE-F5 2999

Figure 6. Typical installation-butterfly valve uniform inlet flow test section.

PFlow-i.-ft

Flat face forward

(FFF)

. ~~~~~~~~~~~~~. Flow - .-.- .

Curved face forward

(CFF)

5 2980

Figure 7. Uniform inlet flow butterfly valve positions.

11 NUREG/CR-5720
Specific Observations

TE-I 021 ao If PT-1 02

PT = pressure

PDT = differential pressure transducer

TE = temperature

V- 02D TV = test valve

Rupture disc

PT-J, TE-J

PT-B TE-D TE-E TE-G TE-H TE F

LPDT J 5 4000

Figure 8. Typical installation-butterfly valve nonuniform inlet flow test section.

FFF-CCW _FF-CCW

FFF-CW IFFFCW 5 2981

Figure 9. Nonuniform inlet flow butterfly valve positions.

NUREG/CR-5720 12
Specific Observations

0

jor

':L an

9U 80 70 60 50 40 30 20 10 0

Valve position 5 2982

Figure 10. Butterfly valve differential pressure to upstream pressure ratio versus valve position.

40

30

0.

-a

' 20

a)

0L

90 80 70 60 . 50 40 30 20 10 0

Valve position 5 2983

Figure 11. Static pressure 15 diameters downstream of a butterfly valve versus valve position.

13 NUREG/CR-5720
Specific Observations

very different from that experienced during

incompressible flow and very different for each

of the valves tested. Specifically, the downstream

pressure, as measured 15 diameters downstream

of each test valve, did not always recover from

the flow perturbation during certain portions of

the valve closure cycle. This indicates that the

measurement location was in a supersonic region.

These results suggest that torque extrapolation

practices using the differential pressure do not

account for supersonic flow downstream of the

valve and its resulting effect on valve torque

during a design-basis accident. Therefore, we

introduced a new parameter (upstream pressure)

and developed plots of valve response, relating

valve upstream pressure, dynamic torque, and the

position of the disc. Figures 12 through 17 are

the response plots for the three valves tested in the

uniform inlet flow configuration. The figures for

the CFF orientation indicate a butterfly valve

responding with a positive or self-closing torque.

In this orientation, the operator must supply

torque to keep the valve from shutting too rapidly.

The figures for the FFF orientation indicate a

butterfly valve responding with a negative or selfopening torque. In this orientation, the operator

must supply torque to close the valve. Therefore,

butterfly valves in the FFF orientation will be

harder to close, and butterfly valves in the CFF

orientation will be harder to open.

Analysis of the response plots shows that the

magnitude of the dynamic torque when the valve

was in the CFF orientation (a positive response)

was greater than the magnitude of the dynamic

torque when the valve was in the FFF orientation

(a negative response). Also, the positive dynamic

torque curves of the three valves in the CFF

orientation, as shown in Figures 12 through 14,

are very similar in appearance. Conversely, the

negative dynamic torque curves of the three valves

in the FFF orientation, as shown in Figures 15

through 17, are very different in appearance. This

provides some assurance that limited extrapolation is possible using the upstream pressure (rather

than the pressure drop) across the butterfly valve

if the valve is in the CFF orientation.

The peak torque for each of the three butterfly

valves tested with a uniform inlet flow configuration was plotted against upstream pressure

in Figures 18 through 20. The results indicate

r- Angle

-50

.0

0

100

- 150L_

Figure 12. Torque versus upstream pressure and angle for Valve 1, the first 8-in. butterfly valve, FFF

orientation.

NUREG/CR-5720 14
Specific Observations

Angle

-50

C,

0~

0

1 -

5 2986

-1o00 o

-150 -

Figure 13. Torque versus upstream pressure and angle for Valve 2, the second 8-in. butterfly valve, FFF

orientation. -

Angle

Or

-500 1

~0

0

I-

-1000 1-

-1500 F-

-2000 L

Figure 14. Torque versus upstream pressure and angle for Valve 3, the 24-in. butterfly valve, FFF

orientation.

15 NUREG/CR-5720
Specific Observations

60

Ic~r

-.0

.0

0

I3

50pI

0

Angle

10

5 2988

Figure 15.

orientation.

Torque versus upstream pressure and angle for Valve 1, the first 8-in. butterfly valve, CFF

100o r

-o

c

50 -

30

0 O L

90 80 70 60 50

Angle

40 30

10

5 2989

Figure 16. .Torque versus upstream pressure and angle for Valve 2, the second 8-in. butterfly valve, CFF

orientation.

NUREG/CR-5720 16
Specific Observations

2000 -

1500

-

-0I

ET

45

1000 _

30

500 15

a I

Angle

Figure 17. I'Torque versus upstream pressure and angle for Valve 3, the 24-in. butterfly valve, CFF

orientation.

200

160

120

80

'a 40

40

0 -40

-80

-120

- 160

* -200

0 10 20 30. 40 50 60 70 80 90

Static upstream pressure (psia) 5 2991

Figure 18. Peak torque versus static upstream pressure for Valve 1, the first 8-in. butterfly valve.

17 NUREG/CR-5720
Specific Observations

200 III I I I I I

160 ;-

120

80 1

-,

-0

0*

0

1-

40 1

C~~~FF

0 0

FFF

0

-40 1

-80

- 120

- 160

-200 I I l

0 10 20 30 40 50 60

Static upstream pressure (psia)

70 80 90

5 2992

Figure 19. Peak torque versus static upstream pressure for Valve 2, the second 8-in. butterfly valve.

3200

-0.

0

Hr

30 40 50 60

Static upstream pressure (psia)

90

5 2993

Figure 20. Peak torque versus static upstream pressure for Valve 3, the 24-in. butterfly valve.

NUREG/CR-5720 18
Specific Observations

that a linear relationship exists between the two

parameters, although the angle where peak torque

occurs varies with pressure. The fact that the

response is linear reinforces the confidence that

the response of a butterfly valve can be

extrapolated using the upstream pressure.

The validity of a diameter ratio cubed term, as

used in typical extrapolation relationships, was

also evaluated. The dynamic torque of the 8- and

24-in. valves from the same manufacturer, with

both valves in the CFF orientation, was compared

from the fully open position (90 degrees) to near

the fully closed position (20 degrees) at upstream

pressures of 15, 30, 45, and 60 psig. The results

of this evaluation are shown in Figure 21. The

extrapolation exponent was always below 3

except for one position at an inlet pressure of

60 psig. These results indicate that use of an

extrapolation exponent of 3 will result in the predicted torque of a larger valve being slightly

greater than the actual torque, provided the inlet

pressure does not exceed 60 psig. The trend of

the data indicates that an extrapolation exponent

of 3 could underpredict the actual torque of the

larger valve at higher inlet pressures.

This evaluation was then repeated for tests with

each valve in the FFF orientation. The results are

shown in Figure 22. This evaluation indicates that

the extrapolation exponent, when the flat face of

the disc is facing upstream, with the valve fully

closed, will frequently be above 3, ranging as high

as 3.47 in these tests. Thus, using an extrapolation

exponent of 3 could result in the predicted torque

of a larger valve being less than the actual torque.

To examine the effect the extrapolation exponent

has on the estimated torque, we calculated torques

for large valves, using an extrapolation of smallvalve data and extrapolation exponents ranging

from 2.8 to 3.5. The results are shown in Figure 23. This figure indicates that the torque

requirements of a larger valve could be seriously

underestimated if an extrapolation exponent of 3

is used when a larger exponent should be used. For

example, if the torque requirements of a 42- to

48-in. valve were being predicted from the results

of a 3- to 4-in. scale-model test valve, an extrapolation range on the order of 10 to 12, the torque of the

larger valve could be underestimated by 50% or

more, if an extrapolation exponent of 3 is used

instead of, for example, 3.2.

Consequently, a diameter-ratio-cubed formulation appears justified if (a) torques are obtained

with the scale-model test valve oriented with the

curved face of the disc facing upstream when the

butterfly valve is fully closed, and (b) the

upstream pressure does not exceed 60 psig. We

developed the following equations to more consistently envelop the response of a larger valve

based on the response of a smaller scale-model

test valve. Note that, in the CFF orientation, the

bearing torque (the torque that must be supplied

to move the valve disc against the resistance of

the bearings and seal only) and the dynamic

torque (the torque that must be supplied to move

the valve disc against the resistance of the flowing fluid only) are acting in opposite directions,

the dynamic torque assisting valve closure and

the bearing torque resisting it. Either equation can

be used, depending on the information available

(i.e., whether the total and bearing torques or just

the dynamic torque of the smaller scale-model

test valve is known). In either case, an estimate of

the bearing torque of the larger valve must be

known.

D3

D d S

- Tbh C (1)

Tb,) - Tb, (2)

where

- valve total torque (dynamic +

bearing torque)

e = a parameter of the large valve

D = valve diameter

s = a parameter of the small, scalemodel test valve

Td = valve dynamic torque only

Tb = valve bearing torque only

19 NUREG/CR-5720
Specific Observations

4 l

z

a,

.0

a,

0

OC 3 - x

a,

0

.-

a,

(D

E

0,,

0- I

Valve position (deg) 6 3200

Figure 21. Extrapolation exponent, butterfly valve oriented with the curved face of the disc facing

upstream. -

4

C

0

0.3

x

a)

C

0

00-

E 2

0

I2x

a)

a,

o0 1

E

c)

n _

90 80 70 60 50 40 30 20

Valve position (deg) 5 3211

Figure 22. Extrapolation exponent, butterfly valve oriented with the flat face of the disc facing upstream.

NUREG/CR-5720 20
Specific Observations

0

0~

0)

0

1 10 100

Valve diameter ratio (D%/ Ds)

6 3201

Figure 23. Actual to predicted butterfly valve torque (percent) as a function of extrap6lation exponent

versus valve diameter ratio.

21 NUREG/CR-5720
Specific Observations

A comparison between the results of the

method presented here and the results of a method

typically used by industry is presented in Table 1.

This table also provides a direct comparison

between the test results from this program and the

methods used by industry to predict the torque

requirements of a 24-in. butterfly valve from the

test results of an 8-in. butterfly valve. The results

indicate that, as expected, torque extrapolations

performed with test data obtained from a valve

oriented with the curved face of the disc facing

into the flow will bound the torque demands of

either orientation. However, extrapolations based

on the results from a test valve with the flat face

of the disc facing into the flow typically do not

bound the data.

We then used the proposed technique, as

reflected in Equation (1), to predict the response

of a 48-in. butterfly valve using both the 8- and

24-in. butterfly valves as the scale-model test

valves. The results, shown in Figure 24, indicate

some differences between the two 48-in. valve

predictions. However, the extrapolations for 8-in.

to 48-in. valves generally result in higher (more

conservative) values than the extrapolations from

24-in. to 48-in. valves. This is particularly true for

the peak torque.

3.2.6 Effect of an Upstream Elbow on the

Torque Requirements of a Butterfly

Valve. We also investigated the effects of system

upstream geometry on the closing torque requirements of a butterfly valve. We compared the test

results for valves located immediately downstream of an elbow to the results with uniform

inlet flow. (The valves were installed as close as

possible to the elbows in order to expose them to

the maximum nonuniform flow anticipated in an

actual installation.) The peak torque at 60 psig

was tabulated for all three valves in each of the six

orientations tested. These torques were then

normalized to the peak torque at 60 psig for each

valve in the uniform flow CFF orientation and

tabulated for easy comparison (Table 2).

Using this table, we can assess the effect nonuniform inlet flow relative to uniform inlet flow

has on valve torque. The worst-case elbow effect

was noted for one of the 24-in. valve orientations,

1.33 times the uniform inlet torque. This was

followed closely by one of the 8-in. valve

Table 1. Comparison of torque prediction methods butterfly versus valve orientation.

Torque

(ft-lbf)

Torque prediction method CFF orientation FFF orientation

Predicted 24-in. valve torque at 900 and 60 psig

Actual 8-in. valve torque 70.2 -52.1

Predicted 24-in. valve torque (INEL method) 1895 -1895

Predicted 24-in. valve torque (Industry method) 1895 -1407

Actual 24-in. valve torque 1754 -1828

Predicted 24-in. valve torque at 800 and 60 psig

Actual 8-in. valve torque 99.2 -58.0

Predicted 24-in. valve torque (INEL method) 2373 -2373

Predicted 24-in. valve torque (Industry method) 2984 -1875

Actual 24-in. valve torque 2257 -1986

NUREG/CR-5720 22
Specific Observations

0 10,000 _

0

-10,000 _

90 80 70 60 50 40 30 20 10 0

Valve position 5 2995

Figure 24. Predictions for a 48-in. butterfly valve based on extrapolating the torques of an 8-in. and a

24-in. butterfly valve at upstream pressures of 15 and 60 psig.

Table 2. Ratio of peak torque to uniform flow peak torque for a butterfly valve in the CFF orientation.a

Valve position Valve 1 Valve 2 Valve 3

FFF 1.06

CFF 1.00

FFF-CCW' 0.90

CFF-CCW 1.29

FFF-CW 1.02

CFF-CW 1.14

a. Valve position identification is given in Figures 7 and 9.

0.81

1.00

0.83

1.00: I

0.84

0.95

0.94

1.00

1.33

1.04

0.87

0.92

23 NUREG/CR-5720
Specific Observationsorientations, which was 1.29 times the uniform

inlet torque. Based on these results, the maximum

torque expected from a nonuniform inlet flow

configuration can be bounded by using 1.5 times

the torque from the uniform inlet flow configuration, if the curved face of the disc is facing

upstream when the valve is fully closed.

Next, a valve response plot was developed

(Figure 25) for the 24-in. butterfly valve in the

CFF orientation, with the shaft of the valve

perpendicular to the plane of an upstream elbow.

The similarity of the shape of this response plot to

the response plot of the same valve without an

upstream elbow (Figure 17) is clear. This

comparison suggests that, although the torques

resulting from a nonuniform inlet flow configuration may be higher, the response can still be

bounded with a factor of 1.5 times the torque

from the uniform inlet flow configuration, if the

curved face of the disc is facing upstream when

the valve is fully closed.

Finally, Figure 26 compares the linearity of

the peak torque versus inlet pressure for the 24-in.

butterfly valve with and without an upstream

elbow. Generally, the torque in the presence of an

upstream elbow is higher, but the response

remains linear. This comparison provides added

confidence that the results of the nonuniform inlet

flow configuration can be bounded.

This study has shown a weakness in the industry's extrapolation procedures for butterfly valves

closing against a compressive fluid. The industry

(a) has not identified a dominant orientation for

the small valve to be tested in, (b) has used

differential pressure (which is influenced nonconservatively by downstream pressure) instead

of upstream pressure, and (c) has not identified

the effect that the upstream piping geometry has

on the torque requirements of these valves, all of

which were shown to be important. Most butterfly valve vendors are aware of the NRC test program, but it is not known to what extent they have

incorporated the test results into their proprietary

calculations. It is known that most large purgeand-vent valves are either closed or blocked at

small angles of opening and that all purge-andvent valves required plant submittals and NRC

Safety Evaluation Reports (SERs) in response to

Three-Mile Island Action Plans NUREG-0660

and NUREG-0737 (NRC, 1980a; NRC, 1980b).

A large percentage of the purge-and-vent valves

were reviewed before these test results were

available, and the status of purge-and-vent valves

replaced in the last five years is not known.

Generic Letter 89-10 will not cause many

re-reviews, because many purge-and-vent valves

are not motor operated.

3.3 Assessment of Wedge-Gate

Valves Closing against

Medium to High Flow

Conditions

Flexwedge gate valves were tested in two NRC

test programs referenced earlier in this report;

those tests were reported in NUREG/CR-5406

(DeWall and Steele, 1989) and NUREG/CR-5558

(Steele et al., 1990). The latter was published in

support of Generic Issue 87. After that report

was published, we developed a technique to

(a) bound the stem force of a 5-degree flexwedge

gate valve closing against medium to high flow

conditions, and (b) validate a low differential

pressure closure test and then bound the stem

force of a flexwedge gate valve closing against

design-basis conditions.

In the two test programs mentioned above, the

authors tested six valves with a total of seven

different internal designs. One valve was tested

with two different discs, one with hardfaced disc

guides (Valve B Phase I), and one without

hardfaced guides (Valve 2 Phase II). Under the

high loadings encountered during the testing of

this particular valve, there was no difference in

the performance of the two discs.

The valves were tested under a broad range of

fluid conditions and flow rates, from normal

system flows to design-basis line break flows.

Initial conditions for each valve tested are shown

in Tables 3 and 4. These conditions were established before the normal system flow isolation

and the design-basis flow isolation portions of

each test. Two of the valves, including the valve

NUREG/CR-5720 24
Specific Observations

2000

1500

n

0-

a, 1 coo

G.

5001-

psIg 5 2996

Figure 25. Torque versus upstream pressure and angle for Valve 3, the 24-in. butterfly valve in the CFF

orientation.

3200

2400 -

.0~~~~~~~~~~~~~~

*1600 o~~~ ~ ~ 80

r--

. 1600 . ~~~CFF-CC /

-800 90

-800 /1III 0 10 *20 30 40 50 60 70 80 90

Static upstream pressure (psia) s 2997

Figure 26. Peak torque versus static upstream pressure for Valve 3, the 24-in. butterfly valve, comparing

the response of the peak torque with elbow and peak torque without elbow orientations.

25 NUREG/CR-5720
Specific Observations

Table 3. Phase I gate valve flow interruption test temperatures and pressures.

Pressure

(psia)

Temperature

Valve Test (OF) Test media

A

A

A

A

A

A

A

A

A

A

2

3

4

5

6

7

8

9

10

11

1000

1000

1000

1400

1400

1400

600

600

600

1000

530

480

400

580

530

450

480

430

350

530

Hot water

Hot water

Hot water

Hot water

Hot water

Hot water

Hot water

Hot water

I-lot water

Hot water

B

B

B

B

2

3

4

5

1000

1400

600

1000

530

580

480

530

Hot water

Hot water

Hot water

Hot water

Table 4. Phase II gate valve flow interruption test temperatures and pressures.

Pressure Temperature

(psig) (OF)

Valve Test Target Actual Target Actual Test media

6-in. Valve Tests

1 1 1000 900 530 520 Hot water

2 1 1000 950 530 520 Hot water

2 2 1000 1040 545 550 Steam

2 3 1000 750 <100 <100 Cold water

2 6a 600 600 300 450 Hot water

2 6b 1000 1000 430 470 Hot water

2 6c 1400 1300 480 520 Hot water

3 1 1000 920 530 520 Hot water

3 5 1200 1100 550 550 Hot water

3 7 1400 1300 580 570 Hot water

10-in. Valve Tests

4 1 1000 750 545 510 Steam

5 la 1000 800 545 520 Steam

5 lb 1400 1040 590 550 Steam

6 la 1000 990 545 580 Steam

6 lb 1400 1400 590 590 Steam

6 Ic 1200 1100 570 550 Steam

NUREG/CR-5720 26
Specific Observations

that was tested with two discs, performed in a

manner we have called predictable. A predictable

valve is one that does not exhibit evidence of

internal damage during testing. In such valves,

the highest stem forces occur when the disc is

riding on the valve body seats just before wedging. Conversely, an unpredictable valve exhibits

evidence of internal valve damage during testing,

characterized by an erratic, sawtooth-shaped stem

force response. In such valves, these high stemforce requirements typically occurwhile the disc

is riding on the guides rather than just before

wedging. Generally speaking, the results from

testing unpredictable valves are not useful for the

kind of evaluation described here. However,

through selective analyses, we were able to

include some of the results from two unpredictable valves, not while the disc was riding on the

guides, but after the disc had transitioned to riding

on the valve body seats.

We initially evaluated the test results with the

standard industry gate-valve stem-force equation.

Although some of the manufacturers modify the

variables in these equations slightly, the

application of the equations is basically the same.

AP = differential pressure across the

valve

= area of the stem

P = pressure upstream of the valve

= packing drag force

= disc and stem weight

= dynamic coefficient of friction

between disc and seat

Aorifice

0

= disc area on which pressure

acts

= seat angle (degrees from stem

axis)

F, = sealing force (wedging force

only)

Rt -= friction factor at torque reaction surface

FS = stem factor

F1,Industry = I'dAdAP ± AP + Fp (3) = distance to torque reaction

surface.

Later in the program after we developed the

INEL correlation, the Nuclear Maintenance

Assistance Center (NMAC) equation was

released. We also evaluated that equation, which

follows:

F., Fp-A, P

usA 0,i,,A P, ±~~~_±F

cosO ± u~sinO

F., NMIAC = (4)

For wedge-type gate valves, the industry has

historically used a disc factor (Wd) of 0.3; more

recently, they sometimes specified a more conservative disc factor of 0.5. The disc factor acts in

conjunction with the disc area and the differential

pressure; the three multiplied together represent

the largest component in the stem force equation.

However, the disc area used in the area term is not

uniformly applied throughout the industry. The

area is based on the orifice diameter by some

manufacturers, on the mean seat diameter by

others, or even on the orifice diameter times one

or more factors to artificially enlarge the area on

which the differential pressure acts. We used a

disc area based on the orifice diameter when we

evaluated the standard industry stem force equation, because it represents the least conservative

use of the term by the industry. This results in a

lower estimate of required stem force than the

other disc area terms.

where

= stem force

Ed

Ad

= disc factor

= orifice area

27 NUREG/CR-5720
Specific Observations

It is important to emphasize that the use of one

disc factor over another, or the comparison of one

disc factor to another, depends heavily on the disc

area assumed. In other words, a disc factor developed empirically with the mean seat area will be

lower and, thus, cannot be used to estimate the

stem force using the orifice area. Likewise, the

point in the closing cycle at which a disc factor is

determined is also of great importance. A disc

factor determined from the closing thrust history

prior to total flow isolation will underestimate the

isolation stem force if any extrapolation is necessary. The disc factor should be defined at flow

isolation, just before wedging.

For the standard industry equation, the estimated stem force is always a positive value, and it

is up to the analyst to differentiate between the

force to open a valve and the force to close a

valve. As a result, the stem rejection force must

be represented as a positive value if the valve is

closing and as a negative value if the valve is

opening. This is because the stem rejection force

always acts in an outward direction relative to the

valve body, resisting valve closure and assisting

valve opening. The packing force is always represented as a positive value because it always

opposes motion. It is typically assumed to be

constant for a given valve, but it varies from valve

to valve, depending on the packing design and the

packing gland nut torque.

For the NMAC equation, a dynamic coefficient

of friction of 0.35 to 0.50 is recommended. More

specific guidance is not given. At the time we

initiated our study, the disc area term in the

NMAC equation was labeled Aoi-,ce and defined

as the area on which pressure acts. The exact

meaning of this term was not clear, but we

initially used the orifice area in our assessment of

the NMAC equation. Unlike the industry

equation, however, the estimated stem force is

represented as a negative value for closing and as

a positive value for opening. The packing force,

the force due to the differential pressure across

the disc, and the sealing force can be either positive or negative, depending on whether the valve

is opening or closing. The weight of the disc and

stem acts in an inward direction relative to the

valve body, so it is always positive. The stem

rejection force acts in an outward direction, so it

is always negative.

Comparisons of the standard industry equation,

Equation (3), with selected test results are shown

in Figure 27. This figure presents the results of

the same valve isolating a break at a common

upstream pressure of approximately 1000 psig,

but with the fluid at various degrees of subcooling. The subcooling ranges from none

(steam) to approximately 400'F (cold water) with

intermediate values of LOTF and 1000F. The

recorded stem force is shown as a solid line; two

calculations of the stem force history, using the

industry equation and real time test data with

standard industry disc factors of 0.3 and 0.5, are

shown as dashed lines. This figure shows that, at

flow isolation, each test required more force to

close the valve than would be estimated using the

standard industry disc factor of 0.3. In fact, for the

tests shown on this figure, the more conservative

industry disc factor of 0.5 ranges from acceptable

(the steam test) to marginally acceptable (the

100F subcooled fluid test) to unacceptable (the

1000F and the 400'F subcooled fluid tests). Note

that, although the results of the industry equation

are presented over the entire closure cycle, the

equation represents a bounding estimate of the

maximum stem force. Therefore, only the

estimated stem force at the final horizontal line,

just before wedging, is applicable. The results are

presented for the entire closure to. aid in identifying trends in the recorded stemr force, not to assess

the equation throughout the closure cycle. Note

also that, although the same valve and operator

were used for each test, the closure durations are

different. Because of facility inventory limitations, some of the tests were initiated with the

valve partially closed.

It is also interesting to note that the shape of the

recorded stem force from flow initiation to flow

isolation varies, depending on the degree of

subcooling of the fluid. In tests with greater subcooling, the stem force during the initial portion

of the closure is lower. In fact, during the test with

cold water, the stem force trace was initially

positive (i.e., the valve was self closing during

this portion of the closure). However, just before

NUREG/CR-5720 28
20000

*__10*000

a)

IL

f--10000

U)

-20000

-30000

I I I I jII I I ,.-

- Valv'e 2, 1000 psi,430°F '(100°F ubcooled) -

-__ '___ Actual _------- Calculated (/L = 0.3)

*-. Calculated (/L- 0.5)

Flow initiation .........

~~ Flow isolation -

0 5 10 15 20 25 30 3

Time (s)

20000

10000

a)

c-, 0

tL

E 10000

-2000 0

-30000

I.

5 0 5 10 15 - 20

Time (s)

25 30 35

k)

20000

10000

_

c. 0

0

UIE.10 0 0 0 0)

U)

-20000

"Annnn II I 11111111111 I 1111 I ~~~~~-V- I -I1

Valve 2. 1000 psi, 530°f (10°F gubcooled)

Actual -------- ...Calculated (,u - 0.3) _ *-. Calculatod (tz = 0.51

_ I

10000

.0

a)

Oc 0

0

IL

Q)

U)

-20000

-30000

' 111|1 I.I ' ' I ' ' ' ' ' ' 1) I l '' 'l- ' ' ' -1-

Valve 2. 100 psi, 545°F (Steam) -~ Actual -------- Calculated (tz = 0.3) - -.---------- Calculated (/L - 0.51

.

-

I I I I I I - IIz

Cr

PO

0

-30000

0 5 10 15 20 25 30 35

Time (s)

0 *5 10 15 20 25 30 - 35

Time (s) MS291 RS-0491-04

Cn

D

0

0_

O

CD

PW Figure 27. Comparison of 0. the standard industry gate valve stem force equation with selected test results.
Specific Observations

wedging, the required stem force is generally

higher in tests with greater subcooling of the fluid

so that closure against cold water requires more

force than closure against steam.

A comparison using the NMAC equation,

Equation (4), is shown in Figure 28. In general,

the estimated stem force using this equation is

very similar to the estimated stem force using the

industry equation. The trends in the stem force

trace during a valve closure are also similar. As

with the standard industry equation, the NMAC

equation represents a bounding estimate of the

maximum stem force. The results of the NMAC

equation were presented for the entire closure to

aid in identifying trends in the recorded stem

force, not to assess the equation throughout the

closure cycle.

3.3.1 Assessment of the Disc Factor Term

in the Industry Equation. As we studied the

test results and analyzed the industry equation, it

became increasingly evident that the disc factor

or friction factor used in the equation was not well

0

CD

0

W

~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~.i ....... _ j..........

.~~ . ... .......... ................. ....................... * ..................... , .............. ....... i

Actual; H

Calc late "g- . 5 .......... al u at d .. 0 0

E

ci

co)

-20000

-30000

0 ....... ..... ... - - -5 10 15 20 25 30 35

Time (s)

Figure 28. Comparison of the NMAC gate valve stem force equation with selected test results.

NUREG/CR-5720 30
Specific Observations

understood. It appeared that the disc factor

depended on parameters not currently being

accounted for, such as the subcooling of the fluid.

Thus, we examined the equation in more detail,

specifically the disc factor term as currently

defined. To perform this evaluation, we used both

the design-basis and applicable parametric testing

to determine what influence pressure and fluid

properties, such as subcooling, had on the

required stem force of a valve.

If all of the parametric studies could have

resulted in just one parameter being varied, then

the tests could have been compared to each other

to determine the effect of that one parameter (e.g.,

fluid properties). That was not the case, however.

It was impossible to provide such precise temperature and pressure control at the valve. This,

along with other facility limitations, such as the

total system supply volume, resulted in tests that

cannot be compared to one another without some

type of normalization.

We normalized the test results using

Equation (3) by solving for the disc factor. This

was possible because we knew the stem force, the

system pressure, and the valve differential pressure throughout the closure cycle. The resulting

equation used in this evaluation was

-5% travel position. The slight slope on the

plateau between -5% and the -9% stem travel is

the result of valve inlet pressure increasing

slightly with flow isolation. The plateau region

represents the disc factor after flow isolation.

Two observations can be made from Figure 29:

(a) the absolute magnitude of the disc factor for

any fluid condition exceeds a 0.3, and (b) the

valve response is affected by fluid properties,

steam having the lowest disc factor and cold

water having the highest disc factor. The fluidproperties effect is evident throughout the closure

cycle but is most pronounced on the plateau

region, when the forces resisting closure have

essentially stabilized. This effect is contrary to

what was expected; one would expect water to be

a better lubricant than steam. The industry

equation and the NMAC equations do not contain

terms for fluid properties effects.

Our next effort was to determine if the disc factor was dependent on pressure. Figure 30 shows

a comparison for Valve 2 using six parametric

tests where the fluid properties remained constant

but the pressure was varied. Although the disc

factor did not exhibit a significant pressure

dependency at the zero stem position, it did from

the minus 5% to the minus 10% stem position,

when the disc was riding on the seats just before

wedging. The figure also indicates that the disc

factor was lowest during the 1400 psig test and

highest during the 600 psig test. This, too, is contrary to what was expected; one would expect a

lightly loaded disc to have a thicker lubrication

film and therefore a lower coefficient of friction

than a heavily loaded disc.

Figure 31 depicts the effect of pressure in the

opening direction, further highlighting inconsistencies with Equation (3). This plot is read from

left to right as the valve opens. Although the disc

factor (a positive value because the valve is opening) did not exhibit a pressure dependency at the

zero stem position, it did from the minus 5% to

the minus 10% stem position. This trend is

similar to what was observed in the closing

direction. The figure also reaffirms our previous

observations that the disc factor is lower during a

high-pressure test and higher during a lowpressure test. The opening disc factor is also

Ft - A, P - Fp

1-d Ad AP (5)

The results of a typical comparison are shown in

Figure 29, where there are two closures at each

fluid condition. The plot is read from right to left

as the valve closes. As time increases, the disc

factor increases in the negative convention (indicating valve closure) and is plotted against stem

position. The zero stem position represents that

point in the valve closure where the horizontal

visual area is blocked. Recognize, however, that,

although the visual area is blocked, fluid can still

flow under the disc and through the valve. At this

point, the disc area term in Equation (5) becomes

constant. From the zero stem position to the

minus 10% stem position, the stem travel

involves full seat contact, flow isolation, and

finally wedging. The forces maximize about the

31 NUREG/CR-5720
QI

n

T4

0

C-)

Il)

ii

0.0

-0.1

-0.2

-0.3

-0.4

-0.5

-0.6

-0.7

_.

CD

0

0

0

cr

_.

CD0

-t

-15 S -10 -5 0 5 10 15

Stem position (% of travel)

20

MS291 RS-0491-06

Figure 29. Disc factor for Gate Valve 2 closing on line break flow, effect of subcooling at 1000 psig.
0.0

-0.1

-0.2

0

0Co

4-)

CO

(

. _

-0.3

-0.4

-0.5

-0.6

-0.7

-15 -10 -5 0 5 10 15

cz

0

tjn

I

20

Stem position (% of travel) MS291 RS-0491-09

en lz2

0

0

0

5c_.

Figure 30. Disc factor for Gate Valve 2 closing on line break flow, effect of pressure at lOOT subcooling.
z

a

PC

;0

1.0

0.9 Effect of

pressure-I-1 1 |-1 as U I -10 Tb-cllng

fpressure at 100°F slbcooling

c:~

0

Cn

la

0

_.

0

0

Opening 0.8

0.7

0

+.

4-)

u)

.

,, .....

0.6G

0.5

0.4

0.3

0.2

0.1

0.0

600 psig

1000

1400

psig

psig

-

-15 -10 -5 0 5 10 15 20

Stem posiLion (% of travel) MS291 RS-0491-10

Figure 31. Disc factor for Gate Valve 2 opening on line break flow, effect of pressure at I 00F subcooling.
Specific Observations

observed to be higher than the closing disc factor

at its peak, non-wedging value.

The previously unaccounted influence of fluid

subcooling and pressure on the disc factor is very

evident. This influence is also contrary to what

one might expect in terms of the effectiveness of

a lubricant. However, what was expected is based

on a lubrication that separates the load-bearing

surfaces with a relatively thick film of lubricant to

minimize metal-to-metal contact. This type of

lubrication is known as thick-film lubrication.

The condition resulting from too little lubrication

is known as thin-film lubrication. The deficiencies of this thin-film lubrication can be

aggravated by valve sliding surface areas that are

too small to carry the maximum load.

When metal-to-metal contact exists, any condition that increases the ability of the lubricant to

penetrate the bearing region will decrease the

friction between the two surfaces. For instance,

the higher the differential pressure across a

bearing region, the more likely a given lubricant

will be forced into this region, thus lowering the

friction between the surfaces. Likewise, a

lubricant in a vapor state is more likely than the

same lubricant in a liquid state to penetrate the

bearing region, thus lowering the friction between

the surfaces. Other researchers have noticed these

same phenomena; however, they attribute this

sensitivity to changes in the temperature of the

fluid and metal. We are still trying to isolate the

reason, but the effect is real and must be

accounted for.

From the results discussed above and from

similar results for the other valves evaluated, it is

apparent that Equation (3) is incomplete and fails

to identify and predict the increasing stem force

due to fluid properties. Equation (4), evaluated

after development of the INEL correlation, is

more complete and can bound the results,

depending on the friction factor used; however,

there is little guidance to selecting a friction factor other than a range of 0.35 to 0.50. In addition-,

the fluid subcooling and pressure dependencies of

the disc factor or friction factor are inconsistent

with past assumptions inherent in the industry's

application of their equation. Thus, we concluded

that the industry equationsfailed to consider

parameters that have an important effect on the

observed responses of the valves.

In response to the above conclusion, we

directed our efforts toward investigating the flow

phenomena through a flexwedge gate valve and

the effect that pressures throughout the valve had

on the resultant stem force. That investigation

eventually yielded a correlation that bounds the

required stem force during closure with more

reliability than the standard industry equation with

either a Q.3 or a 0.5 disc factor, or than the NMAC

equation with a friction factor of 0.35 to 0.50.

Figure 32 shows a cross section of a typical

flexwedge gate valve and identifies those areas on

the disc and stem where the various pressure

forces can act. This figure also shows where we

drilled three pressure measurement ports into

each of the valve bodies before the Phase II testing, to assist in the internal pressure distribution

*study. Figure 33 shows a typical pressure

distribution observed during our testing. The

pressures in both the bonnet region of the valve

and under the disc are lower than the upstream

pressure during most of the valve closure cycle.

This reduction in pressure is due to the Bemoulli

effect, the result of fluid accelerating through a

valve in response to a reduction in the flow area.

This phenomenon depends on the pressure and

subcooling of the fluid and on the magnitude of

the reduction in flow area through the valve.

Thus, the Bernoulli effect will be system and fluid

dependent. The bonnet area also shows a lower

pressure because of the split in the disc and

because of the gap between the disc and the valve

body seats; these structural features provide a

path so that the pressure in the bonnet region can

more closely follow the pressure in the region

under the disc.

However, from the minus 3% to the minus 10%

stem position during this test, the pressures converge. During this portion of the valve stroke,

flow has been isolated and the disc is riding on the

valve body seats; however, wedging of the disc

has not yet begun. It is also during this portion of

the valve stroke that predictable valves exhibit the

35 NUREG/CR-5720
Specific Observations

S

Static

pressure

Flow -

Static

pressure

Static

pressure

M291 rs-0491-07

Figure 32. Gate valve disc cross section showing pressure forces and measurement locations.

1,000

900

I ' ' I 

1

I . . T

U)

a)

800

700

Upstream pressure ; \

Bonnet pressure ...----- 

Under disc pressure

600

500

400 I .. I . . I . . I . . .I

-10 0 10 20 30 40 50

Stem position (% of travel)

60 70 80

M291 rs-0491-11

Figure 33. Gate valve internal pressure distribution.

NUREG/CR-5720 36
Specific Observations

largest stem force. Thus, we concentrated our

efforts on this segment of the valve closure cycle.

Wedging forces were not considered because

these forces are not the result of fluid dynamic

and frictional effects, but instead depend on the

force capabilities of 'a given operator and on the

structural stiffness characteristics of a specific

disc and valve body.

3.3.2 Development of a Correlation to

Bound the Stem Force on a Gate Valve

during Closure. Our first effort was to develop

a relatively detailed free body diagram of the disc

while it is moving in the closing direction (after

the flow has been isolated but before wedging),

with the hope of better understanding the pressure

and area terms that affect the stem force.

Figure 34 presents the results of this effort and

identifies all the nonsymmetrical forces acting on.

the disc. Note that, according to the free body

diagram, the forces acting on the disc ultimatelyreact through the valve body seats, which are at a

slight angle (for a flexwedge gate valve) relative

to the horizontal and vertical valve coordinate

system. To account for this slight seat angle so the

forces are expressed in values consistent with the

definition of a traditional friction factor, we found

it necessary to transform the horizontal and vertical forces into a coordinate system that is normal

and tangent to the valve body seating surfaces.

Following this logic, we theorized the existence of two horizontal forces acting on the disc.

The first horizontal force (FUP) is a result of the

upstream pressure (Pup) acting on that area of the

disc defined by the mean diameter of the downstream seating surface. This surface presents a

Fstem

V

M291 rs-04 9 1 -12;

Figure 34. Gate Valve disc cross section showing unbalanced forces just before wedging.

37 NUREG/CR-5720
Specific Observations

circular profile in the horizontal plane. The mean

seat diameter was selected because it best approximates that area of the disc in contact with the

crown of the downstream seating surface over

which the various pressure- forces act. Fup is

defined as

I70D2 \

F stem rej PUP~\ 4 stem) (1 1)

_ p {:rD~mean) Fup. = Pp 4 (6)

The fourth force is a result of the pressure in the

bonnet region of the valve.(Pup once the flow has

been isolated) acting on the area of the disc

defined by the mean diameter of the seat cast in

the vertical direction. This area term is the result

of the slight angle of the seat in a flexwedge type

gate valve (nominally a 5-degree angle) and

results in an elliptical area over which the pressure acts. The major diameter is equal to the mean

diameter of the seat; the minor diameter is equal

to the mean diameter of the seat times the tangent

of the angle the seat makes with the vertical axis

of the valve.

Resisting this force (Fup) is a horizontal force

(Fdown) from the downstream pressure (Pdown),

also acting on that area of the disc defined by the

mean diameter of the downstream seating surface

F.o~~~r~ = -PdOH.flQ'nmean) down ~ dovvnkr4 | (7)

,7rD2\em, F~0 = P ~m In tan a (12) These two forces represent the only horizontal

forces acting on the disc and provide a far more

realistic estimate of the horizontal force

component than the orifice area term often used

in the industry equation. The net horizontal force

component (H) can be expressed as

Resisting this is a fifth force resulting from the

downstream pressure (Pdown) acting on the area

of the disc defined by the mean diameter of the

seat cast in the vertical direction. This force is

also the result of the slight angle of the seat.

H = F,,p - Fdown (8)

The free body diagram also indicates that there

are actually five vertical forces acting on the disc.

The first vertical force is due to the operator and,

represents the net stem force delivered to the

valve,

Fbotro Pdo. n(7D4 ) tan a Fbottomf = ~down 4 a (13)

The net vertical force component (V) during

valve closure can thus be expressed as

Fstem = stem load (9) V = Fstem - Fpacting

+ Ftop - Fbottom

- F stem rej

(14) The second force is a result of the resistance of the

packing while the valve is closing. The effect of

the disc and stem weights is also included in this

term,

Fpac,, = packing drag - disc and stem

weights (10)

The net horizontal and vertical forces can now

be recast into the plane defined by the valve body

seat and normalized to remove the effect of valve

size. Figure 35 shows these two forces and the

resolution of these into forces normal and tangent

to the seats. Note that the two normal forces

resulting from this transformation act in the same

direction, whereas the two tangent or sliding

forces oppose each other. These can be expressed

as follows for the normalized normal force (Fn)

The third force represents the stem rejection

force, the result of the pressure under the disc (Pup

once the flow has been isolated) trying to expel

the stem,

NUREG/CR-5720 38
Specific Observations

Surface sliding

Net valve vertical load

Normal component

iding component

ig component

lormal component

M291 rs-0491-08

Surface normal load

Net valve

horizontal load

Valve disc/seat,

interface surface

Si

din

Figure 35. Resolution of horizontal and vertical gate valve disc forces into surface normal and sliding

forces.

and the normalized sliding force

respectively:

H cos a + V sin a

F= (3rD2 d

4/

H sina - V cosa

Fs = {a~(~Dnnean)

4

(15)

(16)

The analysis described above allows us to

better characterize the normal and sliding forces

acting on a flexwedge gate valve disc just before

wedging. Our next effort was to determine if the

Phase I and Phase II flexwedge gate valve test

data supported a relationship between these

forces. From our data base, we extracted the test

results of all predictable valves during the closure

cycle when the disc was riding on the seats. We

did not include the results of any testing if severe

internal damage was evident, such as shown in

Figure 36. However, if the valve exhibited

evidence of internal damage while the disc was

riding on the guides but showed no evidence of

such behavior while the disc was riding on the

valve body seats, as shown in Figure 37, we

included the results.

The Phase II data extraction included testing

with three valves representing two valve sizes

(6-in. and 10-in.). The full-flow isolation tests

and normal operating flow isolation tests were

used; upstream pressures ranged from 600 psig

to 1400 psig, and fluid conditions ranged from

steam to 400'F subcooled water. However, the

results of tests when the differential pressure was

less than 20% of the upstream pressure, or when

both the differential pressure and the stem force

were increasing very rapidly while the disc was

riding on the valve seat, were not included. This is

because the magnitude and trends of the resulting

forces are obscured by relatively low loadings on

39 NUREG/CR-5720
c)

C

-0

L.-

L_-

E

01)

.

4 -j

Co

20000

10000

0

-10000

-20000

-30000

C0

D

0

0'

0

a,

CD

c.

M

vW

0

0 6 10 15 20. 25

Time (s)

30 35

S211 KGD-0490-04

Figure 36. Gate-valve stem-force history, indicating severe damage to the guides.
40000

30000

20000

o 10000

a) 0

0

LLL' -10000

E ' .D -20000

U)

-30000

-40000

-50000

-60000

0O 5 10 16 20 25

Time (s)

30 35

z

M

M/

0

S211 KGD-0490-10

Figure 37. Gate-valve stem-force history, indicating damage to the guides but not to the seats.

cn

0l

0

tj
Specific Observations

the valve disc (which affect the transition from

thick-film lubrication to thin-film lubrication),

and by rapid changes in both the stem force and

the differential pressure. This extraction yielded

data from 30 tests, which are listed in Table 5.

The Phase I data extraction included testing

with two valves representing a single valve

size (6-in.). By way of comparison, these valves

were nearly the same as Valves 1 and 2 in the

Phase II testing. Only design-basis flow isolation

tests were used from the Phase I testing.

Upstream pressures ranged from 400 psig to

1400 psig, and fluid conditions ranged from

100F subcooled to 140'F subcooled water. This

extraction yielded data from 12 tests, which are

listed in Table 6.

3.3.3 A Correlation to Bound the Stem

Force on a Gate Valve during Closure. We

can now rearrange the previously developed force

balance and solve for the stem force, based on a

linear relationship between the normalized normal and sliding forces. Using Figures 38 and 39,

the slope of the solid line (the friction factor

between the disc and the seat) can be used to

relate the normalized normal force (F,) to the

normalized sliding force (Fs) as

F, = -f F, ± C (17)

where

f = friction factor, the slope of line

that relates the normal force to

the sliding force and is equal to

The results of both data extractions were used

in the force balance developed as described above

and presented in Figures 38 and 39. The results

reveal two linear relationships between the

normal force on a seat (Fn) and the tangent or

sliding force (Fs) necessary to induce motion.

One is representative of a fluid subcooling of less

than 70'F, while the other is representative of a

fluid subcooling of 70'F or greater. The two

dashed lines on either side of the solid line

represent the limits of the observed data scatter.

The tight grouping of the data scatter lends confidence that not only can we bound the force

requirements of a flexwedge gate valve, but we

can also devise a method where the results of low

differential pressure flexwedge gate valve testing

can be verified and then the design-basis conditions used to bound the maximum stem force.

Note also that the dashed lines do not extend

below a normnalized normal force of 400 lbf/in2.

Because of the limited low pressure and low

differential pressure data available, and because

of the postulated friction mechanism, the applicability of the INEL correlation is currently limited

to normalized normal forces of 400 lbf/in2 and

above. We are working with selected utilities that

have proven diagnostic equipment to extend the

applicability of the INEL correlation.

0.400, if the fluid is less than

70'F subcooled

0.500, if the fluid is 70'F, or

greater, subcooled

C = offset bounding term reflecting

the scatter in the observed data

and is equal to

0 for no offset

50 lbf/in2 to bound the data provided a normalized normal force

of 400 lbf/in2, or greater, exists

according to Equation (15)

(shown as the dashed lines in

Figures 38 and 39).

Substituting the horizontal and vertical components of the normal force [Equation (15)] and the

horizontal and vertical components of the sliding

force [Equation (16)] into the above relationship

yields

V = H (f cos a + sin a) iC C ( )

(cos a - f sin a)

(18)

NUREG/CR-5720 42
Specific Observations I

Table 5. Phase II gate valve test data supporting extrapolation;

Force Pressure

(lbf) (psig) Subcooled

Test Stem - --- fluid

Valve number Step position Stem Normal Sliding Up Delta (OF)

1 25

1 26

6b 18

6b 25

6b 26

6b 1 26

6a 18

6a 25

6a 26

6al 25

6al 26

6c 13

6c 18

6c 25

6c 26

2 18

2 25

2 26

3 25

3 26

1 25

1 26

la l ,, ... 25

la 26

5 25

5 26

1 25

1 26

la 25

la 26

-7.34

-7.38

-7.10

-7.26

-7.08

-7.22

-7.54

-7.37

-7.14

i 0.

-7.07

-7.35

-7.13

-7.08

-7.29

-7.39

*-7.46

-7.06

-7.12

-7.10

-7.27

-7.13

-7.30'1:

-7.17

12929

13578

10477

12173

12453

13654

7880

7859

7703

8688

8866

15626

15046

15798

16452

16083

13096

13079

12751

14512

4355

8628

10804

987

1021

571

888

953

1031

413

550

548

609

633

952

970

1315

1359

1234

1091

1063

793 -

936

247

- 424, 959.6 948.5

-447 1002.4 980.3

-332 887.5 540.6

-404 856.4 851.3

-409 922.4 915.8

-451. 990.6 989.6

-254 601.7 389.7

-260 525.5 526.5

-254 524.7 524.5

-288 585.7 582.8

-294 603.1 606.7

-490 1414.6 907.9

-464 1431.2 927.8

-510 1276.3 1268.3

-533 1317.7 1310.8

-528 1215.7 1186.4

-422 1056.1 1052.5

-424 1027.6 1024.3

*-438 746.4 - 753.2

-496 886.7 890.8

-113 1012.0 237.0

11.5

16.3

95.9

36.2

36.0

48.4

91.9

18.8

10.9

42.5

39.0

124.4

115.4

47.0

39.7

9.5

9.0

8.4

371.7

388.6

17.2

5.9

7.5

7.1

7.3

11.0

9.6

6.9

5.6

7.7

-7.33 ; 11032

-7.18 12235 I .. .

-7.23 13281

-7.47 '24756

-7.28 22547,

-7.21 30474 - .3 29

-7.33 29045

887 -298 856.1 861.0

942 -386 915.1 909.1

928 -396 894.4 893.8

1144 -435 1117.4 1106.8

1177; -477 1140.4 1136.2

904 -336 879.9 876.8

835 -304 813.0 810.9

1055 -421 1026.3 1021.3

1013 -400 985.7 981.2

43 NUREG/CR-5720
Specific Observations

Table 6. Phase I gate valve test data supporting extrapolation.

Force Pressure

(lbf) (psig) Subcooled

Test fluid

Valve number Step Stem Normal Sliding Up Delta (OF)

A 2 5 11550 996 -413 959.3 _-a 10

A 3 5 5347 472 -170 456.3 _-a 60

A 4 5 12884 1040 -470 998.4 _-a 140

A 5 5 17460 1573 -637 1516.3 -a 10

A 6 5 6183 416 -213 397.3 _-a 60

A 7 5 15491 1378 -563 1328.0 -- a 140

A 8 5 7406 557 -258 534.6 _-a 10

A 9 5 7735 593 -270 569.5 -a 60

A 10 5 8183 616 -289 591.1 -_a 140

A 11 5 11743 1047 -417 1009.9 _-a 10

B 2 5 13833 978 -389 938.6 _-a 10

B 4 5 8500 602 -206 580.5 _-a 10

a. Full differential pressure tests were performed; a differential pressure equal to the upstream

pressure was used.

Now, substituting the horizontal and vertical

force components [Equations (8) and (14)] into

Equation (18), limiting the final result to bound

the maximum stem force, and rearranging yields

f = friction factor and is equal to

0.400 if the fluid is less than

70'F subcooled

OlF + 50 (..D2 a)

F -te = Fv + 62

stetn 02-

(19)

0.500 if the fluid is 70'F or

greater subcooled.

where

Fv = Fpacing + F stem tej - Flp + Fbottom

(20)

Equation (19) provides a method that can be

used to bound the maximum stem force requirements of a flexwedge gate valve closing against

medium to high flows and whose operational

characteristics have been demonstrated to be

predictable at design-basis pressures and temperatures. This method will also provide the basis by

which the results of in situ tests conducted on

predictable valves at less than design-basis

conditions can be verified and then the designbasis conditions used to bound the maximum

stem force. Note that the correlation applies only

01 = f cos a + sin a

Fh = Fup - Fdown

02 = cos a - f sin. a

(21)

(22)

(23)

NUREG/CR-5720 44
0

c'J.I-- C\

4-

_0

-J

0)

c: -C

I 's

N

E

o

Z

-100

-200

-300

-400

-500

-600

-700

tAl

-800

0 200 400 600 800 1,000 1,200

Normalized Normal Load (Ibf/in2)

1,400 1,600

cz

_,

Co PO~

0

M455 kgd-1 091-22

Figure 38. Normalized surface sliding force versus normalized surface normal force for gate valves closing against less than 70TF subcooled fluid.

0

CD

0

0

0

Cr

V)

CD

:.

Can
z

n

M

U'

O

0

CD

-o

0

0

0

Cr

v]

CD

cj

0

=3

an

.4-1

-Q

0

C:

N

E

0

z

-100

-200

-300

-400

-500

-600

-700

-800

0 200 400 600

Normalized

800 1,000 1,200

Normal Load (Ibf/in2)

1,400 1,600

M455 kgw-1091-23b

Figure 39. Nornalized surface sliding force versus normalized surface normal force for gate valves closing against 70TF, or greater, subcooled

fluid.
Specific Observations

to flexwedge gate valve closure against medium

to high flows, flows which will result in a normalized normal force of 400 lbf/in2 or greater,

according to Equation (15). This correlation has

not yet been validated for predicting the closing

force requirements when the normalized normal

loading is less, but we are working with selected

utilities to extend the applicability of the INEL

correlation.

The INEL correlation cannot be used in the

opening direction. The following comparison

best explains why. Figure 40 represents the stem

force of a design-basis opening cycle of a smaller

valve. The results are representative of a valve

that experiences the highest stem force demands

immediately after unseating and while the disc is

still riding on the seats. A correlation similar to

the one presented here may be applicable for such

valves, subject to an assessment of the required

friction factor. However, Figure 41 represents

the stem force of a design-basis opening cycle of

a larger valve and clearly shows a very different

stem force response. Here, the highest stem force

occurs when''the valve is approximately 25 to

30% open. This is well outside the assumptions

implicit in the INEL correlation, the NMAC

equation, or the standard industry equation and

dramatizes the influence fluid dynamic effects

can have on valve response. The issue of bounding the maximum stem force to open a flexwedge

gate valve will be investigated in the future.

3.3.4 Use of the Correlation to Bound the

Stem Force on a Gate Valve during Closure. To use the INEL correlation to bound the

closing stem force requirements of a flexwedge

gate valve, the analyst must first determine if the

valve is of the type whose operational characteristics are considered to be predictable. If the valve

is predictable, the following information can be

used to bound the stem force:

* The mean diameter of the valve body seat

(the average of the inside diameter and the

outside diameter of the valve body seats

measured in the plane perpendicular to'the

stem)

* The angle the valve body seat makes with

the vertical or stem axis

* The outside diameter of the stem

- An estimate of the-maximum packing drag

expected, less the effects of the weight of

the disc and stem

* The maximum upstream pressure that

would exist after flow isolation but before

wedging, typically the design-basis pressure

* The maximum differential pressure that

would exist after flow isolation but before

wedging, typically the design-basis

differential pressure

* The subcooling of the fluid at design-basis

conditions, either less than 70'F subcooled

or 70'F or greater subcooled.

With this information, the nef horizontal force

(Fh) can be estimated with Equation (22) and

Equations (6) and (7). The net vertical force (F,)

can then be estimated with Equation (20) and

Equations (10) through (13). The angle between

the seat and the vertical or stemn'axis can then be

used with Equations (21) and (23) to estimate

terms associated with transforming the horizontal

forces into forces on the disc and seat and finally

into ai vertical force. Thus, all the terms in Equation (19) can be determined and the maximum

stem force necessary to isolate flow at the

specified pressure and differential pressure can be

calculated.

By way of example, we will use the upstream

pressure and differential pressure recorded after

flowlisolati6n, but before wedging during one of

the Valve 2 tests, and will bound the stem force

requirements of the valve using the INEL correlation. Similar estimates will also be made using the

industry equation and either an orifice area or a

mean seat area in conjunction with a standard

industry disc factor of 0.3, as well as a more conservative industry disc factor of 0.5. The results

of each estimate will then be compared with the

actual stem force recorded during the test.

47 NUREG/CR-5720
z

C)

MJ

la

0

0

0

Cr

en

20000

10000

-0~

.a)

0

UE

.2a -10000

U)

-20000

-30000

0 5 6 10 -15 20 25

Time (s)

30 35

S211 KGD-0490-31

Figure 40. Stem-force history of a small gate valve during an opening test.
40000 A VPI V `, 1fid 1,L- IjJ £-:iJ lV LV- H 4 tLJ I alll

.asoooo

20000

n 10000

0~~ oI

LL -10000 .

E

2-20000

Cl)

-30000

-40000 Actual

-------------- Calculated (,u = 0.3) 50000 *-- Calculated (mu = 0.5)

-60000

0 5 10 15 20 25 30 35

Time (s) S211 KGD-0490-94

Figure 41. Stem-force history of a large gate valve during an opening test. 0

0D

k)~_

0 >
Specific Observations

The results from Valve 2, Test 3, Step 25 will be used for this comparison. Since this was a cold water

test (400'F subcooled fluid), we will use the 70'F or greater subcooled friction factor in the INEL correlation. Pertinent valve information and the pressure and differential pressure recorded just before wedging are

Stem diameter 1.750 in.

Orifice diameter 5.187 in.

Seat inside diameter 5.192 in.

Seat outside diameter 5.945 in.

Packing drag 200 lbf

Upstream pressure just before wedging 746 psig

Differential pressure just before wedging 753 psid

We can use this information in the INEL correlation as follows:

Dmen (seat ID + seat OD) = (5.192 + 5.945) -5569

Dmean 526 2 2

Dmean =7 (5.569) 2 24.358

Amean 4 4

Astern = *7e - (1-750) - 2.405 4 4

FIp Pu~pAmean = (746)(24.358) = 18, 171

Fdown = PjdAnAmean 0 = (746 - 753)(24.358) = - 171

FIop = PupAmeantanla = (746)(24.358) tan5 = 1590

Fbottom = PdowvrAmeantana = (746 753)(24.358) tan5 = -15

F stem rej = PtpAvstem = (746)(2.405) = 1794

Fh = FUP - Fdow:n = 18,171 - (-171) = 18,342

F, = Ppacklng + FstemrejFtop + Fbottom = 200 + 1794 - 1590 + (-15) = 389

01 = f cos a + sin a = 0.500 cos5 + sinS = 0.585

0- = cos a - f sin a = cosS- 0.500 sinS = 0.953

F =F 01 F + 5A0 mean] = 389 + [(0.585)(18,342) + (50)(24.358)] 12,926.

02

NUREG/CR-5720 50
Specific Observations 1

We can also use this information with the industry equation as follows:

Industry equation with [Ld 0.3

A orifice _ .r(5.187) 2 21.131

orifice - 4 4

_ s____ L. r(I.750) 2 _

Astem = __m = ( *4 ) = 2.405

F, IAorifice Y= AP + Astem Pup + Fpacking = 0.3(21.131)(753) + 2.405(746) + 200

= 4773 + 1794 + 200= 6767

Industry equation with Ed = 0.5

Ft = idAoriice: AP + Ase mPup + Fpacking = 0.5(21.131)(753)

+ 2.405(746) + 200 = 7956 + 1794 + 200 = 9950

During this test, the peak stem force recorded

just before wedging was 12,787 lbf. Thus, the

INEL correlation bounds the actual recorded stem

force. The industry equation using either disc area

term, with either the standard industry disc factor

of 0.3 or the more conservative industry disc

factor of 0.5, underpredicts the actual recorded

stem force just before wedging. Additional comparisons are presented in Table '7 for each

Phase II test evaluated. Examining these results

indicates thatfthe INEL correlation consistently

bounds the maximum recorded stem force before

wedging, whereas the industry equation, using

either disc area term with either disc factor,

typically underestimates the stem force, except

with a disc factor of 0.5 and a mean seat area.

3.3.5 Identifying the Pressure Dependency Contributing to the Stem Force on

a Gate Valve during Closure. Based on the

results of our testing and data evaluation, we

identified both a subcooled and differential pressure dependency on the disc or friction factor. The

subcooled dependency has been previously

shown; not so evident, however, is the differential

pressure dependency of the correlation. Referring

to either Figure 38 or 39, the INEL correlation:

follows the lower bounding curve, a higher resistance for a given normal force, or a nominal friction factor plus an offset. This nominal friction

factor plus an offset can also be reduced to a

single load dependent friction factor. Figure 42

displays the friction factor and offset as shown in

Figure 38. Also shown on this figure is the loaddependent friction factor at two normalized normal forces. The normalized normal force of

400 lbf/in2 results in a 'load-dependent friction

factor of 0.525. The normalized normal force of

1400 lbf/in2 results in a load-dependent friction

factor of 0.436. This effect is the result of the data

offset or bounding term and its relative magnitude

compared to the normal force component. Thus,

embedded in the INEL correlation is the differential pressure dependency observed in our test

data.

3.3.6 Low Differential Pressure Test

Verification and Bounding of the Stem

Force on a Gate Valve Closing against

Design-Basis Flows. Utilities have numerous

flexwedge gate valves in systems throughout a

nuclear power plant; many of these valves must

function in various design-basis events. The capability of these valves to operate at design-basis

conditions usually cannot be verified with in situ

-testing, especially for valves where design-basis

conditions include high pressures and medium to

51 NUREG/CR-5720
Specific Observations

Table 7. Comparison of gate valve stem forces, estimates versus actual.

Stem force

(lbf)

Industry equation

Pressure

(psig) Orifice area Mean seat area

Test

number Step Up Delta Actual INELa [1=0.3 1[=0.5 [t=0.3 rl=0.5

Valve 2

1

6b

6b

6b

6bl

6a

6a

6a

6al

6al

6c

6c

6c

6c

2

2

2

3 .

3

Valve 3

1

la

la

5

Valve 5

1

I

la

la

25

26

18

25

26

26

18

25

26

25

26

13

18

25

26

18

25

26

25

26

959.6 948.5 12929

1002.4 980.3 13578

887.5 540.6 10477

856.4 851.3 12173

922.4 915.8 12453

990.6 989.6 13654

601.7 389.7 7880

525.5 526.5 7859

524.7 524.5 7703

585.7 582.8 8688

603.1 606.7 8866

1414.6 907.9 15626

1431.2 927.8 15046

1276.3 1268.3 15798

1317.7 1310.8 16452

1215.7 1186.4 16083

1056.1 1052.5 13096

1027.6 1024.3 13079

746.4 753.2 12751

886.7 890.8 14512

1012.0 237.0 4355

856.1 861.0 8628

915.1 909.1 10804

894.4 893.8 11032

1117.4 1106.8 12235

1140.4 1136.2 13281

879.9 876.8 24756

813.0 810.9 22547

1026.3 1021.3 30474

985.7 981.2 29045

13429

13845

10556b

12182

13002

13920

7930b

8092

8070

8806

9091

16558b

1 6827b

17434

17975

16455

14718

141355

12934b

1 5 0 4 4 b

5249

11283

11835

11649

141058

14376

29638

27712

33839

32674

8525 12535

8822 12964

5765 8052

7653 11250

8224 12095

8859 13043

4120 5768

4806 7033

4791 7010

5305 7769

5498 8063

9372 13218

9524 13446

11307 16666

11681 17221

10643 15655

9415 13865

9164 13494

6768 9950

7982 11747

3180 4205

7038 10762

7421 11354

7292 11155

8954 13743

9171 14085

18260 27753

16951 25729

21116 32167

20325 30943

9444

9771

6289

8477

9111

9818

4498

5316

5299

5870

6086

10238

10423

12535

12950

11791

10434

10155

7497

8844

3340

7621

8037

7904

9704

9940

20133

18683

23296

22419

14067

14545

8924

12623

13573

14641

6398

7883

7857

8710

9043

14662

14944

18712

19336

17569

15564

15144

11165

13185

4472

11734

12380

12175

14993

15367

30874

28615

35801

34434

25

26

25

26

25

26

25

26 -

25

26

a. A friction factor (f) of 0.400 was used in the INEL correlation, except as noted.

b. 70TF or greater subcooled fluid test, a friction factor (f) of 0.500 was used.

NUREG/CR-5720 52
Specific Observations

0

-100

ro

E

cD

'CDJ

rTD

- -

-300

-400

-500

-700

-800

0 200 400 600 800 1000 1200 1400 1600

Normalized Normal Load

Figure 42. Gate valve load-dependent friction factor using the INEL correlation.

high flows. Usually, only low flow and low differential pressure conditions can be developed near

valve closure under in situ conditions. When such

valves are determined to be predictable by other

means, the INEL correlation offers a method that

can be used to bound the stem force requirements

of the candidate flexwedge gate valve at designbasis conditions. This can be accomplished as

follows:

Step 1: Perform a differential pressure test. Use

the results of the testing to estimate the

normalized normal and sliding loads for

the valve, according to Equations (15)

and (16). The valve is considered to be

representative of the valves tested by

the INEL if (a) the upstream pressure

and differential pressure while the disc

is riding on the seats result in a normalized normal force of not less than

400 lbf/in2, and (b) the resulting forces

fall within the upper and lower bounds

expected for valves of this design

(see Figures 38 and 39). If the resulting loads do not fall within the expected

band, the results of our testing are not

representative of the valve being tested

and the INEL correlation may not be

applicable.

Step 2: If the results of Step 1 fall within the

expected band, use the actual designbasis conditions in the INEL correlation

to bound the stem force requirement of

53 NUREG/CR-5720
Specific Observations

the valve being tested. Then use this

maximum stem force estimate, along

with the other necessary motor operator

sizing calculations, to verify. the size of

the operator and the setting of the torque

switch, and thus ensure that sufficient

stem force is available to operate the

valve at design-basis conditions.

3.3.7 Nuclear Maintenance Application

Center (NMAC) Gate Valve Stem Force

Equation. The current NMAC gate-valve stemforce equation is mathematically equivalent to the

best-estimate portion of the INEL math. Like the

INEL correlation, it is based on a first-principles

analysis, but it does not provide the additional

empirical derived guidance that is contained in

the INEL correlation. The NMAC guide provides

a range of dynamic coefficients of friction that

might be encountered in gate valves. Suggested

values range from 0.35 to 0.50, but more specific

guidance is left to others. Our immediate concern

is with the following footnote, presented in the

NMAC guide, addressing the applicability and

use of a coefficient of friction:

"Extensive testing . . . using Stellite 21

found no statistically significant variation

between static and dynamic friction. This

testing also found that temperature, fluid

type, and contact pressure had no statistically significant effect on friction."

Our data assessment to date has not addressed the

issue of static versus dynamic friction; however,

we have clearly shown that fluid type (subcooling) and differential pressure have a significant

effect on the friction factor. We suggest that this

statement be reviewed in light of the INEL test

results.

The NMAC guide also presents their equation

as being applicable for estimating valve opening

stem forces. One of the basic assumptions stated

in the guide regarding the use of their equations is

that

"The limiting thrust occurs at or near the

seat, during seating or unseating, and no

additional dynamic flow effects contribute

to the required stem thrust."

As mentioned earlier in this report, INEL test

results indicate that some valves experience their

highest stem forces when the valve is 25 to

30% open. Thus, this statement, too, should be

reviewed in light of INEL test results.

NUREG/CR-5720 54
4. UNDERSTANDING DIAGNOSTICS AND DIAGNOSTIC

TESTING OF MOTOR-OPERATED VALVES

4.1 Overview

Recent recommendations made by the NRC in

Generic Letter 89-10 suggest that licensees

develop programs for MOV inspection, maintenance, and testing under pressure and flow

conditions to ensure that the valves will function

when subjected to design-basis conditions.

Development of the licensees' MOV programs is

underway, but in our opinion, it is obvious that

MOV diagnostic testing will play a large part in

these programs. Diagnostic test equipment and

methodologies are continuing to be developed by

both the diagnostic industry and by selected

utilities. Over the years, we have developed a

very sophisticated diagnostic capability to aid in

our MOV research. Admittedly, some of this

capability is needed from a researcher's standpoint and is not necessary for diagnostic in situ

testing. However, we have identified from this'

larger capability what we believe to be the

minimum diagnostic capability necessary for

thorough in situ testing of rising-stem MOVs.

This section of the report presents our thoughts on

measuring and assessing the performance of

MOVs during diagnostic testing and then relates

that performance to 'design requirements.

Because MOV diagnostic test equipment is still

being developed, some of the judgments made in

this section are based on the capability of equipment that was commercially available when this

report was written.

MOV diagnostic systems should serve the following purposes: (a) to identify deterioration. or

faulty adjustment in the valve and operator, (b) to

maintain assurance of design-basis capability that

has been established previously for a particular

MOV, and. (c) to provide information to establish

design-basis capability based on low-pressure

andlow-flow tests and/or from prototype tests.

MOV diagnostic equipment is used in conjunction with one of two types of tests: unloaded static

testing or testing under flow and pressure loads

(dynamic testing). Both position-controlled and

torque-controlled valves will be discussed.

Unloaded static testing of position-controlled

and torque-controlled valves provides measurements of the packing drag and stroke time..

For torque-controlled valves, the static test also

provides measurements of torque and stem force

at torque switch trip. However, the direct

determination of the capability of an operator at

design-basis conditions from an unloaded test of

either a position-controlled or a torque-controlled

valve is highly suspect. These data can be used for

trending the MOV.

Flow and pressure testing provide much more

information on both torque-controlled and position-controlled valves than does no-load static

testing. It is well known that some valves cannot

be design-basis tested in the plant, but the higher

the in situ test loading, '-the more in-depth

information the analyst will have to determine the

performance of the valve and the margins available in each of the MOV components. Again,

more information will come from the torquecontrolled valves than from the positioncontrolled valves. For critical position-controlled

valves that cannot be significantly loaded in situ,

it may be necessary to consider dynamometer

testing of the motor operator.

The thoughts presented here are a product of

our experience in the four full-scale test programs,

discussed in Section 1. The recommendations are

also based on the results of special effects testing

being conducted at the INEL on the MotorOperated Valve Load Simulator (MOVLS),

described in Section 4.3. Our research to date

indicates that, for rising-stem MOVs, ideally a

diagnostic system should be able to accurately

measure the following parameters as a minimum:

position of the motor operator switch (for limit

switches and torque switches), motor current and

voltage, motor operator torque, and valve stem

force. The following discussion explores each of

the recommended measurements and explains its

value to the analyst.

55 NUREG/CR-5720
Understanding Diagnostics

4.2 Understanding MOV

Diagnostic Testing

4.2.1 Motor Operator Switch Position.

4.2.1.1 Limit Switches. For ac- and

dc-powered position-controlled valves (including

those valves that close on torque but open on

limit), the analyst can determine valve stroke time

and stem position in either a static test or a

dynamic test. The point in time when the switch is

activated provides the basis for analyzing many

of the other diagnostic measurements. Figure 43

shows typical dc limit switch histories.

4.2.1.2 Torque Switches. Torque switch trip

(for torque-controlled valves) establishes a timing

point to assess the output of the operator. The stem

force at torque switch trip will typically be lower

than the peak force produced after torque switch

trip, when the motor controller dropout time and

the motor and operator momentum result in further

compression of the stem. The motor controller

dropout time for both ac- and dc-powered MOVs

is small, but its effects are real; the length of the

dropout time is further influenced by the state of

the ac cycle when the torque switch trips. Momentum effects are conspicuous in static tests, but they

can be absorbed by the higher loadings in dynamic

tests. In addition, if the torque switch trips before

the disc is fully seated in a dynamic test, then the

momentum does not produce the additional stem

force that would have been produced in a static

test. Figure 44 is a torque switch history showing

a typical ac-powered motor controller torque

switch trip.

For a torque-controlled valve operating in the

closing direction, the exact trip time (for

ac-controlled circuits this must be interpreted

from the motor controller holding coil ac sine

wave) is important for analysis of all the other

quantitative measurements. It is important that the

data acquisition system have a sample rate that is

fast enough to allow the analyst to correctly determine the time of torque switch trip and all other

data with respect to this time. The faster the sample

rate, the less conservatism the analyst will have to

add to the measurements to account for time

errors. We have found that a sample rate of

1000 samples/second is usually sufficient; it

provides a resolution that is faster than most of the

sensor responses.

4.2.2 Motor Current.

4.2.2.1 Torque-Controlled Alternating

Current Motor Operators. The margin of an

electric motor can be estimated from a motor current history. The concern is that the output from an

ac motor can change quickly when the motor is

operating too close to stall current, where a large

increase in required current and a large decrease in

motor speed correspond with only a small increase

in torque (in some cases, no increase at all). To

determine if the electric motor may have problems, compare the peak motor current just before

torque switch trip with the motor torque/speed

curve for the specific electric motor to help estimate electric motor margins. By way of example,

Figure 45 is a current history for a valve closing

against a design-basis load; Figure 46 is the ac

motor torque speed curve for the operator motor.

Comparing the current drawn in Figure 45 with

the motor torque speed curve, even though the

current transducer is nearly saturated on this

stroke, we see that the motor is operating well out

on the knee of the motor torque speed curve. Figure 47 shows the motor current during the next

valve closure against a design-basis flow load.

During this second test, which was performed

within five minutes of the test shown in Figure 45,

the motor went into a stall. [Careful examination

of the motor current trace from the previous cycle

(Figure 45) would have alerted the analyst to the

impending problem.] Motor heat and voltage

drop, in conjunction with a marginally sized electric motor, were the primary causes of the stall. If

the analyst cannot obtain a motor torque speed

curve for a specific motor, it may be necessary to

conduct dynamometer testing to produce such a

curve to determine motor margins.

4.2.2.2 Position-Controlled Alternating

Current Motor Operators. The analyst must

subject the position-controlled valve to its designbasis loading, by in situ or prototype testing, to

make an analysis that is equivalent to that

previously described for torque-controlled

valves. Again, the analyst may want to consider

dynamometer testing of the motor operator.

NUREG/CR-5720 56
Understanding Diagnostics

2.5

2

. . . . . . . . . . . . . . . . . . . . . . . . . .

U)

t,,

1.5

1

Open limil - --- Close limi

Total stroke time i -- for a position-n' controlled valve

i

.I

0.5

0

-0.5 I I , I i' *. I . . I . I l I I . . . I l l

0 2 4 6 8 10 12 14

Time (s) S314 KGD-0291-09

Figure 43. Limit switch history showing stroke time for a position-controlled valve. Limit switch

actuation points are important when analyzing other data.

0.5

0.4

=: 0.3

.= 02

C,,

CD 0.1

0-

a) 4 - -0.1

0

C3 -0.2

-0.3

-0.4

-0.5

11.70 11.72 11.74 11.76 11.78 11.80 11.82 11.84 11.86 11.88 11.90

Time (s) S314 KGD-0291-10

Figure 44. Torque switch history showing torque switch trip and the termination of current flow to the

motor controller holding coil.

57 NUREG/CR-5720
Understanding Diagnostics

Valve 5, Test 1A, Step 25, 1200 psi, 5700F (Steam) 25

D_

E

-C

C

-1

Co

0

0

20

15

10

- (note: current transducer maximum range - 20 amps)

5

0

0 5 10 15 20 25 30

Time (s)

35

S314 KGD-0291-01

Figure 45. Current history from a design-basis test showing current increasing as the valve closes; the

rapid increase in current indicates torque switch trip. The test began with the valve 75% open.

2000

1600

a1)

m) 1000

CL)

co

L_.

o

0

500

0

60

40

E

Co

30

Co

C.)

20

0

+1

10

10

0 10 20 30 40 50 60 70

Motor torque (ft lb) S314 KGD-0291-11

Figure 46. Motor torque-speed curve showing the speed-torque-current relationship for the ac motor test

results shown in Figure 45. Beyond the knee of the curve, the torque increase is small in proportion to the

speed loss and the current increase. For some motor configurations there is no increase;, for some there is

actually a decrease in torque beyond the knee of the speed curve. (Arrows indicate the applicable axis.)

NUREG/CR-5720 58
Understanding Diagnostics

Valve 5, Test 1A, Step 26,1200 psi, 570'F (Steam) 25

E

-w

~: 0) 4-

C

CD

:t

0

0

20

1

15

10

(note: current transducer maximum range - 20 amps)

5

0 ,,,, . . . .n . . I . . . . . . . . . . . . . . . . . . . . . . . .

0 5 10 15 20 25 30 35

Time (s) S314 KGO-0291-02

Figure 47. Current history from a design-basis test showing the motor going into a stall, saturating the

current transducer. The test began with the valve 30% open.

NOTE: All other readings of motor current for

both torque and position controlled valves are

made for trending. Additionally complicating the

subject, Limitorque has publicly informed the

industry that ac motors of a given model operating under the same conditions can vary up to 20%

in their output.

4.2.2.3 Direct Current Motor Operators. As

in the previous analysis, the dc MOV must be

loaded before torque or limit switch trip. The analyst can then compare the motor current to the

motor torque speed curve to determine how far

out on the curve the motor is operating. Figure 48

shows that, unlike ac motors, dc motors continue

to produce torque linearly even as they approach

stall conditions; however, at the lower motor

speeds that accompany the higher operating currents, the effects of motor momentum are less. As

a result, the lower motor speeds produced under

high loads can affect the stroke time, as shown in

Figure 49. In addition, dc motors operating at

high loads can heat up very quickly, which in turn

reduces their output torque. Figure 50 shows the

drop in output torque over time for a 40 ft-lb dc

motor as it heated. Degraded voltage (because of

line losses) can also reduce the output of the

motor. As the motor turns at lower speeds, the

overall efficiency of the operator is reduced. Figure 51 shows the results of these effects. The test

at the lowest loadings closed the valve, and the

motor tripped out on the torque switches. The

tests at the next two higher loadings did not close

the valve, but the motor was able to generate

enough torque to trip the torque switch. However,

the test at the highest load not only did not close

the valve, but because of the efficiency losses in

the motor circuit, it was also unable to generate

enough torque to trip the torque switch and

therefore stalled the unit. These factors, working

separately or together, can thus produce motor

stall and motor burnout.

4.2.3 Motor Voltage. Voltage drop under load

can alert the analyst to undersized cables in the

power circuit. It has been found on a number of

occasions that some plant circuit designers did

not realize that both ac and dc motors can operate

at four to five times the running current stamped

59 NUREG/CR-5720
Understanding Diagnostics

3,500

3,000

0. E

_0

a)

a)

CL

V)

2,500

2,000

1,500

1,000

150

12 ~ 0

125 a)

a)

E

100 .-

4-.'

C: .. a)

75 '

0

50 *2

2

25

60

MllI kod-0592-01

500

0

0 10 20 30 40 50

Load (lb - ft)

Motor torque

Figure 48. Motor torque-speed curve showing the speed-torque-current relationship for a dc motor. The

torque (load) continues to increase as the speed drops and the current increases. (Arrows indicate the

applicable axis.)

8

6

C:

C:

0

CL)

0

0.

6-

4

2

0

0 5 10 15 20 25

Time (s)

M248 kd-O291-02

Figure 49. Disc position histories from three dc-powered closing tests showing that the stroke time is

longer with higher loads.

NUREG/CR-5720 60
Understanding Diagnostics'

150

125

-6

E

Cu

ci,

0;

100

75

50

2 5 F 1.l

0, 10 20 30 40 50 60

Load (ft - lb)

248 kygd-3O1-05

Figure 50. Motor heating under load caused the current demand of this dc motor to decrease at a rate of

1 amp per second during the 20-second test, resulting in a decrease in the output torque of approximately

I ft-lb every 2 seconds.

8

6

0

0.

C.,

03

Cn

.4

2

0

0 5 10 15 i 20 -25

Time (s)

..248 kgd-02S 1I-0

Figure 51. Disc position histories from four dc-powered MOV closing tests, showing that the stroke time

is longer with higher loads. Three of the tests did not fully seat the valve, and the test at the highest load

stalled the motor.

61 NUREG/CR-5720
Understanding Diagnostics

on the motor nameplate, and therefore specified

less than adequately sized power cables. Not long

ago, a utility traced several burned out dc motors

to this undersized-cable problem. Limitorque has

publicly stated on a number of occasions that, for

the small number of dc motor-operated valves in

the industry (compared to the large number of ac

motor-operated valves), the percentage of

replaced dc motors far exceeds that of ac motors.

This increased incidence is probably a result of

the lower dc voltages and the associated higher

current to perform the same work as the higher

voltage, lower current ac motor.

It should also be noted that in the typical

Limitorque dc motor wiring circuit, it is impossible to accurately determine the voltage drop in the

circuit from one measurement. There are typically four wires that conduct the motor current

and interconnect the motor controller and the

motor. Two of the wires interconnect the armature

and the series field; this interconnection provides

the ability to reverse the direction of the motor.

Since the resistance of each wire contributes to

the voltage drop, any single measurement of

motor voltage will be contaminated by a voltage.

drop caused by a minimum of two wires.

Previously unexplained motor burnouts may be

the result of excessive voltage drop in the power

circuit; a voltage check at the highest valve

loading possible may provide the explanation.

The voltage check will also aid the analyst who

needs to know the actual line losses at normal

voltage conditions so that the additional losses at

design-basis degraded voltage conditions can be

calculated.

4.2.4 Motor Operator Torque. Load-sensitive

motor-operated valve behavior (rate of loading

explained in Section 4.3) and the need to establish

individual valve performance margins provide

the incentive to measure motor operator output

torque. Some of the new-generation transducers

measure motor operator output/stem torque

directly. Where these new transducers are not

available, special-effects testing at the INEL is

providing an increasing data base showing that

the relationship between spring pack force or

deflection and output torque is very constant and

not affected by the loading on the operator.

Figures 52 through 54 show that the ratio of

spring force divided by stem torque (the torque

applied to the stem by the stem nut) does not vary

over a wide range of loadings. Figure 52 is from

a lightly loaded test with a low torque switch setting. Figures 53 and 54 are from tests with

higher loads and with a high torque switch setting. (The stem force histories from those two

tests are shown in Figures 55 and 56.)

Motor operator torque measurements are used

in determining the stem factor (operator torque

divided by stem force) part of the margins assessment for a specific valve and provides a second

reference for the indirect- or direct-force

measurements made with some of today's stem

force measurement transducers. To reduce the

potential errors in indirect torque measurements,

the analyst should obtain detailed calculations for

spring pack force or deflection versus operator

torque. The spring pack should be calibrated to

determine any offset from the published

Limitorque spring constant for each spring pack

assembly. Figure 57 shows a spring pack

calibration. Using this kind of information and

the specific operator moment arm, the analyst can

obtain reliable values for the operator output

torque from many of the commercially available

spring force or deflection transducers. Torque

determined from static testing will not provide

useful information for stem factor determination.

Both torque- and position-controlled valves need

to be tested with a flow and pressure load to

properly load the stem-stem nut interface.

4.2.5 Valve Stem Force. Stem force in both

position- and torque-controlled valves can best be

determined from measurements. While it is

possible to estimate the stem force if the output

torque of the operator is known, our experience

shows that the stem factor varies significantly

with load. However, if the analyst knows the operator torque and the stem force, the stem factor can

be determined by the operator torque divided by

the stem force. Note that the operator torque and

stem force measurements must be obtained just

before the valve hard seats; after the valve starts

wedging, there is almost no motion between the

stem nut and the stem, so no useful value can be

NUREG/CR-5720 62
Understanding Diagnostics

10 f -I l I I I I i I I Il i I I jI I I l l l I .

MOVLS Test 1

.. , .... I., ...... I I

a)

:3

0-

0

E

a)

a)

0

0

0)

Ia

Co

a,

:3

0-

0

FM

9

8

7

6

5

4

' 3

2

1

,Torque swit(

trip

,h

. I

I . . . , I . .*'. . I . .. I Q L

C .................... 1 2 3 4 .5

Time (s)

I . . . . I . . . ..I ... I I , I 1 -i

6 7 8 9 10

S314 KGD-0291-03

Figure 52. Ratio of torque spring force to stem torque from MOVLS Test 1, with a small stem force.

10 .11 ........ I-- ........................ I......... . . I . . . . I

a,

:3

07

4-

E

a)

Co

0

4-

0)

C',

0

or

MOVLe Tst4 4 9

8

7

6

6

4

3

2

1

I.... I .... I....

Torque -

switch -

t ,

I-

,

I 0 1.1-1.1-1 ....

0 1 2 3 4 5 6 7 8 9 10

Time (s) S314 KGD-0291-04

Figure 53. Ratio 'of torque spring force to stem torque from MOVLS Test 4 at a higher stem force.

63 63 ~~~~NUREG/CR-5720
Understanding Diagnostics

10

0)

cr

0

a)

4-

a)

E20

.-_

0) 0)

0-

4-

C',

0) FL

9

8

7

6

5

4

3

I I I . I I I I I I r I I I . I I ' I I T Ir I I I I I . . . I I I . . I I I I I I . I . -. MOVLS Test 5

Torque

switch / trip

2

I I I . 1 , . I I i I i i r I ,~~~~~~~~~~~~~~~~~~~

0

4 6 6 7 8 9

Time (s)

10 11 12 13 14

S314 KGO-0291-05

Figure 54. Ratio of torque spring force to stem torque from MOVLS Test 5, at a high stem force, in

which final seating was not achieved, showing that the average torque spring force to stem torque ratio was

not influenced by the load on the operator.

0

-5000

_ -10000

0)

0

*° -15000

E

0)

u) -20000

-25000

-30000

I . I I I . r I I I . I I I I I I I I I I I I I I I I I I I I r I I I I I I I I I . I I I I I

: ~~~~MOVLS Test 4

> Seating

Margini

0 1 2 3 4 5 6

Time (s)

7 8 9 10

S314 KGO-0291-07

Figure 55. Stem force history from MOVLS Test 4, showing a 3000 lbf margin between the peak stem

force and the stem force at final seating.

NUREG/CR-5720 64
Understanding Diagnostics

0

-5000

- 10000

0)

0

*- -16000

U) -20000

-25000

-30000

4 56 6 7 8 9 10 11 12 13 14

: Time (s) S314 KGD-0291-08

Figure 56. Stem force history from MOVLS Test 5, showing that a 1000 lbf increase in the initial stem

force eliminated the 3000 lbf final seating stem force margin observed in MOVLS Test 4.

0

U-

-i--, . Deflection (in.)

I .1I I

S314 KGD-02911-08.

Figure 57. Typical torque spring calibration plot shows the initial force offset at zero deflection.,This

figure also confirms that forthis spring, the spring force to deflection relationship is linear. The actual values

are omitted from this figure because the data are proprietary until the MOV diagnostic equipment validation

is complete.

65; NUREG/CR-5720
Understanding Diagnostics

obtained for the stem factor. After the analyst has

determined the stem factor and how well the

conversion of operator torque to stem force is taking place, extremely good or bad stem factors can

alert the analyst to either faulty torque or force

measurements or to problems at the stem-to-stemnut interface. Limitorque selection guides (SEL)

provide stem factors at several coefficients of

friction for most popular stem diameter, thread

pitch, and lead combinations. Accurate stem force

measurements can also alert the analyst to excessively high packing loads. Packing loads that are

higher than allowed by the design calculations can

adversely affect the valve's ability to function

when pressure and flow loads are added. Current

and torque measurements can also alert the

analyst to high packing loads, but cannot quantify

the actual packing drag.

For torque- and position-controlled valves, the

stem force measurement also can be used to determine the available stem force margin if the valve

is tested at design-basis flow and pressure loads.

Stem force values obtained from less than fully

loaded tests cannot be used to establish fully

loaded stem force margins. Figures 55 and 56

show why this is so. Figure 55 is a stem force history from an MOVLS test, showing a stem force

margin of approximately 3000 lbf (see Section 4.3

for a discussion of the MOVLS). Figure 56 shows

the stem force history from an MOVLS test with

an additional 1000 lbf initial loading that resulted

in no seating margin. A stem force margin of

3000 lbf was wiped out by a 1000 lbf increase in

the initial stem force loading because the stem factor increased with the increase in load. The stem

factor increased 6% during the second test. This

phenomenon is known in industry as the rate-ofloading effect, and is discussed in the next section

as a load-sensitive behavior.

Although not useful to establish stem force margins, stem force measurement data from unloaded

tests on torque-controlled valves can be used for

trending. Changes in the measured stem force can

alert the analyst to problems with the valve or with

the measurement system. Stem force measurements from unloaded position-controlled valves

will provide the analyst with little useful informnation, as the only'load in the stem is the packing

load.

Inaccuracies in the stem force measurements

do occur; this affects the reliability of data from

testing either torque- or position-controlled

valves. Valve stem force transducers come in

many types and applications. The accuracy of

strain gages depends on accurate characterization

of the materials involved, the expertise of the

installer, the quality of the installation technique,

and in some cases the accuracy of the in-place

calibration. Removing the stem from the valve

and performing a calibration of the strain gages in

a tensile machine can determine their accuracy.

The accuracy of stem collar transducers also

depends on extensive characterization of the

materials involved and the quality of the

installation. All of these instruments measure

very small reaction loads in materials that are

typically specified only by type. Without some

type of calibration, the analyst must choose either

the middle of the material's properties range or a

conservative high or low bound, depending on the

calculation. All of these problems are magnified

when the stem force measurement must be made

in the threaded portion of the stem. Using strain

gages installed on the yoke can result in a

combination of these problems.

Placing load cells between the valve yoke and

the motor operator requires that the attaching bolt

metallurgy and the structural stiffness of the yoke

flange be known. The structural stiffness of the

yoke flange also affects the calibration, if strain

bolts are used to attach the yoke to the motor

operator. Strain bolts without other intrusive

devices will measure the stem force only in the

closing direction.

All of the above technologies measure stem

force with varying degrees of accuracy; the accuracy depends on the degree of characterization of

the geometrical and metallurgical considerations

and the quality of the installation. The sensitivities and overall accuracies of all the current

commercial diagnostic torque and thrust

measuring systems will be better known after the

NRC/Motor-Operated Valve Users Group

completes its validation testing and evaluation.

NUREG/CR-5720 66
Understanding Diagnostics

4.3 Load-Sensitive

Motor-Operated Valve

Behavior

The MOVLS in its simplest configuration was

built to help develop our field test instrumentation

and data acquisition equipment. Following our

Phase II gate valve testing, the MOVLS was

improved to perform special effects testing. It

was again improved for the diagnostic equipment

validation program. Its current configuration,

shown in Figure 58, consists of a Limitorque

operator bolted to a valve yoke, which in turn is

bolted to a hydraulic cylinder. The cylinder

discharges to an accumulator, and the operator

resistive loads are varied by controlling the initial

fluid level and initial gas charge in the accumulator. The output thrust from the operator is

transferred to the hydraulic cylinder through a

valve stem, a thrust bearing, and a load cell. A

torque arm is bolted to the valve stem and serves

two purposes: it acts as an anti-rotation device

and is instrumented to measure the torque in the

stem.

Instrumentation used on the MOVLS includes

3-phase current (both rms and peak to peak),

rms voltage, 3-phase power and power factor,

motor speed, open and close limit switch positions, torque switch position and trip signals,

spring pack position and force, and the stem

force, torque, and position. The data acquisition

system typically runs at 1000 Hz, resulting in a

1-ms time resolution. Loadings in the stem can be

varied from a no-load condition to design-basistype loads. Stem thrust histories obtained from

our MOVLS compare very favorably with stem

thrust histories obtained from full-scale valve

tests we obtained in the field. The MOVLS provides an economical means of producing realistic

valve loadings in the laboratory and allows us to

study operator-related phenomena that we have

observed in our full-scale valve tests.

Load-sensitive MOV behavior describes the

phenomenon in which the maximum output thrust

of a rising-stem MOV at torque switch trip

decreases as the load on the stem increases. This

phenomenon has been observed throughout the

industry and occurred on two occasions during

our full-scale valve test programs. Although we

lacked sufficient instrumentation and analyses to

pinpoint all of the possible causes of this phenomenon at the time it occurred, we were able to

determine that it occurred when the torque switch

was not set high enough to fully seat the valve.

Since then, we have been able to use the INEL

MOVLS to perform the separate effects testing

necessary to isolate the first-order cause of the

phenomenon.

Using the MOVLS, we performed a series of

17 simulated valve closings. The first seven were

setup tests and will not be part of this discussion.

The 10 tests that will be discussed are shown in

Figures 59 through 61, which show the thrust

histories for each of these tests. Each of these tests

had the same torque switch setting. The only variable from test to test was the stem load applied by

the simulator. The first three tests shown in

Figure 59 are low load closures, typical of what

could be obtained during inplant testing while the

valve is under pressure with no flow through it.

These low load tests were followed by the four

tests shown in Figure 60. During these tests, the

closing load was increased before Test 11 and

again before Test 12. Thereafter, the closing load

was not increased. Test 11 is representative of a

valve closing against a pressure and flow load and

shows that increasing the stem load results in a

small decrease in the thrust margin. The thrust

margin is the difference between the thrust at the

time the valve seats and the thrust at torque switch

trip. During Test 12, the stem load is again

increased and results in a larger decrease in the

thrust margin. For comparison purposes, Test 12

is equivalent to slightly less than the design-basis

loading of a typical predictable 6-in. BWR

RWCU flexwedge gate valve.

Tests 13 and 14 represent two additional tests

at the same stem loading .as Test 12 and show

that, with a constant load, the thrust margin continues to decrease. Test 13 has effectively no

thrust margin and the simulator was just able to

seat. Test 14 shows a negative thrust margin and

the simulator does not seat. At the full designbasis loading, we have been able to create the

67 NUREG/CR-5720
Understanding Diagnostics

/z64 Ac 'OO -/O 4L4

i -. - '-.': CtOSc6D PoS* r/27a

,~~~~~~~~~~~~lj .K~ 5( I6A/rOe~5

Ot leQ6e. 4011161

/l O, ,;z ' 4r 

.a~~~~~A e _, S°Att f

ZD4XZ / 4'sUT 16d1/6 r4eAYdt ,,, _ .U *. . 1o.~ CdD f// A

L44-

Figure 58. MOVLS layout drawing.

NUREG/CR-5720 68
Understanding Diagnostics

-o

IE

0)

Cl)

0

- -2,000

-4,000

-6,000

-8,000

-1 0,000

-12,000

-14,000

-1 6,000

-18,000

-20,000

-22,000

-24,000

-26,000

M450 rs.1 091-09

Figure 59. Traces of thrust versus time for Tests 8 through 10 of the MOVLS test sequence. These are

identical low-thrust tests.

Z-o

(n

E

a)

Cl)

0

-2,000

-4,000

-6,000

-8,000

,-1 0,000

-12,000

-14,000

-16,000

-18,000

-20,000

-22,000

-24,000

-26,000

Test 10

r~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~

Test 11 Test 12 Test 13 Test 14

lV

argin . . Margin okar

r : Ad ' ~~~~Margin

... I I I ...........

.9Margin not - | T : seat

A Torque switch trip

I

M450 rs-1091-10

Figure 60. Traces of thrust versus time for Tests 10 through 14 of the MOVLS test sequence.

69 NUREG/CR-5720
Understanding Diagnostics

-o

Cl)

P

E)

CO

0

-2,000

-4,000

-6,000

-8,000

-1 0,000

-12,000

-14,000

-16,000

-18,000

-20,000

-22,000

-24,000

-26,000

M450 rs- 1091-11

Figure 61. Traces of thrust versus time for the final three tests of the MOVLS test sequence. Tests 12-14,

which exhibit load-sensitive MOV behavior, were followed by a series of low-thrust tests (Tests 15-17),

with resistance to closing much like the first three of the series.

Test 14 response in a single step following

Test 11. However, we went through the transition

from normal stem behavior to abnormal stem

behavior so quickly that we were not positive of

the root cause. Performing the test at a lower

loading, and repeating it, provided further insight

about the cause of load-sensitive MOV behavior,

as we will discuss later in this section. We are

aware that three near-design-basis closures back

to back are not a normal design requirement.

We expected this load-sensitive MOV behavior, based on the setup tests. Following the above

tests, we reduced the stem load to the level of

Tests 8 through 10 and repeated the closures

three times. Tests 15, 16, and 17 are shown in

Figure 61. We observed that the thrust margin

during Test 15 was less than the thrust margin

during Tests 8, 9, 10, 16, and 17. This was unexpected, but it did indicate that the MOV loadsensitive behavior was the result of a change in

some parameter that we should be able to isolate.

To better understand this phenomenon, and to

be sure it was not a combination of causes,

kinematic and kinetic evaluations of the motoroperated valve were performed. Based on these

evaluations, we identified potential first-order

causes of this phenomenon. Using measured data

taken from the MOVLS testing discussed above,

we then evaluated the likelihood of each

contributing to the load sensitive MOV behavior.

The kinematic evaluation of the motor operator

revealed that it functions like a planetary gear.

There is one input path for motion, but there are

two possible paths for the output motion, as

shown in Figure 62. The motion input to the unit

is the result of the electric motor; the two possible

output motion paths are at the stem and the spring

pack. Like any planetary gear, the geometry of the

operator (the diameter of the gears, the number of

gear teeth, etc.) does not completely determine

the output motion of each path. Therefore, the

kinetics of the operator were considered for more

insight into the behavior of the unit. We learned

NUREG/CR-5720 70
Understanding Diagnostics

Spring pack

output m11otiOnl

Motor input

motion

1I

Stem Output

motion

Figure 62. The design of the operator allows two potential output motion paths.

that the relative resistance to motion of each path

at any given instant determines how the input

motion is distributed between each of the possible

output motion paths. 'At very low stem loads, the

entire input motion is transferred to the stem.

Motion of the spring pack'does not occur until

there is enough resistance to stem motion

(observed as an increase in the stem thrust) to

overcome the preload of the spring pack. Once

the preload of the spring pack has been overcome,

increasing the resistance to motion at the stem

causes motion in both the spring pack and the

stem. If the resistance to motion in the stem

becomes too high (high valve closure loads or

valve seating), then all the input motion will be

transferred to the spring pack.

Perhaps this conversion of input motion to output motion can best be visualized with examples

of a lightly loaded MOV (Test 8), a moderately

loaded MOV (Test 11), and a heavily loaded

MOV (Tests 12-14). In a lightly loaded MOV,

the two output motions occur, but they occur in

series, one after the other. Prior to seating, the

input motion to the worm drives the worm gear,

resulting in motion of the stem but 'no'motion in

the spring pack. After the valve seats, motion can

no longer occur in the stem, so all the input

motion is transferred to the spring pack' until the

torque switch trips. In this situation, all the kinetic

energy of the operator is transferred to the stem as

the valve seats and results in operator torques and

stem forces that exceed the values expected for

the setting of the torque switch. Increases in stem

thrusts after torque switch trip are the result of

motor controller, dropout time and post-torque

switch trip kinetic behavior. The torque switch

does not interrupt the electrical energy to the

motor, it interrupts the power to the motor controller contact holding coil. Motor controller

dropout time includes the collapse time of the coil

71 NUREG/CR-5720
Understanding Diagnostics

field and the mechanical time for the springs to

open the contacts.

The response of a moderately loaded MOV can

differ significantly. If the resistance to stem

motion is large enough, the increasing torque

levels in the operator can overcome the spring

pack preload before the valve seats. Output motion

will then occur simultaneously in the stem and in

the spring pack. Because there are now two simultaneous output motion paths, the increased frictional losses cause a slight reduction in the overall

efficiency of the unit. As the moderately loaded

MOV seats, the motor controller dropout time and

the kinetic energy of the operator are transferred to

the stem, much like the lightly loaded MOV case.

This results in stem forces that exceed those

expected for the torque switch setting.

The behavior of an MOV that is so heavily

loaded that the torque switch trips before the

valve seats is similar to that of a moderately

loaded operator, except that the stem is not rigidly

restrained because of valve seating. Consequently, the motor controller dropout time and the

kinetic energy of the operator are dissipated

because motion of the stem following torque

switch trip and the high stem thrusts of a rigid

body are not developed.

The above discussion of MOV behavior and

careful analyses of the motor operator internals

generated a number of potential candidates for the

first-order cause of load-sensitive MOV behavior.

These candidates represent one of two types of

effects: frictional losses of the operator mechanism or inertial effects of the operator mechanism

resulting from the stoppage of motion. We used the

MOVLS-generated data to investigate each candidate and determine the single most likely cause of

the observed load-sensitive behavior. To do so, we

grouped each of the inertia and friction effects into

one of four categories, based in part on their possible influence on motor operator behavior and in

part on the type of data available from existing

MOVLS instrumentation. The four categories are

(a) effects that increase the load on the motor but

do not influence the interaction between the torque

switch and the stem, (b) effects that influence the

spring pack force at torque switch trip, (c) effects

that influence the spring pack force needed to

achieve the same level of stem nut torque, and

(d) effects that influence the amount of stem nut

torque necessary to achieve the same level of stem

thrust.

Based on MOVLS testing and our full-scale

valve test programs, load-sensitive MOV behavior was not observed to be associated with motor

stall. Therefore, effects associated with the first

category are not related to the load-sensitive

behavior. Effects from the electric motor slowing

down will be presented in the dc-powered MOV

section.

Small inertial and large frictional possibilities

are associated with the second category that could

cause a change in the spring pack force at torque

switch trip. These include inertia of the spring

pack components and friction among them. Variations of spring pack force at torque switch trip

would cascade through the mechanism and vary

the operator output torque and stem thrust

achieved at torque switch trip. We investigated

this possibility.

Inertial and frictional effects associated with

the third category could cause a change in the

ratio of spring pack force to output torque. Input.

torque to the worm gear/drive sleeve/stem nut

assembly is generated by application of the spring

pack force at a moment arm distance from the

stem centerline (see Figure 63). Elastic effects

(worm shaft bending, for example) could change

this distance a bit under varying spring pack

force, resulting in a change in the ratio of spring,

pack force to output torque. Other effects in this

category include friction between the worm shaft

and worm, a component of the friction force

between the worm and worm gear, and friction in

the drive sleeve bearings. Inertial effects that fall,

into this category include those associated with

acceleration of the worm or deceleration of the.

worm gear/drive sleeve/stem nut assembly. These

additional effects are depicted in Figure 64. The

summed result of all the third category effects

would be an instantaneous difference between the

torque input to the mechanism by the spring pack

NUREG/CR-5720 72
Understanding Diagnostics

Mean radius,

Spring pack force

Stem torque

Figure 63. The application of spring pack force to the operator mechanism at an offset distance from the

stem centerline generates the torque input to the mechanism.

Drive sleeve

bearing friction, Deceleration of ilie

stem nut, sleeve,

and worm gear

Acceleration of

Ithe worn

Figure 64. Potential causes of torque loss in the operator are identified.

73 NUREG/CR-5720
Understanding Diagnostics

force and that output by the mechanism to the

stem nut. Such a difference could cause loadsensitive MOV behavior. We investigated this

possibility.

There is only one inertial or frictional effect in

the fourth category, an effect that increases the

output torque required to achieve the same stem

thrust. This effect would be seen in the stem

factor, defined as the ratio of output stem torque

to stem thrust. The industry uses the traditional

power thread equation to estimate the stem factor:

an equation based on the stem diameter, the

thread pitch and lead, and the coefficient of

friction between the stem and stem nut threads.

This power thread equation does not include any

inertial terms, and none are required. All of the

parameters in the expression, except for the

coefficient of friction, are obviously invariant

between a lightly loaded and a heavily loaded

MOV. This leaves only the friction between the

stem and the stem nut threads as this category's

possible cause of the load-sensitive behavior. We

also investigated this possibility.

Up to this point, we have identified and categorized potential candidates for the first order cause

of load-sensitive MOV behavior. We then used

data from each of the MOVLS tests described

earlier to evaluate the likelihood of each

candidate causing the observed load-sensitive

behavior. In addition to acquiring stem thrust

data, spring pack force data, and stem torque data,

we were also able to calculate an instantaneous

ratio of spring pack force to stem torque, the ratio

of stem torque to stem thrust (stem factor), and

the coefficient of friction between the stem and

the stem nut. All calculations were made on a

real-time basis, using measured data and valve

and operator design information.

The possibility of inertial or frictional effects in

the spring pack causing load-sensitive MOV

behavior is not likely, as shown in Figures 65

through 67. These figures show the same

margins in the spring pack force at torque switch

trip throughout the series of 10 tests as the stem

force plots (Figures 58-60). Torque spring force

margin is defined as the difference in the torque

spring force history between the point where the

operator goes into single-path response and

where the torque switch trips. There are observed

differences in the spring pack force at torque

switch trip, but these are within the 5% advertised

by Limitorque. The most interesting torque

spring history is from Test 15, which shows that a

higher spring pack force was required during the

running portion of the stroke than was required in

Tests 16 and 17, even though the stem load was

the same in all three tests. The stem nut coefficient of friction discussed later in this section

caused the additional spring pack force during the

running portion of Test 15.

The possibility of changes in the ratio of spring

pack force to stem torque causing load-sensitive

MOV behavior is not likely, as shown in

Figures 68 through 70. The variation of the ratio

between individual tests was less than that

associated with the worm interacting with different parts of the worm gear during the same test, as

shown by the oscillations in the figures. Changes

in the ratio indicate that the torque transfer

efficiency of the motor operator mechanism

actually improves somewhat with increasing

load. Such an improvement could result from a

small reduction in the coefficient of friction in the

operator bearing gear surfaces associated with an

increasing normal surface force, a known friction

phenomenon. Regardless, an increase in efficiency is counter to the losses associated with

load-sensitive MOV behavior.

In contrast to the other data, the stem-to-stemnut coefficient of friction changed significantly

and in a fashion that supports such frictional

losses as the most likely cause of load-sensitive

MOV behavior. Figure 71 shows that the stemto-stem-nut coefficient of friction during Tests 8,

9, and 10 remains relatively constant during the

low-load tests. Test 11, as shown in Figure 72,

provides the first truly loaded test and the coefficient of friction is observed to decrease. This is

consistent with an increase in the normal surface

force. However, Tests 12, 13, and 14 are not the

result of normal friction behavior; they are most

likely the result of the lubricant being squeezed

out of the stem-to-stem-nut interface. In other

NUREG/CR-5720 74
Understanding Diagnostics

200

0

Test 8 T Test 10

on

-o

a)

0

LLU-1

C-)

CZ

.

a.cm_

QL

Un

-200

-400

-600

-800

-1,000

t I'I

Margin

* Torque switch

trip

argin MaMrgin

k

-1,200

-1,400

M450 rs-1 091-06

Figure 65. Traces of spring pack force versus time forTests 8 through 10 of the MOVLS test sequence.

These traces correspond to the thrusts of Figure 59.

200

0

.0

I-,

0

IL

C-,

a _

CD

U)

-200

-400

-600

-800

-1,000

-1,200

-1,400

M450 rs-1091-07

Figure 66. Traces of spring pack force versus time for Tests 10 through 14 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 60. Spring pack force does not change

significantly from test to test.

75 NUREG/CR-5720
Understanding Diagnostics

200

0

3-

a)0)

0

0)

C:

0L

C/)

-200

-400

-600

-800

-1,000

-1,200 seat I

-1 ,400

M450 rs-1091-08

Figure 67. Traces of spring pack force versus time for Tests 14 through 17 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 61. Spring pack force does not change

significantly from test to test.

81 '--....... . ,

a)

tioE

a)

U)

a)

a

0

I-.

C-,

EL

cm

. _

LUn

7

6

5

4

3

2

1

YVrWW

Test 8 Test 9 Test 10

- A Torque switch

trip

0 . . . . I . . . . .. I . .

M359 kgd-O7O1-19

Figure 68. The ratio of spring pack force to stem torque for Tests 8 through 10 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 59.

NUREG/CR-5720 76
Understanding Diagnostics

8 .~~~~- - .- . .- . .-. .-. - - ..-... llI

ai)

0

E

a)

n

ci) 2

Co

IL0

CL

tM

CL

C')

7

6

5

4

3

2

1

1 AH

, . . . . . . . . . . . . . . .

 vlll- i

Test 10 s Test 11 Test 12 Test 13 Test 1 4

A Torque switch trip

0

b1359 kod-O7St*20l

Figure 69. The ratio of spring pack force to stem torque for Tests 10 through 14 of the MOVLS test

sequence. These traces correspond to the thrusts of Figure 60.

8

cir 0~

F.

E

a)

(1)

C)

0

o

LL

.w

rL

cm

C:

,()

6

5

4

3

2

Trest 15

Test 15 Test 16 Test17i

Test 14 Test 16 Test 17

4

0

T . I

.A_ Torquo switch rip

b359 0d-0701-21

Figure 70. The ratio of spring pack force to stem torque for Tests 14 through 17 of the MOVLS test

sequence. These traces'correspond to the thrusts of Figure 61. Spring pack force does not change

significantly from test to test. - I

77 NUREG/CR-5720
Understanding Diagnostics

C

0

LL

4-0

C

(1)

0

0

0.20

0.18

0.16

0.14

0.12

0.10

0.08

0.06

0.04

0.02

0.00

Figure 71.

MOVLS test

these tests.

0.20

0.18

0.16

M3S $gd-0791-27

Traces of coefficient of friction in the stem nut versus time for the first three tests of the

sequence. Very little change in coefficient of friction is observed at the low thrust levels of

C

0

L)

U0

0)

a)

0

0:

0

0.14

0.12

0.10

0.08

0.06

0.04

0.02

0.00

FM350 1gd0791-28

Figure 72. Traces of coefficient of friction in the stem nut versus time for Tests 10 through 14 of the

MOVLS test sequence. A steady increase in coefficient of friction is evident in these tests. Coefficient of

friction is highest during Test 14, the test exhibiting load-sensitive MOV behavior.

NUREG/CR-5720 78
Understanding Diagnostics

words, the stem-to-stem-nut interface goes from a

thick-film lubricated surface to near metal-tometal contact. This hypothesis is further substantiated by Test 15, as shown in Figure 73, where

the coefficient of friction remains high during the

first unloaded test following the highly loaded

tests. During this test, the stem-to-stem-nut

interface is relubricating itself. During subsequent Tests 16 and 17, the lubrication returns to

thick-film lubrication.

This behavior has been observed with two stem

and stem nut combinations in the same operator.

Other stem and stem nut combinations in other

operators must be evaluated before any definitive

statements can be made, but if additional research

confirms the hypothesis, this provides an additional concern in bounding load-sensitive MOV

behavior. The coefficient of friction of a lightly

loaded stem varies with stem-to-stem-nut combinations. We have observed values from 0.08 to

0. 18; however, they all increase when the load on

the stem increases. In fact, we have observed that

the coefficient of friction can 'increase from 20

to 50%. A'typical in situ test is a static no-load

test, or at best a low-load test similar to those

shown in Figure 59. During these no-load static

or low-load valve tests, the stem and disc slide

into the valve seat basically unloaded and with

little or no pressure on the lubricated surfaces of

the stem-to-stem-nut interface. Consequently, the

unloaded test fails to generate the load necessary

to evaluate the stem-to-stem-nut coefficient of

friction that will exist when the MOV is highly

loaded. Therefore, additional margin must exist

to ensure closure against a high stem load. Means

to conservatively establish such an acceptable

margin still need to be developed.

Once an MOV has been tested, the no-load

relationships can be baselined and monitored for

degradation. However, it is important to recognize that valves that experience load-sensitive

MOV behavior at torque switch trip are not as

fully seated as those that do not.

These valves are operating in the margin

between successful and unsuccessful closure.

When possible, increasing the torque switch setting enough to fully seat the valve at design-basis

loadings will reduce or eliminate the effects of

load-sensitive MOV behavior.

4.4 Direct Current Powered

Motor Operators

Some of the differences in ac- and dc-powered

motor operators have been discussed in other

sections of this report. This section deals with

dc-powered motor operators with respect to

Generic Issue (GI)-87 concerns. In most BWRs,

one valve out of each pair of GI-87 isolation

valves is dc powered. As part of our GI-87

research, we looked at dc-powered motor operator performance. The results of the work show

that if the dc motor operator is correctly sized for

the valve application and the power cables are

large enough, the dc-powered motor operator will

probably equal or exceed the performance of an

equally sized ac-powered motor operator.

These results were determined from MOVLS

testing and earlier work performed in the field.

The MOVLS is described in Section 4.3 of this

report. The changes to the MOVLS for this testing

included removing the ac motor and installing a

125 Vdc, 40 ft-lb dc motor, and running the dc

motor controller field coil current through the

torque and open limit switches. Appropriate dc

instrumentation was also substituted for ac instrumentation such as voltage and current. The dc

testing used the same Limitorque SMB-0 motor

operator that had previously been used in the

motor-operated, load-sensitive behavior study

cited earlier in this report. This ensured that the

operator would not influence the results. After a

series of check-out tests, the motor operator was

subjected to a 15-cycle test series. Figure 74

represents typical thrust histories from these tests.

Simulated valve loadings included no-load tests

(Test I) to tests that bypassed the torque switch

and nearly stalled the motor (Test 15). As

expected from previous field experience, the

stroke time increased as the motor operator load

increased. The MOVLS monitored motor speed

through a monopole pickup that counts the teeth

on the worm shaft gear. In the highest loaded test

(Test 15), the motor speed was reduced from an

79 NUREG/CR-5720
Understanding Diagnostics

0.30 - ------

0.28 .

0.26-

0.24

c 0.22

0.20.

LL 0.18a

0.16

C: 0.14

0.1 2

~'0.1 0

O 0.08 Test 14 Test 15 Test 16 Test17

0.0 6

0.04

A Torque switch trip 0.02

0.00

M359 b ad0791-29

Figure 73. Traces of coefficient of friction in the stem nut versus time for the last four tests of the

MOVLS test sequence. A full opening/closing low-thrust cycle (Test 15) is needed following the test

exhibiting load-sensitive MOV behavior (Test 14) before the coefficient of friction returns to the levels

recorded for the initial low-thrust level tests.

0

-5,000

-10,000

' -15,000

O -20,000

0

E -25,000

' -30,000

-35,000 DC Test 1 DC Test 12 DC Test 15

-40,000

-45,000

M450 rs-1091-O1

Figure 74. Typical stem thrust histories from the dc-powered MOVLS tests.

NUREG/CR-5720 80
Understanding Diagnostics

unloaded speed of 1900 rpm to below 400 rpm,

at which time the current was manually interrupted. The stroke time, of course, increased

dramatically, going from an unloaded stroke time

of 6.0 seconds (Test 1) to a fully loaded stroke

time of 12.5 seconds (Test 15 of Figure 74). One

of our concerns before testing a dc-powered

motor operator was the effect of the lower stem

nut speed.

As the de motor slows down and increases the

time the stem nut is highly loaded, it may increase

the load-sensitive behavior'of the'motor operator

(rate-of-loading effect). This was not the case.

Figure 75 compares an unloaded coefficient of

friction history, Test 1 with Test 12, which shows

the increase in friction from a heavily loaded

operator, and with Test 15, which represents the

coefficient of friction resulting from an operator

that was subjected to twice the load used in

Test 12. The increase in the stem nut coefficient

of friction appears to have stabilized between

Tests 12 and 15 in spite of the increasing load.

This offers the possibility that the increase in the

stem nut coefficient of friction under load can be

bounded. However, additional testing with more

stem-to-stem-nut combinations will be needed to

verify such a hypothesis.

The load-sensitive behavior of a dc-powered

motor operator is very similar to that observed

during testing with an ac-powered motor operator, discussed earlier in this report. The'results of

the dc-powered motor operator may appear to be

worse because of the increase in the'stroke time;

however, the final thrusts and associated stem nut

coefficients of friction are nearly equal to

similarly loaded ac tests. The output torque of a

dc motor, as shown in the motor torque versus

speed curve presented in Figure 48, is linear and

not significantly influenced as the unit slows

down. On the other hand, the speed-torque relationship of an ac motor is relatively constant until

it reaches the knee in the motor torque versus

speed curve and can thus stall from a relatively

high speed. Consequently, marginally sized

motor operators equipped with an ac motor may

stall more easily than a unit equipped-with a dc

motor. In addition, we observed that this credible

-valve operation did not cause the dc motor to

overheat and adversely influence motor operator

performance any more than with an ac motor.

Direct-current motors do have potential installation problems. The higher current demands of a

dc motor before stalling, up to five times their

nameplate current, were not always recognized,

and the larger cabling necessary to handle these

large currents was not always installed. Units

operating with undersized cables at these higher

currents will experience cable heating, and subsequent voltage drops from cable heatup will

quickly influence the performance of a dc motor.

The difficulty with measuring the voltage drop in

a dc motor operator circuit were discussed earlier

in this report.

Figure 45, shown earlier in this report, shows

how quickly dc motor heating can reduce the output of the motor. Undersized cables and motor

heating from- a marginally sized motor can signif-

" icantly reduce the output of a dc motor-operated

valve. These installation problems have been

observed in operating plants and support the finding identified here.

81 NUREG/CR-5720
Understanding Diagnostics

0.20

0.18

C:

0

L_

C+)

I+-

0

C

C.,

0

C_

E

(D

CO

0.16

0.14

0.12

0.10

0.08

0.06

0.04

0.02

0.00

M450 rs-1 091-02

Figure 75. Typical stem nut coefficient of friction histories from the dc-powered MOVLS test. Note the

increase in friction between low-load Test 1 and high-load Test 12. The load was doubled between Tests 12

and 15 without a significant increase in stem nut coefficient of friction.

NUREG/CR-5720 82
5. WRITING A PERFORMANCE STANDARD

5.1 Overview standard, was finalized at the QME meeting in

New York in 1991.

Development of a performance standard is a

two-fold process: original qualification and

maintenance of qualification (in situ testing). The

current original qualification standard for motoroperated valves is ANSI/ASME B16.41, which

was developed in 1983 and reaffirmed in 1989.

This standard is being outdated by the results of

NRC and other valve testing. The valve subcommittee of ASME Qualification of Mechanical

Equipment (QME) has rewritten the document

under the title "QV, Functional Qualification

Requirements for Power-Operated Active Valve

Assemblies for Nuclear Power Plants" (same title

as B 16.41). QV will be a section in the QME

family of qualification standards. The rewritten

standard is an improvement over B16.41, but

incorporates the NRC test results findings only up

to the start of the analysis of Phase I gate Yalve

test results (DeWall and Steele, 1989). Subsequent analysis of Phase I testing, the results from

the Phase II testing, and the results of separate

effects testing on the INEL MOVLS have not

been incorporated.

The flow interruption section of QV, Annex G,

has been improved to incorporate the NRC butterfly and early gate valve test results, but it still

allows family groupings and extrapolation that

our test results have shown to be faulted. The

qualification philosophy is still go/no-go without

guidance for margins determination, which is a

definite concern for aging. The standard should

also address Generic Letter 89-10 concerns, as

there are currently no guidelines that would aid a

utility in procuring a replacement valve that

would meet the recommendations of the generic

letter. The QV document has been approved by

subcommittees and the main committee, over the

objections of INEL and NRC representatives. On

the second vote, the document was approved over

NRC objections. Subcommittee and main committee chairmen did agree to start the rewrite of

the standard to consider incorporating the recent

findings with regard to NRC and other testing.

This new task group, formed to rewrite the

Maintenance of qualification (in situ testing of

valves) is the responsibility of the ASME Operation and Maintenance (OM) Committee, which is

split into many subcommittees and working

groups with varying responsibilities. Those of

major interest here are OM-8 and OM-I0; of

minor interest are OM-18 and OM-19. OM-8 is

the working group on motor operators, OM-10 is

the working group on valves, and OM-18 and

OM-19 are the working groups on air and

hydraulic operators.

These OM groups have two major problems:

(a) they must, consider the in situ valves qualified

and write in situ tests accordingly, when in actuality every possible level of valve qualification

exists in the plants, and (b) the valve cannot be

considered separate from its operator, as it is by

OM. The three operator types, referenced above,

the valves, and their functions are different

enough that they should be considered as separate

integrated units instead of being separated by

operator type and valves. The OM-10 standard,

currently known as ISTC, has been issued and is

referenced in Section XI of the 1989 ASME

Code. Utilities are required to update their

in-service inspection plans every ten years, so, by

1999, all utilities will be using OM 10 for valve

testing requirements. Unfortunately, the requirements of OM-10 are only marginally better than

the previous Section XI requirements. OM-10

still treats stroke time as the measure of MOV

functionality. At the working group level, OM-10

is trying to add a requirement for motor current

signature analysis to the stroke time requirements. However, this is not expected to be

adopted. OM-8, driven by Generic Letter 89-10,

has redrafted their document for motor operators

into an integrated MOV standard. We believe this

is an improvement over anything else that has

been done for in situ testing of MOVs. Representatives of the INEL and NRC sit on the OM-8

committee and have informed the committee of

83 NUREG/CR-5720
Writing a Performance Standard

NRC-sponsored and other testing., The standard

reflecting this input is now in the balloting stage.

5.2 Detailed Observations

-The valve problem starts with qualification and

continues with the maintenance of qualification.

Most of the plants were built or in construction

before we had a valve qualification standard.

Each valve manufacturer had its own proprietary

methods. (mostly analytic) for determining

qualification and for sizing the operator. Globe

valves are more likely to be challenged in daily

operation than gate or butterfly valves. Early

globe valve problems included plug guidance

problems and bent stems. Most of these have been

corrected. As discussed earlier in the report, we

are not certain whether a gate valve is predictable

or nonpredictable without a design-basis test. We

also do not know if the extrapolation of butterfly

valve response was always performed in the

conservative configuration we identified in our

tests. The OM standards cannot help with these

problems. The: original qualification requirements for a valve and operator must be upgraded

for replacement valves. The upgraded qualification requirements can then serve as guidelines for

the OM standards to provide for the maintenance

of qualification.

NRC, utility, European, and Electric Power

Research Institute test programs may provide

insights on how to quantify the capability of

current in-place valves and quantify the risk that

might be involved with those designs. The recommendations contained in Generic Letter 89-10

and the improved diagnostic testing described

earlier in this report can help to identify some of

the more obvious problems. It is clear that designbasis tests cannot be performed for every valve.

We expect that, during the next year, more

information on these remaining problems will

become available. The INEL considers the

development of adequate standards a necessary

part of this work that must be performed to

improve the reliability of MOVs in the operating

plants.

NUREG/CR-5720 84
6. REFERENCES

Allis-Chalmers Corporation, Test Report on an Allis-Chalmers 6" STREAMSEAL Butterfly Valve in Air

Concerning Nuclear Containment Isolation Valves, VER-0209, December 17, 1979.

DeWall, K. G., and R. Steele, Jr., BWVR Reactor Water Cleanup System Flexible Wedge Gate Isolation lValve

Qualification and High Energy Flow Interruption Test, NUREG\CR-5406, EGG-2569, 1989.

MPR Associates, Inc., Application Guide for Motor-Operated Valves in Nuclear Powver Plants, NP-6660,

March, 1990.

NRC, NRC Action Plan as a Result of the TMI-2 Accident, Volumes I and 2, Revision, NUREG-0660,

August, 1980a.

NRC, Clarification of TMI Action Plan Requirements, NUREG-0737, November, 1980b.

Steele, R., Jr., and J. G. Arendts, SHAG Test Series: Seismic Research on an Aged United States Gate Valve

and on a Piping System in the Decommissioned Heissdampfreaktor (HDR), NUREG/CR-4977,

EGG-2505, 1989.

Steele, R., Jr., K. D. DeWall, and J. C. Watkins, Generic Issue 87 Flexible Wedge Gate Valve Test Program,

Phase II Results and Analysis, NUREG/CR-5558, EGG-2600, 1990.

Watkins, J. C., R. Steele, Jr., R. C. Hill, and K. G. DeWall, A Study of Typical Nuclear Containment Purge

Valves in an Accident Environment, NUREG/CR-4648, EGG-2459, 1986.

85 NUREG/CR-5720
NRC FORM 335 U.S. NUCLEARREGULATORYCOMMISSION 1. REPORT NUMBER

(2-W (A,9Iigned by NRC. Add Vol.. Suap., Rev., NRCM 1102, end Addendum Numbers. if eny.I

3201,3202 BIBLIOGRAPHIC DATA SHEET

(See instru::,ons on the reversel NUREG/CR-5720

2. TITLE AND SUBTITLE EGG-2643

Motor-Operated Valve Research Update 3. DATE REPORT PUBLISHED

MONTH Y YEAR

June 1992

4. FIN OR GRANT NUMBER

A6857, 135529

S. AUTHORJSI 6. TYPE OF REPORT

Robert Steele, Jr.

John C. Watkins Technical

Kevin G. DeWall 7. PERIOD COVERED s-ceusy Dales/

Mark J. Russell

8. PERFORMING OAGAN I Z ATIO -NA MEUDAD R SS (I NR C. mefuon~ademailinnssiton. ttc AginU..NcerReuaory Co- iowide

nane and mailing addrr. e

Idaho National Engineering Laboratory

EG&G Idaho, Inc.

P.O. Box 1625

Idaho Falls, Idaho 83415

9. SPONSORING ORGANIZATION - NAME AND ADORESS (I AIR C. tyPe 'Su- as Joue-: i consjactor., proide NRC DicIision. Oftice or Region. UV. N.~ucteRegstfrogtComnisi

and ma;Iing aorddesJ

Division of Engineering

Office of Nuclear Regulatory Research

U.S. Nuclear Regulatory Commission

Washington, D.C. 20555

10. SUPPLEMENTARY NOTES

1 1. ABSTRACT 1200 words or less

This report provides an update on the valve research sponsored by the U.S. Nuclear Regulatory Commission (NRC)

that is being conducted at the Idaho National Engineering Laboratory. The update focuses on the information

applicable to the following requests from the NRC staff:

* Examine the use of in situ test results to estimate the response of a valve at design-basis conditions

* Examine the methods used by industry to predict required valve stem forces or torques

* Identify guidelines for satisfactory performance of motor-operated valve diagnostics systems

* Participate in writing a performance standard or guidance document for acceptable design-basis tests.

The authors have reviewed past, current, and ongoing research programs to provide the information available to

address these items.

12. KEY WORDSIDESCR:PTORS (List woro's or phrases thai will assist researchers in ocaring rhe rworro. 13. AVAILABILITY STATEMENT

Motor-operated valve (MOV) Unlimited

Gate valve-stem force 14. SECURITY CLASSIfICATION

Butterfly valve-stem torque (rAhs Pagel

MOV diagnostic systems Unclassified

IrhAs Repoen

Unclassified

15. NUMBER Of PAGES

16. PRICE

aRC FORM 335 124191
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