NUREG/CR-5807

KEI No. 1721

Improvements in Motor Operated

Gate Valve Design and

Prediction Models for

Nuclear Power Plant Systems

SBIR Phase I Final Report

September 1990-April 1991

Prepared by

J. K. Wang, M. S. Kalsi

Kalsi Engineering, Inc.

Prepared for

U.S. Nuclear Regulatory Commission
I

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NUREG/CR-5807

KEI No. 1721

Improvements in Motor Operated

Gate Valve Design and

Prediction Models for

Nuclear Power Plant Systems

SBIR Phase I Final Report

September 1990-April 1991

Manuscript Completed: July 1991

Date Published: May 1992

Prepared by

J. K. Wang, M. S. Kalsi

W. S. Farmer, NRC Project Manager

Kalsi Engineering, Inc.

745 Park Two Drive

Sugar Land, TX 77478

Prepared for

Division of Engineering

Office of Nuclear Regulatory Research

U.S. Nuclear Regulatory Commission

Washington, DC 20555

NRC FIN L1667
ABSTRACT

This research is aimed at improving the performance of gate valves at nuclear power

plants (1) by developing improved predictive models and (2) by identifying design

improvements that overcome problems/limitations of the current gate valve designs.

Phase I research is aimed at developing improved operating thrust models for the most

common types of gate valves in use at U.S. nuclear power plants. The research completed

under Phase I addresses shortcomings in the current motor operated gate valve performance models by investigating localized contact stresses under disc tilting caused by fluid

flow, by predicting inertial thrust overshoot, and by providing a comprehensive review of

friction/galling data for gate valves. Instrumented valve test data provided by Duke

Power Company were used to make limited comparisons with the analytical predictions.

The areas that require systematic testing to further refine the predictive models are

identified.

iii
TABLE OF CONTENTS

Page

EXECUTIVE SUIMMARY 1

1. INTRODUCTION

1.1 Background

1.2 Objectives

1.3 Technical Approach and Summary

2. GATE VALVE TYPE, GEOMETRY, AND ITS EFFECTS ON OPENING AND

CLOSING THRUSTS

2.1 Stem Thrust for Solid, Flexible, and Split Wedge Gate Valves

2.1.1 Closing Stem Thrust to Overcome Gate Differential Pressure

2.1.2 Opening Stem Thrust to Overcome Disc Differential Pressure

2.1.3 Stem Wedging Load - Closing

2.1.4 Stem Unwedging Load - Opening

2

2

2

3

5

7

7

8

8

9

2.2 Stem

2.2.1

2.2.2

2.2.3

Thrust for Parallel Expending Gate Valves

Stem Thrust to Overcome Gate Differential Pressure

- Closing and Opening

Stem Wedging Load - Closing

Stem Unwedging Load - Opening

9

9

10

10

2.3 Stem Loads for Parallel Sliding Gate Valves - Closing and Opening 11

2.4 Total Stem Thrust Requirements

2.5 Effective Gate Sealing Diameter

2.6 Disc Tilting Due to Fluid Flow

2.6.1 Estimating Flow-Induced Load on Disc in Mid-Travel Position

2.6.2 Disc Tilting and Its Interaction with Seats and Guides

2.6.3. Contact Stresses

12

13

14

15

19

22

v
3. COEFFICIENT OF FRICTION AND THRESHOLD OF GALLING STRESS 25

3.1 Relationship Between Disc Factor and Coefficient of Friction 25

3.2 Data for Coefficient of Friction from Principal Investigators' Experience 27

3.2.1 Background 27

3.2.2 Results from Room Temperature Water Tests 27

3.2.3 Results from High Temperature Water and Steam Tests 28

3.2.4 Long-Term Surveillance Tests on SIS Valves Under Flow and

AP 29

3.3 Contact Stress and Threshold of Galling 30

3.3.1 Threshold of Galling for Stellite vs Stellite and Other Valve

Trim Materials 30

3.3.2 Average and Local Contact Stresses 33

3.4 Duke Power Data for 4-inch Borg-Warner Flexible Wedge Gate Valve 34

3.4.1 Summary and Comparison of Data for

Carbon Steel vs Stainless Steel Valves 36

3.5 KWU-Siemens Test Data 38

3.6 UK PWR Valve Testing 39

3.7 NRC-Sponsored INEL Test Data 39

3.8 Conclusions from Presently Available Friction and Galling Data 41

4. PREDICTION OF THRUST OVERSHOOT DUE TO INERTIA 43

4.1 Description of the Inertial Overshoot Phenomenon 43

4.2 Available Energy After Torque Switch Trip (TST) 45

4.2.1 Motor Work After Torque Switch Trip 45

4.2.2 Kinetic Energy of Moving Components 46

4.3 Stored Energy in Valve Components After Torque Switch Trip 46

4.4 Energy Dissipated After Torque Switch Trip 48

4.5 Energy Balance and Final Thrust Prediction 49

4.6 Comparison of Predicted Final Thrust Against Test Results 50

vi
5. FACTORS AFFECTING OPENING THRUST REQUIREMENTS 51

5.1 Effect of Wedging Force From the Previous Closing Operation 51

5.2 Effect of Higher Bonnet Pressure on Some Gate Valve Designs 53

5.3 Effect of External Piping Loads 53

5.4 Effect of Temperature Changes on Opening Thrust 54

6. CONCLUSIONS i

7. REFERENCES M

APPENDIXA: Opening and Closing Stem Thrusts for Variation in Disc Design

APPENDIX B: Disc Load Calculations for a Gate Valve in a Pump Flow System

APPENDIX C: Analysis of Local Contact Stresses

APPENDIXD: Analysis of Stem Thrust Overshoot for 4-Inch Borg-Warner Valve

Tested at Duke Power Flow Loop

APPENDIX E: Analysis of Piping Load Effect on Opening Thrust

APPENDIX F: Analysis of Temperature Effect on Opening Thrust

vii
LISI OF FIGURES

No. Page

2.1a Conventional Solid Wedge, Flexible Wedge, and Split Wedge Gate Valves 5

2.1b Parallel Expanding Gate Valves 6

2.1c Parallel Sliding Gate Valve 6

2.2 Gate Equilibrium Under AP Load During Closing 7

2.3 Gate Equilibrium Under AP Load During Opening 8

2.4 Gate Equilibrium under Wedging Load During Closing 8

2.5 Gate Equilibrium under Unwedging Load During Opening 9

2.6 Gate Equilibrium under AP Load During Closing/Opening 9

2.7 Gate Equilibrium under Wedging Load During Closing 10

2.8 Gate Equilibrium under Unwedging Load During Opening 11

2.9 Gate Equilibrium under Under AP Load During Closing 1

2.10 Overall Stem Load Equilibrium 12

2.11 Effective Gate Sealing Diameter 13

2.12 Flow Resistance Coefficient for Gate Valves Based on an Average

from Several Designs 16

2.13 Balance of Available Pump Head and Piping System Pressure Drop 17

2.14 Typical Pressure Drop Across a Gate Valve as a Function of Gate

Position in High Pressure Pumped Flow System 18

2.15 Point Contact Against the Downstream Seat Due to Disc Tilting at a

Typical Mid-Travel Position in 4", 1500# Borg-Warner Gate Valve [15] 20

2.16 Ideal Gate Slide 21

2.17 Tilted Gate Contacting Guides 21

2.18 Tilted Gate Contacting Seat 22

ix
No. Page

3.1a Radial Seat Contact Stress Variation 33

3.lb Circumferential Seat Contact Stress Variation 34

3.2 Increasing Friction Trend During 4-Inch Borg-Warner Gate Valve

Cycle Testing by Duke Power, Reference 15 35

4.1 Typical MOV Closing Sequence for Gate Valve-Stem Thrust Versus

Time [251 44

x
ACKNOWLEDGEMENTS

We are grateful to the United States government for the $50,000 funding provided under the

Small Business Innovation Research (SBIR) Program which made this research possible.

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

encouragement throughout the project.

Duke Power Company and Neal Estep of Duke Power deserve special recognition for

giving permission to use their flow loop test data and the results of analyses performed for

them.

The authors would like to thank Bobbie Lambert for patiently working through many

revisions and for her painstaking attention to detail in preparing this manuscript.

xi
EXECUTIVE SUMMARY

This report documents the results of Phase I research proposed and conducted by Kalsi

Engineering, Inc. to improve the operability of motor-operated gate valves in nuclear

power plants. Phase I research, funded by the Small Business Innovation Research (SBIR)

program, resulted in the following major accomplishments:

* Opening and closing thrust equations for the common types of gate valves used in

U.S. nuclear power plants have been developed and documented.

* An analytical methodology to predict inertial thrust overshoot in an MOV gate valve

has been developed from first principles. Comparisons against data supplied by

Duke Power Company have confirmed that the methodology is sound, and there is

good quantitative agreement between analytical predictions and actual test results.

* The results of a comprehensive review of friction and galling data are documented

in this report to provide a rational basis for selecting an appropriate coefficient of

friction for a given application.

* The concept of index of contact stress severity has been introduced to determine

whether or not a gate valve will behave predictably under fluid flow forces.

Preliminary analysis approaches to calculate localized contact stresses at the discto-guide contact and at the disc-to-downstream seat contact under disc tilting

conditions have been developed.

* Significant factors that affect the opening thrust requirements of a gate valve have

been identified, and quantitative methods that can be used to diagnose valve opening

problems have been documented.

* Improvements in gate valve designs to make them less sensitive to pressure/thermal

transients and external pipe loads have been identified, and some quantitative

examples are included to show the degree of improvement achievable.

In summary, the Phase I research has been successful in completing the preliminary

development of improved gate valve operability models. This can serve as an excellent

foundation to continue further analytical and experimental development that is necessary

to provide reliable and proven gate valve operability models to the nuclear power industry.

-1-
1. INTRODUCTION

1.1. Background

Operability problems with motor operated valves (MOVs) in the U.S. nuclear power plants

have been extensively documented in references 111* through 1121. The Nuclear Regulatory

Commission's (NRC) safety concerns regarding the operability of the M1OVs were further

reinforced by the failure of several gate valves to close within the manufacturer specified

thrust requirements under simulated high energy pipe break conditions in the recently

conducted NRC tests [4,23]. Even though several of the earlier reports and surveys have

described the operability problems in detail 17,8,9:1, it was the issuance of NRC IE Bulletin

85-03 [1], Generic Letter 89-10 121, and the recent gate valve tests [4,23] that finally resulted

in the industry-wide recognition of the significance of the MOV problems.

Kalsi Engineering, Inc., having been intimately involved in solving valve problems for

the utilities for over 13 years, submitted a proposal under the Small Business Innovation

Research (SBIR) Program to initiate a systematic research directed at improving the motor

operated gate valve designs and operability prediction models. This proposal was selected,

thus giving the principal investigators an opportunity to document the key technical

approaches that have been developed by Kalsi Engineering, Inc in solving problems and

making improvements in the design and performance of MIOVs.

1.2. Objectives

The overall objectives of Phases I and II of this SBIR project defined by the principal

investigators are to improve the operability of the motor operated gate valves in the nuclear

power plant safety system by: (1) developing more comprehensive and reliable models for

predicting operability; (2) identifying improvements that can overcome the deficiencies

observed in the current gate valve designs.

The overall objectives stated above are quite broad and would require extensive analytical

research as well as testing to fully accomplish these goals. The more specific objectives

under SBIR Phase I funding constraints were limited to: (1) providing closing and

opening thrust equations (based on ideal free-body diagrams) for the types of gate valve

designs which are in common use in the U.S. nuclear power plants; (2) developing a

preliminary analytical methodology to quantitatively assess the effect of fluid flow forces

imposed on the disc on the. valve performance; (3) developing an analytical model to

predict thrust overshoot due to inertia; (4) documenting friction and galling data,

including principal investigators' experience, relevant to gate valve operability; (5)

documenting factors that affect the gate valve opening thrust requirements; and (6)

identifying possible improvements in the gate valve design.

* Numbers in brackets denote References listed in Section 7.

-2-
1.3. Technical Approach and Summary

The objectives of Phase I research were accomplished by first developing stem thrust

requirements for closing and opening the gate valves of several different designs that are

in use in the U.S. nuclear power plants: (1) conventional gate using solid, flexible and

split wedge designs, (2) parallel (expanding) wedge gates of through-conduit and doubledisc design, and (3) parallel slide gate. These results are summarized in Section 2 with

detailed derivations included in Appendix A.

An important factor that has not been adequately considered in the gate valve designs by

many manufacturers is the effect of fluid flow forces on the disc along the flow direction.

In some valve designs, the fluid forces can cause titling of the disc in mid-travel position

resulting in localized loading at the edges of the disc guides or between the disc and the

downstream seat. As evidenced in the NRC sponsored and other tests [4,21,23], severe

damage can occur to the internals of such valve designs by the high fluid flow forces under

blowdown conditions. Even though the potential for damage is expected to be considerably

lower when these valves are operated under less severe, pumped flow conditions,

quantitative approaches to predict their performance have not been available.

Section 2 of this report summarizes a preliminary analytical approach that has been

developed to estimate the mid-travel disc loads and local contact stresses due to fluid forces

at the potential areas of contact. These preliminary local contact stress calculations are

based on simplifying assumptions of linear, small displacement elasticity equations.

The limitations of this approach and further refinements that are needed in this area are

identified. The concept of using these results as an index of contact stress severity is

introduced, which can be used as a design guide in making comparisons between different

valve geometries and loading conditions. To accurately account for non-linear behavior

at the contact due to localized yielding and wear, further refinements using large

displacement elastic/plastic finite element analysis, as well as systematic testing to

obtain empirical correlations over a wide range of parametric conditions are needed.

Another deficiency in predicting the operability of the MOVs has been the lack of

analytical techniques to quantify the effect of inertia on the thrust overshoot. Currently,

inertial overshoot problems are detected only by MOV testing. The Phase I research

overcomes this deficiency by documenting an analytical approach, based on first

principles, that has been developed to quantitatively predict the inertial overshoot. Results

show good comparisons against actual test data provided by Duke Power Company from

their flow loop testing [151. Section 4 and Appendix D present these results.

Coefficient of friction between the disc and seats is one of the dominant factors that determines the overall thrust requirements for gate valves. Even though Stellite hardfacing

-3-
has been used for years as a standard overlay material at the seating faces by most valve

manufacturers, test data for its coefficient of friction performance span an extremely wide

range. This is largely due to the fact that tribological behavior at the sliding contacts in the

gate valve is affected by several factors that are not explicitly understood and controlled

during valve tests. Section 3 of this report presents a summary of the coefficient of friction

and galling data based on the principal investigators' laboratory testing and field experience, with a focus on those factors that have the most impact on the performance of gate

valves. Condition of the contact surfaces, e.g. the undetected presence or absence of galling

or absorbed contaminant layer of lubricants (even of molecular scale) can account for very

large differences in the "apparent" coefficient of friction. Systematic testing using real

gate valve internals is needed to determine the onset of galling in local areas of contact,

and to determine conditions under which continued cycling results in increases in

apparent friction and seizure or stable frictional behavior after local wear.

Lastly, the opening thrust requirements for wedge gate valves are influenced by several

factors that are not easily quantifiable in actual MOV applications. The unwedging thrust

during opening is influenced by the wedging force from the previous closing cycle,

external pipe load causing disc pinching, effect of higher pressure trapped in the body

cavity resulting in energization of both upstream and downstream discs in some of the

gate valve designs, and thermal binding caused by temperature transients. The practical

approach to minimize the effect of these variables on operability performance has been to

bypass the torque switch during the initial portion of the opening stroke, thus making

maximum actuator output available. The problems are detected only when the magnitude

of these effects exceeds the actuator output. Section 5 and Appendices E and F discuss these

factors and present analytical methods that can be employed by the utility engineers and

valve manufacturers to detect and eliminate such problems.

In summary, Phase I research has shown that the operability predictions of MOV gate

valves can be significantly improved by properly taking into account a number of factors

that have been largely ignored in the past. Further analytical refinements and extensive

testing are needed to systematically address the areas identified in this report to develop

reliable operability prediction models that cover the wide range of variations in the valve

designs present at the nuclear power plants.

-4-
2. GATE VALVE TYPE, GEOMETRY, AND ITS EFFECT ON

OPENING AND CLOSING THRUSTS

There are five different types of gate valves that cover most of the applications in nuclear

power plants in the United States. The key features of these designs are shown in Figure

2.1. Variations in the most commonly used gate valves include solid, flexible, and split

gates (Figure 2.1a). The two types of parallel expanding wedge gates shown in Figure 2.1b

are also used, but their population is smaller. Parallel sliding gate valves shown in

Figure 2.1c are relatively uncommon in the United States, but are widely used in European

nuclear power plants. The advantages and disadvantages of various design features for

these valves are discussed in detail in Reference [131

C I

Flexible Wedge Solid Wedge Split Wedge Gate Gate Gate

Figure 2.1a

Conventional Solid Wedge, Flexible Wedge, and

Split Wedge Gate Valves

As shown in these figures, the designs vary significantly in gate geometries. Other

important variations that affect performance are related to gate guide arrangements and

their dimensions; clearances at critical locations between gate, guides, and seats; seat

contact widths; and materials and surface finish in the disc guide sliding interfaces.

Section 2 presents the gate thrust requirements for the above-described variations in gate

geometries. This section also addresses the potential for disc tilting during mid-travel

due to fluid forces across the disc. Disc tilting causes localized loading between the disc

and the downstream seat, or between the disc and the guides. A preliminary analysis

approach to determine the localized contact stresses is presented in this section to

determine the loading severity based upon valve design and operating conditions.

- 5 -
Preliminary analyses of localized contact stresses between disc and seats as well as disc

and guides used in typical wedge gate valve designs are presented in this section. The

preliminary approach presented here needs further analytical refinement and empirical

correlations to develop improved predictive models. Detailed derivations of the equations

summarized in this section are included in Appendices A, B, and C.

n j>;--Stem

g a b W W~~~~~Upers 2 stream

Upstream Lower__

DiscWeg

Body~ t et I Stop Pad

Segment Gate

Figure 2.1b

Parallel Expanding Gate Valves

-Stem

Retaining

Figure 2.1c

Parallel Sliding Gate Valve

-6 -
2.1. Stem Thrust for Solid, Flexible, and Split Wedge Gate Valves

Even though there are differences in the performance of solid, flexible, and split wedge

gate valves as related to their sensitivity to external piping loads and thermal binding [131,

the equations for their stem thrust requirements based upon free body considerations are

the same. Subsections 2.1.1 through 2.1.2 summarize the stem thrust requirements to

overcome only the differential pressure load across the disc. Subsections 2.1.3 and 2.1.4

give the stem wedging and unwedging thrust requirements to close and open the gate,

respectively. The total stem thrust requirements to close and open the gate are provided in

Section 2.4, which include other components such as stem packing load, stem rejection

force (also referred to as blowout force or piston effect force), and stem and gate weight.

2.1.1. Closing Stem Thrust to Overcome Gate Differential Pressure

As shown in Section A.1.1 of Appendix A, the stem thrust at the gate to overcome the

differential pressure during closing can be expressed as:

Fs=# (> 0 A On 0Js Fp (Eq. 2.1)

Fs = stem load at gate, lb

Fp = disc pressure load due to upstream/downstream

differential pressure, lb

= AP x (effective seat area)

Figure 2.2 = coefficient of friction between gate and seat

Gate Equilibrium Under 0 = 112 of gate wedge angle, deg

AP Lead During Closing

The disc pressure load, Fp, is the product of AP and seat area based on effective disc sealing

diameter as discussed further in Section 2.5.

From Equation 2.1 the relationship between the commonly-used term disc factor (sometimes called value factor) and coefficient of friction, gl, can be derived:

Disc Factor= A- (Eq. 2.1a) cos0- g sin 0

For typical wedge gate valves that use a total wedge angle of around 10 degrees (or 0 = 50)

and a normal range of coefficients of friction, the difference between the disc factor and the

coefficient of friction is practically negligible, as discussed in Section 3.1. The disc factor

calculated in the closing direction can be as much as 5 percent higher than the coefficient of

friction for typical values of 0 and g that are encountered in practice.

- 7-
2.1.2. Opening Stem Thrust to Overcome Disc Differential Pressure

As derived in Section A.1.2 of Appendix A, stem thrust during opening of a wedge disc

against a differential pressure is given by:

F. = ( 1LJFp (Eq. 2.2)

YL X F,,

From this one can derive the equivalence between the disc

factor in the opening direction and the coefficient of

friction:

4X 0 Disc Factor = 1

cos 0 + [I sin 0 (Eq. 2.2a)

Figure 2.3

Gate Equilibrium Under

AP Load During Opening

The disc factor in the opening direction is slightly less

than the coefficient of friction for typical ranges of wedge

angles and coefficients of friction (within 5 percent of the

coefficient of friction), as discussed in Section 3.1.

As stated earlier, the stem force calculated in Equation 2.1 or 2.2 is the force required to

overcome the differential pressure resistance only.

2.1.3. Stem W~edgingLoad - Closing

i S

4-

The stem wedging load is related to the normal seat

contact force, Fn, as shown in Section A.1.3 of

Appendix A:

KL F= 2 (sin 0 + p± cos 0) F, (Eq. 2.3)

It should be noted that this equation applies to the

case when there is no differential pressure across

the gate. When differential pressure is present,

the stem force Fs in this equation is the net stem

force after subtracting the differential pressure

load.

Figure 2.4

Gate Equilibrium under

Wedging Load During Closing

In some cases, the limit switch instead of the torque switch is used to stop the disc travel in

the closing direction. Where acceptable from the shut-off standpoint, this approach can be

used to reduce, and in some cases eliminate, the wedging load, Fn.

-8-
2.1.4. Stem Unwedging Load - Opening

Section A. 1.4 of Appendix A shows that the unwedging load to overcome the seat contact force, Fr, is

given by:

F=2 ( cos 0- sin 0) F, (Eq. 2.4)

F1 YLf

Figure 2.5

Gate Equilibrium under

Unwedging Load During Opening

The seat contact force, Fn, that is to be overcome during the opening cycle is developed by (1) wedging

load from the previous closing cycle, including

inertia overshoot, (2) external piping loads, or (3)

differential thermal effects between the valve body

and disc. Section 4 provides an analytical methodology to predict stem thrust due to inertia overshoot,

and Section 5 discusses external pipe load and thermal effects that may influence the normal load, Fn.

2.2. Stem Thrust for Parallel Expanding Gate Valves

This Subsection 2.2 summarizes the stem thrust requirements for closing and opening

directions for the two types of parallel expanding gate valves shown in Figure 2.1b. The

same stem thrust equations apply to both types of parallel expanding gate valves shown in

this figure. The typical wedge angle used in the through-conduit type is 15 degrees, and for

the double-disc type is 25 degrees. It should be noted that for coefficient of friction of 0.47

(= tan 250) or less, the 25-degree angle between the wedge surfaces (also referred to as back

angles) provides a non-locking condition between the wedges.

2.2.1. Stem Thrust to Overcome Gate Differential Pressure- Closing and Opening

F_

As shown in Section A.2.1 of Appendix

A, the following equation applies to

both closing and opening stem thrusts

to overcome gate frictional force due to

AP load:

Fs = it Fp (Eq. 2.5)

where

= coefficient of friction

between seat and disc

Fp = disc pressure load due to

upstream/downstream

differential pressure, lb

= AP x (effective seat area)

Closing Opening

Figure 2.6

Gate Equilibrium Under hP Load During

Closing/Opening

-9-
2.2.2. Stem Wedging Load - Closing

The stem wedging load for a parallel expanding gate valve is shown in Section A.2.2 of

Appendix A to be given by:

| =F (It + sin 0 + iu'cosO) F. (Eq. 2.6) FS ~~~~~~~~cos 0 - p.' sin 0

whvl ere

\OnY ip. = coefficient of friction between seat and

Fir \ p x disc

F'%fi 11 = coefficient of friction between wedge

.IC faces

i 0 = parallel gate total wedge angle, deg

Fn = normal force between gate and seat due to

Figure 2.7 wedging, lbs

Gate Equilibrium Under

Wedging Load During Closing

This equation makes allowance for the fact that the coefficients of friction at the seat-todisc interface may be different than that at the wedge interface. Typically the seat faces

have a finer surface finish and are overlaid with Stellite hard-facing, whereas the wedge

faces have a rougher surface finish and are not hard-faced.

If the coefficient of friction at the seat faces and the wedge faces is assumed to be the same,

p.' = p., and this equation reduces to

(sin 0(1- p2) + 2. cos 0 F= I Fn (Eq. 2.6a) cos0-pgsin0n 1

Equation 2.6a shows that the stem load is proportional to the seat contact force, Fn.

2.2.3. Stem UnwedgingLoad- Opening

The stem unwedging load to overcome the seat contact force, Fn, for a parallel expanding

gate valve is given by (reference Section A.2.3, Appendix A):
hI_

F. (L'- 1)sinO+(iI+ 4')cos 0F (Eq. 2.7)

cos 0 + 4' sin 0

i-L' For 11 = g', this equation reduces to:

fake) f (~~~~~~sin 0(A 2_1) + 2pcos 0

9 F- (sine + si Fs (Eq. 2.7a) ~~^\-6 ~~~~cosO+psinO

Figure 2.8

Gate Equilibrium Under

Unwedging Load During Opening

As discussed in Section 2.1.4, the seat contact force Fr to be overcome is determined by

adding the wedging force from the previous closing cycle to the resultant force from

external piping loads and differential thermal expansion loads between the body and disc.

2.3. Stem Loads for Parallel Sliding Gate Valves - Closing and Opening

Most parallel sliding gate valves are equipped with a preloading spring to maintain proper

contact and provide a low pressure seal between the disc and seats. As shown in Appendix

A, Section A.3.1, the required stem thrust to overcome AP and spring load friction can be

expressed as:

F. =2 gRp + g Fp (Eq. 2.8)

where F5p = disc spring load, lb

Fp =AP x (effective seat area), Ibs

rF;

Don fLIp i*eom sate soevm ged

Figure 2.9

Gate Equilibrium Under AP Load During Closing

- 11 -
The above equation applies to both closing and opening directions, and the pressure load is

applied to the downstream disc only. It should also be noted that, since the seat faces and the

disc faces are parallel, there is no wedging or unwedging load associated with this type of

design.

2.4. Total Stem Thrust Requirements

The total stem thrust requirements for a gate valve stem can be determined by a

summation of all the loads applied to the stem. For stem thrust requirements in the closing

direction, a summation of the forces yields:

F = 0

l1 F1

where

Ai C

= F2+F3+F 4 +F5+F6-Fwl-Fw2

F1 = required stem thrust for closing, lbs

F2 = stem packing load, lbs (see Section 8)

F3 = stem piston load, lbs

= itI4 dS2 x AP, where ds is the stem

diameter

(Eq. 2.9)

i. 3

t Few

F4 = stem load to overcome gate AP as

summarized in Sections 2.1 through

2.3 for different types of gate valves, lbs

F5 = stem wedging load as summarized in

Section 2.1 and 2.2, lbs

F6 = stem torque reaction load, lbs

d

where M3 = Ml - M2 = stem torque from actuator

- stem packing torque, in-lbs

g "= coefficient of friction at the torque reaction

contact surface (usually at the gate guides)

d = moment arm for the torque reaction forces, in.

Fwi = stem weight, lbs (often negligible compared to

other forces)

FW2 = gate weight, lbs (often negligible compared to

other forces)

Figure 2.10

Overall

Stem Load

Equilibrium

Detailed derivations for each load component are given in Appendix A and summarized

in Sections 2.1 through 2.3.

The overall stem force equilibrium for valve opening is basically the same with F1 , F2 , F4,

F5, and F6 in reverse direction:

F1 = F2 - F3 + F4 + F5 + F6 + Fwl + Fw2 (Eq. 2.10)

- 12-
2.5. Effective Gate Scaling Diameter

The pressure loads calculated in the previous sections are computed based on the gate

differential pressure multiplied by the effective pressure area. The pressure area depends

on the effective sealing diameter, ds. It is an imaginary diameter that seals the upstream

high pressure from leaking into the downstream seat inside diameter. The higher

upstream pressure surrounds the upstream side of the gate and the downstream side of the

gate up to the effective sealing diameter. The area inside the sealing diameter, ds, on the

downstream side of the disc is considered exposed to the lower downstream pressure.

Based on this definition of effective sealing diameter,

ds, the pressure load on the gate can be expressed as:

_p =t AP (4 d) 2(Eq. 2.11)

Sea~t- 4

Without actual testing, the sealing diameter estimate

can be based on engineering judgement considering

the differences in various designs and experience.

Figure 2.11 The key factors that influence the sealing diameter

Effective Gate Sealing Diameter are discussed below:

* Disc stiffness. Elastic deflection of the gate under a differential pressure load

creates a higher local contact stress closer to the seat inside diameter as shown in

Figure 3.1. This tends to bias the effective sealing diameter towards the seat I.D.

The high local contact stress keeps the higher pressure fluid from leaking to the lower

pressure on the downstream side. The disc flexibility effect usually is more

pronounced on large size gates.

* Seat edges. The seat edges at the inside and outside diameters are usually

chamfered. The intersections between chamfers and seat faces are normally

rounded and polished to remove sharp edges. The actual operation of a valve under

differential pressure also causes localized wear or yielding of the seat edges

(especially on the inside diameter) under high local contact stresses. This can result

in some increase of the effective sealing diameter above the seat inside diameter.

* Uneven seat contact. Disc and seat deflections under pressure usually result in

uneven circumferential seat contact due to their uneven stiffness and support, as

shown in Figure 3.1. This uneven seat contact pressure distribution also affects the

equivalent sealing diameter used in the pressure load calculation.

- 13-
The exact contribution of all of these factors on the effective seat diameter is hard to

quantify without testing. In the absence of additional data, the following simple

approach has been found to be adequate, and is recommended:

- Use the mean seat diameter for a narrow seat:

ds = 1/2 (Seat I.D. + Seat O.D.) (Eq. 2.12a)

- Use a sealing diameter closer to the seat inside diameter for a wide seat. An

approximation sometimes used for wide seat faces is

ds = Seat I.D. + 1/3 (Seat O.D. - Seat I.D.) (Eq. 2.12b)

It is not uncommon to see seat I.D. being used as the effective sealing diameter in some

cases, with the objective of determining a conservative coefficient of friction value, as in

the case of KWU-Siemens data presented in Section 3.5. In using any of the coefficient of

friction data to predict operating forces in other gate valves, it is important to use the same

assumptions regarding the effective sealing diameter that were used in reducing the test

data. This is sometimes overlooked, thus adding unnecessary conservatism.

2.6. Disc Tilting Due to Fluid Flow

During valve closing, the disc moves into the flow stream and interrupts the steady flow.

The projection of the disc into the flow stream behaves like a blunt body in the flow stream

and is subjected to fluid dynamic forces along the flow direction. As the disc advances, it

increases the flow path resistance, thus resulting in an increase in differential pressure

load on the disc. Although the actual pressure distribution on the disc in mid-travel

position is difficult to accurately quantify without computational fluid dynamic analysis

or instrumented valve testing, it is certain that the resulting fluid dynamic force will push

the disc in the downstream direction. In some gate valve designs, this fluid force tends to

cause tilting of the disc during mid-travel. The magnitude of this force depends upon the

flow rate and hence the differential pressure across the disc. A tilted disc sliding under a

high differential pressure load in mid-travel position (such as encountered during

downstream pipe rupture) can develop high local contact stresses in the valve components.

As observed in the NRC-sponsored high energy pipe break tests [4,23,251, severe galling can

occur in disc guide slots, disc guides, and/or seat faces depending upon guide clearances

and other specific features of the individual valve design. Under pump flow conditions,

the magnitude of the pressure drop across the disc is significantly lower than that

experienced under blowdown conditions related to a downstream pipe rupture.

- 14 -
The effect of flow on disc closing was analyzed for a 4-inch Borg-Warner flex wedge gate

valve tested at Duke Power Company's flow loopl. Several key factors that affect disc

tilting are discussed and quantified in this section using available valve and flow data.

This analysis shows how the internal parts of this 4-inch gate valve interact during opening or closing under differential pressure. The design details of disc guide length, guide

strength, clearances, and corner radii can significantly affect the valve performance, as

discussed in the following analysis. The analysis also points out the need for better

quantification of gate loads in the mid-travel position. Appendix B includes detailed

calculations and procedures used in the analysis of the 4-inch Borg-Warner valve.

Additional description of this valve and the testing that was done on it by Duke Power are

given in Section 3.4 of this report. A description of the analysis and key conclusions is

given below.

2.6.1. Estimating Flow-Induced Load on Disc in Mid-Travel Position

The fluid flow forces to which the disc is subjected during closing or opening of the valve

have not been addressed adequately in many gate valve designs. This section presents a

simplified approach to estimating the fluid dynamic force imposed on the disc during

closing under typical pump flow conditions.

For gate valves, the average flow resistance coefficient as a function of disc opening from

a number of gate valve designs is derived in Appendix B based on References 30 and 31.

The results are shown in Figure 2.12.

1 J. K. Wang and M. S. Kalsi. Valve Factor Analysis for 4- and 6-inch Borg Warner Flex Wedge Gate

Valves, proprietary report for Duke Power Company, KEI Document No. 1646, April 1990.

- 15 -
260

240

220

200

_ 180

*3 160 -

.4)

0 140

t 120 -

a 100 -

C) 80 t

¢ 60 - 0

40 -

20-

0 T

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0

Gate Position, h/Do

Figure 2.12

Flow Resistance Coefficient for Gate Valves Based

on an Average from Several Designs

The flow resistance decreases rapidly as the disc opens. At any disc opening, pressure

drop across the disc can be estimated using the flow resistance coefficient and the fluid

flow velocity by using the equation below:

V2

AP= K p (Eq. 2.13)

where AP = differential pressure across the disc, Ibs/ft2

K = flow resistance coefficient at a given disc opening

p = mass density of the fluid, slug/ft3

V = flow velocity, ft/sec

- 16 -
The fluid flow velocity in a piping system depends on the pump flow characteristics and the

piping system resistance. As shown in Figure 2.13, the discharge pressure of a typical

centrifugal pump decreases as the pump flow rate increases. On the other hand, pressure

drop across the piping system increases as the flow rate increases. The intersection point

of the two curves satisfies both the pump characteristics and the piping system flow

resistance, and is the solution point for a given disc position.

AVAILABLE/

0

U,

C:

LIU

cc

CL

IL /TOTAL PIPING SYSTEM

PRESSURE DROP

GA10M

FLOW RATE, 0

Figure 2.13

Balance of Available Pump Head and Piping System Pressure Drop

Using the above described approach, a typical change in the pressure drop across the disc as

a function of disc opening in a pump flow system is shown in Figure 2.14. This figure is

based upon test data for the 4-inch Borg-Warner valve tested at the Duke Power flow loop

(see Footnote 1 on page 15). The pressure drop across the disc decreases rapidly as the valve

opens, and after 30 percent of disc opening, the differential pressure drop across the disc is

very low. This shows that the significant disc load in pumped flow systems is encountered

during the final 30 percent of disc closing. The actual magnitude of the pressure drop

across a disc will vary depending upon individual pump characteristics and system flow

resistance.

- 17 -
Ca) 1400 - (-

a) 1300 -

(> 1200

up 1100

2 1000 I

C 900

CL

o 800

700

n 600

Cn 500i

X. 400 \

300

200 -

100_

0- ,

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0

Gate Position, h/Do

Figure 2.14

Typical Pressure Drop Across a Gate Valve as a F unction of Gate Position

In High Pressure Pumped Flow System

A comparison of AP versus the disc opening curve from the pump system in Figure 2.14 and

blowdown tests performed by INEL L23,24.1 shows that the differential pressure across the

disc, which directly affects the load on the disc, is significantly higher under blowdown

conditions. For example, at 25 percent gate opening, the differential pressures ranged

from 50 percent to 90 percent of the fully closed AP in various INEL tests as compared to

approximately 5 percent predicted in the pumped flow results shown in Figure 2.14.

Therefore, localized contact stresses and the propensity of galling damage to the valve

internals is also much higher for valves subjected to blowdown conditions.

- 18 -
In order to make a preliminary estimate of the flow-induced load on the disc in midtravel, it is assumed that the load on the disc is proportional to the differential pressure

across the disc and the percentage of disc opening, as shown below:

F =AP (100 - % of disc opening) x full disc area (Eq. 2.14) K 100

where Fp = pressure load on disc in mid-travel position

AP = differential pressure across the disc

Disc Area = t (effective sealing dia)2

4

The pressure loads across the disc calculated from this equation were used in the

evaluation of localized contact stresses at the disc guide interface and disc-to-seat

interface as detailed in Appendix C and discussed in Subsection 2.6.3.

2.6.2. Disc Tilting and Its Interaction wvith Seats and Guides

Figure 2.15 is a scaled drawing of a disc tilted in the mid-travel position due to forces

imposed on it by the fluid flow. This figure was developed from actual design dimensions

of a 4-inch ANSI Class 900 Borg-Warner flex wedge gate valve used in the Duke Power

tests [151. The geometrical interaction between the disc, seats, and guides was investigated

in detail at several disc openings, and with the extreme combinations of tolerances of these

components. In this figure, the disc is shown at an opening of approximately 25 percent.

The disc guides in this valve design do not limit the disc tilting under fluid forces, thus

allowing point contact to occur between the disc and the downstream seat face. High local

stresses are developed at these points when disc tilting results in downstream seat contact.

Also, the relative magnitude of these stresses is significantly higher than those

encountered in a line contact that occurs when disc tilting is constrained by the guides.

The actual magnitude of the contact stresses calculated for the 4-inch Borg-Warner gate

valve are discussed in Subsection 2.6.3.

- 19 -
Figure 2.15*

Point Contact Against the Downstream Seat Due to Disc Tilting at a

Typical Mid-Travel Position in 4, 1500# Borg-Warner Gate Valve

* See Footnote 1 on page 15. This figure is nonproprietary and is used by permission from Duke

Power Company.

-20 -
In general, the equilibrium position of the disc is defined by the resultant load vectors of the

pressure load, stem thrust, seat contact force, and guide force as well as the geometry of the

disc, guide, and seat area. The following three cases show the extreme variations that can

occur between the disc guide and seat interaction:

1. Ideal disc slide

GA 106

FB .4- B

Fp

A 4 Gate Slot

As shown in Figure 2. 16, if the disc guide design is

such that the resultant pressure load acts within the

two extremes of the disc guide slots (Points A and

B), full contact at the guide surface is achieved

without tilting. This disc orientation results in the

lowest contact stresses due to disc loads imposed in

mid-travel. Sliding under a full-surface contact is

the ideal disc guide design condition to withstand

mid-travel pressure loads.

FAI |

~ NGate Guide

Figure 2.16

Ideal Gate Slide

2. Tilted disc contacting guides

If the resultant pressure load acts below the lower

GA 107 end of the guide contact point A, the disc will tilt, as

shown in Figure 2.17. Depending on the guide

clearance and other dimensions of the disc, guides,

and seats, the disc may resist the resultant load by

contacting the guides, thus preventing any disc-toseat contact in mid-travel. From disc force equilibrium considerations along the flow axis, one

it can see that the contact load at Point A in this case

will be at least as high as the pressure load Fp, and

e it may be higher depending upon the actual location

of the resultant load vector Fp, below the guide. The

Guides location of Fp with respect to the disc will vary as

the disc travels towards the closing position.

Fp_ -j

_ -'Gate SIo

Gate Guide

Figure 2.17

Tilted Gate Contacting (

-21 -
3. Tilted disc contacting seats

In a valve design with large guide clearances

and/or short guide length, disc tilting under the

pressure load can result in point contact with

the upstream and downstream seats as shown

in Figure 2.18. As in the previous case of the

disc contacting the guides, if the pressure load

resultant acts below Point A, the contact load at

A will be greater than the pressure load Fp. The

actual load magnitude can be determined by

force equilibrium along the flow axis. Contact

stresses at Point A in this case can be very high

because of point contact between two curved

surfaces, as discussed in Section 2.6.3 and

Appendix C.

GA 108

Figure 2.18

Tilted Gate Contacting Scat

The three extreme cases discussed above show that the pressure load magnitude, location,

and dimensions of the disc guide and seats are essential in determining the actual

configuration acquired by the disc in mid-travel. Reference 47 presents a mathematical

model to facilitate the evaluation of disc-seat interference for a gate valve during valve

closure. Better quantification of the pressure-induced load is needed to more accurately

determine the disc equilibrium and quantitatively assess the adequacy of a valve to operate

properly and without causing damage to valve internals in mid-travel position.

It should be noted that the disc equilibrium discussion in the above three cases was limited

to pressure load acting along the flow axis only. Disc equilibrium along the stem axis can

also be affected by the resultant pressure load and downstream seat friction load acting in

the stem axis direction. The overall disc equilibrium equations can be further refined by

including these load terms.

2.6.3. Contact Stresses

Gate tilting in mid-travel position can create high localized contact stresses in the disc,

seat, and guide interfaces. The magnitude of contact stress depends on the geometrical

shapes of the two contacting surfaces and their material properties. For simple, welldefined geometries such as spherical, elliptical, cylindrical, and plane surfaces, the

contact stresses between the two surfaces can be calculated using the Hertzian general

linear elastic solution for doubly curved surfaces as shown in Appendix C. For more

complicated contact surfaces and loading situations, computer-aided numerical methods

- 22 -
such as finite element analysis can be used. Finite element analysis can also provide a

more accurate assessment of the localized contact stresses by avoiding small displacement

assumptions and accounting for elastic-plastic behavior.

In Appendix C, local contact stresses for the 4-inch Borg-Warner valve under pump flow

conditions were investigated by using the closed-form Hertz contact stress equation for

point and line contacts (the performance of this valve in Duke Power tests is discussed in

Section 3.4). Localized contact stresses were calculated for two cases: (1) when the disc

tilting causes a contact against the downstream seat, and (2) the case in which disc tilting

causes a line contact at the guide surfaces. The results show that, based on linear elastic

assumptions, the calculated contact stresses vary widely from near the material threshold

of galling stress to an order of magnitude higher than that. As expected, highest localized

stresses were found in the case of disc tilting which results in point contact against the

downstream seat. With an assumption of 0.125-inch edge radii at both the disc outside

diameter and seat inside diameter contacting surfaces, theoretical stresses based on the

linear elastic assumption exceeded 300 ksi at the point contact. Local yielding, load

redistribution, and localized material wear under repeated cycling are likely to occur,

which will flatten or enlarge the contact area and reduce the contact stresses. However,

these high stresses, which are significantly above the thresholds of galling for the sliding

materials, will initially result in local galling. Depending upon the magnitude of the

loads present and the local geometry, a stable frictional behavior without further galling

can be attained after a progressive increase in the local contact area due to yielding as well

as material removal by wear or galling occurs. On the other hand, if the loads are too high

to be supported by area spreading achieved by local wear, continued galling and increase

in frictional forces can be expected. Extensive testing is needed to quantify this

mechanism for various load geometries and loading conditions.

It should be pointed out that the theoretical local contact stress based on simplifying

assumptions cannot be used by itself to predict the valve performance. It can be used more

as an index of contact stress severity, which can be related to actual performance, and to

make relative comparisons to compare different local geometries and load magnitude

The contact stress analysis clearly shows that significant improvements can be made in

the detailed design of valve components in possible contact areas by increasing local radii

and resisting disc tilting by line contact instead of point contact, as shown in Appendix C.

The ultimate goal of an improved gate valve design should be to incorporate a guide

geometry that results in ideal disc slide (discussed earlier in Subsection 2.6.2), which

eliminates high local contact stresses.

- 23 -
3. COEFFICIENT OF FRIICTION AND THRESHOLD OF GALLING STRESS

Coefficient of friction between the disc and seats is the dominant factor in determining the

operating thrust requirements for most gate valve applications. Valve manufacturers

have standardized on Stellite hardfacing alloys for the disc and seat sliding surfaces

because of their excellent resistance to corrosion, wear, and galling in the unlubricated

state, even at elevated temperatures. In spite of the fact that Stellite has been in widespread

use in gate valve applications for several decades, published data for Stellite vs. Stellite

coefficient of friction show a significant lack of uniformity under seemingly similar test

conditions. Reported data span, even under laboratory conditions, a wide range from 0.12

[441 to 0.48 1401, and sometimes even higher.

It should be pointed out that, even though several variables can affect the coefficient of

friction results [29,42,431, the principal investigators have found that the most important

factors responsible for the wide scatter in the reported data are (1) the differences between

the size and geometry of the test specimen, (2) the presence, absence, or gradual removal of

an absorbed layer of lubricant at the sliding surfaces, and (3) presence or absence of

galling of various levels due to high localized contact stresses in some areas of contact.

These factors should be kept in mind while reviewing test data and results reported by

various sources.

During the 1970's, the principal investigators were involved in a gate valve development

effort under which extensive testing and evaluation of coefficient of friction and galling

data was done. A summary of the important results from these tests is presented in Section

3.2. This is followed by the recent results from the principal investigators' involvement in

assisting Duke Power Company to perform a root cause analysis of the 4-inch BorgWarner flexible wedge gate valves that failed to close under high differential pressure

conditions in a pumped flow system 1151. Other recent test data for parallel slide gate

valves reported by KWU-Siemens 126] and British National Power Division of the CEGB

[211 are discussed next. Finally, the important results from the recent NRC-sponsored

INEL blowdown tests 14,233 are also presented here for comparison, and for drawing

overall conclusions from the presently available data.

Before proceeding with a discussion of the results from these sources, it is important to show

the relationship between the commonly used term disc factor, sometimes called valve

factor, to the coefficient of friction. This is presented in the next section.

3.1. Relationship Between Disc Factor and Coefficient of Friction

The common industry equation for determining valve thrust requirements for actuator

sizing is:

-25 -
Minimum stem thrust requirement = disc factor x differential pressure

load across the disc

+ stem packing friction force

± stem rejection force

where the positive and negative signs apply to the closing and opening thrust requirements

respectively. A disc factor of 0.3 has been commonly used in actuator sizing in the past.

The disc factor used in the above equation exactly equals the coefficient of friction between

the gate and seat for a parallel slide gate valve.

For a conventional wedge gate value of solid, flexible, or split disc design, the disc factor is

not exactly the same as the coefficient of friction. The relationship between the two, as

derived in Section 2.1, is shown below:

Disc Factor = 1

cos 0 ± + sin 0

where 1t = coefficient of friction between disc and seat

0 = one-half of total included wedge angle

+ sign in the denominator applies to valve opening, and

- sign in the denominator applies to valve closing

The difference between disc factor and coefficient of friction for conventional wedge gate

valves is usually small. Typical wedge gate valves use a wedge angle of (0 = 5°), for which

the ratio between disc factor and coefficient of friction over the typical range of coefficients

of friction computed for the above equation is within ± 5 percent as shown in Table 3.1. For

all practical purposes, the difference between the two is much smaller than variations in

the coefficient of friction data, and can often be ignored wvithout much impact on the overall

conclusions. Some valves use higher wedge angles, for which the differences become more

significant.

Ratio = disc factor

Strohe Direction coefficient of friction

=0.2 1 _=0.3 _J 1 =0.4 1 11=0.5

Open 0.99 0.98 0.97 0.96

Closing 1.02 1.03 1.04 1.05

Table 3.1

Relationship Between Disc Factor and Coefficient of Friction I

for a 100 Wedge Gate l

- 26 -
3.2. Data for Coefficient of Friction from Principal Investigators' Experience

3.2.1. Background

During the 1970's, the principal investigators worked in the research and development

department of a major U.S. valve manufacturer. This valve manufacturer had been a

dominant supplier of valves for oil field, pipeline, and petrochemical applications for

several decades. In the early 1970's, this manufacturer decided to pursue the nuclear

power, geothermal, coal gasification, and synthetic natural gas markets. In order to meet

the technical challenges posed by the development of valves for these applications, major

additions to the test facilities were made. Significant additions pertinent to the nuclear

power valve development effort included a 1,500 psi air/water high energy blowdown

system; a 600'F, 12,500 lbs/hr steam generator; an external pipe load simulator capable of

applying bending moments of up to 2 x 106 ft-lbs; and a Falex friction/wear test machine.

Additionally, the laboratory was equipped with a 100-channel strain gage data acquisition

system, various load and torque cells, pressure transducers, and a tension/compression

test machine. This environment provided an excellent opportunity to the principal

investigators to be involved in a systematic research and development effort on gate valves

for high temperature applications.

A common test frequently performed on a valve under development consisted of several

hundred cycles (1,000 cycles was the usual goal) of opening and closing under maximum

design differential pressure generated by small positive displacement pumps using room

temperature city water as the flow medium. Stem thrust measurements using strain gage

load cells were performed on several gate valves ranging in size from 2 inches to 16 inches

during this developmental testing effort. Testing was also done on valves for high

temperature service using saturated steam with pressures up to 1,500 psi and temperatures

up to 6001F. Most of the test data were primarily used to support the in-house development of

the new valve designs and were considered proprietary at that time; therefore, no data were

published in the open technical literature. With the virtual disappearance of the nuclear

power market in the early 1980's, this valve manufacturer stopped its nuclear valve production. Valuable technical data that had been developed were never published. The

principal investigators have taken this opportunity to present highlights of the most

significant results from these tests relating to the coefficient of friction, which are

summarized below:

3.2.2. Results from Room Temperature Water Tests

Using room temperature, ordinary tap water as the test medium, the typical range of

coefficient of friction for Stellite-6 overlaid gate and seats from a large number of tests on

several different sizes of valves up to 16 inches was found to be between 0.15 and 0.25 over

several hundred cycles of operation. These tests were performed on parallel expanding,

through-conduit gate valves of the type shown in Figure 2.1b. As discussed below, two

important features of this design are that the gate cannot physically tilt, and the contact

stresses at the seat faces are well below the threshold of galling.

- 27 -
As can be seen in Figure 2.1b, the gate length in this type of through-conduit valve design is

approximately twice as long as the ones used in the conventional solid wedge or flex wedge

gate valves. The lower (extended) part of the gate has a bore through it, which lines up with

the seat bore when the valve is in the fully open position. Under the action of fluid flow

forces on the gate in the mid-travel position, this type of gate assembly is simply pushed

down against the downstream seat face, thus providing a surface contact instead of the

point contact that can occur in conventional wedge gates which have excessive guide

clearance as discussed in Section 2.6. Thus, the through-conduit design of Figure 2.1b

avoids high localized contact stresses at the gate-to-seat faces which have the potential to

cause galling. Furthermore, the design of these valves was based on limiting the average

seat face contact stress to 10,000 psi or less under maximum differential pressure. This is

well below the threshold of galling as discussed in Section 3.3.1. The absence of galling at

the seat faces was confirmed by many tests under high differential pressure conditions. In

these cycle tests, no extraordinary effort was spent to remove any residual lubricants used

during assembly, except what is automatically removed by normal abrasive wear due to

sliding under high contact pressure. The coefficient of friction was calculated using

mean seat diameter as the effective seating diameter.

In summary, coefficients of friction values for Stellite vs. Stellite were typically found to

range between 0.15 and 0.25 for several hundred cycles of testing with gate valves using

room temperature ordinary tap water, with average seat contact stress of 10,000 psi or less,

and a gate design which prevented any galling at the seating surfaces. It is important to

note that the valves were not exposed to high temperatures prior to these tests. As discussed

later, this has been reported to be a factor that can cause an increase in the coefficient of

friction (Sections 3.5 and 3.7).

3.2.3. Results from High Temperature Water and Steam Tests

An extensive series of tests using saturated steam and hot water in temperatures up to 600'F

were performed on the same type of parallel expanding, through-conduit gate valves

(Figure 2.1b) as used in the cold water tests discussed in Section 3.2.2. Valves up to 12

inches in size were tested with a maximum differential pressure of 1,500 psi. As

mentioned in the previous section, this type of gate design prevents mid-travel gate tilting

and maintains a surface area contact against the downstream seat. The total number of

cycles under various levels of differential pressures during these steam tests ranged from

10 to 30 ; and not hundreds of cycles as in the cases of cold water tests. The highest value for

the coefficient of friction between Stellite and Stellite during any of these tests was found to

be 0.39 using water or steam at 600'F. This coefficient of friction evaluation was based on

using the mean seat diameter as the effective seating diameter The calculated average

seat contact stress in these test valves was less than 10,000 psi under the maximum

differential pressure conditions, and no evidence of galling was found at the seat faces.

- 28 -
Based on these results performed on actual gate valves (not friction test specimens), the

principal investigators have found the value of 0.4 for coefficient of friction for pure sliding between Stellite seat and Stellite disc faces for high temperature steam and water

applications to be a reliable result provided it is ensured that the seat faces are free of

galling.

Conversely, the principal investigators have also used this data successfully in root cause

analyses of valves to identify potential galling, component interference, or other problems

with the valve internals when the coefficient of friction, based on measured thrust, significantly exceeds 0.4 in steam or high temperature water application. The above results for

the coefficient of friction are in agreement with the results reported in Sections 3.5 and 3.6

by others using actual gate valves of improved designs that are free of galling damage.

3.2.4. Long-Term Surveillance Tests on SIS Valves Under Flow and AP

In 1981, the principal investigators were involved in the root cause analysis investigation

and modification of two safety injection system (SIS) gate valves at a PWR plant following

their failure to open when challenged1 [51]. The problem was attributed to an increase in

the required thrust to open the valves due to galling of the seat faces (see Footnote 1). After

the modifications proposed by the principal investigators and the utility were implemented

and demonstrated to be successful, NRC required a periodic surveillance testing of these

valves under differential pressure and flow. The objectives of these NRC-imposed longterm surveillance tests were to ensure that (1) the root cause of the failures is indeed

understood and has been corrected, and (2) there is sufficient margin in the actuator force

to account for degradation of the valve internals and possible increase in friction due to

long-term set effect caused by constantly applied differential pressure across the disc.

A total of six dynamic tests were performed under hot standby conditions (plant operating

Mode 3) with temperatures up to 330'F on each of the two valves between November 1981 and

August 1985. During these surveillance tests, both the differential pressure across the

valves and the actuator force required to open the valves were measured. The valves were

operated by hydraulic actuators, and the actuator force was calculated from the pressure

measured on both sides of the piston, plus the stem rejection force due to pressure inside the

valve. All of the pressure measurements were done using calibrated pressure transducers

in accordance with controlled test procedures. In 1986, the principal investigators

performed a detailed evaluation of these surveillance test results2 . It was concluded that

both of the SIS valves had continued to perform consistently and successfully with

sufficient margin below the maximum capabilities of the actuator.

1M. S. Kalsi. Independent Review of Operability Failure Problems with the Safety Injection System

Valves HV-851 A and B at SONGS 1, Kalsi Engineering, Inc. proprietary report to Southern

California Edison, KEI 3.2.0, September 1981.

2 M. S. Kalsi and J. K. Wang. Independent Evaluation of Safety Injection System Valve Surveillance

Test Results and the Proposed Periodic Testing at Songs 1, Kalsi Engineering proprietary report to

Southern California Edison Company, KEI-919, May 1986.

-29 -
One of the important conclusions drawn from these 12 tests performed on the two valves

over a span of approximately four years is that the coefficient of friction between the Stellite

disc and seats ranged from 0.17 to 0.34, including the long-term set and other degradation

effects. Stem packing friction was not subtracted from the total opening thrust to obtain

conservative estimates for the coefficient of friction. As stated earlier, these tests were

performed under dynamic flow and differential pressure with temperatures up to 3301F

during Mode 3 hot standby conditions. It should also be pointed out that the average seat

contact stress for these modified valves under the maximum differential pressure was

around 7,500 psi, well below the threshold of galling for Stellite against Stellite (as

discussed in Section 3.3); and thrust measurements did not indicate a trend of continuing

increase in friction.

3.3. Contact Stress and Threshold of Galling

As the contact stress between the sliding surfaces is increased, a threshold is reached

beyond which the required sliding force between the mating materials increases rapidly

due to significant material transfer through localized welding, tearing, and digging of the

surfaces. The contact stress at which this behavior is initiated is called threshold of

galling stress. Unlike normal wear, the damage to the materials due to galling goes well

beyond the surface in just a few strokes. Under galling conditions, the coefficient of

friction is unpredictable because sliding between the two surfaces involves significant

shearing and tearing of the cold-welded junctions formed between the mating materials.

3.3.1. Threshold of Galling for Stellite vs. Stellite and

Other Valve Trim Materials

The threshold of galling stress for Stellite against Stellite in the unlubricated condition is

reported by some investigators to be over 50 ksi [42, 45, 46], and by others to be over 70 ksi [48,

49] based on the maximum limits of their respective test apparatuses. This is based on a

single forward rotation of 360 degrees using block and button test specimens in which the

load is gradually increased until first signs of galling appear. Using slightly larger

specimens (0.5 inch diameter instead of 0.375 inch), and using one forward rotation of 360

degrees, one reverse rotation of 360 degrees, and another forward rotation of 360 degrees,

Schumacher, the author of Reference [46], has found the threshold of galling stress for

Stellite to be 47 ksi. This situation more nearly duplicates a valve cycling application

rather than an unidirectional 360-degree test. In the experience of the principal investigators, the threshold of galling stress is somewhat subjective, and a reasonable margin

should be provided against threshold values to achieve satisfactory performance. Stellite

vs. Stellite was still found to rank in the category of materials having the best resistance to

galling. Comparatively, the galling resistance of stainless steels and carbon steels without any hard surface treatment is much lower, as shown in Tables 3.2 and 3.3 reproduced

from Reference 42.

-30 -
TABLE 4.16 Galling Resistance of Stainless Steels

Button material

630

Condition and (17-4 Nitronic Nitronic

Block material nominal hardness (BHN) 410 416 430 440C 303 304 316 PH) 32 60

Type 410 Hardened and stress relieved (352) 21 28 21 21 28 14 14 21 320 350+

Type 416 Hardened and strcss relieved (342) 28 90 21 145 60 165 290 14 310 350+

Type 430 Anncaled (159) 21 21 14 14 14 14 14 21 21 250

Type 440C Hardened and stress relieved (560) 21 145 14 75 35 21 250 21 350+ 350+

Type 303 Annealed (153) 28 6( 14 35 14 14 21 21 350+ 350+

Type 304 Annealed (140) 14 165 14 21 14 14 14 14 210 350+

Type 316 Annealed (150) 14 290 14 255 21 14 14 14 21 260

Type 630 (17-4 P11) 11 950 (415) 21 14 21 21 14 14 14 14 350+ 350+

Nitronic 32 Annealed (235) 315 310 55 350+ 350+ 210 21 350+ 210 350+

Nitronic 60 Annealed (205) 350+ 350+ 250 350+ 350+ 350+ 260 350+ 350+ 350+

Note: Values shown atc threshold galling stress (NIN)a condition and hardncss appli to both the button and the

block matcrial: tests wcrc terminated at 350 MNa, so valucs given as 350+ indicate the samples did not gall.

Source: Adaptcd from Anonymous (1978d).

Table 3.2

Galling Resistance of Stainless Steels (1 MPa = 145.14 psi)

Handbook of Tribology; Bharat Bhushan and B. K. Gupta, 1991, McGraw-Hill, Inc.

This material is reproduced with permission.
TABLE 4.17 GCiline Resistance of Alloys

Allvs .n contact' Threshold galling stress.t lPa

Silicon bronze 1:101 vs. silicon bronze (200) 28

Silicon hronze (120t1 vs. AISI (1)4 1140) 3N

AISI ht) tA.SN 12701 vs. A254 127U) 21

AISI 4337j S4) vs. AISI .1337 i415) 14

AISI 10134 115) vs. AISI 1034 (415) 14

Waukesha II-t41) vs. AISI 203 (ISO) 35(1-

Wauksha S.S 1141) vs. AISI :1)11202) 350+

WVaukcsha, SS 1411 vs. AISI 316 t200) 350.

W.aukesha i8 I 1411 vs. AISI 6UO (405) 350-

W.iukcsha S (141) vs. 2UCr-gONi (ISO) 350-

A;SI 111 12U I vs. AISI 011 1:(12) 1(15

A\iSI 01 i2 12) vs. AISI 2014 t411) 14

\ISI 21)1 !; I vs. AISI 30 117.-4 PH) (382) 14

.\ASI 011 202) vs. Nironic 32 '231) 250

AISI 011 II)9) vs. AISI 416 i3421 21

AISI 301I (169) vs. AISI 440C 1560) 21

AISI 41011322) vs. AISI 420 14 2) 21

AISI 416 13421 vs. AISI 416 1372) 90

AISI 416 1372) vs. AISI 410 t322) 28

AISI 416 i42) vs. AISI 43O (10) 21

AISI 416 t342) vs. 2()Cr-SONi (ISO) so

AISI 44flC t56(11 vs. AISI 440C 1604) SO

A1SI 6311 11-4 PHI t3111 vs. AISI 304 (140) 14

AISI t)3 i 1 -4 PHI 13S1O1 vs. Nirronic 32 (401) 21

-\ISI 4311 4351 vs. AISI 011411411 14

.USI 43(1 t410t) vs. AISI 31 .-(1() 21

AISI 6;3()14351 vs. AISI 631 1435) 14

.Na4tronic 32 i235) vs. AISI 30 ;13S01 75

Nitronic 32 4t)1) vs. Nitronic 32 1401) 235

Nitronic 32 (2i5) vs. Nitronic ;2 (401) 235

Nitrnic 32( 2235) vs. AISI 3(4 (140) 50

Nitronic 32 (401) vs. AISI 304 (140) 90

Nitronic 32 (205) vs. AISI 1034 (205) 14

Nitronic -O (2105 vs. Nitronic 50 (205) 14

Nitronic 51(321) vs. Nitronic 50 (321) 14

Nitronic 50 1205) vs. Nitronic 32 (401) 90

Nitronic '0 1321) vs. Nitronic 32 (235) 55

Nitronic 50 1:U5) vs. AISI 304 (140) 2S

Nitronic 60 1205i vs. AISI -01 (169) 350+

Nitronic 60 (205) vs. AISI 420 (472) 350-

Nitronic 60 1213) vs. AISI 630 (313) 350+

Nitronic 60 (205) vs. AISI o0i (332) 350+

Nicrrcnic 60 (205) vs. Nitronic 50 (205) 350+

Nitronic 60 (205) vs. AISI 4337 (44S) 350+

Nitronic 60 (205) vs. AISI 660 (A2S6) (270) 350+

Nitronic 60 (205) vs. 20Cr-SONi (180) 250

Nitronic 60 (2(15) vs. Ti-6AI-4V (332) 350+

Nitronic 60 (205) vs. SieIlite 6B (415) 350+

Sccllite 6B (415) vs. AISI 304 (140) 240

Stellite 6B (415) vs. AISI 316 (140) 25

Stc!litc 6B (415) vs. Stcllite 6B (415) 350+

Stcllitc 6B (415) vs. Tribaloy 400 (54 HRC) 350-

Stellitc 6B (415) vs. Tribaloy 700 (47 HRC) 350

Tribaloy 401 (54 HRC) vs. Tribaloy 400 (54 HRC) 350+

Tribaloy 700 (47 HRC) vs. Tribaloy 700 (47 HRC) 185

'Numbers in parentheses following alloy designations are nominal hardness (Brinell).

tValucs given as 35(1- indicate the samples did not gall. Source: Adapted from Anonrvmous (197Sd) and Foroulis (1984).

Table 3.3.

Galling Resistance of Alloys (1 MPa = 145.14 psi)

Handbook of Tribology; Bharat Bhushan and B. K Gupta, 1991, McGraw-Hill, Inc.

This material is reproduced with permission

-32 -
3.3.2. Average and Local Contact Stresses

In the design of valve seating surface, the average contact stress based upon maximum

differential pressure and full face contact should be kept well below the threshold of galling

stress. A margin is necessary in practice to allow for the higher localized stresses that are

caused by elastic displacement of the disc, body, and seats, resulting in a non-uniform

distribution. Figure 3.1a qualitatively shows that the highest localized contact stresses

occur in the gate valve seat faces around the 3 o'clock and 9 o'clock positions circumferentially, and near the seat inside diameter radially. It should be pointed out that this

non-uniform distribution is not the result of disc tipping as discussed in Section 2.6, but is

due to the uneven stiffness of a gate valve body and flexure of the gate. In Section 2.6 it is

shown how disc tilting can also result in high localized stresses in the guide areas or at the

downstream point contact. If the localized stresses exceed threshold of galling, galling in

local areas is initiated. In our review of differences in various manufacturers' valve

designs and their performance, we have found that the average seat face contact stress of 20

ksi should not be exceeded, and 15 ksi or less is preferred in order to achieve repeatable

performance using Stellite hardfacing.

- ~~~~~~~~ORIGINAL 0\O@

= ;~~~~~~~~A DIS DISC>

UNDER

Figure 3.1a

Radial Seat Contact Stress Variation

-33 -
,STEM AXIS CIRCUMFERENTIAL PEAKS

SAVE

B

h30o

r 7 \ LOCATION: A B C D A / 6 NS EAT WIDTH

0 = 0 \ 90' 180° 270' 360'

SEAT FACE SEAT CONTACT STRESS

Figure 3.1b

Circumferential Seat Contact Stress Variation

Whether this local galling, once initiated, progresses continuously until seizure or heals

itself by spreading the load over a larger area during successive cycling depends upon the

load magnitude and the local geometry of the two contact solids. Actual testing of fullscale or near full-scale test specimens under realistic simulation of actual loading

conditions has been found by the principal investigators to be the most reliable way to

determine the limits of operation without galling.

When analyzing the test results from any valve to determine the applicable coefficient of

friction, it is extremely important to establish that the surfaces are free of galling damage.

Otherwise, "apparent" coefficient of friction values much higher than those obtained in

pure sliding behavior can be erroneously concluded from the test results and applied to

other valves that are of different, healthy designs.

3.4. Duke Power Data for 4-Inch Borg-Warner Flexible Wedge Gate Valve

Following the failure of a 4-inch Borg-Warner flexible wedge gate valve to fully close

under a differential pressure of 1,800 psi at Catawba Nuclear Power Plant, Unit 2 in 1988

[14], Duke Power Company undertook a systematic root cause analysis investigation of the

failure. The valve that failed was a 4-inch, ANSI 1500, flexible wedge, carbon steel gate

with a U-shaped guide fitted into a milled slot in the body at the bottom and pinned to the

bonnet at the top. Duke Power performed extensive differential pressure tests at their

Riverbend Steam Station flow loop on another Borg-Warner carbon steel valve identical in

design to the one that failed at Catawba. Tests were also performed on a stainless steel

valve of the same size and design. The seat and disc faces were overlaid with Stellite in

all of these valves. Stem force data were obtained in these tests using stem strain gages.

- 34 -
Kalsi Engineering, Inc. assisted Duke Power in the root cause evaluation of these valves

which failed to perform under a manufacturer-specified valve factor of 0.3. With Duke

Power's permission, the important results from this investigation (see Footnote 1 on page

15) are presented here.

The coefficient of friction extracted from the carbon steel valve tests is presented in Figure

3.2 for 24 consecutive cycles, starting with a newly refurbished valve. During cycle

testing, the differential pressure was varied between nominal values of 500, 1,000, 1,500,

and 2,000 psi. These tests spanned a period of four to five days during November 1988.

' Opening Cycles

* Closing Cycles

0.48

0

UI.

0

4-,

C:

01)

0

C0.44

0.40

0.36

0.32

0.28

0.24

0.20

0 2 4 6 8 10 12 14 16 18 20 22 24 26

Test Cycle Number

Figure 3.2*

Increasing Friction Trend During 4-inch Borg-Warner

Gate Valve Cycle Testing by Duke Power, Reference

* See Footnote 1 on page 15. This figure is nonproprietary and is used by permission from Duke

Power Company.

-35 -
The plotted results shown are obvious overall central tendency of gradual increase in the

coefficient of friction, along with the expected variation around the mean. Coefficient of

friction magnitude increased from 0.36 to 0.49 in the closing direction, and 0.24 to 0.47 in

the opening direction during cycling. As discussed in Section 2.6, the disc guide design of

this valve has large clearances, and it allows the disc to make point contact against the

downstream seat face. Inspection of the disc clearly shows evidence of contact against the

downstream seat at 4 o'clock and 8 o'clock positions, and two localized areas of relatively

minor wear that have spread to a width of about 1/16-inch at the outside edge of the disc. Itis

believed that the progressive wear of this area, even though minor, contributed to the

removal of the contaminant layer and a gradual increase in friction aggravated by local

galling.

It can be conjectured wvhether or not this trend of increasing friction would have continued

until seizure occurs or stabilized at a certain value if the cycling had been continued.

Based on the magnitude of loads involved in this valve application under pump flow

conditions, minor localized wear and the spreading of the load bearing area results in a

substantial decrease in localized contact stress to below the threshold of galling stress.

Once the localized areas have spread to an equilibrium condition below the threshold of

galling stress, stable performance at some lower coefficient of friction value than that

encountered at the end of this test can be expected from this valve under continued cycling

under the same differential pressure and flow condition.

3.4.1. Summay and Comparison of Duke Power Data for

Carbon Steel vs. Stainless Steel Valves

The average, minimum, and maximum values for the coefficient of friction for the first

ten cycles for this carbon steel valve test (results plotted in Figure 3.2) are summarized in

Table 3.4. It should be noted that two organizations that have performed extensive

blowdown tests on isolation valves have typically specified five test cycles for their

operability qualification 126,391.

-36 -
Test Valve Coefficient of Friction

Disc and Seat Material Pe Average MinimumlMaximum

Carbon Steel with Stellite Closing 0.401 0.362/0.426

Opening 0.302 0.239/0.338

Stainless Steel with Stellite Closing 0.288 0.19/0.349

Opening 0.256 0.12/0.348

Table 3.4*

Comparison of Coefficient of Friction Results for Carbon Steel and

Stainless Steel Valves for First Ten Cycles

The results from another 4-inch Borg-Warner flexible wedge valve of the same design, but

of stainless steel material, which was tested by Duke Power in their Riverbend flow loop,

are also shown in this table for comparison. Both of these valves had Stellite hardfacing

overlay on the disc and seat faces. The comparison shows that the results for the stainless

steel valve are significantly lower than the carbon steel. Duke Power suspected that the

differences in material of the overlay, due to the differences in the iron content, may be

responsible for the differences in their coefficient of friction behavior. This seems

plausible, especially since differences in the composition of various Stellite alloys (e.g.,

Stellite-1, -6, and -12) exhibit different coefficients of friction as reflected by Foroulis t40]

and Rockwell Edwards [27].

Foroulis [40] has reported that the coefficient of friction under relatively light contact stress

(approximately 50 psi) in clean water for Stellite-1, Stellite-6, and Stellite-12 materials in

self-mated tests were 0.28, 0.48, and 0.24 respectively. In non-self-mated tests, when

Stellite-6 was tested against Stellite-1 or Stellite-12, the coefficient of friction was 0.28.

This is significantly lower than for the self-mated case of 0.48.

Rockwell Edwards has also reported significant differences in the coefficient of friction

values for Stellite during their Equiwedge gate valve development program [27]. They

selected Stellite-21 based on its overall performance, including lower coefficient of

friction. However, they did not report actual values from their tests.

Based on the above comparisons, it is clear that further testing under controlled conditions,

using an appropriate range of contact stresses, is needed to evaluate the effect of iron

content and possibly other alloying elements in the cobalt-based Stellite alloys.

See Footnote 1 on page 15. This table is nonproprietary and is used by permission from Duke Power

Company.

-37 -
3.5. KWU-Siemens Test Data

KWU-Siemens has recently reported results of their high pressure blowvdown testing on a 6-

inch parallel slide gate valve under cold water, hot water, and steam conditions. Friction

coefficients were calculated for consecutive cycles and tabulated as shown in the following

table [261.

Max Coefficient of

Friction during

Avg Coefficient of

Medium Stale Friction during

Opening (5,)1 Closing (5)

_ _

Opening Closing

Water t = 860F

Po = 123 bar (1,784 psi)

112 bar (1,624 psi)

0.13 Rises con

0.15 tinuously with the

0.17 number of

0.18 tests

0.21

Water t = 290 0

C (556-F)

Po = 120 bar (1,740 psi)

112 bar (1,595 psi)

0.39

0.39

0.38

0.36

0.36

0.41

0.41

0.41

0.41

0.38 0.41

Steam Sat. Steam 0.38 0.41 0.36 0.39

Po = 100 bar (1,450 psi) 0.36 0.40

90 bar (1,305 psi) 0.38 0.38

0.33 0.38

0.33 0.38

Table 3.5

Friction Coefficients of a Parallel Disc Gate Valve Subjected to

High Pressure and High Flow Tests [261

The above frictional coefficients were conservatively estimated by KWU-Siemens using

the seat inside diameter for the pressure area calculation. The coefficients of friction

ranged from 0.33 to 0.41 after the initial cold water cycles. Initial low friction coefficients

for cold water cycles were attributed to the original good surface finish of the contact

surfaces according to the authors of the paper 1261. After testing with hot water and steam,

the coefficient of friction obtained from the subsequent cold water test remained high, and

-38 -
approximately in the same range as measured during hot water or steam tests. The same

phenomenon has been observed in the NRC-sponsored blowdown tests on 6-inch and 10-

inch valves by INEL [4,23,24]. The differences in the coefficients of friction from KWUSiemens tests for hot water and steam are insignificant, as reflected in Table 3.5. The

highest stem thrust occurred just after the flow isolation, as expected.

3.6. UK PWR Valve Testing

Results of high energy line break tests on parallel slide gate valves at 2,275 psi and 620'F

flow conditions have also been reported by National Power Division of the CEGB [21].

Conventional parallel slide gate valves of original design, which use a round disc,

suffered severe galling damage and required higher thrust to close than predicted by the

manufacturer [21,411. The problem was caused by tilting of the disc by fluid flow forces

which results in point contact against the downstream seat, much in the same fashion as

experienced in the conventional wedge gates tested under NRC-sponsored INEL tests.

Subsequently, tests were performed on an improved valve design, in which the lower part of

the disc is made rectangular to provide a line contact at the lower edge of the disc, thus

eliminating disc tilting and the potential of galling. (Independently, the same approach

had been recommended-by the principal investigators to Duke Power (see Footnote 1 on

page 15).) The new valves have been successfully tested under the U.K. PWR valve

qualification program with repeatable performance and no galling damage. The

coefficients of friction found during these tests with improved parallel slide gate valves

have been reported to be around 0.35 at the MOV User's Group meeting in Jupiter, Florida

in January 1991. These results are in general agreement with the KWU-Siemens results

summarized in Section 3.5 as well as the principal investigators' experience summarized

in Section 3.2.

3.7. NRC-Sponsored NESI Test Data

Results of NRC-sponsored testing by INEL on 6-inch and 10-inch valves under high

energy pipe break conditions are reported in detail in References 4 and 23. Two 6-inch

flexible wedge gate valve designs were tested under Phase I. Testing was extended to

Phase II, which included three 6-inch valves and three 10-inch flexible wedge gate designs

made by four different U.S. valve manufacturers. The overall conclusions from these

tests were that disc friction factors required to close the disc and achieve flow isolation

were higher than the 0.3 that had been used in the standard industry sizing equation used

by most valve manufacturers. Significant differences in the performance and in the

amount of damage to the valve internals were found, which were due to differences in the

specific design features used by the four manufacturers. Two of the valves, made by the

same manufacturer, were found to have the most severe galling damage to the disc and

-39 -
seat faces because of excessive disc guide clearance. These two valves were concluded to

have unpredictable behavior, and were not used by INEL in their coefficient of friction

evaluation.

INEL presented results from their detailed review of the data obtained from Phase I and

Phase II testing at the MOV User's Group meeting held in Jupiter, Florida in January 1991.

The results were presented in the form of a ratio of Normalized Sliding Load/Normalized

Normal Load, which is the same as the coefficient of sliding friction. The average

coefficient of friction was reported to be 0.4 for less than 70'F subcooling and 0.5 for water

that is subcooled by more than 70'F. However, usable data for the > 70° subcooled testing

were available for only two valves, whereas data for < 700F subcooling were from six

different valves.

This dependency of the coefficient of friction of Stellite vs. Stellite on the degree of

subcooling of the flow media has not been reported by others, and it requires confirmation

by additional testing. This testing should be done using valves that are clearly free of

problems that tend to affect the assessment of coefficient of friction at the disc-to-seat

interface, i.e., disc-to-seat galling, disc-to-guide galling, insufficient clearances between

the guides and disc, the inaccuracy of alignment between the seat wedge plane and the disc

wedge plane, etc.

As reported by KWU-Siemens (see Section 3.5), INEL testing also showed that the

coefficient of friction between seat faces tested with room temperature water shows an

irreversible increase after the valves are exposed to high temperature water or steam tests.

A detailed review of the NRC-sponsored Phase I and Phase II blowdown test results was

also performed by EPRI to determine the applicability and limitations of these results to

other operating conditions [24]. The disc factor during closing from these test results

under disc sliding conditions to achieve flow isolation was reported by EPRI to range from

0.28 to 0.48 for the different gate valve designs. For the opening direction, the coefficient of

friction was reported to range from 0.25 to 0.52. It is also stated in the EPRI report that the

highest values of 0.48 and 0.52, which were encountered with only one of the valves, may

not be due to simple sliding friction. The overall condition of the seating faces in this

valve was found to be excellent. Even though a possible mechanism for the higher values

was proposed qualitatively in the report [24], no conclusive quantitative explanation was

given. This valve design needs to be reviewed further to derive more definite conclusions

regarding the reason for apparently higher-than-expected coefficient of friction values

based on the overall condition of the valve.

-40 -
3.8. Conclusions from Presently Available Friction and Galling Data

The discussion on coefficient of friction and galling for Stellite vs. Stellite materials for

gate valve applications can be concluded with the following observations and conclusions:

1. In our experience, the coefficient of friction of Stellite for normal sliding behavior

without galling of the surfaces can range from 0.12 to 0.5 based on the presence or

absence of an absorbed layer of lubricant at the sliding surfaces. Based on our

assessment of the test results in Section 3, the typical range using cold water (without

prior exposure of specimens to high temperature) is from 0.15 to 0.25, and with high

temperature water or steam is from 0.3 to 0.4.

Based on our experience, we have found that it is necessary to perform testing on

actual components that duplicate the geometry and size of the sliding contact to obtain

applicable friction data. Test specimens that have markedly different geometries

and size can produce significantly different results than valves in actual

application.

2. The absorbed contaminant layer of lubricant is sometimes only a few molecular

layers in thickness; however, it can significantly alter the surface traction. The

absorbed layers can be removed by abrading the two surfaces against each other in

distilled water or by exposing the surfaces to high temperatures. Chemical solvents

are typically not effective in removing the absorbed layer of lubricant. In gate valve

applications, the absorbed surface layer may be gradually removed during cycling

under differential pressure which causes high enough pressure at the contact. This

can result in a gradual increase in coefficient of friction approaching values

obtained with clean unlubricated surfaces.

3. Even though we have stated in Conclusion 1 that the normal range of coefficient of

friction for Stellite can be up to 0.5, in our experience values above 0.4 are usually

associated with some type of surface damage such as galling, excessive localized

wear, significant change in surface roughness due to scratches, etc.

4. Threshold of galling stress for Stellite vs. Stellite for cyclic sliding applications is

reported to be around 47 ksi. The average contact stress under the contact should be

kept well below this to allow for higher local peaks in the contact area since the stress

distribution is rarely uniform. To obtain valid galling data, it is important to

faithfully duplicate the actual geometry, loading, and cycling conditions.

5. The initiation of localized galling at the sliding contact surface can either (1) result

in continued damage and deterioration of the sliding surfaces along with an

increase in friction forces or (2) heal itself by spreading until the average contact

-41 -
stress falls below the threshold of galling. Whether or not localized galling will

continue to spread until seizure or stabilize to a repeatable sliding behavior depends

upon the local geometry and the magnitudes of the loads involved.

6. Under relatively low contact stresses, there is a significant difference in the selfmated coefficient of friction values for different alloys of Stellite, i.e., Stellite-1,

Stellite-6, Stellite-12, and Stellite 21. Stellite-6 in self-mated tests is reported to have

the highest coefficient of friction, and Stellite-6 against any of the other Stellite alloys

has a significantly lower value. Howvever, under the higher contact stresses

typically encountered in valve seats, there is not much difference in their

coefficients of friction.

7. Duke Power tests showed significantly lower coefficient of friction performance

when using stainless steel disc and seats with Stellite overlay, instead of carbon steel

disc and seats with Stellite overlay. The difference in performance may be due to

iron content in the overlay caused by dilution from the base metal. The effect of iron

content or other significant alloying constituents should be investigated in

controlled tests.

8. When comparing coefficient of friction data obtained from valve tests performed by

different organizations, it is important to distinguish whether the seat inside

diameter or mean diameter was used; and wvhether the valve factor or coefficient of

friction is being reported. The combined effect of these variations can easily amount

to as much as 10 percent difference in the reported results.

-412 -
4. PREDICTION OF THRUST OVERSHOOT DUE TO INERTIA

This section presents an analytical methodology to predict inertial thrust overshoot in

motor operated gate valves. The industry has relied on some rules of thumb, and mostly

experience, to make estimates of thrust overshoot. Actual testing using MOV diagnostic

devices is the only method used to reliably quantify the magnitude of the thrust overshoot at

the present time. In some cases, this results in unexpectedly higher thrusts that exceed the

manufacturers' ratings of the valve or actuator components.

The principal investigators have developed an analytical methodology from first principles that can be used to predict thrust overshoot due to inertia. The predictions using this

methodology have been compared against actual test data for the 4-inch Borg-Warner

flexible wedge gate valve obtained by Duke Power Company in their Riverbend Steam

Station flow loop. The overall comparison between the predictions and the test results show

very good agreement, thus confirming that the methodology is sound. Additional comparisons against test data should be made to further validate and/or refine the analytical

approach presented here. In the meantime, an analytical tool has been developed which the

industry can use to improve the MOV reliability and performance.

The following sections present the details of the methodology. Appendix D documents the

detail calculations used in comparison against Duke Power test data.

4.1. Description of the Inertial Overshoot Phenomenon

Figure 4.1 taken from Reference 1251 showvs a typical wedge gate valve stem thrust versus

time curves for closing sequences under pump generated flow and no flow conditions. The

stem thrust during the running portion of the closing stroke is low and nearly constant for

both full flow and no flow conditions. As the disc approaches the closed position, the disc

friction force begins to increase because of differential pressure buildup across the disc.

After flow interruption, the disc friction force remains high and relatively constant

during the time that the disc is sliding against the downstream seatwith differential

pressure across the disc. Stem thrust builds up rapidly after the wedge makes solid contact

with both upstream and downstream seats. During the gate wedging action, the torque

switch trips at a preset value to de-energize the actuator motor.

-43 -
Legond

Under Flow Conritions

signal

)4otor starts

-moving St-im

T.nsi iIe

rtnrust

(1-

Hamner 31cw (Stem nut starts to move.)

Stemn tension

rel i eed

- - - Under No Load ConlItions

ning Load (Packing friction load)

Running Load (Packing friction+ piston

P effect load)

Stem f ves to pick up

tne clearance between

stem and disc

rSpring I.

Ipack 4, re] ie',

S I em Cooo~sS i v

Thrijs.

I -)

Time

Oisc solid

in seat

Steen and disc

begin motion

St em,-t a- -

stemI nut

C~ earar..ca

Stem-to-sten nut

clearance closed

(Stem nut starts to

push on stem.) Switch trips

Motor turns off - Final

_, Thrust

-Notes:

1. illustrated trace Is for packing friction load

--- ceedIlng piston effect load.

2. Nit to sceale.

Figure 4.1

Typical MIOV Closing Scquence for Gate Valve-Stem Thrust Versus Time [25]

- 44 -
Inertia of the drive train between the motor through the worm gearing, stem nut, stem, and

gate assembly can cause stem thrust overshoot after the torque switch trips and cuts off the

current to the electric motor. Overshoot beyond the torque switch tripping point depends

upon the kinetic energy of the system and the additional energy added to the MOV by the

motor during the time delay of the contactor switch to cut off the electrical supply to the

motor. The system energy available after the torque switch trips is used to overcome the

disc frictional drag for the remainder of travel, and the excess is converted into strain

energy in the system. The magnitude of stem thrust overshoot can be calculated

considering the energy balance to account for various factors including inertia of the

components, AP across the disc, component stiffness, and the distance of disc travel from

torque switch trip point to final seating position, as shown in the following subsections.

Appendix D documents the actual calculations for a 4-inch Borg-Warner flex wedge gate

valve used in the Duke Power flow tests (see Footnote 1 on page 15).

4.2. Available Energy After Torque Switch Trip (TST)

After the torque switch trips, the motor continues to run for a short time due to time delay

normally associated with the contactor dropout. Therefore, the energy available to wedge

the disc further during the final stage of closing is the sum of motor work after the torque

switch trips and the kinetic energy of the actuator and valve components at that instant.

The method of estimating the available energy components are discussed in Sections 4.2.1

and 4.2.2.

4.2.1. Motor Work After Torque Switch Trip

The motor work after the torque switch trips can be estimated based on the time delay in

contactor dropout and the motor running speed and torque at trip as:

W = wT ATr (Eq. 4.1)

where W = motor work after TST, in-lb

co = motor shaft rotating speed at TST, rad/sec

T = motor running torque at TST, in-lb

ATr = time delay in contactor dropout, sec

The motor work is calculated at the motor shaft location. Using the same method, the

available work at the stem nut location can be estimated by replacing motor speed and

torque with stem nut speed and torque. The available energy calculated at the stem nut

location is expected to be lower than the available energy at the motor shaft location due to

additional frictional loss from the motor shaft to the stem nut. The time delay in contactor

dropout, depending on the specific motor design, can typically vary between 10 and 30

milliseconds.

-45 -
4.2.2. Kinetic Energy of Moving Components

The kinetic energy stored in both actuator and valve components can be estimated using

the following equations.

Rotating Components

The major rotating components in the valve assembly are the motor shaft, gears, worm,

and worm gear assembly. The kinetic energy for the rotating components can be

estimated as:

KE= 1Wc2 (Eq. 4.2) 2

where I = mass moment of rotating inertia, in-lb-sec2

0) = angular speed, rad/sec

Rectilinear Moving Components

Kinetic energy for rectilinear moving components such as stem and gate can be

estimated as:

KE= 1MV2 (Eq. 4.3)

2

where m = mass of the disc and stem, lb-sec2/in

V = disc velocity, in/sec

4.3. Stored Energy in Valve Components After Torque Switch Trip

Load-transmitting components from motor shaft to valve disc and seats experience

different levels of stress and strain. The stored energy in the MOV components can be

estimated using the following equations.

Axial Load

After torque switch trip (TST), the stored energy in an axially loaded valve component

such as a stem is estimated as:

SE = SEf-SEt (Eq. 4.4)

where SE = stored energy in component after TST, in-lbs

SEf = stored energy in component at final thrust, in-lbs

SEt = stored energy in component at TST, in-lbs

or

SE= 2 EA (F FL ) (Eq. 4.5)

-46 -
where L = effective component length, in.

E = elastic modulus, psi

A = cross-section area, in 2

Ff = final axial load, lb

Ft = axial load at TST, lb

Torsional Load

The stored energy for torsional load is estimated as:

SE = SEf-SEt (Eq. 4.6)

L T2-T2 (Eq. 4.7)

where L = effective component length, in.

G = modulus of rigidity

= E/2(1 + v), psi

E = elastic modulus, psi

V = Poisson's ratio

J = polar moment of inertia, in4

Tf = final torque, in-lb

Tt = torque at TST, in-lb

Valve components such as motor shaft, worm, and stem are subjected to torsional loads

during the final stage of valve closing.

General Spring Load

The axial load and torsional load cases can be considered as special cases of a general

spring load case. Any linear elastic MOV component can be analyzed as a general

spring as long as the component stiffness is known. This approach may be used in

calculating the strain energy stored in the actuator spring pack, a disc spring, valve

disc, or any other highly loaded/strained component of the MOV. The component

stiffness may be derived from the available closed form solution, through experimental

testing, or by performing detailed finite element analysis.

Stored energy in a linear elastic component is given by:

SE = 2 Fr or - I Ft t (Eq. 4.8) 2 2

or

SE = 2K (F -1; 2) (Eq. 4.9)

-47 -
where &f = final component deformation, in.

St = component deformation at TST, in.

K = component stiffness, lb/in

4.4. Energy Dissipated After Torque Svitch Trip

Other than the strain energy stored in the valve components, one of the major sources of

energy consumption in the final stage of valve closing is the frictional loss between

sliding components. The key areas where energy is dissipated after TST are summarized

below:

Stem Paching Frictional Loss

Energy consumption due to stem packing friction is estimated as:

Ll = F1 Ad (Eq. 4.10)

where L1 = stem packing frictional loss, in-lb

F1 = stem packing frictional force, lb

Ad = distance of stem travel after TST, in.

The stem travel distance after TST, Ad, can be expressed in terms of stem thrust loads

and gate stiffness and geometry as:

Ad = Fr - Ft (Eq. 4.11) 2Ksin 0(sin 0+gcosO)

where Ff = final stem thrust, lb

Ft = stem thrust at TST, lb

K = disc and seat assembly stiffness (one side), lb/in.

0 = one-half wedge angle, deg

St= coefficient of friction for disc and seat interface

Work Against Stein Rejection Force

During valve closing, the stem thrust is required to work against the stem rejection

force. The energy loss after TST is:

L= F2 Ad =-d 2 P Ad (Eq. 4.12) 4

where d = stem diameter at packing, in.

P = pressure inside the valve body, psi

-_8 -
Frictional Loss Due to Disc Fri ction unuler AP

Energy loss due to the disc sliding against the downstream seat under AP is calculated

as:

L3 = F3 Ad (Eq. 4.13)

where F3 = APxl d2 (Eq. 4.14)

ds = effective disc sealing diameter, in.

Frictional Loss Due to Disc Wedging

The disc frictional loss during final wedging can be estimated as:

L4 = F4 Ad (Eq. 4.15)

where F4 = 2 (Fr + FL) - (F1 + F 2 + Fi) (Eq. 4.16)

Frictional Loss in Worm/lor-m Gear and

StemlStem Nut Connections

Energy loss due to friction between the worm/worm gear or stem/stem nut interfaces can

be estimated by using the worm gear/threaded connection efficiency as shown below [361:

Output Energy = e x input energy

where ec - On - [ t x (Eq. 4.17) CoOSO + P cot X

On = one-half of the thread angle, deg

11 = coefficient of friction

X = lead angle, deg

4.5. Energy Balance and Final Thrust Pr'ediction

After all of the energy components are known, the final thrust, Ff, can be calculated by

considering the overall energy balance after torque switch trip:

Total available energy = total stored energy + total dissipated energy

Total available energy is the combination of kinetic energy of the moving components and

the motor work due to contactor dropout delay after torque switch trip, as described in

Section 4.2. Stored and dissipated energies are calculated using the equations given in

Sections 4.3 and 4.4. Appendix D documents detailed calculations using the above

described energy balance approach to predict stem thrust overshoot. The overall results of

this comparison are presented next.

-49 -
4.6. Comparison of Predicted Final ''lirust Against Test Results

The calculations along with the relevant assumptions documented in Appendix D show a

predicted final thrust of 20,535 pounds for the 4-inch Borg-Warner flex wedge gate valve

tested under a maximum differential pressure of 2,000 psi, with torque switch tripping at

17,615 pounds. This final thrust of 20,535 pounds is in good agreement with the actual final

thrust that was measured to be 20,963 pounds using strain gages on the stem. The overall

comparison over 31 test cycles and the corresponding analytical predictions was found to

be within 9 percent for this valve.

From this comparison, it is concluded that the overall analytical methodology presented in

this report to predict stem thrust overshoot is sound. Further comparison should be made

against test results for valves of other sizes and other manufacturers to validate and refine

the assumptions used in the inertia overshoot model. This model presents the analytical

capability to predict final thrust due to inertia overshoot in the MOVs, for the first time.

This predictive capability can be used to improve the sizing and avoid overthrusting of

MOVs.

- 50 -
5. FACTORS AFFECTING OPENING THRUST REQUIREMENTS

Opening thrust requirements for gate valves to overcome the disc friction load were

presented in Section 2 for different types of gate valve designs. Unlike in the case of

closing operations, the disc friction load in the opening direction is not only dependent

upon the differential pressure across the disc , but is also significantly affected by other

factors. These factors are:

1. Wedging force the previous closing cycle, including the effect of inertia overshoot.

2. High body cavity pressure, resulting in the energization of both the upstream and

the downstream discs.

3. External piping loads causing disc pinching/sticking.

4. Temperature transients causing thermal binding of the disc.

Even though all of these factors can have significant impact on the operability of the valve

during opening, they do not lend themselves to reliable quantification and are therefore

not used in actual sizing calculations. In practice, the effect of these variables on the

operability performance has been minimized by bypassing the torque switch during the

first portion of the opening stroke, which makes the maximum actuator output available to

initiate opening. The problems caused by these factors, however, do surface when the

magnitude of the disc friction force increase due to these effects exceeds the actuator output.

This section provides an insight into these problems and how they affect the gate valve

operability in the opening direction. It also summarizes some analytical approaches that

have been used by the principal investigators to quantitatively investigate the magnitude of

these problems and assess -design modifications or alternative valve designs to solve

them. The method of analysis can assist the valve manufacturers in making design

improvements.

5.1. Effect of WedgingForce F-om the Pi-exious Closing Operation

The final wedging force, FN, from the previous closing cycle, affects the magnitude of the

opening force required. For the same switch settings on a MOV, the final wedging force

can vary because the inertia overshoot is affected by the magnitude of the differential

pressure across the disc. Typically, the highest wedging force is introduced when the valve

is closed without any differential pressure. The equations given in Section 2 of this report

can be used to quantify the unwedging force during opening if the wedging force is known.

The inertia overshoot effect on the final wedging force can be quantitatively addressed by

analytical techniques presented in Section 4, or by actual testing. Regardless of the

approach used, the dependence of the opening thrust during unwedging on the wedging

- 51 -
force from the previous cycle should be assessed to ensure that the actuator output is

sufficient. This is currently not a standard practice in MIOV sizing.

5.2. Effect of Higher Bonnet 1iressurae on Some Gate Valve Designs

Gate valves of the types known as flexible wedge, spilt wedge and double disc (See figures

2..1 a , b and c) have the ability to seal against both seats at the same time. The problem is

related to only these types of valves where the two disc seating surfaces can move

independently of each other; therefore, solid wedge gates do not experience this problem.

Two types of conditions can arise that lead to a higher pressure in the body cavity or bonnet

area than in either upstream or downstream piping:

1. When the valve is closed, fluid may be entrapped in the body cavity, and if the system is

then heated up, an uncontrollable rise in pressure in the body cavity can result. The

reported effects of such pressure increase range from inability to open the valve, to the

structural failure of the internal parts, or failure of the valve bonnet [501.

2. When the valve is in its closed position under a certain upstream pressure, and this

upstream pressure is subsequently reduced, e.g. due to the tripping of an upstream

pump, the original (higher) upstream pressure may remain trapped in the body cavity.

This can result in an increase in the opening thrust requirements due to the

energization of both the upstream and downstream disc seating faces. This condition

is referred to as double disc friction or double disc drag /16/.

Regardless of which of these two factors are responsible, the total disc friction force for a

double drag condition can be expressed as:

Fdt = Fdl d Fd2

= 1 . d2 (Pbi-P1) + It d2 (Pb - P2) 4 4

=r -d2 (2Pb - P, - Ps,)) 4

where Fdt = total disc drag force, lb

Fdi = upstream disc drag force, lb

Fd2 = downstream disc drag force, lb

t= coefficient of friction for the disc/seat interface

dS= disc sealing diameter, in.

P1 = upstream pressure, psi

P2= downstream pressure, psi

Pb= valve body cavity pressure, psi

- 52 -
The effect of lowering the upstream pressure before opening the gate on the stem thrust can

be illustrated in the following example.

Let P1 = 1,250 psi for the case of steady upstream pressure

P'1 = 350 psi for the case of lowered upstream pressure before opening

Pb = 1,250 psi for valve body cavity pressure

P2 = 0 psi for downstream pressure after closing

Then Fcit _ Lowered upstream pressure case - 2 Pb - P- P2

Fdt No change in upstream pressure case 2 Pb- PI- P2

2 x 1,250 - 350 - 0 2,150 - = 172% <

2 x 1,250- 1,250- 0 1,250

In this example, the stem force required to overcome the disc drag is increased by 72 percent

when the upstream pressure was lowered from 1,250 psi to 350 psi [161.

Provision must be made to eliminate the possibility of this excessive pressure build-up in

the body cavity to avoid excessive disc drag as well as structural damage. Simple methods

of mitigating body pressure increase to avoid double disc drag condition can be achieved by

equalizing the body cavity and upstream pressures with a single drilled hole, an external

by-pass piping, an internal or external relief valve, etc., as detailed in Reference 16. Each

of these options have their advantages and limitations which must be carefully reviewed

before selecting the one which is the most appropriate one for the specific application being

considered.

5.3. Effect of External Piping Loads

Variations in pressure, temperature, and piping load in a nuclear power plant piping

system can exert significant forces at the ends of the valve body, resulting in valve body

and seat plane distortions. Normally the valve body is much stronger than the connecting

pipe; therefore, external pipe loads on the valve are not of concern from the standpoint of the

structural integrity of the valve. The main concern about external piping loads on a gate

valve is the valve operability under these loads, especially for certain types of wedge gate

and expanding gate valves. As a wedge gate valve is closed, the space between the seats is

taken up by the relatively stiff gate with metal-to-metal contact on both upstream and

downstream seats.

Any changes in external piping load that tend to reduce the distance between the seat faces

after the valve is closed can cause gate "pinching" or "binding". Depending upon the

magnitude of those external loads and the stiffness of the valve members, the stem thrust

- 53 -
required to open the valve may increase. Load components that have the most significant

effect are the axial compressive loads, and bending moments. Torsional and shear loads

have negligible effect on the seat face distortions and opening thrust requirements in gate

valves.

The effect of these external piping loads is most pronounced on solid wedge type gate

valves. Some of the parallel expending gate valve designs also have high gate stiffness

and therefore exhibit high sensitivity to piping loads. The use of flexible wedge disc

reduces the sensitivity of the valve to piping loads. However, it should be noted that the

actual axial stiffness of the disc needs to be significantly less than that of the valve body to

ensure that its sensitivity to piping loads is negligible. Appendix E compares the predictions in opening thrust increase for a relatively stiff gate design used in an 18-inch

parallel expanding gate valve against a flexible wedge design. Based on the assumptions

stated in that appendix, it is shown that the stiff gate design could experience approximately

a 32 to 36 percent increase in opening stem thrust, whereas a flexible wedge gate opening

thrust is predicted to increase by only 12 percent under the applied axial load.

In summary, it is important to know that even a flexible wedge disc will experience some,

no matter how small, increase in opening thrust under compressive pipe load. Therefore,

simply specifying a flexible wedge gate design does not ensure immunity from the effect of

external piping loads. The valve manufacturers should quantitatively design the disc

stiffness and body stiffness that ensures that resultant increase in opening stem thrust

under anticipated pipe loads is less than a specified percentage of the normal operating

load due to differential pressure. At least one of the U.S. valve manufacturers has

published quantitative test results that show that increase in opening stem thrust is 5

percent or less in their valve design [273.

Another approach that eliminates the effect of external piping loads on opening stem thrust

requirements is to close the valve with no wedging, if the seat leakage requirements permit

it.

5.4. Effect of Temperature Changes on Opening Thrust

Some wedge gate valves exhibit "thermal binding" problems when they are closed hot, and

then allowed to cool down with the disc in the wedged position. Thermal binding is caused

by the body cooling down at the higher rate than the disc, thus causing an increase in the

seat contact force. The magnitude of increase in the seat contact force depends upon the

change in temperature, the difference in coefficient of thermal expansion between the body

and gate, the stiffness of the valve body and gate, and the distance between the seat faces.

Quantitatively, this increase in seat contact force, AF, can be evaluated from the following

equation (see Appendix F for details):

-54-
AF = L( a-(Y.g) AT K1) K9

(IK b +K 9)

where Lo = distance between seat faces

(Xb = coefficient of thermal expansion for valve body material

o.g = coefficient of thermal expansion for gate material

Kb = valve body stiffness along the flow axis

Kg = gate stiffness along the flow axis

The contact force increase can be negligible or high depending upon the actual materials,

valve and gate stiffness, and temperature variations. The opening thrust increase is the

product of AF and the coefficient of friction.

Another phenomenon related to the temperature effects that can also cause thermal binding

is the net growth of stem length when the valve is closed. This is caused by a net increase

in the overall length of the stem due to the previously exposed area of the stem being

inserted into the higher temperature environment inside the body.

Appendix F shows a quantitative example of a 3-inch solid wedge gate valve subjected to

temperature changes as described above. For the assumption stated in the appendix, an

increase of 3,943 lbs, which is equivalent to 60 percent of the stem thrust needed to overcome

the differential pressure load is predicted. The example also shows that gate stiffness, gate

and seat materials, and temperature variations of the valve components have a significant

influence on the calculated seat contact force. To reduce the thermal binding effects, one

may consider the use of a flexible disc; favorable material combinations of disc, seat, and

stem; and minimizing temperature changes in valve components. Actual testing of the

valve designs to qualify their relative immunity to thermal binding effects is the most

reliable approach to address these problems. Some of the valve manufacturers have

performed such tests. In order to ensure freedom from thermal binding problems, the

valve manufacturer should be requested to provide quantitative date to support their valve

design.

It should be pointed out that parallel slide discs are relatively immune to the thermal

binding problems, as well as gate pinching problem due to external loads.

- 55 -
6. CONCLUSIONS

All of the objectives stated for the SBIR Phase I effort have been fulfilled. Several

improvements in the analytical models have been made, as discussed in Sections 2

through 5. A significant contribution has been made by the principal investigators by

providing an analytical methodology to predict inertial thrust overshoot in an MOV gate

valve. A comprehensive review of friction and galling data applicable to gate valves is

provided. A preliminary analytical approach to quantify localized contact stresses due to

disc tilting has been developed. The concept of an index of contact stress severity has been

introduced which can be used to determine the potential for unpredictable behavior in gate

valves. However, an extensive matrix of tests is needed to empirically correlate the index

of contact stress severity to the actual performance of gate valves over a wide range of

conditions.

It is concluded that the Phase I effort was successful in completing the preliminary

development of improved analytical models to predict operability of the motor operated gate

valves. This can serve as a good foundation for continued analytical and empirical

development that is necessary to provide reliable and proven gate valve operability models

to the nuclear power industry.

-57 -
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-59 -
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49. P. Crook, The Development of a Series of Wear Resistant Materials with Properties

Akin to those of the Cobalt-Chromium Alloys, Wear of Materials, American

Society of Mechanical Engineers, New York, N. Y., 1981.

50. U. S. NRC IE Circular 77-05: Fluid Entrapment in Valve Bonnets, March 29, 1977.

51. NRC Licensee Event Report, San Onofre Nuclear Generating Station, Unit 1,

September 11, 1981.

-61 -
Appendix A

Opening and Closing Stem Thrusts for Variations in Disc Designs
AppendixA

Opening and Closing Stem Thrusts for Variations in Disc Designs

This appendix provides a detailed analysis of stem thrust required to close or open the disc

in the gate valve designs shown in Section 2 of this report. The analysis is based on free

body equilibrium considerations of the disc.

A. 1. DISC FORCE EQUILIBRILJM:

STEM LOAD FOR SOLID, FLEXIBLE, AND SPLIT WEDGE GATE VALVES

A.1. 1. Stem load to overcome AP - Closing

EFx = °,

Fpcos0 - FNcos0+Frsin0=0,and

Ff = pI FN

I . FN = ( cos= os 0 )Fp

Fp' ~~~~~~cos 0-~t sin 0)

W~~~f E ~~~~Fy = 0,

_y_ FsFrcos 0+FNsin0-FP sin 0, and

r 11N 11( cos 0 - gl sin 0 )

Figure A.1

Gate Equilibrium F,

under AP Load During Closing .- s=cos 0 -1t sin 0 F

where Fp = disc pressure load, lb differential pressure x area

Fs = stem load, lb

FN = seat normal load, lb

Ff = seat frictional load, lb

g= coefficient of friction between disc and seat

0 = one-half of wedge angle, deg
A.1.2. Stem load to overcome A1' -Opening

Y x

IFX = 0,

FP cos 0 - FNcos - Fr sin O = 0, and

Fr = ji FAI

*- FN=(cos0+ 0 sin 00. F

F,= Fp sin 0- FN sin O+F pcos0, and

-4I0

Figure A.2

Gate Equilibrium under

AP Load During Opening

Fr~g FN ( coss 0+ Fp (cos 0 + gsin 0)

F cosO+ sin 0

A.1.3. Stemwedgingload-Closing

1iS

4-0

The stem wedging load is the required stem

thrust to overcome the seat frictional resistance of a wedge gate.

X Fy =0,

FS = 2 (FN sin 0 + Fr cos 0), and

Fr = Wt FN

F,1N = 1 Fs

2(sin 0+ cos0)

or FB = 2 (sin 0 + g cos 0) FN

4-

Figure A.3

Gate Equilibrium under Wedging Load

During Closing
A.1.4. Stem unwedging load- Opening

A, Fy = 0,

Ft = 2 (FC COS 0 - FN sin 0), and

Ff = 11 FN

1 - . FN =2( c (0- i )

or F,= 2(jicosO- sin 6)FN

&

Figure A.4

Gate Equilibrium under Unwedging Load

During Opening
A.2. STEM LOAD FOR PARALLEL EXPANDING GATE VALVES

-THROUGH CONDUIT AND DOUBLE DISC

A.2.1. Stem load to overcome AP -Closing and Opening

v

F-' ifLA

Y Fx = 0,

Fp = FN

Z Fy = 0

Ff - g FN = Fp FIf

C/osing This equation applies to both gate

closing and opening conditions. Figure A.5

Gate Equilibrium under AP Load During

Closing/Opening

A.2.2. Stem wedging load -Closing

Y Fx =0,

Fhcos0- Fsin O=FN, and

F '7I _ nL5 F14cosO - g'sinOFN

>QFy = 0

F, = F~ sin 0+ Fjcos,0+ Ff and

F f

&9

Figure A.6

Gate Equilibrium under

Wedging Load During Closing

Ff = t FN, F; = 1' FN

F. =(1-. )sin 0+(p + p') cosO FN

s cos 0 - .' sin 0

For p. = ,u', this equation becomes:

F = sin 0(1- 2p) + 2g cos 0 8F

cos0- lsin n

- A.4-
A.2.3. Stem unwedging force-Opening

.El'YL

EFx = 0,

Fj cos O + Ff sin 0 = FN, and

Fj = g' FN

. Fjj - 1 -F

cos 0 + p'sin 0N

£ FY = 0,

= Fjcos 0- Fh sin 0+ Ff and

'F,

L9

Figure A.7

Gate Equilibrium under Unwedging

Load During Opening

Fr = A FN, F= g' FN

1. F _4 (O - 1)csin 0 + (+ s') Cos

S ~~cos 0+ g'sin0 FN

For p = pl', this equation becomes:

F. _r(,u -1l)sin 0+241coso3)F

S cosO+4sinO n
A.3. STEM LOAD FOR PARALLEL SLI)ING GATE VALVES

A.3.1. Stem load to overcome AP -Closing and Opening

1⁄4s~ ~~11w.h ~

Dw-n s*ram y0e

Figure A.8

Gate Equilibrium under AP Load During Closing

Assume that the coefficient of friction on both sides of the seats are the same; then for the

downstream disc, we obtain

£ FX = 0,

FN = Fp + FIp

where

Fsp = spring load between parallel gates, lb

EFy = 0

F, = Fr + F,

where F,,= upstream gate resisting force, lb
And for the upstream disc, we obtain

XFF =0,

F& = Fp

YFy =

F'= F=0

.-. F.=Fr+F.=Fr+F(=1i(FN+Fk)

Fs=2 F.p + lFp

This equation applies to both opening and closing directions.
Appendix B

Disc Load Calculations for a Gate Valve

In a Pump Flow System
Appendix B

Disc Load Calculations for a Gate Valve

In a Pump Flow System

Based on available test data for a 4-inch Borg-Warner flexible wedge gate valve tested

under typical pump flow conditions at Duke Power Company's flow loop (see Footnote 1 on

page 15 of main text), a simple analysis is presented in this Appendix to approximately

quantify the forces acting on the disc in mid-travel. Section 2.6 and 3.4 describe test

results and their analysis for this valve.

The pressure drop across the disc is estimated by using the gate valve flow resistance data

from References [301 and [31]. The flow velocity used in the pressure drop calculations

depends on pump characteristics and the overall piping system design. Finally, the

pressure load is computed based on the calculated pressure drop and the gate opening

position.

B .1. GATE VALVE FLOW RESISTANCE COEFFICIENT

The flow resistance coefficient as defined in Section 2 is given by:

K 2AP

pV2

where K = flow resistance coefficient

AP = differential pressure across the valve, lb/ft2

p = mass density of the fluid, slug/ft 3

V = flow velocity, ft/sec

Flow resistance coefficient data available from References [30] and [31] for gate valves are

summarized in Table B.1. Overall average values from these data are plotted in Figure

B.1 for reference.
Gate Position, h/Do*

Gate Valve Description 0.1 0.15 0.2 0.3 0.4 0.5 0.6 0.7 | 0.8 0.9 1.0

Gate valve in straight

pipe without a recess for 150 62 35 10 4.6 2.06 0.98 0.44 0.17 0.06 0

the disc

Gate valve in straight

pipe with a recess for the 200 77 33 11 4.7 2.35 1.23 0.67 0.31 0.11 0.05

disc

Gate valve at the exit 200 64 36 14.2 7.1 3.85 2.3 1.4 0.75 0.21 0.11

Gate valve with reduced 200 77 34 12.5 6.5 3.65 2.15 1.35 0.71 0.24 0.07

portII

Gate valve tested at U. of 212 155 90 23 15 4.5 4 3.5 2.5 1.5 0.38

Conventional gate valve 100 72 43 15 7.5 3.5 2 1.5 0.6 0.4 0.25

Disc type gate valve** 190 72.9 35.7 12.7 5.7 2.7 1.34 0.6 0.24 0.07 0

Total Average 179 82.8 43.8 14.1 7.3 3.23 2.0 1.35 0.75 0.37 0.12

* Gate position is represented by h (gate opening) and Do (pipe inside diameter at gate)

** Average flow resistance coefficient for the valve type. Data may be interpolated/extrapolated

to other gate positions without specific test data.

Table B.1

A Comparison of Gate Valve Flow Resistance Coefficients

(Data from References 30 and 31)
260 -

240 -

220 -l

200-

_180 -

160 -

o 140 -

0D 120-

U

C

X 100-

0 80-

3 60 V

0

40 -

20

0~

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0

Gate Position, h/Do

Figure B.1

Average Flow Resistance Coefficient Variation as a

Function of Disc Travel for Gate Valves
2100 -

2000

F) 1900 \

< 1800

; 1700 -

wo 1600

1500

1400

1300

1200

GA 111

1100 .. . . . . . . . . . . .

0 1 00 200 300 400 500 600 700 800 900 1000 1 100 1200 1300

FLOW RATE, GPM

Figure B.2

Estimated Pump Performance Curve Used in

Gate Valve Pressure Drop Calculations

-B. 4-
B.2. PRESSURE DROP ACROSS GATE VALVE IN PUMP FLOW SYSTEM

Pressure drop across a gate valve in a pump driven flow system was computed using an

example case for 4-inch flexible gate valve tested at Duke Power Company's flow loop (see

Footnote 1 on page 15 of main text).

In a pump driven flow system, the overall piping system pressure drop should equal the

available pump head in a steady flow condition. Therefore, the system pressure drop can

be expressed as:

APpump = APsystem = Ktotal Pv2

where APpump = available pump head at a specific flow rate, psi

APsystem = total system pressure drop at a specific flow rate, psi

Ktotal = total system flow resistance coefficient

= Kvalve + Kothers

Kvalve = valve flow resistance coefficient

Kother other components' flow resistance coefficient

518 (estimated from Duke Power flow loop data)

Based on a given pump flow characteristics curve, as shown in Figure B.2, the pressure

drop across a gate can be estimated as:

APpipe =APsystem = (KvaIve +518) v

2

where Kvalve and p are known for a given gate position and fluid. Using a numerical

iterative solution approach, the system flow rate and pressure drop across the gate valve

can be calculated. Iterative solutions may be started with a low estimated value of flow rate

for the piping system to calculate the system pressure drop. If the system pressure drop is

lower than the available pump head given in Figure B.2 for the same system flow rate, then

the estimated flow rate should be increased to match the available pump pressure head.

This process is repeated until the pump head and system pressure drop are equal. Table B.2

summarizes the results for the 4-inch Borg Warner valve tested in Duke Power flow loop,

using the above described procedure.
Gate Positon, System Flow Rate, Gate Valve Pressure

LID, gpm Droppsi

0 0 2,000

-0.1 778 487

0.15 843 258

0.2 870 145

0.3 890 49

0.4 895 25.6

0.5 899 11.5

0.6 900 7.1

0.7 900 4.8

0.8 900 2.7

0.9 901 1.3

1.0 901 0.4

Table B.2

Estimated Flow Induced Pressure Drop Across a 4-Inch Gate Valve at

Different Valve Openings

B.3. PRESSURE LOAD ON DISC

A simple and maybe conservative way of estimating the pressure load on a gate disc is by

assuming that the pressure load on the gate is proportional to the pressure drop across the

valve and the percentage of gate closing as:

dp ( 100 - % of gate opening x Adisc = s-" ~100 xA

where Fp = pressure load on gate in mid-travel position, lbs

AP = pressure drop across the valve, psi

Adis, = disc area

= . x (disc sealing dia)2

4

Using a nominal disc sealing diameter of four inches, the pressure loads at different valve

opening positions are summarized in Table B.3. These loads can be used to calculate

localized contact stresses between the disc and guide or disc and seat to determine the

potential for galling damage (as shown in Appendix C).
Percent of Gate Pressure Drop Pressure Load

Opening, % Across Gate, psi lb

0 2,000 25,133

10 487 5,508

15 258 2,756

2) 145 1,458

30 49 431

40 25.6 193

50 11.5 72

60 7.1 36

70 4.8 18

80 2.7 7

90 1.3 2

100 0.4 0

Table B.3

Estimated Flow InducedLoad on a 4-Inch Gate at Different Valve Openings
Appendix C

Analysis of Local Contact Stresses
Appendix C

Analysis of Local Contact Stresses

Contact stresses between two surfaces can be estimated using closed-form linear elastic

solution for single, well-defined geometries such as spherical, elliptical, cylindrical, and

plane surfaces. In this section, two contact stress cases are investigated using Hertz

contact stress equation for point and line contacts. The contact loads used in the analysis

are the estimated pressure loads caused by disc tilting as shown in Appendix B, and Section

2.6. The results of contact stress analysis are compared with material threshold of galling

stresses for the study of disc tilting effect in mid-travel position in Section 3.4. It should be

pointed out that this contact stress analysis is based on linear elastic, small displacement

assumptions using original (unworn) geometry of contact.

Contact stresses based on these simplifying assumptions should be used as an index of

contact stress severity and used as design guides for comparison against actual test data or

to make design improvements.

Hertz equations for contact stress distribution

The intensity of contact stress, p, over the

surface of contact between two general

curved surfaces can be calculated from the

following equation derived by Hertz [38]:

1x 2 y2 (C. 1)

p p ~ a2 b2

where p = contact stress at a selected

location (x, y)

po = peak contact stress, psi

x' x = X axis coordinate, in

y = Y axis coordinate, in

Figure C1: Contact stress distribution a = major semiaxis, in

between two general curved surfaces [38[ b = minor semiaxis, in

P 2n abp (C. 2)

3

where P = total load, lb

7X a b = contact area (ellipse), in2
where

a = m'3 P

a 4A

b = n 3PA

4A

A 1'v- + 1LY1

El E2

A= 1 01 + 1 + 1 1 1

B= 1 ; 1 - 18 +( 1 1, +2 1 - 1' t 1 - 1,)cos 2]

(C.3)

(C.4)

(C.5)

(C.6)

(C.7)

m=f (cos'(B/A))

n=f' (cos l(B/A))

(see Table C.4 below)

(see Table C.4 below)

(C.8)

(C.9)

Also, v, E, R, and R' denote Poisson's ratio, Young's modulus, and minimum and

maximum radii of curvature of the unloaded contact surfaces. Subscripts 1 and 2 denote

contacting bodies 1 and 2. W is the angle between the planes containing curvatures 1/Rl

and 11R2.

cos 1 (B/A) 300 350 40° 450 50° 550 60°

m 2.731 2.397 2.136 1.926 1.754 1.611 1.486

n 0.493 0.530 0.567 0.604 0.641 0.678 0.717

cos-1 (B/A) 650 700 750 800 850 900

m 1.378 1.284 1.202 1.128 1.061 1.000

n 0.759 0.802 0.846 0.893 0.944 1.000

Table CA

Values of m and n for Hertz Equations [381
Example 1: Point Contact Between Disc and Downstream Seat

This example simulates a tilted gate contacting the edge of the seat inside diameter with

two point contact, as shown in Figure C.2.

I Gate

Contact Area See Enlarged Detail

Enlarged Detall

Figure C.2

Tilted Disc Contacting Seat

The following data case was used for contact stress calculations:

R =R 2 = 0. 125 in.

R =2 in., R2 =-2 in. (concave)

P = 41 = 215.5 lb (from Appendix 13 @ 30% opening) 2

El= E2=30 x 1o6 psi

V = V2=-3

The actual angle, N, between the curved surfaces present at the radiused edge of the disc

outside diameter and the radiused edge of the seat inside diameter with the gate in the tilted

position is somewhat difficult to calculate. However, it can be seen from Equation C.7 that,

for values of v between 0° and 150 (or 2 = 00 to 300), the magnitude of B does not

significantly change. Therefore, 4f = 150 is used in the following calculations.
Applying the Hertz equation, we have

A=2x 1-0.09

A 302x 30 x 10 =6.067 x 108 106

A= 1( 2 + 2) =8

2 0.125 2

B =2 [(8 - 0.5)2 + (8 _0.5)2 + 2 (5 -0.5)2 + cos 300] = 7.728

cos-1 (B /A)= 15°

m = 4.2 (extrapolated from Table C. 4)

n = 0.39 (extrapolated from Table C.4)

. ~5.5x6.067x 10-8

a= 4 .2 ] 4x8 =0.04495inch

b = 0.39 3 x 215.5 x 6.067 x 10= 0.004174 inch

4x8

3 x 215.5 548,411 i

2 Tcx 0.04495 x0.04174 584pi

Pave = -o = 365,608 psi

The estimated local contact stress at the downstream disc-to-seat point contact is almost

one order of magnitude higher than the Stellite material threshold of galling stress of 47

ksi as discussed in Section 3.4, based on linear elastic, small displacement, and original

unworn geometry assumptions elastic stress analysis. The calculated results a and b,

which are .045 inch and .0042 inch respectively, are the dimensions of the elliptical contact

at the seat to disc interface under load at 30 percent mid-travel disc position. Local

yielding, load redistribution, and material wear are likely to occur before the contact

stress actually reaches this level.

Example 2: Line Contact Between Disc and Guide

In some gate valve designs, line contact between the disc guide and disc guide slot resists

the flow-induced load imposed on the disc during mid-travel. This example presents an

estimate of contact stresses based on the following design and material data assumptions:
Rl= 0. 1875 in.

Disc Guide Slot

R2 =R1 =R 2

L = 0.375 (contact width) in.

R1~ ~~~~4=

431 P= 2-=215.5lbs

I~~~~~~~

El=E2 =30x106 psi

Vl V = 0.3

Figure C.3

Math Model for

Line Contact between Disc and Guide

Applying the Hertz equations for parallel cylinders ([381, p. 374):

A = 6.067 x 10-

P A 1.5 x 6.067 x10-8 b= L 13 = 113 = 0 002889in.

Po 2P 2 x215.5 1263ps -

X L b -t x 0.375 x 0.002889 126,633 psi

Pave = 99,458 psi v

This estimated contact stress with line contact between the disc guide slot and the guide

surface is an order of magnitude lower than in the case of disc making a point contact

against the downstream seat. Also, it is in the same order of magnitude as the material

threshold of galling stress for Stellite. As mentioned at the beginning of this appendix, the

contact stresses based on simplifying assumptions used in the analysis can be used as an

index of contact stress severity. Comparisons against actual tests under controlled

conditions should be used to obtain an empirical correlation between these theoretical

values and actual performance.
Appendix D

Analysis of Stem Thrust Overshoot for

4-Inch Borg-Warner Valve Tested at Duke Power Flow Loop
Appendix D

Analysis of Stem Thrust Overshoot for

4-Inch Borg-Warner Valve Tested at Duke Power Flow Loop

This appendix documents the analysis performed for stem thrust overshoot estimation

using the energy balance method described in Section 4 of the main report. The actuator

and valve data selected for the calculations are Rotork 16NAX1, 57 rpm actuator and BorgWarner 4-inch flex wedge gate valve used in Duke Power Company's flow loop tests. The

analysis results compare well with the actual test data, as shown in this appendix.

D.1. AVAILABLE ENERGY AT TORQUE SWITCH TRIP

D .1.1. Motor Work After Torque Switch Trip

Motor work for Rotork 16NAX1 actuator is estimated as

W= oT At

Xa = motor speed = 3,600 rpm x 0.9 = 3,240 rpm

T = motor torque = 8 ft-lb = 96 in-lb

Ax = contactor dropout delay = 10 ms = 0.01 sec

27r W 1= 3,240x - x-96 x0.01=326 in -lb

60

D.1.2. Kinetic Energy of Moving Components

1. Rotor

Based on the estimated rotor dimensions shown in Figure D.1, the rotor kinetic

energy is:

KE= 1 (3600X2x)2x2 ( 8.3 XL552 + 2.46 x05332)

2 60 4 386.4 386.4

= 1,911 in -lb
m 7' 3. l0" (b 1.066" -0

E 1 %MX \ \\\\\\\\ rpm = 3600

fl, -m2 =2.96 lb

Hi 3.875" j m t= m 1+ m 2-11.26 lb

15.5" I ~~~~~~~~~~~~~~~~~~~~~~~~~~~

I I

Figure D.1

Worm Shaft Weight and Dimensions

2. Worm gear/stem nut assembly

Based on the estimated dimensions shown in Figure D.2, the kinetic energy for the

assembly is:

( 7r'2 1 2 . E 2 = (57 x 2 x( 4 x 1752+ 587 x L252J

60 024 386.4 386.4

= 0.29in -lb <- small

..........

m 2

Estimated Weight & Dimensions:

1 =3.5 0

02=25 *

ml =1.21b

rE i 1111 M2=5.7llb

rpm = 57

Figure D.2

Worm Gear/Stem Nut Assembly Weight and Dimensions
3. Valve stem and gate assembly

The estimated valve stem and gate dimensions and weight are:

Stem Dimensions = 32" x 1.375"'

Stem Weight = 32 x 1.3752 x 0.7854 x 0.283 = 13.45 lbs

Gate Dimensions = 2.7" thk x 4.5" 4

Gate Weight = 0.7854 x 4.52 x 2.7 x 0.283 = 12.2 lbs

The estimated valve closing speed is 0.5 in/sec. Therefore, the kinetic energy for the

stem and gate assembly is:

(135 3.4 + 12.2 x 0.52

386.4 )

= 0.0083 in -lbs v- very small

D.2. STORED ENERGY IN VALVE COMPONENTS

AFTER TORQUE SWITCH TRIP

Theoretically, all of the load transmitting components store some strain energy. For this

example case, only the three members (stem, disc, and spring pack) that have significant

strain energy stored during valve closing are considered here.

D.2.1. Stem

The stored stem energy under axial load is:

SE1 = 2EA (F2 - F

for Ft = 17,615 lbs thrust at TST

L = 32 in.

E = 30 x 106 psi

d = 1.375 in.

SE1 =3.5917x 10 7 (F - 310.288 x 106)

The stored energy in torsion is:

SE2 L (T- - T2) 2GJ ~r t
fo r G = 1.15 x 107 psi

J = 32 (L375)4 =0.3509 in4

32

Tt = 200 ft-lbs = 2,400 in-lbs

SE2 = 3.965 x 10-6 (To - 2,4002)

= 3.965 x 10- 2,400 ) Fr - 24 0021 V 17,615) 0

D.2.2. Disc

The stiffness of the Borg-Warner flex wedge disc may be estimated using the following

simplified mathematical model.

Assume the gate (one side) is a circular disc with the outer diameter uniformly loaded and

the central circular area fixed or free as the two extremes.

SEA T LOAD , sE4 T LOAD, v 7"' UNIFORMLY LOADD 0 UNIpRoeALY IzAP6D

-o

wS _ir_1 ,~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~

* CC3U7XAL C/~cULAR~ AR6A4

06 41NIFOV-1LY 1LOADGOD

0. 75 "

Model 1 Model 2

Figure D.3

Simplified Math Models for Gate Disc Flexibility Calculations

Based on the above disc dimensions, disc stiffness is estimated using closed-form

equations as given in Reference [371, pages 338 and 366:

K1 = 2.0525 x 106 lb/in (Model 1)

K2 = 1.434 x 106 lb/in (Model 2)
The lower stiffness of 1.43 x 106 lb/in is selected for the stem thrust overshoot calculations.

Mote precise gate stiffness may be obtained through the finite element analysis or testing.

The stored energy in the gate is:

I 2 2x(Fr2 - F. 2

21(2

F'2 - F'2

=

K2

where F' Ff;(F1 + 2+F 9 ) final gate lateral load due to wedging 2 (sin 0 + ii cos 0)

F- - 3,935

0.92146

Ft Ft - (F1 + c2 + Fo ) = gate lateral load at TST

2 (sin 0 + jt cos 0)

17,615-(891 + 2,970 + 9,819)

2 (sin 5° + 0.375 cos 50)

= 14,846 lbs

SE 3 - 1 6 [(L0852 Fr - 4,270) 14,84621

D.2.S. Spring Pack

The stored energy in the spring pack after torque switch trip is:

SE4 = K. (S r- _S)

where Ks = spring pack stiffness, lb/in

Sf = final spring pack load, lb

St = spring pack load at TST, lb

The estimated spring pack stiffness and load at torque switch trip are:

Ks = 6,000 lb/in

Sf = 600lbs

: SE 1 600 F _ 6002

S4 = 12,000 (17,615 Fr) 60
D.3. ENERGY DISSIPATED [N VALVE COMPONENTS

The energy dissipated in valve components after torque switch trip is estimated as follows:

D.3.1. Stem Packing Frictional Loss

Ll = F 1 Ad

where F1= = Kp7rdf)

2

0. lx 15 x 2,000 (nt x L375 x L375)

2

= 8911bs

Ad= Fr - F. 2K sin 0 (sin 0 + vi cos 0)

Fr- 17,615

2 x 1.434 x 10('sin5°(sin5°+0.375cos5°)

L1 = 7.737 x 10-3 (Fr- 17,615)

D.3.2. Work Against Stem Rejection Force

L2 = F2 Ad

where F2 d2 p = x L 375 2 x 2,000 = 2,970 Ibs 4 4

L2 =2,970 x 8.683 x 106 (Fr-17,615)

= 0.01875 (Fr - 17,615)

D.3.3. Frictional Loss Due to Disc Friction Under AP

L3 = F3 Ad

where F= 4 pt s i )

cos 0 cos 0 - RL Sill 0

T1 42

-4 x 2,000 x cs5 0.375si5

cos 5° cos 5° - 0.375 sin 5°

=8,919 lbs

L3 = 8,919 x 8.683 x 10-6 (Fr - 17,615)

= 8.5258 x 10-2 (Fi - 17,615)
D.3.4. Frictional Loss Due to Disc Wedging

L4 = F4 Ad

where F4 = 2 (Fl + F ) - (F1 + F> + F:1) 2

= j (Fr + 17,615) - (891+ 2,970 + 9,819)

=0.5Ff-4,873

.L4= (0.5Ff -4,873) x8.683x 10-(Fr- 17,615)

D.3.5. Efficiency of WormtWorm Gear

cos On - g tan X

Cos On + P cot X

Using the following design data:

On= 14.50

, 0.08

= 40

cos 14.5°- 0.08 tan 4°

cos 14.50 + 0.08 cot 4°

Therefore, rotor kinetic energy and motor work are reduced to 46 percent of the original

magnitudes when they pass through the worm/worm gear connection.

D.3.6. Efficiency of Stem/Stem Nut

e= cos On 0 - tan cos On + t cot X

For the stem design of:

On = 14.50

= 0.08

X=tan- t -tan 1 0.5 =7.0880

ndm tx L28

e cos 14.5° - 0.08 tan 7.0880 59%

cos 14.5° + 0. 08 cot 7.0880
D.4. ENERGY BALANCE AND FINAL THRUST PREDICT'ION

The overall energy balance for the example problem can be summarized as:

[(Xl+ KEB- SE4 ) el + KE2 ] e2 + KE 3 = SE,+ SE2 + SE3 + Ll + L2 + L3 +L4

Solving the above equation, we have

Ff = 20,535 lbs

The calculated stored energy and losses in the valve and actuator components are:

SE1 = 40 in-lb

SE 2 = 8 in-lb

SE3 = 73 in-lb

SE4 = 10 in-lb

Ll = 23 in-lb

L2 = 75 in-lb

L3 = 249 in-lb

L4 = 137 in-lb

The calculated final stem thrust of 20,535 pounds is in good agreement with the measured

stem thrust of 20,963 pounds. Thrust predictions were compared for 31 test cycles, and the

average deviation was found to be within 9 percent.
Appendix E

Analysis of Piping Load Effect on Opening Thrust
AppendixE

Analysis of Piping Load Effect on Opening Thrust

The piping load effect on opening stem thrust is illustrated using a through-conduit 18 x 16

x 18 parallel expanding gate valve as an example. This type of valve design has a relatively stiff gate assembly, similar to solid wedge gate valves which tend to be particularly

sensitive to gate pinching under external pipe loads. Moreover, valve bodies tend to be less

stiff in larger sizes, which accentuates the gate pinching problem in larger valves.

Calculations are also performed for flexible wedge gate design to show the relative

improvement. Even though bending moment in the plane of the valve stem and flow axis

has the worst potential for affecting the valve performance from the standpoint of gate

pinching, the example shown here addresses the effect of axial compressive load only to

keep the illustrations simple.

Body Stiffness

As discussed in Section 4, the magnitude of gate pinching can be estimated through the

valve body and gate stiffness calculations. For this 18 inch valve example case, the valve

body stiffness along the flow axis is known through a detailed 3-D finite element analysis

as documented in a reportl by Kalsi Engineering to W-K-M.

Kb = 5.757 x 107 lb/in

The overall valve assembly stiffness can be expressed by a simple math model as shown in

Figure E.1.

Kb12

Fp|| Fp F A /\D Fp

Seat Gate Body

GA 110

Valve Section Plot Simplified Valve Stiffness Math Model

Figure E.1

Simplified Math Model for Gate Valve Stiffness

J. K. Wang and M. S. Kalsi. Finite Element Analysis of the 18" x 16" x 18" ANSI 900 Feedwater

Isolation Valve Under Faulted Condition Nozzle Loads, Kalsi Engineering, Inc. proprietary report to

W-K-M, KEI 8.4.3, February 23, 1981.
The valve body stiffness obtained from the 3-D finite element analysis represents the valve

body stiffness without the seat and gate in place. The overall valve stiffness for a gate

valve at the closed position can be estimated based on the combined stiffness of body, seat,

and gate as shown in the following calculations.

Gate Stiffness

The gate type is a parallel expanding gate valve (Figure 2.1B). When this gate assembly is

wedged closed, all of the space between the seat faces is taken up by two relatively stiff

pieces of wedges, in the same fashion as in a conventional solid wedge design. Stiffness

for the two-piece expandable gate assembly can be estimated by using equivalent stiffness

of a pipe section with two assumed pipe thicknesses. More precise estimation would require

detailed finite element analysis or actual testing.

Case 1.

Gate Stiffness approximated by a pipe thickness equal to the seat contact width:

AE

' (172_ 14.6252 )x3Ox 10"

10.25

- 1.727x 108 lb/in

Case 2.

Gate stiffness approximated by a pipe of thickness equal to twice the seat contact width:

K 2 AE

= 2x 1.727 x 108

= 3.453x 108 lb/in
Seat Stiffness

Seat stiffness is also estimated using the dimensional data from KEI 8.4.3 (referenced in

Footnote 1 on page E.1), which can be approximated as an equivalent pipe section:

AE K 8 =-

E (172 _ 14.6252) x 30 x 106

1(19.5- 14.625) 2

=7.26 x 108 lb/in

Conbined Stiffness of Gate and Seats

Overall stiffness through the gate and seats is:

1 1 1 1 ___+ _+_

Kgs Ks Kg K8

2 1

7.26 x 108 L727 x 10"

Kg= L17x 108 lb/in

1 1 1 1

K's K8 Kg Kg

2 1

7.26 x ll 3.453 x 108

K' =L77x108 lb/in

Valve Stiffness

The overall valve stiffness is the sum of body stiffness and combined disc/seat stiffness:

Ktotai = Kb + Kgs

= L746x 108 lb/in (usingCaselgatestiffhess)

Ktoatl = Kb + K'gs

=2.346 x 108 lb / in (using Case 2 gate stiffness)
Gate Pinching Force -

Force through the gate assembly under a compressive external piping load can be

expressed as:

Fgs = Kg. Fp

total

where Fgs = force through gate assembly, lb

Fp = external piping load, lb

Based on this estimated gate and valve stiffness, the gate pinching force is:

F= 117 Fp;11 gs174.6

= 0.67 Fp for Case 1 gate stiffness, and

F 177

234.6 '

= 0.754Fp for Case 2 gate stiffness

The above estimate shows that 67 to 75 percent of the compressive external pipe load passes

through the gate assembly. This increases the gate/seat contact load, thus resulting in

higher stem thrust requirement. The amount of stem thrust increase depends directly on

the magnitude of the compressive external piping load. The following assumed piping

load and valve differential pressure are used in illustrating the external piping load effect

on stem thrust increase.

Assuming that for this 18 inch gate valve:

Compressive external pipe = 100,000 lb (approximately 23 percent of the

maximum estimated axial pipe load

under transient pipe rupture condition

(reference Footnote 1 on page E.3). This

corresponds to approximately 2,200 psi

axial compressive stress in the piping

connected to the valve end.)

Maximum differential pressure = 1,250psi

Then the gate pinching force is:

Fgs = 0.67 x 100,000 =67,000 lb for Case 1, and

F'gs = 0.765 x 100,000 = 75,400 lb for Case 2.

-E.{-
Gate/seat contact force due to maximum differential pressure is:

FAp= It(15)2X 1,250=212,058 lbs

4

Therefore, the stem thrust increase due to external pipe load can be expressed as a percentage of the required stem thrust to overcome the gate drag due to differential pressure load

as:

R= F2 =316%J, Fp

F'

R' = 5--= 35.6%

FAp

Reduction in Gate Pinching Effect by Flexible Wedge Gate

The above example used a parallel expanding gate valve with high disc stiffness. Many

gate valve designs employ flexible gate concept to reduce disc pinching effect under

external pipe loads. The following calculations show the effect of gate flexibility on the

stem thrust by replacing the parallel expanding disc with a flexible disc having the

following assumed dimensions shown in the math model:

F

016.0

<-2.75

Flexible Gate Math Model for 1/2 Flexible Gate

GA 112

Figure E.2

Flexible Gate Math Model

The gate assembly stiffness can be estimated using a closed-form solution as given in

Reference 37, page 338:

Kg = 2.127 x 10 7 lb/in
Using the same approach as given in the parallel expanding gate case, we have

Kgs= 2.1146 x 107 lb/in

Note that the stiffness magnitude of the flexible disc and seat assembly is less than 1/8th the

stiffness (1.77 x 108 lb/in) of the relatively stiff gate assembly used in the first example.

Ktota = 7.8716 x 107 lb/in

Fgs = 0.2686 Fp

= 26,860 lbs (for 100,000 lbs of external pipe load)

R 26,860

212,058

= 12.7%

Therefore, the stem thrust increase is approximately 12.7 percent of the required stem

thrust to overcome disc drag under differential pressure load, instead of 31.6 to 35.6 percent

stem thrust increase calculated for the high stiffness parallel expanding disc case.
Appendix F

Analysis of Temperature Effect on Opening Thrust
Appendix F

Analysis of Temperature Effect on Opening Thrust

This appendix provides the derivation of the seat contact force equation used in Section 5

discussions and an example case of the temperature effect on operating thrust. This

example is based on a root cause analysis performed by Kalsi Engineering, Inc. The

main purpose of the example is to show how to quantify various contributing factors that

effect the opening stem thrust due to temperature changes.

F.1. EFFECT OF TEMPERATURE CHANGES ON SEAT CONTACT FORCE

As shown in the following simplified math model, the valve body and gate expansions (or

contractions) due to temperature changes can be estimated as:

GA,113 8b = Lo axb.

8g = Lo axg,

where 5b =

5g =

ST (F. 1)

ST (F.2)

body expansion, in

gate expansion, in Original Contact

I Kg I

| i~- A Sbt2 Body Free Expansion

I+I

I - 1>2 Gate Free Expansion

Figure F.1

Math Model for Body and Seat Stiffness

Lo = distance between seat

faces, in.

b = thermal expansion

coefficient for valve

body, in/in/0 F

otg= thermal expansion

coefficient for gate,

in/in/0 F

AT = temperature change, OF

Therefore, the net difference in valve

body and gate expansion is:

A8 = o - i8b (positive for

interference)

= Lo AT (ag-xb ) (F.3)

The differences in body and gate expansion will generate a seat contact force if AS is

positive. Its magnitude can be calculated using the following relationships:
Valve body and gate deformations under seat contact force are:

AF

A~b =-AKb

AS = AF

g Kg

where A~b = valve body deformation under a seat contact force of AF, in.

A~g = gate deformation under a seat contact force of AF, in.

AF = seat contact force, lb

Kb = valve body stiffiness, lb/in

Kg = Gate assembly stiffiness, lb/in (including seat stiffness)

as = (, - 5b) =b + KAF =AF (Kb Kg

Kb Kg Kb Kg)

Th erefore,

AF (5g -b) ( Kb +Kg)

= Lo AT (5g - 5b) Kb+K (F.4)

F.2. EFFECT OF STEM GROWTH

For a wedge gate valve, the average stem temperature before closing may be lower than the

valve body temperature. After the valve is closed, the seat contact force may be increased

by either of two possible conditions:

F.2.1. Increasing stem temperature.

If the valve temperature remains the same after closing, then the stem growth may be

estimated as:

8S = ts ctS ATs (F.5)

where S5 = stem growvth, in.

Is = the stem length which is subject to an average stem temperature

change of ATs after the valve is closed,in.

0! = stem thermal expansion coefficient, in/in0 F

ATS = average stem temperature change, "F
F.2.2. Decreasing both stem and body temperature.

If both the valve and stem temperatures decrease after closing and the valve body

contraction is greater than the stem contraction, then the seat contact force may be

increased due to this differential body and stem contraction. The relative stem growth

may be estimated as:

8 s = lb ab ATb - IF (Xs ATs (F.6)

where lb = body length with average temperature change of ATb, in.

cxb = body thermal expansion coefficient, in/in/0F

ATb = average body temperature change, 'F

if = stem length with average temperature change of ATs, in.

as = stem thermal expansion coefficient, in/in/0 F

ATs = average stem temperature change, 'F

The calculated stem growth, ES, can be converted into disc lateral interference as

5g = 5s sin 0 (F.7)

where 5g = disc lateral interference (one side) due to stem growth, in.

0 = one-half of the disc wedge angle, deg

It should be noted that the above equation does not include stem or valve topworks

flexibility in the stem growth calculations. If the stem and valve flexibilities are

considered, only a fraction of the stem growth due to differential valve component

expansion/contraction will be converted into disc lateral interference. Net stem

growth that can be converted into disc lateral interference, including the structural

flexibility, may be expressed as:

8's = 5s. 8 sk

where 5'5 = net stem growth including structural flexibility, in.

03 = stem growth excluding structural flexibility, in.

°sk = structural deformation (stem and body) due to stem

compression, in.
F.3. EXAMPLE CASE

To illustrate the temperature effect on valve operating thrust, a simplified case from one of

our root cause analysis investigations is used in the following calculations.

Ll = 1.84 in. (gate width)

L2 = 1.25 in. (seat length, one side)

Isl= stem length outside valve body = 21 in.

Is2= stem length inside valve body = 3.375 in.

ab = 8.9 x 10-6 in/in/0 F (SA-182)

ag = 9.7 x 10-6 in/in/0 F (SA-351)

as = 6 x 10-6 in/in/IF (SA-564-630)

Kb = 71 x 1061b/in

K'g = 19 x 106 lb/in (gate stiffness only)

Kseat= 78 x 106 lb/in

ATb = 650 - 70 = 580'F

ATg = 635 - 70 = 5650 F

ATS1 = 175 - 70 = 105'F (outside valve body)

ATs2 = 450 - 70 = 3800F (inside valve body)

F.3.1 Combined Gate and Seat Contraction

Using the dimensions and coefficients of thermal expansion for the gate and the two seats,

we can calculate the combined thermal contraction from Equation F.2:

5g = 1.84 x 9.7 x 106 x (635 - 70) + 2 x 1.25 x 9.7 x 10-6 x (650 - 70)

= 0.02415 in.

F.3.2 Body Contraction

°b = (1.84 + 2 x 1.25) x (650 - 70) x 8.9 x 10-6

= 0.0224 in.

-F.4-
F.3.3 Net Gate Expansion

to = 0.0224 - 0.02415 = -0.00175 in.

Since, in this example case,the net expansion of the gate and the over the body dimensional

changes is negative, no seat contact load is generated due to differential thermal

contraction between valve body and gate alone.

F.3.4 Net Stem Growth

Using equation (F.6):

o, - [21 x 8.9 x 10-6.x (175 - 70) + 3.375 x 8.9 x 10-6 x (650 - 70)]

+ [21 x 6.0 x 10-6 (175 - 70) + 3.375 x 6.0 x 10-6 x (450 - 70)]

= 0.02093 - 0.03705 = 0.01612 in.

To account for the net gate width contraction from F.3.3 and F.3.4, the net stem growth is

5, = 0.01612 0.00175 1 =0.00612 in. 2 tan 50

F.3.5 Stem and Valve Topworks Stiffness

AE

where

A =-Cd2 = I x11252 in 2

4 4

6 E = 30 x 106, psi

L = equivalent length (including estimated valve topworks flexibility) 30 in.

.K=9.94 x 105 1b/in

Based on the above estimated stem stiffness, the adjusted net stem growth using the result

in F.3.4 and equation (F.8), is

I F.

o, =0.00612- K

K

where Fs = stem force = 2 AF (sin 0 + g cos 0) (see Appendix A, Section A.1.3)

AF = seat contact force

For a typical wedge angle of 50 and an estimated value of g of 0.4, the above equation gives

Fs = 2 (sin 50 + 0.4 cos 50) AF = 0.97 AF
Therefore, the adjusted net stem growth is

Ss =0.00612- 0.97AF =0.00612-9.76x 10- 7 AF

9.94 x 105

From this, the equivalent disc lateral interference caused by the adjusted net stem growth

can be calculated using equation (F.7) and is

AS = 2 83 sin a = 2 x(0.00612 - 9.76 x 17AF) x sin 5°

F.3.6 Seat Contact Force

Using the same approach as in Appendix E, the combined gate and seat stiffness is

1 1 2 1 2 -8

Kg Kseat a 19X 106 78 x 106

Kg = 12.8 x 106 lb/ in

The seat contact force, using the equation (F.4) derived in Section F.1, is

AF =0.1743(0.00612-9.76 x 10-7 AF) 71x 10 6 X 12.8x 106 (71x 106+ 12.8x 1061

= 11,553- L842AF

AF = 4,065 lbs

F5= 0.97 x AF = 3,943 lb -

This can be compared to the disc friction load of 6,505 lbs which is based on a coefficient of

friction of 0.4 and a pressure of 2,100 psi across a mean seat diameter of 2.85 inches. The

seat force increase of 3,943 lbs represents 61% of this disc friction load.

It should be pointed out that, the estimated stem thrust is very sensitive to the temperature

and stiffnesses of the valve components. The estimated thrust can vary significantly

based on a specific valve temperature profile and the actual stiffness of the flexible wedge.

The main purpose of this appendix was to show how to quantify the effect of temperature

changes on the opening thrust.
NRC FORM 335 U.S. NUCLEAR REGULATORY COMMISSION I. REPORT NUMBER

11401 (Assaned by NRC. Add Vol., Supp., Re.. VWICM 1102, a~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~wd Adclendum Nvjnberm. If anv.) Ml.202 BIBLIOGRAPHIC DATA SHEET

[See instructions on the reverse NIUREG/CR-5807

27. TITLE AND SUBTI% KEI 1721

Improvements in Motor Operated Gate Valve Design and Prediction

Models for Nuclear Power Plant Systems 3. DATE REPORT PUBLISHED

MONTH YEAR

SBIR Phase I Final jeport May - 1992 4.FIN OR GRANT NUMBER Septemnbr 1990-!April 1991 L1667

5. AUTHOR IS) 6. TYPE OF REPORT

J.K. Wang, M:S. Kalsi Technical

7. PERIOD COVERED (iclusive Oares)

B. PERFORMING ORGMN(ZATION - NAME AND ADDRESS tll VRC. provide DiVison., Office or Relion, U.& NuclearRegulstory Cormnhission, and mailing addren if con tracror, prowide

-w- w&adloig addyssJ

Kalsi Engineering, Inc.

745 Park Two Drive

Sugar Land, TX 77478

O. SPONSORING ORGANIZATION -NAME AND ADDR ESS lif NRC, type 'sme asabo"e is contractor. proide INRC DiWiion, Office or Region, U.S. Nasder Regulatory Comm-iion,

Sa,-- adiw&1

Division of Engineering

Office of Nuclear Regulatory Research

U. S. Nuclear Regulatory Commission

Washington, DC 20555

10. SUPPLEMENTARY NOTES

11. ABSTRACT f20 weon or sAm

This research is aimed at improving the performance of gate valves at nuclear power

plants (1) by developing improved models for operability prediction and (2) by

identifying improvements that overcome problems/limitations of the current designs.

Phase I research is aimed at developing improved operating thrust models for the

most common types of gate valves in use at U.S. nuclear power plants. Instrumented

valve test data provided by Duke Power Company will be used to develop/compare the

analytical predictions. Specifically, Phase I research will address shortcomings in

the current techniques by investigating localized contact .stresses under disc

tilting caused by fluid flow, by predicting inertial thrust overshoot, and by

providing a comprehensive review of friction/galling data for gate valves.

12. KEY WORDSlDESCR!PTORS List set phraws thatrwllassisttewarchert in locating the report.) 13. AVAILABILITY STATEMENT

gate valves unlimited

prediction models 14. SECURITY CLASSIFICATION

valve test data (This Pag)

friction/galling data unclassified

'This Report)

unclassified

15. NUMBER OF PAGES

16. PRICE

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