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ennieoin usa ,

A MODEL FOR PREDICTING THE PERFORMANCE OF SPRING-LOADED SAFETY VALVES

Avtar Singh

Electric Power Research InstitutePa'o Alto, California 94303

Arthur M. HechtMilton E. Teske

Continum Dynamics, Inc.Princeton, New Jersey 08540

ABSTRACT D pipe diameter

f friction factorSpring-loaded self-actuating safety valves areemployed as part of the overpressure protection systems F back pressure force on the valve disc

Bin various industrial applications. In order to designF upward disc force

and predict their perfomance it is required to study D

the valve dynamic behavior over a range of fluid and h enthalpy of steamsystem conditions. It has been observed that certain k spring ratespring-loaded safety valves show unstable dynamic be-havior known as chattering under subcooled liquid and ( correlation coefficient for back pressurevapor flow conditions. This behavior is the result of m mass flow rateacoustic coupling between the valve motion and pressurewaves propagated into the upstre?m piping. Probabi. M mass of valve disc and moving partslity of this occurring increases with length of upstream p pressurepiping. A one-dimensional theoretical model has been valve exit pMssuredeveloped to study tne ef fects of dif ferent valve and pgsystem parameters influencing valve dynamic behavior.

E valve body bowl pressureThe model is capable of investigating the effects of cvalve geometric and dynamic characteristics and up- pf pressure on lip area of guide ringstream pipe length and diameter on perfomance under p valve upstream stagnation pressuresteam discharge conditions. Valve geometric parameters o

p ,g set or opening pressum of Oe vaheinclude valve dimensions which control flew areas and 3disc and seat forces. Back pressure area and magnitudet timeeffects are included. Adjustable parameters such as

set pressure and adjusting ring effects are accounted- U, velocity of steamfor. The analytical and empirical foundations of the Y '0I "'valve model are doc e ented here and comparisons are

x valve stem displacementmade between simulations run using the model and fullscale steam tests of safety valve types used in Pres-

8 coefficient of restitution|

surized Water Reactor safety discharge piping systems.y isentropic exponent of steam

NOMENCLATURE 8 angle between the horizontal and unit vectorgnomal to A,

a sound speed

A, valve exit flow area Fig. 2 0, en between the horizontal and the exit, j

A lip area of the guide ringf a steam densi y

A* choked or minime flow area in the valveINTRODUCTION

A norrie flow areayA seat top surface area High pressure vessels in various industries are

protected against overpressure by self actuating spring-C damping coefficient of the valve stem motion loaded safety valves. The valve opens when the upstreamC flow discharge coefficient of the valve pipeline or vessel pressure exceeds a certain allowable

d value, thus releasing the fluid to maintain the system'

8407130214 840423PDR FOIAJORDAN 83-765 PDR

I.

.

i,

below a certain design value. In order to otherwise be gained using real steam flow, i'design the overpressure protection systen for Ii

transients, it is necessary to understand the VALVE DYNAMIC MODELbehavior of these safety valves. For example, a

ning valve may not release sufficient quantity The cross-section of a typical safety valve is showr.f to control the rise in system pressure and main- in Figure 1. The main functional components of the valveunder the design value. On the other hand, a include the nozzle, seat, disc, spindle, spring bellowsning or cycling may cause substantial dynamic and adjusting rings. The spring holds the disc, at-on the discharge piping. In any case, it is tached to the spindle, against the no221e seat. If

mous to have available a method to predict safety system pressure exceeds the valve set pressure, the discefonnance. is lifted off the seat and an additional force due todate, little work has been done to analyze the steam flow through the seat annulus and deflection of ;

behavior of a safety valve and its interaction the flow by the adjusting rings pushes the disc furtherprotected system. In most cases it is assmed upward. A properly operating valve opens to maximm ,

~se valves actuate within a certain opening time lif t and is held there while blowdown proceeds. Decreas-upstream pressure exceeds an adjustable value ing system pressure allows the disc assembly to leave

|ne Set Pressure. When the valve is fully open, maximm lift at some point and lift of the disc decreases;;ected to allow a desired flow rate and when the unbl the design blowdown is reached. Ihe valve shouldupstream of the valve decreases by a designated then close rapidly and decisively. Initial lif t and

lled the Blowdown the valve is ass med to close. closing behavior and percent blowdown is regulated by

||order to design and evaluate the performance the adjusting ring settings. These rings adjust theristics of a safety valve it is necessary to angle of steam flow with respect to the flow leaving the |

.nd its dynamic behavior under different fluid seat region.modynamic conditions. The objective of this Bellows are designed to perfom two functions; j

to develop such a coupled thermal hydraulic and firstly, to isolate the back of the disc from the back |

~ ass system model that can be utilized as part of pressure existing in the body bowl and secondly, to pre-!'ell system interactive model. vent leakage of fluid into the atmosphere. It should be

t of the existing theoretical work related to noted that the bellows do not cover the entire area atnvolves analysis of steady state mass flow rate, the back of the disc. However, a part of the area loc-.ent piping sizing (4) and transient discharge ated around the outer periphery of the disc may be ex-is (5). Fowler et al. (6) attempted to simulate posed to the back pressure,ty valve dynamics considering only the mass-ffect on pressure surges in a heat exchanger. Control Volme Analysisconsidered the interaction between the inlet The dynamic model of valve behavior is based on ,

aid dynamics and the spring mass system of a pop- one-dimensional fluid flow through a control volume in.e to analyze the valve dynamic stability. Ray the nozzle and seat region es shown in Figure 2. A'ulated a nonlinear, semi-enpirical model of a momente balance then gives the force Fo pn the valve.alve and dynamic equations were derived from disc in tems of the fluid mass flow rate me , spring.tal principles of rigid body motion and fluid force Fspring , back pressure force FB , seat force

The analytical model developed in the present Fseat , and the entrance and exit flow properties and |.

based on a similar approach. A resultant disc velocities. The disc force is dependent upon the angle, derived based on a control volse momente and of deflection se , increasing as ee increases. Ring41ance between the valve inlet, disc, seat and adjustment is denoted in " notches" (axial grooves in !

t to the body bowl. A one-dimensional variable- adjustment or guide rings used to rotate rings). The '

x approximation has been adopted for this region axial position of the guide ring with respect to theth several other simplifications. The highly seat is determined by the neber of notches per revolu-flow patterns that probably occur preclude a tion of the guide ring and the pitch of the threads ofct analysis, patterns that may include separa- the ring, both fixed geometrical details for any givenshocks in a two-phase, compressible, three- valve.

nal, quasi-steady flow. The model. which The fluid force on the disc, Fo , is obtained fromonly the required basic physics, is calibrated a moment e balance for the control vol s e shown by

11y by leaving free one parameter which is set dashed lines in Figure 2. With the valve disc in an;ng measured perfomance for any particular open position the momentm equation in integral fom isfel. (Geometric and/or familial similarity be-sted and untested valves controls confidence in

g e value of this parameter from the formerg )dA (1)cV U)n dA +pU,dV = - nog j

net disc force derived from the control volse yis utilized in a spring / mass /darper system to

where p , Ui and ei are the fluid density, velocitythe valve stem motion given the reservoir pres- and pressure / shear str sses, respectively. For quasi-the are ha cnr ay be steady flow the integral on the lef t side of Eq. (1) is35 functions of flow geometry and valve charact- neglected.. the flow within the valve is quast-steady; Neglecting shear, the total force acting on thehigh quality and may be represented by the valve disc is obtained by evaluating the pressure and

ic perfect gas approximation; one-dimensional flow variables at the control volume surfaces to yieldustics. The flow may be considered quasi-ste*dye valve internal dimensions are small compared F0 * epi *Pi N* e ' SI" ' ~ IB*ISPd"9Unaracteristic pipe acoustic wavelengths. Theseths are much larger than the pipe diameter,19 one-dimensional acoustics. The simplifica- +Fseat + p A sine (2)ee A1evable using the isentropic perfect gas assump-outweigh any increase in accuracy that would

2

T

*.

.

where AN is the inlet nozzle area. The back pressureforce FB is the product of a participating area behindthe disc which is outside the bellows and the pressure u = [2(h - h)]I (5)g9acting on this region. This back pressure strongly in-fluences the valve perfomance and yet is highly design-dependent. It has been found that the valve models of Back Pressure Modeltwo different manufacturers, each analyzed as part of The valve dynamic model development described herethis study, required different foms of back pressure relied on a series of tests of full scale safety valvesmodeling, as will be discussed further along. to provide calibration of the model and to validate code

The seat force Fseat consists of the area inte. predictions. These tests were carried out in thegrated pressure over the seat annulus and lower ring and Combustion Engineering / Electric Power Research Institutein the case of some valve types, a pressure force over a steam test facility with an initial reservoir pressureportion of an adjusting ring having area Af The seat a 2500 psi (17.2 MPa) Saturated steam. The tests were;

'

force contribution from the pressure acting over the carried out on five valve models of two dif'erent manu-annular seat 1s modified by a seat force adjustment fac- facturers. For purposes of identification, these willtor n which is the free parameter of the analysis as be denoted as A, B and C valves of Manufacturer 1, anddiscussed previously. valves D and E of Manufacturer 2. Table 1 sumarizes

sorte of the characteristics of these valves. It was'.

Fm t = n pdA + A pf (3) TABLE 13 f

Spring-Loaded Safety Valves Used In5 Perfomance Simulations

The flow in the region of the seat is highly curved asit exits the nozzle, negotiates the seat region, andexits between the lower and middle or upper adjusting valve Asmt astea mors'e man tew Inlet ai

M08'i M * "'''' """ ** *'d W " 0*"''**'rings. Thus, n accounts in some measure for the ac- 2

releration-induced pressure decrease in the seat region, _ IW Ov5") w2 %) ,.. M in.

and we should expect n to be less than 1.0 . A 212.182 (26.734) 1.e41 0188) o.3s2 (e.7) 3a6

The flow discharge angle B is shown in Figure 2and is defired as the angle between the horizontal and e 420,006 (52.920) 3.644 (23st) 0.538 (13.7) 6m6e

the direction of the velocity of discharge Ue . 08 C 504.952 (63.623) 4.381 (2826) o.591 (15.0) 6s8is obtained geometrically as a function of the lift andthe guide ring setting. Lowering the guide ring in-

o 297.845 (37.528) 2.545 0 642) o.45 01.4) 26 a 6creases Se and thus increases the lift force. '',

#*"'Critical Flow Model

The compressible effects of the steam flow through *"'t"*'"""''*pat,a n.

MJ the valve have been based on an ideal gas model with a

ratio of specific heats y = 1.26 . The one-dimensionalflow area through the valve was obtained by graphical

| construction based on axial symmetry. Flow area at the found that the back pressure characteristics of valvesseat thus becomes a function of valve lift. This shift A to C differed considerably from those of valves 0 and'

of minime area is illustrated in Figure 3. For this E. The blowdown of the first three valves was found to; particular valve the minime area of flow shifts from be independent of back pressure (except in the very;

the upstream nozzlr: region to the inner radius at the highest back pressure range) while the last two were*

? valve seat as the valve lift decreases below 0.503 in. back pressure sensitive. The internal design of the(128 nun). This change in choke position can have an first three valves included an eductor casuunicating'

effect on the disc force. The area versus lif t charact- with the back pressure chamber, which suggested thateristics must be generated for each valve analyzed. the back pressure area of these valves were at ce close

| The mass flow rate is calculated based on the isen- to the critical pressure throughout each blowdown. The,

|tropic perfect gas flow equation last two valves had back pressure that were pressably

closer to the body bowl (or chamber) pressure. This

~L+1

chamber pressure was partially a function of the down.y-Pi stream pipe geometry (due to frictional effects in the

system) with back pressure set by a valve downstream of5 = C A* Yp D [ 2 (4)d oo +1 the valve. It was found that the valves D and E the

chamber pressure correlated as- .

where A* is the minime area of flow reckoned from theK ,m (6)nozzle to the exit area (Figure 3). The position of the p ag

minimun area can shif t to the exit plane for the valve, illustrated in Figure 3, since the effective exit area of where m is the critical mass flow. It was, therefore,

flow for this valve is Ae cose Fo* the valves an- possible to derive Km for each test run of the appro..

he s rge coeff cient Cd at full lift priate model D or E valve since both mass flow and cham-4t ber pressure were available fra the test data. For

k The area profile along the valve yields local pres- valve E, Figure 4 presents a plot of pc versus m forL sure, temperature, and density since these are functions s m ral ts

j only of A/A* for an isentropic perfect gas. The localvelocity can be calculated knowing the local temperature bY ca ulatin9 the chamber pressure pc for a specified3 (or enthalpy), since in the absence of heat transf r mass flow rate, if pc can be related to pn , the back(assumed) the stagnation enthalpy is constant and ffor

,

l pressure at the valve extt. Continam Dynamics, Inc.the proper units).

3.

J

r --

*

.

.

has develooed a caputer code to predict such back pres-sures (9.10). The code provides pressure and flow con. TABLE 2

ditions along discharge piping systems including the Seat Force Reduction FactorCf f' cts of two-phase flow, copressibility, shocks, andfriction. Losses due to pipe fittings are included.Using C.E. test data, a correlation relating pc to pB Vahe Model nfor valve models D and E has been developed:

A 0.14

P = 0.712pg + 0.173p , (7) B 0.26g

C 0.35Thus, it is possible to specify N for any dischargegsystem. D 0.44

During the pressure transient which occurs as thevalve opens, there is a significant time delay for the E 0.44chamber pressure to reach its maximum value. This isdue to a downstream " fill up" time required and to

"

o t o ti on way u th n ra in I t pre p os bu t

completely. They drop to approximately one half liftand then decrease gradually, falling out of lift andclosing at about one-quarter lif t. The mechanism respon.

df , ;in , g (g) sible for closure, therefore, appears to be differentdt out for each set of valves.

Closure of valves A to C appear to be associatedwhere the control volume V is an effective downstream with transfer of the sonic line as the minimtsn flowv:Ime which accounts for condensation effects. Using area shif ts to the seat region with decreasing lift. A

lift force reduction then occurs over the seat areasufficient to permit the valve to close completely,

2 , ygpjp (9) This force reduction occurs for valves D and E but isa not large enougn to oermit closure. This is the initiallift reduction seen in tests of the valve. Complete.

(10) closure appears to occur when the lif t decreases to thepc= (moutextent necessary to generate large friction losses inthe minime flow area in the seat region. These fric-

it may be shown that tion losses reduce the local Mach neber over the entireseat to Mach one, at which point the seat unchokes and

pc = b [1 - e-t/t) (g}} the seat pressure falls to a level close to the backApressure. This reduction in force permits the stem anddisc to fall out of lift and the valve closes. Since

where the time constant t is the seat area friction coefficient is not known and theflow was not calculated in this detail in any case, theempirica11 determined closure at one-quarter lif t wasused for afl calculations presented here for valves D, , T9Y m (12)

,2 and E.

UPSTREAM PIPING ACOUSTICSThis delay permits the valve to open fully and remainopen in both the actual system and in the model. If the The effect of upstream pipe acoustic coupling, inback pressure increase is too rapid it is possiblethat the valve could close imedfately after reach- which pressure waves in a pipe upstream of the valve

interact with the valve behavior, was made p rt of theing full open. The time delay as we have discussed it valve dynamic model in order to investigate the role ofhere is observable in the C.E./EPRI test data. these interactions in instances where valve chatter

might occur.Seat Force Rduction Factor Asstaning that one-dimensional flow is valid, the

The free parameter n was determined for each momentisn and continuity equations Jvalve model by calibrating the dynamic model once foreach valve and using the resulting value of n for sub- -L

sequent calculations for each valve. The values of n au au 1 dp 2fluluII3)that result from this calibration is shown in Table 2 H + u g = p g, - Dfor all the valves analyzed,

n is constant for the D and E valves, which aregemetrically similar, i.e., their linear dimensions 3 3are in the ratio of approximately 4/3. Valves A to C f+fu=0 (14)are not gemetrically similar. and the values of nconsecuently are different,

h p f hWValve Closure Mec'hanism friction coeffient and D is the diameter of the up-

The stem position histories of valves A. 8 and C stream piping. These equations are solved using adiffer markedly fran those for valves D and E. The MacCormick 2 step explicit scheme,first three resin at full lif t and then decrease nearlylinearly with reservoir pressure until they close

4

-

pressure on water is not a reliable indication of theVALVE DYNAMIC EQUATION existing preload on the spring, the runs were made for

The dynamic valve calculations are made by employ- the nminal value. The runs for the remainder of theing a quasi-steady state disc force derived from the valves were made using the actual, observed openingcontrol volse analysis in the equation defining the pressure as the set pressure. The percent blowdown wasvalve disc, spindle and spring assembly acceleration. calculated for both the data and for the simulation asTime integration of this equation yields the valve lif t follows;

for a given reservoir pressure history. The disc force For valves A to C,is exactly balanced by acceleration of the moving inter-i nals of the valve as well as by damping such that 2500 - p

1 Blowdown = 5' x 100 (17a)2435AMi + Cx + kg = A IPt*Pset) + n pdA, + pf f8

N For valves 0 and E,s

ii+mU sine + A p sine (15) 2515 - pej ' x 100 (17b)u-FB' ee e ee A 1 Blowdown = 2500

When the valve disc is on the seat or at the upper stopa coefficient of restitution B is used to reverse the Figures 7 and 8 are plots of percent blowdown for

valves D and E, which are the two ge metrically similarspinale direction valves discussed here. The trend of the data with ringsettings for both these valves and in fact for valve B.

x, = - 8x (16) Figure 6, agree well with the computer simulations.. .

Note that the lower ring settings varied somewhat amongwhere the prime denotes conditions after contact with the runs shown. The actual lower ring settings were

used for each run although not indicated in the figures.the valve seat or upper stops.For the calculations presented here, the dam;,ing Figures 9 to 13 present measured and calculated

coefficient C was set equal to 20% of the critical stem position for one test of each valve, asstr1ing zerovalue, while 8 was taken equal to 0.01 . upstream pioe length. (Neerous additional calculations

for other test runs were made with similar success.)COMPARISON OF SIMULATED VALVE PERFORMANCE WITH TEST DATA

Reservoir pressure and mass flow rate are included fortest run 411, valve A. Figure Sa presents a comparison

The valves discussed in this analysis were tested between the measured and calculated stem position andfor operability for the range of fluid conditions under Figure 9b shows the comparison for the mass flow rate.which they may be required to perfom. These tests Figures 10 to 13 are the remaining stem positionwere carried out in the C.E./EPRI test facility. The plots for valves B to E, respectively, each asseing

zero pipe length. The last two figures illustrate thefacility was designed to accorriodate tests for bothsteam and water flow as well as a loop seal upstream of coupling ef fect between the valve and upstrea9 acoustics.the test valve. Figure 5 illustrates the valve and loop Figure 14 shows calculated stem position versus time forseal arrangement used. The n minal valve set pressure test run 201, valve E, ass eing zero upstream pipefor the tests was 2500 psi (17.2 MPa). The valve code length a.1d for middle and lower ring settings of +34 anddescribed here was run for many of the steam flow tests -20 notches, respectively. Although the stem does notto first calibrate and then validate the valve model remain at full lif t af ter opening (due to the reductionpresented in the previous three sections. The tests in lift force associated with the positive middle ringwere run for a range of adjustment ring settings, reser- setting) no instability is predicted. When the upstreamvoir pressure ramp rates, back pressure settings, and pipe is included. Figure 15, the valve first opens fully,upstream piping lengths. The rmainder of this paper is then falls out of full lift and goes into a dynamic in-devoted to a comparison between measured perfomance and stability. This process is due to the coupling betweencode simulations for a neber of the test runs for the the upstream pipe acoustic waves and the valve dynamicvalves we have been considering. response. The observed valve response for the 201 test

Figures 6, 7 and 8 present measured versus pre- is shown in the inset of Figure 15. Both the actualdicted percent blowdown for valves B D and E, respec- valve motion and its simulatten reach full lif t and re-tively, made for zero pipe length. The Valve Dynamic main there briefly before closing and then oscillating.Model (VDM) code results shown in Figure 6 fur valve B The characteristics of the valve motion and the oscil-blowdown were generated for a set pressure of 2500 psi lation frequency are well represented by the valve(17.2 MPa) and a lower ring setting of -ti8 notches. dynamic model simulation.Percent blowdown is plotted as a function of upper ring,lsetting (Note that the adjusting rings which direct the CONCLUSICMS AND O!$CUSSION

exit flow are denoted as the upper rings on valves A toC, while they are the middle rings on valves 9 and E. A one-dimensional valve dynamic model has beenThese last two valves have an upper ring which adjusts developed considering the effects of various valveports connecting the back pressure region of the valve design parameters that influence the valve dynamic re-to the body bowl. These rings were set equal to -46 sponse and stability under different upstream and down-notches throughout the tests on both valves. The lower stream conditions. Comparisons between actual andrings on all the valves serve primarily to set the pop- simulated valve perfomance have shown that the model ispingactionatopening.) capable of predicting the major trends in valve perfor.

Most of the tests shown on Figure 6 were made for mance, including unstable response due to coupling be-an upstream water seal with the exception of tests 903 tween the valve dynamics and acoustic waves in upstreamand 1411. Consequently, the opening pressure varied piping. The model presented here can be used in severalconsiderably frm the nminal set pressure of 2500 psi ways, such as providing an assessment of perfomance for(17.2 MPa) for most of these tests. Since the opentrg valves which have not been tested, for evaluating an

$

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.

.

ienvelope of stable valve behavior for a range of ring I

settings and upstream piping length, and as a tool use-ful in developing improved valve designs. 7ACKNDWLEDGEMENT

This study has been performed as part of the EPRl/ f" [ IPWR Safety and Relief Valve Testing Program sponsored by p ja

the Electric Power Research Institute and participatingPWR Utilities. Funding for Continuum Dynamics, Inc. in e ,

,;*this effort was supplied by the Electric Power Research|Institute. m

|,

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yREFERENCES

a- m ,

1. Thompson, L. and Buxton, 0.E.: " Maxim e Isentropic W , , , , , . , ,

Flow of Dry Saturated Steam Through Pressure Relief - W

*^ ~hValves," presented at the Pressure Vessel and

Piping Conference, Montreal, Quebec, Canada, July1978.

2. Sallet. D.W.: "The Flow of Liquids and GasesThrough Pressure Relief Valves," presented at the Figure 1. Cross section of a typical safety valveThird National Congress on Pressure Vessels andPiping, San Francisco, California, June 1979.

3. Sale, J.W.: " Safety Valve Mass Flow Rate Calcula- i, .' " ' ' " ". ;"''tions and Cctrection Factors," presented at the g

Third National Congress on Pressure Vessels and jPiping, San Francisco, California, June 1979 //

4. Liao, G.S.: " Analysis of Power Plant $afety and' ' * * 'Relief Valve Vent Stacks," A$ttE Transactions. '

/ *% -

"sJournal of Engineering for Power, October 1975.pp. 484-494. f {. ,' C'

' ' ' "5. Moody, r.J., Wheeler, A.J. and Ward, M.G.: "The i

/ 6,Role of hrious Parameters en Safety and Relief n-,

', J' '"*Valve Pipe Forces," presented at the Third National,e / /t N . ' '" |*

.'

Congress on Pressure Vessels and Piping, . . , ,

' '"', ,,,f, A N**"* *'~San Francisco, California, June 1979.6. Fowler, D.W., Herndon, T.R. and Wehrmund, R.C.: ', I L*".:. , '.

' ' ' '" ,,"/j',qJ"An Analysis of Fotential Overpressure of a HeatExchanger Shell due to a Ruptured Tube " ASME . . . . . . | g,. . . . .' CO

" " * " 'Petrolem Engineering Conference Preprints, (September 1%8.

7. Funk, J.E.: " Poppet Valve Stability," ASME Trans.actions, Journal of Basic Engineering, Paper No. Figure 2. Control volune of the one dimensional fluid62-WA 160. flow model

8. Ray A.: " Dynamic Modeling and Simulation of aRelief Valve," Journal of Simulation, November1978, pp. 167-172.

L/)9. Quackenbush, T.R. and Bliss D.8. " Preliminary .

Report on Computational Simulation of $teady Two.Phase Flow in Safety Discharge Valve Piping Systems,Rev.1," Continuun Dynamics, Inc. Tech Note 80 2,September 1980. [] , o'{j'"

l10. Hecht. A.M., Teske. M.E. and Bliss D.8.; " Quasi- r uSteady Back Pressure for Pressurized Water ReactorSafety and Relief Valves," Vols.1. !! and !!!, 'r., *_+

Continuun Dynamics, Inc. Report Mos. 81-1, 81 5and 82 1, prepared for participating PWR Utilitiesand Electric Power Research Institute March 1982. j g

11. Wolpert, M.J., 111: "EPRI $4fety Valve Test ProgramGeneral Test Request Transnittal " CombustionEngineering Power Systems, P5A 81-260. November |

' " * ' *,,

1981. 9,

! U \,.."./- . . .

1.,4 .

, , , . , " ,-. , . , ,

.

Figure 3. Area of flow vs. station for a particularvalve model1 in. = 25.4 m

6'

k

.

. .

.. .

. M.".T ".*. . . , . ,,

. . N. .._,msw == . w

/, -;;; :: jeite se -. ..., , ,

:tF*' * * ** g, - . .. . . . u . . . .p,

*.soo ' ' *in.. eso. ;ife it.,uc,''

'.- ,. ..

" " ' ' " " * * * " " * *Figure 4. Correlation of chamber pressure as a function;'

of steam flow rate for valve E Figure 8. Valve E tiowdown as a function of middle ring1 psi = 6894 Pa 1 lb/sec = 0.454 kg/sec setting: n.1ata; 0 VDM code results

tvet e., 0 0* -

seem %,awe wemeni

m'attst w t --.e.....

......, ,' ..... ..4* w 4* sen. se _ nosf,0 -* g

""/ ..'n5p _ , ,

v'e'.e usi+ .emc.c o t -

vtNtumi n swe. era g.-

|'e'sen iso imumi - - =

) $ |~

tama Noi " ooi -

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Looe' $tAL |

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Figure 5. Schematic of a typical safety valve test oo to 5 *o 50inlet configuration and instru.ientation ag,g ,,g ,0

1 in. = 25.4 mFigure 94. Valve A stem position for test run 411:

L[IP.14 notches 55notchesi=0tn =.

n t f t = 304.s msa

a ico ,

d ieI gy, Z . ". f" n '' ' y* * * . . . . . , p "* '"""*" ge esa

g eo . ' * . . . . " -* d, . .. . . . . . ,

s - weiea =an= uet=e g |r eo - !

|Figure 6. Valve B blowdown as a function of upper ring *

notch settingt m datan 6.VUN code results I |'' '

'

eU U * 7m',',.'.n,

~*o lo

a o ,,, ,,,8,0.. . * * ' . , , . . o to,

I ,, '

*P""",,.. Figure 9b. Valve A rass flow rate for test run 411.

*a .

*,. = = = , , " " . . 1 lbm/sec = 0.454 kg/sec*==

" . '

%s =. A*.e. .. " "

Figure 7. Valve 0 blowdown as a function of middlering notch setting: odata6 9.V0H coderesults

1

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\"'oo. ---- .- ......... , , , , , , ,'

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!,

Io ''o . .o is ao as o s to is to as sovint escen t =< se ri

Figure 10. Valve 8 stem position for test run 1411: Figure 13. Valve E stem position for test run 603:L,q,, = 0 ; n, = -77 notches n gp p . O i g = 60 notches ;g = -18 notches a p ,g = 2410 psian

g = C notches s p ,g = 2505 psiag ng1 ft = 304.8 m 1 f t = 304.8 m

o os. . . . . . . . . . . . . . . . . , ,,

oos -

-

oo4 -

' . , .=, w te= "S '

c' ,,,,

ct h ooe -

5 "> 6m

$ |oos- -

'E

oos -a

y y$ g oot '

I =ooi :

. cos

!

'o 'o, io is to as os os os os ostist istc> tist istcl

Figure 11. Valve C stem position for test run 1205: Figure 14 Valve E stem position assuming zeroL =0 n = -75 notches & length upstream pipe, test run 201:g

= .18 notches p ,g = 2470 psia % * '34 " I' b5 I "L = .20 notches 6g

1 f t = 304.8 m Pset = 20 notches. I f t = 304.8 mo os ,

,

o 04 , .am,

: m . ..a=g oos b* g ''' '

i '%,N f ; f~se-

,;' 3t

3

,, oon. . . . . , , g

p. .r, . ,3 .

. . . .,

i,,, ,,, , . .._

oo, . ....t

~ . ... . \f i'\. "

oo io so so ao so

taw estci }-,

= %e u1%* ** ,'

Figure 12. Valve O stem position for test run 310:Figure 15. Valve E stem position including 14 ftLpipe = 0 t n, e +1 notch I length of upstres pipe, test run 201:

g = 9 notches n p,,g = 2557 pela % = M mcw , ng = .M mcmn

I f t = 304.8 m Pget = 2484 psia. 1 f t = 304.8 m,

8