Development of Design Procedure for Gasketed Pipe Flange ...

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DEVELOPMENT OF DESIGN PROCEDURE FOR GASKETED PIPE FLANGE CONNECTIONS

BASED ON ALLOWABLE LEAK RATES

Satoshi NAGATA Engineering Technology Division Engineering Management Unit Toyo Engineering Corporation

Narashino, Chiba JAPAN E-mail: [email protected]

Yasuro OYAMA

Bolted Flange Joint Committee Japan Pressure Vessel Research Council

Kuwana, Mie JAPAN

Toshiyuki SAWA

Professor Department of Mechanical Engineering

Hiroshima University Higashi-hiroshima, Hiroshima JAPAN E-mail: [email protected]

Takashi KOBAYASHI

Professor Department of Mechanical Engineering

Numazu College of Technology Numazu, Shizuoka JAPAN

E-mail: [email protected]

Ryou KUROSAWA

Yokogawa Electric Corporation Kofu, Yamanashi, Japan

E-mail: [email protected]

ABSTRACT

The development of a design procedure for bolted gasketed pipe flange connections based on allowable leak rates has been carried out by the technical committee for sealing technology of pressure equipment organized in the High Pressure Institute of Japan. This procedure covers the design of flange connections used at room temperature. The point preeminent of this procedure is that the maximum allowable leak of the flange connection is obtained by the minimum required initial bolt preloads determined by the elastic interaction analysis of flanges, bolts, and gasket considering thier flexibilities and stiffnesses. The paper explains the design process of this procedure and gives the technical background. The paper also shows some of examples obtained through the validation of this design procedure and the design studies for actual applications.

INTRODUCTION Bolted gasketed pipe flange connections are widely used

in refinery, chemical and power plants as well as the other industrial plants. The sealing performance of pipe flange

connections is important in order to achieve proper and safe operation of the plants. Furthermore, attention is increasingly given to leakage from the gaskets of pipe flange connections from environmental points of view. Therefore, the importance of achieving good sealing performance is one of the main criteria for the design of flange connections nowadays.

The behavior of bolted pipe flange connections with gaskets has been investigated by many researchers [1,2]. Several researches in the past decade clarified that reduction of the bolt load and gasket contact stress are essential to describe the sealing performance of flange connections subjected to internal pressure. Thus, in designing flange connections subjected to internal pressure, the bolt load change has to be taken into consideration for determining the operating gasket stress required to achieve the desired leak tightness. This approach seems to be well accepted by most designers as well as the researchers in this technical field.

Currently two different procedures of the leak tightness based flange design method are available. One is the PVRC method[3] and the other is EN1591-1[4]. The former was an epoch-making work that introduced the tightness parameter to

Proceedings of the ASME 2013 Pressure Vessels and Piping Conference PVP2013

July 14-18, 2013, Paris, France

PVP2013-97861

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handle leak tightness quantitatively. The draft standard is under ballot for adoption by ASME and remains unpublished after over 10 years as a draft. The latter is only the leakage based flange design standard currently published. This method is improved compared with conventional methods like the ASME Code[5] on the point that the bolt load reduction is calculated by using the stiffness of flange. However, some problems such as insufficient gasket data, mismatching of tightness levels targeted in the gasket leak test and the flange design are pointed out.

Considering the above situation, the Sealing Technology Of Pressure-equipments (STOP) Committee organized of the High Pressure Institute of Japan (HPI) developed a new comprehensive design procedure for the bolted gasketed pipe flange connections based on the leak tightness[6]. In the previous PVP conference, the paper introduced this leak tightness based design procedure for the gasketed pipe flange connections prepared by the STOP committee[15]. The present paper provides the latest procedure incorporating some updates that is ready for issue as a HPI standard. And the verification results are also shown that the design procedure is applied to some of standard flanges.

SCOPE OF FLANGE DESIGN PROCEDURE The present design method covers the bolted gasketed pipe

flange connection subject to internal pressure. It is also applicable to flange joints for pressure vessels. External forces and moments, as well as the effects of thermal expansion difference among the parts of flange connection are not currently considered in the procedure. The sheet gaskets and spiral wound gaskets are applicable but the metal gaskets are not applicable.

EXPRESSION OF GASKET PROPERTIES

Sealing Characteristics The present design method adopts the assumption[7,8] that

the gas leak rate from a ring gasket subject to internal pressure is proportional to the inner diameter of gasket di and is inversely proportional to the width of gasket (dodi)/2 as shown by Eq.(1).

2

i

o i

dL

d d

(1)

On the basis of this assumption, the leak rate of a gasket L can be represented by the following equation[8];

11 s s

o i

L L kLd d

(Pam3/s) (2)

where Ls is the specific leak rate of gasket. Leak rates L for a particular size of gasket are corrected using the following gasket dimension correction factor k;

11o i

kd d

(3)

The correction factor k equals to 1.0 when the outer diameter do is twice the inner diameter di. Once the specific leak rate Ls is identified for a size of gasket, the leak rate of another gasket with same material but different size can be estimated by using the factor k and the specific leak rate Ls.

The specific leak rates have a close relation with the gasket deflection for several sheet gaskets and spiral wound

gaskets as reported by Kobayashi[8]. The present design method considers that the relationship can be approximately expressed by the equation;

1.5

*d

S

PL ce

P

(Pam3/s) (4)

where c and d are the constants determined by the gasket leak test according to JIS B 2490[9], is the gasket deflection, P* is the test pressure, P is the operating pressure, respectively. Ls is the specific leak rate and the leak rates are expressed by volume leak rates with the unit Pam3/s.

Furthermore, the present design method expresses the leak tightness of the gasket by using the leak rate per unit length at the outside perimeter of the gasket contact area. The leak rates are measured by volume leak rates of Helium gas.

Deformation Characteristics It is well known that the sheet gaskets and spiral wound

gaskets have rather complicated deformation characteristics, so-called loading-unloading curves. The curves show nonlinearity and hysteresis. Kobayashi[8] also reported that the loading-unloading curves of those gaskets can be represented by simple exponential curves based on the regression of experimental results. The present design method assumes those experimental formulas of Eqs.(5) and (6) can describe the loading curve and unloading curve of gasket, respectively. Ab

a (MPa) (5) Bb

F (MPa) (6) where a, bA and bB are the constants obtained by the gasket leak test according to JIS B 2490. F is calculated by the following equation; 1B A B Ab b b b

preF a

(7)

where pre is the stress at which the unloading started. Although a load-displacement analysis of the pipe flange

connection subject to internal pressure is necessary in order to predict its leak rate, further simplification on the loading-unloading curves is considered in the present flange design method using linear approximation as shown in Fig.1. The loading curve is represented by the line connected between the origin and the point A. The unloading curve is also represented by the line connected between the points A and B. The point A is the point on the loading curve where the gasket stress equals to the average gasket stress under the gasket seating condition, in other words, the bolted up condition. The point A is also the starting point of unloading. If the average gasket stress is less than the maximum testing stress max in the step 8 of the gasket leak test according to JIS B 2490, the max should be used for the Point A. The max is specified as 40MPa for sheet gaskets and 100MPa for spiral wound gasket, respectively. The point B is the point on the unloading curve where the gasket equals to the stress 11 in the step 11 of the gasket leak test according to JIS B 2490. Here, the 11 is 0.125max.

Gasket Property Parameters The parameters representing gasket properties for the

sealing characteristics, as well as the other mechanical characteristics are prepared to be used in the present flange design method for several types of gaskets. Table 1 shows those gasket property parameters. The data shown in the table are



derived from the JIS B 2490 gasket leak tests which had been carried out by several gasket manufacturers under the auspices of the STOP Committee.

REQUIREMENTS FOR SEALING PERFORMANCE

Allowable Leak Rates The present flange design method requires that the

predicted leak rates from the pipe flange connection is within the allowable leak rates per unit length as shown in Eq.(8). 31 10a (Pam3/s/m) (8)

This allowable leak rates per unit length was determined as the leak rate were no indication of bubbles forming during a soapy bubble visual examination based on the data obtained from the gasket leak tests which had been carried out by several gasket manufacturers under the STOP Committee. Be noted that the criterion is based on the value per unit length at the gasket outside perimeter not by total leak rates from the whole gasket. Thus, allowable total leak rates from a whole gasket can vary depending on the size of gasket. The allowable total leak rates from the whole gasket can be calculated by the allowable leak rates per unit length multiplied by the length of outside perimeter of the gasket contact area as shown in Eq.(9). 1000a o aL d (Pam3/s) (9)

It is considered that the above allowable leak rates can be applicable to most kinds of fluid used in industrial plants because the criterion is based on the Helium gas leak rates which shows a conservative value compared with the leak rates of actual service fluid under the same condition in most cases. And also the soapy bubble tests have been widely accepted to validate the leak tightness of gasketed flange connections in practice. In order to achieve a consensus on the acceptance criteria for toxic and hydrogen or the other flammable fluids, further discussion will be required.

Minimum Gasket Compression Since the leak rates can be expressed by a function of the

gasket compression as shown in Eq.(4), the required minimum gasket compression is the dominant design criteria of the pipe flange connection in order to achieve the specified allowable leak rates. The required average minimum gasket compression is given by Eq.(10).

1.5

*

1 lnreq

a

ck P

d L P

(mm) (10)

Minimum Gasket Contact Width The required minimum gasket contact width is also

supplementary defined to establish a proper gasket contact. As shown in Table 1, the minimum contact width is determined as 5mm for sheet gaskets and 4.5mm for spiral wound gaskets.

REQUIREMENTS FOR STRENGTH

Limit of Gasket Stress The average gasket stress is limited to avoid crushing

gasket under the bolted up and pressurized conditions. The average gasket stress is calculated as the gasket load Wg at the condition divided by the gasket contact area and should be within the gasket crush stress limit c as shown in Eq.(12).

2 2

4 gg c

o i

W

d d

(MPa) (12)

LIMIT OF BOLT STRESS The nominal tensile bolt stress is limited as shown in

Eqs.(13), (14) and (15). For the bolted up condition, the nominal bolt stress should be within 80% of the specified minimum yield strength of the bolt material as shown in Eq.(13). This criterion can avoid the bolt yield considering the multiaxial stress condition in the bolt tightening with some margin about 10%. By the effect of torsional stress due to the tightening torque, it is considered that the bolt yield occurs when the axial bolt stress comes to around 90% of the yield strength[9]. For the pressurized condition, the nominal bolt stress is limited within the specified minimum yield strength as indicated by Eq.(14). In addition to this, to keep the mechanical integrity as the part of pressure boundary, the bolts should withstand against the pressure thrust force. The bolt nominal stress should be within the allowable stress specified in the applicable design code for pressure vessel or piping as shown in Eq.(15).

0

0 0.8bb by

b e

WS S

n A (MPa) (13)

p

p bb by

b e

WS S

n A (MPa) (14)

2

4i

bt ba

b e

d PS S

n A

(MPa) (15)

where Sb0 and Sb

p are the nominal bolt stress under the bolted up and pressurized conditions, Wb

0 and Wbp are total bolt load

under the bolted up and pressurized conditions, Sbt is the nominal bolt stress due to the pressure thrust force, Sby and Sba are the specified minimum yield strength of the bolt material and the allowable stress according to the applicable design code, nb is the number of bolts and Ae is the tensile area of bolt. The bending stress of bolt is omitted in the above equations because the effects of bending stress are considered to be negligible.

Limit of Flange Stress Stress at Hub Although the stress occurred at the hub of

flange, or at the junction of pipe and flange in the case of no hub, is considered to be critical from a mechanical strength point of view, the stress is categorized as a local stress at the gross structural discontinuity. Therefore, the membrane equivalent hub stress (SH)M is limited within the specified minimum yield strength Sy of the flange material, or of the pipe material in case of no hub as shown in Eq.(16) to protect against the plastic collapse. The membrane plus bending equivalent hub stress (SH)M+B is also limited within twice of the specified minimum yield strength, known as the shakedown criterion, as indicated by Eq.(17). Those criteria are basically identical to most existing pressure vessel codes like the ASME Div.2[11]. H yM

S S (MPa) (16)

2H yM BS S

(MPa) (17)

The equivalent stress eqv can be calculated by the following equation using the stress components. The calculation of those stress components at the hub are shown in Table 2. The bending moment and shear force are calculated by



the load-displacement analysis based on the shell and rigid ring interaction theory representing the behavior of pipe flange as described in the following clause of Load Displacement Analysis.

2 2 2 21 6

2eqv t l l r r t

(18)

where t, l, r and are tangential, longitudinal, radial and shear stresses, respectively.

Stress at Flange Ring The limit of stress on the flange ring is also specified to determine the minimum thickness of flange ring from a mechanical strength point of view. The nominal shear stress (SR)S is limited to 70% of the applicable allowable stress Sfa of the flange material as shown in Eq.(19). The membrane plus bending equivalent flange ring stress (SR)M+B is limited within the specified minimum yield strength Sfy as shown in Eq.(20). The equivalent stress is calculated by Eq.(18). The calculation of stress components is according to Table 3. This stress calculation is based on the bending of circular plate with a center hole inside perimeter fixed supported.

0.7bR faS

f

WS S

Xt (MPa) (19)

where X is the outside diameter of larger hub end or pipe outside diameter in the case of no hub, tf is the flange thickness. R fyM B

S S (MPa) (20)

Limit of Bearing Stress The average contact stress at the bearing area around the

bolt hole is limited within the bearing strength Sac specified material wise as shown in Table 4. The bearing stress Sc is calculated by Eq.(21).

bc ac

b c

WS S

n A (MPa) (21)

where Ac is the contact area between the flange and the nut. The area is calculated by Eq.(22) when the nut contact area is hexagon and by Eq.(23) when the nut contact area is circle.

2 23 38 4c w hA d d

(mm2) (22)

2 2

4c w hA d d

(mm2) (23)

where dw is the width across flat faces of hex nut or the diameter of washer, dh is the bolt hole diameter.

DESIGN PROCEDURE

Design Flow Fig.2 shows the schematic flow of the flange design

procedure. The design flow has two steps, the first step is the leak tightness design step and the second step is the strength design step. The leak tightness design ensures that the requirements of minimum gasket compression and the contact width are established under the given initial bolt load by performing the load-displacement analysis. The bolt load is determined as the required minimum bolt load for the leak tightness. The strength design demonstrates that all members of the pipe flange connection have sufficient strength to withstand internal pressure as well as the expected maximum bolt load

considering a tightening coefficient depending on the bolt tightening method, as shown in Table 5. The tightening coefficient is the ratio of Fmax/Fmin where Fmax is the maximum axial bolt force and Fmin is the minimum axial bolt force. Here, Fmin takes the bolt force based on the required minimum bolt force determined by the leak tightness design.

Load Displacement Analysis The present design method adopts the load-displacement

analysis to predict the variations of bolt load and gasket load in the pipe flange connection due to the application of internal pressure. The analysis is based on the shell and rigid ring interaction analysis developed by Kohmura[12]. Fig.3 shows the modeling of pipe flange employed in the analysis. The model consists of 3 parts as the pipe, hub and flange ring. The hub is replaced with the equivalent cylinder having the average hub thickness. The flange ring is modeled by a ring with no cross-sectional deformation. The bolt load and gasket load are applied to the model and analyzed the displacement and rotation of each part. Also, the bending moment and shear force at the junction of pipe and hub or hub and flange ring are obtained by solving the problem. The bending moment and shear force is used to calculate the flange stress. In addition to the behavior of the pipe flange, the gasket stress distribution, compression, and the contact width can also be calculated according to the extension of Kohmura’s method proposed by Oyama[13].

OTHER CONSIDERATION The present design method recommends the use of the bolt

tightening procedure according to JIS 2251[14] for the assembly of the pipe flange connection to achieve the sufficient uniformity in bolt preloads which is expected in the design calculation.

VERIFICATION OF APPLICABILITY The applicability of present design method is examined by

applying it to the evaluation of leak tightness and strength of existing standard flanges. A series of standard flanges that JPI Class 300 RF WN Flanges with 4.5mm thickness Graphite filler Spiral Wound Gasket clamped is used to this verification study. The calculations are carried out under the condition that internal pressure 5MPa, i.e. the maximum rating pressure for Class 300, is applied as the operating pressure at ambient temperature. Carbon steel flanges and low alloy steel bolts are assumed. Flange bolts are assumed to be tightened by torque wrench which gives the ratio of Fmax/Fmin 1.4.

Figure 4 shows the calculated leak rates under the condition that the flanges are clamped by the minimum bolt preload to achieve the required gasket compression or the required gasket contact width. It is confirmed that the calculated leak rates satisfy 110-3Pa.m3/s/m. Above 26 inch flanges, the calculated leak rates decrease as the flange size increases. In this range, the gasket contact width determines the minimum required bolt preload because the gasket contact width tends to be narrow due to the flange rotation. Figure 5 plots the bolt stress representing the maximum bolt load condition considering the scatter in the bolt tightening. The bolt stress has sufficient margin to the allowable stress of bolt. Figure 6 indicates the bearing stress corresponding to the maximum bolt load condition. Figure 7 illustrates the gasket



stress to check the gasket compressive strength. As mentioned before, the gasket stress increases as the flange size increases in the bolted-up condition. The calculated flange stress is shown in Fig.8 through Fig.10. The membrane plus bending stress in the flange ring is plotted in Fig.8. The stress slightly increases as the flange size increases. The larger hub stress and the smaller hub stress are shown in Fig.9 and 10, respectively. Although the larger hub stress shows almost constant regardless of the flange size, the smaller hub stress varies depending on the flange size. Relatively severe stress at the smaller hub is calculated for the larger size flanges. But the stress has still some margin to the stress limit.

As indicated in all those figures, the requirements of leak tightness and strength are satisfied in these flanges. This can be understood that the present design method proved the adequacy of the design of existing standard flanges. Conversely, the evaluation method adopted in the present design procedure is valid to examine the leak tightness and the strength of the flange connections.

CONCLUSION Outlines of the committee draft for leak tightness based

design procedure of the bolted gasketed pipe flange connections have been introduced. This method can provide the reliable flange design considering the sealing performance as well as the mechanical integrity. The validity of the design method is also confirmed by the calculation results for actual flanges.

ACKNOWLEGEMENT The present flange design method is based on the

committee draft of the leakage based flange design procedure prepared by the Sealing Technology Of Pressure-equipments Committee organized in the High Pressure Institute of Technology Japan. The chair of committee is T. Sawa of Hiroshima University, the vice chair is T. Kobayashi of Numazu College of Technology, and the secretary is R. Kurosawa of Yokogawa Electoric. The present research and development work is done by the contributions and supports of the above 3 executives and the other 28 committee members from plant owners such as power, petroleum and chemical companies, manufacturers, engineering and construction contractors, and technical consultants.

NOMENCLATURE Ac contact area between flange and nut, mm2 Ae tensile area of bolt, mm2 a constant determined by JIS B 2490 gasket leak test, - bA constant determined by JIS B 2490 gasket leak test, - bB constant determined by JIS B 2490 gasket leak test, - c constant determined by JIS B 2490 gasket leak test, - d constant determined by JIS B 2490 gasket leak test, - dh bolt hole diameter, mm di inside diameter of gasket contact area, mm do outside diameter of gasket contact area, mm dw width across flat faces of hex nut or diameter of washer,

mm ELD slope of gasket loading curve used for load-displacement

analysis, MPa EUL slope of gasket unloading curve used for load-

displacement analysis, MPa

F factor in regression of gasket unloading curve, - Fmax maximum axial bolt force in tightened flange bolts, N Fmin minimum axial bolt force in tightened flange bolts, N k gasket dimension correction factor, - L total leak rate for a particular size of gasket, Pa.m3/s La allowable total leak rate from gasket, Pa.m3/s Ls specific leak rate according to JIS B 2490, Pa.m3/s la allowable leak rate per unit length at outer perimeter of

gasket contact area, Pa.m3/s/m nb number of bolts, nos. P operating pressure, MPaP

* gasket leak test pressure, MPa Sac bearing strength of flange material, MPa Sb

0 nominal bolt stress in bolted up condition, MPa Sb

p nominal bolt stress in pressurized condition, MPa Sba allowable stress of bolt material according to applicable

design code, MPa Sbt nominal bolt stress due to pressure thrust force, MPa Sby specified minimum yield strength of bolt material, MPa Sc average bearing stress around bolt hole of flange, MPa Sfa allowable stress of flange material according to applicable

design code, MPa Sfy specified minimum yield strength of flange material, MPa (SH)M membrane equivalent hub stress, MPa (SH)M+B membrane plus bending equivalent hub stress, MPa (SR)S nominal shear stress in flange ring, MPa (SR)M+B membrane plus bending equivalent flange ring stress,

MPa Sy specified minimum yield strength of flange material or

pipe material, MPa tf flange thickness, mm tg initial gasket thickness, mm Wb

0 total bolt load in bolted up condition, N Wb

p total bolt load in pressurized condition, N Wg gasket load in bolted up or pressurized condition, N wmin Minimum gasket contact width, mm X outside diameter of larger hub end or pipe outside

diameter, mm gasket compression, mm

req required minimum average gasket compression, mm 11 gasket stress at step 11 of JIS B 2490 gasket leak test,

MPa c gasket crushing stress limit, MPa

g allowable maximum average gasket stress, MPa max maximum gasket stress in JIS B 2490 gasket leak test,

MPa pre gasket stress at which the unloading started in gasket

unloading curve, MPa req required minimum average gasket stress, MPa

eqv equivalent stress, MPa l longitudinal stress, MPa r radial stress, MPa t tangential stress, MPa shear stress, MPa

REFERENCE [1] Bickford, J.H., "Gasket and Gasketed Joints", Marcel

Dekker Inc., (1998)



[2] Bickford, J.H., "An Introduction to the Design and Behavior of Bolted Joints, 3rd Ed.", Marcel Dekker Inc., (1995)

[3] Committee Draft of "Sec.VIII Div.1 App.BFJ Draft July 16, 2000," Pressure Vessel Research Council, Bolted Flange Committee, (2000)

[4] EN Standards, EN1591-1, "Flanges and their joints, Design rules for gasketed circular flange connections - Calculation method", (2001)

[5] ASME Boiler and Pressure Vessel Code, Section VIII, Division 1 "Rules for Construction of Pressure Vessels," Appendix 2, (2010)

[6] Sawa, T., Kobayashi, T., Tsuji, H., Nagata, S., "Leak Tightness based Design Procedure for Gasketed Pipe Flange Connections," Proceedings of ASME Pressure Vessel and Piping Conference, ASME PVP2011-57738, (2011)

[7] Kobayashi, T., Nishida, T., Yamanaka, Y., "Mathematical Model for Sealing Behavior of Gaskets Based on Compressive Strain," Proceedings of ASME Pressure Vessel and Piping Conference, Analysis of Bolted Joints, ASME PVP-Vol.416, pp.105-109, (2001)

[8] Kobayashi, T., "Characterization of Sealing Behavior of Gaskets for the Leak Rate Based Design of Gasketed Bolted Flanged Connections," Proceedings of ASME Pressure Vessel and Piping Conference, PVP2008-61465, (2008)

[9] Japanese Industrial Standards, JIS B 2490 "Test method for sealing behavior of gaskets for pipe flanges," (2008)

[10] Nagata, S., Kaneda, S., Tsuji, H., Sawa, T., "Three Dimensional Elastic-Plastic Finite Element Simulation on Bolt Tightening beyond Yield," Proceedings of ASME Pressure Vessel and Piping Conference, PVP2009-77689, (2009)

[11] ASME Boiler and Pressure Vessel Code, Section VIII, Division 2 "Alternative Rules for Construction of Pressure Vessels," (2010)

[12] Kohmura, S., "Study on Design of Alminum Pipe Flanges, 1st Report," Transactions of the Japan Society of Mechanical Engineers, Vol.51, No.461, p.196 (1985)(in Japanese)

[13] Oyama, Y., Sawa, T., Kobayashi, T., Nagata, S., "A Simple Analysis and Design Method of Bolted Pipe Flange Connections with Gaskets under Internal Pressure," Proceedings of ASME Pressure Vessel and Piping Conference, PVP2009-77589, (2009)

[14] Japanese Industrial Standards, JIS B 2251 "Bolt tightening procedure for pressure boundary flanged joint assembly," (2008)

[15] Sawa., T., Kobayashi, T., Oyama, Y., Nagata, S., "Leak Tightness Based Design Procedure for Gasketed Pipe Flange Connections," Proceedings of ASME Pressure Vessel and Piping Conference, PVP2011-57738, (2011)

________________________________________________



Table 1 Gasket Property Parameters

Type Material tg a bA bB c d

Sheet Gasket

Rubber Binding Non-asbestos Fibers

1.5 309.62 1.0505 3.5500 0.0570 58.622 3.0 144.53 1.0355 3.8013 0.0821 27.215

Extended PTFE 1.5 630.83 1.8400 7.0771 0.0241 31.245 3.0 97.052 1.6948 8.7873 0.0711 12.317

Graphite and PTFE Composite 1.5 1519.0 1.4776 4.0350 1.1783 205.74 3.0 231.61 0.9061 4.7556 0.0767 60.830

Inorganic and PTFE Composite 1.5 3331.5 1.6719 4.8361 2.0145 178.62 3.0 126.87 0.6786 4.4773 0.0449 59.955

Expanded Graphite 1.5 212.86 3.9330 23.855 0.0189 5.9389 3.0 19.909 3.4534 21.891 0.0278 2.6236

Spiral Wound

Gasket(2)

Graphite Filler 3.2 259.95 0.8353 3.8574 0.0005 12.464 4.5 123.33 0.6710 8.4242 0.0005 6.1194

Inorganic Paper Filler 3.2 224.30 1.1856 8.9856 0.0047 15.961 4.5 99.900 1.5431 15.884 0.1010 10.354

PTFE Filler 3.2 182.56 1.0478 6.9463 0.0002 11.518 4.5 105.08 1.1448 14.523 0.0002 11.518

Table 1 Gasket Property Parameters (Continued)

Type Material tg ELD(3) EUL

(3) c wmin

Sheet Gasket

Rubber Binding Non-asbestos Fibers

1.5 420.9 830.8 200 5.0 3.0 414.9 861.7 200 5.0

Extended PTFE 1.5 268.6 923.2 200 5.0 3.0 202.4 840.6 200 5.0

Graphite and PTFE Composite 1.5 703.3 1528 200 5.0 3.0 833.5 2059 200 5.0

Inorganic and PTFE Composite 1.5 845.1 2115 200 5.0 3.0 657.5 1549 200 5.0

Expanded Graphite 1.5 91.80 962.0 100 5.0 3.0 98.00 946.7 100 5.0

Spiral Wound

Gasket(2)

Graphite Filler 3.2 1004 2109 200 4.5 4.5 615.1 2461 200 4.5

Inorganic Paper Filler 3.2 632.5 2679 200 4.5 4.5 449.7 3207 200 4.5

PTFE Filler 3.2 568.4 1923 200 4.5 4.5 469.9 3082 200 4.5

Notes: (1) under preparation (2) both inner and outer rings equipped (3) in case the gasket stress 40MPa for sheet gaskets and 100MPa for spiral wound gaskets (4) The data shown in the table are derived from the gasket leak tests according to JIS B 2490. The tests had been carried out by several gasket manufacturers under the STOP Committee.



Table 2 Stress Components Calculation for Hub

Tangential Stress t

Longitudinal Stress l

Radial Stress r

Shear Stress

Outside

Surface 2 2

2 2 61 i

p Ew M

D tY t

2 2

61

p M

Y t

0 Q

t

Average 2

2i

i

pD Ew

t D t

2 1p

Y

2p

Q

t

Inside

Surface

2

2 2

1 2 61 i

p Y Ew M

D tY t

2 2

61

p M

Y t

p Q

t

H MS should be calculated by using average values. H M B

S

takes the greater one of the two values calculated by using the inside or outside surface stress components. Symbols are shown as below;

E : Modulus of elasticity (MPa) M : Longitudinal bending moment per unit length (N.mm/mm) Q : Radial shear force per unit length (N/mm) p : Internal pressure (MPa)

iD : Inside diameter of hub or pipe (mm) t : Thickness of larger hub end or smaller hub end (mm) w : Radial displacement (mm) Y : Ratio of outside diameter to inside diamenter 2i iY D t D (-) : Poison’s ratio (-)

Table 3 Stress Components Calculation for Flange Ring

Tangential Stress t

Longitudinal Stress l

Radial Stress r

Shear Stress

Outside

Surface 2

6 t

f

M

h 0 2

6 r

f

M

h

f

Q

h

Average 0 0 0 f

Q

h

Inside

Surface 2

6 t

f

M

h 0 2

6 r

f

M

h

f

Q

h

R M BS

takes the greater one of the two values calculated by using the inside or outside surface stress components. Symbols are

shown as below; tM : Tangential bending moment per unit length (N.mm/mm)

rM : Radial bending moment per unit length (N.mm/mm) Q : Longitudinal shear force per unit length (N/mm)

ft : Flange thickness (mm)

Table 4 Bearing Strength

Flange Material Allowable Compressive Bearing Stress, Sac Carbon Steel and Low Alloy Steel 0.625Sfu

Stainless Steel Sfy

Note Sfu and Sfy are the specified minimum tensile strength and yield strength or 0.2% proof strength of the flange material.



Table 5 Tightening Coefficient for Strength Design

Method of Tightening Tightening Control Tightening Coefficient for

Assessment of Bolt Strength

Tightening Coefficient for

Assessment of Flange Strength

Torque Method No measurement 2.0 1.5

Torque Control 1.4 1.2

Axial Force Control 1.1 1.05

Tensioning Method Axial Force Control 1.1 1.05

Table 6 Tightening Coefficient for Determining Target Tightening Bolt Force

Method of Tightening Tightening Control Tightening Coefficient for Determining Target Bolt Force

Torque Method No measurement 1.5

Torque Control 1.2

Axial Force Control 1.05

Tensioning Method Axial Force Control 1.05

Fig.1 Approximation of Gasket Elasticity

max

11=0.125max

max 11

Gasket Stress, MPa

Gasket Deflection, mm

(Slope of loading curve) =ELD/tg (Slope of

unloading curve) =EUL/tg

A

B

O



Fig.2 Flow of Flange Design

Start

Initial Design

Gasket Property Parameters

Allowable Leak Rate

Suppose Bolt Load for Gasket Seating

Load-displacement Calc for Leak Tightness Check

Gasket Compression and

Contact Width OK?

Requirements of Gasket

Compression and Contact Width

Leak Tightness Design

Flange Stress

Bolt Stress

Bearing Stress

Gasket Stress OK?

Strength Design

if necessary

Ｙ

Ｎ

Ｙ

Ｎ

Load-displacement Calc for Strength Check

Determine Initial Bolt Load

Tightening Coefficient

End



A/2

C0/2

B/2

t0i

θi

thi

P0

Q1i

M1i

Q1i

w

hix

Q2i

P0

Q2i

M2i

Fig.3 Modeling of Pipe Flange

Fig.4 Calculated Leak Rates per Unit Length for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.

Pipe

Hub

Flange Ring



Fig.5 Calculated Bolt Stress for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.

Fig.6 Calculated Bearing Stress for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.

Fig.7 Calculated Average Gasket Stress for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.



Fig.8 Calculated M+B Flange Ring Stress for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.

Fig.9 Calculated M+B Larger Hub Stress for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.

Fig.10 Calculated M+B Smaller Hub Stress for JPI Class 300 RF WN Flanges with Graphite SWG t4.5mm.


Development of Design Procedure for Gasketed Pipe Flange ...

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