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2 OTC 23063

systems. Since API 17D code was specifically formulated for the subsea equipment it will be mandatory to suffice the API material requirements, design allowables and test requirements.

The industry (API/AWHEM) has a long and established history in the development of new standards and procedures for equipment rated at or above 15000 psi. As described by Payne, M., 2010, the original 15000 psi Wellhead specifications were developed in 1952. The first 20000 psi Wellhead systems were developed in 1972, quickly followed by 30000 psi Wellhead systems in 1974. These systems have been successfully deployed both in land and platform based fields in the Gulf of Mexico and North Sea.

In the years from 1960 to 1990, subsea wells up to a working pressure of 5000 psi at water depth of 2000 ft were being explored. In the 1990s the working pressure increased to 10000 psi at water depth of 3000 ft. From 2000 to 2011, the subsea wells up to 15000 psi at water depth of 8000 ft were explored. The trend of the industry is moving towards wells being explored at 20000 psi at around 10000 ft water depth. These wells require the high pressure high temperature rated equipment for oil production. This establishes a need for design methods appropriate for higher pressure and temperature rated subsea production equipment.

The ASME and API codes differ in the material, hydrostatic test, and NDE requirements in particular. The standard hydrostatic test requirement per API 17D is 1.5 times the rated working pressure whereas in ASME BPVC Sec. VIII, Div. 2, the hydrotest requirement is 1.43 times the working pressure. Additionally, the hydrostatic test pressure in ASME varies between Divisions and has also fluctuated over time due to experiences of the ASME community. The material requirements to satisfy various codes or standards are also not common or aligned. The different code specific material requirements are provided in API 6A sec. 5, ASME BPVC Section VIII, Div. 2, Part 3 and Div. 3, Part KM. The primary choice of materials for these design recommendations will be based on the limitation of API 6A, sec. 5. Specific guidance is provided in the paper in order to generate a conservative utilization of ASME methods with API materials. These approaches are based upon the normalization of the hydrostatic test pressures across the various codes. Finally, the NDE requirements in ASME can allow for larger flaw sizes as compared to the API allowable limits.

Design Methods Components or assemblies classified as pressure containing parts per API 17D will be evaluated by the following techniques. In addition, equipment not defined as pressure containing but which experience a pressure greater than their working pressure (e.g. tubing hangers) during testing before deployment in the field will be designed by these methods. These design methods include a) linear and elastic-plastic protection against plastic collapse; b) linear and elastic-plastic protection against local collapse; c) protection against cyclic loading and d) closure and critical bolting design. For protection against collapse from buckling, the design methodology as suggested in ASME Section VIII, Div.2, 5.4 is recommended.

To ensure that all modes of failure are addressed, three load cases should be considered. A hydrostatic load case with test

pressure value of at least 1.5 times of the rated working pressure shall be considered per API 17D requirements. Additionally, a working load case with appropriate load scenarios as given by the design code used for calculations is addressed. A reference table is supplied in ASME BPVC Section VIII, Div.2 Table 5.1. Lastly, a maximum load case for each component or assembly is evaluated for stress and structural stability. Along with each of the three load cases listed, each part must take into consideration any fatigue effects on the design. If a part falls under the designation of heavy section bodies (R/t < 4, or the ratio of inner radius to thickness), special requirements are outlined under the respective design methods. Worst material tolerance conditions shall be used when evaluating structural capacities. To check for proper functionality of the equipment, the deflection of systems under load must be accounted for during the calculation of capacity. As consistent with API, there shall be no de-rating of material strengths for equipment in use up to 250F. For service temperatures above 250F, the material strength should be de-rated per the factors in API 6A, Annex G.

Protection against Plastic Collapse

This design method addresses general plastic collapse in equipment because of gross distortion across its membrane due to applied loads. This failure mode can be analyzed using both linear methods and elastic-plastic analysis methods.

Linear Methods The linear elastic methods include stress analysis which can use handbook solution or other industry accepted methods as

long as these solutions represent the component geometry and loading conditions appropriately. Otherwise, numerical analysis techniques such as the finite element method shall be used to determine the stresses in equipment. The calculation of stresses will be based upon the Stress-Intensity equation based on Tresca Yield Surface and Maximum Shear Stress Theory.

Hydrostatic test design allowable stress values will be based on API 6A section 4.3.3.2 and/or 4.3.3.3. Hydrostatic test

design allowable stress values can also be based on ASME, Section VIII, Div. 2, section 4.1.6.2. Working load design

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OTC 23063 3

allowable stress values will be based on API 6A, section 4.3.3.2. Working load design allowable stress values can also be based on ASME, Section VIII, Div. 2, section 4.1.6.1.

If appropriate, in vessels which are R/t 4, the classification of stress can be performed based on ASME, Section VIII,

Div. 2, 5.2.2.2 and Figure 5.1, the Hopper Diagram. In vessels which are R/t < 4, the classification of stress should not be used. Stresses may still be linearized, but all results should be considered as primary membrane stresses. If stress linearization is used, the peak stress intensity derived through calculations shall not be allowed to exceed yield strength for more than 5% of the section thickness (path distance). This recommendation is based on ASME, Section VIII, Div. 2, 5.2.1.3. This ensures that the stress distribution through the section of the material is not so extreme as to cause local failure in a portion of the section before full section failure might occur. Finally, the stress linearization procedures should be aligned with ASME, Section VIII, Div. 2, Annex 5.A.

Generally the analyst must follow the requirements of ASME BPVC Section VIII Div. 2, Part 5, with the exception that

stress intensity shall be used instead of equivalent stress. Although, ASME BPVC Section VIII, Div.2, Part 5 calculates stresses through equivalent stress equations (von Mises Stress), it is recommended to continue to use Stress Intensity to satisfy API design requirements. Finite element methods should utilize small displacement theory. Thermal loading or thermal results can be calculated separately or together with the structural model depending on the particulars of the problem. ASME BPVC Section VIII, Annex 5.A established the preferred linearization procedure and path methodology. After evaluating the component or assembly using linear elastic criteria for protection against general plastic collapse, further evaluations may be addressed to protect against other failure modes.

Elastic Plastic Method The Elastic Plastic Method can be used to evaluate components and interactions for protection against plastic collapse.

Generally, the analyst must follow the requirements of ASME BPVC Section VIII Div. 2, Part 5. The von Mises yielding criterion and associated flow rules will be used. Material data will be reported as true stress true

strain. It is preferred to utilize actual material test data for the elastic-plastic methods, but a standard (and conservative) estimation of material behavior is given in ASME Section VIII, Div. 2, Part 3. A load and resistance factor design (LRFD), as documented in ASME BPVC Section VIII, Table 5.5 shall be used to determine if a structure is suitable for operation at a specific load case. However, given that the typical hydrostatic test pressure established by ASME BPVC Section VIII, Part 8 is a minimum 1.43 times the design pressure (Maximum Allowable Working Pressure), this is in conflict with the standard API hydrostatic test case. Therefore, it is recommended that the LRFD factors be increased by the ratio established below:

(1.50/1.43) = 1.049 Equation 1

where 1.50 is the API hydrostatic test multiplier, 1.43 is the ASME hydrostatic test multiplier, and the factor 1.049 is the multiplier applied to increase the LRFD factors established by ASME. If the capacity of a component under engineering evaluation is to be established using the elastic-plastic methods, the load case inputs shall be uniformly increased until the solution is non-convergent due to unbounded deformation. The loads at which non-convergence are established are then decreased by the appropriate LRFD factors and the value set in Equation 1.

As an example, if an internally pressurized vessel reached non-convergence at 60000 psi internal pressure. The design

pressure would be established by dividing the 60000 psi by 2.40 per ASME BPVC Section VIII, Table 5.5, Global Criteria 1 and by 1.049 (per Equation 1). This would establish the maximum allowable design pressure for the vessel at 23832 psi, and an API hydrostatic test pressure of 35748 psi (calculated as 1.50 times the design pressure).

After evaluating the component or assembly for protection against collapse, further evaluations must be made to protect

against local collapse. Protection against Local Collapse

This design method addresses local collapse in equipment in addition to protection against gross plastic collapse, as described in the sections above. This failure mode can be analyzed using both linear methods and advanced elastic plastic analysis methods. In general, the analyst must follow the requirements of ASME BPVC Section VIII Div. 2, 5.3 to address protection against local collapse.

Linear Method The linear elastic method is used in addition to the protection against plastic collapse, as described above, to guard against

local failure. ASME BPVC Section VIII Div. 2, 5.3.2 outlines this linear method.

4 OTC 23063

Elastic Plastic Method The elastic plastic method is used along with the protection against plastic collapse to prevent local failure. ASME BPVC

Section VIII Div. 2, 5.3.3 describes the elastic plastic method used for protection against local failure. This method of analysis is for a sequence of applied loads, based on the load cases discussed previously.

Protection against Cyclic Loading

After evaluation for general and local collapse, further evaluation maybe made to determine that a failure mode generated by load fluctuations is not encountered. Examples of these failure modes are fracture formation or ratcheting of non integral connections.

Components or assemblies having undergone evaluation to prevent general plastic collapse shall adhear to the following

criteria to protect against failure due to cyclic loading. ASME Section VIII Div. 2, 5.5 gives the fatigue analysis approach. This is also the recommendation of API 17D, sec 5.1.3.1. ASME Section VIII Div. 3, KD-3 is a traditional fatigue analysis approach based on stress life theory. It shall be used when a leak-before-burst mode of failure is established. ASME Section VIII Div. 3, KD-4 is a fracture mechanics approach to design. It shall be used when a leak-before-burst mode of failure cannot be shown to exist.

ASME Section VIII Div. 3, KD-340 is a traditional fatigue analysis approach based on strain life theory. It shall be used

when welds are present as structural supports. If the assembly has non-integral components in the structural load path that could progressively distort through sequential load cycles, it must also be evaluated for protection against ratcheting failure. Examples of this are screwed-on caps, screwed-in plugs, shear ring closures, and breech lock closures. Other examples in the industry might include collet or clamp style connectors. Evaluation using the procedure outlined in ASME Section VIII Div. 3, KD-234 shall be followed.

Closure Bolting Design

Linear elastic methods outlined in API 17D and paralleled in ASME Section VIII, Div. 2, Part 5.7 are recommended for closure bolting design. In analyzing the stress capacity of closure bolting, the maximum allowable tensile stress should be no more than 83% of the materials yield based upon the root area of the thread (API 6A, 4.3.4). According to API 6A, bolting stresses are to be determined using all loads acting on the closure area, including the pressure acting over the seal area, gasket loads, and any other mechanical and thermal loads. The maximum allowable stress is also determined in all conditions such as working pressure and hydrostatic conditions.

Both closure and critical bolting require a preload to a high percentage of the materials yield strength. Closure bolting of

all API 6BX and 17SS flanges is to be made up to approximately 67% of the bolts material yield stress. Other studs, nuts, and bolting used on end connections of subsea equipment shall be made up to at least 50% of the bolts material yield stress (API 6A, Annex D). Structural bolting is to be made up to the manufacturers written specification. Studs, nuts, and other closure bolting for use in subsea service are often manufactured with corrosion inhibiting coatings or platings which can dramatically affect the stud-to-nut friction factor. The manufacturer shall document the recommended make-up torque for their fasteners.

High strength alloy steel bolts, studs, and nuts shall be evaluated for cyclic operation using elastic stress analysis and

equivalent stress amplitude loading established in ASME BPVC Section VIII, Div. 2, 5.5.3 and Annex 3.F. Stress concentrations at the root of standard ASME B1.1 threads are not to be included in the fatigue evaluation. In cases where loading is shared between components which have deformed, such as a metal gasket, the controlling stress for the fatigue evaluation will be the effective total equivalent stress amplitude. This value is defined as one-half of the effective total equivalent stress range calculated for each cycle in the loading histogram.

As a safe guard to prevent fatigue failure of threaded components, a limit shall be placed on equipment so that no regular

load encountered during working conditions causes a separation of the clamped joint of a threaded fastening system. An exception to this would be extreme scenarios such as a drive off or impact event. Clamped faces which separate during qualification of factory acceptance testing shall have their fasteners either replaces or re-tensioned, as per the manufacturers written specifications.

Non-Destructive Examination For non-destructive examination of the component, rules from API 17D, API 6A and API RP 6HT should be followed.

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OTC 23063

Example anaTo demonstrateThe flange wasensile strength

assumed to be ension-bending

against plastic by evaluating tas fracture mec

Protection agaFinite elem

degree model oshown in Figurhe model and

concern. The sebe validated bywas applied in analyses: flange

Linear elast

for membrane sof yield strengtwas determined26 ksi.

Elastic-plas

calculated usinghe bolts. The ioad. LRFD of

hydrotest factoro be the same

Figure 1: 3-D

To determipressure & extdetermined for he flange. For

for all loads ap

alysis of a APe the applicatios assumed to bh. The bolts w

preloaded at 6g moment capacollapse by linhe design cycl

chanics analysi

ainst plastic coent analysis (Fof the flange wre 1. The gaske

also because erviceability ofy performance

the model to e structural lim

tic FEA for hystress intensityth and membrad to be 30 ksi l

stic FEA was cg the ASME Binternal pressuf 2.4 from ASr of 1.049 (peras linear elastic

Dimensional 18

ine the Pressuternal tension that particularelastic-plastic plied (Pressure

PI 6A, 4 in. 20on of the desige made from F

were assumed 67% of the boacity chart simnear elastic as les for a set of s. Temperature

ollapse FEA) was condwith 3D bolts et was not inclpreviously co

f the gasket is dtesting based osimulate the g

mit and bolt stru

ydrotest pressury and membranane plus bendiimited by the h

conducted usinBPVC Section Vure was determSME Section Vr Equation 1) toc analysis meth

0 degree FEA M

ure-Tension-Beload. The ben

r load case. Seanalysis, LRF

e, Tension and

0 ksi, type 6Bgn methods outF22, low carbonto be made o

olt yield strengmilar to the one

well as elasticload cases on

e conditions we

ducted using coand rigid bottoluded in the 3Dnducted lineardetermined at ton API 17D te

gasket for all anuctural limit.

re load case wne plus bendinging was limitedhub-neck area.

ng the same mVIII, Div.2 An

mined at which VIII, Div. 2, o determine thhods.

Model of the 4 insimulat

ending Momennding momenteveral cases weFD of 2.4 from d Bending Mom

BX Flange tlined above, an alloy steel (2f high strength

gth per the rec in API TR 6A

c-plastic analysthe pressure-teere ignored in

ommercial sofom surface waD model to avor elastic methothe maximum est methods. Analyses perform

as evaluated fog stress intensitd to yield stren The overall ca

meshed model. nnex. 3. D / Div

the model failTable 5.5, Glo

he maximum al

n. 20 ksi API 6BXting bottom flan

nt (P-T-B) capt to fail the flere run and resASME Sectio

ment). Two cha

an API 6A, 4 in2-1/4 Cr. 1 Moh steel with 1

commendation AF is generatedsis. Protection ension-bendingthe analyses ba

ftware Abaqus as used for botoid numerical ods indicated tworking condi

A predeterminermed. Two flan

or the worst strty. The membrngth. The maxapacity of the f

The true stressv.3 KD-231.4.led to convergobal Criteria 1llowable worki

X Flange with 3nge

pacity chart, tflange either bsults were com

on VIII, Div. 2,arts have been

n. 20 ksi, typeo.) with 75 ksi 105 ksi yield s

in API 17D, d using the desagainst cyclic

g moment charased on API TR

6.10.1. A threth linear and einstability cauthat gasket groition, based ond gasket load unge failure crit

ress classificatrane stress inteximum structurflange was lim

s - true strain . Linear elastic

ged, which is c1, was used aing pressure as

3-Dimensional b

the flange waby flange stresmpiled to determ, Table 5.5, Glpresented in th

e 6BX flange isyield and 95 kstrength. The sec. 5.1.3.5. A

sign method ofc loading is demrt by both fatigR 6AF procedu

ee dimensionalelastic-plastic

using prematureoove was not

n contact stresseusing axisymmteria were con

tion line (lineaensity was limiral capacity of

mited by the bol

curve for the c properties wecalled the plastalong with the s 30 ksi. This i

bolts and Rigid S

as subjected tosses or bolt strmine the P-T-Blobal Criteria 1he paper. Figu

5

s analyzed. ksi ultimate bolts were

A pressure-f protection monstrated gue as well ures.

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arized path) ited to 83% f the flange lts and was

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Surface

o a known tresses was B chart for 1, was used ure 2 shows

6 OTC 23063

the P-T-B chart for the overall capacity of the flange taking into account both flange and bolt structural limits, with both elastic-plastic and linear elastic analysis methods. Figure 3 shows the P-T-B chart considering only the flange structural limit for both elastic-plastic and linear elastic methods for comparison.

Figure 2: Overall Pressure-Tension-Bending Moment Capacity of the 4 in. 20K flange using linear elastic and elastic plastic analysis

methods considering both bolt and flange structural limits

Figure 3: Pressure-Tension-Bending Moment Capacity of the 4 in. 20K flange to compare the linear elastic and elastic plastic

analysis methods considering only flange structural capacity

It is observed from Figure 2 that the capacity of the flange is limited by linear elastic analysis for all the load cases

considered. The chart is comparable to the API TR 6AF charts for the 4 in. 20 ksi 6BX type flange and shows slightly higher

0

5

10

15

20

0 50 100 150 200 250 300

Bor

e P

ress

ure,

[ksi

]

Bending Moment, [ft-kips]

T=0 kips

T=200 kips

T=400 kips

T=600 kips

Bolt

FlangeStructural

FlangeStructural

FlangeStructural

FlangeStructural

Bolt

Bolt

Bolt

FlangeStructural

FlangeStructural

Bolt

Bolt

Bolt

Bolt

Bolt

Bolt

Bolt

Bolt

Bolt

Bolt

0

5

10

15

20

0 50 100 150 200 250 300

Bore

Pre

ssur

e, [k

si]

Bending Moment, [ft-kips]

T=0_LE

T=200_LE

T=400_LE

T=600_LE

T=0_EP

T=200_EP

T=400_EP

T=600_EP

T = External Tension Load in kipsLE = Linear Elastic Analysis EP = Elastic-Plastic Analysis

T= External Tension Load in kips

OTC 23063 7

capacities since higher bolt preload was assumed. Only in the cases with no internal pressure and external bending moment applied, the capacity observed is slightly less than that in API TR 6AF. In contrasting these two methodologies, API TR 6AF considers an axisymmetric FEA model to determine the P-T-B chart for this flange. The bolts are evaluated using an axisymmetric model, with the bolt stresses averaged for all 8 bolts. The results reported in Figure 2 use a 3D model to examine the stresses in each of the bolts. With external bending applied, the outermost bolts reach the maximum allowable stress before other bolts. This is the reason why the capacity observed is slightly less than in the API TR 6AF charts.

As seen in the Figure 3, the structural capacity of the flange is slightly higher in most cases when evaluated with the

elastic-plastic methods. Where there is external tension and low pressure, the linear elastic method gives higher capacity than the elastic-plastic method.

Protection against Cyclic Loading

To demonstrate the design methods for cyclic loading the structural capacity of the flange was evaluated for a set of load cases from the P-T-B chart generated for flange stress using linear elastic analysis. It was assumed that for one cycle, all the loads applied after bolt preload, went from zero to the actual value and again went back to zero. For example, for a load case of 20000 psi internal pressure, 600 kips external tension and 120 ft-kips external bending moment, one cycle of loading would be zero loads to 20000 psi internal pressure, 600 kips external tension and 120 ft-kips external bending moment and back to zero loads. Design cycles were calculated for these cases using fatigue analysis and fracture mechanics analysis per ASME BPVC Section VIII, Div 3.

The flange under consideration is assumed to be made from F22, low carbon alloy steel (2-1/4 Cr. 1 Mo.) with 75000 psi

yield strength and 95000 psi ultimate tensile strength. Published material properties have been used to determine the design cycles for both fatigue and fracture mechanics analysis.

Fatigue Analysis Fatigue analysis was conducted per ASME BPVC Section VIII Div.3, KD-3. Since ASME BPVC Section VIII Div.3,

Table KD-320.1, provides material fatigue curve for low carbon alloy steel for UTS 90000 psi, this curve was used for fatigue analysis. Half of the maximum stress intensity from the linear stress analysis results was used for the equivalent stress amplitude and the number of design cycles was calculated. The allowable cycles were taken as 1/10th of the number of cycles calculated using formulation from Div.3. Figure 4 shows the design cycles for the extreme load cases considered.

The reason for lower cycles at zero pressure cases is the contribution of hoop stress at high pressure which shifts the

fatigue evaluation zone.

Figure 4: Design cycles using fatigue analysis for identified load cases of the pressure-tension-bending moment capacity chart

generated using linear elastic analysis

0

5

10

15

20

0 50 100 150 200 250 300

Bor

e P

ress

ure,

[ksi

]

Bending Moment, [ft-kips]

T=0_LE

T=200_LE

T=400_LE

T=600_LE

T=0_EP

T=200_EP

T=400_EP

T=600_EP

5654 cycles 5572 cycles

3884 cycles 3016 cycles

8 OTC 23063

Fracture Mechanics Analysis Fracture mechanics analysis was conducted per ASME BPVC Section VIII Div.3, KD-4. For material properties, the

fracture toughness value of 138 ksi.in. was used from the published report by Young, K., 2008, which corresponds to Seawater with Cathodic Protection environment and was determined per ASME BPVC Section VIII Div.3, by CTOD (Crack Tip Opening Displacement) method. Recommended crack growth rate factors for steel in marine environment with cathodic protection were used from BS7910 as listed below in form of Paris law equation:

. = 5.22 10(Kksi. in. ). Equation 2

For the crack orientation, both circumferential and axial crack orientations were considered. For the load cases identified,

the flange is loaded in such a way that the stresses are predominantly axial stresses and hence a circumferentially orientated crack on the inner surface that has a radial depth yielded the most conservative results. This places the largest initial crack perpendicular to the worst loaded plane, and initializes crack propagation at the highest stress level in that plane. The crack was assumed to be semielliptical in shape with ratio of the depth to surface length of 1:3 per ASME BPVC Section VIII, Div.3, paragraph KD-411. This initial flaw is placed along the failure path determined from the structural linear elastic analyses. Initial flaw length of 3/16 in. was used for the analyses. Crack face pressure was included in the evaluations when internal pressure was considered. Commercial software SIGNAL Fitness-For-Service version 3.0 by Quest Reliability was used for the analysis. The allowable final crack depth criteria per ASME BPVC Section VIII, Div.3, paragraph KD-412 was used to determine the allowable cycles for a particular load case. Figure 5 below shows the allowable design cycles using fracture mechanics analysis for the load cases considered.

Figure 5: Design cycles using fracture mechanics analysis for extreme load cases of the pressure-tension-bending moment capacity chart generated using linear elastic analysis

Application Two primary design methods have been presented and contrasted above, with both of the results approximating each other. During writing the paper and analyzing the example flange, a senior experienced analyst was used for approximately two months for analyzing, post-processing, verification and documentation. During this period, approximately 5 days were spent on the linear elastic and fatigue portion of the work. Approximately 35 days were spent on the elastic-plastic and fracture mechanics portion of the work. With both methods providing approximately similar results in both capacity and cycle life, the primary difference was a 600% increase in manpower needed to perform the elastic-plastic and fracture mechanics analysis. Even with the analysts experience, the elastic-plastic analysis results were often subjected to numerical instability, boundary condition influences, and interpretation of a point of non-convergence due to unbounded deformation.

These results must be accounted for during the planning of a work package where elastic-plastic and fracture mechanics

analysis is required or anticipated. Staffing needs and experience levels must be planned, and the results/capacities must be reviewed. With the proliferation of FEA software and the easy-to-use programs, the generation of results has become trivial.

0

5

10

15

20

0 50 100 150 200 250 300

Bor

e P

ress

ure,

[ksi

]

Bending Moment, [ft-kips]

T=0_LE

T=200_LE

T=400_LE

T=600_LE

T=0_EP

T=200_EP

T=400_EP

T=600_EP

4854 cycles 4109 cycles

8027 cycles 9448 cycles

OTC 23063 9

However, the establishment of the problem (boundary condition) and the interpretation of the results (non-convergence) are still extremely difficult. The burden of establishing, if a component is safe, shifts from the design engineer to the analyst, resulting in a historic reversal of roles.

Given the geometries associated with pressure containing equipment rated for greater than 15000 psi, the classification

and treatment of stress needs to be carefully evaluated. In many cases, the elastic-plastic methods presented above are the only appropriate tool to use for the definition of component capacity. This is due to the fact that for heavy walled vessels, especially around local or discontinuous regions, stress linearization is not appropriate. However, it should also be noted that for the evaluation of more standard load cases (eg, pressure-tension-bending); the linear-elastic methods provide equally conservative results. Both methods can be used together to provide efficient answers, with linear-elastic methods being used initially and elastic-plastic methods used to justify the linear-elastic assumptions.

Reviewing fatigue and fracture mechanics analysis for cyclic loading, both methods provide approximately equal results

for allowable design cycles. The primary contrast between the methods lies in the material data requirements. Fracture mechanics requires more material testing and characterization, even between the same material batches. While fatigue analysis relies on more generally accepted material assumptions with larger safety factors. Both methods are well posed to analyze ductile materials which are required for pressure containing subsea equipment.

Conclusions Several design analysis methods are given in this paper to verify the capacity of components or assemblies under consideration including protection against plastic collapse, protection against local collapse, and protection against cyclic loading. An approach to closure bolting is also presented. Both linear elastic and elastic plastic methods are given for the protection against local and plastic collapse. Using these methodologies, an example design analysis was presented for a 4 in. 20 ksi API 6BX flange.

The linear elastic and elastic-plastic method of analyzing the 4 in. 20 ksi API flange gave the same pressure rating when

only internal pressure was applied. The overall capacity of the flange for hydrostatic test condition is limited by bolt which is analyzed using linear elastic methods. When external loads are used to determine the combined structural pressure-tension-bending moment capacity of the flange, it was observed that the linear elastic methods gave conservative results in all cases. Examining only the capacity of the flange; allowing the bolt stresses to exceed allowables; there are cases where elastic-plastic methods gave lower capacity than the linear elastic methods. The overall capacity chart reported is comparable to the API TR 6AF chart and verifies the results obtained.

Fatigue and fracture mechanics analysis for cyclic loading is reported for the identified load cases on the pressure-tension-

bending moment chart. The reported cycles show that the flange design has a cyclic loading capacity which far exceeds the standard cycle requirements for subsea equipment.

Limitations of Work

The work and results presented in this paper are not meant to provide a definitive solution to building equipment for working pressures over 15000 psi. The purpose of this paper is to show that it is possible to combine the methods of both API 17D and ASME BPVC to achieve working pressures above 15000 psi since there is not a published opinion. The intention of this work is to show that existing API methods are adequate for working pressures up to 25000 psi, based on a 1.5*WP hydrotest. The recommendations presented in this paper are not meant to act as a new standard and in-depth analysis should be performed on each component or assembly to verify that the proper requirements are met. Acknowledgements We would like to specially thank Stuart Harbert and Paul Bunch from Cameron International Corp. for their valuable input regarding analysis using these design methods. An special thanks to Mike Payne from BP for all his teaching and support. Nomenclature API =American Petroleum Institute ASME=American Society of Mechanical Engineers AWHEM=Association of American Wellhead Equipment Manufactures BPVC=Boiler and Pressure Vessels Code CTOD=Crack Tip Opening Displacement FEA =Finite Element Analysis LRFD =Load and Resistance Factor Design NDE =Non-Destructive Examination PER =Pressures Exceeding a Rating of TR = Technical Report

10 OTC 23063

References American Petroleum Institute, Recommended Practice 6HT, 2005, Heat Treatment and Testing of Large Cross Section and Critical Section

Components, 1st Edition American Petroleum Institute, Specification 17D, 2011, Design and Operation of Subsea Production Systems Subsea Wellhead and Tree

Equipment, 2nd Edition American Petroleum Institute, Specification 6A, 2010, Specification for Wellhead and Christmas Tree Equipment, 20th Edition American Petroleum Institute, Technical Report 6AF, 2008, Technical Report on Capabilities of API Flanges Under Combinations of

Load, 3rd Edition American Society of Mechanical Engineers Boiler and Pressure Vessel Code, 2010, Alternative Rules for Construction of Pressure Vessels,

Section VIII, Div. 2 American Society of Mechanical Engineers Boiler and Pressure Vessel Code, 2010, Alternative Rules of Construction of High Pressure

Vessel, Section VIII, Div. 3 British Standards Institution, 2005, Guide to methods for assessing the acceptability of flaws in metallic structures, BS 7910 Payne, M., 2010, HP/HT Challenges, Journal of Petroleum Technology, pg. 70, April 2010 Edition Young, K., 2008, Characterizing Material Performance for Design of High-Pressure High-Temperature (HPHT) Equipment in

Accordance With API RP 6HP Practices, Technology Assessment & Research Program, BOEMRE.

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