For measurement of the gas temperature inside process piping, temperature sensors are installed in a thermowell. In most cases these cylindrical thermowells are installed perpendicular to the flow direction. Declining pressures of many aging oil & gas assets results in higher gas velocities. This will increase the so-called Strouhal frequency and amplitude of the dynamic forcing at the thermowell. If the frequency of this force coincides with the installed resonance frequency of the thermowell, this can lead to fatigue failure. The design of thermowells is comprehensively described in ASME PTC19.3 TW Code. In a recent version the excitation of the so-called in-line resonance condition was introduced, which already occurs at half the flow velocity of the original transverse resonance condition. As a result, many already installed thermowells do not pass the so-called ‘frequency criterion’ anymore. However, when all other criteria in the ASME Code are passed, the thermowell may operate at higher flow velocities, only if the fatigue stress remains acceptable. When calculating the fatigue stress with the tools as provided in the Code, in many cases this will lead to prediction of unacceptable fatigue stresses. Therefore, here a more detailed fatigue assessment method has been developed. It consists of in-situ vibration measurements at the tip of the thermowell and a numerical evaluation with a detailed FE model of the installed thermowell. With this approach the fatigue stress is then calculated and judged against a fatigue limit. In the meanwhile, this detailed assessment method has been successfully applied to numerous high pressure, gas filled installations. In most cases this led to acceptable fatigue stresses and preservation of the desired gas production profiles.

Flanged joints are used in high pressure applications such as process piping, pressure vessels, risers, pipelines and subsea production systems. These flanges are subjected to external loads in addition to pressure. A brief description of high pressure flanges standards is given. Design of high pressure flanged joints are covered in many design codes. A review of allowable stresses, load factors for bolting, flanges and bolt preload requirements has been made for the following codes: ASME VIII-2, ASME VIII-3, ASME B31.3 Chapter IX, API 6A, API 6X, API 17D, API 17TR7, API 17TR8, API 17G, EN 1591-1 and NORSOK U-001. This paper also presents analytically based structural load-capacity (ultimate strength) design equations for flanged joints. The design equations are used to calculate rated working pressure and flange-face separation load-capacity of API 6A type 6BX flanges. Future code recommendations for flange design are provided.

The use of segmented joint technology allows gasket manufacturers to fabricate and supply gaskets in discrete segments, rather than the conventional one-piece construction. Segmented gaskets are required when using sheet type gaskets for flange diameters larger than the available sheet size. Typical gasket sheets are 60” × 60” or 70” × 70” square, for instance. Flanges designed for these lower stress sheet gaskets that are larger than the available sheet size to fabricate the gaskets are typically fabricated as segmented gaskets. Segmented gaskets are also utilized in smaller flange applications to facilitate maintenance activities. When replacing the tube sheet gasket in a shell and tube exchanger the entire tube sheet must be removed in order to utilize a solid, one-piece ring gasket. In some cases this is not practical or desired, so a segmented gasket is utilized which only requires the tube sheet to be removed a short distance to allow access to the sealing surface. This same scenario exists with other common process piping and vessel flanges including thermowells, lance tubes, agitator or mixer shafts, etc. Valid concerns exist when using segmented gaskets as potential leak paths are created at every joint. If the gasket material and joint design are not correctly specified and fabricated the segmented gasket may not provide the same sealing performance and reliability as the solid, un-segmented gasket. This paper will discuss the research and data found that identifies the stress multiplier for segmented/jointed gaskets to achieve the same (similar) leak rate as their solid, unsegmented gasket counterpart. This assembly or design stress multiplier will be a dynamic number that will vary based upon the gasket material and the joining technology used. Two jointing methods are evaluated in this research; the industry standard “dovetail” and the Engineered Interference Tortuous Path (EITP) joint. In practice there is currently no guidance available to guide the design or assembly of these segmented gasket applications. The purpose of this paper is to provide guidance for both design and assembly purposes.

Whenever undesirable dynamic events occur within power plant, refinery, or process piping systems, specialty supports and restraints have the task of protecting the mechanical equipment and connecting piping from damaging loads and displacements. The array of components that may be affected include, but are not limited to, piping systems, pumps, valve assemblies, pressure vessels, steam generators, boilers, and heat exchangers. In particular, the dynamic events can be classified into two distinct types that originate from either internal events or external events. The internal dynamic load generating events include plant system start-up and shut-down, pressure surges or impacts from rapid valve closures such as steam and water hammer, boiler detonations, pipe rupture, and operating vibratory displacements that may be either low frequency or high frequency vibrations. The external dynamic load generating events include wind loads, earthquake, airplane impact to supporting structures and buildings, and explosions. Most of the aforementioned dynamic load generating events can be defined quite simply as impact loads, i.e., forces and moments that are applied over very short periods of time, for example, less than one second. While earthquake loads may be applied over a total time period of an hour or so, the peak loads and resulting displacements occur on a more sinusoidal basis of peak-to-peak amplitudes. One of the most common specialty restraint components utilized in the piping industry to absorb and transfer the dynamic load resulting from impact events is the hydraulic shock suppressor, otherwise known as the snubber. The snubber is a formidable solution to protecting plant piping systems and equipment from impact loading while not restricting the thermal displacements during routine operations. In the dynamic events that may be characterized by an impact type loading, snubbers provide an instantaneous, practically rigid, axial connection between the piping or other component to be secured and the surrounding structure whether it be concrete or steel (for example). In this way, the kinetic energy can be transmitted and harmlessly dissipated. In the vibratory environment, however, neither the impact load scenario nor the rapid translations are imposed upon snubbers, thereby presenting the competing intended application of the snubber to protect against impact loads versus, in many cases, the improper selection of the snubber to dampen vibratory (other than seismic) loads. The details of the hydraulic shock suppressor design are reviewed and discussed to exemplify why a case can and should be made against the use of snubbers in piping systems within an operating vibratory environment.

Commonly used piping vibration screening limits are typically justified by experience and lack a well-documented technical basis. This paper presents technical background for future Level 1 Fitness-for-Service (FFS) vibration screening criteria. The criteria assess the risk of fatigue in process piping due to bending mode type vibrations. Finite element analysis (FEA) of 20,000 randomly generated candidate-piping models and high-cycle welded joint fatigue curves for both constant amplitude and variable amplitude loading form the stress limits and basis for the proposed criteria. Most importantly, the proposed criteria aligns with historically used allowable vibration limits rooted in substantial experience. The allowable stress basis implemented in this paper considers periodic and random vibrations making it applicable to situations of mechanically induced, two-phase flow induced, turbulent-induced vibration of single-phase process fluid, or wind-induced, which may be manifested as either periodic or random. To reduce conservatism, limits are set for butt-welded and non-butt welded mainline piping to prevent use of a single blanket limit that may lead to unnecessary piping support alterations/additions, or costly piping configuration changes and unit downtime. Furthermore, the proposed Level 1-type criteria are consistent with previously proposed FFS Level 2 and 3 piping vibration fatigue evaluations [1] intended for inclusion in the ASME FFS-1/API 579 (API 579) Fitness for Service Standard [2].

Knife gate valves are common in process piping plants in the petrochemical industry. In this study a failure analysis is undertaken on a large knife gate valve attached on top of a Y-piece as part of a purge bin outlet in a polyethylene production facility. A large crack was discovered on the knife gate valve shortly after its installation and operation of the facility. The study outlines a methodology, based on the finite element method, to analyze failure of such units. It starts out by determining the stress intensification factors for a Y-piece by conducting an FEA, which are used in CAESER II piping software to determine the operational loads on the valve. The loads are used as inputs for a non-linear FEA model of the valve accounting for material failure based on continuum damage mechanics approach. The FEA results suggest that the operational loads are not the root cause to the observed failure on the gate valve. It is rather attributed to internal casting defects and the lack of post-heat treatment of the valve.

The use of steam as the fluid to clean steam lines in large commercial steam power plants is the most common and widely used method in the industry. The advantages of using air as the cleaning fluid can be significant and have positive impacts on both overall costs and schedule. While the general methodology for analyzing the air blow cleaning of process piping is very similar to steam blow cleaning, there are some additional computations and considerations needed.

The 2014 Edition of ASME B31.3, Process Piping [1], introduced significant changes to the post weld heat treatment (PWHT) requirements for P-No. 1 carbon steel materials. In particular, PWHT is no longer a mandatory requirement for any wall thickness provided that multi-pass welding is employed for wall thicknesses greater than 5 mm (3/16 of an inch) and a minimum preheat of 95°C (200°F) is implemented for wall thicknesses greater than 25 mm (1 inch). Detailed fracture mechanics analyses have shown that the lack of a mandatory PWHT requirement for thicker P-No. 1 components may result in a significant increase in risk for brittle fracture failures due to near-yield level weld residual stresses. Given the concern throughout the pressure vessel and piping community regarding potential brittle fracture failures, this updated PWHT guidance is examined. Impact testing requirements and exemption curves were introduced in the 1987 Addenda [2] of ASME Section VIII Division 1 (VIII-1) [3] in Paragraph UCS-66 and extended into ASME Section VIII Division 2 (VIII-2) [4]. During the VIII-2 rewrite in 2007 [5], the available technical and historical basis for the UCS-66 exemption curves was examined and improved to reflect modern fracture mechanics standards. The result of that effort was a systematic approach that can be modified for particular geometries and assumed flaws, if desired. The method used the most modern, fracture mechanics approach for welds in API 579-1/ASME FFS-1, Fitness-For-Service, (API 579) [6] based on the failure assessment diagram (FAD). As a result of explicitly accounting for weld residual stress, two separate sets of exemption curves are provided in VIII-2 [4]; one set for as-welded components and another set for PWHT components. In this paper, a similar approach is summarized to generate exemption curves by establishing newer as-welded and PWHT curves using the Fracture Toughness Master Curve (Master Curve) as documented in upcoming Welding Research Council (WRC) Bulletin 562 [7]. The increased propensity for brittle fracture in as-welded components versus PWHT components is clearly highlighted using this approach. The Master Curve, in conjunction with the elastic-plastic fracture mechanics employed in API 579 [6] provides a means to develop exemptions curves anchored in state-of-the-art fracture toughness technology that can be directly tied to different reference flaw sizes. Additionally, commentary on the appropriateness of the current ASME B31.3 [1] PWHT requirements is offered and the effectiveness of using weld preheat in lieu of PWHT as permitted in the National Board Inspection Code (NBIC) [8] is examined using simplified computational weld analysis.

The 2012 edition of ASME B31.3 Process Piping Code [1] included an update to the fatigue assessment rules which limited the allowable stress value used in the fatigue assessment to 138 MPa (20 ksi). This change was made due to the knowledge that higher strength materials, especially those materials with yield strength to tensile strength ratios above 0.8, do not tend to exhibit significantly longer fatigue life as compared to lower strength materials. This change means that for existing high-strength piping in plants that were designed using pre-2012 B31.3 [2] and are in fatigue service, the end-user must now recalculate the piping remaining life based on new requirements. Depending on the size of the plant under consideration and the level of high-strength piping used, such a recalculation may be an extensive scope of work. This paper uses a comparison between the rules in ASME B31.3 and the fatigue life prediction methods outlined in API RP 579-1 / ASME FFS-1 [3] in order to establish a categorization methodology which is used to identify which pre-2012 high strength piping systems will require re-assessment, more frequent monitoring and/or may have a shorter than anticipated life than the original design basis. The methodology allows for easy categorization of the piping systems to allow efforts to be focused on the most appropriate piping systems. The need for this work came from concerns regarding the extensive use of Duplex Stainless Steel piping on a large offshore platform that was designed prior to ASME B31.3 2012.

Vibration induced fatigue failures (VIFF) of pipework small bore connections (SBCs) due to turbulent flow excitation continue to occur in process piping systems, resulting in elevated safety risks and costly interruptions to normal operations. One of the primary causes of these failures is poor design. When the valve and pipe components in an SBC assembly combine to produce a fundamental natural frequency that is low, the SBC can readily be excited into a high vibration condition. In this high vibration condition, the threat of VIFF increases significantly. One of the primary design goals for SBCs should therefore be to achieve a high natural frequency, however, this is very rarely considered in the existing design process. Reasons for this include the lack of understanding of the importance of natural frequency to the vulnerability of the design and a lack of information on how high this target natural frequency must be when it is considered. In this paper, the concept of a cut-off natural frequency is defined, above which the probability of VIFF of SBCs is negligible. If this concept is incorporated into the design process for SBCs a decrease in the number of failures of SBCs due to vibration should be realised.

Motivated by the response of industrial piping under seismic loading conditions, the present study examines the behavior of steel process piping elbows, subjected to strong cyclic loading conditions. A set of experiments is conducted on elbow specimens subjected to constant-amplitude in-plane cyclic bending, resulting into failure in the low-cycle-fatigue range. The experimental results are used to develop a low-cycle-fatigue curve within the strain-based fatigue design framework. The experimental work is supported by finite element analyses, which account for geometrical and material nonlinearities. Using advanced plasticity models to describe the behavior of elbow material, the analysis focuses on localized deformations at the critical positions where cracking occurs. Finally, the relevant provisions of design codes (ASME B31.3 and EN 13480) for elbow design are discussed and assessed, with respect to the experimental and numerical findings.

The 2010 Edition of the American Society of Mechanical Engineers (ASME) B31.3 Procecess Piping Code [1] includes a new chapter: Chapter X High Purity Piping. Chapter X covers industries listed in the scope of ASME B31.3 which use methods of fabrication, examination, inspection and testing different than other industries covered by the Code. Industries which have a need for cleanness and cleanability on a more demanding level, such as the semiconductor industry, which uses Semiconductor Equipment and Materials International (SEMI) Standards [2–4], and the pharmaceutical and bioprocessing industries, which use the ASME Bioprocessing Equipment (BPE) Standard [5], also reference ASME B31.3 for safety requirements. ASME B31.3 now addresses issues common to the semiconductor and biopharmaceutical industries. The new High Purity Fluid Service defined in Chapter X permits weld coupon examination in lieu of the 5% radiography required in Normal Fluid Service when orbital welding is used in fabrication. Industries that may not otherwise be considered as high purity, such as refineries, the chemical processing industry [6–7], solar panel fabrication and nuclear or petrochemical applications that could use tubing rather than pipe, may benefit from the fabrication technology introduced in Chapter X while meeting the safety requirements of the Code.

Pipe bends are common elements in piping system such as power or process piping, and local thinning are typically occurred on pipe bends due to erosion or corrosion. Therefore, it is important to establish the plastic collapse condition for pipe bends having a local thin area (LTA) under combined internal pressure and external bending moment. In this paper, a simplified plastic collapse assessment procedure in p-M (internal pressure ratio and external bending moment ratio) diagram method for pipe bends with a local thin area simultaneously subjected to internal pressure, p , and external out-of-plane bending moment, M , due to earthquake, etc., is proposed, which is derived from the reference stress. In this paper, only cases of that an LTA is located in the crown of pipe bends are considered. The plastic collapse loads derived from the p-M diagram method are compared with the results of both experiments and FEA for pipe bends of the same size with various configurations of an LTA.

Appendix P Alternative Rules for Stress Range Evaluation was first introduced into the 2004 edition of ASME B31.3 Process Piping Code [1] as an alternative method that was conceived to be more computer friendly and possibly a more nearly theoretically defensible method of protecting piping systems individually against fatigue failure and dimensional ratchet than the singular stress range requirements of the Code’s base text (“base Code”). This paper describes the progression of events that led up to the inception of the Appendix and the subsequent revisions to same and detailing the reason for the 2010’s newly coined Allowable Operating Stress to protect against dimensional ratchet and Allowable Operating Stress Range to protect against fatigue failure.

For more than fifty years the oil refining industry has been using American Society of Mechanical Engineers (ASME) B31.3 “Process Piping” for the design of piping systems in hydrogen-containing services. In 2008, ASME B31.12 “Hydrogen Piping and Pipelines” was released, which caused uncertainty within the refining industry about which Code to apply in these services. The foreword of B31.12 states that it “applies to design, construction, operation, and maintenance requirements for piping, pipelines, and distribution systems in hydrogen service. Typical applications are power generation, process plants, refining , transportation, distribution, and automotive filling stations.” Typical refinery services containing hydrogen include a gamut of applications. Some operate at high pressures and/or temperatures; some also contain water, hydrogen sulfide (H 2 S) and/or other corrosives and others are pure hydrogen at mild conditions. This paper describes the various services along with a summary of associated material degradation or cracking mechanisms and the measures used to prevent piping failures. This is followed by a discussion about whether the more rigorous requirements of B31.12 provide safeguards for these potential degradation mechanisms. A comparison of the two Codes in design, materials, fabrication and non-destructive examination (NDE) requirements is provided along with qualitative estimates of the cost differences. This comparison should help new design and construction projects choose which Code to apply. The paper also has suggestions for clarifying the scope of B31.12.

For the first time in the history of the ASA [1], ANSI/ASME, ASME/ANSI, to the 2008 edition of ASME B31.3 Process Piping Code (hereafter referred to as “the Code”) [2], an equation for the Stress Due To Sustained Loads, S L , has been introduced into the Code. From its inception to the current edition, the Code has remained ambiguous in several areas. Even though the displacement stress range, S E , has been explicitly defined by both text and equation, the stress due to sustained loads, S L , is mentioned not with an explicit equation but with a statement that S L shall be limited by the allowable stress at the corresponding operating temperature, S h . This paper describes the equation in detail and the background, the events, and the effort involved that led to the insertion of this equation into the Code.

Nowadays, time saving in piping stress analysis is a major concern among consultant engineering companies in power and process plants. In this paper, we are going to have a quantity review between some important parameters such as sustained, occasional and expanding stresses with total length and nozzle load of a real case according to ASME B31.3. Most of the times, piping stress designers try different and more flexible piping routes completely arbitrary. Longer piping route means greater mass, and it means a big trouble in earthquake time especially for allowable nozzle loading, on the other hand shorter piping route increases thermal nozzle loading and operating stresses. We will try to find a relationship among mentioned parameters to have an optimal piping route in order to save time and material. Finally, different routes will be analyzed using Caesar II to plot related parameters and find optimum criterion.

A significant amount of work has transpired over the past several years in generating consensus-based standards that include ASME PCC-2 and ISO 24817 for developing composite repair systems. The intent in developing these standards has been to provide industry with guidelines for designing composite repair systems to ensure that damaged piping and pipelines are safely and properly reinforced. With the numerous composite repair systems currently available to operators, the importance of evaluating the capabilities of each system cannot be overstated. The fundamental design variables available to manufacturers are stiffness, strength, and thickness of the composite. A properly-designed repair system ensures that strains in the reinforced steel and reinforcing composite material do not reach unacceptable levels. This paper provides a basic overview of the design philosophy embedded into the current design codes, as well as presenting results associated with several specific studies that were conducted to evaluate composite repair performance.

The American Standards Association (ASA) B31.3-1959 Petroleum Refinery Piping Code [1] grew out of an ASA document that addressed all manner of fluid conveying piping systems. ASA B31.3 was created long before widespread engineering use of computer “mainframes” or even before the inception of piping stress analysis software. From its inception until recent times, the B31.3 Process Piping Code [2] (hereafter referred to as the “Code”) has remained ambiguous in several areas. This paper describes some of these subtle concepts that are included in the Code 2006 Edition for Appendix S Example S3. This paper discusses: • the effect of moment reversal in determining the largest Displacement Stress Range, • the impact of the average axial stress caused by displacement strains on the Example S3 piping system and the augmenting of the Code Eq. (17) thereto, • a brief comparison of Example S3 results to that of the operating stress range evaluated in accordance with the 2006 Code Appendix P Alternative Requirements.

The 2004 edition of ASME B31.3 Process Piping Code [1] introduced both Appendix P and Appendix S Example 1; Examples 2 and 3 were included in the 2006 edition. Appendix P is to illustrate a nearly computer stress analysis basis by defining alternative rules for displacement stress range evaluation, and Appendix S contains examples describing how to satisfy the requirements of B31.3 when performing computer based stress analysis. As is possible with any extensive addition to any code, these first offerings contained shortcomings either explicitly or by lack of clarity. This paper discusses some of these issues, provides workarounds when using today’s commercially available piping stress analysis software, and proposes revisions to both Appendices for the 2010 edition of the B31.3 Process Piping Code.