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Pressure Vessels: Types, Design, Supports, Applications, Materials

Written by Anup Kumar Dey in Mechanical,Process

What is a Pressure Vessel

A pressure vessel is a closed leak-tight container (normally cylindrical or spherical) designed to hold fluids (i.e, gases, liquids, or two-phase fluids) at a pressure substantially different (higher or lower) from the ambient pressure. They are usually made from carbon steel or stainless steel and assembled from plates by welding method. However, other materials like Aluminium, copper, non-metals, etc also find usage as pressure vessel material in some specific situations. Even though most pressure vessels are basically long cylinders with two heads at both ends, they can take the shape of a sphere or cone.

Examples of some widely used pressure vessels are columns, boilers, separators, knock-out drums, Towers, Bullet Tanks, Reactors, and Heat Exchangers. All these pressure vessels are extensively used in the oil and gas, petroleum refining, and chemical/petrochemical processing industries, and power plants have varying operating pressures and temperatures. However, note that an atmospheric storage tank is not a pressure vessel. Fig. 1 shows a typical pressure vessel used in the oil & gas industry.

What are the Types of Pressure Vessels?

Pressure vessels are classified based on various different parameters which are covered here.

Based on their installation/orientation they are classified into the following two groups.

  1. Vertical pressure vessel
  2. Horizontal Pressure Vessel

Various methods are used to support pressure vessels, like

Types of ends attached to the vessels are

  • Dish ends
  • Conical ends
  • Flat Ends

Fig. 2 shows the general configuration of a pressure vessel.

Typical Pressure Vessels
pressure vessels used in a process plant
Fig. 1: Pressure Vessels used in a process plant

Inputs for the Design of Pressure Vessels

The design of pressure vessels must be done with utmost care as these operate under immense pressure. A ruptured pressure vessel can cause serious irreversible harm to mankind and properties. Normally the ASME Sec VIII code governs the design of pressure vessels.

The following inputs are required for pressure vessel design

Pressure Vessels Design Formula

The basic formula for designing the cylindrical shell is

σ = PD/2t

Therefore,           t = PD/2σ

Where,

  •                 t = thickness of the shell
  •                 P = internal pressure
  •                 D = diameter of the shell
  •                 σ = tensile stress
General Configuration of a typical pressure Vessel
Fig. 2: General Configuration of a Typical Pressure Vessel

This basic formula is modified in international design codes.

For ASME BPVC Sec VIII, the thickness of the cylinder is calculated by the following formula (Fig. 3)

Formula for calculation of pressure vessel shell thickness
Fig. 3: Formula for calculation of pressure vessel shell thickness

Where,

  • t = thickness of the shell
  • P = internal pressure
  • R = radius of the cylinder
  • S = tensile stress
  • E = joint efficiency

Refer to ASME Sec VIII Div 1 for design formulas for all sections of the vessel/cylinder. The following figure provides a typical flow chart for pressure vessel design steps.

Flow Chart for Pressure Vessel Design Steps
Fig. 4: Flow Chart for Pressure Vessel Design Steps

Type of Pressure Vessel Supports

Based on how the pressure vessels are supported, they can be of the following types.

  • Vessel Supported on Lug Support (Fig. 5)
  • Vessel Supported on Skirt(Fig. 5)
  • Vessel Supported on Leg(Fig. 5)
  • Vessel Supported on Saddle(Fig. 6)
types of pressure vessel supports
Fig. 5: Various types of supports for pressure vessel
Pressure Vessel Supported on Saddle
Fig. 6: Pressure Vessel Supported on Saddle

Pressure Vessel Parts

A pressure vessel consists of the following parts

  • Shell, head
  • Nozzles
  • Flanges
  • Gaskets
  • Internals
  • Platforms & ladders
  • Skirt or saddles
  • Baseplate

Click here to know more about major pressure vessel parts.

Design codes used for Pressure Vessel Design

There are various codes that are used for pressure vessel design, construction, and maintenance purposes. Some of them are listed below:

  • EN 13445: The current European Standard, harmonized with the Pressure Equipment Directive (97/23/EC). Extensively used in Europe.
  • ASME Code Section VIII, in addition, is supported by Sections II (materials), V (NDT/NDE), and IX (welding). Published by the American Society of Mechanical Engineers.
  • ASME Code Section VIII Division 1: US standard, design by the formula. Almost exclusively used in North America, widely used worldwide.
  • ASME Code Section VIII Division 2: Alternative Rules, design by analysis.
  • ASME Code Section VIII Division 3: Alternative Rules for Construction of High-Pressure Vessel
  • BS 5500: Former British Standard, replaced in the UK by EN 13445 but retained under the name PD 5500 for the design and construction of export equipment.
  • AD Merkblätter: German standard, harmonized with the Pressure Equipment Directive.
  • EN 286 (Parts 1 to 4): European standard for simple pressure vessels (air tanks), harmonized with Council Directive 87/404/EEC.
  • BS 4994: Specification for design and construction of vessels and tanks in reinforced plastics.
  • IS 2825-1969 (RE1977): code unfired Pressure vessels

Pressure Vessel Internals

Internals are used to separate the liquid from a mixture of liquid & vapor. Refer to Fig. 6

Pressure Vessel Internals.
Fig. 6: Pressure Vessel Internals.

Example of Pressure Vessels

Fig. 7 shows a few of the various types of pressure vessels that are normally used in plants.

various types of pressure vessels in industrial use
Fig. 7: Various types of pressure vessels.

Material of Construction of Pressure Vessel

For industrial applications, the following materials are widely used for pressure vessel construction. However custom-made pressure vessel fabrication is possible depending on requirements.

Applications of Pressure Vessel

Pressure Vessels are most widely used in the following sectors

  • Refinery and Petro-chemical
  • Fertilizer
  • Oil and Gas
  • Chemical
  • Power Plants

Regular inspection of pressure vessels is very important to avoid pressure vessel failures. API 510 provides guidelines for pressure vessel inspection.

The following video provides a comprehensive basic of the pressure vessel design methodologies:

Mechanical Aspects of Pressure Vessel Design

Few more Resources for you.

Brief Explanation of Major Pressure Vessel Parts
A Presentation on VESSEL CLIPS or VESSEL CLEATS
10 points to keep in mind while using project-specific pressure vessel nozzle load tables during stress analysis
Understanding Pressure and Temperature in the context of Pressure Vessel Design
A video tutorial on “Comprehensive Design Code Coverage for Pressure Vessel and Heat Exchanger Design” by Bentley Institute

Reference

https://faculty.washington.edu/vkumar/me356/pv_rules.pdf

Online Course on Pressure Vessels

If you wish to learn more about Pressure Vessels, their design, fabrication, installation, etc in depth, then the following online courses will surely help you:

Several ASME B31 & EN 13480 Issues Needed to Know for Pipe Stress Engineer

Written by Alex Matveev in Piping Stress Analysis,Piping Stress Basics,Start-Prof

Axial Stress from Pressure Load, Axial Force and Torsion Moment

This issue leads to serious under-estimation sustained and expansion stresses in pipes and fittings in ASME B31.1-2018, ASME B31.4-2016, ASME B31.5-2016, ASME B31.8-2016, EN 13480-2017, ISO 14692, and some other codes.

For example, let’s take ASME B31.1-2018 code that requires using the following equation for stress from sustained loads (104.8.1):

The first problem is not so serious for above ground piping. This equation is suitable only for totally unrestrained pipes, but it is easy to use in manual calculations, and it is always conservative. For more information please refer to the article What is Restrained and Unrestrained Pipes. Also in software, the more accurate equation is recommended to use for axial stress 102.3.2 a (3):

Instead of a simplified equation, that is easier for manual calculations

The second and most serious problem is with the formula (104.8.1). It doesn’t take into account axial force in the pipe. Sometimes, the axial force from sustained loads can be so huge, that axial stress becomes greater than allowable. Engineers can easily overlook this problem when using pipe stress analysis software for big models. For example, it could be very tall vertical pipes or heavy valves on the vertical pipes. The code sustained stress for this model is almost zero (see “Sl” stress in the screenshot below). But real sustained stress is greater than allowable (see “Sl*” stress in the screenshot below).

The same problem with the expansion stress equation (104.8.3), it doesn’t include axial force too

For example for totally restrained pipe code expansion stress range will be zero (see “Se” stress in the screenshot below). But real stress range is greater than allowable (see “Se*” stress in the screenshot below).

Sometimes even experienced piping designers can make a mistake and create a wrong design. For example in the piping system below, the 1-2 pipe is restrained by trunnion 2-4. Code stress range is zero, but if we choose the option to include axial force the stress range is greater than allowable!

To protect users from such mistakes we add the special option in PASS/Start-Prof software that allows taking into account stress from the axial force and torsion moment for sustained stress and for expansion stress range. Users can simultaneously see official code stresses (Sl, Se) and modified stress (Sl*, Se*) in the same table. It also automatically solves the first described problem with axial stress from pressure load for restrained, totally unrestrained, and middle behavior systems. Axial stress from pressure load (Sl*) will also be more accurate. This option may be activated using the “Add axial force and torsion stress” checkbox in Project Settings.

For modified stresses PASS/Start-Prof software use the equations similar to ASME B31.3 code:

We recommend always switching this option on!

Effective Tee Branch Section Modulus Issue in ASME B31.3-2016

This issue led to serious under-estimation of sustained and occasional bending stresses at reducing intersections.

ASME B31.1-2018 code requires to calculate the bending stress from thermal expansion for reducing tee branch by equation (all symbols are taken as per ASME B31.3-2016 code for better understanding):

Z is section modulus, where effective branch wall thickness is

 is the thickness of pipe matching branch,

 is the thickness of the pipe matching run of tee or header.

What is the effective branch wall thickness? The answer is given in L.C. Peng’s “Pipe Stress Engineering” book, 4.5.1. The idea is the following. Bending stresses must be checked in two potential locations of failure:

  • In branch and header pipes junction, zone 1
  • in branch pipe next to the junction, zone 2

The failure will happen in zone 1 if the stress intensification factor is high. But if the stress intensification factor is not significant, the failure will happen in zone 2. The code should check the bending stress in both locations.

ASME B31.1-2018 use simplified equation 104.8.3, 104.8.4C:

Where effective branch wall thickness is

For sustained stress equations for 1 and 2 zones should be:

The simplified equation in ASME B31.1-2018 104.8.1, 104.8.4

Now let’s check ASME B31.3-2016 code. As it uses two stress intensification factors (in-plane and out-plane), the equation for bending stress in locations 1 and 2 should be:

But ASME B31.3-2016 code offers for reducing tee bending stress from sustained loads the equation 23b2

Sustained in-plane and out-plane SIFs:

Effective section modulus of branch

Effective wall thickness is determined according to 319.4.4 (c)

If we convert these equations back to the stresses in 1 and 2 locations, we will have

For location 2 stresses become incorrect:

Let’s assume that Mo=0, in this case, we get underestimated by 25% bending stress in the branch pipe

Instead of correct stress value

If we take Mi=0, we get bending stress in the branch pipe

If we assume that io=1.4, then we get bending stress underestimated by 19%

To partially fix this problem we should change the equation 23b2 for effective wall thickness in 320.2 to

But in this case, the out-of-plane bending stress in branch connection will be overestimated. If we assume that io=3.0 the out-of-plane bending stress in branch pipe will be overestimated 25%:

To fix this problem we should correct the bending stress equation in 320.2 to

Zh – Header section modulus, Zb – Branch section modulus.

Unfortunately, these corrections can’t be made in the software. The ASME B31.3 code revision needed.

Fortunately, a new code ASME B31J-2017 has been released. Using this code we can bypass this problem. We recommend activating the “ASME B31J” option in PASS/Start-Prof software.

If this option is activated, the accurate section modulus will be used for header and branch pipes:

Also if “ASME B31J” option is activated then all tees are automatically modeled with simultaneous use of run and branch springs with flexibilities and stress intensification factors calculated according to ASME B31J code requirement:

Start-Prof software allows activating the “ASME B31J” option for all ASME B31 codes and EN 13480! If some of k-factor becomes less than 1.0 Start-Prof assumes this spring as rigid:

Real Bend Wall Thickness is Greater than the Matching Pipe Wall Thickness

This is a very serious issue. It leads to an underestimated pump, nozzle, and support loads and stresses calculated by ASME B31 codes!

ASME B16.9 and all ASME B31 codes don’t regulate the bend, tee, and reducer wall thickness. Only the pipe wall thickness is regulated. So many people think that the elbows and other fittings have the same or almost the same wall thickness as the matching pipe. But in most cases, the real bend, tee, reducer body wall thickness is greater than matching pipe wall thickness with the same Schedule.

For elbows, the real wall thickness can be 10%-40% greater than the matching pipe. Because bends must have greater wall thickness to hold the same pressure as the connected straight pipe (see 304.2.1 3d).

Manufacturers usually produce the bends with a greater wall thickness than matching pipe, but we can get real bend wall thickness only after contacting the manufacturer or even measure it after delivery.

Click to enlarge

Piping designers usually know nothing about it. And piping stress engineers usually use the pipe wall thickness for elbows when using piping stress analysis software. Leaving the “Fitting Thk” field blank makes software thinking that elbow has the same wall thickness as a connected pipe element. This is a serious mistake!

Click to enlarge

According to the ASME B31 and other ASME B31-based codes bend flexibility factor depends on real bend wall thickness, not on matching pipe wall thickness.

The greater bend wall thickness, the greater is bend stiffness (k-factors) and greater are loads on rotating equipment, nozzles, supports and expansion stresses in piping system.

This problem quite often comes to light when Russian companies try to check the design made according to ASME B31 codes for the Russian market. While rechecking the stress analysis using PASS/START-PROF software according to GOST codes a lot of error messages appear. They say that the wall thickness of the elbows is lower than the minimum required one to hold the pressure because it is usually left blank in CAESAR II and software takes fitting thickness equal to connected pipe WT in the piping stress model. When the real elbow wall thickness entered and model recalculated, the nozzle loads and stresses become much greater than it was calculated in CAESAR II and other software! That’s because the elbow flexibility k-factors used during analysis was incorrect.

But in real practice counterparts usually can’t provide the real bend body WT. They just don’t have this information!

We received different answers to the direct question – “what will be the real wall thickness throughout the whole bend body?”

  1. Some manufacturers do not answer at all;
  2. Others say it’s a trade secret (?!);
  3. Some manufacturers replied that the edge will be 100% consistent with that ordered according to ASME (according to schedule), but the thickness of the wall at the bend may even be 40% greater!

Only after we receive the ordered fittings (bends, reducers, and tees) from the factory, only at this time we can measure and find out what are the real wall thicknesses. The stress analysis model should be changed, nozzle loads become greater. The design should be changed to add more flexibility and reduce nozzle loads and sometimes expansion stresses. And all of this should be done after the design job by our contractors was formally “finished”. Amazing!

For example, Russian standards, which are completely different from ASME B16.9 for bends, tees, and reducers, always provide the real body wall thickness for each fitting. Manufacturers follow the standards. Every piping stress engineer knows the real body wall thickness of bends and other fittings and specifies it in START-PROF while performing piping stress analysis. Also, all the bend properties can be taken from the fitting database (see screenshot below).

Click to enlarge

All RD, GOST, and SNiP stress analysis codes (power, process, oil & gas main pipelines, etc.) provide detailed wall thickness calculation procedure for all fittings including bends, tees, and reducers. On screenshots below, you can see that the calculated bend wall thickness is always greater than pipe wall thickness for the same pressure load.

The real bend body wall thickness should be used in piping stress analysis instead of matching pipe wall thickness. To solve this problem we added the special feature in PASS/START-PROF software that allows calculating the approximate “real” bend wall thickness on-the-fly according to ASME B31.3 304.2.1 and the same requirements in other ASME B31 and EN 13480 codes. Just push the button “C” near the “Wall Thickness” field and it will be calculated according to the code requirements.

Conclusion

  1. Bend, tee and reducer wall thickness should be regulated by ASME B31 codes and provided in ASME B16.9 code. Manufacturers should produce bends with body wall thickness according to the code requirements.
  2. Until the first problem is solved, the manufacturers should provide bend wall thickness in their catalogs. It will allow designers and piping stress engineers to use the real WT in the pipe stress model and to get accurate nozzle loads and expansion stresses.
  3. There should be a special remark in ASME B31 codes that explains how to calculate flexibility k-factors for the elbows if the real body wall thickness is unknown.
  4. If the elbow wall thickness is unknown, then piping stress engineers should use WT calculated by ASME B31 code equations for bend or use pipe wall thickness multiplied by 1.4 factor. This will provide more conservative design, and after the real bend, wall, thicknesses will become available (can be measured) the changes in piping design will not be as critical, as now.

To listen directly from the author and learn refer to the following embedded video:

Applicability of Caesar II for stress analysis of lines having D/t ratio more than 100

Written by Anup Kumar Dey in Caesar II,Piping Stress Analysis

Sometimes while working you may have come across situations when the pipe diameter to thickness ratio becomes more than 100. These normally happen for very large-size low-pressure systems. Due to low design pressure (of the order of 1 bar to 3 bar) calculated thickness is less. For example, consider a flare line of 42-inch (or 48 inches) NPS with 3-bar design pressure. The selected thickness is STD wall thickness i.e, 9.525 mm. So the Diameter to thickness ratio is more than 100. Normally Caesar II is widely used for stress analysis of all piping systems. But when you use the software for stress analysis of such stress systems you will find a warning message similar to as shown in Fig. 1. This warning message is generated for all Bend and Intersections in the pipe.

Warning message of Caesar II while stress analyzing line with D/t greater than 100
Fig. 1: Warning message of Caesar II while stress analyzing line with D/t greater than 100

So whether you should proceed with pipe stress analysis ignoring such warning messages. ASME B31.3 does not provide any equation for calculating the SIF values for lines with D/t exceeding 100 (Also known as thin-walled pipes). This restriction on D/t comes from the B31.1 and B31.3 piping codes because they do not have computations for stress intensification factors above this threshold, largely because the testing that was done by Markl had this same limitation. CAESAR II will continue to use the existing code formulations for SIF calculation when a D/t ratio is greater than 100, but no one knows if these relations are still valid. That is the reason Caesar II displays such a warning.

CAESAR II is a pipe flexibility analysis software package and as with all such packages, there are limitations that should not be exceeded. It is always better to use some sort of FEA analysis for analyzing such systems and you will get accurate results. But considering the number of such stress systems in any project (hardly one or two) purchasing costly FEA software may not get the project approval. So in such a situation, you may not have other options than to use Caesar II. So the analyst can use Caesar II, but he needs to always bear the following points in mind:

Failure of thin wall piping is dominated by the buckling phenomenon and the validity of the flexibility analysis assumptions decreases. CAESAR II and other such flexibility analysis programs (CAEPIPE, AUTOPIPE, etc.) do not compute buckling because it is a localized effect caused by imperfections in the manufacture, uneven corrosion, pitting, or other defects. CAESAR II considers the pipe to be homogeneous in cross-section with an even wall thickness everywhere. So the analyst is encouraged to consider buckling separately from the CAESAR II analysis for such systems.

Another problem with the very thin-walled pipe is local deformation in the region of restraint. When significant local deformation is likely at restraint locations, the load distribution will no longer follow expectations as with homogeneous cross-sections and standard flexibility analysis results should not be relied on. Normally it is a standard practice to use wear pads in all support locations to reduce the possibility of local deformation at restraint points.

It is better to use a lowered allowable stress than what is allowed by the piping code in use. For example, the analyst can reduce the allowable value by 75% of the code allowable.

It is better to multiply the Caesar calculated SIF at intersections by 2.5-3.0 and input the same manually at all three node points (Branch and header).

So whether to use Caesar II or not solely depends on the Pipe Stress Engineer. He must discuss the situation with the client and if the client approves then he may proceed with the same. However, recent FEA packages that come along with the latest Caesar II software editions can accurately calculate the SIF values and can be used.

Few more Resources for you..

Stress Analysis Basics
Stress Analysis using Caesar II
Stress Analysis using Start-Prof
Piping Design and Layout

Valve Inspection & Testing | API 598 | Third Party Valve Inspection

Written by Anup Kumar Dey in Engineering Materials,Pipeline,Piping Design Basics

Valve Inspection and Testing is an important subject for piping professionals as it ensures the integrity and performance of Valves during plant operation. During the shutdowns of operating plants, hundreds of valves tend to have various needs to be inspected. Also before use in new plants (during the design phase), the valves must be inspected thoroughly in the manufacturer’s shop to ensure proper quality.

API STD 598- Valve Inspection and Testing provides guidelines for the inspection and testing of valves and the test is conducted by the valve manufacturer. This Valve Inspection article will provide you with some important points related to the inspection of valves and valve testing in the manufacturing shop as well as in operational plants.

Normally the following tests and examinations are performed on Valves/Components.

Valve Inspection and Testing

Shell Test

Required for all valve types. The shell test validates the strength and soundness of the valve pressure-containing structure. This is actually a pressure test where normally air, inert gas, water, kerosene, etc can be used as the test fluid. The minimum shell test pressure that can use are listed in table 2 of API-598. The pressure is applied inside the assembled valve with the ends closed. The valve shall be partially open and be able to hold pressure for a certain time. The pressure shall not be less than 1.5 times of maximum working pressure. No leakage is allowed.

Backseat Test

Required for all valves that have the backseat feature, except for bellows seal valves. This is a feature that allows valve packing to be replaced while it is in service. Normally gate valves, globe, and parallel slide gate valves are subjected to this test. This pressure test is performed to verify leakage past the stem or shaft to the bonnet seal i.e, backseat. Table 3 of API-598 provides the required test pressures for the backseat test for different valve types. In this test, the valve shall be fully opened and the packing gland shall be loose or not installed. No leakage is permissible in the backseat test.

Low-pressure closure test/ High-pressure closure test

Closure tests are performed to confirm leakage past or through a valve’s closure mechanism. The Closure Test or Valve Seat Leak Test is performed after the successful completion of the Valve Body Test/shell test. One side of the valve inlet or outlet shall be subjected to the hydrostatic pressure and the amount of leak shall be measured on the opposite side of the valve.

To explain it, if you are pressurizing the valve inlet, then you have to measure the amount of leakage in the outlet. For both low-pressure and high-pressure tests leakage through the disc, behind the seat rings, or past the shaft seals is not permitted. However, a limited amount of leakage is permissible at the seat-sealing surface interface which is listed in table 5 of API-598.

Double block and bleed (DBB) high-pressure closure test

Pressure shall be applied successively to each side of the closure and Leakage into the body cavity shall be checked.

Visual examination of castings

A visual examination must be performed on all castings to ensure conformance with MSS SP-55.

High-pressure pneumatic shell test

When specified in the purchase order, a high-pressure pneumatic shell test shall be performed. This test shall be performed after the shell test, using appropriate safety precautions. The pneumatic shell test pressure shall be 110 % of the maximum allowable pressure at 38 degrees C (100 degrees F) or as specified in the purchase order. Visible leakage is not allowed.

Table 1 of API 598 specifies the pressure test requirements for various types of Valves. The test equipment should not apply external forces that affect seat or body seal leakage. The following figure shows some sample test equipment used for valve testing.

Valve Test Equipments
Sample Test Equipment

Codes and Standards for Valve Inspection and Testing

Valve inspection and testing must be performed for the smooth and safe operation of industrial processes. Major codes and standards that guide the inspection and testing procedure of a valve are

  • API 598
  • API 607
  • API 6D
  • BS 759

Third Party Inspection for Valves

Third-party valve inspection is normally carried out by the manufacturer, purchaser, and the representative from the third-party inspection agency. Usually, a list of documents is required to be produced before the valve inspector for his review. These documents are:

  • Purchase Order of the Valve
  • Valve Manufacture Quality Control Plan
  • Valve Inspection and Test plan
  • Data Sheet of the Valve
  • All Approved Drawings drawings for the valve
  • Valve Material Test Reports
  • Welding Specification Procedures(WPS) and Procedure Qualification Records (PQR) for the Valve.
  • Valve Welders Qualifications Reports
  • Valve NDE Personnel Qualifications Reports
  • Required NDE Procedures like Dye Penetration, Magnetic Particle, Radiographic, Ultrasonic, PMI Testing, etc. for the Valve.
  • Heat Treatment Procedure of the Valve material.
  • Valve Calibration Certificates for Test Equipment
  • Hydrostatic Testing Procedure of the valve to ensure leak tightness.
  • Closure Testing Procedure, Backseat Testing Procedure, and Water Quality Document
  • Valve Preparation and Painting Procedure
  • Valve Preservation, Packing, and Shipping Procedure
  • Valve Packing List

Valve Material Inspection by Third Party

The original or authenticated copies of valve material mill certificates need to be available with the manufacturer. The representative of the third-party inspector examines these material certificates for compliance with design specifications or drawings. The review includes inspection and checks on:

  • Certificate No.
  • Heat or cast No.
  • Chemical composition.
  • Mechanical properties.
  • Heat-treated condition.
  • NDE applied and results.
  • Surface finish

Valve material includes the body stem material, shell material, valve trim materials, ball, wedge or flap materials, operating components, support material, gland materials, anti-friction materials, well end, flanges, and any other specified component material. Then the inspector witnesses the materials identification on the certificates against the materials marking which is further verified with the valve drawing datasheet, material list, and other specifications.

Upon completion of the valve material inspection by the third part inspector, he issues an inspection visit report (IVR), that contains the following items:

  • Confirmation of satisfactory document review
  • Record of the endorsement of certification reviewed/witnessed
  • Record of all non-conformities
  • Record of any tests witnessed and the result

Use of Sway Brace, Strut, and Snubber for Pipe Supporting

Written by Anup Kumar Dey in Piping Stress Basics,Piping Supports

Use of Dynamic Restraints

Whenever unplanned dynamic events occur, Dynamic restraints carry the responsibility of protecting the piping and other components from damage. Undesirable abrupt movement of the components in the system can be caused by:

  •       Pressure shocks from valve operation/ PSV
  •       Water hammer
  •       Boiler events
  •       Pipe breakage
  •       Wind load
  • Mechanical vibrations are transmitted from pumps, compressors, turbines, or other process equipment.
  •       Seismic events
  •       Fluid disturbances
  •       Explosions etc.

Dynamic restraints are specially designed to absorb a sudden increase in load from the pipe and transfer into the structure and to dampen any opposing oscillation between the pipe and the structure. These restraints are not intended to carry the weight of pipework and should not impede the function of the supports. Dynamic restraints are required to be very stiff, to have high load capacity, and to minimize free movement between pipe and structure.

Types of Dynamic Restraints

The main supports that make up the dynamic restraints for process piping are-

  1. Sway Braces
  2. Hydraulic and Mechanical Snubbers
  3. Rigid Struts
  4. Pipe Clamps
  5. Welding Clevis etc.

In the following paragraphs, we will discuss in brief about Sway Braces, Rigid Struts, and Snubbers. 

Typical Dynamic restraints
Fig. 0: Typical Dynamic restraints

What is a Sway Brace?

Sway braces can be defined as spring-loaded units mounted on pipework that is used to limit the swaying or vibration induced by external forces (vibration force) by applying an opposing force on the pipe. They are double-acting variable spring units that can handle both tensile and compressive loads. It is commonly used to allow unrestrained thermal movements while “tuning” the system dynamically to eliminate piping vibration. It could be pre-loaded in the cold or installed position so that after thermal pipe movement (growth) it reaches the neutral position and the load on the system in the operating (OPE) condition is negligible (almost zero). 

Working of Sway Brace

The construction is fairly simple, the unit has two piston plates: one on either side of the helical coil compression spring connected by a single-piston rod.

sway brace
Fig. 1: Schematic Representation of Sway Brace construction.

If a tensile load is applied, the top piston plate is pulled down causing the spring to compress & if a compressive load is applied the thrust nut/rod coupling pushes the bottom piston plate push up which causes the spring to compress. Therefore in both of the situations, the spring gets compressed but due to design (see the cutaway section above) the unit is capable of handling both compressive & tensile movements /forces. 

The spring is pre-compressed (usually a full inch =25 mm) providing an initial force (preload) that instantaneously opposes vibration. Whenever any movement from the sway brace neutral position occurs it is opposed by a load equal to the pre-load plus travel from the neutral position times the sway brace spring constant. To explain it further, if the piping load on the sway brace is less than preload then there will not be any line movement. If the load is equal to preload then the line will be on the verge of movement, but then also the line will not move. If the load is more than the preload the line will deflect causing the spring to compress further. The deflection of the spring/pipe in this case will be as given in equation 1. 

Pipe deflection= (piping load – Preload)/spring rate   Equation 1 

So there is no pipe movement if the load is less than the preload and with a load in excess of preload the deflection is as given in Eqn 1.  

When a sway brace with a preload P is installed in a pipe there is no force exerted by the sway brace on the pipe. But for the pipe to have any movement in either direction along the line of sway brace installation it will experience a reactive force equal to P plus travel from neutral position times the sway brace spring constant. It is desired to have no force on the pipe during normal operation of the pipe. So sway braces are normally attached during normal operation or adjusted to the neutral position during the normal operating condition. 

When the maximum allowed travel (usually 3-in. / 75 mm in either direction) is reached the sway brace locks prevent additional movement and act as a rigid restraint.

The preload for LISEGA sway braces can be adjusted as per requirement at the site. But for C&P or others, the unit is shipped after adjusting the required preload. 

The effect of the sway brace on the piping system is to increase the K value in the equation                                       

Mx2(t) +Cx(t) +Kx(t)=F(t)

This in turn will raise the natural frequencies of the vibratory modes & thus normally reduce the response of the pipe to dynamic loads & vibrations.

The force required to restrain the pipework can be calculated as follows:

If the pipework is vibrating with frequency f Hz at a maximum displacement (half amplitude) of x mm then, in simple harmonic motion, the restoring force exerted by the pipework at maximum displacement (kgf) = 4π2 f2 m x/1000 g. Where m is the equivalent mass of the pipework in kg. 

It is likely that a Sway Brace having a preload greater than this value willfully restrains the pipe at the support location, while a Sway Brace for which this value is greater than the preload but less than the maximum load will have a significant effect.  

Manufacturers normally recommend a specific size of sway brace for a pipe’s nominal diameter. If the exact restraining force required to control the piping vibration is known beforehand then a more specific sway brace selection is possible. The energy necessary to control the piping system is proportional to the mass, amplitude of movement, and the external force which is causing the vibration. From this relation, the exact restraining force required to control the piping vibration can be calculated and an appropriate sway brace size can be selected. 

Sway braces need to be installed in operating condition. However, it can be installed in cold conditions. But in that case, when the plant starts operating, the pipe may have thermal movements. This may cause the spring in the sway brace to compress by an amount equal to the thermal movement/displacement. At this point, the sway brace will be exerting a force equal to the pre-load + movement X spring constant. The load needs to be released by doing a “Neutral adjustment”. This can be achieved by rotating the Rod coupling shown above in a direction such that the piston plate gets released & rests against the endplate. In this situation, the sway brace will not exert any force on the pipe. During the shutdown, as the pipe cools & gets into the cold position, the sway brace will exert a force on the pipe as the spring will get compressed. To summarize,

Sustained loads on sway brace = Pre-Load + Hot Deflection * Spring Rate

In OPE case the displacement allows thermal expansion and the sway assumes a neutral position exerting zero or negligible load on the pipe. i. e, Operating case restraint loads on sway brace =~ 0.0 (does not restrain thermal expansion)

Major Application of Sway Braces

Sway Braces are mainly used to reduce pipe vibration amplitude and at the same time do not increase the expansion stress in operating cases. It prevents the pipe from vibrating at its resonant frequency. Typical examples of using the sway braces are in the pipeline feeding the flare stack in a refinery. When gases at very high pressures are passed in the pipeline in the flare stack, it tends to vibrate & the sway brace will try & limit the vibrations. Every time the vibrating force has to act as opposed to the sway brace preload+ the stiffness multiplied by the distance moved from a neutral position. When the line movement exceeds the sway brace becomes rigid and acts as a rigid guide in that direction.

The spring stiffness and preload are fixed depending on pipe size. However, for special applications manufacturer can change those values as per requirement.

Click here to visit Caesar Modelling Procedure for Sway Brace

Strut or Rigid Strut: Definition, Specification, and Selection

When we need to limit the displacement which does not affect an increase of thermal stress in operating conditions or when the disturbed displacement is at an axis normal to the thermal displacement it is preferable and less expensive to use a rigid strut or strut.

Rigid struts are selected to suit the force that they will resist and the space available to fit them. The anchor point to the structure is the most simple to select since it is only dependent on the size of the rigid strut. The pipe attachment is dependent on both pipe size and strut size but it is also influenced by the orientation of the strut relative to the pipe arrangement.  

The strut is often more difficult to specify because it may be resisting forces in the three primary axes, x, y and z. It is therefore necessary to use some simple trigonometry to resolve the given forces into axial force acting on the strut and to calculate the actual length of the strut between the fixing point and the pipe attachment. Because the strut is held between two pinned connections its ability to resist compressive force is greater the shorter the strut is. A long strut will have a lower safe working load in compression than a short strut. However, its length does not affect the tensile load capacity of the strut.

 The strut is therefore selected by considering the direction and magnitude of the axial force and if compressive forces are acting, the length between the fixing pins of the connections.  After the strut size is selected, the welding clevis will automatically specify to suit the strut size. The pipe attachment is selected now by considering the pipe size, the strut size, and the connection requirements between the strut and the clamp. It is essential that the strut can attach to the clamp without obstruction and that any thermal movements are able to occur without the strut interfering with the clamp. Therefore it is very important to consider the transition of the assembly during all expected displacements.

 Major Applications of Rigid Strut

Rigid Struts are used in Turbine and Compressor connected lines near the nozzle connections to take advantage of very little friction. Otherwise, struts can be used as a substitute for guide supports where the structure is not available for using standard guides.

Click here to check the modeling procedure of rigid Struts in Caesar II

Use of Snubbers as Dynamic Restraints

 The use of snubbers (Also called shock absorbers) is preferred in thermally operating piping systems. In a dynamic event, snubbers instantaneously form a practically rigid restraint between the protected component and the structure. Resulting dynamic energy can at once be absorbed and harmlessly transferred while the operational displacements due to thermal expansion and contraction must not encounter any noticeable resistance. Through the special function of the shock absorbers, thermal displacements during normal operation remain unhindered (offers very little resistance to pipe movement). When however a sudden impact load acts upon the snubber internal braking device engage, thus controlling the movement of the pipe. Snubber is said to be “lock-up” and in this condition, the snubber acts as a rigid restraint. When the load has dissipated, the snubber unlocks and again allows gradual movement of the pipe.

Types of Snubber

Depending on the internal mechanism of working snubbers are of two types:

  1. Hydraulic Snubbers and
  2. Mechanical Snubbers

Working of Hydraulic Snubbers

Similar to an automobile shock arrestor the hydraulic snubber is built around a cylinder containing hydraulic fluid with a piston (See Fig. 2) that displaces the fluid from one end of the cylinder to the other.

Displacement of fluid results from the movement of the pipe causing the piston to displace within the cylinder resulting in high pressure on one end of the cylinder and relatively low pressure on the other.

The velocity of the piston will dictate the actual difference in pressure.

The fluid passes through a spring-loaded valve, the spring is used to hold the valve open. If the differential pressure across the valve exceeds the effective pressure exerted by the spring, the valve will close. This causes the snubber to become almost rigid and further movement or displacement is substantially prevented.

The hydraulic snubber is generally used when the axis of restraint is in the direction of the expansion/ contraction of the pipe. The snubber is therefore required to extend/ retract with the normal operation of the pipework. The snubber has low resistance to displacement/ movement at very low velocities. The resistance to normal thermal movements (pipe velocity less than 1 mm/Sec and with an amplitude of vibration less than 3 mm) is less than 2% of the rated load of the snubber.

hydraulic snubber
Fig. 2: Hydraulic Snubbers

Working of Mechanical Snubbers

Whilst having the same application as the hydraulic snubber, retardation of the pipe is due to centrifugal braking within the snubber. A split flywheel is rotated at high velocity which causes the steel balls to be forced radially outwards. The flywheel is forced apart by the steel balls causing braking plates to come together thus retarding the axial movement/displacement of the snubber. Rotation of the flywheel is generated by the linear displacement of the main rod acting on a ball screw or similar device. 

Mechanical snubbers (See Fig. 3) are used in cases of applications where human access is restricted, for instance, due to the high radiation atmosphere in the nuclear plant or due to a high elevation point where no scaffold is available & maintenance work is not easy to do. No maintenance service is required for mechanical snubbers & are designed to generate the required resistance force instantly on reaching threshold acceleration, to restrain a displacement of piping caused by an earthquake or other dynamic events & resume their free movement as soon as the dynamic displacement is suppressed while developing very little (a negligible level of) frictional resistance force during the slow thermal displacement mode of piping.

Selection of Snubbers

 The snubber is influenced by the same factors that the rigid strut is, the magnitude and direction of axial force, but it is also necessary to consider the thermal displacement the snubber has to undergo.

Again it is necessary to use trigonometry for calculating the force and the length of the snubber along with the actual displacement applied to the snubber. Displacements in the primary axes cannot be combined simply to determine the snubber movement/displacement; it is necessary to calculate the overall length of the snubber in the various installed and operating conditions in order to determine the needed stroke.

After calculating the actual stroke it is good engineering practice to take a margin of excess travel at each end of the design travel. So, Always select a snubber that is capable of allowing greater displacement than is theoretically required.

Orientation of the snubber is also important for both hydraulic and mechanical types. Access to either lubrication points or inspection points is normally required and must be considered during the design and installation of the restraint. It may also be required to allow in-situ testing of the snubber for validating its functionality and so access may be a permanent requirement.

mechanical snubber
Fig. 3: Mechanical Snubbers

For selecting the proper Snubber, determine the minimum required stroke by taking the anticipated design movement and adding an allowance for excess travel. This allowance should normally be at least 20% of the anticipated design movement. Then select a snubber where the cylinder stroke is greater than or equal to the minimum required stroke and the applied loadings in tension and compression are less than the allowable maximum loadings in tension and compression for the size and length of snubber as shown in the catalog. For intermediate lengths, allowable compressive loadings may be determined by interpolation. The length of the snubber must be such that the maximum angulations are not exceeded.

 To calculate the required closed centers for the snubber, use the following formulae:

Closed Centres = Installed Centres — X

Where X = (Stroke – Design Movement in Extension) / 2     or

X = (Stroke + Design Movement in Compression) / 2

This method will result in spare travel being distributed evenly on either side of the design movement. 

Main Application of Snubber

Snubbers are normally used for reducing the damaging effects of Earthquake events.

Click here to check the modeling of snubber in Caesar II

Online Course on Pipe Support Engineering

If you want to learn more details about pipe support engineering then the following online course is a must for you:

Smart Tee Model considerations in START-PROF

Written by Alex Matveev in Start-Prof

What is the Tee branch connection?

A tee branch connection is a type of piping fitting that allows for the branching of a pipeline into two or more directions, typically at a 90-degree angle. This configuration resembles the letter “T,” hence the name. Tee branches are essential in various fluid transport systems, including water, gas, and oil pipelines, as they enable the distribution of fluids to multiple locations from a single source. The design ensures that flow can be redirected efficiently while maintaining pressure and minimizing turbulence, which is crucial for the system’s overall performance.

Importance of Tee Branch Connection

The importance of tee branch connections lies in their ability to facilitate complex piping layouts without compromising flow integrity. They are commonly used in industrial, commercial, and residential applications, allowing for flexibility in system design and maintenance. By enabling multiple connections from a single pipeline, tee branches optimize resource utilization, reduce material costs, and streamline installation processes. Properly installed tee connections also contribute to system reliability, helping to prevent leaks and other operational issues that could arise from improper flow management.

This article provides guidelines for modeling tee or branch connections using software START-PROF. The guideline covers modeling for both the Standard Tee and Non-Standard Tee Connection.

Modeling Standard Tee

When you create a standard tee in START-PROF, you can specify tee length (L), tee height (H), header wall thickness, and branch wall thickness.

Fabricated Tee Modeling in Start-Prof
Fabricated Tee Modeling in Start-Prof

When you run analysis simple tee model is automatically replaced by a complex model. Added 6 additional nodes (21, 2, 3, 4, 5, 20). Nodes are hidden for the user. It is visible only in developer mode.

  • 10-20 Element is rigid (H=Dh/2, Dh – header outer diameter)
  • 4-10 and 10-5 elements length is 1 mm (L=2mm)
  • 20-21 pipe element has a wall thickness specified in “Tee” dialog box Tb=10 mm. It can be thicker than the connected pipe
  • 4-3 & 5-2 pipe elements have a wall thickness specified in the “Tee” dialog box tn=10 mm. It can be thicker than the connected pipes
Smart Tee
Tee Connection in Start-Prof

Modeling Non-Standard Tee

For “Nonstandard Tee” and for any standard tee if ASME B31J code is selected, additional run flexibilities are added into nodes 4 & 5, and branch flexibilities are added into node 20. Flexibilities calculated according to ASME B31J

Using ASME B31J Option in Start-Prof
Using ASME B31J Option in Start-Prof

If the option “Consider Tee Branch Flexibility” is selected, then 4 & 5 nodes are not created for standard tees. The flexibilities in node 20 calculated according to ASME BPV SIII div 1 class 1 NB 3686.

Considering Tee Branch Flexibility
Considering Tee Branch Flexibility

Few More Resources for you…

Piping Stress Analysis using Start-Prof
Piping Stress analysis Basics
Piping Stress Analysis using Caesar II
Piping Stress Analysis