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Understanding Pressure and Temperature in the context of Pressure Vessel Design

Written by Anup Kumar Dey in Mechanical,Process

Many of the failures that we experience in the process industry are due to overpressure. Overpressure is the result of unbalance or disruption of normal flows of material and energy. This causes material or energy, or both, to build up in some part of the system. Overheating above the design temperature may also result in overpressure, due to the reduction in allowable stress.

What is the meaning of Pressure in a Pressure Vessel

Let us first understand what pressure is and what is its impact on pressure vessels.

The pressure is force per unit area (Pressure = Force/ Area)

It’s like many small weights sitting on a surface & many small weights add up to a big weight.

So at a given pressure, the LARGER the area, The LARGER the force

Can you imagine 1 barg pressure acting on a 20” manway exerts a 20270 Newton force? This is equivalent to lifting a 2 Tons weight.

Calculation: (100000*PI*(0.508)^2)/(4*9.8) = 20270N

Pressure Vessels (Fig. 1)

We cannot imagine the oil and gas industry without pressure vessels and these include separators, Knock Out drums, columns, steam drums, etc.

The thermodynamic energy contained in a pressure vessel can be very large.

The energy stored in this pressure vessel operating at 50 barg is 0.5*10^10 NM

This is equivalent to the kinetic energy of 13000 cars of 1000 kg each driving at a speed of 100km/hr

Pressure Vessel and Impact of Pressure
Fig. 1: Pressure Vessel and Impact of Pressure

Effects of Pressure:

You might have seen or participated in Tug of war or rope pulling where two teams pull on opposite ends of a rope.

In a similar way if we apply tension or stress to a steel specimen by pulling it to different levels the specimen under stress starts to deform and elongate.

The strain is a measurement that gives the change in length of an object divided by the original length

With increasing stress, if we measure corresponding strains, then we can plot a stress-strain curve for that specimen as shown on the right.

Deformation in the elastic region is reversible i.e. Once the forces are no longer applied, the object returns to its original shape.

Plastic deformation is not reversible.

Yield stress is the minimum stress that causes a material to lose its elastic behavior and deforms permanently.

Ultimate Tensile stress is the maximum stress that a material can handle before breaking.

Pressure inside also subjects equipment to stress. If a vessel designed for 3.5 barg pressure is subjected to higher pressures, it may result in catastrophic failure.

Pressure Vessel Codes

Robust Process and Mechanical Design are the fundamental and first barriers to ensuring safe operations.

At the start of the 20th Century, there were numerous incidents related to pressure vessels and manufacturers started to exchange their know-how and experiences. This finally resulted in the nationalized codes on pressurized equipment. Applying a pressure vessel code provides the equipment with sufficient margins against failure under specified temperatures and pressures.

The pressure vessel code allows pressure vessels to be designed, operated, and manufactured along with rules set up by the industry and is widely used by manufacturers and operators.

The main objectives of the pressure vessel codes

  • Ensure that the vessel will be safe under all foreseeable circumstances.
  • Define minimum protection requirements
  • Clearly define the responsibilities of various parties in design and construction
  • Give requirements for manufacturing and quality control of equipment
  • Clearly define operating windows for safe operations

An important concept of vessel design is to yield stress. if stresses higher than the yield stress (which is temperature-dependent) of a given material are applied then elastic or eventually plastic deformation may occur.

As per Pressure Vessel Code ASME Section VIII Div. 1

Design stress of a pressure vessel = lowest of:

  • Ultimate tensile stress / 3.5
  • Yield stress / 1.5
  • Creep rupture stress / 1.5

Designing for High Pressure

Ideally, all process equipment would be designed to withstand the maximum pressure which can be attained in service during normal operating, upset, start-up, and shutdown conditions. However, in some cases it may not be economically feasible to do this.

Equipment/vessels are usually designed for the pressure which is calculated by adding a margin to maximum operating pressure and full protection is achieved by a relief valve set at or below specified design pressure. The accepted margin is adequate for normal control of the process but not to cope with deviations when control is lost or inadequate. The high-pressure alarm and a high-pressure trip are provided within this margin to avoid unintentional relief valve opening. This keeps the process within the set boundaries “The design envelops.”

Pressure and stress
Fig. 2: Pressure and stress

For example, centrifugal pumps are often not provided with a relief valve as they are designed for the highest possible pressures corresponding to the blocked outlet condition whereas the associated piping and vessels are protected by a relief valve.

Pressure Terminology

It is important to understand the different terms used in the context of pressure:

Operating pressure (OP): The OP is the gauge pressure that prevails inside equipment and piping during any intended operation. The OP is determined by the process engineer.

Maximum operating pressure (MOP): The MOP is typically

  • MOP = OP + 1 (If OP below 20 barg)
  • MOP = 105% of OP (If OP above 20 barg)

Design pressure (DP): The DP is the gauge pressure at the top of the equipment in its operating position. It is used to determine the minimum thickness of equipment parts at the Design Temperature.

The DP is initially selected by the process engineer and finally determined in close consultation with the mechanical design engineer.

The following three scenarios are commonly used for determining DP

  1. For non-liquid full systems with a vapor relief to the atmosphere, DP is normally determined from the MOP by the following rules:
  • DP = MOP + 1 (Below 10 barg)
  • DP = 110% of MOP (Above 10 barg)
  1. Vapor relief to flare system: Equipment that is part of a pressure system protected by a relief valve discharging into a flare system or combined vent system shall have a minimum design pressure of at least 3.5 bar (g).

Thus, in this case, DP = Max of : MOP+1; 110% of  MOP; at least always 3.5 barg. This is because a minimum of 3.5 barg pressure is required to transfer the relieving pressure to the flare tip (against all the back pressure present).

  1. For liquid full systems, DP=Maximum shut-off pressure of the pump.

Since the DP is related to the top of the equipment, for other parts or elements of the equipment the designer shall establish the associated design pressures taking into account the maximum pressure drops caused by flow through the equipment, plus the fluid static head.

Maximum allowable working pressure (MAWP):

The MAWP is used in the ASME Boiler and Pressure Vessel Code and a number of other codes referring to it. The MAWP is the maximum gauge pressure permissible at the top of the equipment with the equipment installed in its operating position and at a designated temperature. For the existing equipment, the MAWP is calculated based on the type of material its wall thickness, and its service conditions.

MAWP ≥ DP

Lower Design Pressure

The Lower Design Pressure is the external design pressure or the sub-atmospheric pressure at the top of the equipment in its operating position. It is used to determine the minimum thickness of equipment parts or stiffening rings at the design temperature.

In vacuum systems, the pressure is pushed inward and comes from the atmosphere. Some processes which require vacuum design conditions

  • When cooldown is expected,
  • Components with a boiling point below 0 °C,
  • Steam side of heat exchangers or vessels with steam.
  • Pumping out in absence of adequate replacement with vapor

Low-pressure storage tanks and railcars are particularly susceptible to damage. If tanks or vessels are not designed for vacuum, it is likely they will be damaged if placed under vacuum.

With the exception of storage tanks that are protected by pressure-vacuum vents, process vessels that can be subjected to a vacuum shall be rated for that vacuum. Vessels that are in steam service should be designed for full vacuum at 150 °C

Here are a few examples of things going wrong:

A tank was being painted and the painters had covered the vent with plastic sheeting. When operations started to empty the tank, it collapsed before the plastic sucked through.

The tanker was being steam cleaned and, at the end of the job, the hatches were closed. With no vacuum breaker fitted, as the steam condensed, the tanker imploded.

Guidelines for Design Temperature

After pressure, it is important to understand the different terms in the context of temperature

Operating Temperature (OT): The OT is the temperature that prevails inside equipment and piping during the pre-dominant intended operation (in0C).

Maximum Operating Temperature (MOT): If operational flexibility is needed the MOT is established higher than the OT, otherwise they are equal.

  • MOT ≥ OT

Sometimes the MOT equals the maximum equilibrium temperature of the composition in the vessel at the Maximum Operating Pressure. Also many times multiple OTs are specified.

Mechanical Design Temperature (MDT) and Upper Design temperature (UDT): The Mechanical Design Temperature (MDT) is often referred to as the upper design temperature (UDT). The UDT is the highest temperature to which equipment may be subjected at the upper and/or lower design pressure. The MDT is typically 10 °C above the MOT

  • DT = MOT + 100C

The design pressure and temperature form the basis for mechanical design for equipment and piping and are used in conjunction (coincident design conditions) for calculating minimum wall thickness for vessel and piping design

Lower Design Temperature (LDT): The LDT is the lowest temperature to which equipment may be safely subjected at its design pressure with respect to brittle fracture control. DEP 30.10.02.31-Gen contains details for designing to safeguard against brittle fractures

Brittle Fracture

Brittle Fracture is a condition that occurs when a material is subjected to temperatures that make it less resilient, and therefore more brittle.

The potential for material to become brittle depends on the type of material that is subjected to these low temperatures. Some materials, such as carbon and low alloy steels will become brittle at low temperatures and therefore susceptible to damage ranging from cracking to shattering or disintegration of equipment.

When a material becomes brittle, the consequences can be very serious. If the brittle material is subjected to an impact or an equivalent shock (ex. rapid pressurization) the combination could potentially lead to catastrophic failure under certain conditions.

A serious brittle fracture incident occurred at the ESSO Longford gas plant in Australia in Sept 1998 when a heat exchanger failed catastrophically due to exposure to low temperatures. The released hydrocarbons caused a massive explosion. Two employees were killed, eight others were injured and supplies of natural gas to domestic and industrial users were halted. So, Any possibility of re-pressurization, or pressurization from connected systems, whilst the equipment is colder than the LDT shall be prevented to avoid brittle fracture failures.

Few more Resources for you..

A short Presentation on Basics of Pressure Vessels
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

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:

Stress Analysis of PSV/PRV Piping Systems Using Caesar II | PSV Piping Stress Analysis

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

PSV or pressure safety valves (pressure relief valves) are a type of valve and are very common in any process industry. To protect any equipment from overpressure PSV systems are used in lines. When the pressure inside the system/equipment exceeds a pre-determined level (normally Set Pressure), they are activated automatically and release the pressure by popping up and bringing the equipment pressure to a safe operating level.

Two types of PSVs are extensively used in process industries:

  • Open discharge PSV
  • Closed discharge PSV

Are PSV Connected Systems Critical?

Due to an uncertain event if the pressure of any equipment becomes higher than the set pressure of the installed PSVs then they pop up and reduce the system pressure. During popping-up activity, the PSVs exert a huge reaction force over the system. During the analysis of PSV-connected stress systems, we have to consider this reaction force. This is the main reason that PSV-connected systems become stress-critical. The following write-up will try to explain the methods used during the analysis of such systems using Caesar II. Click here to gain in-depth knowledge about pressure relief valves.

Required Documents for Stress Analysis

The following documents are required for PRV system stress analysis using Caesar II.

  • Piping isometrics.
  • P&ID and line list.
  • PSV datasheet with reaction force and PSV weights.
  • Equipment GA and datasheet if the equipment is part of the stress system.

PSV Reaction Force Calculation & Application Philosophy

Before we start the actual analysis we should first know the reaction force. Normal practice is to obtain the reaction force from the PSV vendor or manufacturer. However if during the preliminary stage of analysis, data is not available then the reaction force for open discharge PSVs can be calculated using the below-mentioned formula (from API RP 520) for gaseous/vapor services. But later it must be corrected for forces received from the vendor.  

Open Discharge PSV
Fig. 1: Typical Open Discharge PSV Connection

PSV Reaction force at the point of discharge for Gas Services in lbf, F=[(W/366)* √{K*T/(K+1)*M}]+A*P

  • Here, W=flow of any gas or vapor in lbm/hr      
  • K=ratio of specific heats (Cp/Cv) at the outlet condition  
  • T=temperature at the outlet in Degree R
  • M=molecular weight of the process fluid          
  • A=area of the outlet at the point of discharge in inch^2          
  • P=Static pressure within the outlet at the point of discharge in psig.          
  • Cp and Cv=Specific heat at the constant pressure and at constant volume respectively.    

For liquid services, the PSV reaction force (FR) due to the outflow of PSV (or PRV) can be calculated following the AD 2000-Merkblatt standard A2-Safety Devices against Excess Pressure (Clause 6.3.3) using the following equations (momentum theory):

PSV/PRV Reaction Force for Liquid Services in N, FR=(qm*vn/3600)

Where,

  • qm=Mass flow in Kg/hr
  • vn=velocity in the blowout opening=(qm*106)/(3600*ϱn*An)
  • ϱn=density of the fluid in the blow-out opening at the end of the pipe in kg/m3
  • An= Clear cross-sectional area at blow out end of the line in mm2

For closed-discharge PSV systems, there is no specific method to calculate the reaction force. Complex time history analysis can be used to calculate the reaction force for closed-discharge PSV systems.

Parameters Affecting Pressure Relief Valve Reaction Force

From the above equations, it is quite clear that the main parameters that affect the PSV/PRV reaction forces are:

  • Mass Flow Rate: With an increase in mass flow rate the PSV reaction force increases.
  • Outlet Temperature: With an increase in the PRV outlet temperature, the reaction force of the pressure relief system increases for the gaseous medium. However, for liquid medium, the PSV reaction force is not dependent on temperature.
  • PSV Size: With an increase in the PSV size, the PSV reaction forces increase.
  • The ratio of Specific Heats(Cp/Cv): For gas/vapor systems the increase in specific heat ratio, increases the pressure safety valve reaction force value.
  • Pressure
  • Fluid Density
  • Velocity, etc

Applying PSV Reaction Force

The reaction force application philosophy for open discharge PSV (PSV output discharges into the atmosphere) connected systems is the same throughout the process industries. But for closed discharge, PSV connected system the force application philosophy varies from organization to organization. Some organization applies the reaction force for closed discharge PSVs but some organizations do not consider it. So users need to follow the company-specific project guidelines in such cases.

Where to Apply the PSV Reaction Force?

The following figure (Fig. 2) shows the points where the reaction force is required to be applied for open discharge PSVs.

open discharge PSV
Fig.2: Reaction force Application point for open discharge PSV connected systems

Fig. 3 shows the application point (If required) of reaction forces for closed discharge PSV connected systems.  

closed discharge psv
  Fig.3: Reaction force Application point for Closed discharge PSV connected systems

Caesar II Load Cases for PSV Connected Systems:

PSV forces are considered occasional forces. So occasional Stress due to PSV reaction force has to be calculated and to be limited within 1.33 times Sh (As per code ASME B31.3). Here Sh=Basic allowable stress at hot conditions. Based on company practice PSV reaction force is added either with a design temperature case or with an operating temperature case. Also, some organizations have the practice of making One PSV pop up and others stand by load cases. Accordingly, make the load cases as shown below:  

reaction force
  Fig.4: Caesar II methodology to enter the reaction force
load cases
    Fig.5: Caesar II simple load cases for analysis PSV connected system

The following load cases assume two temperatures (T1= operating temperature, T2= design temperature) along with Wind and Seismic load cases:  

Load CaseStress TypeDescription
L1 WW+HPHYD Hydrostatic Case
L2                W+T1+P1                           OPE Operating temperature case
L3W+T2+P1                            OPE Design temperature case
L4 W+T1+P1+F1                     OPE Operating temp+PSV reaction ( PSV 1 popping up)
L5 W+T1+P1+F2                     OPE Operating temp+PSV reaction ( PSV 2 popping up)
L6 W+T1+P1+WIN1 OPE Operating temp+Wind from North
L7 W+T1+P1+WIN2 OPE Operating temp+Wind from South
L8W+T1+P1+WIN3OPEOperating temp+Wind from East
L9 W+T1+P1+WIN4OPEOperating temp+Wind from West
L10W+T1+P1+U1OPEOperating temp+Seismic from North
L11W+T1+P1-U1OPEOperating temp+Seismic from South
L12W+T1+P1+U2OPEOperating temp+Seismic from East
L13W+T1+P1-U2OPEOperating temp+Seismic from West
L14W+P1SUSSustained case
L15L4-L2OCCPure PSV Reaction
L16L5-L2OCCPure PSV Reaction
L17L6-L2OCCPure Wind
L18L7-L2OCCPure Wind
L19L8-L2OCCPure Wind
L20L9-L2OCCPure Wind
L21L10-L2OCCPure Seismic
L22L11-L2OCCPure Seismic
L23L12-L2OCCPure Seismic
L24L13-L2OCCPure Seismic
L25L15+L14 OCC Pure Occasional+Sustained
L26L16+L14 OCC Pure Occasional+Sustained
L27L17+L14 OCC Pure Occasional+Sustained
L28L18+L14 OCC Pure Occasional+Sustained
L29L19+L14 OCC Pure Occasional+Sustained
L30L20+L14 OCC Pure Occasional+Sustained
L31L21+L14 OCC Pure Occasional+Sustained
L32L22+L14 OCC Pure Occasional+Sustained
L33L23+L14OCCPure Occasional+Sustained
L34L24+L14OCCPure Occasional+Sustained
L35L3-L14EXPPure Expansion
L36L2-L14EXPPure Expansion
Table 1: PSV Piping System Stress Analysis Load Cases

Output Study:

  • Check Code stresses for load cases L1, L14, and L25 to L36. It is better to keep stresses for L1 and L14 below 60% and for the rest within 80%.
  • Check forces for load cases from L1 to L14.

Better Engineering Practices

  • It is a better practice to use 3-way restraints in both inlet and outlet piping of PSV-connected systems if feasible (As shown in figures 2 and 3 above).  However if not possible then try to provide a 3-way restraint in the outlet only by layout modification.
  • In a normal operating case Safety valve inlet line temperature will be operating temperature up to the inlet of the safety valve and the Safety valve outlet line will be in ambient temperature up to the header.
  • Sometimes a Dynamic Load Factor (DLF) of  2 is used for calculating PSV reaction force.
  • If any stress failure or abnormal routing changes are required, then a certain local area from the header can be used at an average temperature of 2 meters or 5D which is higher (Safety valve outlet joining at header junction point) and also shall be taken process engineer’s approval.
  • If required stress engineer shall provide an R.F. pad for the trunnion-type support.
  • If the connection of the PSV closed system is emerging from the header with 45˚ put SIF for this tapping. If required tapping point of the outlet line and outlet header shall be reinforced to reduce SIF.
  • In case any safety valve assembly is placed on the top platform of any vessel, Support can be taken either from the top platform or support can be arranged from the top portion of the vessel taking a clip from the vessel. In both cases, the load and locations of support or clip equipment vendors must be informed through the mechanical group along with the clip information.
  • Do not provide a spring below the safety valve inlet line

Few more references for you

Modelling Relief Valve (Pressure Safety Valve) Thrust force
Routing Of Flare And Relief Valve Piping: An article
Various types of pressure-relieving devices required for individual protection of pressure vessels in process plants

References:

Online Course on PSV Piping Stress Analysis

If you still have doubts, then you can enroll in the following video course that explains the stress analysis of the PSV piping system using Caesar II software.

Introduction to Fire Protection System | Fire Fighting System

Written by Anup Kumar Dey in Health & Safety,Piping Design Basics,Piping Interface

Importance of Fire Protection System

Fire Protection systems are a very important part of safety in any operating plant as it provides a reasonable degree of protection to expensive equipment, property, documents, life, and inventory during a fire event. For Oil & Gas, Refinery, or plants those deals with Petroleum or similar flammable products, It must have to be in place to avoid major loss during uneven circumstances.

A Fire requires combustible materials, oxygen, and an energy source (heat) to provide ignition. Three components – fuel, oxygen & heat are referred to as the fire triangle.

Principle of Fire Extinction:

  • Starvation – Removing or blanketing the fuel
  • Smothering – Cutting off or diluting the oxygen supply
  • Cooling –   Removing heat from the fire.
Principles of Fire Extinction
Fig. 1: Principles of Fire Extinction

Types of Fire & Extinguishing Medium:

Types of fire and extinguishing medium
Fig. 2: Types of fire and extinguishing medium

Fire Fighting Agents:

Fire Fighting Agents are those substances that are used to reduce the effect of fire during fire events. Examples are:

  • Sand
  • Blanketing
  • Water
  • Steam
  • Carbon dioxide
  • Dry Chemical Powder
  • Aqueous Film Forming Foam (AFFF)

Properties of Petroleum Products:

  • For all flammable liquids, it is the vapor that burns and not the liquid.
  • Petroleum vapor is heavier than air so it has a tendency to descend on the ground or lower level or sump.
  • Petroleum is immiscible with water. Its specific gravity is less than 1, so they float on water
  • The electric conductivity of almost all petroleum products (except crude oil, ethanol, etc.) is very low and hence it generates static electricity during storage and transportation

Petroleum Products are divided into the following classes

  • Class A – Flashpoint <   23 deg.C
  • Class B – Flashpoint >  23 deg.C & <  65 deg.C
  • Class C – Flash point > 65 deg.C & < 93 deg.C
  • Unclassified – Flash point> 93 deg.C & above

Fire Protection Facilities in Petroleum Installations:

Fire Protection Facilities
Fig. 3: Fire Protection Facilities

Codes and Standards for Fire Protection System Design

The following reference codes and Standards govern the design of Fire Protection System Design.

  • NFPA 24 – Standard for the Installation of Private Fire Service Mains and Their Appurtenances
  • NFPA 13 – Standard for the Installation of Sprinkler Systems
  • NFPA 15 – Water Spray Fixed System
  • NFPA 11 – Standard for Low, Medium, and High expansion foam
  • NFPA 16 – Standard for the Installation of Foam water sprinkler and Foam water Spray system
  • NFPA 20 – Standard for the Installation of Stationary Pumps for Fire Protection
  • NFPA 22 – Standard for Water Tanks for Private Fire Protection
  • NFPA 30 – Flammable and Combustible Liquids Code
  • IP 19 – Fire Precautions at Petroleum Refineries and Bulk Storage installations
  • DEP 80.47.10.31-Gen – Active fire protection systems and equipment for onshore facilities

Fire Protection System Design

  • The firefighting system should be designed based on the Single Fire Scenario.
  • The Facility should be divided into zones
  • The type of Fire Fighting system should be decided.
  • Fire-Water application rate and discharge time should be referred from IP-19 or NFPA standards.
  • Firewater demand for the facility should be calculated.
  • Similarly, the water required for the Foam system should be calculated.
  • The facility with the highest water demand is considered critical and based on this the Firewater storage tank and pump capacity should be determined.

What are the different types of fire protection systems?

  • Fire-Water Tanks: Above-ground storage tanks of adequate nos. to meet the norm of 2 hr. continuous firefighting (As per IP-19).
  • Fire-Water Pumps
  • Hydrant Network
  • Water Spray System
  • Foam Pourer System

Fire-Water Pumps

Fire-Water Pumps (Fig. 4) should be selected based on the largest firewater demand for the facility. Fire-Water Pumps should be selected as per NFPA requirements

Firewater pumps should be three basic types as a minimum:

  • Electrical Driven Pumps – Primary Considered as cost-effective but not mandatory
  • Diesel Driven pumps – Secondary considered but mandatory
  • Jockey pumps – Required to keep the hydrant system pressurized and to check the health of the Fire Water system
Fire Water Pump
Fig. 4: A figure showing a Fire Water Pump

Hydrant Network in Fire Protection System

  • Hydrant Networks (Fig. 5) consist of Hydrants and Monitors.
  • Hydrants and Monitors should be placed at least 15 meters away from the hazard but not more than 45 meters.
  • Proper coverage should be checked for the hydrant and monitors ensure the facility under protection is adequately covered.
  • The hydrant mains can be laid above ground or underground.
  • The hydrant mains should form a closed loop ensuring multi-directional flow in the system. The layout should be such that the facility has access to Fire protection at any given time.
  • Isolation valves should be located at branches and other strategic locations.
hydrant network
Fig. 5: Figure showing a typical hydrant network

Water Spray System as Fire Protection System

  • The water spray system (Fig. 6) is provided for cooling the tank shell, and piping exposed to fire. The system is provided for cooling the structure on fire and exposure protection of adjacent property
  • The system consists of fixed piping with pipe fittings, isolation valves, NRV, and water spray nozzles.
  • In the case of tank and piping, the water spray directly impinges onto the surface of the tank or piping for cooling.
  • A spray ring should be installed between each tank wind girder.
  • NFPA 13 and NFPA 15 requirements should be met
  • Two types: Manual Water Spray system and Automatic Water Spray system
Water Spray System
Fig. 6: Typical Water Spray System

Use of Foam Pourer System (Fig. 7) in Fire Protection:

  • Semi-fixed foam Pourer system – Comprises fixed pipings and pipe fittings, drain valves, foam coupling, foam makers, foam pourer, and deflector plate on tank.
  • Mobile Foam tender is required for actuating the system.
  • Fixed foam pourer system( Manual/Automatic) – Manual system comprises of fixed foam concentrate storage shed, foam supply pumps, proportioning system, pipings, and pipe fittings, isolation valves, drain valves, foam coupling, foam makers, foam pourer and deflector plate on tank.
  • The automatic system requires motor-operated valves at different points and PLC for its actuation based on feedback from the automatic fire detection and alarm system.
Foam Pourer System
Fig. 7: Typical Foam Pourer System

Extinguishing System for Fire Protection

  • Fire Extinguisher – CO2, DCP, Clean Agent
  • Clean Agent System – FM 200, Halon, etc.
  • Sand Buckets

Some more References for you

Piping Design and Layout
Piping Stress Analysis
Piping Interface
Piping Materials
Piping Design Softwares

References:

Methods for Checking Flange Leakage | Flange Leakage Calculation | Flange Leakage Analysis

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

1. What is Flange Leakage?

Flange leakage is a serious problem in the piping industry. It has a tremendous potential to cause severe hazards to operating plants. Hence, the possibility of leakage needs to be investigated during the design stage to reduce leakage possibility during operation.

  • Basically, flange leakage is a function of the relative stiffnesses of the flange, gasket, and bolting.
  • Flanges are designed to remain leak-free under hydrostatic test pressure when cold and under operating pressure when hot.
  • The design of flanges (ASME B16.5) does not take into account the bending moment in the pipe. This generates a wire drawing effect on the mating surface of the flange. Hence, additional flexibility is to be provided when a flange joint is located near a point of high bending moment. So, Leakage checking is required.
Flanges
Flanges Used in Oil & Gas Processing Plants

Process Piping Flanges are designed in accordance with the ASME Boiler and Pressure Vessel Code, Section VIII, Division 1, Appendix 2, using allowable stress and temperature limits of ASME B31.3.

2. Reasons for Flange Leakage

Numerous possible reasons may cause a flange assembly to leak. Some of the common causes of flange leakage problems are:

  • Excessive Piping System Loads:
    • Forces and bending moments: can loosen bolts or distort flanges.
    • Causes include insufficient flexibility, excessive mechanical force, and poor support placement.
  • Incorrect Gasket Size or Material:
    • Wrong size: noticeable during installation.
    • Wrong material: This may cause issues like corrosion or blowouts later.
  • Vibration Levels:
    • Excessive vibration can loosen bolts, resulting in leaks.
  • Thermal Shock:
    • Rapid temperature changes can deform flanges and cause leaks, exacerbated by varying thermal expansions.
  • Improper Gasket Installation:
    • Off-center gasket: This leads to uneven compression and potential leaks.
  • Fastener Issues:
    • Too tight or not tight enough: uneven bolt stress due to improper tightening or loss of bolt tension over time.
    • Overly tight fasteners: excessive pressure on gaskets or fatigue on equipment, especially in high-temperature conditions.
  • Improper Flange Alignment:
    • Misalignment causes uneven gasket compression and potential leaks.
  • Flange Facing and Surface Finish:
    • Deeper serrations: Can prevent proper gasket seating, leading to leakage.
  • Damaged Flange Faces:
    • Corrosion pitting: Creates leakage paths.
    • Contaminants: Dirt, scale, scratches, protrusions, and weld spatter can lead to uneven gasket compression and leakage.

Most of the reasons mentioned above are construction-related, which can be eliminated by following proper best practices during construction activity. However, the major design cause that could result in flange leakage is excessive forces and moments that can be controlled during the design phase. The following section provides flange analysis methodologies to check the suitability of flanges against high forces, which may cause excessive stresses.

3. Flange Leakage Analysis Criteria

The criteria regarding when flange leakage checking is required should be mentioned in the ITB (Invitation To Bid) documents or project specs. But as a general practice, the following can be used:

  • Flanges with a rating of 600 or more
  • Flanges with a rating of 300 and size greater than 16 inch
  • Pipe flanges carrying category M fluid service
  • Pipe flanges carrying Hydrogen or other flammable fluid
  • PSV lines with NPS 4 inch or more
  • All Flanges in Jacketed Piping
  • Flanges where stress engineer finds a very high bending moment

This list is not exhaustive. Always refer to your stress or project guidelines for more details about flange leakage criteria.

4. Flange Analysis Methodology 

Four widely used flange analysis methods are practiced in the prevalent process or power piping industry. These are

  1. Pressure Equivalent method based on ASME B16.5 pressure-temperature rating table and
  2. ASME BPVC Sec VIII Div 1 Appendix 2 method.
  3. NC 3658.3 method
  4. Flange Leakage Analysis Using EN-1591
  5. Other Flange Analysis Methods

4.1 Flange Leakage Checking by Pressure Equivalent Method

In this method, the generated axial force (F) and bending moment (M) on the piping/pipeline flange are converted into equivalent pressure (Pe) using the following equations.

  • Equivalent Pressure for Axial force, Pe1=4F/ΠG2
  • Equivalent Pressure for bending moment, Pe2=16M/ΠG3
  • Here G=diameter at the location of gasket load reaction =(Gasket OD+ID)/2 when bo<=6 mm                                                                                =(Gasket OD-2b) when bo>6 mm. Here bo=basic gasket seating width as given in table 2-5.2  of ASME sec VIII 

These two equivalent pressures are then added with pipe or pipeline system design pressure (Pd) to find the total pressure (Pt=Pd+Pe1+Pe2) and checked against the rated pressure at the temperature for the flange. The rated pressure is found in the ASME B16.5 (ASME B16.47 for flanges with size 26 inches and more) pressure-temperature rating table associated with flange material. If Pt is less than the allowed pressure on the rating table corresponding to the associated temperature, then the flange will not leak.  

Hence the requirement for this method is Design Pressure (P):

Equation for Equivalent Pressure Flange Leakage Analysis

4.1.1 Drawbacks of the Pressure Equivalent Method

The major drawbacks of the pressure equivalent method are:

  • The flange leakage analysis by the pressure equivalent method is too conservative, resulting in significant changes to the piping layout and support and hence impacting cost and schedule. In some cases, the allowable bending moment may be as low as 5% of the pipe yield.
  • It does not address the issue of bolting and gaskets, two important factors for controlling leakage.
  • The pressure equivalent method of flange leakage checking does not compute stresses in gaskets, bolts, or flanges. On the flip side, its overly conservative approach keeps the stresses on these components, subject to the proper selection of material, installation, and fabrication, to low magnitudes.

The steps for the pressure equivalent method in Caesar II software are explained here…

4.2 Flange Leakage Checking by ASME BPVC Sec VIII Div 1 Appendix 2 Method

This method is the widest one used in the industry where rules of ASME SEC III NC3658.3 are not applicable (B16.47 Flanges). In this method, flange stresses (longitudinal hub stress, radial flange stress, and tangential flange stress) are calculated based on ASME code-provided equations and formulas. These calculated stresses are then compared with allowable stresses as given in ASME BPVC Code Sec VIII Div 1 Appendix 2, Clause 2-8.

The mathematical basis for the computation of the stresses is in the work of Rossheim and Waters, where the governing equations for a circular plate (for the flange) and for a cylindrical shell (hub/pipe) under the action of internal pressure were supplemented with deformation compatibility at the interface between the hub and the ring and hub and pipe.

It’s important to note that many B16.5 flanges fail to meet the ASME code requirements, especially when a recommended magnitude of installation bolt stress of 40 Ksi is used. As Rodabaugh explains in Background of ANSI B16.5 Pressure Temperature Ratings (API 54-72).

Also, in the same report, Rodabaugh answers certain critical questions –

If ASME SEC VIII Div 1 Appendix 2 is implemented, then there is no way other than to convert the applied external bending moment and axial force as an “equivalent pressure,” which is what CAESAR II does. This approach can be seen as overly conservative, and the following is taken from ASME PVP-97814.

If however this method has to be implemented ( as many clients demand in their engineering standards), a few relaxations can be done, like checking if the computed bolt stress in seating and operating conditions are within 65% of SMYS for SS and 85% of SMYS for CS; alternately, if using the approach stated in ASM PCC-1 Appendix O, it can be shown that the combination of internal pressure and external bending moment and axial force is within the limits of equation (8), the joint can be qualified as acceptable.

For calculating flange stresses, one needs to calculate the flange moment, which is dependent on bolt load. Bolt load has to be calculated for two design conditions: operating & gasket seating and the most severe will govern. For more details on the equations and calculation methodology, the above-mentioned code can be referred to.

Some more ready references for you:

Flange Selection Guidelines
Pressure Equivalent Method in Caesar II
Flange leakage calculation ASME Section VIII in Caesar II
Flange leakage calculation NC 3658.3 method in Caesar II
Procedure for Flange Bolt Tightening of Various Sizes of Flanges

4.3 Flange Leakage checking by NC 3658.3 Method

In this method, the flanges are evaluated using the ASME BPVC Section III Subsection NC-3658.3 method. The calculated flange moments are compared to some limited values as calculated from the code equations.

The theoretical basis for this method lies in the computation of bolt stress due to pressure and applied bending moment and ensuring that the flange is not overstressed. This brings out the inherent weakness of not addressing the issue of gaskets. This method also presupposes an application of 40 Ksi of tightening stress for B16.5 flanges.

This method is recommended for high-strength bolts (allowable stress >20 Ksi) and for B16.5 Flanges. However, as the basis of computation is a control on bolt stress under pressure and bending moment and ensuring that flanges are not overstressed, its use in B16.47 Flanges can logically be extended, ensuring that the bolts and flanges are not overstressed. Moreover, as long as the applied bolt stress meets the requirements of ASME PCC-1 Appendix O, it ensures that the issue of adequate gasket compression is also taken care of, of course, subject to the condition that the flange is within the size limitation of ASME PCC-1 (48 inches). Some words of caution on the use of NC3658.3 from ASME PVP2013-97814 are shown below

For more details on the Caesar II application of the NC method, click here.

4.4 Flange Leakage Analysis Using EN-1591

The EN 1591-1 calculation code offers a comprehensive method for analyzing flange leakage by considering the behavior of all components involved—flanges, bolts, and gaskets. Unlike simpler methods, EN 1591-1 accounts for factors such as gasket thickness reduction due to flange stress and changes in gasket elasticity with temperature variations. This advanced approach provides not only an allowable stress check on the components but also an indication of the expected leak tightness of the flange assembly. It is applicable to both regular piping flanges and custom-designed body flanges for equipment. The code incorporates gasket characteristics based on EN 13555, including maximum allowable surface pressures, modulus of elasticity, and minimum seating pressure for various tightness classes. Additionally, it factors in the coefficient of thermal expansion of flange and bolt materials

4.5 Some Other Flange Leakage Methods

4.5.1 Blick’s method

In essence, this method stands for -Internal Pressure force on gasket + residual compressive force on gasket + force on gasket due to applied bending moment should be equal to operating stress on each bolt times total bolt area.

4.5.1.1 Derivation of Blick’s formula-

Blick's formula

4.5.1.2 Drawbacks of this method-

  • Relative deformation and stiffness of bolts, gaskets, and flanges are not considered.
  • Flange stress is not discussed.
  • How to calculate operating stress on bolts is not discussed.
  • The basis for the m factor in ASME B&PV code Sec VIII Div 1 Appendix 2 is not known, and hence their accuracy is a question by itself.
  • An improvement of this method can be to compute 𝑆𝑂𝑃𝑇 using ORNL2913.3 [8] and ASME PCC-1 Appendix O

4.5.1.3 Modified Blick’s Formula

Blick proposed modification to the above formula (eon) and the modified version of Blick’s method is

Modified Blick's formula

4.5.2 ASME B31.8 Method of Flange Leakage Analysis

The essence of this method is that when the force on the BJF due to applied internal pressure and applied bending moment (converted to a force) reaches a critical value, residual stress on the gasket equals zero.

As per ASME B31.8, Table E1, Note 14, the moment to produce leakage of a flanged joint with a gasket having no self-sealing characteristics can be estimated by the following equation:

4.5.2.1 Drawbacks of the method-

  • The effect of applied loading on Flange and hub stress is not discussed
  • Gasket stress is not discussed.
  • Bolt preload and the load reduction in the bolt are not discussed.
  • Relative deformation and stiffness of bolts, flanges, and gaskets are not discussed.

4.5.3 New method developed by Integrity Solutions (Warren Brown), Australia

The following is taken from the paper ASMEPVP2013-97814, citing the relevant equation for the new method:

Flange Leakage Checking by Warren Brown Method

5. Recommended Good Practices to Avoid Flange Leakage

Here are some of the recommended practices that can be followed to avoid flange leakage issues:

  • The acceptable limit of bolt stress during tightening- 65% of SMYS for SS and 85% of SMYS for CS.
  • Use of proper lubricant- Molybdenum disulphide.
  • Bolt load scatter should be considered with a typical value of 5-10% of the target value.
  • Although tensioning, in general, is a better approach than torquing, but for l/d ratio or length-to-diameter ratio of bolts to <=3, tensioning is normally not recommended.
  • Two factors govern the loss when using bolt tensioning: Tool load loss factor (TLLF) and Flange load loss factor (FLLF); the latter being the additional case when 100% tensioning is not possible/applied. This aspect has to be considered when applying the right magnitude of bolt stress. The challenge will be to ensure that even with consideration of TLLF and FLLF, the bolt stress should not go beyond 65% of SMYS for SS and 85% of SMYS for CS.
  • Ambiguity appears between sections 5.3.4 and 5.4.2 of B16.5 on the selection of a proper gasket for low-strength bolts. Since the use of low-strength bolting does not allow high gasket stress levels to be achieved during assembly, this issue should be considered in selecting a proper gasket for use with low-strength bolts. The same recommendation should be followed in B16.47 (for this document, the ambiguity on the same issue is exhibited in the paragraphs with the same no’s as in B16.5). For B16.5 the referenced edition is 2013 and for B16.47 the referenced edition is 2010.
  • To avoid problems with sealing, it is recommended that the purchased flanges be having a 3.2*10-3 m concentric groove surface finish with a nominal 5 mm tool nose radius and 0.45 mm pitch. This is in relation to avoiding problems with leakage with spiral finish and 250*10-3 mm surface finish (section 6.4.5.3 of B16.5). The same recommendation should be followed in B16.47 (relevant paragraph 6.1.4.2 in B16.47) For B16.5 the referenced edition is 2013 and for B16.47 the referenced edition is 2010.
  • For accepting flanges with imperfections, it is recommended that the limits outlined in ASME PCC-1-2013 Appendix D be used in lieu of Section 6.4.6 of B16.5; the same recommendation should be followed in B16.47 (the relevant paragraph is 6.1.5). For B16.5 the referenced edition is 2013 and for B16.47 the referenced edition is 2010.
  • It is recommended that the minimum hub height be limited to greater than 75% of the full hub height. The same recommendation should be followed for B16.47.
  • When using B16.20 gaskets with B16.47 Series A Class 150 flanges, custom dimensions should be specified ( the ideal gasket sealing element OD is between 6 mm and 12 mm smaller than the flange seating surface OD) such that the OD of the sealing surface is closer to the raised face O. The dimensions as they are in B16.20 can result in joint leakages.
  • For selecting minimum pipe wall thickness for the use of spiral wound gaskets with inner rings and B16.5 flanges, the minimum specified dimensions in B16.20 (Table 15 in the 2012 edition of B16.20) should be followed. Recommendations in Tables 16 and 17 of the 2012 edition (maximum bore of B16.5 Flanges for use with spiral wound gaskets and maximum bore of B16.47 Series A flanges for use with spiral wound gaskets) should also be followed.

References:

  • ASME B16.5
  • ASME BPVC SEC VIII
  • https://docs.hexagonppm.com/reader/O04dxcwMfibZIK1cyksZvQ/libWcNN9xB96GAfpXDAXlg
  • ASME PCC-1-2013 Guidelines for Pressure boundary bolted flanged joint assembly.
  • Demonstrating leak tight joints during Piping Design -ASME PVP2011-57923 by David Meir, ASME PVP conference , Baltimore 2011.
  • Dissecting the dinosaur: problems with B16.5 and B16.47 Flange standards ASME PVP2013-97813 by Warren Brown, ASME PVP conference, Paris 2013
  • Improved analysis of external loads on Flanged joints, ASME PVP2013-97814 by Warren Brown, ASME PVP conference, Paris 2013.
  • WRC Bulletin 473-External bending moments on bolted flanged joints
  • Behaviour of bolt force changes of flanged connections under thermal loading ASME PVP2013-97730 by M.Hiratsuka, T.Kobayasi
  • FLANGE: A computer program for the analysis of Flanged joints with ring-type gaskets- by E.C.Rodabaugh and H.C.Moore
  • Evaluation of the bolts and flanges of ANSI B16.5 Flanged joints-ORNL 2913.3, E.C.Rodabaugh and E.C.Moore
  • Working bolts near to yield -Theory, experience and recommended practise -Robert Noble, PVP2013-97956, ASME PVP conference, Paris 2013.

Some parts of this article are taken from Mr. Anindya Bhattacharya’s studies on Bolted Flange Joints

Load Cases for Pipe Stress Analysis: Caesar II Load Cases

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

Piping Stress Analysis is simply creating the load cases required for analysis and studying the impact of the same on the behavior of the critical piping systems. A load case can be defined as a set of loads (Weight, Pressure, Temperature, External Forces, Displacements, etc) and boundary conditions for defining a particular loading condition. So Stress Analysis can not be thought of without proper load case creation. Sometimes these load cases are mentioned in the piping stress analysis design basis. In this article, we will learn the basic load cases that are required for stress analysis activity.

Objectives of Pipe Stress Analysis:

The main objectives of stress analysis are to ensure:

  1. Structural Integrity (Design adequacy for the pressure of the carrying fluid, Failure against various loading in the life cycle, and Limiting stresses below code allowable.)
  2. Operational Integrity (Limiting nozzle loads of the connected equipment within allowable values, Avoiding leakage at joints, Limiting sagging & displacement within allowable values.)
  3. Optimal Design (Avoiding excessive flexibility and also high loads on supporting structures. Aim towards an optimal design for both piping and structure.)

What is a Load Case in Pipe Stress Analysis?

In the context of pipe stress analysis, a load case refers to a specific scenario or combination of loads that the piping system may experience during its lifecycle. These loads can arise from various sources such as internal pressure, temperature changes, external forces, and more. By analyzing these load cases, pipe stress engineers can predict how the piping system will behave under different conditions, allowing them to design systems that can withstand these stresses without failure.

A typical example of a load case in Caesar II pipe stress analysis may constitute the thermal, deadweight, and pressure loads together known as the operating load case. Similarly, a sustained load case is composed of dead weight and pressure load. Again, a load case can also be formed by combining the results of other load cases. For instance, a load case might represent the difference in displacements between the operating condition and the installed condition.

Notations Used for Load Cases in Caesar II

To meet these objectives several load cases are required during stress analysis. In this article we will use the following notations for building load cases:

  • WW=water filled weight of the piping/pipeline system,
  • HP=Hydrotest Pressure,
  • W=Weight of pipe including content and insulation,
  • P1=Internal Design pressure,
  • T1=Maximum Design temperature,
  • T2=Maximum Operating temperature,
  • T3= Minimum Design temperature,
  • WIN1, WIN2, WIN3, WIN4: wind loads acting in some specific direction,
  • U1, U2, U3, U4: uniform (seismic) loads acting in some specific direction.

Basic Load Cases for Caesar II Pipe Stress Analysis:

For Stress analysis in Caesar II, Various Load case combinations are used which serve several purposes. While analysis at a minimum, the stress check is required for the below-mentioned cases:

a. Hydrotesting case:

Piping/ Pipeline systems are normally hydro-tested (sometimes pneumatic tested) before the actual operation to ensure the absence of leakage. Water is used as the testing medium. So during this situation pipe will be subjected to water-filled weight and hydro-test pressure.
Accordingly, our first load case will be as mentioned below

1WW+HP HYD
Load Case for Hydrotest Case

b. Operating and ALT Sustained load cases:

When the operation starts working fluid will flow through the piping at a temperature and pressure. Alt Sustained cases are used as Hot Sustained cases which means sustained stress that the system carries during operation. So accordingly our operating load cases will be as mentioned below:

2W+T1+P1 OPE For operating temperature case at maximum design temperature
3W+P1 SUS Alt Sustained case based on operating case 1 (T1)
4W+T2+P1 OPE For a maximum system operating temperature case
5W+P1 SUS Alt Sustained case based on operating case 1 (T2)
6W+T3+P1 OPE For minimum system temperature case
7W+P1 SUS Alt Sustained case based on operating case 1 (T3)
Operating and Alt-Sustained Load Cases

c.  Sustained Load Case:

Sustained loads will exist throughout the plant operation. Weight and pressure are known as sustained loads.  So our sustained load case will be as follows:

8. W+P1 SUS
Sustained Load Case

d. Occasional Load Cases: 

Piping may be subjected to occasional wind and seismic forces. So to check stresses in those situations we have to build the following load cases:

9W+T2+P1+WIN1 OPEConsidering wind from +X direction
10W+T2+P1+WIN2 OPEConsidering wind from -X direction
11W+T2+P1+WIN3 OPEConsidering wind from +Z direction
12W+T2+P1+WIN4   OPEConsidering wind from -Z direction
13W+T2+P1+U1 OPEConsidering seismic from +X direction
14W+T2+P1-U1 OPEConsidering seismic from -X direction
15W+T2+P1+U2   OPEConsidering seismic from +Z direction
16W+T2+P1-U2 OPEConsidering seismic from -Z direction
Occasional Load Cases in Operating Condition

While stress analysis the above load cases from load case 9 to load case 16 are generated only to check loads at node points. Figure 1 shows typical load cases that should be generated during stress analysis

Typical Load Cases
Fig. 1: Typical Load Cases

To find occasional stresses we need to add pure occasional cases with sustained load and then compare them with code allowable values. The following sets of load cases are built for that purpose.

17L9-L4 OCC Pure wind from +X direction
18L10-L4 OCC Pure wind from -X direction
19L11-L4 OCC Pure wind from +Z direction
20L12-L4 OCC Pure wind from -Z direction
21L13-L4 OCC Pure seismic from +X direction
22L14-L4 OCC Pure seismic from -X direction
23L15-L4 OCC Pure seismic from +Z direction
24L16-L4 OCC Pure seismic from -Z direction
25L17+L8 OCC Pure wind+Sustained
26L18+L8 OCC Pure wind+Sustained
27L19+L8 OCC Pure wind+Sustained
28L20+L8 OCC Pure wind+Sustained
29L21+L8 OCC Pure seismic+Sustained
30L22+L8 OCC Pure seismic+Sustained
31L23+L8 OCC Pure seismic+Sustained
32L24+L8 OCC Pure seismic+Sustained
Occasional Load Case Sets

Load cases from 25 to 32 will be used for checking occasional stresses with respect to the code ASME B31.3 allowable (=1.33 times Sh value from code). Use scalar combination for load cases 25 to 32 above and algebraic combination for others as shown in Fig. 2 attached below:

Load Case Combination
Fig. 2: Load Case combination

e. Expansion Cases:

Following load cases are required for checking the expansion stress range as per the code

33L2-L8 EXP
34L4-L8 EXP
35L6-L8 EXP
36L2-L6 EXP
Expansion/Thermal Load Cases

The above load cases (from 33 to 36) are used to check the expansion stress range

The above-mentioned load cases are the minimum required load cases to analyze any stress system. Out of the above load cases, the load cases mentioned in load case numbers 1, 3, 5, 7, 8, and 25-36 are used for stress checks. The load cases mentioned in load case numbers 1, 2, 4, and 6 to 16 are used for checking restraint forces, displacements, and nozzle load checking.

Some additional load cases may be required for PSV-connected systems, systems having surge or slug forces, and rotary equipment-connected systems.

Seismic and Wind analysis may not be required every time. So those load cases can be deleted if the piping system does not fall under the purview of wind and seismic analysis by project specification. However, to perform wind and seismic analysis proper related data must have to be entered in the Caesar II spreadsheet.

If the stress system involves the use of imposed displacements (D) and forces (F) then those have to be added with the above load cases in the form of D1, D2, or F1, F2 as applicable.

Better Engineering Practices for Stress Analysis

It is a better practice to keep:

  • Hydro and sustained stresses below 60% of the code allowable
  • Expansion and occasional stresses below 80% of the code allowable
  • Sustained and Hydrotest sagging below 10 mm for process lines and below 3 mm for steam, two-phase, flare lines, and free-draining lines.
  • Design/Maximum displacement below 75 mm for unit piping and below 200 mm in rack piping.

Video Tutorial for Load Case Creation in Caesar II

The following video tutorial explains the load case creation steps with an example

Caesar II Load Case Video tutorial

Online Course on Pipe Stress Analysis with Practical Example

Complete Pipe Stress Analysis using Caesar II Online Course (30+ Hours)

References (External Links):

Introduction to Pressure Surge Analysis

Written by Anup Kumar Dey in Process

What is Pressure Surge or Water Hammer?

A pressure surge is a pressure wave that is caused by the kinetic energy of the moving fluid when there is a sudden change in flow velocity. Due to the instantaneous conversion of momentum to pressure when flowing liquid stopped quickly this sudden increase or surge of pressure is experienced. Pressure surge is popularly known as Water Hammer, Fluid Hammer, or Hydraulic Surge.

In Piping/Pipeline system networks this phenomenon is a major concern for Piping/Pipeline/Process engineers. As noticed in the below graph (Fig. 1), the pressure spike will continue hitting the pipe/pipeline trying to release the generated excessive energy and therefore the system will be at high risk.

Typical Pressure Surge Curve
Fig. 1: Typical Pressure Surge Curve

ASME B31.3 defines that the pressure rise due to pressure surge and other normal operation variations shall not exceed the internal design pressure at any point in the piping system and equipment by more than 33%.

What Can Cause Pressure Surge?

Pressure Surge or the sudden change in velocity and or pressure can arise due to various reasons. Hydraulic transients occur at changes in flow in piping/pipelines and this could be due to the:

  • Pump start & stop, specifically due to load shedding or sudden power failure
  • Quick operation of Valve (Sudden closure/opening)
  • Sudden closure of the check valve
  • Presence of Air pockets inside piping/ pipeline systems, especially during pump start
  • A sudden release of Air
  • Quick Pipeline filling
  • Pressure Surges can occur in open channels and partly liquid-filled pipes, as well

All of the above causes will generate high-pressure waves that can travel both upstream and downstream from point of origin. Please note that

  • Some pipelines are in transient operations over 75% of the time.
  • Pressure Surge (pressure rise) increases as the pipeline straight length increases since the contained momentum within two direction changes (elbows/Tee) will be higher (more volume).
  • A pressure surge normally consists of multiple events, resulting in up to ten times the normal pipeline pressure. When a surge relief valve opens, it vents the pressure to a safety system.
  • Surge pressure is created during the last 20% of valve closure.

Is Water Hammer Dangerous?

Refer to Fig. 2 to understand what a pressure surge can cause to a Piping System. Pressure Surge of Significant nature creates high pressure and velocity rise that can lead to:

  • Failure of pipe/pipeline fittings
  • Bursting of pipes
  • Damage to the Pump/pumping system
  • Deformation of valves and piping supports
  • Vibration or shaking of the piping/pipeline system
Consequences of Pressure Surge
Fig. 2: Consequences of Pressure Surge

Basic Definitions concerning Pressure Surge:

  • Pressure Surge:– It is basically a pressure wave caused due to a sudden change in flow velocity.
  • Wave speed or acoustic velocity:– The velocity at which pressure waves travel through the liquid/fluid.
  • Joukowsky equation:– Relationship relating head change to velocity change and acoustic velocity.
  • Pipeline Period:– Time required for a pressure wave to traverse the pipe/pipeline length and come back.
  • Pressure Head:– Pressure is measured as the height of fluid (10 m head of water is roughly 1 atmosphere)
  • Effective Valve closure Time: The period over which a Valve reduces the flow from 90% of its steady state to zero. In relation to Total Valve Closure Time, this is typically the last 15% opening for butterfly valves, 25% opening for ball valves, and 30% opening for plug valves. This can be used as a rule of thumb during the initial assessment phases.

Analysis of Water Hammer/Pressure Surge

The most important parameters to estimate the magnitude of transient pressures is:

  • Acoustic wave speed, a
  • Pipe/Pipeline period, T
  • Joukowsky head, Δh

The acoustic wave speed formula depends on the fluid and the pipe characteristics expressed as:

acoustic wave speed formula
  • a = Velocity of the pressure wave
  • K = Bulk modulus of the fluid
  • ρ = Liquid density
  • D = Internal diameter of the pipe
  • E = Young’s modulus of the pipe material
  • e = Wall thickness
  • ϕ = restraint factor (usually taken as 1)
Variation of wavespeed with pipeline characteristics
Fig. 3: Variation of wave speed with pipeline characteristics

The time that a pressure wave takes to travel from its origin through the system and back to its source is defined as the pipe period. For a single pipeline with pipeline Length, L this is provided as given below:

  • T = Critical period
  • L = Length of the pipe
  • a = Velocity of the pressure wave

Events that take place in less than T are called ‘fast’ events and these are likely to cause pressure surge issues.

Joukowsky formula
Fig. 4: Joukowsky formula

As per the Joukowsky formula, the pressure head change (Δh) due to an instantaneous velocity change (ΔV) is expressed as shown above. Here,

  • Δh = head rise
  • ΔV = change in velocity
  • a = wave speed
  • g = acceleration due to gravity

This is a very useful guide that explains the likely severity of a pressure surge event but is not a replacement for a proper surge analysis!

Limitations of the Joukowsky formula

Joukowsky formula is applicable to a limited set of fluid systems.

  • Its application should be limited to situations matching the following criteria:
    • Simple ‘linear’ piping systems i.e. there are no branches by which pressure waves can be reflected back and cause constructive interference in the main line.
    • Valve closure time is significantly shorter than the pressure wave communication time.
    • System frictional losses are similar to that of a water transport system.
  • Joukowsky equation does not consider column separation in its analysis of fluid hammer. Column separation can often result in surge pressures exceeding those predicted by the Joukowsky equation and therefore the Joukowsky equation should not be applied when analyzing a system in which the pipeline pressure can rapidly drop below the fluid vapor pressure.

How to Avoid Pressure Surge

To avoid pressure surge system must be protected. Protection of systems against water hammer can be parted into three groups:

1.0: System Design Solutions:

  • Use of pipework with a higher pressure rating i.e to make the pipework stronger to withstand the effects of surge pressure (Normally followed for radioactive, highly corrosive, or lethal fluids, where no fluid is allowed to escape.)
  • Rerouting of the pipeline avoiding high/low points
  • Changing of piping material, thus altering the wave speed
  • Increase the pipe diameter, thereby reducing the velocity
  • Increase pump inertia by incorporating a flywheel
  • Adding bypass lines
  • Providing Additional Pipe supports: By adding more supports in the piping system, the natural frequency of the system is increased. So, vibration tendency will reduce. Also, providing support near concentrated mass will reduce high local stresses.

2.0 Active Protection:

Piping/Pipeline systems can be protected against Surge impact by using devices during pipeline normal operation like:

  • Variable speed pumping: Variable speed drives provide a reliable means of preventing damage from pressure surge events.
  • Soft starters: The primary purpose of Soft starters is to reduce the electrical load on the power supply to a facility.
  • Slow closing and opening valves: A common form of pressure surge initiation is due to the rapid closing of a valve. Extending the closure times attenuates the pressure surge possibility.

Be informed that these devices require power and during load shedding or power failure cannot be of use.

3.0: Passive Protection (Fig. 5):

Passive Equipments for Surge Protection
Fig. 5: Passive Equipment for Surge Protection

There are several passive protection equipments available in the market that operates without the need for additional power. A few examples of these are:

  • Surge Vessels
  • Surge Shafts
  • Air Valves
  • Vacuum Breakers
  • Pressure Relief Valves/Surge relief Valve: Click here to know more about Surge relief valves
  • Surge Anticipation Valves: A surge anticipation valve is specially designed to provide a diversionary fluid flow during a pressure surge event.
  • Intermediate Check valves: In a long pipeline, an intermediate check valve has the ability to prevent the damaging reverse velocity from reaching a pump station. It effectively reduces the pressure surge to half.
  • Gas Accumulators: The gas accumulator is particularly effective in pressure surge scenarios due to a loss of power situation when downstream of the pump check valve, a negative pressure wave develops immediately. The deceleration of the liquid column is reduced by the residual pressure in the gas accumulator and prevents column separation. However, the gas accumulator should be located close to the boundary element that causes the transient event.
  • Liquid Accumulators: A liquid accumulator is a vessel that has lower elasticity than the pipe itself. The vessel will exhibit strain to a higher degree than the pipe and thus mitigate pressure transient.
  • Using low-modulus thermoplastic materials in combination with ferrous materials can mitigate a pressure surge.

Selection of System for Surge Protection

Refer to Fig. 6 below which provide a flowchart for Surge Protection System Selection.

Selection of System for Surge Protection
Fig. 6: Selection of System for Surge Protection

Pressure Surge Modelling Software

There are currently various software packages that can be used for analysis:

  • HyTran
  • Flowmaster
  • WANDA
  • Hammer
  • AFT Impulse
  • PIPENET
  • PTRAN
  • PASS/Hydro system
  • Flownex Simulation Environment

Methodology (Fig. 7):

Surge Analysis Methodology
Fig. 7: Surge Analysis Methodology

Designing a Pressure Surge Relief System

  • Consideration of a complex range of factors like the potential for pressure increases, the volumes to be passed by the surge relief equipment in operation, and the capacity of the system to contain pressures, etc are required for the design of a complete surge relief system.
  • Control or ESD valve closure times can also affect surge pressures in a pipeline. By increasing the valve closure time, a gradual flow decay can be achieved which will reduce the potential for pressure surge.
  • Control narrative and system interlocks to ensure Staged pump shutdown sequence and linked ship/shore ESDs when your facility is linked to loading berths/jetties.
  • Carry out transient / surge analysis using detailed computer modeling using the software mentioned above to simulate the complex interactions of equipment, pipelines, and fluid to normal, fault, and emergency events.
  • Design piping to withstand maximum surge pressure – MSP.
  • Although many design approaches can help reduce surge pressures in pipelines, going for a higher pipe rating or massive support arrangements aren’t recommended for the associated significant cost, and a surge relief valve was found to be the most feasible option to protect the system.
  • A correctly designed surge relief system will include components to dampen or slow the relief valve on closing, and this often requires sophisticated reverse flow plots.
  • In nitrogen-loaded Surge Relief valves, attention must be paid to the nitrogen gas system. The nitrogen system must supply a constant pressure (set point) to the modulating valve, even under conditions of varying ambient temperatures. Normally, the system is designed to use standard gas bottles and has its own control system to regulate the nitrogen supply pressure.

Conclusions:

  • The Pressure Surge phenomena during transient events are very important as they can put the system’s integrity at high risk.
  • During risk and HAZOP analysis, Pressure Surge Events and the corresponding mitigation devices should be always taken into account.
  • System operations staff must be trained in order to prevent operations likely to damage the system’s integrity.
  • Surge protection equipment must be maintained periodically.
  • It is highly possible to increase the reliability and life expectancy of systems by taking preventive measures for reducing the risk of failure due to pressure surge events,
  • The pipe/Pipeline system should be properly supported with the hold downs, guides, and line stops and the supports along with supporting structures must be designed considering dynamic forces during a Surge event.

Some more Resources for you..

Understanding Centrifugal Compressor Surge and Control
Water Hammer Basics in Pumps
Pipe Stress Analysis from Water Hammer Loads

References:

Frequently Asked Questions

What is Pressure Surge?

Pressure Surge is a pressure wave that is caused by the kinetic energy of the moving fluid when there is a sudden change in flow velocity.

What is the pressure surge in piping?

If the high-velocity flow in a pipe is forced to stop or change direction suddenly, a pressure wave generates and moves back at the speed of sound in the liquid. This can produce huge forces in the piping or pipeline system. This is called Pressure Surge in Piping

What is the difference between Pressure Surge and Water Hammer?

Pressure Surge, Water Hammer, Fluid Hammer or Hydraulic Surge, all these refer to the same event. There is no difference.

What Can Cause Pressure Surge?

The Pressure Surge in a Piping system can be caused by any of the following Events:
1. Pump start & stop, specifically due to load shedding or sudden power failure
2. Quick operation of Valve (Sudden closure/opening)
3. Sudden closure of the check valve
4. Presence of Air pockets inside piping/ pipeline systems, especially during pump start
5. A sudden release of Air
6. Quick Pipeline filling

How to Avoid Pressure Surge?

Pressure Surge can be avoided by the following methods:
1. Rapid Changes in fluid velocity occurs when valves are opened or closed suddenly. So by reducing the fluid velocity or by increasing the time taken for closing/opening the valve it can be avoided.
2. Surge can be avoided by installing Surge Relief Valve, Surge Tank, Viscous Damper, etc in the system.
3. The impact of surge can be reduced by reducing the number of elbows.
4. Eliminate the Presence of Air

What is Surge Analysis?

Surge Analysis is the analysis of pressure changes in the piping system, normally performed by Process Engineers for proper pipe sizing or finding the peak surge pressure.