Geothermal Pipe Flow Systems

Geothermal power plants utilise hot water and steam from deep underground dry steam or hot water wells which is piped upwards and used to generate electricity. Technologies used in geothermal power plants typically includes dry steam, flash steam or binary cycle power plants.

Dry Steam Plant

In dry steam plants, hot geothermal steam is piped from geothermal reservoirs which then drives turbines. The turbine powers a generator which produces electricity and adds to the power distribution network. In dry steam plant, the steam is transferred from the well to the powerhouse via the gathering piping system. This piping system generally consists of one or more large diameter mains pipelines which collects the steam from the smaller diameter pipelines connected to each wellhead. As the main pipe approaches the powerhouse, the mass flow rate increases as each wellhead is connected into this common mains pipeline. One of the main considerations in the design of the gathering system is the selected diameter of the pipelines. For steam transportation pipelines, an allowable maximum velocity of circa 50 to 60 m/s for large diameter lines (>300mm pipes) whereas for smaller diameter pipes the velocity of circa 20 to 25 m/s is recommended. These velocity limits are recommended as the presence of water droplets or fine solid particles carried by high velocity steam can cause erosion of valve seats, bends as well as any other exposed parts.

At the well, there is a valve arrangement with steam purifier (inline axial centrifugal separator) which is designed to remove water droplets and particulate matter from the steam before it enters the piping system. The steam pipelines are fully thermally insulated and typically include expansion loops to accommodate any pipe movement. Steam traps are positioned strategically along the pipeline to remove condensate which is then transferred by a separate piping system to holding ponds before being reinjected. As the main gathering line approaches the powerhouse, a safety relief valve station is provided to permit the safe removal of steam in the event of turbine trip. Safety pressure relief valves can be automatically sized in FluidFlow to ISO & API Standards. Experience shows that it is better to operate the wells in a steady open state instead of changing between open and closed positions.

Flash Steam Plant

Single-flash steam plant is said to be the most widespread type of geothermal power plant. This plant is often the first power plant type installed at a newly developed liquid-dominant geothermal field. In geothermal wells which produce a two-phase mixture of liquid and steam, the single-flash plant offers a simple method of converting the geothermal energy into electricity. The mixture is firstly separated into two distinct liquid and steam phases with a minimum loss of pressure. This step is completed in cylindrical cyclone separators where the water and steam phases separate due to their inherent density differences. Separators are often positioned at the powerhouse, at satellite stations in the geothermal field or at the wellheads.

One of the main considerations in the design of the gathering system is the pressure loss in the steam lines from the wellhead to the powerhouse. The pressure loss in steam pipelines is a function of the configuration of the piping system such as pipe length, diameter as well as the mass flow rate and physical properties of the steam. The most critical design parameter is the pipe diameter since the pressure drop is the function of internal diameter to the fifth power. Installing larger diameter pipes reduces the pressure loss however the extra cost of installing larger diameter pipelines may not be feasible. An optimum pipe diameter should be selected based on a thermodynamic study whilst also giving due consideration to capital cost. Software tools can assist with selecting optimum pipe diameters and FluidFlow offers three automatic pipe sizing approaches; Economic Velocity Sizing, Design Velocity Sizing and finally, Design Pressure Gradient Sizing.

The pressure loss in liquid pipelines is less of a concern since the liquid shall be disposed of by injection. However, high pressure losses may require pumps to maintain sufficient reinjection pressure.

The pressure loss in two-phase (steam-liquid) pipes is a much more complex phenomenon and difficult to predict accurately using tools such as spreadsheets as this approach doesn’t accurately consider the interaction of connected equipment. The solution of two-phase pipe flow systems is further complicated by the presence of different flow patterns which can change with pipe orientations etc. The mechanism for pressure loss is different for each flow pattern and therefore the use of a suitable software tool can assist greatly in this regard. FluidFlow includes several two-phase correlations and determines the flow pattern in each pipeline and solves accordingly. The software also plots the flow pattern map for each pipe segment. This allows the designer to identify the flowing conditions, flow regime and also identify how close the pipeline is to undesirable flow regimes such as slug flow conditions.

The double-flash steam plant is an evolution of the single-flash plant in that it typically can produce around 15 to 25 % additional power output for the same geothermal fluid conditions. The main difference is the presence of a second flash process which takes place on the separated liquid leaving the separator. This process generates additional steam at a lower pressure than that of the primary steam. This system therefore extracts even more energy by comparison to a single-flash system.     

In double-flash steam plant, several different piping arrangements (in comparison to single-flash plant) can be utilised to provide the additional flash process. They can generally be summarized as follows:

  • Two-phase pipe connections from each well to satellite separation stations in the field. High and low-pressure steam lines are then derived from each satellite separation station and piped to the powerhouse. Hot water pipelines from the satellite stations is then reinjected to the wells.
  • Two-phase pipe connections from each well to satellite separation stations in the field. High pressure steam lines and hot water lines derived from each satellite separation station and piped to the powerhouse. The power house includes separators where short low-pressure steam pipelines are connected to the turbines and hot water pipelines from the powerhouse are piped to the injection wells.
  • Separators at each well head with high and low-pressure steam lines piped to the powerhouse. Hot water is piped from each production well to the injection wells.
  • Two-phase pipe connections from each well piped to the powerhouse with separators located at the powerhouse. This system features short lengths of high and low-pressure steam lines to the turbines. The hot water pipelines are piped from the powerhouse to the injection wells.

A combination of the above arrangements may also be considered depending on the field conditions. The final selected design solution will take account of geofluid conditions, field conditions, well positions, topography position of powerhouse etc.

Figure 1 provides an illustration of a double-flash two-phase flow system as modeled in FluidFlow.


Figure 1: Geothermal Double-Flash Two-Phase Flow System.

Flash steam plants are very common and use fluid temperatures in the range of circa. 182oC.

Binary Cycle Plant

Binary cycle power plants use the warm geothermal fluid (typically below 204oC) to heat a secondary fluid (such as isobutane) which exhibits a much lower boiling point. The fluids are kept separate with heat energy transferred from the geothermal fluid to the secondary fluid via a heat exchanger. This causes the secondary fluid to flash into steam which drives the generator turbines. 

Binary cycle power plants are essentially closed-loop systems and virtually only water vapor is emitted to the atmosphere. The most common geothermal resources exhibit temperature below 149oC and as such, a considerable proportion of future geothermal electricity could come from binary cycle power plants.

A combination of both flash steam and binary cycle technologies has also been utilized effectively to take advantage of both technologies. In this plant, the flashed steam is first converted to electricity with a backpressure steam turbine and the low-pressure steam exiting the turbine is then condensed in a binary system. This allows for the effective use of air cooling towers with flash applications and takes advantage of the binary process. The flash/binary system has a higher efficiency where the well field produces high pressure steam while the elimination of vacuum pumping of non-condensable gases allows for 100 percent injection. Examples of these systems include the Puna Geo Venture facility in Hawaii where the geothermal fluid is over 316oC, the Upper Mahiao geothermal facility in the Philippines as well as several power projects in New Zealand.

A cooling system is essential for the operation of modern geothermal power plants. Cooling towers prevent turbines from overheating and prolong facility lifespan. Many geothermal power plants use water cooling systems. Figure 2 provides an illustration of a cooling water system as modeled in FluidFlow.


Figure 2: Cooling Water System.

FluidFlow software is used to design and model geothermal two-phase flow systems with separator stations, injection systems as well as cooling water systems. Companies which use FluidFlow to model geothermal piping systems include Philippine Geothermal Production Company, MTL Consulting Engineers in New Zealand and EFLA Consulting Engineers in Iceland.

FluidFlow is used by EPC’s as well as owner operators to design, develop and optimise piping systems for geothermal, coal and natural gas power plants. The systems modeled include natural gas delivery systems, water and steam systems (single and two-phase), air systems, ash handling systems, chemical systems and water treatment systems. As part of this process, designers utilise the powerful automatic equipment sizing tool to size pipes, fans, pumps, control valves, safety relief valves and orifice plates.

This discussion attempts to outline some of the basic principles of geothermal power plant systems. Further reading is recommended and the references outlined serve as a useful resource.


  1. Pressure Drop in Large Diameter Geothermal Two-Phase Pipelines.
  2. A Guide to Geothermal Energy and the Environment, Geothermal Energy Association.
  3. Flow and Heat Transfer in Geothermal Systems, Aniko Toth, Elemer Bobok.
  4. Geothermal Power Plants Principles, Applications, Case Studies and Environmental Impact, Ronald DiPippo.

Two-Phase Liquid-Gas Flow System Design

The design of two-phase liquid-gas pipe flow systems is a complex phenomenon. It can be common practice for engineers to use template design calculation sheets for two-phase systems. These calculation sheets often utilize rules-of-thumb as part of the template for design. This can result in the failure to acknowledge inherent differences from one project to the next.  In large-scale projects such as geothermal systems, errors in system design such as underestimating friction losses can become magnified resulting in poor design with inherent operational difficulties.

In this paper we will make a hand calculation for a pipe system that has both vertical and horizontal pipe runs and then compare the results to those that can be obtained from a software based solution.

This piping system has 358ft of level pipe, three vertical rises of 10ft each and one vertical rise of 50ft. We wish to evaluate the type of flow and expected pressure drop.

Table 1: Fluid Physical Property Data.

Pipe Data: 3 Inch, Schedule 40 Stainless Steel (I.D. 3.068 in). Pipe Relative Roughness: 0.000587.

A hand calculation has been completed for this system, details of which are outlined below.

This calculation is based on the gas having a viscosity of 0.00127 cP and a density of 1.23 kg/m3.

To make a comparison, a model has been created in FluidFlow software. For convenience, based on the fluid physical properties, the liquid and gas used in the model is water and air. It is worth noting that the viscosity of air and water in FluidFlow is 0.018 and 1.13 cP respectively. This represents a very slight difference to the hand calculation.

FluidFlow Solution

The first step is to build the model using the above data. The model should appear as per Figure 1 below. Note, two Known Flow inlet boundaries have been defined, one with air and one with water. The connecting pipe has been set to 0.5m in length, 3 inch in diameter and the statues of these pipes (nods -1 & -3) have been set to “Ignore Pressure Loss” as, these pipes are only provided for model connectivity purposes and do not form part of the overall calculated system.


Figure 1: System Model


The system consists of a total length of 358ft of level pipe. As we have a “stepped” system i.e., changes in elevation, we have simply split the 358ft into 5 equal level pipe segments of 71.6ft each. Details of the various pipe lengths are set out in Table 2.

Table 2: Pipe Data.

The node elevations have also been set in accordance with the system design criteria. Details of which can be seen in Table 3 below.

Table 3: Node Elevation Data.

This system has been calculated using the various different two-phase pressure loss correlations available in FluidFlow. Table 4 provides details of the calculated results which are compared to the results provided in the literature.

Table 4: Table of Calculated Results.

*The results noted in Table 4 have been generated in FluidFlow v3.35.

Comparison of Results

  1. The first option selected is Whalley Criteria. This option lets the software select the appropriate method from the three correlations available when using this approach i.e., Friedel, Chisholm Baroczy & Lockhart Martinelli. In this example, FluidFlow has analysed the fluid physical properties (fluid viscosity ratio and mass flux) and has applied the Friedel correlation. Note, this correlation uses different equations for horizontal and vertical pipes. We can see that the results compare well with that in the hand calculation.
  2. Chisholm Baroczy: This approach dates back to the 1970’s and uses a two-phase multiplier. This is one of the more simplistic two-phase correlations. However, the results compare well when using this approach.
  3. Lockhart Martinelli: This forms the basis of the solution in the hand calculation. In general, it is accepted that this model should only be applied to horizontal pipe flow. This is also one of the more simplistic two-phase correlations and also dates back to the 1970’s. Again, the results compare well when using this approach.
  4. Drift Flux: This approach is recommended for modelling vertical or inclined pipelines.
  5. Beggs Brill: Recommended not to use this method for vertical upward flow as it under predicts pressure loss. It is difficult to assess the full effect of this in this particular example as only a fifth of this system consists of vertical pipework, i.e. most of the pressure loss occurs across the horizontal pipe segments.
  6. MSH: This approach is more useful when considering heat transfer and modelling single component fluids such as refrigerants. This method tends to lose accuracy when high vapor quality conditions are experienced in a system. This particular system has an outlet vapor quality of 0.75 (75%) and as such, it is considered that the results obtained using MSH may have a slightly lower level of accuracy when compared to, for example, the Friedel method.
  7. HEM: This method uses a simplified approach and averages the fluid density values. It also assumes the velocity, pressure and temperature between phases are equal. Due to the simplification of the modelling approach, it is recommended that this option should only be selected as a means of checking or validating results from the correlations noted above. It is also clear from the HEM approach that the total system pressure drop is significantly lower than that provided in the hand calculation and also when using other available loss correlations in FluidFlow.

Note, interestingly, the three approaches of Friedel, Beggs Brill and MSH which take into account flow orientation or flow regime offer a solution which is closest to the hand calculation example.

FluidFlow generates a flow pattern map for each pipe in the system. Figure 2 below shows the duty point for a sample pipe in this system which indicates the flow regime is annular mist which correlates well with the hand calculation.

Figure 2: Flow Pattern Map.


The literature calculation is based on the gas having a viscosity of 0.00127 cP and assumes gas ideality. For simplicity, the model has been developed using air and water which has a viscosity of approx. 0.018 and 1.13 cP respectively (based on 15oC). This will therefore have a slight effect on the calculated results.

The “hand” calculation is based on ideal gas conditions. FluidFlow does not assume gas ideality but solves for real gas conditions using an equation of state and hence, provides more accurate result.

Based on the above, it is considered that the results provided by FluidFlow correlate well with the hand calculation and offers an accurate reflection of the system operating conditions. It is also considered that the Friedel correlation may be best suited for this particular application owing to the combination of both vertical and horizontal pipes.

Note, solving this system in a suitable software package allows the designer to consider different correlation approaches and also identify the flow pattern map for each pipe in the system.

The design of two-phase piping systems may pose a daunting task to the designer. However, having a useful and easy to use software tool can help eliminate the likelihood of data entry errors, enable better understanding of the operating performance the system, help identify unstable operating conditions and also, consider different design conditions without the need to refactor any hand calculations. Furthermore, once a design has been finalized, you can revisit in the future should the operating system require expansion or alterations.



“MAN Diesel & Turbo’s activities in the power plant sector are based on a well-
established range of diesel engines and a rapidly growing gas engine offering. The
products range from small emergency power generators to turnkey power plants with
outputs of up to 400 MW. We use the liquid, gas, and two-phase modules of FluidFlow
and make extensive use of its simulation capabilities for engineering sub systems as fuel
gas lines or cooling water pressure systems, including for the development of new
systems. Before we bought we carried out extensive product research and chose
FluidFlow because of its completeness and value for money. We have had to make
occasional technical support calls and have been impressed by the responsiveness. We
have found Flite Software to be extremely knowledgeable and helpful, really excellent.”

Norman Kurth, System Engineering, MAN Diesel & Turbo, Germany.




The phenomenon of two-phase gas-liquid flow

Two-phase gas-liquid flow is a common phenomenon in nuclear, geothermal, chemical, oil and gas and refrigeration industries where gas-liquid mixtures are transported in piping systems. Two-phase pipe flow is the simultaneous flow of a gas (or vapor) and a liquid exhibiting different physical properties through the same pipe. A typical example of two-phase flow is a water-steam mixture. The prediction of the fluid physical properties and pressure drop in two-phase pipe flow is of vast importance during the design and operation stages in all applications to ensure the fluid does not deviate from its operational envelope and cause damage to personnel, plant and equipment. Due to the complexities of two-phase flow, it is commonplace to apply modelling techniques to predict the behaviour and characteristics of two-phase flow.

The physical properties of the two-phase fluid are not constant or uniform throughout the piping system and flow-induced pressure drops can cause considerable changes in magnitude of one phase compared to the other. For instance, some of the liquid can evaporate (flash) due to pressure drop across pipelines or through an orifice which results in an increase in gas/vapor. FluidFlow assumes that instantaneous isenthalpic flashing occurs. The software also identifies the physical properties of the fluid at the inlet and outlet of each pipe and fittings. A common feature of two-phase flow systems is that the flowing velocities increase and the mixture density decreases.

In cryogenic applications, two-phase flows are usually a mixture of a cryogenic liquid along with its corresponding vapor. A mixture of liquid helium and helium vapor is a typical example. An advantage of two-phase flows is that it can provide isothermal heat sinks. Due to the latent heat of boiling, heat added to the liquid portion of a two-phase flow will contribute to the conversion of the liquid to vapor in the mixture but will not increase the mixture’s temperature. This characteristic is used in cooling schemes such as those found at the Hadron Collider, HERA and Tevatron magnets.

There are also a number of disadvantages to two-phase flows. In comparison to single-phase pipe flows, these flows typically have a higher pressure drop and flow instabilities may develop that result in pressure surges and vibrations. Incorrectly designed systems can also trap the vapor phase against surfaces resulting in inadequate cooling or energy transfer and many flow components such as pumps and flow meters may not function properly in two-phase flows. Furthermore, in cases where the flow is on an incline, the liquid may flow to the downhill side. This may therefore necessitate the use of shorter flow paths to maintain any desired liquid-vapor ratio.

Due to these issues and complexities, detailed modelling is required whenever two-phase flows are used in cryogenic systems, to ensure proper performance. Care must also be taken that a nominally single-phase (say pure liquid) flow in a system does not suddenly become two-phase due to heat inputs or pressure drops resulting in unexpected operating problems.

Two-phase mixtures flow in pipelines in several configurations or structures called flow patterns or flow regimes. This describes the arrangement and relative fractions of the liquid and vapor in the flow. In the case of horizontal or near-horizontal pipe flows, typical flow patterns are stratified smooth, stratified-wavy, elongated-bubble, slug, annular, wavy-annular and dispersed-bubble. Some of these flow patterns are shown in Figure 1. Typical flow patterns for upward vertical pipe flow are bubble, slug, churn, annular and dispersed-bubble.

Figure 1: Two-Phase Flow Pattern Map (FluidFlow).

The fluid flows in different flow patterns such as slug, churn and annular. Some of these flow patterns can generate dynamic fluid forces which in turn may induce structural vibration. Excessive vibration may lead to component failures which must be avoided as they can produce significant economic, environmental and even human losses.

Therefore, the knowledge of flow conditions including flow pattern as well as flow-induced vibration (FIV) can have a significant impact on the proper design and operation of piping systems.

A Study of Choked Flow in Gas Piping Systems

The phenomenon of choked flow is often encountered in gas piping systems and tends to occur where there is a change in the flow path cross-sectional area. It can, therefore, occur in locations where we have orifice plates, valves, and fittings. It can however, also occur as flow exits a pipe into a vessel or to atmosphere.

Sonic choking is experienced in pipelines where the flowing gas reaches a velocity which is equal to the local sonic velocity of the gas at that given temperature and pressure. The gas, therefore, attempts to accelerate beyond the local sonic velocity and therefore becomes limited or choked.

If we consider air flow systems, choking typically occurs when the downstream absolute pressure (P2) is 52.8% of the upstream absolute pressure (P1), i.e. the absolute pressure ratio is 0.528.

Let’s initially consider a scenario where we have air flowing through a 10M long 2-inch pipeline with an inlet and outlet pressure of 6 & 5 bar a respectively. The air in the pipe stream must expand due to the pressure loss in the line. This results in a reduction in air density which tends to increase the fluid velocity and as such, the actual volumetric flow rate in the system. See Figure 1 below for details.

Screen Shot 2018-06-26 at 14.29.07

Figure 1: Air Flow System.

If we gradually decrease the downstream pressure we can reach the choke point for the system. Based on our rule of thumb, this should occur when the outlet pressure is in the region of approximately 3.168 bar a.  Let’s recalculate the system based on an arbitrary outlet pressure of 2.0 bar a.

In Figure 2, we can see the result of this design scenario. FluidFlow has automatically detected endpoint choking conditions in the pipeline.

Screen Shot 2018-06-26 at 14.48.25

Figure 2: Choked Flow with Outlet Pressure 2.0 bar a.

We can see from Figure 2 that the Mach number is 1.0 and the mass flow rate is 1.7146kg/s. Let’s see what happens when we reduce the outlet pressure even further to 1.0 bar a.

Screen Shot 2018-06-26 at 14.40.23Figure 3: Choked Flow with Outlet Pressure 1.0 bar a.

As we can see, the mass flow is unchanged as it has become choked in this system.

FluidFlow automatically detects choked flow and also plots the resistance, density, static pressure, static temperature & velocity curves for all pipes. Figure 4 provides an illustration of the density curve for the pipe in Figure 3.

Screen Shot 2018-06-26 at 14.33.56

Figure 4: Density vs Pipe Length Curve.

Figure 5 illustrates how FluidFlow detects choking conditions across a size-change component such as an orifice plate. In this case, we have an inlet and outlet pressure in a 10.0M 2 inch pipeline of 2.2 and 1 atm respectively. The orifice diameter is only 10mm. The pressure loss across the orifice plate is quite high at 1.2 bar. As a result, the air must accelerate to satisfy the continuity equations. In this case, the gas attempts to accelerate beyond the local sonic velocity causing choked flow.

Screen Shot 2018-06-26 at 14.35.16

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Why choose the FluidFlow Gas Module ?

FluidFlow does not assume gas ideality but solves for real gas conditions. The software will solve complex systems including gas mixtures (by mole or mass %). FluidFlow is provided with heat transfer functionality and a comprehensive database of fluids and components as standard.

Designing Compressed Air Systems

Inadequate or poorly designed compressed air distribution systems can lead to low productivity, poor air tool performance and perhaps more importantly, high energy bills. In order for a compressed air system to operate properly and cost effectively, it should be carefully designed to meet the needs of your applications. As part of the design process, there are six items which should be considered and factored into the final system design to give optimum results at maximum efficiency. The six items include demand, compressed air quality, supply, storage, distribution and control or management and all six must work together for the system to achieve top performance levels. The following discussion considers a number of these items.

1) Compressed Air Demand
To determine the demand for new systems, the operating pressure requirements and duty cycle of individual equipment should be considered. Compressed air consumers are rated by the manufacturer for optimum performance at a certain pressure and air flow rate. To design a system that delivers uniform pressure, it is necessary to ensure all tools and equipment work efficiently within a narrow pressure range. However, if this is not achievable, the system can be designed to operate at a higher pressure and user regulators to reduce the pressure as required. A booster can also be added to increase pressure for any specific applications requiring higher pressures. One final option available it to design two separate compressed air systems operating at different pressures. Leakage and artificial demand can often represent a significant amount of the system overall demand. All compressed air systems exhibit leaks and leakage can be measured in a number of ways while no pneumatic equipment is in use; measured using the loaded running time of a compressor, timing the pressure drop of the receive tank while all compressors are off or
measuring leakage at the point of use. Compressed air system demand can also include artificial demand caused by excess system pressure that doesn’t increase productivity. Artificial demand can be reduced significantly by installing a regulator at the point of use of a flow controller at the beginning of the distribution network.

2) Compressed Air Quality
The quality of compressed air is determined by measuring three main contaminants which include water vapor content as measured by pressure dew point temperature, oil content as measured by concentration and finally, solid particles as measured by the concentration of by their size. The level of contamination is influenced by the type of compressor, dryer, filtration and other related components. It’s no surprise then that the higher the air quality required, the more expensive the equipment. The International Standard document ISO-8573-1 provides a classification system for the main contaminants of a compressed air system and identifies how other contaminants can be identified in addition to the classification system.

Compressed air for power-tool usage, sand blasting, pneumatic pumps etc is a relatively low-grade quality where water, oil or solid particles in the air are more of an annoyance than a major concern.

Compressed air for instrumentation is a higher quality which is used in more sensitive areas where water or particulates could contribute to significant quality issues in the process. Paint spraying and powder coating are good examples of such a process where contaminants will ultimately affect final product quality. This classification of compressed air would be filtered for solids particles and oil and dried to a higher standard to that of compressed air for say, power-tools.

Process air is often used in food or drug production. Compressed air for use under these circumstances would need to be completely oil-free with almost negligible water vapor.

Higher quality compressed air is used in hospitals or for diving applications where the air must be of a quality suitable for safe breathing.

Compressed Air Supply

The compressed air supply must match the compressed air demand. If the supply, storage, or distribution system are not optimized. Excessive pressure fluctuations can occur resulting in increased operating costs and reduced productivity.

Many compressors are controlled by the line pressure. An increase in demand results in a drop in line pressure which is then rectified by an increase in the output of the compressor(s). A rise in line pressure, therefore, indicates a reduction in downstream demand which causes a reduction in compressor output. There are various forms of compressor control which can be utilized to manage such system operating conditions.

Single compressor installations are generally more suited to smaller compressed air installation where the compressor may often run at full capacity. Multiple compressors installations can offer many advantages such as; an ability to adjust to changing usage patterns, provide flexibility, floor space flexibility, can be centralised or de-centralised and of course, by their nature offer a backup facility in the event of plant failure.

4) Compressed Air Storage

The provision of sufficient storage is vital and represents available energy that can be utilized and replenished as required and at any time. The air receiver tank generally provides the bulk of the total storage capacity. In some cases, the compressor controls depend on storage to limit maximum cycling frequency when the demand is below 100% of supply. A correctly sized tank will prevent excessive cycling. 

Correctly sized receiver tanks can also provide sufficient storage capacity for any peaks in demand. During peaks demand periods, a poorly designed system can experience a drop in pressure as air in excess of system capacity is drawn from the system. As not all compressors in a multi-compressor system remain online at all times, the actual air supply at any time can be less than the total system capacity. During the time required to bring additional compressors online, the stored compressed air can be used to prevent any pressure drop in the system. The quantity of stored capacity needed is dependent on the amount of excess demand in cubic feet, available pressure differential between the compressor station and demand point, compressor start-up time as well as the time available to replenish the stored compressed air.

The installation of a flow controller downstream of the receiver tank is essential for providing additional compressed air when needed without downstream pressure fluctuations. The flow controller works life a precision regulator, increasing or reducing flow to maintain constant line pressure. It also provides the necessary pressure differential between the receiver tank and the system to create storage without changing system pressure downstream.

5) Compressed Air Distribution

An ideal distribution system provides a sufficient supply of compressed air to all demand points at the required pressure. Inadequate or poorly designed compressed air distribution systems can lead to low productivity, poor equipment performance and high energy bills. When designing a compressed air system, it is therefore good practice to consider the factors which help improve the efficiency and reliability of the compressors and ancillary equipment, minimise leakage and pressure drops and improve compressor life-cycle cost.

In general, there are three demands imposed on a compressed air distribution system; low-pressure drop between the compressor and most remote demand point, minimum leakage from the distribution pipework and efficient condensate separation if a compressed air dryer is not included in the system.

The compressed air pipe routing, design and dimensions are important factors for the efficiency, reliability and cost of compressed air production. Occasionally, a large pressure drop in the piping system is compensated by adjusting or increasing the working pressure of the compressor from for example, 7 bar to 8 bar. This results in inferior compressed air economy. Furthermore, when compressed air consumption is reduced, the pressure drop is also reduced and so the pressure at the demand point consequently rises above the allowed level. This therefore is not a recommended operating strategy.

In general, fixed compressed air distribution systems should be sized such that the pressure drop in the pipes does not exceed 0.1 bar (10,000 Pascals or 1.45 psi) between the compressor and most remote demand point. The pressure drop arising from flexible hoses, couplings and fittings should also be included in the pressure drop calculation. However, if possible, in an attempt to reduce pressure losses, the number of bends, valves, fittings or flow obstructions should be kept to a minimum.

A useful approach is to design a ring main or looped piping system to serve the space where the compressed air consumption will take place. This is an effective way to minimise pressure drop in a system. Branch pipe connections are then run from the loop to serve the various demand points. This helps provide for a uniform compressed air supply as the air distributed to the demand point from two directions.

The design of a ring main system is a recommended approach however, may not be entirely suitable in scenarios where there are large compressed air consumers located at a much greater distance from the compressor installation. It is recommended that a separate compressed air main should be routed under these circumstances to serve this equipment.

Compressed air piping system can generally be grouped into four main parts; risers, distribution pipes, service pipes and compressed air fittings. The risers transport the compressed air from the compressors to the consumption zones, the distribution pies split the air across the distribution zone and the service pipes route the air from the distribution pipes to the working areas/final demand points.

At design stage, the starting point is generally to obtain a diagram indicating the location of all compressed air equipment and an equipment list detailing all compressed air consumers. The compressor plant should be in a centralized location in close proximity to all relevant applications and processes in order to minimize the length of piping between the compressor station and the consumers. The consumers should ideally be grouped in a logical manner and supplied from the same pipe.

The location of the compressor must also give consideration to the quality of intake air, ideally clean, dry and cool. The compressors should be positioned clear of any potential steam, chemical vapors, engine exhaust and dust.

The pressure developed by the compressor plant can generally never by fully utilized as a result of pressure losses occurring in the system from pipework and fittings. When designing and sizing the various piping zones of the system, the following rule of thumb can be used for pipe losses.


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The length of the pipes required for the system can be scaled off the design plan and section drawings for the building/facility. The pipes can then be sized using relevant formulae and/or pipe sizing tables or nomograms. The flow rate, pressure, allowed pressure drop and length of pipe usually must be known in order to correctly size the pipework. When sizing the pipework, the equivalent length for each pipe section is estimated in an attempt to take into account the losses from fittings.  It is also often necessary to design and calculated the service pipes, distribution pipes and riser pipes separately, particularly for larger installations.

It is also worth remembering that, horizontal pipe installations should have a slope of around 1 to 2 % toward the point of consumption to aid the transfer of condensate to the predetermined drain points. Note, some would argue that the need to slope the pipework is no longer required when a correctly sized air dryer is installed however, the cost is minimal/negligible and offers added protection in the event of an air dryer failure.

The use of an appropriate software tool helps overcome the difficulties with designing these systems by hand. Using tables, nomograms etc can be time-consuming and occasionally, error prone. Applying the equivalent length approach in a hand calculation also produces a lower level of accuracy by comparison to a software-based solution. FluidFlow helps with the design of such systems and allows the designer to consider dynamic type operating scenarios with fluctuating demand etc.


Compressed Air FF

Figure 1: Compressed Air System Model.

Compressed Air System Performance

Tests and experiments have shown that for every 1 bar in pressure drop across the system, the resulting power consumption is increased by circa 6 – 7 % (plus the additional cost of unregulated users). Accordingly, the distribution network can have a profound influence on the system performance. There are a number of steps which can be taken to optimise the air distribution system.

1. Pressure Drops and Line Improvements.

As outlined previously, a ring main is generally the most efficient type of distribution layout. Air mains are usually sized on velocity and a velocity level of 6 to 9 m/s is common as this is sufficiently low to prevent excessive pressure drop and should also allow reasonable water separation. The local feeding mains can flow up to around 15 m/s. However, in order to prevent adverse pressure drops, the flow velocity in the main header sections should not exceed circa. 6 m/s.

It is best to replace any tee connections for directional angle entry connections or swept tees. Turbulence caused by a 90o tee connection can cause pressure drops resulting in back pressure sending a false “unload signal” to the compressors which can potentially cause excessive cycling of the compressor.

Incorrect pipe sizing and restrictions are a major source of pressure losses in the system. Losses in the interconnecting distribution pipework between the compressor and the header distribution piping are commonplace however, the losses along these lines should be kept to a minimum.

Interconnecting piping between compressors or systems often require close attention. It is vital that they are carefully designed to avoid sending back any false signals to the compressor.

2. Leak Detection

Leaks in compressed air systems are a regular feature. The energy requirements served by compressed air systems are intermittent in nature, however leaks are constant and surprisingly, potentially significant. The monetary cost of leaks can be quite startling and perhaps a little eye watering.

For instance, one 4mm hole in a compressed air distribution pipe can cost €2,005 per annum on a typical compressed air system operating throughout the year and at 8 bar.

In addition to the monetary cost, leaks can cause significant pressure drops resulting in excessive compressor cycling. In an attempt to reduce the pressure loss in a system where excessive leakage is an issue, operators occasionally increase the system discharge pressure. However, this has the effect of exacerbating the problem by increasing the leakage rate and create more leaks in the future.

It is not uncommon for leakage rates to be around 20 – 30%. Leaks can occur at any point in the system with joints, drains, valves regulators etc being the most common sources. Fixing leaks in the most basic form can occasionally be as simple as tightening connections or applying a sealant at a strategic point. However, leaks will be found which require replacement of faulty components. It is worth noting that one of the most effective means of reducing leakage is to reduce the distribution pressure. Note, a 10% reduction in leakage would often be achieved through carrying out an appropriate leak reduction programme. 

3. Dedicated Air Receivers for End Users.

In industrial applications where air pressure is subject to large fluctuations or variations, air receivers are beneficial. In these situations, the increased compressed air requirement is compensated by air from local air receivers thus minimising idling at the generation station.  The receivers subsequently replenish slowly using control valves to minimise peak energy demand on the compressor station. In addition to reduced compressor cycling, air receivers provide protection for end users that require high pressure by minimising the system pressure drop off while supporting the speed of transmission response in supply.

Other useful tips with regard to compressed air system design include:

  • Position one air receiver near the compressor to provide a steady source of control air, additional air cooling and moisture separation. A large storage receiver can be located downstream of the dryer and filters to act as a buffer for demand surges and controlled by a flow controller.
  • Compressors should be positioned in a clean, dry, cool and well-ventilated room. Ensure there is sufficient room around the compressors and equipment for proper air flow. Manufacturer’s will often specify the minimum spacing required around compressors.
  • Piping in a loop is recommended will all piping sloped to accessible drain points. Air outlets should be taken from the top of the main lines to prevent moisture entering the outlet.
  • Under average conditions, every 100 cfm of air compressed to 100 psig (6.89 barg) produces approximately 20 gallons (75.7 litres) of condensate per day which needs to be treated.
  • The minimum amount of storage recommended is one gallon per cfm of capacity. This should be increased to 4 to 10 gallons per cfm of capacity for systems with sharp changes in demand.


  1. Chemical Engineering, March 2018.
  2. SEAI – Compressed Air Technical Guide.
  3. Atlas Copco Compressed Air Manual.
  4. Designing Your Compressed Air System – Kaeser Compressors.

FluidFlow v3.44

General Release summary: This release provides speed improvements at start up and across a network. There are over 250 new components, centrifugal and positive displacement pumps, controllers, fluids and pipes added to the databases. This release also contains bug fixes.

General Release info:


* Hundreds of new database entries. Over 180 centrifugal pump models added from KSB, Warman etc. Over 70 positive displacement pump models from SPXFLOW & Fristam added.

* Steel pipes database has been expanded to include additional B36.10M sizes and classifications for existing and new steel pipe sizes.

* 7 new control valves and new fluids including d-limonene, thionyl chloride, AVCAT, AVGAS, and Duratherm XLT-120

* Additions and corrections to help file.


* Velocity Pressures now shown without gauge or absolute unit reference.


* Supercritical -> Liquid and Supercritical -> Gas phase change via a fixed temperature change at any element sometimes predicts incorrect exit phase and incorrect heat loss for a fixed temperature change.

* Now allowable to have viscosity AND affinity law correction together. Was a bug that prevented the 2 effects from occurring together.

* Tee junctions with a fluid change caused instability and non-convergence. Now fixed.

* Intermittent bug in print preview that sometimes caused an error is fixed.

* Fixed head unit inconsistency in EGL and HGL charts.

* For old files containing Liquid Relief valves, the Kd was sometimes not updated for an autosize calculation.

* Fixed scripting error on start-up when Scripting module is not available.

* Added check to stop file being saved to a non-existent folder.

Joule-Thomson Effect (J-T)

To obtain maximum accuracy in gas calculations, particularly in systems where large pressure losses occur, the Joule-Thomson effect needs to be included in the solution. The J-T effect describes the increase or decrease in temperature of a real gas (as opposed to an ideal gas) or liquid as it flows from an area of high pressure to an area of low pressure without heat transfer to or from the fluid and no external mechanical work extracted from the fluid. An example of such a scenario would be a gas flowing through a perfectly thermally insulated throttling device such as a valve, nozzle or orifice plate.

The temperature change experienced during a Joule-Thomson expansion is quantified by the Joule-Thomson coefficient.

Joule-Thomson Coefficient

The temperature change arising from a reduction in pressure at constant enthalpy in the Joule-Thomson process can be expressed as:

μJT = (∂T/∂P)H

The value of μJT (oC/bar) depends on the gas type as well as its pressure and temperature before expansion. This coefficient may be either positive (cooling) or negative (heating). For all real gases, the value of μJT will equal zero at some point called the inversion point and the J-T inversion temperature is the temperature where the coefficient changes sign (where the coefficient equals zero).

The inversion temperature represents the temperature of a gas at which a reduction in pressure causes no temperature change. Above this temperature, the gas heats on expansion and below this temperature, the gas cools on expansion. When a real gas expands through a throttling device, the gas temperature may either increase or decrease depending on the initial gas pressure and temperature. Figure 1 provides an illustration of a typical temperature inversion curve.

Figure 1

Figure 1: Temperature inversion Curve.

In gas expansion, the gas pressure reduces and thus the sign of ∂P is always negative. The following table explains when the Joule-Thomson effect heats or cools a real gas.

Screen Shot 2018-05-30 at 14.11.04

It is worth noting at this point that, there is no Joule-Thomson effect for an ideal gas. The value of μJT, therefore, equals zero, i.e., in theory, ideal gases neither heat or cool during expansion at constant enthalpy.

FluidFlow takes account of the Joule-Thomson coefficient when solving gas flow systems using real gas conditions. Figure 2 details the calculation of Joule-Thomson coefficient for a pressure control valve in a pipeline transporting a mixture of n-butane and n-pentane at a pressure of 11bar and temperature of 390K.

Figure 2 J-T PRV

Figure 2: FluidFlow J-T Coefficient Calculation.



Pressure Drop in Natural Gas Pipelines

A previous blog post considered the applicability or the ideal gas equations for gas flow modelling and the potential pitfalls of using such a solution approach. This post considers a number of simple natural gas pipe flow systems which have been solved using FluidFlow. FluidFlow software doesn’t make simplifying assumptions and solves for real gas flow conditions using an equation of state.

Figure 1 provides an illustration of a horizontal natural gas pipe line consisting of 50 miles of 26 in OD pipe having an ID of 25.44 in. The temperature of the flowing gas is 60 F. The upstream and downstream pressures are 825 and 565 psia respectively.

horizontal natural gas pipe line

Figure 1: Modeling Horizontal Flow of Natural Gas.

The solution in Figure 1 shows good agreement with published literature (1) where Q was estimated by way of a hand-calculation to be 462 MMft3/day at 14.65 psia (101008 Pa) and 60 F (15.5 oC). FluidFlow has established a flow rate of 458. 7 MMft3/day. The difference in flow can be attributed to the difference in density of the flowing gas and reference volumetric flow conditions, i.e. a 101325 Pa and 15 oC.

Figure 2 provides an overview of a gas-distillate well pressure model where the inlet pressure has been calculated. The vertical pipe length is 5790 ft with a corresponding well depth of 5790 ft. The pipe ID is 1.995 in with an inlet flow rate of 5.153 MMft3/day (STP) and temperature of 83 F.

Figure 2 gas-distillate example page 323

Figure 2: Gas-Distillate Well Pressure Model.

Published literature (1) makes reference to a hand-calculated inlet pressure of 2,556 psia. FluidFlow has established an inlet pressure of 2,526 psia. A plot of the pressure traverse for this vertical system has also been provided (Figure 3).


Figure 3

Figure 3: Pressure Traverse of Flowing Gas Well.

To make a comparison, two further test case scenarios were created in FluidFlow, one with a well depth of 1,000 ft and one at 3,000 ft. Based on the graph plot, the corresponding pressure is approximately 2,195 psia and 2,340 psia. The models for the two cases are outlined in Figure 4.


Figure 4 vertical pipes

Figure 4: Gas-Distillate Well Pressure Model (Well Depths of 1,000 ft & 3,000 ft).

The FluidFlow solution has established calculated inlet pressures of 2193.7 and 2334.7 psia respectively. Again, the solution produces good agreement with published literature.

Figure 5 provides an illustration of a natural gas distribution system as modeled using FluidFlow. The system consists of 77,103 ft of pipework ranging from 2 to 18 inches in diameter. The initial condition is 50 barg and 100 oC and all outlets are set to 140 barg. Four compressor stations have been modeled in the system generating a total flow rate of 2,000,000 kg/h. The analysis of this system included optimisation of the plant and an evaluation of the effect of pipe roughness after many years of use.

Figure 5 natural gas distribution system

Figure 5: Natural Gas Distribution System.


The Calculation of Pressure drop in the flow of natural Gas through a pipe – Fred H. Poettmann.

To read our previous blog post on Compressible Gas Flow In Pipes; please click here –

Compressible Gas Flow in Pipelines

Many equations used for the solution of gas flow in pipelines do not take into consideration rigorously the deviation of natural gas from ideal gas behaviour. For low-pressure gas flow, the error is considered low, however, for high pressures, high flow rates, low temperatures or the closer the gas is to change in phase state, this error can be significant. This deviation from ideality must be taken into consideration. It is therefore important when modeling gas pipe flow systems to use a
software tool which doesn’t make the simplifying assumptions of gas ideality.

FluidFlow does not make these simplifying assumptions but solves for real gas conditions using an equation of state. The simplified application of the ideal gas law (Pv = nRT) makes several assumptions as outlined below;

– Particles in an ideal gas are in constant, random, straight-line motion.
– The volume occupied by the particles is negligible relative to the volume of the enclosure.
– The collisions between particles are elastic with no loss of kinetic energy.
– There are no intermolecular forces acting between the particles in an ideal gas.

An ideal gas assumes a large number of point particles colliding elastically. It neglects any short-range intermolecular forces resulting from repulsion or attraction due to molecular charges and the fact that molecules have a finite volume, i.e. are not infinitely small! This means a real gas is not infinitely compressible whereas an ideal gas has no such limitations.

Although the ideal gas law may be a useful simplified description of gases, all real gases fail to obey this relationship to some degree. The extent to which a real gas departs from ideal behaviour can be seen by rearranging the ideal-gas equation to solve for n (PV/RT = n). This form of the equation tells us that for 1 mol of an ideal gas (n=1), the quantity PV/RT equals 1 at all pressures. However, under the conditions of real gas flow outlined earlier, a product of PV divided by RT is no longer equal to one.

In the graph shown in Figure 1, we can see the deviation of nitrogen gas from ideal gas behaviour. On the y-axis, we have the product of PV/RT. On the x-axis, were have pressure. The dashed blue line shows the behaviour of an ideal gas for which the product of PV/RT is equal to one under any conditions. The orange, grey and yellow dashed lines show the deviation of nitrogen gas from ideal gas behaviour at different temperatures and pressures. These dashed lines show a considerable deviation of ideal gas behaviour. Notice how the conditions that produce this curve are high pressure and low temperature.

Figure 1 nitrogen real gas pv rt graph

Figure 1: Effect of Temperature & Pressure on the Behaviour of Nitrogen Gas.

As high pressures, the deviation from ideal behaviour is considerable and is different for each gas. Real gases, therefore, do not behave ideally at high pressures.

It is noted in many literature sources that the ideal gas equations can be used with some degree of accuracy under certain specific conditions, i.e. at low pressures. Caution should, however, be exercised when applying gas ideality and a thorough check of system operating conditions to ensure the design solution falls within the criteria.

FluidFlow doesn’t make the simplifying and often error-prone assumption of applying gas ideality but solves for real gas conditions using an equation of state. The software, therefore, takes into account the gas compressibility factor (Z) and solves gas flow systems using a marching algorithm for incremental pipe length. This ensures a much higher level of solution accuracy.

The solution of gas flow systems becomes more complex when dealing with gas mixtures and large systems. It is recommended that a suitable software tool should is selected to design and solve gas flow systems in an attempt to eliminate the potential for errors resulting from the simplification of ideal gas flow.

In conclusion, let’s ponder the following quote:

“Perfect gases are like perfect people: They do not exist.”

District Cooling Systems

District Cooling Systems

District Cooling Systems – Many countries experience hot summers which impose a need for air conditioning in order to achieve and maintain a comfortable indoor environment. Air conditioning has traditionally been provided to buildings by electrically powered air conditioning units however, this equipment can have a high power demand. Many cities across the world are exploring alternative and more energy efficient cooling solutions.

One such solution is a District Cooling System (DCS) which involves the production and distribution of chilled water to multiple buildings through a network of insulated underground pipelines. Each building connected to this central system is connected by an Energy Transfer Station (ETS). The ETS typically uses a heat exchanger to chill down the water of the building using the district mains. This chilled water within the secondary building system is then used for space comfort cooling and process cooling. This essentially means that there is no requirement for chilled water plant to be installed locally in each building. Each building operator therefore purchases chilled water from the central provider.

District Cooling Systems include central chiller plant, pumps, valves & distribution pipework. Chilled water is often generated at the central plant station by dedicated chiller units. Free cooling can also be achieved by using water from nearby rivers, deep lakes etc. There are many examples of such systems such as in Paris where water from the river Seine is refrigerated to around 5oC and transported 71km to serve circa. 5 million square meters of offices, hotels, theatres, government buildings and of course, the Louvre. Another notable example case is one of the world’s largest district-cooling system schemes at The Pearl-Qatar.

Figure 1 provides an example of a District Cooling System as designed and modeled using FluidFlow software. This system uses two “central” cooling plant locations to provide chilled water which is then distributed through a network of 17.8 km pipework to serve 48 Energy Transfer Station’s.

District Cooling Systems


Figure 1: 505MW District Cooling System.

District Cooling Systems have been quoted as offering a more energy-efficient approach to achieving cooling as they can reportedly consume 20-35% less electricity in comparison to Individual water-cooled air conditioning systems and traditional air-cooled air conditioning systems respectively.

Other benefits of such systems include; lower CO2 emissions, reduced pollution, air-conditioning provided from a sustainable source, reduced maintenance costs for users, lower operating costs for users, increased reliability and increasing the potential for economics of scale. 

Join us on 14th March 2018 for a 30 minute webinar discussing modelling of closed loop piping systems using FluidFlow here

Fire Sprinkler Systems Technical Paper

Fire sprinkler systems can be found in a wide range of buildings or structures from warehouses to multi-story buildings, ships, oil rigs etc. Sprinkler systems are used to provide a reliable method of automatically detecting and extinguishing fires at an early stage. Many countries or regions have their own specific set of standards for sprinkler system design and installation. It is therefore always necessary to establish at the project outset, what specific set of regulations or standards must be adhered to when designing a sprinkler system. This is necessary in order to meet the approval of the local authority and indeed any specific insurance requirements. Typical standards used over the years include those issued by the Loss Prevention Council (LPC), British Standards Institution, Factory Mutual Insurance and of course & NFPA (National Fire Protection Association). NFPA Standard 13 serves as a benchmark for the design and installation of fire sprinkler systems.

The function of a building or type of work carried out within a building determines the fire risk and the design criteria for a sprinkler system. It is therefore imperative to discuss the design criteria for any given building with the local authority and buildings insurers prior to commencing design.

When designing sprinkler systems for fixed building structures, the local water authority should also be consulted at the preliminary design stage to obtain the relevant information regarding the suitability of the local water supply for supplying the system. In most cases, the local water main cannot be used to feed the sprinkler installation directly and as such, a fixed storage tank and pumping station may be required on site. The pumping station typically is made up of a duty and standby pump arrangement. Either pump must be capable of providing the performance characteristics required to suit the design system hydraulic calculations.

In the event of a fire in a building, once the sprinkler system has detected the fire, the water must be discharged at the required flow rate and spray pattern over the area of protection. The amount of water discharged from a sprinkler can be expressed using the following formula:

Q = K√P


Q is the water discharged in litre/minute.

K is a constant for the sprinkler.

P is the pressure at the sprinkler in bar.

The shape of the sprinkler deflector determines the spray pattern, water droplet size and area of coverage. There are a wide range of sprinklers designed for different purposes. It is therefore essential that the correct sprinkler is selected for the application under consideration.

As mentioned earlier, the function of a building determines the type of sprinkler system required including the spacing of the sprinkler heads required to ensure adequate coverage of the system in the event of fire conditions.

The spacing of sprinkler heads is largely dependent on the hazard type or hazard classification within the building. NFPA Standard 13 sets out spacing requirements which requires careful consideration when designing any sprinkler system.

The NFPA standard provides an example case of a large warehouse facility measuring approximately 200 ft x 130 ft. The building has approximately 200 sprinkler heads and the example case focused on 12 remote sprinklers in operation (see sprinklers within the blue dashed line in Figure 1). The system operates with a design flow rate of 259.6 GPM.

An outline plan and elevation of the building is shown in Figure 1.


Figure 1: Building Plan & Elevation.

Based on this operating condition, the system pipe friction loss calculations have been performed using the Hazen-Williams formula, as follows;image


P is the frictional resistance of the pipe (psi/ft).

Q is the flow in GPM.

C is the friction loss coefficient.

d is the internal diameter of the pipe in inches.

The Hazen-Williams Coefficient used in this worked example is C = 120 with the exception of the underground pipework which has a C = 150. The solution provided within the documentation utilises the equivalent pipe length method to estimate the loss associated with fittings in the system.

A key part of the completion of design calculations for any sprinkler system is the definition of the sprinkler coefficient, K value. This example utilizes a K factor of 5.6 for each sprinkler.

As outlined earlier, the example case focuses on 12 remote sprinklers in operation. The system was modeled in FluidFlow generating results which are a close match to those in the NFPA hand calculation. We can, however, expect differences between the two cases which can be attributed to the simplifying assumptions of using the equivalent pipe length method to determine the pressure loss across fittings. Note, FluidFlow doesn’t make this assumption and calculates the pressure loss according to the fitting characteristics. If for a specific design you have a requirement to use the equivalent pipe length approach, this method is also available in FluidFlow.


Figure 2: NFPA Sprinkler System Model.

The hand calculation notes a flow rate of 19.5 GPM serving the most remote nozzle whereas FluidFlow has determined a flow rate of 19.7 GPM. The static pressure calculated by the software is 12.1 psig which is an exact match of the hand calculation. The results for the branch pipes are noted as 40.2, 62.1 GPM in the hand calculation which again, compare well to the software result of GPM 40.2 & 61.3 respectively.

The minimum operating pressure for each sprinkler has been noted as 7 psi (0.5 bar). This minimum setpoint can be defined in FluidFlow and as such, any sprinklers which operate below this minimum pressure will be highlighted and clearly identifiable.

Sprinkler systems can be solved for any number or sequence of sprinklers activated at any given time. This solution can also be automated using the Scripting Module (Dynamic Analysis).

Using the pre-installed script option, it was established that this entire system has circa. 2421 ft of pipework with a total pipe volume of around 1.03m3.

The next example case is for a yarn processing facility (Figure 3). The design system flow rate, in this case, is 1639.9 l/min. The design is based on STD steel piping with a Hazen Williams pipe coefficient of 120. Each sprinkler head has an associated K factor of 115.



Figure 3: Yarn Facility Sprinkler System Model.

This system was solved with just 15 remote sprinkler nozzles in operation. As can be seen in Figure 3, all other nozzles are turned off. The flow rate in the most remote nozzle is calculated as 81.5 l/min and the associated pressure at the system inlet based on the defined design flow rate is 5.5 barg.


  1. NFPA 13 Standard for the Installation of Sprinkler Systems.
  2. Building Services Pocket Book (Second Edition) J Knight & P Jones.

Control Valve Sizing

Control Valve Sizing. The successful operation of process plant and systems involves the highest level of measurement and control performance. Control valves therefore perform a key function in this process. The performance of control valves can have a dramatic effect on plant operating efficiency, overall profitability and asset life cycle costs.

A correctly sized control valve can provide significant quantifiable savings as well as increase process availability, reduce process variability and reduce maintenance costs. Correctly sized control valves also last longer in comparison to unmatched or incorrectly sized valves. Careful consideration should therefore be given to the correct sizing of system control valves.

The Importance of Sizing

Control valve sizing procedures are based on accepted mathematical methods such as those detailed in ISA-75.01.01-2007 – Flow Equations for Sizing Control Valves. These methods can be used to develop accurate valve sizes however, during the course of plant design, control valves are often sized based on a future maximum design process load plus a safety factor. This can result in the specification, procurement and maintenance of a larger valve than necessary thus producing imprecise control, poor production performance in addition to the operational issues discussed above.

When sizing a control valve, the most common approach is to calculate the flow coefficient, Cv which is a measure of the capacity of the valve body and trim. A valve Cv can be described as the number of gallons per minute (GPM) at 60oF (15.5oC) that will pass through a valve with an associated pressure drop of 1 psi. In simple terms, a fully open control valve with a Cv of 14 usgpm/psi passes 14 usgpm of fluid with 1 psi pressure drop. The flow coefficient is discussed in further detail below.

Flow Coefficient (Cv)

The flow coefficient or valve coefficient denoted by “Cv”, is used to determine the valve size that will best allow the valve to pass the required flow rate while providing stable control of the process fluid. Most valve manufacturers publish Cv data in product catalogues for various valve styles. Table 1 provides an example of an equal percentage valve curve characteristic as published by Crane Engineering.

Table 1 Crane curve data points

 Table 1: Equal Percentage Characteristic Curve Co-ordinates

The performance curve for the tabulated co-ordinates is shown in Figure 1.

Figure 1 crane curve

Figure 1: Equal Percentage Characteristic Curve

Crane Engineering noted a value of 50% capacity at 82.3% travel which is denoted by the dashed line in the capacity curve. 

The tabulated coordinates for this performance curve have been defined in the FluidFlow control valve database. The control valve performance curve is automatically generated by the software and can be viewed in Figure 2.

Figure 2 crane equal curve DB

Figure 2: Control Valve Performance Curve (FluidFlow)

In an attempt to make a comparison with the Crane Engineering data, a test-case piping system was created in FluidFlow piping system software (Figure 3).

Figure 3 crane eq model

Figure 3: Crane Equal Percentage Control Valve Modeled in FluidFlow Software.

The control valve has been modeled in the above liquid flow system whereby the duty point is near the Crane Engineering case, 50% capacity at 82.3% travel. As shown in Figure 3, the calculated FluidFlow result matches exactly with the Crane Engineering example.

If the Cv for a control valve is not calculated correctly or accurately, the resultant selected valve will experience diminished performance. If the Cv is too small for the process, the valve itself or the trim inside the valve will be undersized resulting in the system being “starved” of the process fluid. Undersized valves exhibit a higher pressure drop across the valve to maintain adequate flow and exhibit limited flow capacity. Furthermore, since the restriction in the valve can cause a build-up of upstream pressure, higher back pressures created before the valve can lead to damage in upstream pumps or other upstream equipment. Although relatively uncommon, an undersized control valve generally cannot deliver sufficient flow rates under maximum load conditions.

If the Cv calculated is too high for the system requirements, the result is a larger oversized control valve is usually selected. In addition to the more obvious issues with an oversized valve such as a larger cost, weight and size, when in throttling service, significant control instability can occur. Usually the closure element, such as the valve plug or disk, is positioned just off the valve seat which leads to higher pressure drops across the valve and higher fluid velocities which can cause cavitation, flashing or erosion of the valve trim elements.

Cavitation can occur in liquid systems when high velocity reduces the static pressure inside the valve to below the pressure level at which the liquid stats to boil and produce vapor bubbles. These vapor bubbles collapse whenever the downstream pressure is higher than the vapor pressure causing high pressure waves. These implosions result in very high noise levels and can cause considerable damage to the valve body or trim parts under prolonged service.

Oversized control valves are quite common and are highly sensitive to operating conditions with even the smallest of adjustments in valve position causing significant changes in fluid flow rate. It is therefore extremely difficult to achieve the exact flow rate required under these conditions. Oversizing of control valves can also have a domino effect. Safety relief valves must be sized to match the capacity of the control valve. Within bypass configurations, isolation valves, bypass valves and drain valves must all be larger which can impact on the size of the piping and associated structural supports.

The most common and basic form of the equation for liquid applications is;



Cv is the Flow Coefficient which describes how much fluid will flow (GPM) through the valve for a given pressure drop of 1 psi.

Q is the Flow Rate (GPM).

P1 is the Upstream Pressure (psi).

P2 is the Downstream Pressure (psi).

SG is the liquid specific gravity.

The ISA-75.01.01-2007 standard includes equations for predicting the flow coefficient of compressible and incompressible fluids through control valves. The standard uses the following flow sizing equation for Newtonian liquid flow;

C = Q/N1 * √ (ρ10) / (∆P)


C is the Flow Coefficient which describes how much fluid will flow through the valve for a given pressure drop.

Q is the Flow Rate (m3/h, GPM, SCFH).

∆P is the differential pressure between upstream and downstream pressure taps (kPa, bar, psi).

ρ10 is the relative density.

N1 is the numerical units constant (Refer to Table 1 in ISA-75.01.01-2007 standard).

These equations are widely accepted for sizing valves in liquid flow systems.

Note, the equations detailed in the ISA-75.01.01-2007 standard enable the control valve size to be determined in absence of any connected piping and fittings. An even better solution is to apply these equations whilst also including the interactions of the pipework and fittings in a full connected piping system. This will provide a much more accurate solution and provide a baseline for correct valve selection.

Control Valve Sizing – Rules of Thumb

When sizing control valves, a general rule of thumb noted in many engineering publications is to size the valve such that it operates between 20 to 80% open at maximum required flow rate. It is also recommended to have the minimum opening no less than 20% to provide a safety margin at the minimum flow rate required. This approach ensures that as much of the valves control range as possible is used while maintaining a reasonable (but not excessive) safety margin.

As a guide, correctly sized globe valves are typically one size smaller than the line size and properly sized butterfly, full-ball and segment-ball valves are typically two sizes smaller than the line size. Note, this statement should be used as an indicative guide only and not as design criteria or a design rule. Furthermore, this guide only serves as a useful rule of thumb if the connected lines have been sized accurately and correctly.

 Control Valve Sizing Example

The ISA-75.01.01-2007 standard provides an example liquid case which is summarised as follows;

Fluid: Water.

Inlet Temperature: 363K.

Fluid Density: 965.4 kg/m3.

Inlet Absolute Pressure: 680 kPa.

Outlet Absolute Pressure: 220 kPa.

Flow Rate (Q): 360 m3/h.

Pipe Size: 150mm.


C = Q/N1 * √ (ρ10) / (∆P)

C = 165 m3/h/bar for Kv

Modeling this scenario yields the following results.

Figure 4 ISA Liquid Example1

Figure 4: ISA Liquid Control Valve Sizing Example.

The control valve in this example has been automatically sized using FluidFlow software based on the design conditions presented. The modeled case provides a C or Kv of 165 m3/h/bar which matches the ISA-75.01.01-2007 example.

Energy Saving Opportunities

Control valve inefficiencies in plant processes offer opportunities for energy savings and reduced maintenance costs. Valves which consume a large fraction of the total pressure drop for the system or are excessively throttled can present considerable opportunities for energy savings. Pressure drops in liquid systems increase the energy requirements of these systems. Pressure drops are caused by resistance or friction in piping and in pipe bends, joints etc as well as throttled control valves. The power required to overcome the pressure drop is proportional to both the fluid flow rate and magnitude of the pressure drop. If the valve is oversized, the valve will be throttled excessively and also, if the valve is undersized, the pressure drop will be unnecessarily excessive which of course, can increase pump energy requirements significantly. Control valves should therefore be carefully sized and selected.

There is also scope for energy savings in existing systems featuring control valves. When reviewing design control valve sizes, it is important to understand the causes of any valve sizing errors. Emerson research previously identified several major contributing factors to valve sizing errors. These include selecting line-size valves, out of date process data resulting from changes in process conditions or operating conditions which differ to the original system design. To correct systems with incorrect valve sizes, its essential to obtain accurate process data at all expected operating conditions. The control valve can then be sized correctly.


The process of “control valve sizing” is a procedure where the dynamics of the system are matched to the performance characteristics of the valve. This produces a control valve of an appropriate size and type that will best meet the needs of managing flow within the process system. Every attempt should be made to carefully and accurately size and select a control valve for the required application. The topic of control valve selection will be discussed in a separate blog.

This discussion attempts to outline only the basics of control valve sizing. Further detailed reading is recommended when completing control valve sizing for liquid or even gas flow systems along with consideration of other design factors which should be considered as part of the sizing process such as Liquid Pressure Recovery Factor (FL), etc.

Note, FluidFlow solves the various forms of the equations described in ISA-75.01.01-2007, taking into account Liquid Pressure Recovery Factor (FL) etc.



An Insider’s Guide to Valve Sizing & Selection by Jon F. Monsen.

ISA-75.01.01-2007 Flow Equations for Sizing Control Valves.

Valve Handbook, Philip L. Skousen.

The Definitive Guide to Control Valves, Crane Engineering.

Closed Loop Systems

Join us on 14th March 2018 for a 30 minute webinar discussing modelling of closed loop piping systems using FluidFlow here

In general, there are two main types of systems in which pumps can be installed, open-loop and closed loop systems. Open-loop systems are circuits in which the pumped fluid is exposed to the local atmosphere at some point in the circuit. A typical open-loop system would be a cooling tower system where the flowpath is linear, i.e. fluid transfer between two vessels. Conversely, closed loop systems as the title suggests are closed piping circuits where the pumped fluid circulates in a closed loop without any exposure to the local environment and typically without transfer of fluid into or out of the closed loop. Examples of closed loop systems include hot oil circuits, cooling/chilled water systems, hot water heating and air conditioning systems. Figure 1 provides an illustration of a closed loop fresh water cooling system consisting of heat exchangers, circulating pumps, orifice plates and over 300 M of pipework.

Figure 1

Figure 1: Closed Loop Fresh Water Cooling

One of the unique aspects of closed loop piping systems is that the static elevation is not accounted for in head-pressure calculations as these systems are largely unaffected by static pressure. However, just like open flow systems, we still need to make checks for sufficient NPSHa, that the static pressure throughout the system does not fall below the fluid vapor pressure and therefore induce cavitation etc. Any pump selected for the closed loop system, must be able to transport the fluid to the highest point without flashing or inducing vacuum and the lowest point must also be evaluated for pump shut-off pressure. A closed loop circuit will exhibit only friction losses. Pumps operating in closed loop systems are therefore only required to overcome dynamic friction losses.

Let’s consider the operating conditions of a closed loop system with an elevation change of say, 10.0 M. The system circulating pump is required to transport the fluid from the bottom of the system (0 M) to the top at 10.0 M. It would initially appear that the pump must overcome the height difference of 10.0M however, due to gravitational effects, this isn’t the case as for every metre of fluid pumped vertically upwards, a corresponding 1.0 M of fluid drops on the return side of the system.

When the system is motionless, i.e. the fluid is not being circulated, the pumps suction and discharge exhibit the same pressure exerted by the two separate 10.0 M columns of fluid which are connected at the top. An alternative way of considering this is, the suction pressure available at the pump is equal to the discharge pressure required to move the fluid to the top of the system.

Regardless of where the pump is positioned in the loop, the differential head developed by the pump will always be the same.

As you would expect, the head required to maintain flow in a closed loop system decreases as flow decreases and becomes greater as flow increases. In well-designed systems, the friction losses will decrease proportionally with a decrease in flow rate.

Large scale closed loop systems can become complex to manually design as they often have many multiple branch lines or sub-loops.

When selecting a constant speed centrifugal pump for closed loop systems, the best efficiency point (BEP) on the pump efficiency curve should fall between the design minimum and maximum flow points on the pump capacity curve. This ensures the pump operating efficiency is maximised across the expected pump operating conditions.

The performance curves shown in Figure 2 show that, the centrifugal pump achieves the maximum flow condition of 1400 m3/h at 55.0 M TDH and rises just 5.0 M at the minimum design flow rate of 700 m3/h. Note, the flow experienced in this sample system at any point in time is based on the demand of the system.

Figure 2

Figure 2: Closed Loop – Constant Speed Pump

Note, a “flat” pump capacity curve profile is preferred for closed loop systems with variable flow due to the energy savings which can be achieved at lower flow conditions.

Although choosing a centrifugal pump with a “flat” capacity curve profile offers potential energy savings, much higher energy savings can be achieved by selecting a suitable VFD controlled pump (Figure 3). Since the system resistance curve reduces steadily from maximum to minimum flow (in this case 1400 to 700 m3/h), frequency control can be used to achieve the desired operating condition at low flow and in doing so, achieve much higher savings in comparison to that of a constant speed pump using valve throttling to adjust or regulate flow.


Figure 3: Closed-Loop – VFD Controlled Pump

The reason a much higher reduction in power is achieved by using the VFD pump is, a significant reduction in pump operating speed is achieved and secondly, the pump BEP efficiency isomer closely follows the system curve. As such, at 700 m3/h, the control frequency at the lower operating speed has almost the same efficiency as it does when the pump is operating at the speed required to achieve maximum flow.

If you are considering a VFD controlled centrifugal pump for an application, select the highest efficiency pump with a BEP that falls at or just to the left of the maximum flow condition. Keep head rise to minimum flow as low as the application will allow and test your pump selection using an appropriate software tool. This will allow you to evaluate the pump performance at different operating conditions.


The Practical Pumping Handbook, Ross Mackay.

Pump Selection for VFD Operation, Joe Evans.


Maximizing Steam Plant Performance Technical Paper

 What are the different types of steam?

When water is heated beyond its boiling point it vaporizes into the gaseous phase or as we more commonly know it, steam. The properties of steam vary considerably and are greatly dependent on temperature and pressure.

Before we consider why we use steam, lets firstly consider the different types of steam.

Saturated Steam (Dry Steam): This form of steam occurs when water undergoes sensible heating to raise the fluid to its boiling point and then vaporizes with additional latent heating. Heating this steam further, i.e. above the saturation point, produces superheated steam.

Superheated Steam: Steam in this form occurs when saturated or unsaturated steam is heated beyond the saturated steam point. The result is steam at a higher temperature and lower density by comparison to saturated steam at the same pressure.

Unsaturated Steam (Wet Steam): This is probably the most common form of steam which is usually generated by steam boiler plant. It often contains non-vaporized water molecules which are carried over to the steam being transported throughout the piping distribution network. These water molecules can affect the performance of the steam plant and as such, it is for this reason steam systems are fitted with condensate removal equipment.

Why do we use Steam?

In its simplest terms, steam offers an excellent means of transferring a large mass of heat energy, usually via a piping distribution network. In terms of the operation of steam distribution systems, it’s the heat energy, pressure, temperature and flow rate of steam at each demand point which is of importance. This therefore places key emphasis on the design and operating performance of the piping distribution system.

Steam for power plant applications is often required at relatively high pressures and frequently needs to be superheated (dry gas). The presence of superheat improves the thermodynamic efficiency of the turbine operating cycle and also minimises condensation which can lead to erosion, leaking joints and glands. This in turn reduces the overall mechanical efficiency of the power plant. In general terms, a good system design will be based on a steam operating condition which yields the lowest pressure that will provide the heat output required, saturated and as dry as practicably possible.

An effective steam plant system design will produce steam at the desired quality (as dry as possible at the point of demand), will ensure the steam isn’t required to do more work than necessary, will be based on the process plant and equipment having a large enough heating area and ensure that the effective heat transfer from the steam to the process takes place at the highest practicable rate. The overall design can be enhanced further by ensuring that any heat given off from the process itself is harvested and utilised elsewhere on perhaps another application.

Can we Reduce Steam Demand?

Surprisingly, a considerable amount of heat energy is wasted in steam being required to carry out more work than necessary. When reviewing your steam plant, perhaps a useful starting point is to consider all components in the system between the steam generator and the demand point(s) which unnecessarily add to the total system heat demand. In general, when designing new steam systems, a good design approach is to consolidate all equipment with a steam requirement to one area and locate this area as close as possible to the steam main and ideally, steam generation plant.

There are a number of areas which can give rise to a reduction in steam demand. For instance, any steam pipework becoming redundant should be isolated or disconnected from the steam main as having redundant lines is particularly wasteful of heat energy, particularly if the piping is unlagged and/or exhibits signs of leakage. Let’s consider steam pipe sizes. If a steam pipe is oversized given the steam flow rate required, it becomes a continuous point of energy wastage. If the insulation on a steam pipe is in poor condition, has been applied incorrectly or is of a poor quality, the pipe will incur increased radiation losses which essentially means that some of the heat energy inherent in the steam line is simply wasted. A greater volume of condensate will be formed due to the greater heat loss which means wet steam will be delivered to the point of demand. To overcome this, additional steam trapping is required. Oversized lines are also by their nature, more expensive to install owing to the cost of the larger pipe, relatively higher cost of piping supports, fittings valves, insulation, and labour. These larger pipes can also contribute to a lower quality of steam due to the formation of additional condensate in the lines as noted earlier.

If steam piping is undersized, the system will incur higher pressure drops and insufficient steam flow to the demand point(s). There is also a higher risk of water hammer, noise and erosion. The enemy of the steam plant, water hammer, can cause damage to steam pipes, equipment and personnel and thus the likelihood of this phenomenon should be given due consideration and be eradicated from the system.

In older steam system installations, there can be bypass arrangements installed at steam traps. These bypass connections are often unnecessary and left open giving rise to significant heat loss. Furthermore, the correct and strategic installation of steam traps with appropriate air venting can eliminate the requirement for these bypasses and thus, the associated energy wastage.

In general, the lowest permissible operating temperature should be selected for a steam system. This obviously will optimize the amount of fuel required to generate the steam and maintain the design operating temperatures.

Giving due consideration to the points discussed, the overall system heat demand can be optimized to ensure the system is operating as effectively and efficiently as possible and consequently, reduce the overall site steam demand and associated plant fuel costs.

Can the Steam Distribution be Improved?

The dryness and quality of steam delivered to any demand point in a steam system depend heavily on how the steam is generated and of course, distributed throughout the site. In steam systems which are required to generate saturated steam, the steam begins to condense the moment it begins its journey through the distribution pipework. In general, the wetter the steam, the lower the quality for downstream process heating requirements. In such scenarios, steam carrying moisture can create a water film on the inside pipe surface. At this point it’s worth noting that, the water film is a very poor conductor of heat which of course, can inhibit the performance of the installation.

The steam distribution system can be improved by ensuring there is a correctly sized steam trap immediately downstream of the boiler plant. This trap will collect condensate from the lines thus improving the overall quality of the steam and improve efficiency.

Another point for consideration is the level of insulation fitted to the distribution pipework. Poorly insulated steam lines give rise to the formation of condensate in the steam, hence reducing the steam quality. The steam lines and in particular, flange connections and valves should be properly insulated to maximise the steam quality.

When selecting the insulation thickness for distribution pipework, the optimum insulation thickness is the economic thickness. Economic thickness considers the installed cost (material & labour) of the insulation as well as the ongoing value of energy savings over the lifetime of the plant. It is essentially defined as the thickness of insulation that minimizes the total lifecycle cost.

The following graph illustrates the lost energy cost over the lifespan of the system which decreases as the insulation thickness increases. The total cost curve relationship represents the sum of the installed insulation cost and the cost of the lost energy. The total cost curve shows a minimum value at the apex which represents the economic thickness of insulation.


Economic Thickness of Insulation (FluidFlow)

Overall, any water/condensate which develops in the main infrastructure pipework should be collected as early as possible at system low points.

In existing steam installations, it is very often the case that the installed pipework becomes too small or too large owing to changing usage patterns or changing demands imposed on the system which can develop and change over the years since the plant was originally installed. As a rule, pipes which are too small can result in lower than required steam flow at the demand point and also, pipes which are too large cause excessive heat loss due to the larger than necessary surface area/radiation surface.

A particular issue for consideration is elements of the piping system which have sagged due to lack of adequate piping supports. This sagging can result in water lodging at the base of the pipeline at the lowest point. This water can be a source of water hammer in a steam system which can be detrimental to the operation of the plant. Also, the passing steam can collect moisture from the water-logged point which if not removed by steam traps downstream, will be delivered to the demand point and once again, lower the quality of the delivered steam.

Another area of concern is the presence of air in the steam lines. Air gets into the lines when the boiler plant is off. This air should be vented from the distribution mains to improve the transfer of steam through the system. The air can be removed through the use of appropriately positioned thermostatic air vents. As a minimum, these vents should be fitted at terminal ends of various pipe runs.

It was noted earlier that the water film on the surface of the pipe is a poor conductor of heat. The air film however much more drastic. In fact, it is widely documented that air is over 1,500 times more resistant to heat transfer than steel piping material. It stands to reason that both the water and air films should be eliminated from the piping distribution system as rapidly as possible.

Steam Condensate

Another point of notable interest is the condensate and its recovery from the steam system. Condensate is essentially purified/distilled water which usually includes chemical treatment – ideal for use a boiler feed water. Condensate is of high monetary value owing to its inherent heat content and the fact it is of high quality, i.e. purified hot water. It should therefore be harvested at every available opportunity in the steam process. After all, it is much more cost effective to re-heat hot condensate into steam than it is to heat cold make up water into steam. In fact, condensate can be almost one third of the cost of generating steam. Take Note !

When developing or designing a condensate system, it should be borne in mind that condensate is a two-phase fluid and as such, sizing of the condensate lines is a much more complex process. In such cases, the fluid velocity requires careful consideration. Traditionally, condensate lines were sized on the basis they were liquid carrying pipelines. However, as we now know, these lines can carry both liquid and flash steam.

So what can I do with the recovered condensate? Firstly, we have seen that it is perfect as boiler feed water. However, it can also be used for space heating systems. The scenario which suits your site may be dictated based on your specific set of site conditions and requirements.

Giving consideration to the above, the performance and efficiency of a steam system can be radically improved through careful and considered recovery and usage of hot condensate.

CR System


Two-Phase Condensate Recovery System (FluidFlow)



This document attempts to outline some key points to be considered for when reviewing the performance and efficiencies of new and existing steam systems. Ensuring steam plant is consolidated to a specific zone, optimizing the amount of steam required including its temperature and pressure as well as its quality will increase plant efficiencies. Careful selection of pipe diameters will also have a considerable benefit in ensuring the plant operates effectively and efficiently.

A useful approach to take when reviewing the design and operation of any new or existing steam system is to think of the heat energy desperately trying to escape at every available opportunity! What can you do to prevent this phenomenon and develop a more effective and efficient system? Developing such a system will go a long way to reducing the overall fresh water demand, heat demand, associated fuel costs and site emissions.


  1. Steam & Energy Conservation – Spirax Sarco.
  2. The Proper Use of Steam – Spirax Sarco.



Hazen Williams vs Moody Friction Factor Pipeline Pressure Loss

This example case considers the flow of water in a 6 inch Schedule 40 pipeline and compares the results when using the Moody Friction Factor vs the Hazen Williams formula for the same flowing conditions. A hand calculation will initially be carried out to determine the pressure loss for the two scenarios. A model will then be developed to make a comparison with the hand calculation.

Problem Statement:

720 GPM of water at 95 F flows in a 6 inch Sch 40 steel pipe at a velocity of 8 ft/s. The pipeline has a total length of 1000 ft (equivalent length). Determine the friction loss using the Moody Friction Factor and the Hazen Williams formula.


Re = ρvD/μ


ρ               is taken as 62.057 lbm/ft3

μ               is taken as 0.000484 lbm/ft sec.

v                is 8 ft/s.

D               is 6.065 in.

Re = (62.057)(8)(6.065)/(0.000484)

Re = 5.2 x 105

The relative roughness of new steel pipe can be calculated as;

ϵ/D = (0.0002 ft / 6.065 in)(12 in/ft)

ϵ/D = 0.0004

Considering the Re of 5.2 x 105 and relative roughness of 0.0004, we arrive at a friction factor of 0.017 approximately.

The head loss due to friction can be determined as follows;

Hf = f (L/D) (V2 / 2g)

Hf = (0.017) (1000 ft / 6.065 in) (12 in/ft) (8ft/s)2 / (2(32.2 ft/s2)

Hf = 33.4 ft

Modeling this scenario yields the following results.


Figure 1: 6 Inch Pipe – Moody Friction Factor.

The modeled value for friction loss is 32.3 ft fluid which is close to the hand calculation value of 33.4 ft. The difference can be attributed to the simplifications of the hand calculation by comparison to the modeled solution.

Let’s now consider the Hazen Williams approach using the following equation:

H = Cf L Q1.852 / C1.852 D4.87


Cf              is the unit conversion fact (4.72).

L                is the pipe length (1000 ft).

Q               is the volumetric flow rate (8 ft/s)(0.20063 ft2) = 1.605 ft3/s.

C               is the Hazen Williams coefficient, initially considered to be 100 for steel pipe. Note, this is a conservative value and allows for future scaling of the internal pipe surface. The American Iron & Steel Institute’s Committee for Steel Pipe Producers recommends a C value of 140.

H = (4.72) (1000) (1.605)1.852 / (100)1.852 (0.505)4.87

H = 11337 / 181.6

H = 62.4 ft.

Modeling this scenario yields the following results.



Figure 2: 6 Inch Pipe – Hazen Williams (C = 100).

The modeled solution (62.1 ft fluid) compares well with the hand calculation results of 62.4 ft.

Note, if we use a C value of 140 for clean or new steel pipework, we get;

H = 11337 / (140)1.852 (0.505)4.87

H = 33.5 ft.

Modeling this scenario yields the following results.



Figure 3: 6 Inch Pipe – Hazen Williams (C = 140).

The modeled value for friction loss is 33.3 ft fluid which is close to the hand calculation value of 33.5 ft. Again, the difference can be attributed to the simplifications of the hand calculation by comparison to the modeled solution.

We have already calculated a friction loss of 33.4 ft using the Moody Friction Factor. Conversely, if we use a roughness value for corroded steel pipe (0.013 ft), we get;

ϵ/D = (0.013 ft / 6.065 in)(12 in/ft)

ϵ/D = 0.03

From the Moody diagram, the friction factor is approximately 0.0265. This increases the friction loss by a factor of 0.0265 / 0.017 which yields,

H = 33.4 ft (0.0265/0.017)

H = 52.1 ft

Review this scenario in a model yields the following results.



Figure 4: 6 Inch Pipe – Fixed Friction Factor of 0.0265.

Let’s summarise the results.

Screen Shot 2017-09-19 at 12.01.30

This study demonstrates the following;

  1. There is good agreement between the Darcy-Weisbach and Hazen Williams equations for clean pipework.
  2. The results are sensitive to the roughness and C factor values for corroded or old pipework installations.


  1. Piping Systems Manual (Brian Silowash).

Net Positive Suction Head (NPSH) Technical Paper

When considering Net Positive Suction Head, it is useful to differentiate between available net positive suction head (NPSHA) and required net positive suction head (NPSHR).

NPSHA is a characteristic of the system in which a centrifugal pump operates and represents the difference between the absolute suction head and the fluid vapor pressure at the prevailing temperature.

NPSHR is a function of the pump design and represents the minimum required margin between the suction head and fluid vapor pressure.

The way NPSHA is calculated depends on the system configuration. The following Figures help illustrate this point.

Figure 1 outlines how NPSHA is calculated for a given capacity of water at 80 F based on a system with suction lift (15ft).

Figure 1: NPSHA Calculation with Suction Lift.

The hand calculation of NPSHA for this scenario (at sea level) would be:

NPSHA = 2.31 (Ps – Pv) / SG + Z – Hf


Ps         is the pressure above the liquid surface (psia).

Pv         is the vapor pressure of the liquid (psia).

Z          is the static head (ft)

Hf         is the friction losses (ft). The friction losses total 2.38 + 0.46 + 0.16 = 3 ft.


NPSHA = 2.31 (Ps – Pv) / SG + Z – Hf

NPSHA = 2.31 (14.7 – 0.5) / 1.0 – 15 – 3

NPSHA = 14.8 ft.

The hand calculation produces a NPSHA of 14.8 ft which match the modeled result.

Let’s consider a scenario whereby the system inlet features a pump taking its suction from a pressurized tank at an elevation of 10ft. The fluid, in this case, is once again water at 80 F.


Figure 2: NPSHA Calculation with Suction from a pressurized tank.


The hand calculation of NPSHA for this scenario would be:

NPSHA = 2.31 (Ps – Pv) / SG + Z – Hf

NPSHA = 2.31 (14.7 + 5 – 0.5) / 1.0 + 10 – 4.001

NPSHA = 2.31 (14.7 + 5 – 0.5) / 1.0 + 10 – 4.001

NPSHA = 50.4 ft

The hand calculation produces a NPSHA of 50.4 ft which matches the modeled result.

Note, both NPSHA and NPSHR vary with capacity. With a given static pressure or elevation difference at the suction side of a centrifugal pump, NPSHA is reduced at larger flow rates by the friction losses in the suction piping and fittings. On the other hand, NPSHR being a function of the velocities in the pump suction passages and at the inlet of the impeller, increases basically as the square of the capacity. The NPSHR for a given pump is provided by a pump vendor. Many factors affect the estimation of NPSHR such as eye diameter, number of impeller vanes, suction area of the impeller, shape of the vanes, shaft and impeller hub diameter, impeller specific speed, shape of suction passages, etc. it is therefore not recommended that a designer estimates a value of NPSHR for a pump but obtain this specific data from the pump vendor.

The NPSHR describes the amount of pressure required at the inlet of a pump to prevent air bubbles from forming inside the pump unit. If a scenario arises whereby the NPSHA is lower than the NPSHR by the pump, air bubbles are allowed to form. These bubbles can implode violently inside the pump causing significant damage. This effect is known as cavitation.

Pump cavitation becomes evident when there is one or more of the following conditions present in a system; noise, vibration, drop in pump head-capacity and efficiency curves and with time, damage to the impeller by pitting and erosion.

A useful point to note is that NPSHR curves provided by pump manufacturer’s are usually based on using cold water for the pump test conditions. Thus, it might be assumed that the NPSHR by a centrifugal pump for satisfactory operation is independent of liquid vapor pressure at the pumping temperature and of course, this is not true. The NPSHR for a given capacity can vary appreciably for different fluids over a range of temperatures. In its simplest form, even when pumping water the NPSHR decreases when the water temperature increases.

What options are available to me to increase the NPSHA:

  1. Reduce the resistance in the suction side of the system and hence the associated pressure losses.
  2. Raise the height of the supply tank in open systems.
  3. Raise the liquid level in the supply tank.
  4. Lower the pump installation height.
  5. Increase the surface pressure of the liquid in a closed system.
  6. Monitor and control the fluid temperature.
  7. Lower the pump operating speed.
  8. Use a larger impeller eye area.
  9. Use several smaller pumps in parallel.

Item 9 above may initially appear to be a costly solution however, in many cases, three half capacity pumps of which one may be a spare, are often no more expensive than one full capacity pump plus its spare. In many cases, just two half-capacity pumps can be installed without a spare since part-load can still be carried if one pump has failed and is temporarily out of service. In addition, if the demand varies widely, operating a single pump during light-load conditions will save energy.

This technical paper attempts to briefly describe NPSH, outline some of problems which can arise when the NPSHA is lower than the NPSHR and what measures can be considered to remedy such conditions. In the end, careful design together with liaising with the pump vendor regarding the NPSH requirements for a proposed pump model will go a long way to eliminating any potential operational problems.


  1. CIBSE Guide B1 2016.

Petrojarl 1 FPSO (Marine Systems, Floating Production, Storage and Offloading Unit)

Back in November 2014, Nevesbu and Iv-Oil & Gas joined forces and started the Pre-Contract Engineering for the contract to modify and upgrade the Petrojarl 1 FPSO. It was anticipated that the upgrade will take around a year to complete. On completion of the project, the FPSO will be used as an early production system (EPS) unit on the Atlanta field located around 185 kilometres offshore from the Brazil coast. This project presented many challenges and was also unique in that an EPC redeployment upgrade contract had not been carried out in Europe before.

Teekay Offshore awarded Damen Shiprepair Rotterdam (DSR) the complete EPC contract for the modification and upgrade of the Petrojarl 1 FPSO. DSR in turn contracted Nevesbu to take care of the Marine Systems, Structural (design and engineering) and Naval Architecture for all modifications. Iv-Oil & Gas was then requested by Nevesbu to execute the Process, Mechanical, Electrical, Instrumentation, Structural, Piping Design and Piping Engineering. Frames Process Systems (FPS) was also selected for its expertise in the field of separation. They provided a large amount of the skid mounted package units (skids) and worked closely together with Nevesbu and Iv-Oil & Gas for integration of the skids with the vessel.

On completion of the initial demolition works, in April 2015, the Petrojarl 1 was moved into DSR’s dry dock, where the refurbishment, modification, and upgrade of the ship was carried out.

Fast-forward to August 18, 2017, Damen Shiprepair Rotterdam (DSR) successfully delivered the FPSO Petrojarl 1 to Teekay Offshore following a complete redeployment project which took place over the past 2 ½ years.




Scope of work 
The self-propelled FPSO spent 14 months in Dock No. 8 (300 x 50m) undergoing refurbishment of its marine systems, underwater hull, seawater system, crane booms, heating coils in the cargo tanks and specialised steelworks in the upper and lower turret areas, which needed to be completely revised and adapted to suit the 1500 metre deep mooring location. Simultaneously, new designed high quality, prefabricated equipment skids containing heating, cooling, separation, compression, boilers, centrifuges as well as a new E-house with electrical equipment were placed on board. Interconnecting piping and cabling was subsequently installed to complete the topsides and connect it to the remaining facilities.

The process installation consisted of the following:

  • Crude oil, produced water and associated gas separation;
  • Crude oil dehydration and desalting;
  • Produced water treatment;
  • Associated gas treatment and fuel gas distribution;
  • Fresh water washing system to meet the maximum allowable salt content in the stabilised crude oil;
  • Upgrade of existing seawater (cooling) system;
  • Steam system to heat the produced crude oil and generate fresh water from seawater;
  • Upgrade of chemical injection systems;
  • Upgrade of existing metering systems to comply with Brasilian requirements;
  • Pigging facilities to enable the launching/receiving of pigs from pigging operations;
  • Modify the existing piping arrangement on the manifold deck;
  • Upgrade of existing subsea controls;
  • Well services/diesel flush system to allow for diesel to be used for flushing crude oil from production, service risers or flowlines on shutdown, for preservation and also for heating prior to startup;
  • Upgrade of well kill system to reduce drilling rig workover time and to reduce produced crude oil handling in the rig using brine.

According to DSR, the project involved more than 450,000 engineering hours and more than 50% of the process equipment was removed and replaced by new and additional equipment.

Teekay had operated the Petrojarl 1 for 28 years in the North Sea, but it is now destined for the Atlanta field offshore Brazil.

Located in block BS-4 in the Santos basin, Atlanta is a postsalt oil field, 185 km (115 mi) offshore Rio de Janeiro in a water depth of 1,500 m (4,921 ft). First oil is expected in 1Q 2018.

Queiroz Galvão Exploração e Produção S.A. is the operator of the block with a 30% ownership along with consortium members OGX Petróleo e Gás S.A. (40%) and Barra Energia do Brasil Petróleo e Gás Ltda. (30%).

We are proud that FluidFlow software was used to design and model the upgraded seawater cooling for topsides and diesel transfer system of the Petrojarl 1 FPSO vessel together with checks on other piping systems required for the successful operation of the FPSO.



FluidFlow v3.43

General Release summary: The thermodynamic capabilities have been extended for single phase fluids. FluidFlow can now accurately cover all phase regions and enthalpy paths through and from supercritical phase region. Calculations within supercritical region are now valid. This release also contains some minor bug fixes.

General Release info:

* Extended the thermodynamic capabilities so that the product can accurately cover all regions and enthalpy paths through supercritical phase region. Calculations within the supercritical region are also available.

* Supercritical phase state now fully detected and shown in results.

* Added marine fuel oils to fluids database.

Valve Authority – Technical Paper

Control Valve Authority

Control Valve Authority – Technical Paper

Valve Authority is a term used to describe the basis on which a control valve is selected. Quality in design represents the correct consideration of many engineering decisions and sizing control valves based on authority is an important factor. If the control valve is oversized, the performance, life span and reliability of the valve and other equipment items will be reduced. If we undersize then we may have the authority but cannot achieve the design capacity. Additionally, designs should consider energy optimization but this often leads to poor authority.

The Valve Authority (N) is generally defined as the ratio of the pressure drop across the fully open valve compared to the pressure drop across the entire circuit (including the valve) at design flow conditions. Valve Authority is expressed using the following equation:

N =ΔPvalve /ΔPtotal


N              is the Valve Authority.

ΔPvalve    is the pressure drop across the valve in the fully open position.

ΔPtotal     is the total pressure drop across the circuit.

In order to develop good control, it is recommended that a control valve is selected to achieve a valve authority of 0.5 or greater. An authority below 0.25 gives unstable control; 0.25 – 0.5 gives fair to good control whereas 0.5 -1.0 gives excellent control. The higher the authority, the greater the energy wastage.

Let’s consider a simple water circuit example where the available pump pressure is 13 kPa at maximum flow conditions, i.e. the valve is in the fully open position. This 13 kPa represents the total frictional resistance of the circuit including the control valve which has a pressure drop of 6 kPa in the fully open position.

Valve Authority


Considering the pressure drop across the control valve, the pressure drop across the remainder of the circuit will be 13 – 6 = 7 kPa. The valve authority can therefore be calculated as follows;

N = 6 / 6 + 7

N = 0.46

Let’s consider a design example to explore the phenomenon of valve authority a little further.

Problem Statement:

It is intended to model a section of pipework downstream of a heat exchanger transporting Methyl Diethanolamine (MDEA) going to a column. The system pressure shall be controlled by a pressure reducing control valve with a design valve authority of approximately 0.5. The engineer is also required to ensure that the pipelines in the system are to be sized such that the gas superficial velocity (vapor velocity) exiting the system is between 10 & 15 m/s.

Design Data:

Temperature of MDEA at system inlet: 106.5 °C.

Vessel Surface Pressure: 5 bar a.

Elevation to Vessel Base: 3 m.

Outlet Elevation: 31.27 m.

Pressure at outlet: 1.8 bar a.

Design Flow at system outlet: 367,000 kg/h.

Fluid Vapor Quality Approx. 0.06.

Solution Approach:

Let’s develop this outline scheme design such that the pressure drop across the control valve is minimised which is consistent with good control. Good control of course means that the design valve authority should be around 0.5 and in all circumstances, must be above 0.2.

We know the vessel surface pressure is 5 bar a and the design outlet pressure is 1.8 bar a. The total system pressure loss between the inlet and outlet of the circuit can be established as 3.2 bar. This means that, based on a design valve authority of 0.5, we can estimate the valve pressure loss at the design flow rate of 367,000 kg/h to be around 1.6 bar (3.2 bar x 0.5 = 1.6 bar). This data can be used as the basis of a preliminary design as outlined in Step 1.

Step 1

A preliminary design of the system has been developed based on the design flow rate provided. A Flow Coefficient (Kv component) can be used in the first instance to represent a pressure reducing valve as at this stage, we simply wish to define a fixed pressure drop of 1.6 bar across the component. Figure 2 provides an illustration of the system.


MDEA Circuit

Figure 2: MDEA Circuit – Control Valve represented by a Flow Coefficient (Kv).

When this initial design is calculated, the MDEA in the system is detected as being a two-phase fluid with a vapor quality of approximately 6.7%. For two-phase systems, a useful rule of thumb is to keep the gas superficial velocity around 15 m/s at this vapor quality.

At this point, it’s often useful to know what flow regime is encountered in the system. In this particular design case, the flow in the lines is in the annular mist regime which is preferable to slug flow which of course should be avoided. This undesirable slug flow regime can be avoided by keeping the gas superficial velocity relatively high.

Annular Mist Flow Pattern Map

Figure 3: Annular Mist Flow Pattern Map.

Various pipe diameters can be applied to this system in an attempt to achieve a flowing velocity which is as near as possible to the design velocity of 10 to 15 m/s. This preliminary design case has been solved using 300mm pipes which generates an initial velocity of 14 m/s. It is worth noting at this point that as the fluid flows along the pipelines, the frictional resistance of the various pipes and fittings results in an increase in the vapor quality and hence, gas volumetric flow rate and flowing gas velocity. This phenomenon would at this early stage indicate that it may be prudent to increase the pipe diameters in the system – perhaps downstream of the pressure reducing valve. Indeed, a check on the velocity at the system outlet indicates the velocity increases to 112 m/s. This velocity is clearly much too high and should be reduced as the design evolves.

Step 2

The outline design completed in Step 1 can now be developed further. The Kv component can be changed to a PRV (pressure reducing valve) and this valve can be automatically sized. When sizing the valve, the outlet pressure in the MDEA system can be set to 1.8 bar a (based on the design data provided). The design setpoint pressure of the PRV can then be adjusted until the design flow rate of 367,000 kg/h is achieved in the system.


MDEA Circuit with Pressure Reducing Valve

Figure 4: MDEA Circuit with Pressure Reducing Valve (PRV).

Step 3

At this point, it is worth considering different pipe diameters downstream of the PRV. The PRV setpoints pressure can also be adjusted until we obtain the required design flow for each pipe diameter considered. The calculated results are outlined in Table 1.


valve authority PRV figures

Table 1: Review of Valve Authority for different Line Sizes & PRV Set-points.

We can see that the preliminary line sizes of 450 – 500 mm produces a sufficiently high valve authority (N = 0.525 & 0.56).

The next stage is to select the correct line size whilst giving consideration to the limiting velocity criteria of 15 m/s for the vapor phase. The most suitable pipe diameter is therefore the 500 mm pipe. Smaller pipe will also work, however we need to avoid choked flow in the system. We therefore need to keep the gas velocity below a max of 50 m/s. If capital cost was the main design criteria, you could consider using a smaller pipe size.

Figure 5 provides an illustration of the system at this stage of development.


guide to Valve Authority

Figure 5: MDEA Circuit – Pipe Sizes Optimised.

So far, the calculations have produced the valve Cv along with fluid physical properties. This data can be provided to the valve supplier so that they can make the final valve selection for their range of equipment.

Step 4
Finally, in this step the automatically sized pressure reducer can be swapped for the actual vendor control valve (Figure 6). In this final case, the pressure loss across the valve is calculated to be 1.5 bar which produces a valve authority of 0.47 (1.5/3.2 = 0.47).

Valve Authority

Figure 6: MDEA Circuit – Vendor Control Valve.

This example helps us consider the application of valve authority in a real system design whilst giving due consideration to other aspects of the system such as fluid physical properties, valve Cv, design flow rates, pressures etc.

FluidFlow v3.42

Fixed a bug that occured when selecting “mm Water g” OR “m Water g” units.This was originally fixed in V3.39 but regressed in V3.41. New procedures in place to eliminate code regressions.

FluidFlow v3.41

General Release info:

Improved pipe heat loss calculation with the addition of new correlation for estimating outside film heat transfer coefficient. Bug fixes.


* Crane Tee Junction TP410 pre 2009 relationships. New recommendation is to use for existing/legacy calculations ONLY. Input editor text now reflects this.

The relationships are just too simplistic to realistically predict the pressure losses over all possible operating conditions and totally ignore any pressure recovery effects.

* All Heat Transfer Coefficients in pipe results were shown as W/m C where units shown are W/m2 C. Coefficient values now normalised to W/m2C by multiplying original value by the log mean radius.

This is more in line with most literature sources. Change was made not because answers or calculations are incorrect but because it is now easier compare values directly with common literature sources.

* Following a suggestion from some of our German customers, we incorporated a new improved method for estimating the outside coefficient in pipe heat loss/gain calculations.

The literature source for the improved method is the Springer publication “VDI Heat Atlas (VDI Wärmeatlas)”.

Previously FluidFlow used ASTM Standard C680. It is still possible to use the original ASTM Standard C680 relationship, selected from the Options -> Calculations Dialog -> Global Settings tab.

* Co2 gas density and specific heat definitions have been adjusted to provide more accurate results over a wider temperature and pressure range.


* Diffusers at choked conditions show the inlet results as outlet results.

* In networks containing centrifugal pumps and more than one fluids with viscosities of > 800 cP, sometimes Pump viscosity corrections were made when this should not be the case.

* Outside heat loss convection was overestimated in rare cases for gas flow along medium/long pipes.

* Improved convergence testing to prevent the solver converging too soon. For gas calculations this sometimes resulted in incorrect downstream temperatures.

* Reservoir No Flow – Bug in creation of accumulator for solver.

* Pump derating and speed changes were not working when both effects occurred together.

FluidFlow v3.40

General Release info:

Control Valves: Improved consistency of calculated valve position and valve coefficient over the complete operating range.

Polyethylene pipes, new pipe sizes added.

Petroleum fraction properties, NBP050 to NPB450 range added to fluids database.


  • Added back a directional definition for all flow control valves. This requires the user to specify the flow direction through the valve and is necessary to reduce interaction between flow control valves in large networks
  • Removed Text Import and Export menu options because format is now out of date and is not supported in the future.


  • Multiple flow controllers in a network could occasionally cause convergence BEFORE the network has actually converged. See changes above.
  • Liquid line with heat transfer in buried pipes, sometimes calculated out phase state as two-phase when phase state should be liquid.
  • Pipe scaling was not read in properly from old files (pre V3.30).

FluidFlow v3.39

General Release info:

Calculation procedures for viscosity correction method for centrifugal pumps has been updated to HI 2015 guidelines

Bug Fix:

Units m Water g and mm Water g went missing in V3.38. Now reinstated.

FluidFlow v3.38

Contains fixes for all reported bugs up to end of September 2016.

General Release info:
No new enhancements, bug fix maintenance release only

None in this release

* Fixed a rounding error (<2%) that occurred when converting to m Water and m Water gauge.
* Improved the chart visibility of slug region in two-phase flow pattern maps, previously part of the area was overwritten by elongated bubble regime.
* Fixed a bug in outlet velocity calculation when phase change occurs within a pipe.
* Fixed a bug that caused phase change within a pipe when Heat Loss Model = “Ignore Heat Loss/Gain”, without taking into account heat of vaporization.
* Fixed a bug that caused flow reversal at known flow nodes out of a network at very high specified flows i.e. many times greater than sonic flow.
* Fortis Only – Overcome the instability caused by the discontinuities in the Universal Gas Sizing Equation for control valves.
* Fixed a bug which shows gauge pressure results incorrectly if atmospheric pressure is changed.

Two-Phase Flow


Gas-liquid two-phase pipe flow is of significant importance in a wide range of engineering industries such as steam generators, chemical process plant, distillation processes and heat transfer systems. The design of these systems is often a complex phenomenon

Using suitable engineering simulation software can help the engineer design efficient and effective systems, understand plant performance and quickly evaluate alternative design scenarios.

In two-phase flow, the vapor mass fraction is often not constant and there is mass transfer between the fluid phases. FluidFlow takes this into account in your model solution. In fact, you can see the results for inlet and outlet vapor quality for all pipes and elements in your system. Flow pattern maps are generated automatically for all pipe in the system, helping you identify flow regimes and any undesirable operating conditions.

FluidFlow is used successfully by engineers to calculate pressure losses and flow distribution in two-phase pipe flow systems. The simulation software will automatically track fluid phase-state throughout the piping distribution system and the software is provided with a comprehensive database of two-phase fluids, boosters and associated piping equipment. Automatic control valve and equipment sizing is included helping you to accelerate the design process.

FluidFlow is easy to use and new users are provided with a Designer Handbook meaning you can tackle those design projects instantly.

For more information on Two-Phase Flow click here

Compressible Flow Systems

compressible flow

FluidFlow customers use the software to design and optimize a wide range of compressible pipe flow systems including; natural gas transmission pipelines, steam distribution systems and compressed air systems.

For compressible fluid flow in pipes, the pressure and temperature conditions continuously change as a gas or vapor flows along a pipeline. This means that the physical properties of density, viscosity, heat capacity, thermal conductivity, velocity etc, change with pipe length.

FluidFlow uses a number of compressible flow equations, and incorporates the Joule Thomson effect to obtain a rigorous solution which is accurate for both low and high velocity flow systems.

By using FluidFlow, engineers can accurately calculate compressible flow through an orifice plate, control valve, relief device, nozzles, valves and all common piping components. You can also automatically sizing pipes, pumps, ducts, fans, compressors, control valves, relief devices (ISO & API), orifice plates and nozzles.

The software can also calculate heat loss/gain from pipes and model buried pipe heat transfer. The software is provided with a library of insulation materials as standard and engineers can select the desired insulation thickness. Convection, conduction and radiation losses are calculated. This means you can use FluidFlow to optimize energy use by selecting the economic insulation thickness.

For more information on Compressible Flow click here

FluidFlow Pressure Drop Calculator

FluidFlow is a pipe flow calculator which is used to perform fluid flow analysis in piping systems featuring heat exchangers, orifice plates, control valves, pumps, venturi flow meters and equipment items.

The software is easy to use and supported by an experienced team of engineers which are always willing to lend a helping hand.


FluidFlow is a modular pipe flow calculator

Available calculation modules include

  • Liquid
  • Gas
  • Two-Phase
  • Non Newtonian and Settling Slurry
  • Scripting (Dynamic Analysis)


FluidFlow Liquid Module

The FluidFlow Liquid Module is a water flow calculator which allows you to define vendor equipment to a database such as pumps, control valves, pumps, venturi flow meters etc. When defining pumps, you can enter your vendor-specific pump curves to the database which will be stored for all your modeling projects. You can then model the performance of the pump in your system.


The software enables designers to complete fluid flow simulation studies in an instant and automatically size equipment, taking the pain out of your design projects.


Flow Measurement In Piping Systems

FluidFlow is a flow calculator which allows you to understand flow measurement in your piping systems. You can input data obtained from a site pressure test to a model and understand exactly how your piping system is performing.


FluidFlow solves the continuity of mass, energy and momentum equations. You can clearly view result for flow rate, pressure, pressure loss, velocity, density, viscosity, temperature, Reynolds number and friction factor.

FluidFlow Overview

FluidFlow is primarily a maintenance release addressing reported bugs and adding new features requested by our users. It is recommended that all users upgrade to this release. We have updated our control valve calculation code to reflect the very latest Instrumentation Society America ISA-75 Guidelines, this includes choking detection for both liquids and gases.


  • Large networks containing many tee junctions now converge quicker. Note, it is still important to place the branch orientation (red dot) correctly.
  • Improvements made to equipment sizing consistency.
  • All Control Valve Equations now updated from ISA 1985 to ISA 2007 guide. For choked valves a warning is now provided. Added calculation of Cv at liquid choked flow conditions.
  • Added network data caching to improve performance of network version.
  • Added the ability to use visual elements on script forms (labels, list boxes, combo boxes, tabs, grids, etc.).


  • Open pipe exit pressure reverted back to pre 3.31 where stagnation pressure is assumed to be atmospheric.
  • Removed mouse wheel support for the input editor to prevent clashes with flowsheet occurring. Flowsheet now works better using mouse wheel.

FluidFlow v3.37

Primarily a release for Fortis, including all requested logging and database enhancements, bug fixes regarding gas reducers, and series choking.

General Release info:


  • Added the ability to automatically adjust atmospheric pressure for altitude.
    Options -> Calculation-> Global Settings.


  • Fixed an overwrite of value 101325 Pa a, for atmospheric pressure that sometimes occurred when resetting defaults.
  • Improved network convergence for multiple flow controllers.
  • New warning advising when a pump is acting as a turbine.
  • PD Pump auto-sizing bug fixed.
  • Fixed a bug which caused a program crash when changing to French language.

FluidFlow v3.33

Version 3.33 is primarily a maintenance release addressing reported bugs and adding new features requested by our users. It is recommended that all users upgrade to this release.

We have added the ability to make buried pipe heat loss calculations. To support this calculation, different pipe coatings, soil, and backfill types have been added to the Insulation database. In addition, the thermal conductivity as a function of temperature relationship for all insulation materials and soils has been improved.

For gases flowing at or near the saturation point in pipes with heat loss on, we have added an option to include condensate traps. If this option is “on”, the flow will stay in the gas phase at the saturation point and the software will report the mass of condensate to be removed. By default, this option is “off”, and so some of the gas will condense and become two-phase as it flows down the pipe. This means the vapor quality decreases along the flowpath because the condensate is not removed.


  • Added the ability to make a buried pipe calculation. Added different pipe coating, soil, and backfill types to the insulation database.
  • Added phosphine gas to physical property database.
  • Improved insulation thermal conductivity as a function of temperature relationship and added more data for insulation materials.
  • Added the ability to assume steam traps are present for a condensing gas. This option (‘Options | Calculation…’ menu item; Gas page) prevents gasflow from developing into 2-phase flow when heat loss is included.
  • Bill of Materials now subdivides pipes into schedules.
  • Added ability to model expansion loops in one component, instead of drawing out each individual expansion loop.
  • Added more calculation examples to QA tests and updated help file.


  • For Open Pipes and Open Boundaries with a resistance, the exit static pressure is now assumed to be atmospheric pressure. In earlier releases, the exit stagnation pressure was assumed to be atmospheric.
  • Restricted tee junction K values to maximum and minimum values: Max = 90 and Min = -15. This aids convergence without limiting practical values.
  • Added the ability to include Joule Thomson Coefficient in “Do Heat Loss Calculation”, via the ‘Options | Calculation…’ menu item; Gas page.

FluidFlow v3.31

With version 3.31, we have really focused on empowering you to build your models faster. We’ve implemented unique, cutting edge tools that will not only reduce the time it takes you to build your model on the flowsheet – we’ll also do the design for you!

Automatic Pipe Sizing: In addition to the economic pipe sizing feature, designers can now specify a desired pressure gradient or nominal velocity for each pipe element in the model. FluidFlow will then calculate the pipe diameter required to achieve that design constraint

Automatic Equipment Sizing: We already have auto-sizing in place for safety relief valves and burst disks to both API & ISO standards for liquid, gas, steam and two-phase flow systems. Now FluidFlow can auto-size all your key equipment items. Pumps, compressors, fans, orifice plates, nozzles, venturi tubes, pressure and flow controllers can now all be sized automatically, saving you time and effort.

Automatic Flow Balancing with Orifice Plates: Orifice plates can now be sized based on design pressure loss or flow rates. No more iteration in your design as you adjust orifice diameter to achieve your desired flowrate – FluidFlow will do it all for you.

Flowsheet Improvements: Getting your flowsheet built quicker is key to efficient modelling. So we’ve changed our directional components. No more “red dot” to define the flow direction of a controller. Just drop the pump or controller on the flowsheet and FluidFlow will work it all out for you. You can now also hold down the CTRL key when adding a template to continue adding multiples of the same template.

Reporting Upgrades You can now view your flowsheet data report from directly within FluidFlow. No need to export to excel to look at the raw data. The report has its own dialog box, so you can put it on a separate screen, or have it side by side with your flowsheet. We’ve improved the print quality of our exports too, so your .pdf exports are now crystal clear.


  • Ability to autosize the following equipment items: Pipes, Centrifugal Pumps, Compressors, Fans, PD Pumps, Pressure Reducers, Pressure Sustainers, Differential Pressure Controllers, Flow Controllers, Orifice Plates, Inline Nozzles, Venturi Tubes, Safety Relief Valves, and Bursting Disks.
  • Added 2 additional pipe sizing criteria options: Pressure Gradient and a user-entered Velocity.
  • No longer necessary to define the flow direction of pressure controllers.
  • No longer necessary to define the flow direction of flow controllers.
  • Hold down CTRL when Inserting a Template to continue insertion. Same logic applied to Paste.
  • Components Bar (‘View | Components Bar’)
  • New Data Report. Flowsheet Data/Data Report. (‘View | Flowsheet Data’)
  • Improved Print and Export Quality.
  • Integrated Joule-Thomson derivative into existing Equations of State and compressible flow calculations.
  • Added JT Coefficient, slurry volume fractions for each component (4 component and Liu models) to results table, printouts, etc.
  • Increased accuracy of calculation of Control Valve Cv and %Opening for gas systems.
  • Added 2-phase loss correlations for control valves. Not covered by ISA guide. Used Parcol method.
  • Default Folders for License, Data, Preferences, and Templates moved to “C:\Users\Public\Documents\Flite\FluidFlow” for NEW installations.

FluidFlow v3.23

  • Added Flowsheet Undo facility.
  • Added ability to plot two-phase flow pattern charts for ALL pipe inclinations.
  • Added ability to plot composite charts for boosters in parallel and series. Access via New flowsheet toolbar button.
  • Added ability to plot HGL and EGL charts.
  • Updated Pipe Sizing Data.
  • DATABASE ADDITIONS – Over 50 new fluids added, 40+ Pumps and valves + many more fluid equipment items.
  • CHANGE: Density results shown for all phase states are now static density. Previously Stagnation Density was shown.

FluidFlow v3.22 Build 6

  • The calculated K value is now an available result (in export, excel, tables, etc.) for the following elements: GenericK, Inline Filters, Cyclones, Expansion Bends, Known Resistance Exits, Bends, & Mitre Bends.
  • Buried Pipe Calculation now uses an iterative solution process, which results in a more accurate solution. Log Mean temperature is calculated more accurately.
  • Now possible to “Check for Updates” from the Help menu.
  • NPSHa now shown for auto boosters and PD pumps.
  • Improved heat transfer in two-phase flow.
  • Heat Loss from Pipes – Instead of R values (heat transfer resistances) being displayed, individual transfer coefficients (Inside Film, Pipe Wall, Insulation, Outside Film) are now displayed. This change was requested by several users.

FluidFlow v3.32

With version 3.32, we have improved our physical property predictions and have introduced the ability to model petroleum fractions from ASTM D86 or True Boiling Point Curve data.

We have improved the accuracy of our two-phase flash calculations, added pressure recovery effects and generally improved the two-phase solution algorithms.

Our pre-release Quality Assurance tests have been expanded (we now test over 700 worked examples, covering all phase states).

For slurry calculations a new deposition calculation method “Oroskar and Turian” has been added.

The flowsheet results presentation has been improved and this is also reflected in the generated reports.

We continue to work hard on our upcoming Version 4 product which is due to be released later this year.


  • Added ability to model Petroleum Fractions to the Physical Property Estimator and the property database.
  • Added pressure recovery into 2-phase flash calculations. The effect of this becomes apparent with liquids at their boiling point.
  • Added new correlation for estimating Settling Slurry Critical/Deposition Velocity. Oroskar and Turian.
  • Added ability for a pump to handle low two-phase quality mixtures at suction.
  • Improved Two-Phase flow calculations by tracking enthalpy along pipe; this allows for more accurate flash calculations.
  • Added ability to show control valve charts from the flowsheet.
  • Improved flowsheet text and fly-by formatting.
  • Added ability to activate over a network.
  • Addition of an automatic backup to the DATA folder. Occurs every 30 days by default.
  • Added more QA examples. We now have over 700 examples that are fully checked before each new release: 85 Two-Phase, 176 Compressible, 79 Equipment Sizing, 251 Incompressible, 79 Non Newtonian, 11 Petroleum Fraction, 8 Saturated gas and Two-Phase and 22 script examples.


  • Velocities and Pressure for ALL size change elements are now based on the actual size of the element.
  • Added additional checks for out of range atmospheric pressure changes made by the user via the ‘Calculation Options’ dialog.
  • Removed ability to customize the Components Palette.

FluidFlow v3.30

  • Added NEW settling slurry calculation method. 4-Component model based on a 2007 paper by Sellgren and Wilson.
  • Added NEW settling slurry calculation method. Liu Dezhong method.
  • For settling slurries, size distribution charts can now be viewed at the Supply Node(s).
  • Added the ability to Size Relief Valves and Bursting Discs to API520 or ISO4126
  • Flowsheet – Added the ability to drag an OpenPipe node to make a connection to any other node (provided node can accept another pipe connection) OR to drag an OpenPipe into an existing pipe.
  • Flowsheet – Added the ability to rotate any node, in 90° increments for ortho mode and in 30° increments for iso mode.
  • Flowsheet – Added the ability to store a flowsheet as a template. Templates can be inserted into existing networks and connected via drag connect.
  • Added a resources window, that connects back to website for examples and videos.
  • Improved the temperature and pressure range and the accuracy of the internal relationships used to predict air physical properties.
  • Expanded the pipe heat loss calculation method, to enable direct entry of U values. Full calculation of U value is also available.
  • Added Specific Heat Capacity, Flow Cross Sectional Area and Out Flow Cross Sectional Area to results table, flowsheet properties, fly-bys etc.
  • Data added for PE100 Polyethylene Pipes, Non Newtonian fluids (phosphate clays, red muds, fly ash, sugar processing fluids and various foodstuffs).
  • Created a Licence Manager that also gets installed with the software, allowing greater licence flexibility.

FluidFlow v3.22 Build 5

  • Added ability to calculate heat loss in buried pipes.
  • Improved calculation consistency for non-Newtonian Casson and Hershel Bulkley fluids.
  • Input Inspector now highlights properties that have been edited (i.e., changed from default values).
  • All calculation modules are now available for each calculation, but user has a restricted view of results if a calculation module is missing.
  • Improved gas mixture physical property prediction. You can now use mole or volume fraction to define a mixture.
  • Improved gas calculations with heat transfer. Better integration of heat transfer equations into loss calculations and improve consistency of results.
  • New warning added to slurry module if user attempts to input a solids concentration greater than the packed bed voidage.

FluidFlow v3.22 Build 4

  • Added back the ability to decide on the Vsm calculation method for settling slurries.
  • Added results to text output.
  • Expanded scripting section of the help file.
  • Added some new PP and PPF Pipes and some new PP diaphragm valves into databases.
  • Stopped an “Out of Resources” error occurring in the scripting window if margins were shown.
  • Fixed a heat transfer bug that occured when 2 exchangers were used in series, with the Heat Transfer Option set to “heat transfer into network”.
  • Fixed a bug that occured if a dialog box was shown off screen.

FluidFlow v3.22 Build 3

  • Networks can now be imported and exported via text files.
  • Added ability to size pipes automatically – Beta Release.
  • Over 50 new fluids added bringing total of fluids in the database to over 1050.
  • Settling Slurry Calculations – Method extended to included inclined pipes.
  • Added ability to run help files locally from the network release.
  • Database mixtures can now have a fixed phase state, this is useful for users who do not have the 2 phase module.

FluidFlow v3.22 Build 2

  • The V3.22 upgrade now supports Paper/Pulp Stock calculations as a standard part of the Slurry and non-Newtonian module.
  • ASME and ISO 4126 calculation methods for predicting flow across safety Relief Valves are also available across all modules.
  • This release also provides the ability to export and import network designs via text files, meaning it is now possible to interface FluidFlow to other applications.
  • The dynamic analysis and scripting module has undergone a major internal reorganisation to speed improvements and additional functionality. Users can now write scripts in Basic as well as Pascal, for instance, users can easily call Excel directly from script with methods available to set and get Excel data and charts.
  • In response to user requests, the V3.22.2 release has improved UI speeds and additional abilities such as the specification of stagnation or static pressure at boundaries.

FluidFlow v3.22 Build 2

  • The V3.22 upgrade now supports Paper/Pulp Stock calculations as a standard part of the Slurry and non-Newtonian module.
  • ASME and ISO 4126 calculation methods for predicting flow across safety Relief Valves are also available across all modules.
  • This release also provides the ability to export and import network designs via text files, meaning it is now possible to interface FluidFlow to other applications.
  • The dynamic analysis and scripting module has undergone a major internal reorganisation to speed improvements and additional functionality. Users can now write scripts in Basic as well as Pascal, for instance, users can easily call Excel directly from script with methods available to set and get Excel data and charts.
  • In response to user requests, the V3.22.2 release has improved UI speeds and additional abilities such as the specification of stagnation or static pressure at boundaries.