Flow Simulation in Pump/Compressor Design

Pumps & Systems, February 2009


Computer simulations of pumps and compressors can now serve the same function as hardware testing. These simulations can be done in less time with less cost while providing engineering data of similar quality. Furthermore, computer modeling can be performed directly by the engineer doing the hardware design, thus providing a tight link between analysis and design optimization.

Computer Pump/Compressor Simulations

We will start by defining pump simulation. Modeling and simulation can take many forms, but in this article, pump simulation refers specifically to 3-D computational fluid dynamics (CFD), an example of which is shown in Figure 1.

Figure 1. CFD model of an external gear oil pump (surface pressures and x-y data plots)

In this context, Computational Fluid Dynamics applies to liquid or gas, compressors or pumps, and can be used to model fluid motors and a wide range of other fluid components. Such CFD codes usually start with a CAD model of the geometry and then create a 3-D numerical mesh representing the flow path through the device. This mesh is subsequently used to model the dynamics of the flow based on fundamental laws for conservations of mass and momentum. The output of these models includes plots and three-dimensional maps of flowrates, loads, head-rise, power, pressure ripples, velocities and torques, depending on whether it is for a pump, compressor or motor.
For liquid applications, the more advanced codes include aeration and cavitation. Aeration refers to the presence of non-condensable gases, such as air; and cavitation refers to the formation of vapor from the liquid. Both can have a significant effect on pump performance and life.
If temperature influences the properties and performance, such as in compressors, the code must also solve conservation of energy equation (Figure 2).

Figure 2. CFD predicted air temperature (on a cutting plane) in a lobe compressor

To be effective, CFD codes should provide the same data as a hardware test. The question is often raised, however, whether these simulation tools are sufficiently accurate, easy-to-use and fast to be used reliably as virtual hardware tests.
The Advantages of Simulation

The general advantages of computer simulations have been well established. For the majority of pump/compressor developers, computational stress simulations have been an integral part of their design process for decades. This is true for axial, centrifugal and positive displacement (PD) pumps. In contrast, the use of CFD simulation has lagged behind other computer simulation tools because of difficulty-of-use, slow turnaround time, and inaccuracies or deficiencies in the results.
Within the last few years, however, state-of-the-art pump modeling has evolved to where setting up a detailed model of even a complex pump takes less than an hour. This reduction in set-up time has been accomplished, in part, by new meshing algorithms that start with a CAD file and automatically generate the numerical grid needed for CFD. With the newer software, running a simulation on a standard desktop now takes between one hour to overnight, depending on the type of pump and the extent of the system. These statistics include both axial/centrifugal pumps and positive displacement (PD) pumps, the latter of which are more challenging to model from a numerical perspective.
Using state-of-the-art results from a good CFD code should be within 10 percent accuracy or better compared with experimental data. Accurate and quick simulation can reduce the time and cost associated with hardware testing by providing the same type of data that would be collected during testing. With simulation, it is easy to add "measurement" probes in locations that would be extremely difficult to access in a physical experiment. Perhaps most importantly, numerical simulation offers flow visualization that allows the engineer to "look" inside the pump during operation and provides invaluable insight as to how the pump operates and how it might be improved.
Validation and Application

The engineering data of interest for the modeler are the same as those of the designer and test engineer. These include flowrates, pressure ripple, efficiency, torque, power, head rise, loads, cavitation damage and cavitation onset. These same data are in turn the focus for validation of CFD tools. Figure 3 provides an example where CFD is used to predict flowrate vs. RPM in a crescent pump [1]. Comparison with experiments indicates the model accurately captures the drop in performance at higher RPMs due to the combined effect of aeration and cavitation. This type of analysis is used to improve the porting and/or system integration for a given pump design or to compare pump designs.

Figure 3. Flow rate predictions for a crescent pump

CFD can also predict head-rise, power and efficiencies for PD pumps and axial/centrifugal pumps as effectively as a hardware test. Figure 4, for example, compares the predicted efficiency for a centrifugal pump and shows excellent correlation with measured data [2].

Figure 4. CFD simulation of a centrifugal pump

CFD is also used to analyze cavitation effects-perhaps even more effectively than a hardware test because CFD enables the analyst to easily visualize the cause and location of bubble formation. Cavitation can have a detrimental effect on efficiency, and it can also affect pump life through damage caused by high pressures generated as vapor pockets collapse. Cavitation has historically been a difficult problem to model, because the high density ratios between the liquid and gas can cause numerical convergence problems and accuracy issues. As a result, many codes only model up to the onset of cavitation. However, recent advances in CFD now make it possible to model robustly deep cavitation, predicting the instantaneous volume fraction of vapor and non-condensable gases (e.g. air) throughout the pump [1, 3].
Figure 5 shows a correlation between the locations of predicted cavitation (indicated by the crescent -shaped red spot inside the yellow circle) in an axial piston pump and the actual damage observed during pump testing [4]. Given the long duration it sometimes takes for cavitation damage to appear during testing or operation, numerical prediction can be a much more efficient means of assessing possible cavitation damage.

Figure 5. CFD prediction of cavitation in an axial piston pump

All the simulations in the examples above were accomplished by applying fundamental principles (conservation of mass and momentum) coupled with physical models for cavitation and/or turbulence models. These basic principles and models have been available for decades. What has changed is their implementation using more sophisticated algorithms. CFD codes for pumps and compressors have now evolved to the point where a detailed 3-D model can be run in less than a day and produce accurate engineering data that compare well with experiments over a wide range of operating conditions.

References

[1] Y. Jiang and D. Zhang, "A Three-Dimensional Design Tool for Crescent Oil Pumps," 2008 SAE conference, Detroit, Michigan.
[2] WANG Xiu-yong,WANG Can-xing, Zhejiang University, Hangzhou 310027, China "Performance Prediction of Centrifugal Pump Based on the Method of Numerical Simulation," Fluid Machinery 2007, volume 35 (10), page 9-13.
[3] Singhal, A. K., Athavale, M. M., Li H., and Jiang, Y. 2002. Mathematical Basis and Validation of the Full Cavitation Model, Journal of Fluid Engineering, Vol. 124, Issue 3, pp. 617-624, 2002
[4] O. Meincke and R. Rahmfeld, "Measurements, Analysis and Simulation of Cavitation in an Axial Piston Pump", 6th International Fluid Power Conference, Dresden, 2008.
Samuel A. Lowry is president of Simerics Inc., 303 Williams Ave., Suite 123, Huntsville, AL 35801, 256-489-1480, ext. 111.
http://pump-zone.com/compressors/compressors/flow-simulation-in-pumpcompressor-design.html

PUMP AND PUMP SYSTEM GLOSSARY

[table]
[TR]
[TD]Absolute pressure: pressure is measured in psi (pounds per square inch) in the imperial system and kPa (kiloPascal or bar) in the metric system. Most pressure measurements are made relative to the local atmospheric pressure. In that case we add a "g" to the pressure measurement unit such as psig or kPag. The value of the local atmospheric pressure varies with elevation It is not the same if you are at sea level (14.7 psia) or at 4000 feet elevation (12.7 psia). In certain cases it is necessary to measure pressure values that are less then the local atmospheric pressure and in those cases we use the absolute unit of pressure, the psia or kPa a.


pa(psia) = pr(psig) + patm(psia), patm = 14.7 psia at sea level.

where pa is the absolute pressure, pr the relative pressure and patm the absolute pressure value of the local atmospheric pressure.

and in the metric system

pa(kPa a) = pr(kPag) + patm(kPa a), patm = 100 kPa a at sea level.

[/TD]
[TD][/TD]
[/TR]
[/table]


[hr][/hr] [table]
[TR]
[TD]Accumulator: used in domestic water applications to stabilize the pressure in the system and avoid the pump cycling on and off every time a tap is opened somewhere in the house. The flexible bladder is pressurized with air at the pressure desired for acheiving the correct flow rate at the furthest point of the house or system. As water is pulled from the tank the bladder expands to fill the volume and maintain the pressure. When the bladder can no longer expand the water pressure drops, the pressure switch of the pump is activated on low pressure, and the pump starts and fills the water volume of the accumulator. The bladder keeps the air from entering into solution with the water resulting in less frequent re-pressurisation of the accumulator.
[/TD]
[TD][/TD]
[/TR]
[/table]

[table]
[TR]
[TD][/TD]
[TD][/TD]
[/TR]
[/table]

Pumps are often sold as a package with an accumulator.
[hr][/hr] Affinity laws: the affinity laws are used to predict the change in diameter required to increase the flow or total head of a pump. They can also predict the change in speed required to achieve a different flow and total head. The affinity laws can only be applied in circumstances where the system has a high friction head compared to the static head and this is because the affinity laws can only be applied between performance points that are at the same efficiency. see affinity laws.pdf

The following figure shows a system that has a friction head (curve A) higher than its static head for which the affinity laws apply, as compared to curve B, a system with a high static head as compared to the friction head where the affinity laws do not apply.

Domain of application of the affinity laws for an axial flow pump.

The affinity laws are expressed by the three following relationships where Q is the flow rate, n the pump rpm, H the total head and P the power. You can predict the operating condition for point 2 based on the knowledge of the conditions at point 1 and vice versa.



The process of arriving at the affinity laws assumes that the two operating points that are being compared are at the same efficiency. The relationship between two operating points, say 1 and 2, depends on the shape of the system curve (see next Figure). The points that lie on system curve A will all be approximately at the same efficiency. Whereas the points that lie on system curve B are not. The affinity laws do not apply to points that belong to system curve B. System curve B describes a system with a relatively high static head vs. system curve A which has a low static head.


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CENTRIFUGAL PUMP SYSTEM TUTORIAL 2

What is friction in a pump system (cont.) Another cause of friction is all the fittings (elbows, tees, y's, etc) required to get the fluid from point A to B. Each one has a particular effect on the fluid streamlines. For example in the case of the elbow, the fluid particles that are closest to the tight inner radius of the elbow lift off from the pipe surface forming small vortices that consume energy. This energy loss is small for one elbow but if you have several elbows and other fittings the total can become significant. Generally speaking they rarely represent more then 30% of the total friction due to the overall pipe length.

Figure 9
[hr][/hr] Energy and head in pump systems
Energy and head are two terms that are often used in pump systems. We use energy to describe the movement of liquids in pump systems because it is easier than any other method. There are four forms of energy in pump systems: pressure, elevation, friction and velocity.


Pressure is produced at the bottom of the reservoir because the liquid fills up the container completely and its weight produces a force that is distributed over a surface which is pressure. This type of pressure is called static pressure. Pressure energy is the energy that builds up when liquid or gas particles are moved slightly closer to each other and as a result they push outwards in their environment. A good example is a fire extinguisher, work was done to get the liquid into the container and then to pressurize it. Once the container is closed the pressure energy is available for later use.

Elevation energy is the energy that is available to a liquid when it is at a certain height. If you let it discharge it can drive something useful like a turbine producing electricity.

Friction energy is the energy that is lost to the environment due to the movement of the liquid through pipes and fittings in the system.

Velocity energy is the energy that moving objects have. When a baseball is thrown by a pitcher he gives it velocity energy also called kinetic energy. When water comes out of a garden hose, it has velocity energy.



Figure 9a

In the figure above we see a tank full of water, a tube full of water and a cyclist at the top of a hill. The tank produces pressure at the bottom and so does the tube. The cyclist has elevation energy which he will be using as soon as he moves.


As we open the valve at the tank bottom the fluid leaves the tank with a certain velocity, in this case pressure energy is converted to velocity energy. The same thing happens with the tube. In the case of the cyclist, the elevation energy is gradually converted to velocity energy.


The three forms of energy: elevation, pressure and velocity interact with each other in liquids. For solid objects there is no pressure energy because they don’t extend outwards like liquids filling up all the available space and therefore they are not subject to the same kind of pressure changes.


The energy that the pump must supply is the friction energy plus the elevation energy.


PUMP ENERGY = FRICTION ENERGY + ELEVATION ENERGY


Figure 9b

You are probably thinking where is the velocity energy in all this. Well if the liquid comes out of the system at high velocity then we would have to consider it but this is not a typical situation and we can neglect this for the systems discussed in this article.

The last word on this topic, it is actually the velocity energy difference that we would need to consider. In figure 9c the velocities at point 1 and point 2 are the result of the position of the fluid particles at points 1 and 2 and the action of the pump. The difference between these two velocity energies is an energy deficiency that the pump must supply but as you can see the velocities of these two points will be quite small.

Now what about head? Head is actually a way to simplify the use of energy. To use energy we need to know the weight of the object displaced.


Elevation energy E.E. is the weight of the object W times the distance d:


EE = W x d


Friction energy FE is the force of friction F times the distance the liquid is displaced or the pipe length l:


FE = F x l


Head is defined as energy divided by weight or the amount of energy used to displace a object divided by its weight. For elevation energy, the elevation head EH is:


EH = W x d / W = d


For friction energy, the friction head FH is the friction energy divided by the weight of liquid displaced:


FH = FE/W = F x l / W (see Figure 9b)


The friction force F is in pounds and W the weight is also in pounds so that the unit of friction head is feet. This represents the amount of energy that the pump has to provide to overcome friction.



I know you are thinking this doesn’t make sense, how can feet represent energy?

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CENTRIFUGAL PUMP SYSTEM TUTORIAL 1

What is total head
Total head and flow are the main criteria that are used to compare one pump with another or to select a centrifugal pump for an application. Total head is related to the discharge pressure of the pump. Why can't we just use discharge pressure? Pressure is a familiar concept, we are familiar with it in our daily lives. For example, fire extinguishers are pressurized at 60 psig (413 kPa), we put 35 psig (241 kPa) air pressure in our bicycle and car tires.For good reasons, pump manufacturers do not use discharge pressure as a criteria for pump selection. One of the reasons is that they do not know how you will use the pump. They do not know what flow rate you require and the flow rate of a centrifugal pump is not fixed. The discharge pressure depends on the pressure available on the suction side of the pump. If the source of water for the pump is below or above the pump suction, for the same flow rate you will get a different discharge pressure. Therefore to eliminate this problem, it is preferable to use the difference in pressure between the inlet and outlet of the pump.

The manufacturers have taken this a step further, the amount of pressure that a pump can produce will depend on the density of the fluid, for a salt water solution which is denser than pure water, the pressure will be higher for the same flow rate. Once again, the manufacturer doesn't know what type of fluid is in your system, so that a criteria that does not depend on density is very useful. There is such a criteria and it is called TOTAL HEAD, and it is defined as the difference in head between the inlet and outlet of the pump.

You can measure the discharge head by attaching a tube to the discharge side of the pump and measuring the height of the liquid in the tube with respect to the suction of the pump. The tube will have to be quite high for a typical domestic pump. If the discharge pressure is 40 psi the tube would have to be 92 feet high. This is not a practical method but it helps explain how head relates to total head and how head relates to pressure. You do the same to measure the suction head. The difference between the two is the total head of the pump.

Figure 25

The fluid in the measuring tube of the discharge or suction side of the pump will rise to the same height for all fluids regardless of the density. This is a rather astonishing statement, here's why. The pump doesn’t know anything about head, head is a concept we use to make our life easier. The pump produces pressure and the difference in pressure across the pump is the amount of pressure energy available to the system. If the fluid is dense, such as a salt solution for example, more pressure will be produced at the pump discharge than if the fluid were pure water. Compare two tanks with the same cylindrical shape, the same volume and liquid level, the tank with the denser fluid will have a higher pressure at the bottom. But the static head of the fluid surface with respect to the bottom is the same. Total head behaves the same way as static head, even if the fluid is denser the total head as compared to a less dense fluid such as pure water will be the same. This is a surprising fact, see this .


For these reasons the pump manufacturers have chosen total head as the main parameter that describes the pump’s available energy.

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[QUOTE=Esam;176603]

CONNECTING and INTERPRETING LIMIT SWITCHES

© Walter ********, P. Eng., 2000 May 20. walter(at)********(dot)ca

First published in Intech , January 1993 as "Limit Switches Key to Valve Reliability"

This Adobe® file is available for download.

[IMG]http://www.********.ca/limitsw/LS-1.jpg[/IMG]INTRODUCTION. There is a great variety of possible combinations for installing and connecting limit switches on valves. The number of switches depends on the particular control objective and may be influenced by redundancy considerations. The way they are connected depends on the safety and reliability requirements.
In order to clarify this discussion, diagrams like Figure 1 will be used. All signals, switch positions, etc. are shown with the valve at the center of travel. No limit switches are actuated, all are shown in their shelf position as determined by their internal springs. Imagine the valve to be like a guillotine where the stem travels upward to open the valve and downwards to close it. The diagrams show the switches connected to indicating light bulbs but the logic is identical if a DCS or other form of MMI is used.
The limit switch that is actuated when the valve is fully open is labeled ZSO. The one at the extreme opposite end is labeled ZSC.
The terminals on the electrical switches are labeled Common (C), Normally Open (NO), and Normally Closed (NC). This unfortunate choice of terminology has nothing to do with the state of the valve nor even the "normal" position of the switch. It refers to the state of the switch when nothing is pushing on it.

[IMG]http://www.********.ca/limitsw/LS-2.jpg[/IMG]SINGLE SWITCH, DIRECT APPROACH. A single limit switch at the OPEN end of the valve (ZSO), as shown in Figure 2, will tell us when the valve is fully open. It cannot tell us if the valve is fully closed. The problem is that the term "open" is a bit ambiguous. Question: Is a half-open valve open, closed, neither open nor closed, or both open and closed? This discussion will use the following definitions:
OPEN = Partly or fully OPEN
CLOSED = Partly or fully CLOSED
Not OPEN = Fully CLOSED
Not CLOSED = Fully OPEN
According to these definitions the half-open valve is both open and closed. A single ZSO switch can only tell us if the valve is "fully open" and "not closed". It cannot tell us if the valve is partly open.
Example 1: We need a limit switch and a status light to tell the operator that the fuel gas to a furnace is OPEN. If so, it is not safe to begin the light-off sequence. A ZSC switch at the closed end of travel is used so that we can be sure the valve is "fully closed" and "not open" even a little bit. The correct contact is NC. If the valve is even the slightest bit open, the OPEN light comes on.
Example 2: We need a limit switch and a status light to tell the operator that the fuel gas to a furnace is CLOSED. If so, it is safe to begin the light-off sequence. This is exactly the same limit switch as before: ZSC. We want to know if the valve is "fully closed". The only difference is that Example 1 uses the NC contact of the switch to turn off an OPEN light when the valve is not fully open while Example 2 uses the NO contact to turn on a CLOSED light when the valve is fully closed.

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Optimum Settings For Automatic Controllers
the American Society of Mechanical Engineers, 1942.

First published in "Transactions of the A. S. M. E.", November 1942.

An exact facsimile on Adobe ® is available for download.
"Optimum Settings For Automatic Controllers" – is the name of one of the most important publications in the history of automation, instrumentation, and control systems. Written by Ziegler and Nichols and published in the November 1942 issue of the Transactions of the American Society of Mechanical Engineers, it gave the "hit and miss" art of tuning controllers a practical basis. Ziegler and Nichols developed and published tuning rules for pneumatic PID controllers while working in Rochester, NY for Taylor Instruments, now a part of ABB. Although developed for pneumatic controllers, their rules are still widely used as a comparison for other methods. When Nichols died in April 1997, at age 82, and Ziegler not long after on December 9, at age 88, a chapter in industrial automatic control came to an end.
The ASME graciously granted permission for me to reproduce "Optimum Settings for Automatic Controllers", © 1942, on my website, walter(at)********(dot)ca. To generate this reproduction, a photocopy of the printed original has been scanned, run through an optical character recognition (OCR) program, formatted in MS Word, and edited to duplicate the original as closely as possible. The diagrams were scanned and embedded in the Word file in .jpg format. Finally, this MS Word "forgery" was converted into .html and .pdf formats.

I thank the ASME for allowing this important work to be made so freely available. The society can be reached at www.asme.org. I would also like to thank my daughter, Mika ********, for her patience with the OCR work.

Optimum Settings for Automatic Controllers

By J.G. ZIEGLER¹ and N. B. NICHOLS² • ROCHESTER, N. Y.

In this paper, the three principle control effects found in present controllers are examined and practical names and units of measurement are proposed for each effect. Corresponding units are proposed for a classification of industrial processes in terms of the two principal char-acteristics affecting their controllability. Formulas are given which enable the controller settings to be determined from the experimental or calculated values of the lag and unit reaction rate of the process to be controlled. These units form the basis of a quick method for adjusting a controller on the job. The effect of varying each controller setting is shown in a series of chart records. It is believed that the conceptions of control presented in this paper will be of assistance in the adjustment of existing controller applications and in the design of new installations.
A purely mathematical approach to the study of automatic control is certainly the most desirable course from a standpoint of accuracy and brevity. Unfortunately, however, the mathematics of control involves such a bewildering assortment of exponential and trigonometric functions that the average engineer cannot afford the time necessary to plow through them to a solution of his current problem.
It is the purpose of this paper to examine the action of the three principal control effects found in present-day instruments, assign practical values to each effect, see what adjustment of each does to the final control, and give a method for arriving quickly at the optimum settings of each control effect. The paper will thus first endeavor to answer the question: "How can the proper controller adjustments be quickly determined on any control application?" After that a new method will be presented which makes possible a reasonably accurate answer, to the question: "How can the setting of a controller be determined before it is installed on an existing application?"
Except for a single illustrative example, no attempt will be made to present laboratory and field data, to develop mathematical relations, or to make acknowledgment of material from published literature. A paper covering the mathematical derivations would be quite lengthy as would also a paper covering laboratory and field-test results. Work on these phases of the subject is still under way, and it is expected that the results will be published at a later time when convenient. It is believed advisable to publish the present paper without delay in order to make the information available for use by the many persons interested in the application of automatic-control instruments. To these persons the present subject matter is of much greater interest than the other phrases of the study which are being omitted.
To simplify terminology wewill take the mostcommon type of control circuit in which a controller interprets the movement of its recording pen into a need for corrective action, and, by varying its output air pressure, repositions a diaphragm-operated valve. The controller may be measuring temperature, pressure, level, or any other variable, but we will completely divorce the measurement portion of the control circuit and speak only of the pen movement in inches; 1 in. of pen movement might represent 1 or 1000 deg F, or a flow of 1 or 1000 gpm. The actual graduation will be of no moment in a study of control.
Ourcontroller will translate pen behavior into behavior of a valve; the relation between the two behavior patterns is determined by the setting of each control effect. The term valve covers any similar device, i.e., a damper or rheostat which must be operated by the controller in order to maintain correct process conditions.

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Originally Posted by Esam

THE COMPRESSOR MONITORING SKETCH

© Walter ********, P. Eng., 2000 August 20. walter(at)********(dot)ca

First published in Intech , July 1990

This Adobe® file is available for download.


INTRODUCTION. Review meetings between mechanical engineers, control systems engineers and equipment manufacturing representatives do not always provide a final resolution of all outstanding technical issues relating to the precise scope of supply. Without some means of focusing the discussion a number of scenarios are possible.
Scenario number 1: Everybody thinks they understand.
A typical discussion may run like this:
Equipment Engineer, "Where will you provide bearing monitoring?"
Vendor Representative, "On all the bearings."
The control systems engineer writes this down in his note book and goes off to list and tag all the monitoring points. Some time later he finds out that there were more bearings than anyone thought, that axial probes appear on both sides of the thrust bearings and the key phasor is missing. At this point he realizes that his monitoring package is too small and that the proper monitor will not fit into the compressor control panel that has been ordered.
Scenario number 2: Everybody is totally mystified.
A typical discussion may run like this:
Equipment Engineer, "Where will you provide bearing monitors?"
Vendor Representative, "Radial X/Y probes will be provided on both the inboard and the outboard bearings but axial probes only on the active side of the outboard bearings."
Equipment Engineer, "Are those on the driven end or on the opposite from driven end?"
Controls Engineer, "Is that on the north end or the south end of the compressor?"
Vendor Representative, "It's on the high speed shaft of the gear box.'
Controls Engineer, "Are the RTDs SAMA or DIN calibration?"
Vendor Representative, "They are all type J thermocouples except in the motor windings where they are 10 Ohm copper. The gear box is made in Europe so they are probably European type RTDs."
Scenario number 3: Everybody understands what is wanted.
The discussion starts off like either scenario number 1 or 2. Then the controls engineer (our hero!) goes up to the blackboard and draws out the sketch shown below. He shows an outline of all major pieces of equipment: The motor, the gear box and the compressor. Every bearing is identified. The equipment engineer first finds out there is a bearing he didn't know about. The vendor representative finds out that axial probes are wanted on both sides of the thrust bearings. The controls engineer finds out that there is no accelerometer on the gear box. And everybody has a great lunch at the vendor's expense with the satisfaction of a job well done.
Once instrument tags and details are added, the Compressor monitoring sketch can serve as a basis for a Process and Instrumentation Diagram. Later, it can be included as a graphic display on a DCS or it can be engraved on laminated plastic and attached to the local compressor panel below the vibration and temperature monitors.

[IMG]http://www.********.ca/vibetemp/VT-1.jpg[/IMG]

The author has never experienced any confusions or misunderstanding where this sketch has been used. It certainly saves on long distance conference calls.

[QUOTE=Esam;176596]CONTROLLING FIRED HEATERS © Walter ********, P. Eng., 2000 May 20. walter(at)********(dot)ca

First published in Hydrocarbon Processing , April 1997.

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[IMG]http://www.********.ca/ce5_fh/5-1.jpg[/IMG]INTRODUCTION. The purpose of a fired heater is very simple: To add heat to a process fluid. Its representation on a process flow diagram is also very simple. But, of course, fired heaters are among the most complex pieces of process control equipment. Each furnace is, after all, at least two pieces of equipment in one. Firstly, it is a special variant of the shell and tube heat exchanger since its purpose is to exchange heat. Secondly, it is a chemical reactor in which fuel and air undergo extremely exothermic reactions to produce the required heat.

In previous articles of this series^{1, 2, 3, 4}, the process aspects of controlling a piece of equipment were presented before dealing with protection and safety. This time the topics will be reversed: In the case of fired heaters, it must be safety first!
SAFETY. If fired heaters had not been invented and were being proposed for the first time, I would probably say, "You've got to be kidding. That thing will blow up in your face the first time you throw a match in it." However, at least a half a billion gas fired heaters are in service around the world (according to the American Gas Association). Most of them are operated by people with no technical experience whatsoever; few heaters blow up. Still, the average domestic water heater is not in the same league as a hydrogen reformer furnace. The fact that accidents and disasters are as few as they are, is due to the long experience the human race has in dealing with fire. A million years, I'm told. For the last century, this experience has been embodied in various codes and standards that have been written into law and are en-forced by inspectors around the world.
THE CODE. The most popular, or notorious, of these codes in North America is NFPA 85⁵ and 86⁶ issued by the National Fire Protection Association. These have been considerably updated in recent years, especially in terms of clarity. Nevertheless, there is still the problem of interpretation. The code is not at all easy to read as it combines many facets of construction, instrumentation and operation in a single document. Not only that, but the code⁵ contains the following disclaimers:
It is not possible for these standards to encompass specific hardware applications, nor should these be considered a "cookbook" for the design of safety systems.
and:
This standard applies to boilers with a fuel input of 12,500,000 Btu/hr (3663 kW) or greater. This standard applies only to boiler-furnaces using single burners firing:
a) Natural gas only as defined in Chapter 3.
b) Other gas with a BTU value and characteristics similar to natural gas.
c) Fuel oil of No. 2....
and:
Furnaces such as those of process heaters used in chemical and petroleum manufacture, wherein steam generation is incidental to the operation of a processing system, are not covered in this standard.
What is an engineer to use for a guide when the furnace is not a boiler, but a feed heater; does not exceed 12½ million Btu/hr, but is only four million; does not burn natural gas as defined in Chapter 3; but refinery off-gas with a high hydrogen content? Despite the disclaimers, the NFPA series is still an excellent guide to the instrumentation and control of any furnace.

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[QUOTE=Esam;176591]

CONTROLLING STEAM HEATERS

© Walter ********, P. Eng., 2000 May 20. walter(at)********(dot)ca

First published in Hydrocarbon Processing , November 1996.

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INTRODUCTION. Steam Heaters are simply heat exchangers in which one of the media is steam being condensed while the other is a process fluid being heated. In doing this, there is a phase change which puts special demands on the process control system. It is difficult to generalize about the various options for control. Special system requirements often put unexpected constraints on the process. Even the orientation of the exchanger can have peculiar and unexpected results.
[IMG]http://www.********.ca/ce4_sh/4-1.jpg[/IMG]A SIMPLE STEAM SPACE HEATER. Figure 4-1 shows a steam heater such as those used to heat a warehouse. This simple example demonstrates many of the characteristics of steam heaters of all sizes and applications. Steam enters the heater at the top. As the moving air draws away the heat, the steam condenses. The condensate flows down the tubes, through the steam trap, and into the condensate drain header.
The function of the steam trap is to prevent steam from blowing through into the condensate system. It is the one essential part of any steam heater and will receive further attention later. For now it is sufficient to say only that it passes condensate and blocks steam.
This system tends to be rather self-regulating. The moving air rises to some temperature approaching that of the steam and draws away as much heat as it can. Colder air will draw more heat, and warm air will draw less. The steam trap is essentially a level controller with a set point of zero.
This arrangement can be compared to a shell and tube exchanger where the room itself is the shell and the air is the process stream. The fan draws some of the air through the heater and then blends it with the remaining air in the room. The first level of control complexity is to add a thermostatic switch to control the fan. As with any exchanger on bypass control, the sensing element must be placed at a point where the two stream has mixed sufficiently to provide an representative temperature (not directly in front of the fan, as the drawing shows). When the temperature in the room reaches the setpoint, the fan will stop and the air immediately around the tubes will rise to the steam temperature. The heat withdrawn will be reduced until only a small amount of steam is condensed.
If it were practical to stop all air circulation and to fully insulate the heater so that no heat is transferred out of it, steam condensation would cease and no condensate would flow through the steam trap. This is not practical, so on a hot day any amount of steam that still is condensed by air convection is a complete waste. Furthermore it adds to the heat in the room. Thus the next level of complexity is to block the steam to the heater. When this is done, the steam already in the heater condenses, the temperature drops to room temperature and the pressure drops to the corresponding vapour pressure. Condensate will not flow through the trap once the pressure drops below that of the condensate header. Because of the higher density of water, a given volume of steam condenses to a much smaller volume of condensate. The final equilibrium is reached with a pressure of about 2.8 kPa_abs (0.4 psia), essentially full vacuum, and with the tubes about 0.15% full of water. (The steam supply in this example is assumed to be at 170 kPa_ga (25 psig), fully saturated.)
The simple system described above, minus the fan, is used for many non-process heating applications such as steam tracing or open tank heating.
STEAM TRAPS. As steam condenses, the resulting water drains downward. A steam trap is placed at the low point of the system. It is a valve that opens to allow the water to drain out into the condensate system but closes when all the water has been drained and steam tries to pass through. There are numerous varieties of steam traps operating on various principles. A detailed discussion of various types can be found in the article Steam Traps, Key to Process Heating¹ by Haas.
CONTROLLING A PROCESS HEATER. The parameter of interest in any process heater is the temperature of the process stream at some particular point in the process. There are essentially only three means of control:
· Bypass a fraction of the process stream around the exchanger and blend it with the fraction that has passed through.
· Vary the effective surface area of heat exchange. This is accomplished by restricting the outlet and partially flooding the exchanger with condensate.
· Vary the temperature of the heating medium. This is accomplished by throttling the steam and dropping the pressure of the steam in the exchanger.
Each of these is discussed in turn below.


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CONTROLLING POSITIVE DISPLACEMENT PUMPS

© Walter ********, P. Eng., 2000 May 20. walter(at)********(dot)ca

First published in Hydrocarbon Processing , May 1996.


[IMG]http://www.********.ca/ce2_pdp/2-1.jpg[/IMG]
INTRODUCTION. The positive displacement pump is in some ways an even simpler device to control than the centrifugal pump discussed previously¹. It has the same function, namely to provide the pressure necessary to move a liquid at the desired rate from point A to point B of the process. Figure 2-1 shows a 'generic' process with a positive displacement pump (in this case a gear pump) connected to deliver liquid from A to B.
There is a great variety of positive displacement pumps. They are divided into two broad categories: Rotary and reciprocating. From the controls point of view, however, they are all similar. Their characteristic curve is so simple that it is rarely drawn. It is essentially a straight vertical line, as shown in Figure 2-2. (For some reason PD pump curves are usually shown with the pressure and flow axis exchanged. I will not follow that convention in this article.) All are constant flow machines whose pressure rises to whatever value is necessary to put out the flow appropriate to the pump speed. If the discharge is blocked, the pressure will rise until something yields -- preferably a relief valve. Close examination of the curve shows a slight counter clockwise rotation. This is due to internal leakage.
[IMG]http://www.********.ca/ce2_pdp/2-2.jpg[/IMG]For positive displacement pumps the major cause of leakage is the small amount of reverse flow that occurs before a check valve closes and possibly past the check valve after it is closed. Leakage past the piston is negligible. Diaphragm operated PD pumps have no cylinder to leak past. Rotating PD pumps, such as gear pumps or progressing cavity pumps have internal clearances which permit a small reverse flow, called "slip" or "blowby". There is another reason why the curve may rotate to slightly lower flows at higher discharge pressures: The driver may slow down as the load increases. None of these have a significant affect in curving the slope of the characteristic enough that this slope can be used for control. For most practical purposes the slope is vertical. The system curve of the process is also shown on Figure 2-2. Its intersection with the pump characteristic defines the operating point.
As always, the process controls engineer has the responsibility of matching the capacity of a specific piece of equipment to the demands of the process at every instant in time. Rarely does the actual system curve fall exactly on the one used for design and selection. As with any two port device, there are three locations in which a control valve can be placed: On the discharge, on the suction, and as a recycle valve.
[IMG]http://www.********.ca/ce2_pdp/2-3.jpg[/IMG]DISCHARGE THROTTLING. Discharge throttling does not work! Looking at the process from the point of view of the pump, discharge throttling rotates the system curve counter clockwise so that the modified system curve intersects the pump curve higher up. The additional pressure is dropped through the valve so that the pressure and flow to the process is (almost) exactly the same as before. The "almost" is due the small increase in internal leakage that results in an equally small reduction in flow. An increased wear rate and a shortening of the life of the machine are the only results of this approach. If the pump is seen from the point of view of the process so that the valve is considered part of the pump, the same result is obtained. To obtain a modified pump characteristic curve, the pump curve must be rotated clockwise around the intersection with the pressure axis. The problem is that this hypothetical intersection is far off the top of the operating range. It is the point where the pressure is so high that 100% internal leakage occurs. The machine would self-destruct from excess pressure if one were stubborn enough to attempt to find this point. The rotation of the curve can still be performed on paper and it amounts to a slight shift to the left. Shown in Figure 2-3, it is virtually identical to the unmodified curve. To cut a long story short, you can't control a PD pump with discharge throttling.

[IMG]http://www.********.ca/ce2_pdp/2-4.jpg[/IMG]
SUCTION THROTTLING. Suction throttling has the same effect on the characteristic curve as discharge throttling and doesn't work either. PD pumps have a Net Positive Suction Head Required (NPSHR) just as centrifugal pumps do. In fact their requirements are even more stringent. Therefore restrictions and pressure drops in the suction lines must be similarly avoided.


RECYCLE CONTROL. This leaves recycle control as the only means of using a valve to control a PD pump. The valve is installed in a line teeing off from the discharge and leading back to the source of the liquid, possibly a surge tank. It must be fail open , of course. Figure 2-5 shows its effects on the characteristic curves. Viewing the process from the point of view of the pump, its effect is to rotate the system curve clockwise around its intersection with the pressure axis. Note that the little "tail" at the bottom left of the modified system curve is due to the flow through the recycle valve before the discharge check valve has opened. The flow through the pump is essentially as before but the pressure to the process has been reduced. Process flow will, of course, also be reduced by the amount flowing through the recycle line.
[IMG]http://www.********.ca/ce2_pdp/2-5.jpg[/IMG]Viewing the pump from the process gives a different perspective on the same phenomenon. This time it is the pump curve that is rotated counter clockwise around its intersection with the flow axis. This modified pump curve gives the effect of greatly increased internal leakage. From the point of view of the process, this is exactly what is happening. Note that I have not used the same operating points in Figure 2-3 as I did in Figure 2-5. It is simply impossible to show any significant reduction in flow on a curve representing the effects of discharge throttling.
Recycle control is an efficient method of control for PD pumps. Since the flow rate is essentially constant, the power requirement is roughly proportional to discharge pressure. Since the effect of recycle is to drop the discharge pressure, it results in significant reductions in power requirement. Nevertheless there is still wasted power in proportion to discharge pressure times recycle flow.
Recycle valves experience rather severe service if the pressure drop is high. Cavitation will destroy them if they are not appropriately selected. Two approaches exist to deal with this problem: The first solution is to drop the pressure in many small stages through the use of many twists and turns in the valve trim. The second is to tolerate the resulting cavitation by shooting the liquid as a jet through a small hole in the middle of a disk. The jet then blasts directly into the discharge piping. The line diameter is often increased immediately downstream of the valve and the wall thickness is also increased. In this way the jet cavitates down the middle of the pipe. It makes a terrific racket.
In either case it may be necessary to put a fixed restriction downstream of the valve. It should be sized so that the ratio of the high to intermediate pressure is the same as the ratio of intermediate to low pressure. Keep in mind that the restriction will reduce the rangeability of the valve by making it act like a quick opening valve. This is because the restriction becomes the dominant factor in the line once the valve is about half way open. From that point on, the valve has little control.
Recycle lines for PD pumps should be run back to the suction vessel. This allows any entrained bubbles to escape. If they do not, they can build up to the point where pump capacity is impaired. It may even vapour lock.


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Controlling positive displacement pumps

CONTROLLING VESSELS and TANKS © Walter ********, P. Eng., 2001 Sept 07. walter(at)********(dot)ca

First published in Hydrocarbon Processing , July 1995.


INTRODUCTION. It would seem that controlling a vessel should be a very simple matter -- They don't really do anything! But then, if they didn't do anything why are there so many of them? And why do they have so many different names? Going through a typical set of Piping and Instrumentation Diagrams (P&IDs) I see the following vessels:

· Degassing Drum · Gas Separator · Storage Tank
· Feed Flash Drum · Reflux Accumulator · Day Tank
· Surge Drum · Suction Scrubber · Slug Catcher
· Lube Oil Separator · Head Tank · Deaerator

Although each of these is essentially a simple vessel or tank without any special internal structure, each serves a different purpose. Once it is clear what the purpose of a piece of equipment is, and how it functions, it will also be clear how to control and protect it. Different purposes require different controls.
SURGE TANKS. The most common function of a vessel or tank is to match two flows that are not identical in time but are expected to average out over the long run. Take a feed surge drum, for example. Flow into the unit is more or less steady but is subject to interruption. The flow to the processing unit should be as constant as possible, avoiding sudden change. Nevertheless, it, too, may be subject to interruption due to downstream conditions.
The purpose of the surge drum is to maintain sufficient inventory to feed the process and to maintain sufficient void capacity to continue receiving feed as it arrives. Clearly the tank must be large enough to accommodate any normal discrepancies between input and output over a reasonable period of time. Between the upper and lower bound, the exact value of the level does not matter.
Two separate control parameters are implied: Level and flow. Level control is no problem. Greg Shinskey ¹ refers to "The easy element -- capacity". A high gain, level controller connected to a valve at either the inlet or the outlet will maintain the level very accurately at its setpoint. The only problem with this approach is that it absolutely defeats the purpose of the vessel. The same effect would be achieved by blocking in the vessel and bypassing the inlet directly to the outlet.
To control flow alone is also quite simple. A flow controller at the outlet, properly tuned, will maintain a steady flow to the process. Unfortunately, there is nothing to make this flow equal to inflow. It will not even equal the average inflow unless there is something to make it do so.
What we need is an instrument that measures the accumulated error between inflow and outflow. The tank itself is that instrument!

Level = Starting Level + [IMG]http://www.********.ca/ce6_v&t/int.jpg[/IMG] (Inflow - Outflow) dt / Tank Area

(To a process controls engineer, every piece of equipment is just a big, non-tuneable instrument!) The level transmitter only transmits the process value to the control system. If we now cascade the output of the level controller to the flow controller, we have a system that has one process variable: Accumulated flow imbalance. It has only one point of control: Outflow to the process.
To start this simple process:
[IMG]http://www.********.ca/ce6_v&t/6-1.jpg[/IMG]· Fill the tank about half full.
· Give the level controller the current level as its set point. (PV tracking does this automatically.)
· Switch the flow controller to automatic with an estimated average flow as its setpoint.
· Switch the flow controller to cascade.
· Switch the level controller to automatic.

The control system will keep the flow "constant" but that constant varies in response to the imbalance between outflow and inflow. It is not important that the initial estimate of average flow be exact. A low guess will result in the tank level rising a little. A new, higher, estimate will result and the outflow will be adjusted accordingly. In the long term the average flow out is not an independent variable at all. It will be exactly equal to the average flow in. This can be accomplished at any arbitrary tank level. The level setpoint is based on the operator's estimate of the nature of the flow interruptions and whether the most probable upset will require additional flow or void capacity.
Should a pump be necessary to transfer the liquid from the vessel to its destination it should be placed between the vessel and the flow measurement. Further information on the control of pumps is found in Controlling Centrifugal Pumps² . This article also includes a section titled "On/off Control" for less critical level applications.
There is a long discussion on the special requirements for level control of steam heat exchangers and condensate receivers in Controlling Steam Heaters³.
Surge drums are sometimes used for gas. The abrupt flow variations of a Pressure Swing Absorption (PSA) unit, for example, often need to be smoothed out before the tail gas can be introduced into a down-stream process. In these cases, pressure takes the role that level has in a liquid process. That is, a pressure/flow cascade is the appropriate solution.
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Thread: CONTROLLING VESSELS and TANKS]1[/h]

CONTROLLING CENTRIFUGAL PUMPS

[IMG]http://www.********.ca/ce1_cp/1-1.jpg[/IMG]INTRODUCTION. The centrifugal pump is one of the simplest pieces of equipment from the controls and instrumentation point of view. It is a two port device with a well defined characteristic. Its purpose is to provide the necessary pressure to move liquid at the desired rate from point A to point B of the process. Figure 1-1 shows a 'generic' process with a centrifugal pump connected to deliver liquid from A to B.
Figure 1-2 shows the characteristic curve of an actual pump (a single stage vertical turbine pump) together with the characteristic curve of the process, known as the system curve. The intersection of the two curves defines the operating point of both pump and process. It would be fortunate indeed if this operating point is the one actually specified for the process. It is impossible for one operating point to meet all desired operating conditions since the operating point is, by definition, exactly one of an infinity of possible operating points. In fact the entire point of controlling the pump is to modify its characteristic so that its actual operating point is the one that is required at every instance in time.

Several definitions are presented in order to discuss the diagram:
[IMG]http://www.********.ca/ce1_cp/1-2.jpg[/IMG]
P_o = Differential pressure, or head, at the operating point of the pump and also of the process.
Q_o = Flow rate, at operating point, of the pump and also of the process.
P_pm = Maximum differential pressure across the pump (at shutoff).
Q_pm = Maximum discharge flow of the pump.
P_lm = Static (Minimum) differential pressure between points B and A of the process.


The minimum static differential pressure of the process is frequently zero, as in a closed, circulating system. If the pump is in parallel with other pumps that are maintaining the system pressure, then P _lm is greater than zero. It is clear from the outset that if P _lm is greater than P_pm, no amount of process control can force the two curves to intersect. The pump is simply inadequate. How is process control like cutting off a rope? You can always cut off more, but you can't cut off less.
Assuming the pump is more than adequate for the process requirements at the moment, what is the best way to trim it back to the desired operating point, P₁, Q₁? There are three possible locations to place a valve: At the discharge, at the suction, and as a recycle valve. Each will be discussed in turn.
DISCHARGE THROTTLING. Since the pump exists to serve the requirements of the process, and one of the primary purposes of instrumentation is to adapt the equipment to the process, let us consider the pump from the point of view of the process. It can be viewed as a constant pressure device with an internal restriction. It is the restriction that gives it the "curve". It seems natural to put a valve on the discharge to further restrict the pump. This has the effect of rotating the curve of the pump/valve system clockwise around P_pm, as can be seen in Figure 1-3.
[IMG]http://www.********.ca/ce1_cp/1-3.jpg[/IMG]At this point I must warn the reader that we are about to encounter a paradigm shift. (!) The combination of pump and valve will be presented as a "black box" with a single characteristic curve which I shall term the "modified" pump curve.
The more traditional way of looking at the situation is from the point of view of the pump. It sees the process system curve as having rotated counter clockwise around P_lm. Figure 1-3 shows that the flow, Q₁, is the same for both cases. The difference between the two pressures is the Delta P across the valve. Since the purpose of the pump is to serve the process requirements, and the purpose of the valve is to adapt the pump to the process, it makes sense to consider the valve to be part of the pump system and to use the modified pump curve rather than the modified system curve in our discussion. In any case it can be seen that a discharge valve can be used to achieve any operating point on the system curve so long as the point is below the pump curve.
SUCTION THROTTLING. The second possibility for control using valves is to place the valve in the pump suction line. This would have an identical effect on the characteristic curve, but the method has a fatal flaw – cavitation. Cavitation is a phenomenon that occurs when the pressure of a liquid is reduced below its vapour pressure and brought back up above the vapour pressure again. Bubbles of vapour form in the liquid and then collapse upon arriving at the higher pressure region. The collapse occurs at sonic speed ejecting minute jets of extremely high velocity liquid. Wherever these jets impinge on a solid surface extreme erosion occurs. Over time even the hardest materials will be destroyed. Therefore it is of utmost importance that this pressure reduction never occurs. It is prevented by having sufficient pressure available at the pump suction so that the pressure drops that occur as the liquid is drawn into the eye of the impeller are at all times above the vapour pressure of the liquid at its current temperature.
An explanation of the term Net Positive Suction Head (NPSH) is in order. This is the pressure of the liquid at the pump suction in terms of feet or meters of liquid head above the vapour pressure of the liquid. The actual NPSH under operating conditions is called NPSHA and the minimum required by the pump to prevent cavitation is called NPSHR. Clearly NPSHA must be greater than NPSHR to avoid cavitation. It is safe to leave a margin of about one meter.
These peculiar definitions are very reasonable in terms of the pumps actual characteristic but they cause some problems to the controls engineer. It means that the gauge pressure equivalent of a given NPSHA is proportional to the density of the liquid and is also affected by its temperature. The vapour pressure can rise dramatically as the temperature rises. This means that the NPSHA can fall without a noticeable change in pressure.
Anything that would reduce the net positive pressure at the pump inlet below the NPSHR must be absolutely avoided. Thus suction throttling is never used to control pump flow.
[IMG]http://www.********.ca/ce1_cp/1-4.jpg[/IMG]RECYCLE CONTROL. The third remaining possibility for pump control with valves is to bleed some of the discharge flow back to the pump suction or to some other point on the supply side. Once again we can view the result as a modified system curve or as a modified pump characteristic. Figure 1-4 shows both. Each curve is a rotation of the original: The modified system curve as a clockwise rotation around P_lm. Note the little "tail" at the left of the modified system curve. This represents the flow through the recycle valve before the discharge check valve opens to the process. The modified pump curve has a counter clockwise rotation around the hypothetical intersection of the pump curve with the flow axis.

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CENTRIFUGAL PUMP SYSTEM TUTORIAL



What is total head

Total head and flow are the main criteria that are used to compare one pump with another or to select a centrifugal pump for an application. Total head is related to the discharge pressure of the pump. Why can't we just use discharge pressure? Pressure is a familiar concept, we are familiar with it in our daily lives. For example, fire extinguishers are pressurized at 60 psig (413 kPa), we put 35 psig (241 kPa) air pressure in our bicycle and car tires.For good reasons, pump manufacturers do not use discharge pressure as a criteria for pump selection. One of the reasons is that they do not know how you will use the pump. They do not know what flow rate you require and the flow rate of a centrifugal pump is not fixed. The discharge pressure depends on the pressure available on the suction side of the pump. If the source of water for the pump is below or above the pump suction, for the same flow rate you will get a different discharge pressure. Therefore to eliminate this problem, it is preferable to use the difference in pressure between the inlet and outlet of the pump.

The manufacturers have taken this a step further, the amount of pressure that a pump can produce will depend on the density of the fluid, for a salt water solution which is denser than pure water, the pressure will be higher for the same flow rate. Once again, the manufacturer doesn't know what type of fluid is in your system, so that a criteria that does not depend on density is very useful. There is such a criteria and it is called TOTAL HEAD, and it is defined as the difference in head between the inlet and outlet of the pump.

You can measure the discharge head by attaching a tube to the discharge side of the pump and measuring the height of the liquid in the tube with respect to the suction of the pump. The tube will have to be quite high for a typical domestic pump. If the discharge pressure is 40 psi the tube would have to be 92 feet high. This is not a practical method but it helps explain how head relates to total head and how head relates to pressure. You do the same to measure the suction head. The difference between the two is the total head of the pump.

Figure 25

The fluid in the measuring tube of the discharge or suction side of the pump will rise to the same height for all fluids regardless of the density. This is a rather astonishing statement, here's why. The pump doesn’t know anything about head, head is a concept we use to make our life easier. The pump produces pressure and the difference in pressure across the pump is the amount of pressure energy available to the system. If the fluid is dense, such as a salt solution for example, more pressure will be produced at the pump discharge than if the fluid were pure water. Compare two tanks with the same cylindrical shape, the same volume and liquid level, the tank with the denser fluid will have a higher pressure at the bottom. But the static head of the fluid surface with respect to the bottom is the same. Total head behaves the same way as static head, even if the fluid is denser the total head as compared to a less dense fluid such as pure water will be the same. This is a surprising fact, see this .


For these reasons the pump manufacturers have chosen total head as the main parameter that describes the pump’s available energy.

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CENTRIFUGAL PUMP SYSTEMS TIPS

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When a viscous fluid is handled by a centrifugal pump

• brake horsepower requirement increases
• the head generated is reduced
• capacity is reduced
• efficiency of pump is reduced and the Best Efficiency Point - BEP - is moved

The head, flow and capacity at other viscosities than used in the original documentation can be modifying with coefficients.
[h=Flow]3[/h]

q_v= c_q q (1)
where
q_v = flow compensated for viscosity (m³/h, gpm)
c_q = viscosity flow coefficient
q = original flow according pump curve (m³/h, gpm)

[h=Head]3[/h]

h_v= c_h h (2)
where
h_v = head compensated for viscosity (m, ft)
c_h = viscosity head coefficient
h = original head according pump curve (m, ft)

[h=Efficiency]3[/h]

μ_v= c_μ μ (3)
where
μ_v = effciency compensated for viscosity
c_μ = viscosity efficiency coefficient
μ = original efficiency according pump curve

[h=Power - SI units]3[/h]

P_v= q_v h_v ρ_v g / (3.6 10⁶ μ_v) (4)
where
P_v = power compensated for viscosity (kW)
ρ_v = density of viscous fluid (kg/m³)
g = acceleration of gravity (9.81 m/s²)

[h=Power - Imperial units]3[/h]

P_v= q_v h_v SG / (3960 μ_v) (5)
where
P_v = power compensated for viscosity (bhp)
SG = specific gravity of viscous fluid

Originally Posted by Esam

A pump does not completely convert the kinetic to pressure energy. Some of the energy is always lost internal and external in the pump.
Internal losses

• hydraulic losses - disk friction in the impeller, loss due to rapid change in direction an velocities through the pump
• volumetric losses - internal recirculation at wear rings and bushes

External losses

• mechanical losses - friction in seals and bearings

The efficiency of the pump at the designed point is normally maximum and is called the

• Best Efficiency Point - BEP

It is possible to operate the pump at other points than BEP, but the efficiency of the pump will always be lower than BEP.