UNIVERSITY OF NAIROBI
COLLEGE OF ARCHITECTURE AND ENGINEERING
SCHOOL OF ENGINEERING
DEPARTMENT OF MECHANICAL AND MANUFACTURING ENGINEERING
PROJECT TITLE:
INSTALLATION AND TESTING OF AN ENGINE DYNAMOMETER
PROJECT NUMBER: JAN 02/2010
A FINAL YEAR PROJECT REPORT SUBMITTED IN PARTIAL FULFILLMENT OF
THE REQUIREMENTS, FOR THE AWARD OF THE BACHELOR OF SCIENCE
DEGREE IN MECHANICAL AND MANUFACTURING ENGINEERING OF THE
UNIVERSITY OF NAIROBI.
BY:
Musyoka K. Chris
F18/2019/2005
Mukuna F. Phillip
F18/2022/2005
Alega Kennedy
F18/2031/2005
SUPERVISED BY: DR NYANG’AYA
2009-2010
i
DECLARATION
We declare that this project is our own original work and has not been presented in this or any
other university for examination or any other purpose.
Musyoka K. Chris.
F18/2019/2005
Signature……………….
Date……………….
Mukuna F Phillip.
F18/2022/2005
Signature…………………
Date…………………
Alega Kennedy
F18/2000/2005
Signature…………………
Date…………………..
Supervisor
This project has been submitted for examination with my approval as the project supervisor.
Dr. Nyang’aya.
Signature……………………
Date…………………….
ii
DEDICATION
This project is dedicated to our parents for being there for us always, our siblings and colleagues
for all their support during our undergraduate studies.
You are so special to us and may the almighty God Bless You ABUNDANTLY.
iii
ACKNOWLEDGEMENT
We would like to thank the Almighty God, for safeguarding our lives and health in the
University. We are also grateful to our parents and families for providing financial support and
encouragement throughout our stay at the University.
We would like to sincerely thank our supervisor Dr. Nyang’aya, Senior Lecture in the
Department of Mechanical and Manufacturing Engineering of the University of Nairobi, for his
invaluable guidance, comments, constructive criticism and priceless advice which facilitated the
compilation of this report.
Last but not least, we appreciate the support we got from the University of Nairobi Mechanical
and Manufacturing Engineering Workshop technical staff.
May God bless you all for your kindness and patience.
iv
LIST OF SYMBOLS
ND – nominal diameter (mm)
hf – friction head loss (moW)
hts –static head (moW)
H – total pressure head in the pipe (moW)
hm – loss due to bends and fittings (moW)
m – mass flow rate ()
ρ – Density of water (kg/m3)
Q – Discharge (m3/sec)
Cp – specific heat capacity of water (J/Kg K)
∆T – temperature difference of water between the dynamometer’s inlet and outlet (oC)
A – inside area of pipe (m2)
d – Nominal diameter of selected pipe (mm)
t – Wall thickness of the selected pipe (mm)
k – inside roughness of selected pipe material (mm)
di - inside diameter of the selected pipe (mm)
v – Mean velocity of flow (m/s)
Re – Reynolds number for pipe flow (dimensionless)
υ – Kinematic viscosity (m2/s)
λ – Coefficient of fluid friction or Friction factor (dimensionless)
g – Gravitational acceleration
L – Length of pipe (m)
p - Pressure (N/m2)
v
z – static head (m)
hL – total Head in the Pipeline (moW)
Qd – demand flow rate (m3/s)
Qg – flow rate due to gravity (m3/s)
Qp – flow rate to be supplied by the pump (m3/s)
vi
ABSTRACT
The title of the project was installation and testing of an engine dynamometer. Due to wide scope
of the whole project, it had to be done in phases. In this first phase the following objectives were
set to be achieved.
•
Selecting an appropriate site for the installation of the dynamometer in the workshop.
•
To design the system with proper interaction of components
•
To select appropriate pipe material and sizes to meet the demand discharge required by
the dynamometer to be installed.
•
To select an appropriate pump type and pump size to meet the demand discharge.
•
To select appropriate valve type and valve sizes to control the flow of water into and out
of the dynamometer and engine during engine testing.
The first step was selecting a suitable site for the dynamometer in the workshop. Selecting the
site was a fundamental step in installing the dynamometer because once the decision is taken and
the installation done, it is costly and difficult to change. A physical survey of all the potential
sites in the engine shop was carried out and the most efficient site and layout was selected. The
factors considered during the site selection were safety, flexibility, maximum accessibility and
minimum investment.
Constructions of the system and interaction of components of the dynamometer was also
considered. The following components were found important for the proper functioning of the
dynamometer : pressure gauge, water pump, strainer, cradle, coupling, exhaust system, the intake
duct system, the engine cooling system, thermometers, throttling valves and engine control
system. AutoCAD drawings showing the layout of components and their interactions were made
The next step was the process of designing for a pipe size and for an appropriate pump to ensure
sufficient supply of water to the dynamometer. The water flow rate required was established
from an energy balance equation between the inlet and out of the dynamometer. A boost pump
was found necessary to boost the flow rate to meet the dynamometer water demand. The pump
chosen was specified in terms of the flow rate and the total head to be overcome by pumping
system. Both throttle valves and shutoff valves identified, selected and their sizes established.
The results obtained were discussed and concluded that objectives were achieved. Finally
recommendations were made.
vii
TABLE OF CONTENTS
Declaration………………....…..………………………………………….………………. ii
Dedication…………………….…….……………………………….…………………….. iii
Acknowledgement…………………………………………………………...……………. iv
List of Symbols……………………………..………………………………………..……. v
Abstract……………………………………………………………………...…………….. vii
Table of Contents………………………………………………………………………….. ix
CHAPTER ONE: INTRODUCTION
1.1 Background of dynamometers……………………………………………………....……1
1.1.1. PRINCIPLES OF OPERATION OF ABSORBING DYNAMOMETERS……………..….…. 1
1.1.1.1. Constant Force……………………………………………………….….….. 1
1.1.1.2. Constant Speed……………………………………………………..……….. 2
1.1.2. Types of Dynamometers…………………………………………….……..…. 2
1.1.3. Designs of Dynamometers………………………………………….…...……. 2
1.2 Project Justification…………………………………………………………..……….…. 3
1.3 Objectives……………………………………………………………………..……….… 3
CHAPTER TWO: LITERATURE REVIEW
2.1. CONSTRUCTION AND OPERATION OF A DYNAMOMETER………………..… 4
2.2. TORQUE AND POWER CHARACTERISTIC OF A DYNAMOMETER…….……. 5
2.3. PURPOSE OF WATER IN A DYNAMOMETER…………………………….…..…… 6
2.3.1. Water Tank…………………………………………………………………….….. 7
2.3.2. Booster Pump……………………………………………………………..……. 7
2.3.3. Water Pressure Regulator…………………………………………..……….. 8
2.3.4. Load Control Valve…………………………………………………..….…… 8
2.3.5. Plumbing………………………………………………………………....……….. 8
2.4. CALCULATING FLOW RATE……………………………………………..….….….. 9
2.5. PUMPING SYSTEM………………………………………………………..….………. 9
2.5.1. DEFINITION OF PROBLEM………………………………….……….…….….. 9
2.5.1.1. Specification of the discharge flow rate required………………………..… 9
viii
2.5.1.2. Specifications of the total pressure head to be overcome by pumping system...10
2.5.1.2.1. Total static head………………………………………………...……10
2.5.1.2.2. Pressure loss due to friction head in the pipeline………………..…...10
CHAPTER THREE: INSTALLATION SITE
3.1 SELECTION OF INSTALLATION SITE AND LAYOUT…………….….………. 11
3.1.1. Safety………………………………………………………………….….…….… 13
3.1.2. Flexibility…………………………………………………………….….……..…. 13
3.1.3. Overall Integration………………………………………………….…….….…… 13
3.1.4. Effective Use of the Available Space……………………………….….…….....… 14
3.1.5. Minimum Movement..……………………………………………….….………… 14
3.1.6. Maximum Accessibility…………………………………………….……...……… 14
3.1.6. Minimum Investment…….……………………………………….………..……… 14
3.2 CONSTRUCTIONS OF THE SYSTEM AND INTERACTION OF COMPONENTS.…15
3.2.1. The Test Cell………………………………………………………………….…… 15
3.2.2. Pressure Gauge………………………………………………………………..…… 15
3.2.3 Water Pump…………………………………………………………….…….…….. 15
3.2.4. Strainer…………………….……………………………………………………… 16
3.2.5. Cradle……………………….…………………………………………………….. 16
3.2.6. Coupling…………………………………………………………….…….………. 16
3.2.7. Exhaust System……………………………………………………….….…….….. 17
3.2.8. The Intake Duct System…………………………………………….….……….…. 17
3.2.9. The Engine Cooling System………………………………………….…………… 18
3.2.10. Thermometers…………………….……………………………….………...…… 20
3.2.11. Throttling Valves…………….……………………………………………..……. 21
3.2.12. Engine Control……………….………………………………………………..…. 21
3.2.13. Fuel System…………………………………………………………………....…. 21
3.2.14. Water Flow……………………………………………………………………….. 22
CHAPTER FOUR: PUMPING SYSTEM
4.1. DESIGN PROCEDURE FOR PUMPING SYSTEM…..………………………………. 23
4.1.1. Survey of the site…………………………………………………..……………… 23
ix
4.1.1.1. Pipeline route and length……………………………………….……..…….. 23
4.1.1.2. Determination of static head………………………………….……….….… 23
4.1.2. PRELIMINARY SELECTION OF PIPE SIZE TO MEET DEMAND….…..…... 23
4.1.2.1. Selection of pipe size………………………………………………..………. 24
4.1.2.2. Selection of pipe material……………………………………………...….… 26
4.1.3. CHARACTERISTICS FOR THE SELECTED PIPELINE (Actual)……….……. 28
4.2. PRELIMINARY SELECTION OF PUMP………………….……………………….…. 32
4.2.1. DETERMING FLOW RATE DUE TO GRAVITY…….………………….……... 32
4.2.2. SELECTION OF BOOST PUMP………………………………………….……… 41
4.2.2.1. Flow Rate to Be Supplied By the Pump…………………...……….……...... 41
4.2.2.2. Total Head in the Pipeline…………………………………….…….……….. 42
4.2.2.2.1. Static Head to Be Overcome By the Pump……………….……….. 42
4.2.2.2.2. Head loss due to friction in the pipeline ………………….……….. 42
4.2.2.2.3. Head Loss Due To Fitting and Bends…………………….……….. 50
4.2.3. SELECTION OF PUMP TYPE………………………………………….………… 50
4.3. VALVES………………………………………………………………………………… 52
4.3.1. SELECTING A VALVE TYPE………………………………………….…………52
4.3.1.1 Throttling valves………………………………………………………….….. 52
4.3.1.2 Shut-off valves……………………………………………………………….. 53
4.3.2. SELECTION OF GATE VALVES SIZES.……………………………………….. 54
CHAPTER FIVE
DISCUSSIONS…………………………………………………………………….………… 55
CHAPTER SIX: CONCLUSIONS AND RECOMMENDATIONS
6.1 CONCLUSIONS……………………………….………………………………………… 61
6.2 RECOMMENDATIONS………………………………………………………………… 61
REFERENCES………………………………….…………………………..……………… 62
APPENDICES……………………………….……………………………………………… 63
Appendix 1……………………………………………………………Project CAD Drawings
Appendix 2…………………………………………………………………………………. 63
Appendix 3…………………………………………………………………………………. 64
x
CHAPTER ONE
1. INTRODUCTION
1.1 Background of dynamometers
A dynamometer is an instrument for measuring power, force or energy, such as the power
developed by an internal combustion engine or electric motor, or the current voltage or power in
an electric circuit.
It can also be used to determine the torque required to operate a driven machine such as a pump.
In that case, a motoring or driving dynamometer is used. A dynamometer that is designed to be
driven is called an absorption or passive dynamometer. A dynamometer that can either drive or
absorb is called a universal or active dynamometer.
1.1.1. PRINCIPLES OF OPERATION OF ABSORBING DYNAMOMETERS.
An absorbing dynamometer acts as a load that is driven by the prime mover that is under test
(e.g. Pelton wheel). The dynamometer must be able to operate at any speed and load to any level
of torque that the test requires. It is usually equipped with some means of measuring the
operating torque and speed.
The dynamometer absorbs the power developed by the prime mover. The power absorbed by the
dynamometer is converted into heat and the heat is generally dissipated into the ambient air or
transferred to cooling water.
Absorption dynamometers can be equipped with two types of control systems to provide
different main test types.
1.1.1.1. Constant Force
The dynamometer, which has a "braking" torque regulator, is configured to provide a set braking
force torque load while the prime mover is configured to operate at whatever throttle opening,
fuel delivery rate or any other variable it is desired to test. The prime mover is then allowed to
accelerate the engine through the desired speed or RPM range.
1
1.1.1.2. Constant Speed
If the dynamometer has a speed regulator, then the dynamometer provides a variable mount of
braking force (torque) that is necessary to cause the prime mover to operate at the desired single
test speed or RPM. The braking load applied to the prime mover can be manually controlled or
determined by a computer. Most systems employ eddy current, oil hydraulic or DC motor
produced loads because of their linear and quick load change ability.
1.1.2. Types of Dynamometers.
Most dynamometers are classified in one of these two categories:
•
Engine dynamometers
They are designed for coupling directly to the driveshaft of an engine under test.
•
Chassis dynamometers
They measure the power output of a drive train by using rollers turned by the tires of a
vehicle under test.
1.1.3. Designs of Dynamometers
•
Eddy current dynamometers: harness the magnetic flux between fixed and rotating
electromagnets spun by the engine under test.
•
Powder dynamometers: create flux through the application of a fine magnetic powder
between the rotor and coil.
•
Electric Motor Testing Systems: Electric Motor Testing Systems are designed to provide
maximum reliability, excellent durability and performance; available for testing electric
motors.
•
Fan, hydraulic and water brakes: use air, water or hydraulic fluid to provide an indication
of the power applied to the system.
2
1.2 Project Justification
It is crucial to understand the importance and necessity of subjecting an internal combustion
engine to complete and thorough test of efficiency and durability before it is put in use. In this
interest therefore, a dynamometer is an inevitable device in any engine manufacturing industry or
other industry which deals with the maintenance and repair of internal combustion engines.
Internal combustion engines are characterized by many moving parts which are subjected to
wear and tear and therefore their original operating conditions cannot be maintained. This
depreciation in quality is marked with a corresponding decrease in efficiency and for this reason;
there must be ways and means of measuring the power output in order to know how much the
engine is deviating from the original operating condition. A dynamometer provides a reliable
means of measuring the power output by coupling it to the engine.
1.3 Objectives
•
To select an appropriate site for the installation of the dynamometer in the workshop.
•
To construct the system with proper interaction of components
•
To select appropriate pipe material and sizes to meet the demand discharge required by the
dynamometer to be installed.
•
To select an appropriate pump type and pump size to meet the demand discharge.
•
To select appropriate valve type and valve sizes to control the flow of water into and out of
the dynamometer and engine during engine testing.
3
CHAPTER TWO
2 LITERATURE REVIEW
2.1. CONSTRUCTION AND OPERATION OF A DYNAMOMETER
The Heenan and Froude dynamometer type under study is a water brake. It consists of a rotor,
driven by the engine or machine being tested, with this rotor being fitted inside a water-filled
casing (defined as a stator) that is free to rotate a limited distance. The rotor forces the water
against the casing, and thus the torque is transmitted through the water to the casing. By
measuring the effort needed to prevent the casing from rotating, the output torque, and thus the
power, can be determined. The power is dissipated by the heating and circulation of the water.
The amount of momentum transferred at a given revolutions per minute (RPM) is a function of
the mass of water that is accelerated from near the center of the dynamometer rotor to the
tangential velocity of the rotor tip per unit time. The factors determining this torque at a
particular RPM are the rotor diameter (which determines tangential velocity), the efficiency of
the transfer of the tangential velocity water from the rotor to the stator and then the stator
converting the tangential water velocity to axial water velocity, and the percentage of water
capacity inside the dynamometer at that moment.
The majority of the water inside a water brake dynamometer is re-circulated back from the stator
to the rotor. Water is transferred from the rotor tangentially, generally near the circumference of
the rotor, with the water being received by the stator and its direction being changed so that it is
re-circulated back to the rotor in an axial direction generally near the center of the rotor. The
process of momentum transfer from the rotor to the stator generates heat in the water.
There must be a water flow through the dynamometer sufficient to remove the heat generated or
the water will eventually flash to steam. The flow required is calculable from the power being
absorbed and the desired outlet water temperature or desired temperature rise above the inlet
water temperature.
Outlet temperature control is not critical, but a general guideline of 60°C maximum will reduce
lime deposits within the dynamometer, 70°C is permitted for short durations and 80° C is the
absolute maximum for very short durations. Excessive mineral deposits will result from
4
sustained operation in the 70°C - 80°C range. Generally most dynamometer installations produce
acceptable outlet temperatures (60°C - 70°C) when the water flow is adjusted for a 10°C - 20°C
rise above the inlet water temperature.
Water brake absorbers are relatively common, having been manufactured for many years and
noted for their high power capability, small package, light weight, and relatively low
manufacturing cost as compared to other, quicker reacting power absorber types.
Their drawbacks are that they can take a relatively long period of time to stabilize their load
amount and the fact that they require a constant supply of water to the water brake housing for
cooling. Environmental regulations now prohibit flow- through water and large water tanks must
be installed to prevent contaminated water from entering the environment.
2.2. TORQUE AND POWER CHARACTERISTIC OF A DYNAMOMETER.
The word horsepower was introduced by James Watt, the inventor of the steam engine in about
1775. Watt learned that "a strong horse could lift 150 pounds a height of 220 feet in 1 minute."
One horsepower is also commonly expressed as 550 pounds one foot in one second or 33,000
pounds one foot in one minute. These definitions include force (pounds), distance (feet), and
time, (minute, second). A horse could hold weight in a static position but this would not be
considered power, it would be similar to torque. Adding time and distance to a static force (or to
torque) results in power. RPM, revolutions (distance) per minute (time), is today's equivalent of
time and distance.
Power can be directly measured. However there is a problem directly measuring power of
modern day internal combustion engines because they produce rotary motion, and not linear
motion, and unless the engine is geared down, the speed at which they do work (time and
distance or RPM) is too great for practical direct measurement of power. The solution is to
directly measure torque (rotational force expressed in pounds at one foot radius) and RPM (time
and distance, i.e. distance in circumference at the one foot radius) and from these calculate
power. Torque and RPM are easily measured directly.
The H&F D.P.Y.5 dynamometer in the present project uses water brake as a device to load the
engine. A torque arm is attached to the brake's stator. The brake's rotor is then coupled to the
5
engine's crankshaft. A spring scale is connected the torque arm to the stationary fixture holding
the engine and brake.
The dynamometer measures torque and RPM and then from these, the power is calculated. The
dynamometer takes more water to increase the load on the engine being tested. As the test
engine's torque rises more water flow is needed. As the test engine's torque drops less water flow
is needed. The dynamometer’s water brake does not respond to power. Major adjustments to
water flow are needed as an engine crosses its torque peak but none are needed as it crosses its
power peak. In other words the water flow to the brake during a dynamometer test follows the
engines torque curve and not its power curve. Torque is what twists the shaft. Power helps us
understand an amount or quantity of torque. (Torque × time and distance)
2.3. PURPOSE OF WATER IN A DYNAMOMETER
Water serves two purposes in a water-brake dynamometer application:
•
It provides the means by which “load” (opposing force to the rotation of the engines’
crankshaft) is applied to the dynamometer.
•
Water also carries away the heat generated by the process of absorbing torque, and thus
power.
The Nairobi City Council water systems often have inconsistent water performance of the
dynamometer. Typically, the supply pressure from the City Council is not enough to provide the
water brake dynamometer with the amount of water required at its maximum absorption capacity
which results in a lack of controllability.
Insufficient water supply can also contribute greatly to internal wear of the absorbing elements
due to cavitations. An adequate water system is therefore the biggest factor in getting a
dynamometer to run optimally. The best practice to ensure proper water delivery pressure and
volume is to build a system that includes a source of water that isn’t affected by outside
influences.
6
A basic water system includes the following:
2.3.1. Water Tank
In the Mechanical Engineering workshop, there exists an overhead water tank which supplies the
workshop with water. This tank will be the source of water for the H&F D.P.Y.5 dynamometer
to be installed. The capacity of the tank was established as 1232 litres.
The water brake dynamometer requires draining to gravity at atmospheric pressure so there
exists a gravity drain tank to collect the exhausted load water. An installed transfer pump returns
it to the overhead supply tank.
Since the water is used to absorb the torque output of the engine which is being converted to
heat, the dynamometer outlet water has to be cooled. In the workshop, this is done with in a
radiator with an electric fan designed specifically for this purpose.
If the hot, effluent water from the dynamometer is dumped directly back into the supply tank, the
inlet water temperature will continue to rise. This change in inlet temperature will cause the
density of the inlet water to decrease which will change the absorption profile of the
dynamometer.
2.3.2. Booster Pump
In cases where the suction tank is located above the dynamometer, the need for the pump has to
be determined. This is accomplished by examining the flow rate that that can be delivered by
gravity and comparing it with the required demand discharge. If the demand discharge exceeds
the discharge due to gravity, then a pump is required to boost the flow rate to meet the demand.
The type of the pump is chosen depending on the flow rate and the total head to be overcome by
pumping. If additional equipment, such as a cooling tower for the engine is used, the additional
water usage has to be accounted for. It is always good practice to oversize the water system and
its’ components to allow for future upgrades.
The distance between the booster pump/ pressure regulator outlet and the dynamometer inlet
control valve should be minimized to ensure the pressure and volume delivered is not affected by
the head loss associated with the distance traveled through a piping system.
7
If the pump fails to provide at least the minimum required supply pressure that the dynamometer
is designed to operate with, this will effectively decrease the absorption capacity of the brake
unit.
2.3.3. Water Pressure Regulator
Depending on the type of dynamometer selected for the application, the water delivery pressure
requirement will vary. The output pressure of a centrifugal pump is usually pretty steady by
design, but the pressure must be regulated to an amount consistent with the power absorption
characteristics of the dynamometer. As just like any other machine, water-brake dynamometers
have water pressure and volumetric flow-rate specifications that coincide with the units’ optimal
performance characteristic.
Optimal performance of a water-brake dynamometer exists when all conditions are met to
achieve linear control throughout the power range of the unit. A water pressure regulator, when
paired with a sufficient booster pump, will ensure the proper water delivery criterion is met. The
water pressure regulator should be mounted at the outlet of the pump, as close to the load control
valve as possible.
2.3.4. Load Control Valve
Controlling the amount of water to and from the dynamometer is done by throttling a valve in the
inlet and/or outlet to achieve the parameters such as RPM and/or Torque desired. Manually
operated globe valves, ball valves, butterfly valves, and needle valves are all examples of a good
means by which to throttle load water and control a water-brake dynamometer. Gate valves are
not designed to throttle water. The design of a gate valve is for isolation only. Attempting to
control load with a gate valve will result in a turbulent, non-linear flow profile and erratic
behavior and/ or lack of control of the dynamometer. All load control valves should be located
as close to the dynamometer as possible, without sacrificing personnel safety.
2.3.5. Plumbing
The process of the plumbing is just as important as the equipment used to support the
dynamometer test cell. It is always important to minimize joints and bends in the supply line to
the dynamometer inlet.
8
The pump should not be mounted in a position that it has to draw water up from its supply tank.
In a situation where a 900 bend must be made prior to entry into a system component such as a
pump, regulator, strainer, or valve, at least 300mm of straight pipe or hose should be installed
prior to entry into one of the above mentioned components to allow the turbulent flow created by
the rapid fluid direction change while traveling through the bend to dissipate. Flexible, noncollapsible hose may be used to make otherwise sharp bends more gradual and limit the pressure
drop normally experienced with the use of cast fittings.
2.4. CALCULATING FLOW RATE
Both the water level and the speed of the dynamometer determine how much torque resistance
there is (more torque is attained at a higher speed for a given water level). Once the
dynamometer is full at a given speed, its torque capacity is maximized. If the engine speed
increased then the flow rate through the dynamometer should be increased, while keeping it full,
to keep the water temperature acceptable.
The flow required is calculable from the power being absorbed and the desired outlet water
temperature or desired temperature rise above the inlet water temperature.
2.5. PUMPING SYSTEM
2.5.1. DEFINITION OF PROBLEM
The requirement to be met by the pumping system is specified in terms of:
•
The discharge flow rate for transfer of water from the suction to the dynamometer inlet.
•
Total pressure head to be overcome by the pumping system.
2.5.1.1. Specification of the discharge flow rate required
The discharge flow rate required was stated in cubic meters/sec (m3/s). It was determined by a
study of water demand required by the dynamometer.
9
2.5.1.2. Specifications of the total pressure head to be overcome by pumping system
The total pressure head (H) to be overcome by the pumping system was stated in meters of water
(moW). This total pressure head (also referred to as the dynamic head), is the sum of the static
head and friction head. It was referred to as dynamic because it incorporated the head loss due to
fluid friction in pipe line, which arose only during the dynamic conditions of fluid flow.
H = hts + hf
Where,
H – Total or dynamic pressure head to be overcame by the pumping system
hts – total static head to be overcome by pumping
hf – total pressure head due to fluid friction in pipe line
The two components of pressure were as elaborated below:
2.5.1.2.1. Total static head (hts).
The total static head in the pumping system was the water level difference between the suction
and delivery to the dynamometer inlet.
2.5.1.2.2. Pressure loss due to friction head in the pipeline (hf).
The friction head was the total pressure head lost due to fluid friction which occurred as fluid
flowed through the pipe line. This friction head loss included that in pipe work and fittings
starting from the suction inlet fittings, through to the discharge pipe outlet. For a given discharge
flow rate this friction loss depended on the pipe material, size, length and the type and number of
fittings. It was computed once these pipeline specifications were determined.
10
CHAPTER THREE
3 INSTALLATION SITE
3.1 SELECTION OF INSTALLATION SITE AND LAYOUT
Selection of installation site refers to the allocation of a space convenient for the installation of
the dynamometer and other equipments essential for its operation within the engine shop in such
a manner that the overall installation and operating costs are minimized.
Selecting the location site is therefore the fundamental step in installing the dynamometer in the
present project because once the decision is taken and the installation done, it is costly and
difficult to change. A physical survey of all the potential sites in the engine shop was carried out
and the most efficient site and layout identified as shown in Fig 1 below.
11
Fig. 1 Installation site
12
The following factors were considered in selecting this convenient space and layout for the
installation of the dynamometer and its components.
3.1.1. Safety
Due consideration to industrial safety methods is necessary and therefore care was taken not only
of the persons operating the dynamometer but also of any other person within the test cell. To
achieve this, enough space was left between the dynamometer and the wall and the other
machines to ensure that any individual can stand in front or behind the equipments without being
hurt. Still, hazardous facilities such as the exhaust pipe were located in a position such that it
cannot burn anybody during test. In addition to this, all pipes were placed in the drain channels
and those not in drain channels were placed as close to the engine and dynamometer base as
possible. This ensured that persons in the workshop could not be injured by the pipes running on
the floor.
3.1.2. Flexibility
A flexible site is one in which the facilities could be rearranged at a minimum cost and least
inconvenience.
The dynamometer in the present project can be coupled to an engine on both sides and therefore
the design and construction of the site selected in the engine shop along with its plumbing had to
provide a repeatable environment for all coupling and testing on either side of the dynamometer.
On these grounds, the site selected as shown in Fig 1fulfilled this requirement in that enough
space was left on either side of the dynamometer for the installation of the cradle.
3.1.3. Overall Integration
As a basic requirement, all facilities of the system should be fully integrated into a single
independent operating unit to achieve maximum efficiency and minimum cost of operation.
Installation of the dynamometer in the present project required coordination of its components
without being affected or affecting the operation of other machines in the workshop. In addition
to this, interaction with existing or planned facilities in the workshop and in the selected space
13
was essential in reducing the total installation cost. These facilities included drain channels,
utility routings and engine cooling water tank.
3.1.4. Effective Use of the Available Space
The layout was selected to make an effective use of the available space both horizontally and
vertically. This was achieved by locating all the components in strategic places where they could
be operated and maintained easily.
3.1.5. Minimum Movement
Minimum travel of the operator is essential in improving operating efficiency. In this present
project, minimum distance of travel during engine testing was achieved by locating the engine
control system as close to the engine as possible. In addition to this, both throttling and gate
valves were located near the engine, pump or the dynamometer in order to ensure that they can
be reached with ease.
3.1.6. Maximum Accessibility
A good layout is one that makes all servicing, maintenance and control components easily
accessible. The layout is designed such that there is enough clearance from the wall and other
machines in the engine shop. This also allows the engine to be coupled or decoupled to the
dynamometer easily. All gauges are located at points where they are readily observable hence
they could be read and adjusted easily.
3.1.6. Minimum Investment
All the available facilities were utilized in an optimum manner to result in minimum initial
investment without affecting the proper functioning of the dynamometer to be installed or any
other existing machine. These facilities include the overhead suction tank, engine water cooling
tank and some of the existing pipeline as shown in the Appendix 1.
14
3.2 CONSTRUCTIONS OF THE SYSTEM AND INTERACTION OF COMPONENTS
3.2.1. The Test Cell
The test cell is the mechanical engineering workshop where the dynamometer is installed. It has
a strong concrete floor sufficient enough for the foundation of the dynamometer. This meets the
requirement for the dynamometer to be firmly fixed to a substantial foundation to facilitate
steady running and eliminate vibrations.
Ventilation and lighting are typically the biggest issues with an indoor test cell. This is because
engine testing processes delivers a lot of heat and exhaust fumes to the surrounding and must be
dealt with in order to guarantee safety and comfort of the operators. The aim is to remove
polluted air regularly and to replace it with clean air. If conditions are such that natural
ventilation is not sufficient, it should be supplemented with artificial ventilation which involves
the intake of fresh air by fans. Adequate lighting which is of the right quality is important for
efficient operation of the dynamometer. Good lighting reduces eye strain in reading
measurements, prevents accidents and promotes efficiency and high quality work.
The high enough roof, use of glass in the outer walls and wide doors provides better and reliable
use of natural lighting and ventilation. The wide doors also facilitate wheeling of the engine
stand through the workshop.
3.2.2. Pressure Gauge
The pressure gauge is used to indicate the pressure at which the water enters the dynamometer.
3.2.3 Water Pump
This is a centrifugal pump driven by an electric motor and is used to boost the flow rate to meet
the demand discharge. It pumps water from the overhead tank through existing 65mm and 50mm
nominal diameter pipes to the pump inlet via a selected 50mm ND pipe. It pumps water to the
dynamometer and to the mixing tank once the manual valve for the latter is opened. The pump is
located strategically near the tee junction where the water is tapped from the existing pipe and
close to the pump in order to shorten the lengths of the pipes. The specifications of the pump and
the electric motor are described later in this section
15
3.2.4. Strainer
The cleanliness of the water from the over head tank is not guaranteed therefore a strainer was be
installed between the dynamometer inlet and the pump. This is because foreign materials such as
metallic pieces can damage the dynamometer.
3.2.5. Cradle
It is used for supporting the engine and should be of rigid design, adequately bolted to a suitable
foundation in correct alignment with the dynamometer. It is adjustable and therefore convenient
for use in testing a variety of engines.
3.2.6. Coupling
Coupling an engine to the dynamometer is crucial because the high bending moments caused by
the use of improper shafting and misalignment between the engine and dynamometer can lead
serious engine and dynamometer bearing damage. For this reason, it is not advisable to have a
strong and rigid drive shaft connecting the engine to the dynamometer. This is because when the
high stresses are combined with the high fatigue cycles, catastrophic failure may occur. It is
therefore advisable to use a flexible coupling with the lightest possible construction and in
perfect dynamic balance which prevents whirling.
The flexible coupling should be mounted in such a way that it does not overhang from the
bearings of the dynamometer shaft. If adaptors are necessary, then they should be of the smallest
diameters possible. Starting shaft coupled to the dynamometer should be supported in its own
independent bearing and connected to the dynamometer through a flexible joint
Another design consideration for the coupling shaft is that it should permit a universal coupling
arrangement such that several engine models may be quickly coupled and tested without the need
for special adaptor so that as soon as one engine is uncoupled from the dynamometer another
engine may be coupled thereto with a minimum of time lapse. This can be achieved by using an
adjustable length shaft.
Cardan shaft which consists of two universal joints mounted back to back, with an intermediate
shaft is preferred for the coupling. The second U joint cancels the velocity errors introduced by
16
the single joint, and so they act as a constant velocity joint provided both the driving and the
driven shaft are parallel and the two universal joints are correctly aligned with each other. The
cardan shaft allows for some misalignment between the engine crankshaft and the dynamometer
input flange. The advantages claimed for their use is that they are available in a variety of sizes
depending on engine type and are dynamically balanced. However they require use of a drive
shaft guard.
Fig 2 Cardan shaft with two universal joints
3.2.7. Exhaust System
The exhaust gases from the engine are routed out of the workshop through a vertical discharge
pipe to the atmosphere. A stainless steel pipe is preferred to prevent rust and of adequate
diameter to avoid restriction and back pressure which is a common course of loss of engine
power. High temperature flexible pipe with quick disconnect flanges and locking clamps should
be used to ensure tight fittings which prevent the exhaust fumes from permitting into the
workshop and facilitate the connection to the exhaust manifold of the different engines to be
tested. A silencer needs to be fitted in the exhaust pipe to reduce undesired sounds which can
lead to lack of concentration of the operator and barrier to communication.
Exhaust gasses are hot and hence the exhaust system pipes therefore they should be located in a
position that guarantees the safety of the operator.
3.2.8. The Intake Duct System
The intake duct is used to route the atmospheric air into the engine intake manifold. It should be
fitted with an air filter to eliminate foreign materials from entering the engine. The orientation of
the air intake duct is crucial and should be in a position to allow fresh and cool atmospheric air
into the engine.
17
An engine can only draw in a certain volume of air depending on the engines size and operating
conditions but modern engines are designed to accommodate large changes in air volume and
pressure by incorporating a control unit which adjusts the amount of fuel required as a result of
these changes. Air restriction when the filter is loaded with dust can result into reduced air
pressure which is a major cause of loss of power and therefore the filter should be often cleaned
or replaced as it might be required.
Air contains water vapor and therefore the duct should be fabricated from a rust resistance
material but strong enough to withstand the vibrations in the test cell (stainless steel). In addition
to this, the fittings should be flexible to a void sharp bends which can interfere with the smooth
flow of air and quick disconnect flanges as well as locking clamps should be provided to
facilitate coupling to the intake manifold of the different engines to be tested. The intake duct for
this dynamometer is installed facing the wide door of the workshop and it is recommended to
have the door open during test to allow the flow of fresh air.
3.2.9. The Engine Cooling System
The engine cooling system is undoubtedly important in an engine test cell in order to prevent
over heating of the engine and its corresponding consequences such as detonation, pre-ignition
and blown head gaskets. One key requirement is that an engine fails if just one part overheats
therefore, it is vital that the cooling system keep all parts at suitably low temperatures. Liquidcooled engines are able to vary the size of their passageways through the engine block so that
coolant flow may be tailored to the needs of each area. Locations with either high peak
temperatures (narrow islands around the combustion chamber) or high heat flow (around exhaust
ports) may require generous cooling. This reduces the occurrence of hot spots, which are more
difficult to avoid with air cooling. Most engines today are designed to operate within a
temperature range of about 90oC to 104oC. A relatively constant operating temperature is
absolutely essential for proper emissions control, good fuel economy and performance.
In addition to this, the clearances in most of today's engines are much closer than those in
engines built earlier. Piston-to-cylinder clearances are much tighter to reduce blow by for lower
emissions. Valve stem-to-guide clearances also are closer to reduce oil consumption and
emissions, too. Plus, many engines today have aluminum heads with overhead cams. Such
18
engines don't handle higher than normal temperatures well, and are very vulnerable to heat
damage if the engine gets too hot. Due to this reason the maximum temperature should be
avoided.
The cooling system adopted for this dynamometer is a closed loop water system which consists
of a small mixing tank fixed above the engine. To this tank is connected a 20 mm ND rising pipe
carrying the hot water from the engine and a 40 mm ND return pipe carrying water to the engine
and the water is pumped by the engine water pump. A flexible high temperature rubber pipe
fitted with locking clamps is used to connect the pipes to the different engines to be tested. A
cold water supply pipe of 15 mm ND is connected to discharge visibly into the tank at a point
near the return pipe and remote from the overflow pipe. The cold water supply pipe picks water
from the pump delivery through a reducer and is fitted with a manually operated gate valve V1.
In order to avoid wastage of water when dismantling engines, gate valves are fitted to each of the
pipes connected to the engine. An over flow pipe is fixed near the hot water inlet to the tank.
Under operating conditions, the water leaving the engine is maintained within the limits specified
by manufacturers by controlling the gate valve. This is illustrated in Fig. 3.
19
Fig 3 Engine Water Cooling System
3.2.10. Thermometers
The main parameters used in calculating the power output are torque, dynamometer constant K
and speed of rotation and therefore the temperatures indicated by the thermometers are control
variables used to maintain the operating conditions of the engine and the dynamometer within
the temperature limits recommended by the manufacturer. This improves efficiency, accuracy
and prevents damage to the engine as well as the dynamometer.
The following thermometers are used:
Thermometer 1
It is a mercury-in glass thermometer.
It’s used to measure the inlet water temperature to the dynamometer.
The water is expected to enter the dynamometer at room temperature (25oC) but some additional
heating occurs due to pumping therefore the temperature can rise to a maximum of 35oC hence
the thermometer temperature range is 20 oC - 35 oC.
Thermometer 2
It is a mercury-in glass thermometer.
It’s used to measure the outlet water temperature from the dynamometer.
The water enters the dynamometer at 60° C but could rise to70° C which is only permitted for
short durations, while 80° C is the absolute maximum for very short durations; hence the
thermometer temperature range is 60 oC - 80 oC.
Thermometer 3
It is a mercury-in glass thermometer.
It’s used to measure the inlet water temperature to the engine. The warming up temperature of
the engine is the limit and is always specified by the engine manufacturers.
20
Thermometer 4
It is a mercury-in glass thermometer.
It’s used to measure the outlet water temperature from the engine.
The water outlet temperature from the engine varies with the operating conditions and the engine
type and therefore the limits recommended by the manufacturers should be adhered to.
Thermometer 5
This is also mercury in glass thermometer and is used to measure the temperature of the water in
the engine water tank. Its limit depends on the temperature of the water coming from the engine
and is always specified by the manufacturers.
3.2.11. Throttling Valves
Throttle valves are used as the main control system of the dynamometer to match the various
operating conditions of the test engine by increasing or reducing the flow rate of the water into
the dynamometer. The valves used here are manually operated by the operator and are fitted both
on the intake (as close to the inlet into the dynamometer as possible) and exitt pipe of the
dynamometer.
3.2.12. Engine Control
The engine control is made to regularly vary the operating conditions of the engine so that the
out put can be tested on the dynamometer in the same manner in which the accelerator operates
in the vehicle. The control unit is operated manually by rotating the handle which adjusts the
length of the chord attached to the throttle valve of the engine. It is located as close as possible to
the control of the dynamometer to enhance minimum time lapse in making adjustments.
3.2.13. Fuel System
The fuel system consists of a small tank located above the engine to provide the necessary head
for the fuel flow as required by the engine. It is located on the intake manifold side of the engine
and feeds the engine through a flexible pipe. Locking clamps are important to prevent leakages.
21
3.2.14. Water Flow
Water from the overhead supply tank flows through the 65 mm ND pipe, 50 mm ND pipe and
then to the 50 mm ND pipe into the pump. It is then pumped to the dynamometer via the strainer
through the 50 mm ND pipe. The globe valve on the dynamometer’s inlet is adjusted to allow the
required flow rate into the dynamometer.
Once the casing is filled up, the water circulates as the cardan shaft rotates and absorbs the
engine power. The globe valve on the 40 mm ND outlet pipe of the dynamometer is adjusted to
maintain an almost constant water temperature difference inside the casing. The heated water
flows out to the drainage from where it is directed into the underground collection tank.
From this tank, it is pumped back to the existing overhead supply tank via the radiator (the
radiator fan is run by an electric motor) where it is cooled to atmospheric temperature and the
cycle is repeated till the test is over.
As the engine runs, water is pumped from the mixing (engine water cooling) tank to the engine
by the engine water pump. Hot water from the engine flows back to the mixing tank and the
process is repeated till the test is over.
When the engine is disconnected, the gate valve V3 on the inlet pipe is closed to prevent water
from the mixing from flowing out.
If the temperature of the water in the mixing tank goes beyond the limits specified by the
engine’s manufacturer (when the engine is overheating), the gate valve V1 on the 15 mm ND
pipe is opened and water from the pump flows to the tank to reduce the temperature. To prevent
the over heated water from re-circulating back to the engine and to speed up the cooling process
of the water, gate valve V4 is opened and the hot water flows out into the drainage.
Once the temperature reduces to the required value, the gate valve V4 is closed and the flow
from the pump is cut off.
22
CHAPTER FOUR
4 PUMPING SYSTEM
4.1. DESIGN PROCEDURE FOR PUMPING SYSTEM.
The design of the pumping system therefore proceeds in three steps:
a) Survey of the installation site.
b) Selection of a pipeline.
c) Selection of a pump.
4.1.1. SURVEY OF THE SITE.
This step determined the opportunities and constraints of the environment at which the pumping
system was to be located. The essential data determined in this physical survey of the site was
pipeline length and static head to be overcome by pumping.
4.1.1.1. Pipeline route and length.
The specification of pipeline length was determined through a survey of the intended pipeline
route. This was established as 36.232 m and runs from the suction tank to the dynamometer inlet.
4.1.1.2. Determination of static head.
This is the level difference between suction reservoir and delivery inlet of the dynamometer.
This was established as 3.74 m above the workshop flow.
4.1.2. PRELIMINARY SELECTION OF PIPE SIZE TO MEET DEMAND
A preliminary selection of the pipeline was made using a recommended flow velocity for water
pipelines. This flow velocity recommended for preliminary design of water pipelines was chosen
such that a pressure loss due to fluid friction in pipeline was kept within acceptable limits. This
ensured that pumping equipment size and costs were also kept within certain limits.
23
4.1.2.1. Selection of pipe size
The pipe size was selected such that the flow velocity when the pipeline delivered the design
flow rate remained within a specified range.
The recommended flow velocity was an empirical guide as per the Ministry of Water and
Irrigation specifications manual 2005aimed at the compromise of ensuring that the pressure loss
due to fluid friction in the pipeline was not too high while the discharge flow through the
pipeline was not too low.
Operating duty of proposed pumping system
Maximum power
= 750 Hp (as indicated on the H&F D.P.Y.5 dynamometer to be installed)
Converting Hp into Kw,
= 750 * 0.7457 Kw
= 559.275 Kw
But power
P
= m Cp ∆T
but m = ρQ
= ρQCp∆T
Where,
m – Mass flow rate (Kg/s)
ρ – Density of water (Kg/m3)
Q – Discharge (m3/sec)
Cp – specific heat capacity of water (J/Kg K)
∆T – temperature difference of water between the dynamometer’s inlet (T1) and
outlet (T2) (oC)
T1 – ambient temperature of water at dynamometer’s inlet (25oC)
T2 – permissible temperature of water at dynamometer’s outlet (60 oC)
Hence,
Q
Substituting for
= P/ ρCp∆T
P = 559.275 Kw, ∆T = (60 - 25) oC, ρ = 1000 kg/m3, Cp = 4.184 Kj/kg K
= 559.275 * 1000/ (1000 * 4.184 * 1000 * 35)
Discharge, Q = 3.819 * 10-3 m3/s
24
The recommended range of flow velocities for water pipelines to be applied during preliminary
design was between 1 and 3 m/s (according to the Ministry of Water and Irrigation specifications
manual 2005), after this preliminary state, the design specifications guided further decisions.
Therefore, choosing a flow velocity, v = 2 m/s
Theoretical pipe size to meet discharge requirements
From the continuity equation:
Q
= Av
Where,
A – area of pipe
= π * d2
4
Where,
d = diameter of pipe
Q
= π *d2 * v
4
d
= √ (4 * Q ⁄ π *v)
Substituting for Q = 3.819 * 10-3 m3/s, v = 2 m/s
d
= √ (4 * Q ⁄ π *v)
= √ (4 * 3.819 * 10-3 ⁄ π * 2)
d
= 49.31mm
Theoretical nominal diameter = 49.31mm
25
4.1.2.2. Selection of pipe material
The pressure loss due to fluid friction in pipeline depends on the pipe size, material, length,
fittings and flow velocity. The second step in the selection of pipeline was to select the pipe
material. There are two broad classifications of pipes namely metallic (steel, cast iron and ductile
iron) and non metallic (UPVC). In Kenya only UPVC and steel pipes are available according to
Kenya Bureau of Standards.
Metallic pipes are stronger and harder to break; they are resistant to high pressures; they are easy
to connect, install, operate and maintain; they can withstand shocks and vibrations, but are more
conductive to heat and electricity and less corrosive resistance than none metallic pipes. The non
metallic pipes commonly used are the plastic pipes. Some of the advantages claimed for these
types of pipes are low cost, light weight, easy to cut and join and corrosion resistance. However,
plastic pipes are not as strong as metal pipes, deform easily, expands when subjected to high
temperatures, soften or burns at high temperatures and became brittle in very cold weather.
Plastic pipes are unsuitable for use in the piping system of the dynamometer because of the
operating conditions of the dynamometer such as the mechanical vibrations, high exit water
temperatures as well as exposure to breakage of the pipes lying on the workshop floor by falling
objects and operators stepping on them.
On this account, medium steel pipes are selected and their dimensions in metric series are given
in TABLE 1 below.
26
TABLE 1: DIMENSIONS OF MEDIUM STEEL PIPES – METRIC SERIES
Nominal
Outside diameter
Bore
(mm)
Wall
Mass per unit length
thickness
Max.(mm) Min. (mm)
(mm)
Plain end
Screwed &
pipes (Kg/m)
socketed pipes
(Kg/m)
8
14.0
13.2
2.35
0.65
0.65
10
17.5
16.7
2.35
0.85
0.86
15
21.8
21.0
2.65
1.22
1.23
20
27.3
26.5
2.65
1.58
1.59
25
34.2
33.3
3.25
2.44
2.46
32
42.9
42.0
3.25
3.14
3.17
40
48.8
47.9
3.25
3.61
3.65
50
60.8
59.7
3.65
5.10
5.17
65
76.6
75.3
3.65
6.51
6.63
80
89.5
88.0
4.05
8.47
8.46
100
115.0
113.1
4.50
12.1
12.4
125
140.8
138.5
4.85
16.2
16.7
150
166.5
165.3
4.85
19.2
19.8
27
From TABLE 1 (Dimensions of medium steel pipes – metric series) above, the theoretical
nominal diameter calculated above (49.31mm) lies between 40mm ND and 50mmND. The
larger nominal diameter (50mm) was selected since the smaller nominal diameter (40mm) cannot
meet the required discharge.
The inside diameter of the selected pipe will therefore be given by:
di = d – t
Where,
d – Nominal diameter of selected pipe
t – Wall thickness of the selected pipe
From TABLE 1,
d = 50 mm
t = 3.65 mm
di = 50 – 3.65
= 46.35 mm
4.1.3. CHARACTERISTICS FOR THE SELECTED PIPELINE (Actual)
The selected pipe is specified in terms of demand discharge Q, internal diameter di, velocity of
flow in the pipe v, relative roughness (k/ di), Reynolds number Re, coefficient of friction λ and
total pressure head H.
These characteristics data were calculated as follows:
Relative roughness of selected pipe material and size
= (k/ di)
Where,
k – inside roughness of selected pipe material
(for steel pipes, k = 1.0 mm)
di - inside diameter of the selected pipe
28
= 1.0/46.35
= 0.02157
Mean velocity of flow (v) at demand discharge
v
= Q/A
Where,
A – Inside cross section area of pipe
= π *di2
4
Where,
di = inside diameter of pipe
v
= (4 *Q)
(π * di 2)
= (4 * 3.819 * 10-3)
(π * {46.35 * 10-3}2)
= 2.26 m/s
Reynolds number for pipe flow
Re = v * di
υ
Where,
v - Mean velocity of flow at demand discharge
di - Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
29
TABLE 2: VARIATION OF KINEMATIC VISCOSITY OF WATER
Water temperature oC
0
20
25
30
40
60
100
Kinematic viscosity υ * 10-6 (m2/s)
1.78
1.0
0.91
0.83
0.66
0.48
0.3
At a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s) (from TABLE 2: Variation
of kinematic viscosity of water)
Re
= 2.26 * 46.35 * 10-3
0.91 * 10-6
= 115.11 * 103
Coefficient of fluid friction
λ = f (Re, k/ di) at demand discharge
From the Moody diagram (Appendix 2)
λ = 0.0525
Total pressure head in the pipe
The total pressure head was established by summing up the total static head (hts), total friction
head loss (hf) and loss due to bends and fittings (hm).
H = hts + hf + hm
Pressure loss due to friction in pipe (hf) at demand discharge was calculated using the Darcy
equation as shown below.
From the Darcy equation hf = λ * L * v2
di 2g
Where,
L – Length of pipe from T junction 2 to dynamometer’s base
v – Velocity of flow
di – inside diameter
30
g – Gravitational acceleration
λ – Friction factor
L = 2.647 m, v = 2.26 m/s, di = 46.35 mm, g = 9.81 m/s2, λ = 0.0525
hf =
0.0525 * 2.647 * 2.262
46.35 * 10-3 * 2 * 9.81
hf =
0.5575 moW
Head loss due to bends and fittings, hm = 10% of the head loss due to friction
hm = 0.05575 moW
Static head, hts = 0.78 moW
Total head, H = hts + hf + hm
= 0.78 + 0.5575 + 0.05575
= 1.39325 moW
TABLE 3: CHARACTERISTIC DATA OF SELECTED PIPE
Q
di
v
m3/s
mm
m/s
(*10-3)
3.819
Re
λ
hf
hm
hs
H
moW
moW
moW
{hf +hm + hs}
(* 103)
46.35
2.26
115.11
moW
0.0525
0.5575
0.05575
0.78
1.39325
31
4.2. PRELIMINARY SELECTION OF PUMP
Total flow rate required = 3.819*10-3 m3/s
In the case of this present project, the suction tank is located above the dynamometer; hence the
need for installing a boost pump has to be determined. This is accomplished by examining the
flow rate that that can be delivered by gravity and comparing it with the required demand
discharge. If the demand discharge exceeds the discharge due to gravity, then a pump is required
to boost the flow rate to meet the demand.
4.2.1. DETERMING FLOW RATE DUE TO GRAVITY
To determine the flow rate due to gravity, a steady and uniform flow condition is assumed from
the tank to the dynamometer.
Then, from Bernoulli’s equation,
p/ρg + v2/2g + z = Constant
Taking point 1 as water surface of the suction tank and point 3 as inlet to the dynamometer, then
the above equation can be written as:
p1 + v12 + z1 = p3 + v32 + z3 +hL …………………………………..………………. (i)
ρg 2g
ρg 2g
p1 = pressure at the inlet of the dynamometer = atmospheric pressure
p3 = pressure at the surface of the tank = atmospheric pressure
v1 = velocity at point 1 = 0
v3 = velocity at point 3
z1 – z3 = difference in static head between points 1 and 3 = 3.74 – 0.983
= 2.757 m
32
hL = hf + hm
Where,
hL – total head loss
hf – friction head loss in the pipes
Hm – losses due to bends and fittings
Rearranging equation (i)
p1 - p3 + v12 - v32 + z1 - z3 = hL ……………………………………………………. (ii)
ρg
2g
p1 = p3 and v1 = 0, equation (ii) reduces to
v32 = 2g ({z1 – z3} – hL) ………………………………………………………….. (iii)
Calculating the friction head losses in the pipeline
The pipeline from the suction tank to the dynamometer inlet consists of three steel pipes:
• 65 mm ND (from the suction tank to the T junction 1)
• 50 mm ND (from T junction 1 to T junction 2 , then to the dynamometer’s base)
• 40 mm ND (from the base of the dynamometer to the dynamometer inlet)
The total friction head loss in the whole pipeline will be the sum of the individual friction head
loss in the three pipes mentioned above.
The friction head loss is calculated using the Darcy equation.
From the Darcy equation hf = λlv2/2dig
Where,
λ = coefficient of fluid friction
L = length of pipe
v = velocity of flow
33
di = internal diameter of the pipe
g = acceleration due to gravity
Calculating the coefficient of friction (λ)
The coefficient of friction (friction factor) is a function of Reynolds number (Re) and relative
roughness (k/d), and is found on the Moody diagram.
λ=f (Re, k/d)
Where,
Re = is the Reynolds number
k = internal roughness of the pipe
di = internal diameter of the pipe
i.
For the existing 65 mm nominal diameter pipe (from the suction tank to the T junction 1),
wall thickness, t = 3.65 mm (from TABLE 1)
The inside diameter of the pipe will be given by:
di = d – t
Where,
d – Nominal diameter of pipe
t – Wall thickness of the pipe
From TABLE 1,
d = 65 mm
t = 3.65 mm
di = 65 – 3.65
= 61.35 mm
34
In order to calculate Reynolds number, a flow velocity of between 1 and 3 m/s is chosen. A flow
velocity of 1m/s was chosen so that the pressure loss due to fluid friction in this pipe was kept
within acceptable limits.
Chosen velocity of flow, v1 = 1m/s,
Reynolds number of the flow
Re = v1 * di
υ
Where,
v1 – Velocity of flow
di - Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
From TABLE 2, at a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s)
Re = 1 * 61.35 * 10-3
0.91 * 10-6
=17.42*103
Relative roughness of existing 65 mm pipe
= (k/ di)
Where,
k – inside roughness of pipe
(for steel pipes, k = 1.0 mm)
di - inside diameter of pipe
= 1.0/61.35
= 0.0163
35
Coefficient of fluid friction λ = f (Re, k/di)
From the Moody diagram (Appendix 2)
λ = 0.048
ii.
For the existing 50 mm nominal diameter pipe (from T junction 1 to base of
dynamometer), wall thickness, t = 3.65 mm (from TABLE 1)
The inside diameter of the will be given by:
di = d – t
Where,
d – Nominal diameter of pipe
t – Wall thickness of the pipe
From TABLE 1,
d = 50 mm
t = 3.65 mm
di = 50 – 3.65
= 46.35 mm
From continuity equation Q1 = Q2
v1A1= v2A2
Therefore,
v2= v1 (A1/A2) = v1 ({di1}2/ {di2}2)
Where,
v2 – velocity of flow
A1 – Inside cross section area of 65 mm nominal diameter pipe
A2 – Inside cross section area of 50 mm nominal diameter pipe
di1 – Internal diameter of 65 mm nominal diameter pipe
di2 – Internal diameter of 50 mm nominal diameter pipe
36
v2
= 1 * (61.35)2
(46.35)2
= 1.75 m/s
Reynolds number for pipe flow
Re = v * di
υ
Where,
v - Mean velocity of flow at demand discharge
di - Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
From TABLE 2, at a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s)
Re = 1.75 * 46.35 * 10-3
0.91 * 10-6
= 89.13 * 103
Relative roughness of existing 50 mm pipe
= (k/ di)
Where,
k – inside roughness of pipe
(for steel pipes = 1.0 mm)
di - inside diameter of pipe
= 1.0/46.35
= 0.0216
Coefficient of fluid friction λ = f (Re, k/ di) at demand discharge
From the Moody diagram (Appendix 2)
λ = 0.0514
37
iii.
For the existing 40 mm nominal diameter pipe (from dynamometer base to dynamometer
inlet), wall thickness, t = 3.25 mm (from TABLE 1)
The inside diameter of the will be given by:
di = d – t
Where,
d – Nominal diameter of pipe
t – Wall thickness of the pipe
From TABLE 1,
d = 40 mm
t = 3.25 mm
di = 40 – 3.25
= 36.75 mm
From continuity equation Q1 = Q3
v1A1= v3A3
Therefore v3= v1 (A1/A3) = v1 ({di1}2/ {di3}2)
Where,
v1 – velocity of flow = 1m/s
A1 – Inside cross section area of 65 mm nominal diameter pipe
A3 – Inside cross section area of 40 mm nominal diameter pipe
di1 – Internal diameter of 65 mm nominal diameter pipe
di3 – Internal diameter of 40 mm nominal diameter pipe
v3
= 1 * (61.35)2
(36.75)2
= 2.79 m/s
38
Reynolds number for pipe flow
Re = v * di
υ
Where,
v – Mean velocity of flow
di – Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
From TABLE 2, at a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s)
Re
= 2.79 * 36.75 * 10-3
0.91 * 10-6
= 112.67 * 103
Relative roughness of existing 40 mm pipe
= (k/ di)
Where,
k – inside roughness of pipe
(for steel pipes = 1.0 mm)
di - inside diameter of pipe
= 1.0/36.75
= 0.0272
Coefficient of fluid friction λ = f (Re, k/ di) at demand discharge
From the Moody diagram (Appendix 2)
λ = 0.0571
39
Substituting back in the Darcy equation
Total head loss due to friction, hf, will be given by the sum of the friction head losses in each of
the three pipes mentioned above.
hf = λ1L1v12 + λ2L2v22 + λ3L3v32 …………………………………………. (iv)
2gdi1
2gdi2
2gdi3
But from continuity,
v1A1 = v2A2 = v3A3
Hence,
v1 = v3 * (A3/A1) = v3 * (di32/di12) and v2 = v3 * (A3/A1) = v3 * (di32/di22)
Substituting back for v1 and v2 in equation (iv)
hf = λ1L1v32 di32 + λ2L2v32 di32 + λ3L3v32 …………………………………. (v)
2gdi23
2gdi3
2gdi13
λ1 = 0.048
λ2 = 0.0514
λ3 = 0.0571
L1 = 1.7 m
L2 = 34.532 m
L3 = 1.765 m
di1 = 61.35 mm
di2 = 46.35 mm
di3 = 36.75 mm
g = 9.81 m/s2
hf = 0.048 * 1.7 * v32 * (36.75 * 10-3)2 + 0.0514 * 34.532 * v32* (36.75 * 10-3)2 + 0.0571 * 1.765 * v32 * (36.75 * 10-3)2
2 * 9.81 * (61.35 * 10-3)3 2 * 9.81 * (46.35 * 10-3)3
2 * 9.81 * (36.75 * 10-3)3
= 0.02432552 v32 + 1.22701909 v32 + 0.028887719 v32
= 1.28023 v32
Losses due to bends and fittings = 10 % of the total frictional head loss = 0.128023 v32
Total head loss
= 1.28023v32 + 0.128023 v32
= 1.408253 v32
Substituting back for hL in equation (iii)
40
v3 2
= 2 * 9.81 ({2.757} – 1.408253 v32)
= 54.09234 – 27.62992 v32
28.62992 v32 = 54.09234
v32 = 1.8894
v3 = 1.375 m/s
Flow rate due to gravity, Qg = v3A3
Qg
= 1.375 * π * (36.75 * 10-3)2
4
= 1.459 * 10-3 m3/s
This calculated flow rate due to gravity (1.459 * 10-3 m3/s) is less than the demand flow rate
(3.819*10-3 m3/s). It can therefore be concluded that, a pump is required to boost the flow rate
due to gravity to meet the required demand flow rate.
4.2.2. SELECTION OF BOOST PUMP
A pump is specified in terms of the flow rate it supplies and the total head it overcomes.
4.2.2.1. Flow Rate to Be Supplied By the Pump (Qp).
The flow rate to be supplied by the boost pump is established from the difference between the
demand flow rate and flow rate due to gravity
Flow rate to be supplied by the pump, Qp = Qd – Qg
Where,
Qp
Qd – demand flow rate
= 3.819*10-3 m3/s
Qg – flow rate due to gravity
= 1.459 * 10--3 m3/s
= 3.819*10-3 – 1.459 * 10--3
= 2.36 * 10-3 m3/s
41
4.2.2.2. Total Head in the Pipeline (hL)
The total head to be overcome is established by summing up the total static head (hs), total
friction head loss (hf) and loss due to bends and fittings (hm). The total friction head loss is
calculated for each of the three pipes forming the pipeline route from the suction tank to the
dynamometer’s inlet as described in 4.2.1. above.
Total head loss, hL = hf + hm + hs
Where,
hf – head loss due to friction in the suction pipe
hm – head loss due to fitting and bends
hs – static head to be overcome by the pump
4.2.2.2.1. Static Head to Be Overcome By the Pump (hs)
Static head to be overcome by the pump from the center of the pump to the dynamometer inlet
hs = 0.707 moW
4.2.2.2.2. Head loss due to friction in the pipeline (hf)
The head loss due to friction is established from the Darcy equation. To do this, flow velocities
in each of the pipes in the pipeline was established first.
From the Darcy equation hf = λlv2/2dig
Where,
λ = coefficient of fluid friction
L = length of pipe
v = velocity of flow
di = internal diameter of the pipe
g = acceleration due to gravity
42
Velocity of Flow in the Pipes
Since the flow from the suction tank to the dynamometer is continuous, the equation of
continuity was applied to establish the velocities in each pipe as shown below
From the equation of continuity, Q = vA
Where,
Q - Flow rate
v - Velocity of flow
A - Area of pipe
Therefore,
v1= QP/A1
v2= QP/A2
v3=QP/A3
Where,
Qp – Flow rate to be supplied by the pump
v1 – velocity of flow in 65 mm nominal diameter pipe
v2 – velocity of flow in 50 mm nominal diameter pipe
v3 – velocity of flow in 40 mm nominal diameter pipe
A1 – Inside cross section area of 65 mm nominal diameter pipe
A2 – Inside cross section area of 50 mm nominal diameter pipe
A3 – Inside cross section area of 40 mm nominal diameter pipe
di1 – Internal diameter of 65 mm nominal diameter pipe
43
di2 – Internal diameter of 50 mm nominal diameter pipe
di3 – Internal diameter of 40 mm nominal diameter pipe
But A = π * di2
4
Hence,
v=
QP * 4
π * di 2
Substituting for QP = 2.36 * 10-3 m3/s
di2 = 46.35 * 10-3 m
v1
di1 = 61.35 * 10-3 m,
di3 = 36.75 * 10-3m
= ( 2.36 * 10-3 ) * 4
π*(61.35 * 10-3 )2
= 0.798 m/s
v2
= ( 2.36 * 10-3 ) * 4
π*(46.35*10-3 )2
= 1.399 m/s
v3
= ( 2.36 * 10-3 ) * 4
π*(36.75 * 10-3 )2
= 2.225 m/s
Calculating the coefficient of friction (λ)
The coefficient of friction (friction factor) is a function of Reynolds number (Re) and relative
roughness (k/d), and is found on the Moody diagram ( Appendix 2)
By definition λ=f (Re, k/d)
Where,
Re = is the Reynolds number
k = internal roughness of the pipe
di = internal diameter of the pipe
44
i.
For the existing 65 mm nominal diameter pipe (from the tank to the T junction 1), wall
thickness, t = 3.65 mm (from TABLE 1)
The inside diameter of the pipe will be given by:
di1 = d – t
Where,
d – Nominal diameter of pipe
t – Wall thickness of the pipe
From TABLE 2,
d = 65 mm
t = 3.65 mm
di = 65 – 3.65
= 61.35 mm
For v1 = 0.798m/s
Reynolds number of the flow
Re = v1 * di1
υ
Where,
v1 – velocity of flow
di1 - Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
From TABLE 2, at a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s)
Re
= 0. 798* 61.35 * 10-3
0.91 * 10-6
= 53.799 * 103
Relative roughness of existing 65 mm pipe
= (k/ di1)
45
Where,
k – inside roughness of pipe
(For steel pipes = 1.0 mm)
di1 - inside diameter of pipe
= 1.0/61.35
= 0.0163
Coefficient of fluid friction λ = f (Re, k/di)
From the Moody diagram (Appendix 2)
λ = 0.0457
ii.
For the existing 50 mm nominal diameter pipe (from T junction 1 to dynamometer base
via T junction 2), wall thickness, t = 3.65 mm (from TABLE 1)
The inside diameter of the will be given by:
di2 = d – t
Where,
d – Nominal diameter of pipe
t – Wall thickness of the pipe
From TABLE 1,
d = 50 mm
t = 3.65 mm
di2 = 50 – 3.65
= 46.35 mm
46
Reynolds number for pipe flow
Re = v2 * di2
υ
Where,
v2 - Mean velocity of flow at demand discharge
di2 - Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
From TABLE 2, at a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s)
Re
= 1.399 * 46.35 * 10-3
0.91 * 10-6
= 71.25 * 103
Relative roughness of existing 50 mm pipe
= (k/ di2)
Where,
k – inside roughness of pipe
(For steel pipes = 1.0 mm)
di2 - inside diameter of pipe
= 1.0/46.35
= 0.0216
Coefficient of fluid friction λ = f (Re, k/ di) at demand discharge
From the Moody diagram (Appendix 2)
λ = 0.0556
47
iii.
For the existing 40 mm nominal diameter pipe (from the base of dynamometer to the
dynamometer inlet), wall thickness, t = 3.25 mm (from TABLE 1)
The inside diameter of the will be given by:
di3 = d – t
Where,
d – Nominal diameter of pipe
t – Wall thickness of the pipe
From TABLE 1,
d = 40 mm
t = 3.25 mm
di = 40 – 3.25
= 36.75 mm
Reynolds number for pipe flow
Re = v3 * di3
υ
Where,
v3 – Mean velocity of flow
di3 – Inside diameter of pipe
υ – Kinematic viscosity (m2/s)
From TABLE 2, at a temperature of 25 oC (room temperature), υ = 0.91 * 10-6 (m2/s)
Re
= 2.225 * 36.75 * 10-3
0.91 * 10-6
= 89.856 * 103
Relative roughness of existing 40 mm pipe
= (k/ di)
Where,
k – inside roughness of pipe
48
(for steel pipes = 1.0 mm)
di - inside diameter of pipe
= 1.0/36.75
= 0.0272
Coefficient of fluid friction λ = f (Re, k/ di) at demand discharge
From the Moody diagram (Appendix 2)
λ = 0.0588
Substituting back in the Darcy equation
Total head loss due to friction, hf, will be given by:
hf = λ1L1v12 + λ2L2v22 + λ3L3v32 …………………………………………. (v)
2gdi2
2gdi3
2gdi1
λ1 = 0.0457
λ2 = 0.0556
λ3 = 0.0588
L1 = 1.7 m
L2 = 34.532 m
L3 = 1.765 m
di1 = 61.35 mm
di2 = 46.35 mm
di3 = 36.75 mm
g = 9.81 m/s2
v1 = 0.798 m/s
v2 = 1.399 m/s
hf
v3 = 2.225 m/s
= 0.0457 * 1.7 * 0.7982 + 0.0556 * 34.532 * 1.3992 + 0.0588 * 5.266 * 2.2252
2 * 9.81 * 36.75 * 10-3
2 * 9.81 * 61.35 * 10-3 2 * 9.81 * 46.35 * 10-3
= 0.04110 + 4.1322 + 2.1254
= 6.2987moW
49
4.2.2.2.3. Head Loss Due To Fitting and Bends (hm)
hm = 10% of the head loss due to friction
hence,
hm = 0.62987 moW
Therefore:
Total head loss,
hL
= hf + hm + hs
hL
= 6.2987+ 0.62987 + 0.707
= 7.63557 moW
4.2.3. SELECTION OF PUMP TYPE
Pumps are classified according to their specific speeds. The specific speed is a function of the
flow rate Q and total head to be overcome by the pump hL, and these are tabulated in TABLE 4.
50
TABLE 4: GUIDE FOR PREDICTION OF PUMP TYPE IN WATER PUMPING SYSTEMS
PERFORMANCE REQUIREMENTS TO BE MET ROTODYNAMIC TYPE INDICATED
SPECIFIC SPEED OF <12.5 METRIC UNITS MULTI-STAGE (RADIAL OR MIXED) FLOW
SPECIFIC SPEED OF 12.5 - 96 METRIC UNITS RADIAL (CENTRIFUGAL) FLOW TYPE
SPECIFIC SPEED OF 81 - 175 METRIC UNITS MIXED FLOW TYPE
SPECIFIC SPEED OF >175 METRIC UNITS AXIAL FLOW TYPE
MAXIMUM EFFICIENCY FOR ROTODYNAMIC MACHINES OCCURS IN SPECIFIC SPEED RANGE 29 – 58
SPECIFIC SPEED OF IN METRIC UNITS FOR VARIOUS PRESSURE HEAD AND DISCHARGE FLOWS
OPERATING SPEED OF 1450 RPM
3
Flow Q m /s
(Q) ½
(H)3/4 Head
31.6 moW
100
0.0001
0.001
0.01
0.1
1
2
0.01
0.03
0.1
0.32
1
1.41
Multistage or Positive Displacement Radial Flow Machines
3
4
5
1.73
2
2.24
2.45
2.65
Mixed Flow Machines
6
7
8
9
10
2.83
3
3.16
0
1
5
15
46
65
79
92
103
112
121
130
136
145
95
107
117
126
135
143
151
30.4
29.2
95
90
0
0
2
5
15
48
67
83
2
5
16
50
70
86
99
111
122
131
140
149
157
28
85
1
26.7
80
1
2
2
5
5
16
17
52
54
73
77
90
94
104
108
116
121
127
133
137
143
147
153
155
163
164
171
25.5
75
1
2
6
18
57
80
99
114
127
139
151
161
171
180
24.2
70
1
2
6
19
60
85
104
120
134
147
159
169
180
189
22.9
21.6
65
60
1
1
2
6
20
63
90
110
127
142
155
168
179
190
200
2
7
21
67
95
116
135
150
165
178
190
202
213
20.2
55
1
18.8
50
1
2
2
7
8
23
24
72
77
102
109
124
134
144
154
161
172
176
189
190
204
203
218
215
231
227
244
17.4
45
1
3
8
26
83
118
145
167
187
204
221
236
250
264
15.9
40
1
3
9
29
91
129
158
182
204
223
241
258
273
288
35
30
1
1
3
10
32
101
143
175
202
225
247
267
285
302
319
4
11
36
113
331
196
468
253
277
299
320
339
358
11.2
25
1
9.5
20
2
4
5
13
15
41
48
130
153
183
217
225
266
259
307
290
343
318
376
343
406
367
434
389
460
410
485
7.6
15
2
6
19
60
190
269
330
380
425
466
503
538
571
602
5.6
10
3
8
26
82
258
365
447
516
577
632
682
729
774
815
3.3
1
5
1
4
15
14
43
137
434
613
751
867
970
1062
1147
1227
1301
1371
46
145
459
1450
2051
2511
2900
3242
3552
3836
4101
4350
4585
14.4
12.8
51
From TABLE 4, at H = 7.63557 moW and Q = 2.36 * 10-3 m3/s, the radial (single stage
centrifugal) pump was selected at a standard operating speed of 1450 RPM.
TABLE 5: CHARACTERISTIC DATA OF THEORETICAL SELECTED PUMP TYPE
Type
Head (H)
moW
Radial (centrifugal)
7.63557
Flow rate (Q)
Operating speed
m3/s
(RPM)
2.36 * 10-3
1450
From Appendix 3, at H= 7.63557 moW and Q=2.36 * 10-3 m3/s a radial centrifugal pump with a
total head of 8 moW and a net positive suction head (ns) of 10 moW
TABLE 6: CHARACTERISTIC DATA OF SELECTED PUMP SIZE
Type
Radial (centrifugal)
Head
Net positive suction head
Flow rate
Operating speed
(H)
(ns)
(Q)
(RPM)
moW
moW
m3/s
8
10
2.36 * 10-3
1450
4.3. VALVES
4.3.1. SELECTING A VALVE TYPE.
Valves needed for proper functioning of the dynamometer are manually operated throttle valves
and shut-off valves.
4.3.1.1. Throttling Valves
Throttle valves are used as the main control system of the dynamometer to match the various
operating conditions of the test engine by increasing or reducing the flow rate of the water into
the dynamometer. The valves used here are manually operated by the operator and are fitted both
52
on the intake (as close to the inlet into the dynamometer as possible) and exit pipe of the
dynamometer.
The valve identified for throttling in the present project was globe valve.
Fig. 4 Globe Valve
4.3.1.2. Shut-Off Valves
Shut-off valves are fitted to each of the pipes connected to the engine in order to allow water to
circulate into and out of the engine cooling water tank. The shut-off valves are also fitted to the
cold water supply pipe that picks water from the pump delivery through a reducer. The shut-off
valves are also fitted to each of the pipes connected to the engine to avoid wastage of water when
dismantling engines.
53
Fig. 5 Fixed Spindle Gate Valve
The valves chosen for shut-off in the present project were gate valves.
4.3.2. SELECTION OF GATE VALVES SIZES.
Gate valves are connected at three different positions for various reasons in the pipeline; one
(gate valve V2 = 20 mm ND) is connected to a 20 mm ND rising pipe carrying the hot water
from the engine to the engine water cooling tank, another (gate valve V3 = 40 mm ND) is
connected to a 40 mm ND return pipe carrying water to the engine from the engine water cooling
tank. Another gate valve (V4 = 15 mm ND) is connected to the cold water supply pipe of 15 mm
ND which picks water from the pump delivery through a reducer and discharges the cold water
into the engine water cooling tank.
Sizes of gate valves are determined by the sizes of the pipes they are fitted onto.
54
CHAPTER FIVE
5 DISCUSSIONS
The site selected for the installation of the dynamometer provides a convenient and safe
environment for the operation of the dynamometer. This is because enough distance of 1.718 m
is left between the wall and the dynamometer and a distance of 1.485m is left between the
dynamometer and the main drainage at the centre of the workshop. This space is large enough to
permit several individuals to surround the dynamometer and engine without being burned or
injured during engine testing. This space also provides free movement and coupling of the
engine to the dynamometer as well as carrying out maintenance on the dynamometer.
In addition to this, safety of individuals is ensured by placing the pipes supplying water to the
dynamometer from the main pipeline and those for the engine cooling system in the drain
channels so as not to hinder movement of persons around the dynamometer and engine after
installation. However, the design and construction of drain channels in the site selected did not
provide for the placement of all the pipes inside the channels. The risks of injuries were reduced
by placing the pipes running on the floor of the workshop as close to the dynamometer’s base
and engine cradle as possible.
The drain channels in the workshop are designed to route the exit water to the underground
collection tank as well as provide routes for the pipes. However, some exit drains where utilized
for placing engine cooling water pipes. In order to reduce the effect of the hot water in the drains
which includes corrosion of the steel pipes and heating of the cooling water, the pipes were
placed at half the channel depth.
Since the dynamometer in the present project can be coupled to engines on either of its sides, the
site selected in the engine shop provides enough space for the installation of the engine cradle on
either side of the dynamometer. Due to this reason, a closed end is provided for the completion
of the engine cooling water system on the other side of the dynamometer if another engine is
coupled.
The dynamometer considered in this present project has a maximum absorption power of
559.275 Kw as indicated on the dynamometer. This value is almost twice the capacity of the
55
existing dynamometers in the engine shop and therefore, a larger pump is required to provide the
extra flow rate. For this reason, the system is integrated into a single independent operating unit
by providing its own pump and ensuring coordination of its own and the shared components such
as the suction tank and the engine cooling water tank, without being affected or affecting the
operation of other machines in the workshop. This is achieved by providing new pipelines and
gate valves to isolate flow from the other parts of the workshop and direct it to the dynamometer
when it is running. The space selected provides a convenient location of the pump due to the
availability of an electric terminal switch to run the water pump motor and existing pipeline
terminal on the T junction which provides connection to the pump’s inlet. In addition to this, the
selected site provides a convenient path for pipes from the engine cooling water tank to the
engine.
Through a physical survey of the selected site, the capacity of the overhead tank was established
to be 1232 litres which is sufficient for the proper operation of the dynamometer since the water
system adopted in this present project is a closed loop. The pipeline route consists of 65 mm ND
(from the suction tank to the T junction) and a 50 mm ND (from T junction 1 to T junction 2,
then to the dynamometer base) and their corresponding lengths are 1.7 m and 34.532 m
respectively. It is also noted that there exists a 40 mm ND pipe from the dynamometer base to
the inlet of length 1.765 m.
The surface of water in the overhead suction tank when full is found to be 3.74m above the
workshop flow. This means that there is some flow due to gravity which must be established
after establishing the demand discharge.
The total (dynamic pressure) head that the pump is supposed to overcome partly depends on the
pressure loss due to fluid friction in the pipeline. This pressure loss due to fluid friction in the
pipeline depends on several factors which include pipe size, material, length, fittings and flow
velocity.
In the present project, steel pipes are selected because they are stronger and harder to break; they
are resistant to high pressures; they are easy to connect, install, operate and maintain and they
can withstand shocks and vibrations.
56
All the power absorbed is assumed to be dissipated into the water that flows into the
dynamometer. The maximum power which can be measured by the dynamometer was used to
design for the demand flow rate which the dynamometer requires. By energy balance the demand
flow rate is established using the maximum power, 559.275 Kw (this is indicated on the
dynamometer) and the temperature difference of water between the dynamometer’s inlet and
permissible temperature of water at dynamometer’s outlet.
The water is assumed to enter the dynamometer at ambient temperature of 250C. This
temperature is chosen because the water in the suction tank is expected to be cooled to this
temperature. A small temperature variation of inlet water is expected due to variation in ambient
conditions and due to the pump, but the effect of this variation can be neglected. The exit water
temperature is chosen to be 600C. This is chosen on the basis that temperatures above 600C will
increase lime deposits within the dynamometer and also at temperatures above 600C the water in
the dynamometer could easy flash into steam. Lime deposits will increase wear of the rotor
blades and reduce the amount of water in the stator and subsequently reduce the efficiency of the
dynamometer. Flashing of water into steam reduces the braking effect of the water and hence
reducing the efficiency of the dynamometer. The demand discharge is established to be 3.819 *
10-3 m3/s and this is used in the preliminary selection of the pipe size.
The theoretical nominal diameter of the pipe is found to be 49.31 mm. It can be seen from
TABLE 1 that this value lies between the 40 mm nominal diameter and 50 mm nominal
diameter. The larger nominal diameter (50mm) is selected since the smaller nominal diameter
(40mm) cannot meet the required discharge. The pipe size was selected such that the flow
velocity when the pipeline delivered the design flow rate remained within a specified range of
between 1 m/s and 3 m/s which is an empirical guide as per the Ministry of Water and Irrigation
specifications manual 2005. This is aimed at ensuring that the pressure loss due to fluid friction
in the pipeline is not too high while the discharge flow through the pipeline is not too low.
57
CHARACTERISTIC DATA OF SELECTED PIPE
Q
di
v
m3/s
mm
m/s
(*10-3)
3.819
Re
λ
hf
hm
hs
H
moW
moW
moW
{hf +hm + hs}
(* 103)
46.35
2.26
115.11
moW
0.0525
0.5575
0.05575
0.78
1.39325
The selection of the pipeline is normally done before that of the pump, because pipeline data
influenced the selection of the pump.
Since the suction tank is located above the dynamometer, the need for installing a boost pump in
this present project is determined by examining the flow rate that is delivered by gravity. This
flow rate delivered by gravity is then compared with the required demand discharge. This
calculated flow rate due to gravity was established as 1.459 * 10-3 m3/s which is less than the
demand flow rate of 3.819*10-3 m3/s. It is therefore concluded that, a pump is required to boost
the flow rate due to gravity to meet the required demand flow rate.
In the selection of pipes and pump, the trade-off is therefore between the size of pump and size
of the pipe. The total head to be overcome dictated the size of pump. The size of pump therefore
depended on the friction loss in the pipe and this in turn varied inversely with the pipe size. The
size of pump therefore varied inversely with the size of pipe.
A pump is specified in terms of the flow rate it supplies and the total head it overcomes. The
flow rate to be supplied by the boost pump was established as 2.36 * 10-3 m3/s, from the
difference between the demand flow rate (3.819*10-3 m3/s) and flow rate due to gravity (1.459 *
10--3 m3/s), while the total head to be overcome by the pump was established as 7.63557 moW.
This was established by summing up the total static head (hs = 0.707 moW), total friction head
loss (hf = 6.2987moW) and loss due to bends and fittings (hm = 0.62987moW). The pipeline is
composed of several bents and fitting s of various sizes and therefore the losses due to this are
58
small compared to total friction head loss due to length of pipeline. These losses due to bends
and fittings are approximated to be 10 % of the total friction head loss.
Pumps are classified according to their specific speeds which are a function of the flow rate Q
and total head to be overcome by the pump hL, and these are established as Q = 2.36 * 10-3 m3/s
and hL = 7.63557 moW. At a standard operating speed of 1450 RPM, a radial (single stage
centrifugal) pump was selected.
From Appendix 3, it can be seen that there is no standard pump which matches the total head H
= 7.63557 moW and Q = 2.36 * 10-3 m3/s established above. Therefore the next larger standard
radial pump with a total head of 8 moW and a minimum net positive suction head (ns) of 10
moW is selected. If a smaller pump size is selected, it will not meet the total head H = 7.63557
moW established.
CHARACTERISTIC DATA OF SELECTED PUMP
Type
Radial (centrifugal)
Head
Net positive suction head
Flow rate
Operating speed
(H)
(ns)
(Q)
(RPM)
moW
moW
m3/s
8
10
2.36 * 10-3
1450
In this present project valves needed either for throttling or shutting off flow.
Throttling valves were selected to match the various operating conditions of the test engine by
increasing or reducing the flow rate of water into the dynamometer. The valves selected for
throttling were globe valves because of their ability to accurately control the flow, since they
are efficient in throttling, and also because they can be frequently operated.
Shut-off valves are selected to be used in the present project for on-off service. The design is
not suitable for throttling duty because the sealing surfaces can easily suffer from erosion when
low flows are being maintained against high differential pressures and the design give very poor
flow control characteristics. In this present project, the valves selected for shut-off are gate
59
valves because of their high capacity ability, they are less costly, and they give a tight shutoff, and offer little resistance to flow.
60
CHAPTER SIX
6 CONCLUSIONS AND RECOMMENDATIONS
6.1 CONCLUSIONS
In conclusion therefore, the site of the installation of the dynamometer together with its
components was selected as the space bordered by the main drainage at the centre of the
workshop, the two gas engines and the outside engine shop wall with large window panes.
The system was constructed with proper interaction of components as shown in Appendix 1.
A medium steel pipe of nominal diameter 50 mm was selected to connect the dynamometer to
the existing pipeline on the wall.
A radial (single stage centrifugal pump) with total head of 8 moW and a net positive suction
head (ns) of 10 moW was selected to supply the demand flow rate of 3.819 * 10-3 m3/s at
operating speed of 1450 RPM.
Two throttling valves of size 40 mm ND were identified on the inlet and outlet of the
dynamometer. Four shut off valves (gate valves) of sizes 50mm ND, 40mm ND, 20mm ND and
15mm ND were selected.
6.2 RECOMMENDATIONS
With the first phase of the project completed, the following recommendations are proposed for
the continuity of this project:
1. A complete design of the proposed components of the system.
2. Since the dynamometer has not been in use for a long time, it should be reconditioned.
3. A cost analysis of all the components required for the system should be done.
4. Actual installation and testing of the engine dynamometer should be undertaken.
61
REFERENCES
1.Rosaler, Robert C., Standard Handbook of Plant Engineering, McGraw-Hill, New York, 1995.
2.Purcell, Michael K., "Easily Select and Size Control Valves", Chemical Engineering Progress.
3.Winther, J. B. (1975). Dynamometer Handbook of Basic Theory and Applications. Cleveland,
Ohio: Eaton Corporation.
4.Martyr, A; Plint M (2007). Engine Testing - Theory and Practice (Third ed.). Oxford, UK:
Butterworth-Heinemann.
5.Basusbacher, E. and Hunt, R., ‘Process Plant Layout and Piping Design’, Auerbach
Publishers, Boston, 1990.
6.Sule, D.R., ‘Manufacturing Facilities: Location, Planning, and Design’, Boston, MA: PWSKENT Publishing Co., 1988.
7.Eng. G.O. Nyang'asi Engineering Design 1 and 2 University of Nairobi Department of
Mechanical Engineering.
8.Ministry of water and irrigation manual 2005
9.Heenan dynamatics dynamometers instruction manual 1985
10. http://www.dyno.com.au/dyno/controller
11.John Dinkel, "Chassis Dynamometer", Road and Track Illustrated Automotive Dictionary,
(Bentley Publishers, 2000)
12.'Rankin Kennedy, The Book of Modern Engines, Vol VI', 1912
62
Appendix 2 Moody Diagram
63
APPENDIX 3 GUIDE FOR RADIAL FLOW TYPE WATER PUMP (N = 1450 RPM)
64