Design of Plumbing System And Variation Of Hydrostatic ...

Technological University Lashio Journal of Research & Innovation Vol. 1, Issue: 4

TULSOJRI September, 2020 48

Design of Plumbing System And Variation Of Hydrostatic Pressure For Eighteen Storey High

Rise Building

Khin Zar Ni Myint(1), Thant Zin Oo(2)

(1)Technological University (Thanlyin), Myanmar

(2) Technological University (Thanlyin), Myanmar

Email; [email protected] ABSTRACT: This paper suggests a technique for

including pressure dependent demand and leakage

terms in simulation models for water distribution

systems. In this paper, the plumbing system is designed

for Yadanar Hninsi Residence, Yangon, Myanmar. It

was eighteen storeys, building floor area (669.37 m2)

and building height (61.4 m).1st level included 3 rooms

for commercial rooms. 2nd level included 12 rooms for

residential rooms. From 3rd to 9th and 11th to 18th level

of each floor include 8 rooms for residential rooms. In

10th level included 6 rooms. Among them 4 for

residential and others were halls. There were (696)

persons in this building. According to the calculated

results, the ground tank size and roof tank size are

(8m×7m×3m) and (7m×6m×2m). The observed values

of transfer pipe size and distribution pipe size are 80

mm. The flow rate of transfer pump and booster pump

are (36 m3/ hr). The power of transfer pump and booster

pump are (13 hp) and (4.3 hp). The transfer pump head

and booster pump head are 68.68m and 20.02m

respectively. The hydrostatic pressures are increasing

from 18th floor to 1st floor, when before using the

Pressure Reducing Valve and Booster Pump. The

pressures of 18th , 17th , 16th floors decrease below

1bar and 9th to 1st floors increase above 3 bars.

KEYWORDS: High rise building, Water supply

system, Pumps, Pressure reducing valve, MATLAB

1. INTRODUCTION

Water is a natural resource that is necessary for the

provisions of life and environmental systems, and a key

resource to social and economic development, national

and international organizations from all over the world.

Pipes, drains, fittings, valves and fixtures are installed

for the distribution of water at plumbing system,

heating and washing and waterborne waste removal.

Plumbing system is included a water supply and

distribution pipes, plumbing fixtures, traps, and water

using equipment. In this paper, the plumbing system is

designed for Yadanar Hninsi Residence, Yangon,

Myanmar. These building floor area is (669.37 m2) and

building height is (61.4 m).This building is eighteen

storey residential building, which consists of 55.7418

m2 and 111.4836m2 units. There are 138 residential

rooms and 3 commercial rooms. The plumbing system

includes water supply and distribution pipes: riser, up

feed or down feed distribution pipes, underground tank,

overhead tank, plumbing fixtures, pumps and pressure

reducing valve.

In general, water is one of the key reasons for

human survival and society. Water is careful to be the

most important factor behind the survival of life on

Earth. Daily water requirement for one person is 40

gallons per day according to Singapore Standard CP 48

from Table -1 [1]. Increasing population density and

land prices in cities make high-rise building an

attractive construction option. Such buildings have to be

supplied with water, which requires so-called transfer

pump system and booster pump system. The booster

pump is used to increase the water pressure in order to

reach the uppermost floors in high rise buildings. The

typical layout of pressure boosting systems for water

supply in high-rise buildings is characterized by the

interconnection of components from different physical

domains. The pressure reducing valves are installed in

this high rise building. By using the pressure reducing

valves to prevent excessive pressure. Flow and pressure

demands at any point of the plumbing system are

determined by the calculation are as follows: (i) Sources

of Water, (ii) Water Distribution, (iii) Water Demand,

(iv) Water Storage, (v)Pressure Requirements, (vi) Pipe

sizing,(vii) Pump Selection and (viii)Pressure Reducing

Valve.

2. METHOLOGY

The high rise building tendency increases energy

required for water supply. In this paper, the plumbing

system for the eighteen storeyed high rise building is

designed as per CP 48 code and CQHP (Committee for

Quality control of High-rise building construction

Projects).The result of before and after using Pressure

Reducing Valve and Booster Pump are compared by

using MATLAB.

3. DESIGN CALCULATION

Water Reducing Valves and Pressure Reducing

Valves are installed in residential water systems to

reduce and stabilize inlet pressures from mains water

supplies or boosted water systems, which generally are

too high and variable for domestic appliances to

function correctly. The water from the ground tank is

pumped to the overhead tank by using the transfer

pump. A booster pump is a pump that increases fluid

pressure while maintaining a specified flow rate. The

pressures under 1.5 bar are raised by using the booster

pump. The process of plumbing cycle consists of

source, treatment, supply, distribution, use, collection

and disposal section are shown in Figure - 1.

mailto:[email protected]



Fig 1.Plumbing cycle

3.1 Water supply system and tank size

Treated water is used for daily activities such as

cleansing, washing and plumbing. A proper water

distribution system is needed to ensure a constant flow

of water supply. The type of distribution system largely

depends on the structure of the area. There are three

types of distribution systems which are gravity

distribution system, pumped distribution system and

gravity and pumped combination system. The gravity

and pumped combination system is the most commonly

used system. It is economical, efficient and reliable

system. It uses a pumped system to get the water from

the sources to the treatment plants, the ground tank and

the overhead tanks and then changes to a gravity

distribution system to supply water to the service area

(15th floor to 1st floor). The booster pump is used not to

be enough the water pressure in order to the uppermost

three floor (18th floor to 16th floor) in high rise building.

Human life, as with all animal and plant life on the

Earth, is dependent upon water. Not only the water is

needed to grow the food, generate the power of body,

and run the industries but it is a basic part of the daily

lives. The water is need for the bodies every day to

continue functioning. ‘Basic needs of about 40 gallons

per day according CP 48’ from Table-1 [ 1 ].

Table 1. Recommended Minimum Storage of Cold

Water for Potable Purposes

Type of building Storage in litres

per head

Storage in

gallons per

day

Dwelling houses

and flats

150 40

Hostels

90 24

Hotels

135 36

Schools

15 4

The designed building has eighteen floors. The

first floor has 3 commercial rooms (55.7418 m2).The

second floor to eighteen floor (55.7418 m2 and

111.4836 m2) have 76 and 62 residential rooms. The

number of occupants has 5 people in each residential

room [1]. So the total number of occupants was 696

persons. So total daily water required for whole

building was 27840 gallons per day. It includes the

need for water domestic hygiene sufficient to maintain

health and to maintain a basic standard of personal. The

effects of insufficient water supply cause sickness, high

unit costs, time and energy expended in daily collection,

etc. Provision of basic daily water needs is yet to be

regarded by several countries as a human right finished

water is rising, more in some areas of the world than in

others. [1]

The effective capacity of ground tank and roof tank

sizes can be calculated by equation as follow:

Effective capacity of ground tank =Daily water requirement

0.5 (1)

So the required ground tank size is (8m × 7m × 3m).

Roof tank capacity = 50 % of total domestic demand

per day (2)

Effective capacity of roof tank =0.75

capacity tank Roof (3)

The required roof tank size is two units of (7m length ×

6m width × 2m height) tank. [2]

3.2 Loading units

In this high rise building, there are three

commercial rooms (55.7418 m2), 76 residential rooms

(55.7418 m2) and 62 residential rooms (111.4836

m2).The first floor has three commercial rooms. The

total number of loading units for three commercial

rooms is 15 units [3]. The second floor to eighteenth

floor has 76 residential rooms (55.7418 m2) and 62

residential rooms (111.4836 m2). The loading unit each

of the 55.7418 m2 residential room is 12 units and

111.4836 m2 residential room is 25 units. So the total

number of loading units for second floor to eighteenth

floor is 2462 units. Therefore, the total number of

loading units for the whole building is 2477 units

[3].Table - 2 describes the loading units and flow rates

for each residential room and commercial room [1].

Table 2. Loading units and flow rates

Room

(m2)

Total Loading

Units

(units)

Flow Rates

(m3/s)

55.7418

m2,commercial 5 0.315 × 10-3

55.7418 m2

,residential 12 0.641 × 10-3

111.4836 m2

,residential 25 0.798 × 10-3

3.3 Pipe sizes

The transfer pipe size is used to transfer of the

water from ground tank to overhead tank. The time

taken to fill the overhead tank is assumed 2 hours.

Transfer pump flow rates can get from the roof tank

capacity divided by refilling time. To find the pipe sizes

the flow rates is added 5 % extra use for safety [4].



Transfer pipe size can be calculated by the continuity

equation as follow:

Q = A V (4)

Where, Q = flow rate of water (m3/s)

A = area of the pipe (m2)

V = velocity of the fluid (m/s)

The velocity of water is assumed as 2.1 m/s for

distribution pipe sizes and 0.6 m/s for branch pipe sizes

[3]. The distribution pipe sizes from roof tank to

plumbing fixtures are calculated by the continuity

equation. The gravity flow is used for distributing the

water from roof tank to fixtures (15th floor to 1st floor).

The distribution pipe sizes from booster pump are used

for the three uppermost floors. The booster pump is

used to distribute the water because of the hydrostatic

pressure is not to be enough the three uppermost floors.

The branch pipes from distribution pipe are distributed

to the corresponding rooms. Branch pipe size can be

calculated by the continuity equation.

Table 3. Legends for plumbing fixtures

Symbol Description

WC Water Closet

WB Wash Basin

SH Shower

SK Sink

WM Washing Machine

SK Sink

3.4 Pump power and pump size

Pump is a device that increases pressure while

maintaining a desired flow rate. The transfer pump is

used to transfer the water from ground tank to roof tank.

A booster pump is a pump that increases fluid pressure

while maintaining a specified flow rate. The pressures

under 1.5 bar are raised by using the booster pump. The

water pressure is supplied by using the booster pump

cannot reach the three uppermost floors on the building.

To calculate the transfer pump power, the daily water

demand of the building, the operating hours of the

pump, total head to be pumped and the flow rates of

water must be considered. Normally the filling times of

the pump is taken as 2 to 4 hours. The shaft power of

the pump can be determined by dividing water power

by the pump efficiency. The power of transfer pump

and booster pump can be calculated by:

Transfer and Booster pump power = 𝜌𝑔𝑄𝐻/ƞ (5)

Where, ρ = density of water (kg / m3)

G = acceleration due to gravity (m/s2)

Q = flow rate of water (m3/s)

H = Total head (m) [5]

To get the pump power above the equation ( 5 )

must be know the total head , the flow rates and pump

efficiency .The pump efficiency is assumed 65% .

The total head for the transfer pump can be

calculated by :

Total head = Static head + Total friction loss (6)

Fig 2.Heads in transfer pump system

The total head for the booster pump can be calculated

by :

Total head = Static head + Total friction loss +

Design pressure (7)

[4]

Fig 3.Heads in booster pump system

3.5 Pressure reducing valves and water reducing

valves at inlet of each floor

Water Reducing Valves and Pressure Reducing

Valves are installed in residential water systems to

reduce and stabilize inlet pressures from mains water

supplies or boosted water systems, which generally are

too high and variable for domestic appliances to

function correctly. They are often designed to prevent

excessive pressure at water outlets such as taps, basins,

toilets, dishwashers and other appliances. Hydrostatic

pressure exerted by a fluid at hydrostatic equilibrium on

the contact surface due to gravity. Hydrostatic pressure

increases in proportion to depth measured from the

water surface. The weight of the fluid increased because

of exerting downward force from above. The floors

below have the value of the static head reaching more

than 3 bars, which is not allowable. To avoid this high

pressure need additional pressure reducing valve. The

pressures over 3 bars are controlled by adding the

pressure reducing valve. On the other hand, floor 9 and

the floors below it, have the value of the static head

reaching more than 3 bars, which is not allowable. To

avoid this high pressure need additional pressure

reducing valve. So the pressure reducing valve

additional at 9th floor .The hydrostatic pressure can be

calculated by:

P = γh (8)

Where,

P = Hydrostatic pressure in bar

γ = Specific weight in N/m3

h = Pressure head in m

The pressure reducing valve is set up to decrease

the fluid pressure from 9th to 1st floors. The variation of



hydrostatic pressure after using booster pump and

pressure reducing valve can be calculated by:

P = Hydrostatic pressure + Design Pressure of booster

pump (9)

P = Hydrostatic pressure + Out put pressure of pressure

reducing valve (10)

4. RESULT AND DISCUSSION

In this paper, the plumbing system is designed for

Yadanar Hninsi Residence, Yadanar Hninsi Street,

Tharkayta Township, Yangon Division, Myanmar. . It

is 669.37 m2 (Area) and 61.4 m (Height).After

calculating the daily water requirement, tank sizes, pipe

sizing, pump flow rate, and pump selection, it can be

compared the differences between existing and results.

It is shown in Table 4, 5, 6 and 7.

Table 4. Comparisons of calculated results and existing

data for branches pipe sizes

No Description Symbol Result

Data

Existing

Data

1

Branches pipe size

for (55.748 m2,

commercial room)

(Dbr)1 32mm 25mm

2

Branches pipe size

for (55.748 m2,

residential)

(Dbr)2 40mm 32mm

3

Branches pipe size

for (111.4836

m2,residential)

(Dbr)3 50mm 50mm


data for distribution pipe sizes for transfer pump


Data

Existing

Data

1

Standard

transfer

pipe size

Dtr 80 mm 100mm

2

Distribution

pipe size

(15th to 3rd )

(Ddi )1 100mm 100mm

3

Distribution

pipe size

(2nd to 1st )

(Ddi )2 40mm 50mm


data for distribution pipe sizes for booster pump


Data

Existing

Data

1

Horizontal

distribution

pipe size for

booster pump

(Ddi )3 80mm 100mm

2

Distribution

pipe size for

booster pump

(18th to 16th )

(Ddi )4 50mm 80mm


data for pump flow rate, pump head and pump power

for transfer pump and booster pump


Data

Existing

Data Unit

1

Transfer

pump flow

rate

Qtr 36 60 m3 /

hr

2

Booster

pump flow

rate

Qboost 36 42 m3 /

hr

3 Transfer

pump head Htr 68.68 96 m

4 Booster

pump head Hboost 21.22 31 m

5 Transfer

pump power Ptr 9.6 15 KW

6 Booster

pump power Pboost 3.23 4 KW

The pipe sizes changed in above the comparisons

of Table 4, 5 and 6, because it has different area of the

rooms and the usages of the fixtures are not the same. In

Table 7, the existing data of transfer pump flow rate, the

booster pump flow rate, the transfer pump head, the

booster pump head, the transfer pump power and the

booster pump power are larger than the calculated

results by using equation ( 5 ) , ( 6 ) and ( 7 ). If the

transfer pump flow rate Qtr and the booster pump flow

rate Qboost are changed, the pump sizes, pipe diameters

and water velocities will be changed. Transfer pump

can be replaced by a booster pump. However the

booster pump cannot be a substitute for a transfer pump.

Transfer pumps are used in that area where systems

need more pressure (low to high). Booster pumps are

used in that area where system need low pressure.

Fig 4.Before using the pressure reducing valve and

booster pump

Fig 5.After using the pressure reducing valve and

booster pump



The result of the variation of pressure before and

after using the booster pump and pressure reducing

valve are compared by using MATLAB software. After

that the result charts of comparison are shown in Fig.4

and 5. In Fig. 4, the hydrostatic pressures are increasing

from 18th floor to 1st floor, when before using the

Pressure Reducing Valve and Booster Pump. The

pressures of 18th, 17th , 16th floors decrease below 1bar

and 9th to 1st floors increase above 3 bars when before

using the Pressure Reducing Valve and Booster Pump.

From Fig .5, the pressure of 18th, 17th, 16th floors

increase and 9th to 1st floors decrease by using booster

pump and pressure reducing valve. By using the booster

pump and pressure reducing valve, water pressure in all

stories reach to the optimum water pressure range

(between 1 bar and 3 bar). This Fig .4 and Fig.5 are by

making the comparison of before and after booster

pump and pressure reducing valve.

5. CONCLUSIONS

Water is a basic human need. Water supply system

is very important for any building because anybody

daily activities need water. Plumbing design is a fluid

mechanic problem because of the flow through pipes.

This involves precision pressure head and friction loss,

calculation and accuracy in appliance of various pipes,

fittings, pumps, etc. Based on the study and research

conducted in this paper the flow of water distribution in

this high-rise building is from tube well to ground tank

then use Transfer Pump to pump the water up to roof

tank. Water services in this building are efficient and

suitable to distribute water for the entire building. The

numbers of storage tanks are enough usage for the

people occupying the building. As only two transfer

pumps are used. It is easier to maintenance and also

reducing the cost and energy used. In this paper, the

total number of occupants, the transfer pipe size, the

distribution pipe size, the transfer pump size, the

booster pump size and the condition of before and after

using booster pump and pressure reducing valves are

calculated. Firstly, the flow rates are calculated to get

the pump sizes. The total number of occupants is

calculated by using CP-48. After calculating in each

section, the pump sizes are selected. The booster pump

is used to increase the fluid pressure from 18th floors to

16th floors, where the fluid pressure is less than one bar.

The pressure reducing valves are used to decrease the

fluid pressure from 9th to 1st floors, where the fluid

pressure is greater than three bars. The schematic

diagram of eighteen storey high rise building is shown

in Fig. 6.

ACKNOWLEDGEMENT

The author wishes to express her heartfelt thanks to

each and every one who assisted in completing this

paper.

Finally, the author deep gratitude and appreciation

go to her parents for her moral supports, patience,

understanding and encouragement.

REFERENCES

[1] Anonymous. Code of practice for water

service CP 48. Singapore Productivity and

Standards Board, Building and Industrial

Standard Council, 1989.

[2] T.Christopher Dickenson F.I .Mgt. Valves,

Piping and Pipelines Handbook.3rd ed , 1999.

[3] Anonymous. Plumbing Engineering Design

Handbook, Plumbing Components and

Equipment. U.S.A.: American Society of

Plumbings Engineers, 2008.

[4] Dr.Ali Hammoud,A.“Mechanical and

Electrical Engineering for Building

services”.Plumbing Systems Lecture Notes .

March.

[5] 2017<http://www.scribd.com

[6] Chawlwin, Mechanical and Electrical

Engineering for Building services. Domestic

water supply system, July 2017.

[7] Tayza Zaw. Plumbing & Fire Protection

Systems In Building. Yangon: Future Engineer

Generation, 2015.

[8] C.R. Mohan, and Vivek Anand. Design And

Practical Handbook On Plumbing. Delhi:

Standard Publisher Distributors, 2007.

[9] Anonymous. Committee For Quality Control

Of High-Rise Building Construction Projects

CQHP-01-Sanitary Guide lines for High-Rise

Building. Yangon: Archetype Myanmar Ltd,

2007.

[10] Anonymous. Code of practice for water

service CP 48. Standardisation Department,

Singapore, 2005.

[11] Khin Zar Ni Myint ,2020 . Plumbing and

Sanitation System for High-Rise Building. TU

(Thanlyin), Master Thesis.



Design and Structural Analysis of Impeller Blade for Centrifugal Pump

San San Yi(1), Mi Mi San(2), Yu War Myint(3) (1)Technological University (Maubin), Myanmar

(2) Technological University (Maubin), Myanmar (3) Technological University (Maubin), Myanmar

Email: sansanyi.ptu.mechanical @gmail.com

ABSTRACT: Centrifugal pump is the device for

converting of mechanical energy of the drive shaft to

hydraulic energy of the handling fluid to get it to a

required place or height by using the centrifugal force.

A centrifugal pump can be used for lifting highly

viscous liquids such as oils, muddy and sewage water,

paper pulp, sugar molasses, chemicals etc. This paper

is intended to calculate the design of backward type

impeller for centrifugal pump and structural analysis.

This pump is driven by 17 horse power electric motor

and its speed is 2929 rpm. It can be developed by the

head and pump flow rate are 34.17m and 1.083m3/min

respectively. The impeller design is based on Kyushu

method principally for centrifugal pump. In this paper,

the design of impeller for single-suction centrifugal

pump consists of mainly design and finite element

analysis. A parametric model and analysis of structural

for this design is applied statics pressure and using

Solid Works software simulation. The design data is

taken from Han Sein Thant Engineering, Pump and

Accessories Company limited, Yangon.

KEYWORDS: Centrifugal Pump Impeller Blades

Design, Finite Elements Analysis, Structural Analysis of

Impeller Blades, Aluminum Bronze, SolidWorks

Simulation.

1. INTRODUCTION Pump is machine equipment which is required

to lift from low level to high level or to flow liquid low pressure areas to high pressure areas. Centrifugal pumps are so called due to rotation in an important factor in their operation. The pump consists of associate blade rotating among a case. Water enters the impeller at the center impeller eye, flows radially outward and is discharge from the circumference into the case. During this flow through the impeller the water has received energy from the vanes resulting in an increase both in pressure and velocity. The kinetic energy of water is converted to pressure energy in the casing volute chamber [2].

A Syam Prasad, Bvvv lakshmipathi Rao, A Babji, Dr P Kumar Babu, ‘‘Static and dynamic Analysis of a Centrifugal Pump Impeller’’ Alloys are playing major role in many engineering applications. They offer outstanding mechanical properties flexibility in design capabilities, and ease of fabrication. Additional advantages include light weight and corrosion resistance, impact resistance, and excellent fatigue strength. In this paper study of static and model analysis of a centrifugal pump impeller which is made of three different alloy materials.(viz, Inconel alloy 740.Incoloy alloy 803,Warpaloy0The best material for design of

impeller is Inconel 740.Specific modulus of Inconel 740obtained in static analysis is 10% higher than other material. The nature frequency I modal analysis is 6% higher than other material. The deformation of Inconel 740 in static analysis is reduced by 12%.

S.Rajendran and Dr.K Purushothaman ‘‘Analysis of centrifugal pump impeller is using ANSYS-CFX” In this paper analysis of centrifugal pump impeller design is carried out using ANSYS-CFX. It is most common pump used in industries and domestic application. The complex internal flow in centrifugal pump impeller can predicted by ANSYS-CFX .A centrifugal action of impeller accelerates the liquid to high velocity, transferring mechanical (rotational)energy to the liquid. The flow pattern, pressure distribution in blade passage and blade loading of centrifugal pump impeller are discussed in this paper. Centrifugal pump impeller without volute casing is solved at designed mass flow rate is high. Total efficiency of pump is 30% increases.

I. PROCEDURE FOR PAPER SUBMISSION

Centrifugal pumps are classified supported the way

within in which fluid flows through the pump. The three

varieties of flow through a pump are radial flow, axial

flow, and mixed flow.

A. Radial Flow Pumps

In a radial flow pump, the liquid enters at the

middle of the vane and is directed out on the vane blades

during a direction at right angles to the pump shaft.

B. Axial Flow Pumps

In an axial flow pump, the blade pushes the liquid

in an exceedingly direction parallel to the pump shaft.

Axial flow pumps area unit generally known as

mechanical device pumps as a result of they operate

basically a similar because the mechanical device of a

ship.

C. Mixed Flow Pumps Mixed flow pumps borrow characteristics from both radial flow and axial flow pumps. As liquid flows through the vane of a mixed flow pump, the vane blades push the liquid out off from the pump shaft associated to the pump suction at an angle larger than ninety degrees.

II. THE WORKING OPERATION OF

CENTRIFUGAL PUMP Centrifugal pumps are using to raises the water or

liquid from a lower level to a higher level. Fluid enters axially through eye of the casing, is fixed within the vane blades, and is whirled tangentially and radially outward till it leaves through all circumferential



components of the vane into the diffuser a part of the casing. The working principle is that slurry enters the pump through the eye of rotating impeller which imparts a circular motion [3].

Fig 1 Centrifugal working

III. BLADES OF CENTRIFUGAL PUMPS

IMPELLER

There are three types of Blades;

1. Forward curved blade

2. Backward curved blade

3. Radial blade

Fig 2 Types of Blades

A. Specific Speed The speed in revolutions per minute at associate vane

is operated at that if reduced proportionately in size on deliver one unit of capability against one unit of total head. It is mathematically expressed as [5]

4

3H

QNsn

=

(1)

B. Input Power

η

HρgQL s=

(2)

Rated output of an electric motor

1000trη

L)aF(1rL

+=

(3)

Where,

L = input power (W)

Lr = rated output of an electric motor (kW)

Fa = allowable factor (0.10.4)

= transmission efficiency, 1 (for direct coupling)(%)

Fig. 3 Overall Efficiency Curve

C. Shaft and Hub Diameters

The diameter of the end of the main shaft can be

expressed by the equation:

3N

rLshKcd = (4)

Where,

dc = diameter of the end if main shaft (mm)

Lr = maximum input power (kW)

Ksh =0.112 (the main shaft is made of mild steel) The diameter of hub at the impeller eye is

( )

shd2.0~1.5

hD = (5)

The length of hub at the vane eye measure typically determined from

( ) shd2.0~1.0hL = (6)

The diameter of impeller eye D0 is calculated from the following equation:

2

hD

moVπ

'sQ4oD +

= (7)

Where, Do = diameter of impeller eye

Dh = diameter of hub at the impeller eye

The Velocity at the Eye Suction is

m1momo V)0.35.1(2gHKV == (8)

( ) s0.00023N0.11~0.07moK +=

(9)

Kmo = impeller eye velocity constant

Vmo = velocity at the eye suction (m/s)

g = acceleration due to gravity (m/s2)

H = head (m)

Dh = diameter of the hub (m)

D. Input Power Volumetric Efficiency

The volumetric efficiency is due to the leakage loss

which is a loss of capacity through the running

clearances between the rotation elements and the



stationary the running clearances between the rotation

elements and the stationary casing parts. The volumetric

efficiency is defined as

2/3sn

1.1241

1vη

+

= (10)

Where,

vη = volumetric efficiency (%)

Q = actual discharge at pump outlet (m3/s)

Qs’ = theoretical volume flowing through the impeller

(m3/s)

sn = specific speed (m3/s)

E. Input Power Volumetric Efficiency

The Stepanoff Chart is widely used to decide the

impeller geometry. The parameters Ku, Km1, Km2 are

obtained according to the value of ns. Then, U2, Vm1and

Vm2 are calculated by using the equations [7].

2gHKU u2 = (11)

2gHKV m1m1 = (12)

2gHKV m2m2 =

U2 = vane outlet peripheral velocity (m/s)

Vm1 = vane inlet velocity (m/s)

Vm2 = vane outlet velocity (m/s)

Ku = vane outlet peripheral velocity constant

Km1 = vane inlet velocity constant

Km2 = vane outlet velocity constant

g = acceleration due to gravity (m/s2) H = head (m)

Fig 4 Stepanoff Chart

The impeller outlet diameter, D2 is calculated by the

equation:

N π

2 U60

2D =

(12)

Where,

D2 = impeller outlet diameter (mm)

n = rotational speed (rpm)

The impeller inlet diameter D1 is calculated by the

equation:

1D

2D

2D1D = (13)

Where,

D1 = impeller inlet diameter (mm)

The following calculations are made;

2

1sD1h

D

1mD1D+

== (14)

Where,

D1s = (1.0 ~ 1.1)Do D1h = (0.7 ~ 0.9)Do

Thus, the peripheral velocity at the inlet is

60

N1D π

1U = (15)

Where,

U1 = vane inlet peripheral velocity (m/s) [8]

E. Inlet and Outlet Blade Angles of the Impeller

The blade inlet angle b1β (deg) is obtained by the

equation;

6)(0U

Vtan

U

VKtanβ

1

m11

1

m1b11

b1 +

= −−

(16)

Where,

Kb1 = velocity constant

b1β = blade inlet angle (degree)

The blade outlet angle b2β may be selected within

fairly limits. The angle b2β is usually made between 15˚

and 35˚. It is usually made slightly larger than the inlet

angle b1β to obtain a smooth and continuous passage.

F. Blade Number

The number of blades Z is based upon experience and

fixed after the vane shape has been determined. There

should be enough blades to secure proper guidance of

the liquid. The number of blades generally used is

between 5 and 8.The number of blades is decided using

the Pfleidier formula

+

−

+

2

b2β

b1β

sin

1D2D

1D2D6.5Z

(17)

(19)

Where,

Z = number of blades

D1 = impeller inlet diameter

D2 = impeller outlet diameter



G. Velocity Triangle of Centrifugal Pump

The shape of such vector diagrams is triangular and

they are called velocity triangles. It can be drawn for

any point of the flow path through the impeller, but

usually attention is focused on the entrance and

discharge triangles.

Fig 5 Inlet Velocity

Triangle

Where,

V1 = Absolute velocity at inlet

U1 = Tangential blade velocity

Vr1 = Relative velocity

Vf1 = Radial velocity

Vw1 =

Tangential velocity [8]

Fig 6 Outlet

Velocity Triangle

Where,

V2 = absolute velocity at outlet

U2 = tangential blade velocity

Vr2 = relative velocity

Vf2 = radial velocity

Vw2 = tangential velocity

The water is assumed to enter the vanes radially, so

that the absolute velocity 1α is

90 .

Vf1 = V1, Vw1 = 0

Fig 7

Inlet andOutlet Velocity Triangle

1

1b1

U

Vβtan =

(18)

From Outlet velocity diagram,

m2

f2b2

V

Vsinβ =

(19)

2tanβ

f2V

2UV w2 −=

(20) Impeller Outlet Velocity,

2

w2

2

f22 VVV += (21)

H. Inlet and Outlet Passage Width

The width at the inlet b1 and that of outlet b2 are

respectively decided based on the following equations.

Inlet passage width, [5]

=

Z1S-1πD

1πD

m1V1πD

'sQ

1b

(22)

Where,

'Qs= Flow rate through impeller

( )b1

sinβ1δ1S = (23)

( )b2sinβ2δ2S = (24)

Where,

S1 = water passage width (mm)

S2 = water passage width (mm)

1δ and

2δ are blade thicknesses near the leading

edge and trailing edge respectively. Moreover, S2 can

also be determined by the following relationship

equation [8].

ZS-πD

πD

11

1=

ZS-πD

πD

22

2 (25) (27)

Outlet passage width,

=

ZS-πD

πD

VπD

'Qb

22

2

m22

s

2

(26)

2. SPECIFICATION AND CALCULATION

RESULT OF CENTRIFUGAL PUMP IMPELLER

BLADE

Table 1. Specifications Data of Centrifugal Pump

Parameters Values

Pump head, H 34.17 m

Discharge,Q 1.083 m3/min

Rotational speed, N 2929 rpm

Density of water, ρ 1000 kg/m3

Table 2. Result of Calculation Data for Centrifugal

Pump’s Impeller

No Descriptions Symbols Result

1

Blade Thickness

Near the Leading

Edge

1 2 mm

2

Blade Thickness

Near the Trailing

Edge

2 7.42 mm

3 Inlet Blade Angle βb2 23

4 Smooth Variation in

Velocity S2 12 mm

5 Inlet Passage Width b1 20 mm

6 Outlet Passage b2 12.3 mm

V1

β1 1

U1 Vw1

Vm1

α2

β2

Vw2

U2

V2



Width

7 Hub Length Lh 40.5mm

8 Outlet Impeller

Diameter D2 180.6 mm

9 Inlet Impeller


10 Inner Hub Diameter D1s 95.95 mm

11 Impeller Eye

Diameter Do 91.39 mm

12 Inlet Impeller

Diameter D1m 84.98 mm

13 Outer Hub Diameter D1h 74.02 mm

14 Hub Diameter Dh 47.25 mm

18 Hub Section Shaft

Diameter dsh 27mm

3. FINITE ELEMENT ANALYSIS The finite element technique (FEM) rapidly grew

because the most helpful numerical analysis tool in many fields of engineering and technology. The main advantages are that it will be applied to arbitrary shapes in any vary of dimensions. When the FEM is applied to a specific field of analysis like stress analysis, thermal analysis, or vibration analysis it is often referred to as finite element analysis FEA. SolidWorks Premium CAD software provides the advanced capabilities. A. Mesh Types and Shapes

The shells are triangular with three vertex nodes or

three vertex and three mid-edge nodes in Fig 8(a). The

solids are tetrahedral with four vertex nodes or four

vertex and six mid-edge nodes in Fig 8(b) (c) and (d).

Solid and membrane shell elements use linear and

quadratic interpolation two or three nodes on an edge

[4].

(a) (b) (c) (d)

Fig 8 Solidworks Simulation (a) Shell Element Types

(b) (c)(d) Solid Element Types

B. Structural Component Failure

Structural components can be determined to fail by

various modes determined by buckling, deflection,

natural frequency, strain, or stress. Strain or stress

failure criteria are unit completely different depending

counting on whether or not they are unit thought of as

brittle or ductile materials. Solidworks simulation, and

most finite part systems, default to assumptive a ductile

material and show the distortional energy failure theory

that is typically referred to as the Von Mises stress, or

effective stress, even though it is actually a scalar[6].

4. SOLID MODELING USING SOLIDWORKS

The finite element model of centrifugal pump

impeller blades constructed for the dimensions used the

inner diameter is 86.68mm and outer diameter is 180.6

mm. Inlet blade angle is 18.92 ̊ and outlet blade angle is

23 ̊ with the number of blades as 7. Hub Length is 40.5

mm, hub diameter is 47.25 mm and outer hub diameter

is 74.02 mm respectively. Impeller eye diameter is

91.39 mm and inlet impeller diameter is 84.98 mm. The

impeller is made of aluminum bronze that it can be cast

easily and prevented from corrosion.

Fig 9 Modeling of Centrifugal Pump Impeller Blades

using SolidWorks Software

5. STUDY RESULT OF CENTRIFUGAL PUMP IMPELLER

BLADES

Here we are applying the pressure distribution on

the model is applied statics pressure P = ρgH. Then the

input pressure 335207.7 N/m2 applied the blade curves

and the result for bronze are described as shown in

figures and tables.

Table 3. Material Properties for Impeller Blades

Parameters Aluminum Bronze

Poisson’s Ratio (μ) 0.3

Young’s Modulus (E) 110 GPa

Thermal expansion

coefficient

1.7×10-5/K

Mass Density (ρ)

kg/m3

7400 kg/m3

Tensile Stress (t) 551.485 MPa

Yield Stress (y) 275.742 MPa

Table 4. Volumetric Properties of Impeller Blades

Document

name and

reference

Treated

As Volumetric

properties

CirPattern1

Solid

body

Mass:3.38207 kg

Volume:0.000457036

m^3

Density:7400 kg/m^3

Weight:33.1443 N

C. Study Result of Mesh Information

Fig 10 Mesh Modelling of Impeller Blades



Table 2. Result of Mesh Information

No Descriptions Symbols Result

1

Blade Thickness

Near the Leading

Edge

1 2 mm

2

Blade Thickness

Near the Trailing

Edge

2 7.42 mm

3 Inlet Blade Angle βb2 23

4 Smooth Variation in

Velocity S2 12 mm

5 Inlet Passage Width b1 20 mm

6 Outlet Passage

Width b2 12.3 mm

7 Hub Length Lh 40.5mm

8 Outlet Impeller


9 Inlet Impeller


10 Inner Hub Diameter D1s 95.95 mm

11 Impeller Eye

Diameter Do 91.39 mm

12 Inlet Impeller

Diameter D1m 84.98 mm

13 Outer Hub Diameter D1h 74.02 mm

14 Hub Diameter Dh 47.25 mm

18 Hub Section Shaft

Diameter dsh 27mm

D. Study Result of Static Von-misses Stress

Fig 11 Von-misses Stress

The Fig 11 illustrates the variation of von-misses

stress in the blades applied load. The value of maximum

stress of impeller blade using Aluminum bronze

material is found to be 22.86 MPa and minimum stress

is found to be 0.01285 MPa.

E. Study Result of Displacement

Fig 12

Displacement Result

The Fig 12 illustrates the total deformation of the

centrifugal pump impeller blade. The value of

maximum deformation of impeller blade using

Aluminum bronze material is 0.02752 mm and

minimum deformation is 0.000 mm as shown in figure.

Fig 13 Equivalent Strain Result

The Fig 13 illustrates the variation of effective

strain by using Aluminum bronze material in the

centrifugal pump impeller blades. The value of

maximum strain found to be 0.001060and minimum

strain is found to be 0.0000000008811. In this paper,

structural behaviors of impeller blade using Aluminum

bronze material is analyzed by SolidWorks.

6. CONCLUSION

The calculated impeller design has inlet diameter

86.68mm and outlet diameter 180.6mm. Inlet blade

angle 18.92 ̊ and outlet blade angle 23 ̊ with the number

of blades is 7. The inlet and outlet width are 19.1 mm

and 13 mm respectively. The impeller is made of

aluminum bronze that it can be cast easily and

prevented from corrosion. The max von misses stress

location for Aluminum bronze is at the fixed end of the

blade. Max von misses stress impact at the tip of the

blade because of acting high pressure. Aluminum

bronze is more suitable than other materials for

centrifugal pump impeller blade design. By using

aluminum in the existing blade, weight is reduced and

then, minimum deformation is formed. Their structural

analysis shows that the maximum and minimum von

misses stress, displacement and strain.

ACKNOWLEDGMENT

The author is also wishes to extend special thanks to

Dr. Kyawt Khin, Rector, Technological University

(Maubin) for her permissions of this paper. I would also

like to thank to head of Mechanical Engineering

Department Prof. Mi Mi San and all faculty members

who have timely helped to make my work successfully.

Finally, the authors are thankful to everyone who

assisted in completing this paper.

REFERENCES

[1] pa A Syam Prasad, BVVV Lakshmipathi Rao, A

Babji, Dr P Kumar Babu , “Static and Dynamic

Analysis of a Centrifugal Pump Impeller”

International Journal of Scientific & Engineering

Research, Volume 4, Issue 10, October-2013,

pp966-971



[2] Ausstin, H. Church, “Centrifugal Pump and

Blower”, 1972, John Wiley and Sons, Inc,

New York.

[3] Cho Cho khaing, Nyi Nyi, “Design of Single

Suction Centrifugal Pump and Performance

Analysis by Varying the Speed of Impeller”, July,

2018.

[4] B Karthik Matta, Kode Srividya, Inturi Prakash ,

“Static and Dynamic Response of an Impeller at

Varying Effects” IOSR Journal of Mechanical and

Civil Engineering (IOSRJMCE) e-ISSN: 2278-

1684,p-ISSN: 2320-334X, Volume 11, Issue 1 Ver.

III pp 101-106 (Jan. 2014).

[5] Kyushu Institute Technology, Training Course,

“Fluid Mechanics of Turbomachinery”, Japan,

1996.

[6] S.Rajendran and Dr. K Purushothaman “Analysis

of centrifugal pump impeller using NSYS-CFX”

International Journal of Engineering Research &

Technology (IJERT) Vol. 1 Issue 3, May - 2012 pp

1-6

[7] Stepanoff. A. J, “Centrifugal and Axial Flow

Pumps”, 1957.

[8] Touzson, J, “Centrifugal Pump Design”, John

Wiley and Sons, Inc, 2000.



NUMERICAL ANALYSIS OF CANTILEVER BEAM’S DEFLECTION FOR VARYING

LOADS BY USING SOLIDWORKS

Win Kyaw Thu(1), Khin San Htay(2), Zaw Htway(3)

(1)Technological University (Meiktila), Myanmar

(2)Technological University (Hpa-An), Myanmar (3)Technological University (Lashio), Myanmar

Email: [email protected]

ABSTRACT: The classical problem of maximum

deflection and von-mises stress analysis of a cantilever

beam, under the action of an external vertical

concentrated load at free end and a uniformly

distributed load along its length, is solved by

numerically and analytically. Among the different types

of beams, a cantilever beam has been taken and two

cross-sections are selected (I, H). The design materials

for cantilever beam are Alloy steel, Aluminium alloy

and Malleable cast iron.

The theoretical calculations are done based on the

general Euler-Bernoulli’s Beam Equation. The

numerical analysis is done on Solidworks Simulation

Software. The maximum deflection of beam which

occurs at the point of the applied load is recorded.

Comparing the theoretical results with Solidworks

simulation results, choose the material and design

which gives the minimum deflection and the maximum

yield strength. From all the numerical results and

analytical results, it can be concluded that I-section

beam with alloy steel is the best to do for cantilever

beam than any other materials and cross-sections in this

paper. Although H-section beam has higher yield

strength than I-section beam, it has more deflection than

I-section. This is the main reason of H-section beam is

not used as a cantilever beam. Therefore I-section is the

best design for cantilever beam as it deflection is

minimum. As alloy steel and malleable cast iron have

nearly the same deflection, therefore malleable cast iron

can be used as a cantilever beam instead of alloy steel in

order to save the economic costs.

KEYWORDS: Cantilever beam, Simply Supported Beam,

Overhanging Beam, Maximum deflection, Moment of

Inertia, Varying Load Analysis

1. INTRODUCTION

Basically in general terms, a cantilever beam is a

beam which is fixed at one end. In this paper, behavior

of beams and solid elements have been analyzed on the

basis of deflection and von-misses stress occurred on

cantilever beam due to various types of load i.e point load at free end and uniformly distributed load over the

whole beam. These loads are applied on I-section, H-

section cantilever beams. The materials used for the

beams are Alloy steel, Aluminium alloy and Malleable

cast iron. In this paper, while the beam gets deflected under

the loads, bending moment occurs in the same plane

due to which stresses are developed. Here the deflection

of the beam element is calculated by using Euler-

Bernoulli’s beam equation. Solidworks software has

been used to do the computational analysis. Cantilevers

are widely found in constructions, notably in cantilever

bridges and balconies [3,4,8].

A . CANTILEVER BEAM

A cantilever beam is a beam supported on only one

end. The beam transfers the load to the support where it

has managed to the moment of force and shear stress.

Moment of force is the tendency of a force to twist or

rotate an object. Shear stress is defined as a stress which

is applied parallel to the face of a material.

. In other words, the beam bears a specific weight

on its open end as a result of the support on its closed

end, in addition to not breaking down as a result of the

shear stress the weight would generate on the beam’s

structure. Cantilever construction allows for

overhanging structures without external bracing pillars

[12].

Figure 1.Cantilever Beam Source: [12]

2. EULER BERNOULLI’S BEAM EQUATION

Euler Bernoulli beam theory, also known as

engineer’s beam theory or classical beam theory, is a

simplification of the linear theory of elasticity which

provides a means of calculating the load-carrying and

deflection characteristics of beams. Euler Bernoulli’s

equation describes the relationship between the beam’s

deflection and applied load [11].



Maximum Deflection for Point Load at Free End,

maxδ =

3W L

3 E I

Where,

W = Point load (N)

L = Length of the beam (mm)

E = Young’s modulus (N/mm2)

I = Moment of inertia (N/mm4)

maxδ = Maximum deflection (mm)

Maximum Deflection for Uniformly Distributed

Load over the whole beam,

maxδ =

3ω L

8 E I

ω = W × L

Where,

ω = Uniformly distributed load (N/mm)



I = Moment of inertia (4mm )

maxδ = Maximum deflection (mm)

Cantilever Bending Method

E =

3

3

4×m×g×L

δ×w×t (3.5)

Where,


m = Applied load (g)


w = Width of the beam (mm)

g = Acceleration due to gravity (m/s2)

t = Thickness of the beam (mm)

δ = Deflection (mm)

Moment of Inertia for I-Section Beam

I =

3 3

2 1wh -(w-t )×(h-t )

12

Where,


h = Height of the beam (mm)

1t = Thickness of flange (mm)

2t = Thickness of web (mm)


Moment of Inertia for H-section Beam

I =

3 3

1 1 22(t ×h )+(w-2t )×t

12 Where,


h = Height of the beam (mm)

1t = Thickness of flange (mm)

2t = Thickness of web (mm)


Table 1. Dimensions of I-Beam, H-Beam

Dimensions I Beam H Beam

w 100mm 200mm

h 200mm 100mm

t1 8mm 8mm

t2 5.5mm 5.5mm

L 3000mm 3000mm

W 1000N 1000N

ω 1 N/mm 1 N/mm

Table 2. Mechanical Properties of Alloy Steel,

Malleable Cast Iron and Aluminium Alloy [11]

Mechanical

Properties

Alloy Steel Mallesble

Cast Iron

Aluminium

Elastic

Modulus

2.1×1011

N/m2

1.9×1011

N/m2

7.4×1010

N/m2



(a)Deflection of I-Beam on Alloy Steel by Point Load

(b)Deflection of I-Beam on Alloy Steel by Uniformly

Distributed Load

(c)Deflection of I-Beam on Malleable Cast Iron by

Point Load

(d)Deflection of I-Beam on Malleable Cast Iron by

Uniformly Distributed Load

(e)Deflection of I-Beam on Aluminium Alloy by Point

Load

(f)Deflection of I-Beam on Aluminium Alloy by


(g)Yield Strength of I-Beam on Alloy Steel by Point

Load

(h)Yield Strength of I-Beam on Alloy Steel by


(i)Yield Strength of I-Beam on Malleable Cast Iron by

Point Load

(j)Yield Strength of I-Beam on Malleable Cast Iron by




(k)Yield Strength of I-Beam on Aluminium Alloy by

Point Load

(l)Yield Strength of I-Beam onAluminium Alloy by


(a)Deflection of H-Beam on Alloy Steel by Point Load

(b)Deflection of H-Beam on Alloy Steel by Uniformly

Distributed Load

(c)Deflection of H-beam on Malleable Cast Iron by

Point Load

(d)Deflection of H-Beam on Malleable Cast Iron by


(e)Deflection of H-Beam on Aluminium Alloy by Point

Load

(f)Deflection of H-Beam on Aluminium Alloy by


(g)Yield Strength of H-Beam on Alloy Steel by Point

Load

(h)Yield Strength of H-Beam on Alloy Steel by




(i)Yield Strength of H-Beam on Malleable Cast Iron by

Point Load

(j)Yield Strength of H-Beam on Malleable Cast Iron by


(k)Yield Strength of H-Beam on Aluminium Alloy by

Point Load

(l)Yield Strength of H-Beam on Aluminium Alloy by


4. RESUIT AND DISCUSSION

Table 3.Maximum Deflection of Alloy Steel, Malleable

Cast Iron and Aluminium Alloy by Point Load and

Uniformly Distributed Load on I-Beam

Materials

Young’s

Modulus

(N/mm2)

Maximum Deflection

( mm )

Point

Load

Uniformly

Distributed

Load

Alloy Steel 2.1×105 2.43 2.74

Malleable Cast

Iron 1.9×105 2.69 3.03

Aluminium

Alloy 7.4×104 6.91 7.8

Table 4.Maximum Deflection of Alloy Steel, Malleable

Cast Iron and Aluminium Alloy by Point Load and

Uniformly Distributed Load on H-Beam

Materials

Young’s

Modulus

(2N mm )

Maximum Deflection (

mm )

Point

Load

Uniformly

Distributed

Load

Alloy Steel 2.1×105 32 36

Malleable

Cast Iron 1.9×105 35.3 39.8

Aluminium

Alloy 7.4×104 90.7 102.1

Table 5. Comparison of Deflection and Von-mises

Stress Results

Res

ult

s

Cro

ss-S

ecti

on

Alloy Steel Malleable Cast

Iron

Aluminium

Alloy

P.L U.D.

L

P.L U.D.

L

P.L U.D.

L

Def

lect

ion (

mm

) I 2.4

3 2.74 2.69 3.03 6.91 7.8

H

32 36 35.3 39.8 90.7 102.1

Von-m

ises

Str

ess

(MN

/m2)

I 17.

04 25.55 17.04 25.55 17.04 25.55

H

112

.3 168.4 112.3 168.4 112.3 168.4



In this paper, the maximum deflection of alloy steel

is 2.43 mm, the maximum deflection of malleable cast

iron is 2.69 mm and the maximum deflection of

aluminum alloy is 6.91 mm for I-section beam. The

maximum deflection of alloy steel is 32 mm, the

maximum deflection of malleable cast iron is 35.3 mm

and the maximum deflection of aluminum alloy is 90.7

mm for H-section beam. From all the numerically and

analytically results, it can be concluded that I-section

beam with alloy steel is the best to do for cantilever

beam than any other materials and cross-sections in this

paper. The yield strength of H-section is 112.3 MN/m2

and I-section is 17.04 MN/m2. Although H-section

beam has higher yield strength than I-section beam, it

has more deflection than I-section. This is the main

reason of H-section beam is not used as a cantilever

beam. Therefore, I-section is the best design for

cantilever beam as its deflection is minimum. As alloy

steel and malleable cast iron have nearly the same

deflection, therefore malleable cast iron can be used as

a cantilever beam instead of alloy steel in order to save

the economic costs.

5. CONCLUSIONS

In this study the deflection of I- beam and H-Beam

are compared with analytical and numerical analysis.

The deflection of I-section beam with alloy steel is 2.43

mm and its has the yield strenght of 17.04 MN/m2.

Therefore, I-section is the best design for cantilever

beam as its deflection is miniumm.

ACKNOWLEDGEMENT

I would like to thank my thanks the reviewers for

their valuable comments and suggestions and would

also like to thank all teachers for their valuable support

and encouragement.

REFERENCES

[1] Victor Debnath, Bikramjit Debnath, Deflection

and Stress Analysis Of A Beam On Different

Elements Using Ansys APDL, (IJMET)

Volume 5, Issue 6, June (2014), pp 70-79.

[2] Eduardo Valencia Morales, Alloy Steel

Properties and Use First-Principles Quantum

Mechanical Approach to Stainless Steel

Alloys.

[3] Ashis Kumar Samal, T. Eswara Rao, Analysis

of Stress and Deflection of Cantilever Beam

and its Validation Using ANSYS, Vol. 6, Issue

1, (Part - 4) January 2016, pp.119-126.

[4] K.Varundeep, V.S.Gangadhar, B.B.Vasanth,

A.Gophichand, A.Krishna Priya,

Determination of Stress and Deflection of

Cantilever Beam for Various Cross-Sections.

[5] Yousif K. Yousif, Experimental & Theoretical

Analysis of Composite (Polyester & Silicon-

Carbide) Cantilever Beam, Al-Khwarizmi

Engineering Journal, Vol. 8, No. 3, PP 12-

23(2012).

[6] Tarsicio Belendez, Cristian Neipp and Augusto

Belendez, Numerical and Experimental

Analysis of a Cantilever Beam: a Laboratory

Project to Introduce Geometric Nonlinearity in

Mechanics of Materials, Int.J.Engng Ed,

Vol.19, No.6, pp.885-892, 2003.

[7] A.Kimiacifar, N.Tolou, A.Baran and

J.L.Herder, Large deflection analysis of

cantilever beam under end point and

distributed loads, Journal of the Chinese

Institute of Engineers, 2014 Vol. 37, No. 4,

438–445.

[8] Shang-Hsi Tsai and Heng-Chuan Kan, The

exact solution of the load-deflection model of a

uniformly loaded rectangular cross-section

cantilever beam, J. Phys. D: Appl. Phys. 41

(2008) 095502 (6pp).

[9] Solidworks Software (2016)

[10] J.R. Davis, p351-416, Aluminium and

Aluminium Alloys.

[11] Strength of Materials (Second Edition), S S

Rattan.

[12] https://www.quora.com



Design and Analysis of Piston for Four Stroke Single Cylinder Engine by Using Different

Materials

Arkar Htun (1), Zin Mar Nwe (2), Thin Thin Oo (3)

(1) Department of Mechanical Engineering, TU (Magway), Myanmar

(2) Department of Mechanical Engineering , TU (Magway), Myanmar

(3) Department of Mechanical Engineering, TU (Magway), Myanmar

Email; [email protected]

ABSTRACT: Piston is the major part of an internal

combustion engine, which converts the chemical energy

of the fuel into the mechanical energy obtained at the

crankshaft through the connecting rod. An internal

combustion engine is acted upon by the pressure of the

expanding combustion gases in the combustion

chamber space at the top of the cylinder. This force then

acts downwards through the connecting rod and onto

the crankshaft. The piston design is for 150cc 4-stroke

petrol engine in which the various dimensions of piston

is calculated by analytical method considering

maximum pressure condition and the material

Aluminum alloy 4032-T6 is used. In this paper of the

piston consists of mainly design and analysis. A

parametric model and analysis of piston is using

Solidworks 2019 software. The existing one in the

design and analyzed the different material Aluminum

alloy 2014-T6. Then the analysis becomes completed

on the different parameters (temperature, stress,

deformation) and easily analysis the result. After the

analysis of the different material piston it analyzed that

the Aluminum alloy 4032-T6 is suitable for I.C Engine

piston.

KEYWORDS: Design of Piston, Solidworks, Model of

Piston, Analysis thermal stress, Aluminum alloy 4032-

T6, Aluminum alloy 2014-T6.

1. INTRODUCTION

Engine pistons are one of the most complex

components among all automotive or other industry

field components. The engine can be called the heart of

a car and the piston may be considered the most

important part of an engine. The purpose of the piston is

to transfer the energy to crankshaft via connecting rod.

The piston ring is used to provide seal between the

cylinder and piston.

The main function of the piston is to transfer force from

gas in the cylinder to the crank shaft through connecting

rod. It is very important to calculate temperature

distribution on the piston in order to control thermal

stresses and deformation in working condition.

It is very essential to check out or analyze the stress

distribution, temperature distribution, heat transfer,

thermal load, mechanical load in order to minimize the

stress and different loads on working condition of the

piston.

Fig 1. Piston components for I.C. engine [1]

2. PISTON MATERIALS

The pistons are made of different materials such as

carbon steel, cast iron, aluminum alloys etc. Since the

aluminum alloys used for pistons have high heat

conductivity, therefore, these pistons ensure high rate of

heat transfer and thus keeps down the maximum

temperature difference between the centre and edges of

the piston head or crown. For aluminum alloy pistons,

TC is about 260°C to 290°C and Te is about 185°C to

215°C, but they lose their strength (about 50%) at

temperatures above 325°C. [6]

3. THEORETICAL BACKGROUND

3.1 Design Consideration for Piston

In designing a piston for I.C engine, the following

points should be taken into consideration:

1. It should have enormous strength to withstand the

high gas pressure and inertia forces.

2. It should have minimum mass to minimize the

inertia force.

3. It should form an effective gas and oil sealing of

the cylinder.

4. It should provide sufficient bearing area to prevent

undue wear.

5. It should disperse the heat of combustion quickly

to the cylinder walls.

6. It should have high speed reciprocation without

noise.



7. It should be of sufficient rigid construction to

withstand thermal and mechanical distortion.

8. It should have sufficient support for the piston

pin.[6]

3.2 Design Procedure

The design procedure of piston consists of the

following parameters:

1. Maximum Gas Pressure (P)

2. Thickness of piston head (th)

3. Heat flows through the piston head (H)

4. Radial thickness of the ring (t1 )

5. Axial thickness of the ring (t2)

6. Width of the top land (b1 )

7. Width of other ring lands (b2)

8. Maximum Thickness of Barrel (t3 )

9. Piston wall thickness towards the open end

10. Piston pin diameter (do)

3.3. Maximum Gas Pressure

The maximum gas pressure (P) can be calculated

using the following equation,

= I.P

BP (1)

I.P= Pb ×4

2πD

× L × 2

N (2)

P=10Pb (3)

Its range 9 to 10 times of Pb.

Where,

η = mechanical efficiency

BP = brake power of the engine per cylinder (kW)

IP = indicated power produced inside the cylinder (kW)

N = engine speed (rpm)

L = length of stroke (mm)

D = cylinder diameter (mm)

Pb = brake mean effective pressure (MPa)

P = maximum gas pressure or explosion pressure (MPa)

3.4 Thickness of Piston Head

The piston thickness of piston head calculated

using the following Grashoff’s formula.[1]

th =

t σ

PD

16

23

(4)

Where,

P = maximum gas pressure (MPa)

D = cylinder bore/outside diameter of the piston (mm).

1 = permissible tensile stress for the material of the

piston (MPa) and th = thickness of piston head (mm)

3.5 Heat Flow through the Piston Head

The heat flow through the piston head is

calculated using the equation,

H= C × HCV × M × BP (5)

th = )eTc(TK.

H

−5612 (6)

Where,

H = heat flow through the piston head (kJ/kg hr)

C=constant heat supplied to engine

HCV= higher calorific value of petrol (kJ/kg)

M = mass of fuel used per cycle( kg/BP s)

k = thermal conductivity of material (W/mK)

Tc = temperature at centre of piston head (°C)

Te = temperature at edges of piston head in (°C)

3.6 Radial Thickness of Ring

Ring dimension from the front face touching the

cylinder wall to the back or inside face of the ring.

t1 = D ×

tσ

wP3 (7)

Where,

D = cylinder bore (mm)

Pw = pressure of gas on the cylinder wall(MPa)

σt = allowable bending tensile stress (MPa)

3.7 Axial Thickness of Ring

The thickness of the rings may be taken as

t2 = 0.7t1 to t1 (8)

Where,

t2 = axial thickness of ring (mm)

t1 = radial thickness of ring (mm) [3]

3.8 Width of the Top Land

The width of the top land varies from

b1 = tH to 1.2tH (9)

Where,

b1 = width of the top land (mm)

tH = thickness of piston head (mm) [3]

3.9 With of the other Lands

Width of other ring lands varies from be

calculation from the equation (7).

b2 = 0.75t2 to t2 (10)

Where,

b2 = width of other ring lands (mm)

t2 = thickness of the rings (mm) [10]

3.10 Maximum Thickness of Barrel

t3 = 0.03D + b + 4.5 (11)

Where,

t3 = maximum thickness of barrel (mm)

b = radial depth of piston ring groove.) [3]

3.11 Piston Wall Thickness towards the Open End

t4 = 0.25t3 to 0.35t3 (12)

Where,

t4 = piston wall thickness towards the open end (mm)

3.12 Piston Pin Diameter

do = 0.28 D to 0.38 D (13)

Where,

do = piston pin diameter (mm) [3]



4. PISTON DESIGN CALCULATIONS

The design calculations for the piston is considered

under the maximum pressure condition over the piston.

Table 1. Engine Specifications [4]

Parameters Values

Engine Type Four stroke, Petrol Engine

Induction Air cooled type

No. of cylinder single cylinder

Bore 50 mm

Stroke length 70 mm

Speed 4250 rpm

Brake power 6.5 kW

Compression ratio 8.4

Table 2. Piston Design Calculation Results

Parameters Values

Maximum Gas Pressure 1.7 MPa

Thickness of Piston Head 4.6 mm

Heat Flow through the Piston Head 3.38 mm

Radial Thickness of Ring 1.6 mm

Radial Thickness of Ring 1.12 to 1.6 mm

Width of the Top Land 4.6 to 5.52 mm

Width of the Other Land 0.84 to 3 mm

Maximum Thickness of Barrel 8.00 mm

Piston Wall Thickness Towards 2.0 to 2.8 mm

Piston Pin Diameter 14 to 19 mm

Table 3. Materials Properties of Piston [7]

Parameters Al alloy

4032-T6

Al alloy

2014-T6

Poisson’s Ratio (μ) 0.34 0.33

Young’s Modulus (E) 79 GPa 104.8 GPa

Thermal

Conductivity (k)

138 W/mK 155 W/mK

Thermal expansion

coefficient ()

1.94×10-5/K 2.3×10-5/K

Mass Density (ρ)

kg/m3

2680 kg/m3 2800 kg/m3

Tensile Stress (t) 380 MPa 47 MPa

Yield Stress (y) 315 MPa 415 MPa

5. PISTON MODELING

The design of the piston based on the analytical

calculations is modelled in the software solidworks and

material assigned for this design is of Aluminum 4032 –

T6 material. Drawing and 3D Modelling of the piston is

done using Solidworks software. Design calculations

are done mathematically using various formulae for

finding the different dimensions of the piston. [3]

5.1 Model Information of Aliminum Alloy 2014-T6

using Solidworks Software

Fig 2. Piston Modeling Solid Body Mass : 0.149637 kg Volume : 5.58347e-05 m^3

Density : 2680 kg/m^3

Weight : 1.46644 N

6. ANALYSIS THERMAL STRESS for PISTON

The static and thermal analysis for the piston was

done by finite elements method using solidworks

software. In present examination work we have used

FEA for the Thermal and Structural analysis of piston.

Stress analysis of the piston model has been

performed to obtain the value and parameters at which

the piston would be damaged. Solid works software is

used to prepare the piston modelling and the finite

element analysis. By using this softwar, simulation

results are obtained for two different material. [8]

6.1 Study Result of Piston Head Applied

Temperature 500 C

Here, the temperature distribution on the piston

head is applied 500℃. The result for Aluminum alloy

4032-T6 are described.

Fig 3. Mesh Modelling of Aluminum Alloy 4032-T6

Piston Head applied Temperature 500°C

Table 3. Result of Mesh Information Study of

Aluminum Alloy 4032-T6 Piston Head Applied Temp

500 C

Mesh type Solid Mesh

Mesher Used: Standard

mesh



Jacobian points 4 Points

Element Size 3.82334 mm

Tolerance 0.191167 mm

Mesh Quality Plot High

Total Nodes 17753

Total Elements 9766

Maximum Aspect Ratio 17.251

% of elements with Aspect Ratio < 3 93.3

% of elements with Aspect Ratio > 10 0.236

Time to complete mesh(hh;mm;ss): 00:00:02

Fig 4. Von-misses Stress Result 1

The figure. 4 illustrates the variation of von-misses

stress in the piston head. The value of maximum stress

found to be 1043 MPa. The value of minimum stress is

found to be 0.0278 MPa.

Fig 5. Displacement Result 1

The figure 5 illustrates the total deformation of the

piston. The value of maximum deformation is 0.1304

mm .The value of minimum deformation 000 mm,

which is occurred at the center of piston head as shown

in figure.

Fig 6. Equivalent Strain Result 1

The figure 6 illustrates the variation of equivalent

strain in the piston. The value of maximum strain found

to be 0.008026. The value of minimum strain is found

to be 0.00000001656.

6.2 Study Result of Piston Wall Applied

Temperature 300 C

The temperature distribution on the piston wall is

applied 300℃. The result for Aluminum alloy 4032-T6

are described as shown in figures.


The figure 7 illustrates the variation of von-misses

stress in the piston wall. The value of maximum stress

found to be 516.2 MPa. The value of minimum stress is



The figure 8 illustrates the total deformation of the

piston. The value of maximum deformation is 0.01792

mm. The value of minimum deformation 0.000000 mm

at the piston wall as shown in figure.





to be 0.000003866.

6.3 Study Result of Aluminum Alloy 2014-T6 Piston

Head Applied Temperature 500°C

The temperature distribution on the piston head is

applied 500℃. The result for Aluminum alloy 2014-T6

are described as shown in figures and tables.



Fig10. Mesh Modelling of Aluminum Alloy 2014-T6

Piston Head Applied Temperature 500°C

Table 4. Result of Mesh Study Aluminum alloy 2014-

T6 Piston Head Applied 500 C

Mesh type Solid Mesh

Mesher Used: Curvature-based

mesh

Jacobian points 4 Points

Maximum Element Size 6.97759 mm

Minimum Element Size 1.39552 mm

Mesh Quality Plot High

Total Nodes 6037

Total Elements 3039

Maximum Aspect Ratio 25.3

% of elements with Aspect

Ratio < 3

59.4

% of elements with Aspect

Ratio > 10

0.79

Time to complete

mesh(hh;mm;ss):

00:00:02


The figure 11 illustrates the variation of von-

misses stress in the piston head. The value of maximum

stress found to be 1064 MPa. The value of minimum

stress is found to be 0.02149 MPa.


The figure 12 illustrates the total deformation

of the piston. The value of maximum deformation is

0.1498 mm. The value of minimum deformation

0.00000 mm, which is occurred at the center of piston

head as shown in figure.


The figure 13 illustrates the variation of

equivalent strain in the piston. The value of maximum

strain found to be 0.008892. The value of minimum

strain is found to be 0.00000008881.

6.4 Study Result of Aluminum alloy 2014-T6 Piston

Wall Applied Temperature 300°C


The figure 14 illustrates the variation of von-misses

stress in the piston wall. The value of maximum stress

found to be 544.6 MPa. The value of minimum stress is





The figure 15 illustrates the total deformation of

the piston. The value of maximum deformation is

0.1837mm. The value of minimum deformation

0.000000 mm occurred at the piston wall as shown in

figure.





to be 0.000003217.

7. COMPARATIVE PERFORMANCE RESULT

TABLE 5. TWO DIFFERENT ALUMINUM ALLOY AT

PISTON HEAD AT TEMPERATURE 500C

Parameters Al alloy

4032-T6

Al alloy

2014-T6

Von-misses stress

(MPa)

1043

0.002078

1064

0.02149

Total Deformation

(mm)

0.1304

0000

0.1498

0.000

Equivalent strain

(MPa)

0.008026

0.00000001656

0.008892

0.0000008881

TABLE 6. TWO DIFFERENT ALUMINUM ALLOY AT

PISTON WALL AT TEMPERATURE 300C

Parameters Al alloy

4032-T6

Al alloy

2014-T6

Von-misses stress

(MPa)

516.2

0.4954

544.6

0.314

Total Deformation

(mm)

0.01792

0000

0.1837

0.000

Equivalent strain

(MPa)

0.003425

0.000003866

0.003633

0.000003217

8. CONCLUSION

After doing comparative analysis result through

this work, it is concluded that stress occurred by using

Al-alloy 4032-T6 is lower than the permissible stress

value, so that is best material for piston. The structural

analysis shows that the maximum von Mises Stress

occurs at Aluminum Alloy 2014-T6 at piston head at

displacement 0.1498. Maximum temperature is found at

the center of the top surface of the piston head.

Depending on the thermal conductivity of the materials,

heat transfer rate is found maximum in Al alloy piston.

For the given loading conditions, Al alloy 4032-T6

piston is found most suitable. But when the loading

pattern changes, other materials may be considered.

ACKNOWLEDGEMENT

The authors wish to express their deep gratitude to

Dr. Tin San, Rector, Technological University (Lashio)

for his kindly advice and permission to carry out this

paper. Furthermore, special thanks are extended to the

Technological University (Magway) for their

deliberations and disclosures on the paper.

REFERENCES

[1] Deovrat Vibhandik, Ameya Pradhan, Sampada

Mhaskar, Nikita Sukthankar, Atul Dhale, (2014) ,

Design Analysis and Optimization of Piston and

Determination of its Thermal Stresses Using CAE

Tools, 3(5), pp.273-277.

[2] Gantla Shashidhar Reddy and N. Amara

NageswaraRao, (2013), Modeling and analysis of

diesel engine piston, International journal of

Mathematics and Engineering, 2, pp. 199 – 202.

[3] Jan Filipczyk, Zbigniew Stanik, (2012), Piston

damages-case studies and possibilities of early

detection, Journal of KONES Powertrain and

Transport, 19(4), pp. 179-184.

[4] Lokesh Singh, Suneer Singh Rawat, Taufeeque

Hasan, Upendra Kumar, (2015), Finite element

analysis of piston in ansys, 02, pp. 239-241.

[5] M. T. V and D. S. P. M, "Structural and Thermal

Analysis of Rotor Disc of Disc Brake," International

Journal of Innovative Research in Science,

Engineering and Technology, vol. 2, no. 12, pp.

7741-7749, 2013.

[6] Swati S. Chougule, Vinayak H. Khatawate, (2013),

Piston Strength Analysis Using FEM,International

Journal of Engineering Research and Applications, 3,

pp.124-126.

[7] Vaibhav V. Mukkawar, Abhishek D. Bangale,

Nititn D. Bhusale, Ganesh M. Surve, (2015), Design

analysis and optimization of piston using CAE tools,

International Conference, Pune, India.

[8] S. Srikanth Reddy, Dr. B. Sudheer Prem Kumar,

(2013), Thermal Analysis and Optimization of I.C.

Engine Piston Using Finite Element Method,

International Journal of Innovative Research in

Science, Engineering and Technology, 2, pp. 319-

323.

[9] Vaibhav V. Mukkawar, Abhishek D. Bangale,

Nititn D. Bhusale, Ganesh M. Surve, (2015), Design

analysis and optimization of piston using CAE tools,

International Conference, Pune, India.

[10] Vivek Zolekar, Dr. L. N. Wankhade, (2013),

Finite Element Analysis and Optimization of I.C.

Engine Piston Using RADIOSS and Optistruct, Altair

technology conference.



Design Calculation of a Double Suction Centrifugal Pump (Impeller) by Using MATLAB

Dr Aung Kyaw Soe(1) (1)Technological University (Toungoo), Myanmar

Email: [email protected] ; [email protected] ABSTRACT: In this paper, the basic theory and

detailed design calculation of double suction centrifugal

pump are presented. A double suction centrifugal pump

is chosen to carry out design calculation because it can

be widely used in the field of drainage system,

irrigation work and agriculture.

In this study, the impeller of the double suction

centrifugal pump is designed for delivering the water

flow rate of 5.05 m3/min and 54 m head. The rotational

speed of driving shaft is 2980 rpm. The outer and inner

diameters of designed impeller are 0.220 m (220 mm)

and 0.108 m (108 mm) respectively. Calculated number

of blades is eight and the calculated specific speed is

238. All feasible losses are neglected in design

calculation. Pump efficiency is 77.13 % in this design.

And, design calculation is done by using MATLAB

programming. So, design calculation for different input

data such as head, flow rate and rotational speed can be

carried out with the aid of the MATLAB program.

KEYWORDS: centrifugal, design, impeller, double,

MATLAB

1. INTRODUCTION

In general, pumps are devices which impart

energy to a flow of liquid. Although there are different

types of pumps based on the flow direction, blade

designs, and so on, centrifugal pumps are in the

majority of those used in cleaning systems. Centrifugal

pumps are simple, efficient, reliable, relatively

inexpensive, and easily meet the needs of most cleaning

system requirements including spraying, overflow

sparging, filtration, turbulation and the basic function of

moving liquids from one place to another using

pressure. A centrifugal pump uses a combination of

angular velocity and centrifugal force to pump

liquids. The pump consists of a circular pump housing

which is usually made up of metals, (stain steels etc.)

solid plastic, or ceramics. The outlet extends

tangentially from the diameter of the pump housing.

Inside the pump housing there is a rotating component

an “impeller” which rotates perpendicular to the central

axis and is driven by a shaft secured to its center of

rotation. The shaft, powered by an electric motor,

enters the pump housing through a liquid tight seal

which prevents leaking. Liquid entering the pump

through the inlet is swirled in a circular motion and

displaced from the rotation center of the impeller by

centrifugal force. The combination of the swirling

action (angular velocity) and centrifugal force (radial

velocity) pushes the liquid out of the pump through the

outlet [9]. According to the design and principal of

operation, pumps may be classified into two general

categories: rotodynamic pressure pumps and positive

displacement pumps [5].

The centrifugal pumps may be either single or

multi-stage. In addition, the single-stage centrifugal

pumps may be either single-suction or double-suction.

In single-suction centrifugal pumps, the water enters on

one side of the casing and impeller. The most common

type of multi-suction design in centrifugal pump

engineering is a double-suction pump whose impeller

pair is arranged back to back on the pump shaft. A

double-suction pump with impellers arranged in parallel

can be used to increase the flow rate at a constant head.

Double-suction pumps are employed when the flow rate

required of a centrifugal pump becomes too large for

the inlet cross-sections of one impeller or when the flow

velocity in the inlet cross-section of the first impeller

has to be reduced to prevent cavitation [10]. Figure 1

shows main components of centrifugal pump with

double suction impeller.

Fig.1 Main Components of Centrifugal Pump with

Double Suction Impeller [2]

2. LITERATURE REVIEW

In 2018, design and flow simulation for a

centrifugal pump with double-suction impeller are

carried out with the aid of SolidWorks Flow Simulation

by Eugen-Vlad Năstase. In 2013, computational flow

analysis through a double-suction pump impeller of

centrifugal pump was carried out by Deepak Kumar

Kalyan in India. A double suction centrifugal pump

impeller was modeled and analyzed with the aid of

CFD commercial software in this research. In 2016,

Nwe Ni Win performed design and flow analysis of

double suction centrifugal pump impeller in Myanmar.

In this research, the double-suction impeller blade 3D

model was generated with SolidWorks software and

analyze in flow simulation of the impeller. In 2017,

CFD Analysis of double-suction centrifugal pump with

double volute is carried out by Pranav Vyavahare. In

this research, Impeller vane profile is generated by

tangent arc method and CFD analysis is performed for

1st stage of vertical pump out of 15 stages. Moreover,

head developed by this impeller is calculated and





compared with the required value by using ANSYS-

CFX. In 2009, design and calculation of multi-stage

centrifugal pump impeller was carried out by Aung

Kyaw Soe in Myanmar. In this research, design

calculation of impeller was performed by using

MATLAB programming.

2.1 Specification

The following specifications are used to carry out

the design calculation of single stage double-suction

centrifugal pump (impeller) for fresh water.

Flow Rate, Q=5.05 m3/min

Head, H=54 m

Speed, N=2980 rpm

Density ρ=1000 kg/m3 (for water)

Gravitational Acceleration = 9.81 m2/sec

2.2 Design Theory of Centrifugal Pump

Specific Speed, ns =𝑁𝑄𝑠

1/2

𝐻3/4

(1)

where, ns = specific speed

Q = discharge in (m3/s)

N = pump speed in rpm

H = head per stage in m

Input Power, L=gQsH/ (2)

Rated output of an electric motor Lr(kW) is decided

from the following equation.

Lr=(1+𝑓𝑎)×𝐿

𝜂𝑡𝑟

(3)

Where, fa is the allowance factor and 0.1~0.4 for an

electric motor and larger than 0.2 for engines. tr is the

transmission efficiency, and 1.0 for the direct

coupling, and 0.9~0.95 for the belt drive.

Hub and Shaft Diameters

The diameter of the end main shaft can be calculated by

the equation:

dc=Ksh √Lr

N

3 (4)

Where, Ksh=permissible shear stress factor

dc=diameter of the end of main shaft in m

Lr=output power of an electric motor in KW

The shaft diameter at the hub section should be

selected so as to satisfy dsh>dc. The diameter at the hub,

dh is made from 1.5 to 2 larger than the shaft diameter,

depending on the shaft size. It is usually decided from

the equation:

dh=(1.5~2.0)dsh (5)

The length of the hub, lh at the impeller is also

calculated by the equation:

lh=(1.0~2.0)dsh>dh (6)

Volumetric efficiency, 𝜂𝑣

=1

{1+(1.124

𝑛𝑠2/3)}

(7)

Flow Rate of impeller eye, Qs' = 𝑄𝑠

𝜂𝑣

(8)

The diameter of impeller eye, Do=√4𝑄𝑠′

2𝜋𝑉𝑚𝑜+ 𝑑2

ℎ (9)

Where, Velocity of impeller eye, Vmo =Kmo√2𝑔𝐻

Velocity Coefficient of impeller eye,

Kmo= (0.07~0.11)+0.00023 x ns (10)

The Stepanoff Chart is widely used to decide the

impeller geometry, if the blade outlet angle 2 near 22.5

degree is selected. The parameter Ku (speed constant),

Km1, Km2 and D1/D2 are obtained, since ns is given.

U2= Ku√2𝑔𝐻 (11)

Vm2=Km2√2𝑔𝐻 (12)

Vm1=Km1√2𝑔𝐻 (13)

Where, Vm1 = Meridional velocity at impeller profile

entrance ( m/s)

Vm2 = Meridional velocity at impeller profile outlet m/s)

Km1 = Design speed constant at impeller profile

entrance

Km2 = Design speed constant at impeller profile outlet

Ku = Velocity coefficient of impeller outlet

U2 = Peripheral velocity at outlet (m/s)

The impeller outlet diameter, D2= 60𝑈2

𝑁𝜋 (14)

The outlet vane angle 2 is usually made between

15∘and 35∘. It is usually made slightly larger than the

inlet angle to obtain a smooth continuous passage. And

average value of 22.5∘ of vane outlet angle 2 can be

called 'normal' for all specific speeds.

The Impeller Inlet diameter, D1=D2

D1

D2 (15)

If the inlet edge of the vane is sloped, an average value

is used for D1. D1≈ 𝐷1𝑚 =𝐷1ℎ+𝐷1𝑠

2 (16)

D1s= (1.0~1.1) Do (17)

D1h = (0.7~0.9)Do (18)

The peripheral velocity at the inlet ,U1=𝜋𝐷1𝑁

60 (19)

The blade angle, 1 =tan-1 𝐾𝑏1×𝑉𝑚1

𝑈1 (20)

Where, Kb1=1.1~2.5

Blade Number

Z= 6.5 ( 𝐷2+𝐷1

𝐷2−𝐷1 ) sin (

𝛽1+𝛽2

2 ) (21)

Passage Width

𝑏1 = [𝑄𝑠

′

2𝜋𝐷1𝑉𝑚1] [

𝜋𝐷1

𝜋𝐷1−𝑠1𝑍] (22)

𝑏2 = [𝑄𝑠

′

𝜋𝐷2𝑉𝑚2] [

𝜋𝐷2

𝜋𝐷2−𝑠2𝑍]

(23)

S1=𝛿1

𝑠𝑖𝑛 𝛽1

, S2=𝛿2

𝑠𝑖𝑛 𝛽2

(24)

Where, 1=blade thickness near the leading edge

2= blade thickness at the trailing edge

Impeller Blade Shape

A method to draw the impeller blade by three

circular arcs is used for the present design. Each radius

is given by the corresponding equation. Figure 2 shows

the drawing of impeller blade shapes.

𝜌𝐴

=𝑅𝐴

2 − 𝑅𝐵2

2(𝑅𝐴𝑐𝑜𝑠𝛽𝑏2

− 𝑅𝐵𝑐𝑜𝑠𝛽𝑏)

(24)

𝜌𝐵

=𝑅𝐵

2 − 𝑅𝐶2

2(𝑅𝐵𝑐𝑜𝑠𝛽𝑏

− 𝑅𝐶𝑐𝑜𝑠𝛽𝑐)

(25)

𝜌𝐶

=𝑅𝐶

2 − 𝑅𝐷2

2(𝑅𝐶𝑐𝑜𝑠𝛽𝑐

− 𝑅𝐷𝑐𝑜𝑠𝛽𝑏1

) (26)

Where,

RA=D2

2 , RB=RA-

RA−RB

3,RC=RD+

RA−RD

3, RD=

D1h

2



Fig. 2 Drawing of Impeller Blade Shape [7]

3. CALCULATION BY MATLAB

clc

Q = input('Enter Desired Flow Rate Q (m^3/min) : ' );

H = input ( ' Enter desired Head (H) in m : ');

N = input ( ' Enter Motor RPM ' ) ;

rho = 1000 ;

Ns = ceil ( N * sqrt ( Q/2) / H^0.75 ) ;

Qs = Q/60 ;

Eff = 0.77 ; % From Pump Efficiency Chart (Ns & Q)

g = 9.81 ;

L = rho * g *Qs * H*1e-3 / Eff ;

% if kW <= 1.5

fa = 0.4;

% elseif kW <= 3.7

% fa = 0.39;

% fa = input (' Choose the value of allowance factor

0.4~0.25 : ');

% elseif kW <= 37.3

% fa = 0.25;

% fa = input (' Choose the value of allowance factor

0.25~0.15: ');

% end

etr = 1 ;

Lr = (1 + fa) * L / etr ;

Ksh = 0.125;

dc = 1000 * Ksh * ( Lr/N)^ (1/3) ;

dsh = 1.02*dc;

dh = 1.5 * dsh ;

lh = 1.8 * dsh ;

etv = 1/ (1 + (1.124/Ns^(2/3))) ;

Qsp = Qs/etv ; % m^3 /sec

Kmo = 0.1 + 0.00023 * Ns ;

Vmo= Kmo * sqrt ( 2 * g * H )

Do = sqrt ( 4 * Qsp/(2*pi * Vmo) + (dh/1000)^2 ) ;

Ku = 1.054; % from Stepanoff Chart

U2 = Ku * sqrt ( 2 * g * H) ;

D2 = ceil (1000 * 60 * U2 / (pi * N)) ; % mm

Di = stpnD ( Ns); % from Stepanoff Chart

D1 = D2 * Di ;

D1s = 1.1 * Do ;

D1h = 0.9 * Do ;

U1 = pi * (D1/1000) * N/60 ;

Km1 = 0.1614; % from Stepanoff Chart

Vm1 = Km1 * sqrt ( 2 * g * H) ;

Km2 = 0.119: % from Stepanoff Chart

Vm2 = Km2 * sqrt ( 2 * g * H) ;

beta2 = 22.5;

Kb1 = 1.2 ;

beta1 = ceil ( atand ( Kb1 * Vm1 /U1)) ;

Z = ceil( 6.5 * ( ( D2 + D1) / ( D2 - D1)) * sind ( (beta1

+beta2)/2)) ;

if D2 < 200

St = 2.5 ;

d1 = 2 ;

else

St = 2.5 ;

d1 = 3 ;

end

S1 = ceil( d1 /sind ( beta1)) ;

S2 = ( pi * D2 - ( pi * D2 * ( pi * D1 - S1 * Z)/ ( pi *

D1) ) )/ Z ;

d2 = S2 * sind ( beta2) ;

b1 = Qsp * pi * D1 / ( 2*pi * D1 * Vm1 * ( pi * D1 - S1

* Z)) * 1000 ; % m

b2 = Qsp * pi * D2 / ( pi * D2 * Vm2 * ( pi * D2 - S2 *

Z)) * 1000 ; % m

Ra = D2 / 2

Rd = 1000 * D1h / 2

Rb = Ra - ( Ra - Rd)/3

Rc = Rd + ( Ra - Rd)/3

if beta2 > beta1

bb = beta2 - 1 ;

bc = bb - 1 ;

else

bb= beta2 + 1 ;

bc = bb + 1 ;

end

rhoA = ( Ra^2 - Rb^2) / ( 2 * ( Ra * cosd (beta2) - Rb*

cosd( bb)))

rhoB = ( Rb^2 - Rc^2) / ( 2 * ( Rb * cosd (bb) - Rc*

cosd( bc)))

rhoC = ( Rc^2 - Rd^2) / ( 2 * ( Rc * cosd (bc) - Rd*

cosd( beta1) ))

disp(' ')

fprintf ( ' RESULT TABLE \n\n ' )

fprintf ( 'No. of Blades, Z = %10d Nos. \n \n ' , Z)

fprintf ( 'Shaft Diameter, d_sh = %10.2f mm \n \n ' ,

dsh)

fprintf ( ' Hub Diameter, d_h = %10.2f mm \n \n' , dh)

fprintf ( ' Hub Length, L_h = %10.2f mm \n \n ' , lh)

fprintf ( 'Diameter of Impeller Outlet, D_2 =

%10.2f mm \n \n ' , D2)

fprintf ('Diameter of Impeller Inlet, D_1 = %10.2f

mm \n \n ',D1)

fprintf ( 'Outlet Blades Angle, Beta_2= %10.2f degee

\n \n ', beta2)

fprintf ( 'Inlet Blades Angle, Beta_1 = %10.2f

degree \n \n ' , beta1)

fprintf ( 'Inlet Blade Thickness, S_1 = %10.2f

mm \n \n ' , S1)

fprintf ('Outlet Blade Thickness,S_2 = %10.2f mm

\n \n',S2)

fprintf ( 'Inlet Width, b_1 =%10.2f mm\n \n ' , b1*1000)

fprintf ('Inlet Width, b_2=%10.2f mm \n \n ' , b2*1000)

fprintf ('Pump Efficienc = %10.2f %% \n\n ' , Eff*100)

fprintf ( ' Shaft Power = %10.2f kW \n\n ' , L)

fprintf ( ' Input Power = %10.2f kW \n\n ', Lr)

fprintf ( ' Specific Speeds (Ns)= %10.2f rpm \n\n ', Ns)



4. RESULTS AND DISCUSSIONS

Table 1 shows the calculated result of the impeller

dimension for the sample model. The results are

numerically computed by the MATLAB programming.

Table 1 Results of Impeller Dimension

Shaft Diameter, dsh 38.32 mm

Hub Diameter, dh 57.49 mm

Hub Length, Lh 68.98 mm

Diameter of Impeller Outlet,

D2

220 mm

Diameter of Impeller Inlet,

D1

108 mm

Shaft Power 57.81 kW

Input Power 80.93 kW

Pump Efficiency 77 %

Specific Speed 238

Outlet Blade Angle, 2 22.5 deg

Inlet Blade Angle, 1 21 deg

No of Blades, Z 8 blades

Blade Thickness at Inlet, S1 9 mm

Blade Thickness at Outlet, S2 18.32 mm

Inlet Width, b1 30.83 mm

Outlet Width, b2 41.08 mm

Impeller Blade shapes are the following.

RA= 110 mm , RB= 91 mm, ρA=113 mm ,

RC= 72 mm , RD= 54 mm, ρB= 90 mm ,

ρC= 67 mm

Figure 3 shows two-dimensional view of the

impeller blade profile. The three-dimensional view of

impeller with double suction is shown in figure 4.

Fig. 3 Impeller Blade Profile (2D View)

Fig.4 Three-Dimensional View of Double Suction

Centrifugal Pump Impeller

Table 2 Comparisons of Different Results of Impeller

Dimensions at Different Heads

Head 50 54 60 m

Flow Rates 5.05 5.05 5.05 m3/min

Rotational speed 2980 2980 2980 rpm

Pump Efficiency 77.34 77.13 76.68 %

Shaft Diameter,

dsh

37.32 38.32 39.75 mm

Hub Diameter, dh 55.98 57.49 59.62 mm

Hub Length, Lh 67.18 68.98 71.55 mm

Diameter of

Impeller Outlet,

D2

212 220 232 mm

Diameter of

Impeller Inlet, D1

106 108 110 mm

Shaft Power 53.38 57.81 64.5 kW

Input Power 74.73 80.93 90.3 kW

Specific Speed 252 238 220 -

No of Blades, Z 8 8 7 blades

Blade Thickness

at Inlet, S1

9 9 9 mm

Blade Thickness

at Outlet, S2

17.95 18.32 18.87 mm

Inlet Width,b1 31.84 30.83 28.47 mm

Outlet Width, b2 43.25 41.08 36.94 mm

Table 2 shows the different results of impeller

dimension for the double suction centrifugal pump. In

this study, heads are varied and other input data keep

constant. According to the result table, it can be

observed that the higher the desired head, the larger the

size of the impeller diameter. Also, the values of shaft

and input power have increased when the amount of

required head is increased in design calculations,

whereas impeller diameters and required input power

have decreased when the desired head is reduced.

5. Conclusion

Double-suction pumps are an important type of

centrifugal pump. They are used for large flow rate.

Double-suction pumps have two suction chambers. In

this study, design calculation of impeller is done for the

double-suction centrifugal pump used in water pumping

projects. In the impeller design, diameter of impeller

eye, outlet and inlet diameters, blade angles, number of

blades are computed for the sample parameter of 54 m

of head and 5.05 m3/min of flow rate. The rotational

speed of the drive shaft is 2980 rpm. The outer and

inner diameters of impellers are 0.220 m (220 mm) and

0.108 m (108 mm).

The impeller inlet angle, β1 usually falls in the

range from 10 degree to 25 degree. The calculated value

of inlet blade angle (β1) is 21 degree. And the impeller

outlet blade angle (β2) is assumed as 22.5 degree. The

outlet angle of β2 is made slightly larger than the inlet

angle β1 to obtain a smooth and continuous passage.

The number of blades, Z, depends on the blade angles

and it may be between 5 and 12 blades. According to

the Pleiderer formula, numbers of blade are eight. In

this design, pump efficiency chart from Fluid

Mechanics of Turbomachinery (Kyushu) is used to



estimate the efficiency. The shaft power and input

power are 57.81 kW and 80.93 kW respectively. The

useful technical data from this study can be provided

for designing and manufacturing a double suction

centrifugal pump.

ACKNOWLEDGMENTS

The author is deeply grateful to his parents and to

all teachers that teach us from childhood up to now. The

author also wishes to express his gratitude to those who

share their knowledge and to all of friends who help

him.

REFERENCES

[1] Htay Htay Win, " Design of Double-Suction

Centrifugal Pump Impeller and Casing", Vol.2 ,

Issue 7, June,2019.

[2] Eugen-Vlad Nastae, “Design and Flow Simulation

for a Centrifugal Pump with Double-suction

Impeller 178”, MATEC Web of Conferences 178,

05018(2018).

[3] Nwe Ne Win, “Design and Flow Analysis of

Double-suction Centrifugal Pump Impeller”, Vol:2,

Issue 11, November, 2016.

[4] Deepak Kumar Kalyan, “Computational Flow

Analysis Through a Double-suction Impeller of a

Centrifugal Pump”, 2013.

[5] K.M. Srinivasan," Rotodynamic Pumps (Centrifugal

and Axial), 2008.

[6] Aung Kyaw Soe. “Design and Calculation of Multi-

stage Centrifugal Pump Impeller”, August,2009.

[7] Kyushu Institute Technology, 1996, Training

Course, "Fluid Mechanics of Turbo machinery,

Japan: Kyushu Institute of Technology.

[8] Austin, H,Church, “Centrifugal Pump and Blower ”,

Wiley and Sons, Inc. 1972.

[9] https://blog.softinway.com/ An Introduction to

Centrifugal Pumps

[10] https://www.ksb.com/centrifugal-pump-

lexicon/double-suction-pump

[11] Double suction Split casing Pumps,50Hz, Swiss

Pump Company AG, Switzerland

https://www.ksb.com/centrifugal-pump-lexicon/double-suction-pump

https://www.ksb.com/centrifugal-pump-lexicon/double-suction-pump



Design and Construction of Pretreatment Water Purification System for Technological

University (Lashio)(500L/hr.)

Than Tun Aung(1), May Thu Aung(2), ThihaKyaw(3)

(1)Technological University (Lashio), Myanmar

(2)Technological University(Hmawbe), Myanmar (3)Technological University (Lashio), Myanmar


ABSTRACT: This paper presents is to provide

purified drinking water for Lashio Technological

University at Hopate village with pretreatment of water

purification system. The water source of Hopate is

mostly dissolved calcium. So, the original water result

is considered. Then, three vertical water tanks for

supply source are modified, so that it can treat and

reduce calcium percentage in water.

During this project, Chemical properties of

components (such as active carbon, zeolite, sodium

chloride etc.) are studied and the diameter of the pipes

which connect to vertical tanks by using the theory of

fluid mechanics and gravity flow is calculated. Then

necessary solutions (flow rate of water, power of pump)

the project are calculated.

Water purification is the process of removing

undesirable chemicals, biological containments,

bacteria, suspended solid and gases from water. The

methods used for water purification include such

physical processes as filtration, sedimentation and

gravity flow. In the project, filtration method and

biological filtration processes such as slow sand filters,

biological active carbon and zeolite (Resin) filter were

mainly used.

It is expected to have about 500 liters of purified

water per hour. This project can solve some water

problems in environment.

KEYWORDS: filtration, sedimentation, water flow

rate, rural area.

1. INTRODUCTION

Water is a basic necessity of man along with food

and air. Fresh water resources usually available are

rivers, lakes and underground water reservoirs. About

71% of the planet is covered in water, yet of all of that

96.5% of the planet’s water is found in oceans, 1.7% in

groundwater, 1.7% in glaciers and the ice caps and

0.001% in the air as vapor and clouds. Only 2.5% of the

Earth’s water is fresh water and 98.8% of that water is

in ice and groundwater. Less than 1% of all freshwater

is in rivers, lakes and the atmosphere. [1]

Water purification is the process of removing

undesirable chemicals, biological contaminants,

suspended solids and gases from water. The goal is to

produce water fit for a specific purpose. Most water is

disinfected for human consumption (drinking water),

but water purification may also be designed for a

variety of other purposes, including fulfilling the

requirements of medical, pharmacological, chemical

and industrial applications. The methods used include

physical processes such as filtration, sedimentation, and

distillation; biological processes such as slow sand

filters or biologically active carbon; chemical processes

such as flocculation and chlorination and the use of

electromagnetic radiation such as ultraviolet light. [3]

Purifying water may reduce the concentration of

particulate matter including suspended particles,

parasites, bacteria, algae, viruses, fungi, as well as

reducing the concentration of a range of dissolved and

particulate matter. The standards for drinking water

quality are typically set by governments or by

international standards. These standards usually include

minimum and maximum concentrations of

contaminants, depending on the intended purpose of

water use.


In fluid, dynamic Bernoulli’s principle states that

an increase in the speed of a fluid occurs

simultaneously with a decrease in pressure or a decrease

in the fluid's potential energy. The principle is only

applicable for isentropic flows: when the effects of

irreversible processes (like turbulence) and non-

adiabatic processes (e.g. heat radiation) are small and

can be neglected.

Bernoulli’s principle can be derived from the

principle of conservation of energy. This states that, in a

steady flow, the sum of all forms of energy in a fluid

along a streamline is the same at all points on their

streamline. This requires that the sum of kinetic energy,

potential energy and internal energy remains constant.

Thus an increase in the speed of the fluid – implying an

increase in its kinetic energy (dynamic pressure)-

occurs with a simultaneous decrease in (the sum of) its

potential energy (including the static pressure) and

internal energy. If the fluid is flowing out of a reservoir,

the sum of all forms of energy is the same on all

streamlines because in a reservoir the energy per unit

volume (the sum of pressure and gravitational potential)

is the same everywhere.

Fluid particles are subject only to pressure and their

own weight. If a fluid is flowing horizontally and along

a section of a streamline, where the speed increase it




can only be because it has moved from a region of

lower pressure to a region of higher pressure.

Consequently, within a fluid flowing horizontally, the

highest speed occurs where the pressure is highest and

the lowest speed occurs where the pressure is lowest.

[4]

Bernoulli’s equation is as follow,

2

2

22

1

2

11 Z2g

V

γ

PZ

2g

V

γ

P++=++

(1)

Where,

Atmospheric pressure = 1P

Initial velocity = 1V

Specific weight = γw

Height of tank 1 = 1Z

Pressure (outlet) = 2P

Outlet velocity = 2V

Height of tank 2 = 2Z

Specific gravity = g

Volume flow rate AVQ = (2)

Where,

V = flow velocity

A = cross-sectional vector area/surface

When laminar flow exists in a system, the fluid

flows in smooth layers called laminae. A fluid particle

in one layer stays in that layer. The layers of fluid slide

by one another without apparent eddies or swirls.

Turbulent flow, on the other hand, exists at much higher

flow rates in the system. In this case, eddies and

vortices mix the fluid by moving particles tortuously

about the cross section. The existence of two types of

flow is easily visualized by examining results of

experiments performed by Osborne Reynolds. A

dimensional analysis, when performed and combined

with the data, shows that the criterion for distinguishing

between these flows is the Reynolds number:

υ

ρVDRe = (3)

where: V = average velocity of the flow

D = inside diameter of the tube

For straight circular pipes, the flow is always

laminar for a Reynolds number less than about 2100.

The flow is usually turbulent for Reynolds numbers

over 4000. For the transition regime in between, the

flow can be either laminar or turbulent, depending upon

details of the apparatus that cannot always be predicted

or controlled. We will sometimes need to have an exact

value for the Reynolds number at transition. We will

arbitrarily choose this value to be 2100.

Table 1. Comparison of laminar flow and turbulent

flow[4]

Laminar flow Parameter Turbulent flow

Parabolic Velocity

distribution

Determined

from

experimental

data

For Re less than

or equal 2100

For between

5x105 and 107

2.1 Specification of pretreatment systems

The components of water purification systems are

(1) Raw Water Tanks

(2) Filters Tanks

(3) Sands

(4) Silica Sands, Sediment

(5) Gravel stone

(6) Active Carbon

(7) Birm

(8) Zeolite

(9) Brine

(10) Purified Water Tank

(11) PVC Pipes

(12) Pipe Sockets

(13) Floating Valves

(14) Frames

Silica (SiO2) is the name given to a group of

minerals composed solely of silicon and oxygen.

Extracted ore undergoes considerable processing to

increase the silica content by reducing impurities. It is

then dried and sized to produce the optimum particle

size distribution for the intended application.

For industrial and manufacturing applications,

deposits of silica-yielding products of at least 95% SiO2

are preferred. Silica is hard and chemically inert and has

a high melting point, attributable to the strength of the

bonds between the atoms. These are prized qualities in

applications like foundries and filtration systems. [6]

Fig 1.Sands Fig 2.Silica Sand

Fig 3.Gravel stone Fig 4.Active Carbon

The gravel itself provides mechanical filtration by

catching large free-floating particles. The gravel, as

well as the filter plate, tank bottom, and lift tubes,

provide a bed for the bacteria of a biological filter.

Carbon filtering is a method of filtering that uses a

bed of activated carbon to remove contaminants and

impurities, using chemical adsorption. Active charcoal

filters are most effective at removing chlorine particles

such as sediment, volatile organic compounds (VOVs),

taste and odor from water. Each particle, or granule, of

carbon provides a large surface area, or pore structure,

allowing contaminants the maximum possible exposure

to the active sites within the filter media. One gram of

activated carbon has a surface area in excess of 3,000

m2 (32,000 ft2).



2.2 Design Calculation for pipe diameter

(Tank1,Tank2 and Tank3)

Fig 5. Tanks

Calculation for shaft diameter (Tank 1)

0.914m

0.304m

Fig 6. Water Tank 1

Atmospheric Pressure, 2

0 0N/mP =

Initial velocity, m/s 0V0 =

Specific Weight, 39810N/mγw =

Height of tank (from outlet pipe) 0.8255m=

Pressure,

wshP == (4) 28.0982kN/m2551x9810x0.8 ==

Bernoulli’s equation,

4.2m/sV

17.69V

2x9.81

Vx

9810

8.09820.8255

Z2g

V

γ

PZ

2g

V

γ

P

1

21

22

1

211

0

200

=

=

=

++=++

Table 2. Discharge Value of Tank 1

Diameter (d)

(m)

Area (A)

(m2) 2d π/4A =

Discharge (Q)

(m3/s)

d1 = 1 in

= 0.0254

A1 = 5.07 x 10-4 Q = A1V1

= 0.0021

d2 = 1.5 in

= 0.038

A2 = 1.14 x 10-3 Q = A2V2

= 0.0047

d3 = 2 in

= 0.0508

A3 = 0.002 Q = A3V3

= 0.0085

So we select suitable diameter to get the flow rate of

0.0021m3/s for tank 1 (Outlet)

d = 1 in = 0.0254 m

Calculation for specific weight (Tank 2 )

0.0889m Water 0.0254m

0.228m Sand

0.127m Active Carbon

0.127m Gravel(small)

0.127m Gravel(Large)

0.127m Gravel(Large)

0.0889m 0.508m

Fig 7. Water Tank 2

Insert materials are; (a) Sand, (b) Active Carbon, (c)

Gravel (small, large)

Their specific gravity are described below Table,

Specific Weight, γm= Sγw

Table 3. Specific Gravity and Weight

Material Specific gravity

(s)s

Specific Weight

(γ)N/m3

Sand 2.65 26.094 x 103

Water 1 9.81x 103

Active Carbon 1.32 12.949x 103

Gravel (small) 2.1 20.6x 103

Gravel (Large) 2.5 24.525x 103

Calculation for pressure, Pressure, P = γh

Table 4. Calculation for pressure

No. Material Specific

weight (γ)

N/m3

Height

(h)

(m)

Pressure

(P)

(N/m2)

1 Sand 26.094×103 0.9779 9.5932

2 Water 9.81x103 1.2065 31.48

3 Active

Carbon

12.949×103 1.3335 17.27

4 Gravel

(small)

20.601×103 1.4605 30.08

5 Gravel

(Large)

24.525×103 1.5875 38.93

Table 5. Calculation of head and pressure for the whole

process are

Tank Head (m) Pressure(kN/m2)

Tank 1outlet 0.8255 8.0928

Tank 2 inlet 1.1303 11.088

Tank 2 outlet -10.8 -106.672

Tank 3 inlet -10.49 -102.906

Tank 3 outlet -9.7534 -95.68

Note:We calculate with SI units

Bernoulli’s equation,

Tank 1

Tank 2

Tank 3



54

25

24

5

255

4

244

VV

2x9.81

V7534.9-7366.0

2g

V49.10

Z2g

V

γ

PZ

2g

V

γ

P

=

+=++−

++=++

To find the velocity and pipe diameter, we should know

the head losses and it depend on Reynold’s Number and

type of flow.

Viscosity of water is 0.0091 Poise. The flow cannot be

exceed by 1x105 because Reynold’s no is 194.19 D.

Therefore, by using Bernoull’s formula,

Table 6. Result of pre-treatment tank

Tank Diameter (d)

(m)

Velocity (V)

(m/s)

3 0.0127 0.8047

3 0.0254 0.2011

3 0.0381 0.0894

3 0.0508 0.0502

3 0.0635 0.0322

3 0.0762 0.0223

So we select suitable diameter for tank 3(Outlet)

d = 0.0381m=1.5in, Q = 1.388x10-4 m3/s and with the

velocity of 0.0894 m/s.

3. EXPERIMENT

Original Water Result

Name : TU(Lashio)

Lab Code No. : 034718

Date of Receipt : 4.7.2018

Date of Report : 5.7.2018

Source of Water : TU(Lashio)Hopate Village

Area : LASHIO

No. Post Result Maximum

Permissible

Level

Unit

1. Appearance Slightly

Turbid

-

-

2. Colour(Plati

num,Cobolo

t scale)

7 50 Units

3. Turbidity(Si

lcoda Scale

Unit)

- 25 NTU

4. PH Value 7.7 6.5 to 9.2 mg/l

5. Total Solids 803 1,500 mg/l

6. Total

Hardness (as

CaCO3)

340 500 mg/l

7. Total

Alkalinity

(as CaCO3)

520 950 mg/l

8. Calcium as

Ca

120 200 mg/l

9. Magnesium

as Mg

10 150 mg/l

10. Chloride as

CL

20 600 mg/l

11. Sulphate

asSO4

78 400 mg/l

12. Total Iron as

Fe

Nil 1 mg/l

remarks :chemical portable

check by :ministry of health and sports public health

laboratory(Mandalay)

3.1 Results Data of Pre-treatment Tanks

Pretreatment Tank

Name : TU(Lashio)

Lab Code No. : 1851019

Date of Receipt : 22.10.2019

Date of Report : 23.10.2019

Source of Water : Pretreatment Tank

Area : LASHIO

No

.

Post Result Maximum

Permissib

le Level

Unit

1. Appearance Clear - -

2. Colour(Platinum,

Cobolot scale)

5 50 Unit

s

3. Turbidity(Silcoda

Scale Unit)

- 25 NT

U

4. PH Value 8.1 6.5 to 9.2 mg/l

5. Total Solids 602 1,500 mg/l

6. Total Hardness (as

CaCO3)

40 500 mg/l

7. Total Alkalinity (as

CaCO3)

455 950 mg/l

8. Calcium as Ca 8 200 mg/l

9. Magnesium as Mg 5 150 mg/l

10. Chloride as CL 40 600 mg/l

11. Sulphate asSO4 39 400 mg/l

12. Total Iron as Fe Nil 1 mg/l



4. ANALYSIS

Fig 8.Comparession of final water results with initial

water results

In tank 1 and 3 there is no negative upward force to

water flow but in the tank 2 head losses and pressure

losses occurs because of inserting materials (i.e., sand,

active carbon, etc.). So, the velocity of water is

changing depend on pressure and head. The velocity

and pipe diameter can change depend on flow rate. To

get to the flow rate of 500l/hr. with the velocity 0.08m/s

we should choose 1.5in diameter.

At the end of water treatment operation, water is

ready to drink. Final water result is checked by ministry

of health and sports (Public Health Laboratory,

Mandalay). This treatment system can mainly reduce

Total solid from 803 mg/L to 660 mg/L as 17 %,Total

Hardness from 340 mg/L to 20 mg/L as 94.1 %,

Sulphate (SO4) from 78 mg/L to 59 mg/L as 24.36 %

and Calcium amount from 120 mg/L to 4 mg/L as 96.67

% .Remark is Chemical Portable. Solar power supply is

absolutely enough for system. So, it is helpful to utilize

in condition with no electricity. It can also impact on

environmental cleanness.

5. CONCLUSIONS

This research project, pretreatment of water purification

system is a modifying type of water treatment. Before

passing raw water through each layer, raw filtration for

local water is considered. So, three vertical water tanks

were used for this process. This is main modifying

method of this project. Due to this method, it can rather

reduce the working rate of machine. This is mainly

considered in local water source. The sources water in

Northern Shan State involves large amount of calcium.

Thus, local people are infected by kidney related

diseases. By this project, it can reduce the health

problem and maintain water and sanitation.

ACKNOWLEDGEMENT

Firstly, the authors wish to express the deepest

gratitude to Dr. Tin San, Pro rector (acting),

Technological University (Lashio), for his kindness,

idea and support for the completion of this research

project.

The authors are deeply grateful and sincerely

thankful to their all teachers, for their suggestions and

valuable discussions during the research project.

Finally, the authors are thankful to everyone who

assisted in completing this research project.

REFERENCES

[1] Chittaranjan Ray and Ravi Jain(Auth) 3rd

ed.,Butterworth Heinemann,Press,2014.

[2] Rajindar Singh, Hybrid Membrane for Water

Purification: Technology, System Design and

Operations, 1ed, Elsevier Science,2006.

[3] Nikolaj Gertsen, Linus Sonderby, Water

Purification ( Air, water and Soil pollution

Science and technology series ), first edition,

press 2009.

[4] William S. Janna, Fluid Mechanic, Fourth

edition, University of Memphis Memphis,

Tennessee, U.S.A., 2010.

[5] Flow of water from tank.

http://www.wikepedia.com

[6] Water purification system.

http://www.waterpurification.com

0

100

200

300

400

500

600

700

800

900

Colo

ur

PH

Val

ue

Tota

l S

oli

d

Tota

l H

ardnes

s

Alk

alin

ity

Cal

cium

Mag

nes

ium

Chlo

ride

Sulp

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mg/l

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Comparession of final water results with

initial water results

Iniitial water results Final water results

http://www.wikepedia.com/

http://www.waterpurification.com/



Design, Construction and Performance Testing of Pico Hydro Turbine for Rural

Electrification

Ei Ei Mon(1), Cho Cho Khaing(2)Aung Zaw Lynn(3)

(1)Technological University (Mandalay), Myanmar

(2) Technological University(Mandalay), Myanmar


ABSTRACT: The generation of electricity from water

is the most effective and the cheapest way to get

energy. The pico scale renewable energy is to bring

electricity to remote villages that are not near

transmission lines. In hydropower plant water turbine is

one of the most important parts to generate electricity.

The main purpose of this project is to develop the living

standard in rural areas and to reduce the use of non-

renewable energy. In this research paper, 1 kW pelton

turbine design is based on head and flow rate of

Department of Research and Innovation. The available

head and flow rate are 10 m and 0.02 m3/sec. The

pelton turbine is a tangential flow impulse turbine.

There are two main components of this turbine namely,

runner and nozzle. A 1 kW medium head hydro

concrete turbine is constructed in Taung Da Gone

Industrial 3 at Yangon. The pelton turbine project was

tested at Department of Research and Innovation.

KEYWORDS: Nozzle, Pelton turbine, Pico Hydro,

Renewable Energy, Runner.

1. INTRODUCTION

Nowadays, renewable energy such as bio fuel energy,

wind energy, solar energy, geothermal energy, biomass

energy and hydropower energy is very important.

Firstly, renewable energy technologies are clean energy

that has a much lower environmental impact than

conventional energy technologies. Small scale hydro

power generation is one of the types of renewable

energy. Hydropower systems use the energy in flowing

water to produce electricity or mechanical energy. In

hydro power plants the kinetic energy of falling water is

captured to generate electricity. A turbine and a

generator convert the energy from the water to

mechanical and then electrical energy.

There are several classifications related to the

dimension of hydropower plants. An actually useful

classification is the following.(A Harvey and Brown

1992)

(i) Large hydropower > 100 MW

(ii) Medium hydropower 15 -100 MW

(iii) Small hydropower 1 MW-10 MW

(iv) Mini hydropower 100 kW-1MW

(v) Micro hydropower 5-100 kW

(vi) Pico hydropower up to 5kW

Typical hydroelectric plant is shown in figure 1.

Figure 1. Typical Hydroelectric plant (Sharma, R.K.

2003)


A turbine is a rotary mechanical device that extracts

energy from a fluid flow and converts it into useful

work. The work produced by a turbine can be used for

generating electrical power when combined with a

generator. Turbines are also divided by their principle

of operation and can be divided into impulse and

reactions turbine. The impulse turbine generally uses

the velocity of the water to move the runner and

discharges to atmospheric pressure. Reaction turbines

are pressure type turbines that rely on the pressure

difference between both sides of the turbine blades. For

micro-hydro applications, Pelton turbines can be used

effectively at head down to about 20m. Draft tubes are

not required for impulse turbine since the runner must

be located above the maximum tail water to permit

operation at atmospheric pressure. Impulse turbines are

usually cheaper than reaction turbines because there is

no need for a special pressure casing or for relatively

high heads.

A Pelton turbine is a hydraulic turbine where the

runner is rotating from the impulse of water jet on its

buckets. The Pelton wheel is a special type of axial flow

impulse turbine and is used for very high heads. In large

scale hydro installation, Pelton turbines are normally

only considered for heads above 100m. A Pelton

turbine consists of a set of specially shaped buckets

mounted on a periphery of a circular disc as shown in

figure 2. The runner consists of a circular disc with a

Intake

Penstock

Power House

Tail Race

Head Race



number of buckets evenly spaced round its periphery.

The rim of the runner disc are fastened bucket-shaped

blade which are for a better discharge of the water

divided by a ridge or splitter into two symmetrical parts.

The water jet is deflected by the bucket and thus

transfers its energy to the wheel. In order to achieve the

most efficient position of the bucket for the impinging

water, a notch is made into the edge of the bucket at the

largest radius.

Figure 2. Components of Pelton turbine

2.1 Selection of Turbine

The choice of water turbine depends on the site

conditions, notably on the head of water H and the flow

rate Q. figure 3 indicates which turbine is most suitable

for any particular combination of head and flow rate.

Reaction turbines suited for low head and high flow

rate. Pelton turbine is suitable for high head and low

flow rate.

Figure 3. Choice of Turbine in Terms of Head and

Flow rate

Site selection is a very important factor. The

amount of power obtained, the expense of installation,

and even the applications for which the power can be

used may be determined by the quality of the site. A

careful analysis of the site is necessary in order to

determine the feasibility of the site for use of any kind

and the amount of power obtainable from the site.

Specification data of 1Kw pelton turbine are head 10 m

and flow rate 0.02 m3/sec.

2.2 Design Consideration of Pelton Turbine

The effective head and power available of this Pelton turbine is considered at 10m and 1 kW. The power developed by a turbine is given by the following equation.

P = ηo ρ g Q H (1)

The required shaft power is 1.47 kW.

The specific speed can be calculated from the following equation.

Ns = 45(H)

PN

(2)

The speed of the turbine can be calculated from the following equation.

N=147.7√𝐻 (3)

The following points should be considered while designing a Pelton turbine. The absolute velocity of water at inlet can be obtained by using this equation

V1 = Cv 2gH

(4)

The tangential velocity of wheel is determined the

following factors.

k u =

1V

u (5)

The mean diameter or the pitch circle diameter of the Pelton turbine is known from this equation.

u =60

πDN (6)

Number of nozzle is single jet. Thus the jet diameter of

1000 kW Pelton turbine can be calculated from this

equation

Hz

Q0.545d

0

0 = (7)

Jet ratio is a size parameter for the turbine. This value can be obtained by using this equation.

m = d

D (8)

The number of buckets required for the efficient operation of the Pelton turbine is calculated from this equation.

z = 15+ 0.5m (9)

Table 1, which presents a variation of number of

buckets with jet ratio.



TABLE 1. Approximate Number of Buckets for a Pelton Turbine

Jet ratio 6 8 10 15 20 25

No: of

bucket

17-

21

18-

22

19-

24

22-

27

24-

30

26-

33

The blade pitch p1p2 on the pitch circle can be obtained

by using this equation.

z

R 2πpp 21 = (10)

The relative velocity with the direction of motion of the vane at outlet is 15˚ Velocity triangle from figure 4.

= 15˚

Figure 4. Inlet and outlet velocity diagram of Pelton

Turbine

(i) Relative velocity of water at inlet

Vr1 = V 1 - u1

(11)

(ii) Whirl velocity of water at inlet and outlet

Vw1 = V1

Vw2 = Vr2 cos - u2 (12)

(iii)Flow velocity of water at outlet

sin =

r2

f2

V

V (13)

(iv) Angle at exit runner

tan β =

w2

f2

V

V (14)

The force exerted by the jet of water to the direction of

motion is given as

Jet force on the runner, F = aV1 (Vw1+Vw2) (15)

Jet force on the bucket can be obtained by using this

equation

( )

( )cosα1V

uVρaVP

1

2

110 −

−= (16)

The centrifugal force on the bucket can be calculated.

C.F = F - P0 (17)

Weight of the bucket is

gR

GuC.F

2

= (18)

In a Pelton turbine design, two parameters are

important.

(i) the ratio of the bucket width to the jet diameter

and

(ii) the ratio of the wheel diameter to the jet diameter

If the bucket width is too small in relation to jet

diameter, the fluid is not smoothly deflected by the

buckets and in consequence, much energy is dissipated

in turbulence and the efficiency drops considerably.

Table 2 shown in the main dimensions of the Pelton

wheel bucket.

TABLE 2. Dimension of Bucket with Respect to Jet

Diameter

Item Minimum

Value

Maximum

Value

Bucket length, L 2.28 do 3.3 do

Bucket width, B 2.8 do 4 do

Notch depth, S 0.44 do 0.625 do

Notch width, M 1.12 do 1.6 do

Bucket depth, E 0.8 do 1.2 do

Bucket height, A 1.75 do 2.5 do

The dimension of Pelton turbine bucket is shown in

figure 5.

Figure 5 Dimension of pelton turbine bucket

2.3 Results Data of Bucket

Table 3 shown the minimum value and maximum

value of the bucket dimensions.

TABLE 3 Result Data of Bucket

Item Minimum Value Maximum Value

Bucket

length, L

52.4mm 75.9mm

Bucket

width, B

64mm 92mm

Notch depth,

S

10.12mm 14.375mm

Notch width,

M

25.76mm 36.8mm

Bucket

depth, E

18.4mm 27.6mm

Bucket

height, A

40.15mm 57.5m



3. Production Process of Pelton Turbine Runner

Production process of pelton turbine runner are

design and drawing, pattern making, casting and

assembling. Mechanical drawing can be drawn by using

design data. The next step is patter making for casting.

figure 6 show that pattern is made by CNC milling

machine.

Figure 6. Pattern making

After pattern making, the next step is mould making.

The proper mould for this pattern is clay mould as

shown in figure 7.

Figure 7 Mould making

And then, casting for pelton turbine bucket is chosen

by available local aluminum because the weight of the

bucket is 0.2kg. Figure 8 shows the casting process and

Fig 9 shows pelton turbine bucket.

Figure 8 Casting Figure 9 bucket

Final step is assembly of runner as shown in figure

10. To construct the turbine runner, these materials are

required. They are resin, colouring, pigment, talc

powder, fiber mat and hardener. Figure 11 shows

assembly of pelton turbine.

Figure 10 Runner

Figure 11 Assembly of Pelton Turbine

3.1 Performance Testing of Pelton Turbine

Before being tested the turbine and generator

assembly must be set firstly. The turbine is started at no

load conditions. The speed of turbine is measured by

using tachometer. The speed of the turbine is 500 rpm.

And then the load is gradually increased and result are

recorded. The generator output can easily be measured

by using various load. The turbine is tested at five

different loads and head and flow rate are constant. The

water passed through nozzle and is guided to the runner.

For that turbine permanent magnet type 4 pole

generator is used with belt drive which speed increases

three times. During the test, the voltage of the turbine

and the speed was found to decrease depending on the

increasing load. It is the reason that there is no gate

valve to control the water to be increased for increasing

the load. The higher the load, the more flow rate is

required at constant head to get the higher power. But at

that testing, the flow rate is constant at all varying load.

Therefore, the flow rate must be increased to get the

required power.

3.2 Result Data

TABLE 4. Test Result Table No Head

(m)

Speed

(rpm)

Volt

(V)

Load

(watt)

Quan

tity

Power

(watt)

1 10 600 270 100 1 100

2 10 550 230 100 2 300

3 10 500 200 100 2 500

4 10 450 150 100 1 600

5 10 450 130 100 1 700

Total Power 700 watt



Figure 12. Test Result

4. CONCLUSIONS

This turbine can be used for household in remote areas

to produce 1 kW power for 1 household are easily and

inexpensively. This turbine can be used to

demonstration for hydropower training at DRI. The

required head is 10m to generate 1kW output power.

The flow rate of this turbine is 0.02 m3/sec and the pitch

circle diameter is 0.248m. Number of poles of generator

is 4 pole and the speed of turbine is 500rpm. The

diameter of jet is 23mm and jet ratio is 10. The number

of blade is 20. The performance testing is made at DRI

which is located in Yankin Township. In this research,

the design power is 1kW. The test result is 700 W. At

that testing, the flow rate is constant at all varying load.

Therefore, the flow rate must be increased to get the

required power. So ,this test results are the correct

design of the runner. The turbine can be manufactured

by any simple workshop. It can also be quickly and

easily removed temporarily during flooding of other

adverse condition. The design of Pelton turbine for

others hydro power plant can be calculated by using the

similarity law.

The micro and pico hydro power plant are easily

established at low cost. So, the micro and pico hydro

power generation is the best method for rural

electrification.

ACKNOWLEDGEMENT

The author is grateful to Dr. Kyaw Aung, Professor

and Head, Technological University(Mandalay), for

giving the permission to summit the paper.

The author grateful to her supervisor Daw Cho Cho

Khaing, Associate Professor , Department of

Mechanical Engineering, Technological University

(Mandalay) for her encouragement, patient guidance,

invaluable supervision, kindly permission and

suggestions throughout the research and continuous

guideline.

The author wishes to express gratitude to Dr. Aung

Zaw Lynn, Professor, Department of Mechanical

Engineering Technological University (Mandalay) for

his guidance and suggestion for this paper. The author

wishes to extend her gratitude to her parents, husband

and lovely daughter.

REFERENCES

[1] A Harvey and Brown. 1992. “Micro-hydro Design

Manual” ITDG Publishing.

[2] Sharma, R.K. 2003. A Text Book of Water Power

Engineering. S. Chand and ompany

Ltd.

[3] Khurmi. R.S. 1979. A Text Book of Hydraulic,

Fluid Mechanics and Hydraulic Machines.S.

Chand and Company Ltd. Ram Nagar, New Delhi.

[4] Celso Penche and Inge niero de Minas. 1998.

Layman’s Guide Book. How to Develop a Small

Hydro Site, 2rd. ed. France.

[5] Miroslav Nechleba, Dr Techn., M. 1957.

Hydraulic Turbines. Their Design and

Equipment.

[6] Khin Maung Aye, U. 2001. Fluid

Mechanics. Mechanical Engineering

Department, YTU.

[7] Franke, G.F., D.R. Webb, R.K Fisher, D.Mathur,

P.N Hopping, P.A. March.

1997. “Development of environmentally

Advanced Hydropower Turbine System

Concepts” Voith Hydro, Inc. Report No.: 2677-

0141.

[8] Egual U.Y. 1963. Pelton Turbine, Theory and

Research and Calculation.

[9] Bansal. R.K. 1983. Fluid Mechanics and

Hydraulic Machine. Laxmi Publication,

London, UL.



Design of Hydraulic Ram Pump and Performance Analysis on Different Parameters

Cho Cho Khaing(1), Aye Aye Khaing(2), Kyaw Aung(3) (1)Technological University (Mandalay), Myanmar

(2) Technological University (Mandalay), Myanmar (3) Technological University (Mandalay), Myanmar


ABSTRACT: Hydraulic ram pump plays the important

role in Myanmar since Myanmar is rich in water

resources. This paper forecasts the performance

characteristics of the designed pump by varying the

supply pipe length, waste valve travel, delivery head,

waste valve diameter and delivery valve diameter. The

designed pump can develop a delivery head of 3.1 m

and deliver 7.728 litre/min with 127 beat/min. Based on

the drive pipe length 6 m and the supply head 1.4 m, the

Rankine’s efficiency is nearly 55% and the

D’Aubuisson’s efficiency is about 64%. The

designed pump can also fulfill for water supply to

remote area.

KEYWORDS: head losses, maximum velocity,

time/beat, flow rate

1. INTRODUCTION The hydraulic ram (or) hydrum is a type of pump in

which the energy of large quantity of water falling through small height is utilized to lift a small quantity of this water to greater height. No external power is required to operate this pump. It can be used for water supply to country side and remote area where a source having large quantity of water at some height is available.[1]

1.1 Components of hydrum

The main parts of a ram pump are hydrum

body, waste valve, delivery valve, air chamber and

relief valve. Ram pumps have a cyclic pumping action

that produces their characteristics beat during operation.

Fig 1. Hydraulic ram pump

1.2 Working principle

The working of a hydraulic ram is based on the

principle of water hammer or inertia pressure developed

in the supply pipe. Initially as the water flows down the

supply pipe into the valve chamber, the waste valve

being open, the water flows through it to the waste

water channel. As the rate of discharge past the waste

valve increases, the flow of water in the supply pipe

accelerates. [1]

Due to the accelerating flow in supply pipe and the

static column of water in the supply tank, the pressure

in the valve chamber rapidly increases and acts the

lower face of the waste valve. Then, the waste valve

almost instantaneously closes due to the force is greater

than the weight of the waste valve.

The instantaneous closing of waste valve brings the

water in the supply pipe suddenly to rest and causing

the pressure rise in the valve chamber due to inertia.

Then, the delivery valve is forced open and the water

flows through the delivery valve into the air vessel and

delivery pipe.

Thus some of the flowing water is directly supplied

to the delivery tank and some of it is stored in the air

vessel. An air vessel assists in providing a continuous

delivery of water at a uniform rate.

The flow of water through the delivery valve

continues until the momentum of the water in the

valve chamber is destroyed, the delivery valve then

closes and the waste valve opens, thus again causing the

water to flow from the supply tank to the waste water

channel. This constitutes one cycle of operating or one

beat of the hydraulic ram. [1]

2. BASIC THEORY OF HYDRAULIC RAM PUMP 2.1 Conservation of energy in fluid mechanics

The Bernoulli’s equation can be considered to be a statement of the conservation of energy appropriate for flowing fluids. The energy equation (or) the extended Bernoulli’s equation taking into account gains and losses of head is given by

lpump hzg

VhHz

g

Vh +++=+++ 2

2

221

2

11

22

(1)

Where h = pressure head (m)

V = average velocity of fluid (m/s)

z = elevation head (m)

Hpump = head added by pump (m)

hl = head loss (m)

g = acceleration due to gravity (m/s2)

https://www.nuclear-power.net/nuclear-engineering/fluid-dynamics/bernoullis-equation-bernoullis-principle/



2.2 Head losses

The Darcy-Weisbach Equation of head loss in a pipe

is

gD

fLVhl

2

2

= (2)

Head loss in pipe fitting may be expressed as

g

Vkhl

2

2

= (3)

in which the coefficient k depends on the size, shape

and type of fitting. The value of k and f can be found

from standard reference handbooks.

The total head loss coefficient for driving section and

delivery section may be assumed by

im kkD

fL

A

A+++

=

2

2

1lh (4)

Where km is minor total losses coefficient and ki is

impulse valve loss coefficient.

In this study, the supply pipe diameter, waste valve

diameter and delivery valve diameter are selected the

same. Friction factor f is assumed 0.024 for smooth

pipe. From the Moody diagram, Reynold’s number is

estimated and the maximum velocity for turbulent flow

can be determined by

s

mD

VRe

= (5)

Where υ is the kinematic viscosity of water.

2.3 One dimensional unsteady flow equations Since the head contributed to water acceleration in the

driven pipe, the effective head available to accelerate the liquid in the pipe during the drive cycle is given by

dt

dV

g

L

g

Vh s

ls =−2

H2

(6)

Similarly, the one dimensional unsteady flow during the delivery section may be expressed

dt

dV

g

L

g

Vh s

ld =+2

h2

(7)

Where H = supply head

h = delivery head

hls = total head loss coefficient in drive section

hld = total head loss coefficient in delivery section

Ls = supply length of pipe

By integrating (6), the time required to close the

waste valve (or) the time during which the velocity in

the supply pipe builds up from zero to Vm, t1 and the

volume of water in drive section are determined by

2gH

Vhtanh

gHh

2Lt

2

mls1

ls

2

s

1

−= (8)

( )

−

=

2gH

Vh1

1ln

h

ALVol

2

mlsls

ss

s (9)

Similarly, the volume of pumping water in delivery

section and the time for which the waste valve remains

closed (or) the delivery valve remains open, t2 are

determined by integrating (7).

+=

gh

Vh mld

21ln

h

AL(Vol)

2

ld

ssp

(10)

2gh

Vhtan

ghh

2Lt

2

mld1

ld

2

w

2

−= (11)

So, the pumping flow rate can be determined by

( )

( )21 tt

VolQ

p

p+

= (12)

2.4 Efficiency of hydraulic ram

The Rankine’s efficiency is given by

HQ

hQη

s

p

r = (13)

and the D’Aubuisson’s efficiency can be determined

by

( )

( )pQ+

+=

s

p

dQH

QhHη (14)

3. DESIGN PARAMETERS OF HYDRUM

Design specifications of hydrum are as shown

in Table 1.

Table 1. Design Specifications of Hydrum

No Design Parameters Symbol Value unit

1 Supply head H 1.4 m

2 Supply pipe diameter Ds 0.0381 m

3 Waste valve diameter Dw 0.0381 m

4 Delivery valve diameter

Ddv 0.0381 m

5 Delivery pipe diameter Dd 0.0190 m

6 Waste valve travel b 0.0120 m

7 Length of waste valve Lw 0.0762 m

8 Supply pipe length Ls 6 m

9 Delivery head h 3.1 m

Table 2 shows the designed results of Hydraulic Ram

Pump.



Table 2. Designed Result of Hydrum

No Parameter Symbol Value Unit

1 Maximum velocity Vm 0.782 m/s

2 Time for which the waste valve remains open during each beat

t1 0.4705 s

3 Time for which the waste valve remains close during each beat

t2 0.0018 s

4 Total time for each beat

t 0.4723 s

5 Supply flow rate Qs 0.5169 Litre/s

6 Pumping flow rate Qp 0.1288 Litre/s

7 Rankine’s Efficiency

ηr 55.17 %

8 D’Aubuisson’s efficiency

ηd 64.12 %

The three dimensional hydraulic ram pump

with the designed data is created by Solid works Software.

Fig 2. Detail drawing of designed Hydrum

4. PERFORMANCE ANALYSIS ON DIFFERENT

PARAMETERS

Fig. (3) shows the variation of efficiency, time for

one beat and pumping flow rate by changing the supply

pipe.

According to Fig. (3), the efficiency and pumping

flow rate tend to decrease and the time required for one

beat increases if the supply length is longer.

Fig 3. Efficiency, time for one beat and pumping flow

rate on various supply pipe

Fig. (4) presents the efficiency, time for one beat and

pumping flow rate at different waste valve travel.

Fig 4. Efficiency, time for one beat and pumping

flow rate on various waste valve travel

If the waste valve travel is higher, the efficiency and

pumping flow rate are more and time for one beat is

shorter.

Fig. (5) describes the efficiency, time for one beat

and pumping flow rate on various delivery head.


rate on various delivery head

Fig.(5) can be seen that the delivery head increases,

the efficiency also increases although the time for one

beat is approximately constant.

Change of the waste valve diameter and delivery

valve diameter doesn’t affect significantly on the

efficiency, time for one beat and pumping floe rate as

shown in Fig.(6) and (7).


rate on various waste valve diameter 5 6 7 8 9 10 11 12 13

0

0.2

0.4

0.6

0.8

1

1.2

1.4Efficiency, time for one beat and pumping flow rate on various supply pipe length

supply pipe length(m)

eff

icie

ncy,

tim

e /

beat,

pum

pin

g f

low

rate

efficiency

time for one beat,s

pumping flow rate,L/s

5 6 7 8 9 10 11 12 13

x 10-3

0.1

0.15

0.2

0.25

0.3

0.35

0.4

0.45

0.5

0.55

0.6

Efficiency,time for one beat and pumping flow rate on various waste valve travel

waste valve travel(m)

eff

icie

ncy ,

tim

e/b

eat,

pum

pin

g f

low

rate

efficiency

time for one beat,s


2 2.5 3 3.5 4 4.5 5 5.5 6 6.50

0.1

0.2

0.3

0.4

0.5

0.6

0.7Efficiency,time for one beat and pumping flow rate on various delivery head

delivery head (m)

eff

icie

ncy,t

ime/b

eat,

pum

pin

g f

low

rate

Efficiency

time for one beat,s


0.025 0.03 0.035 0.04 0.045 0.05 0.0550.1

0.15

0.2

0.25

0.3

0.35

0.4

0.45

0.5

0.55

0.6

Efficiency,time for one beat and pumping flow rate on various waste valve diameter

diameter of waste valve (m)

eff

icie

ncy,t

ime/b

eat,

pum

pin

g f

low

rate

efficiency

time for one beat,s





rate on various delivery valve diameter

5. CONCLUSIONS

The designed hydraulic ram pump is aimed to use in

domestic. It can develop 3.1 m head and deliver 7.728

litre/min without using electricity power (or) fuel. The

designed efficiency is analyzed based on Rankine’s

efficiency equation. The designed hydraulic ram pump

can be used in village which locate on hillside and

remote areas which doesn’t have the electricity grid.

Since the mechanism of a hydrum is simple and doesn’t

require maintenance cost, it is crucial for small scale

water supply schemes in developing countries.

ACKNOWLEDGEMENT

The author would like to express her special thanks

to Dr. Aye Aye Khaing, Professor, Department of


(Mandalay) for her encouragement, suggestion and Dr

Kyaw Aung, Professor and Head, Department of


(Mandalay) for his valuable guidance of this paper.

REFERENCES

[21] Modi P.N., Hydraulics and Fluid Mechanics

[22] Victor L. Streeter and E. Benjamin, Fluid

Mechanic 7th Ed,(1951)

[23] David, J.P. and Edward, H.W., Schaum’s Outline

of Theory and Problems of Fluid Mechanics and

Hydraulics, SI (Metric)Edition , McGraw Hill

Book Company,Singapore,1985.

[24] Watt, S.B.,. Manual on a Hydraulic Ram for

Pumping Water, Intermediate Technology

Publication Ltd., London, 1982

[25] Design and Fabrication of A Hydraulic Ram

Pump, Nari Suraj, “International Journal of

Engineering Research and Technology” Special

Issue, Conference Proceedings,2018.

0.025 0.03 0.035 0.04 0.045 0.05 0.0550.1

0.15

0.2

0.25

0.3

0.35

0.4

0.45

0.5

0.55

0.6

Efficiency,time for one beat and pumping flow rate on various delivery valve diameter

delivery valve diameter (m)

eff

icie

ncy,t

ime/b

eat,

pum

pin

g f

low

rate

efficiency

time for one beat,s




Design and Heat Dissipation of Internal Combustion Diesel Engine Exhaust Valve

Aung Zaw Lynn(1), Kyaw Linn Soe(2)

(1)Technological University (Mandalay), Myanmar (2)Technological University (Mandalay), Myanmar


ABSTRACT: This paper mainly describes the design

and heat dissipation of internal combustion engine

exhaust valve. As the airflow passes various

components and stages of the intake system, inlet valve

does not undergo excessive heat from working engine.

But the exhaust valves are hot enough to cause

immature failure. Therefore this paper intends to study

the design and temperature distribution of the exhaust

valve by using related theoretical concepts. To consider

the design, the specifications of 72 hp four cylinder four

strokes diesel engine are used. The design of the

exhaust valve is calculated and valves are used to

control gas flow to and from cylinder of automotive

internal combustion engines. The most commonly used

valve is the poppet valve. The valve itself consists of a

disc shaped head having a stem extending from its

center at one side. The valve-seat angle of exhaust valve

kept 45 degrees in this design. Although both inlet and

exhaust valves receive heat from combustion, the inlet

valve is cooled by incoming air, whereas the exhaust

valve experiences a rapid rise of the temperature in the

valve head, seat insert, and under head area from hot

exhaust gases. To make the design of valve, the

dimensions of bore and stroke of 100 mm and 115 mm

from 72 hp four cylinder diesel engine was adopted.

The material of exhaust valve is AISI 4340 low alloy

steel and this paper also describes the process of

temperature distribution of exhaust valve at peak

combustion temperature 1616 K for three different

materials.

KEY WORDS: exhaust valve, temperature

distribution, heat power, combustion gas, ic engine

1. INTRODUCTION

The purpose of the internal combustion engine is

the production of mechanical power from the chemical

energy contained in the fuel. The fuel-air mixture or

fresh air is actual working fluids. The work transfers

from working fluids which provide the desired power

output by moving mechanisms. About thirty five

percent of the heat generated is lost into the

surroundings of combustion space, remainder being

dissipated through exhaust and radiation from the

engine. Additional heat is produced by friction between

the moving parts.

Heat moves from the area of high temperature to

the area of low temperature. The heat from the engine is

moved to the atmosphere by cooling with water and air.

If it is not removed from engine, engine components

will damage cause of excessive temperature. The heat

from combustion process affected mainly to the piston

head, cylinder wall, exhaust valve and exhaust pipe.

The exhaust valve may become hot enough to cause

preignition or may fail structurally. Besides, preignition

would increase the cylinder head temperature further

until engine failure or complete loss of power results.

The charge to burn will contact the wall of the

combustion chamber and upper cylinder head area. If

this charge temperature of these parts will reduce the

delay period, it may cause knocking. For the side of

inlet valve, the fresh air which is at the atmosphere

temperature and pressure is entered through the inlet

valve. Therefore, the exhaust valve suffers from damage

more than the inlet valve because of excessive heat.

The valve face angle (with the plane of the valve

head) is generally kept 45° or 30°. A smaller face angle

provides greater valve opening for a given lift, but poor

sealing because of the reduced seating pressure for a

given valve spring load. Due to this reason, in some

engines, the inlet valve face angle may be kept 30° or

45° whereas the exhaust valve face angle is only 45°, as

this increases its heat dissipation. The temperature

distributions are affected by heat conduction in valve

seat and heat transferred by convection and radiation

from combustion gases.

Fig 1. Configuration of Inlet and Exhaust Valves [13]

2. DESIGN CONSIDERATION OF EXHAUST

VALVE

2.1 Description of Main Parameters

B. The design parameters of exhaust valve are

as follow;

• Thickness of the valve head

• Diameter of valve face

• Diameter of valve stem

• Maximum valve lift

The exhaust valve design of internal combustion

engine in this paper is used as the flat-headed type

Spring

Piston

Head

Cylinder

Wall

Exhaust Valve

Inlet

Valve



poppet valve. To design the valve, the port diameter is

known data which is got from the engine specifications.

The valve stem diameter (ds), the valve face diameter

(df), the valve disc thickness (t), and the maximum

valve lift (hlift) can be considered from the calculation of

poppet valve design. The sketch of poppet type valve is

shown in Fig 2.

Fig 2. Conical Poppet Valve in the Port [9]

The thickness of the valve head (t) can be

determined empirically from the following relation.

σ

Pdkt 2

p1= (1)

where,

t - thickness of valve disc in mm

k1 - constant for metal

dp - diameter of exhaust port in mm

P2 - maximum gas pressure in N/mm2

σ - permissible bending stress in N/mm2,

(100 N/mm2 ~ 120 N/mm2 for alloy

steel) [9]

The constant for metal (k1) varies for alloy steel

and for cast iron which are 0.54 and 0.42 respectively.

The maximum gas pressure is the pressure which may

be taken as nine to ten times of the mean effective

pressure (Pi). The permissible bending stress for the

material of valve may be taken as 100MPa which is for

alloy steel. The valve face diameter can be determined

from the following equation.[9]

α)]sin(902[tdd pf −+= (2)

where,

df - diameter of valve face in mm

α - valve seat angle in 45 degrees

The diameter of valve stem can be determined from

the following equation.

6.35mm8

dd

p

s += (3)

where,

ds - diameter of valve stem in mm

The maximum valve lift can be obtained by using

the following equation.

hlift =0.25 dp

cos∝ (4)

where,

h - maximum valve lift in mm

dp - diameter of exhaust port in mm

α - valve seat angle in 45 degrees

2.2 Specifications of 72 HP Four Cylinder Diesel

Engine

The following table shows specifications which are

used to design the valve.

Table 1. Specification of 72 hp Four Cylinder, Four

Stroke DI Diesel Engine

No

.

Parameter Value Unit

1 Cylinder bore 100 mm

2 Piston stroke 115 mm

3 Piston displacement 0.903×10

-3

m3

4 Compression ratio 20:1 -

5 Power output 72(53.71

2)

hp(kW

) 6 Rated speed 4000 rpm

7 Atmospheric

temperature

298 K

8 Atmospheric pressure 0.1013 MPa

9 Weight of the valve 0.12 kg

10 Max: suction pressure 0.025 MPa

11 Cylinder pressure

when exhaust valve

open

0.3004 MPa

12 Exhaust port diameter,

dp

31.82 mm

13 Valve length 130 mm

The exhaust valve from 72 hp four cylinder diesel

engine is made up of low alloy steel. This low alloy

steel is ranged at the standard of AISI4340. From this

internal combustion engine, the intake process

temperature and pressure are at the standard condition.

The starting pressure of compression stroke is 0.8~0.9

of atmospheric pressure. From the adiabatic process, 1-

2, the temperature is 987.71 K and the pressure is 5.3

MPa. The combustion peak temperature is 1616 K and

the combustion pressure is 5.3 MPa at constant pressure

process. The exhaust temperature is 751 K which is

calculated from the p-V and T-s diagrams. 2.3 Estimation of Combustion Gas Temperature

The volume of the cylinder can be determined as a

position of crank from the compression ratio, stroke,

bore and connecting rod length.

df



Fig 3. Basic Geometry of Reciprocating IC Engine

The difference between the maximum and

minimum volume is known as the displacement

volume, Vd, displacement volume can also be

represented as a function of the bore and stroke.

The displacement volume is

LD4

πV 2

d = (5)

= 9.032×10-4 m3

The following parameters define the basic

geometry of a reciprocating engine. They are used to

obtain the gas temperature.

Compression ratio, rc = 20

c

cdc

V

VVr

+= (6)

Vc = 4.7537×10-5 m3

where,

Vd - the displacement volume

Vc - the clearance volume.

Crank radius, r can be calculated from the

following Equation (7).

2

length Stroker = (7)

= 0.0575 m

Equivalent length of the connecting rod, Le, can be

obtained from the following equation (8).

λ

rLe = (8)

= 0.21296 m

where,

λ - the constant value

The range of λ is 0.25 to 0.5. Use the value of λ is 0.27.

Distance between the crank axis and piston axis, s

s = rcosθ + (Le2 + r2sin2θ)1/2 (9)

= 0.1747 m

where,

θ - the crank angle is 125 degrees measure

from BDC during compression process.

The cylinder volume Vcyl will be computed to

consider the average temperature at the crank angle 125

degrees.

s)r(LD4

πVV e

2

ccyl −++= (10)

= 7.996×10-4 m3

The temperature of the gas, Tg will be average

temperature. The temperature will be found from the

following ideal relationship.

M = 176 kg/kmol

ρa = 1.2 kg/m3

R = 8.3413 kJ/kmol K

Tg =PcylVcylM

ρaR[

rc−1

rcVd] (11)

= 1500 K

where,

M - Molar mass of fuel, kg/kmol

ρa - density of air, kg/m3

R - universal gas constant, kJ/kmol K

Pcyl - Pressure at crank angle 125 degrees

Tg - Temperature at crank angle 125 degrees

2.4 Determination of Heat Transfer Coefficient for

Air at 1500 K

The heat transfer coefficient is determined by using

the thermophysical properties at 1500 K. This value is

practically applied to find the thermal resistance of the

exhaust valve. The thermophysical properties of exhaust

gas are shown in the following TABLE 2.

Table 2. Thermophysical Properties of air at 1500 K[14]

No Parameter Symbol Value Unit

1 Density ρ 0.2322 kg/m3

2 Thermal

conductivity к 100 x 10-3 W/mK

3 Specific heat

capacity cp 1.230 kJ/ kgK

4 Viscosity μ 557 × 10–7 Pa-s

5 Prandtls No: Pr 0.685 -

6 Kinematic

viscosity N 240 × 10–6 m2/s

The heat transfer model of the valve stem is based

on the flow upon the cylindrical rod. The value of

Reynolds number is shown as Equation (12).

Re =ρv(dp −ds)

μ = 8.06 x 103 < 5 x 105 (12)

where,

ρ - Density of exhaust gas, kg/m3

µ - Dynamic viscosity of exhaust gas, Ns/m2

v - Velocity of exhaust gas, 90 m/s ~ 100 m/s

for high speed engine [9] This condition is laminar flow. Therefore, the

Nusselt number defines as Equation (13).[12]

NuD = 0.664 ReD1/2

Pr1/3 (13)

The heat transfer coefficient for exhaust gas can be

obtained from Equation (14).



l

kNuh = (14)

h = 40.43 W/m2K

where,

h - Convection heat transfer coefficient,

W/m2K

2.5 Distribution of Heat Transfer Rate for Exhaust

Valve

The distribution of heat transfer rate on exhaust

valve considers by heat conduction and radial effects as

the following Fig 4. The heat flow transfers from valve

face to along the valve stem. The temperature of the

valve stem edge will be predicted by solidworks

simulation. It is considered when exhaust valve is

opening period after combustion.

Fig 4. Sketch of Thermal Resistance

Total Thermal Resistance can be computed by

equation (15).

∑RT = 1

hi Ai+

L1

kA1+ [

11

Ahi

+ 11

Ahi

+1

L2kA2

]

−1

+1

h0 A0

(15)

where,

k - thermal conductivity, 44.6 W/m-K for

alloy steel.

Area of valve head,

A1 = (df+dp

2) t

(16)

Area of valve stem,

A2 =π

4ds

2 (17)

Area of gas flow between exhaust port and valve

stem,

A = π

4(dp

2 − ds2) (18)

Area of valve tip, Ai

Ai = π

4di

2 (19)

Area of valve stem tip, Ao

Ao = π

4do

2 (20)

Heat power for exhaust valve can be calculated by

equation (20).

q =Ti−To

∑ RT (21)

where,

q - Heat power for exhaust valve, W

3. DESCRIPTION OF DESIGN PARAMETERS AND

THERMAL DISTRIBUTION

The following results from Table 3 shows the

comparison of existing and theoretically based data of

the exhaust valve. Material of the exhaust valve is the

same for the existing valve and designed valve.

Table 3. Design Parameter of Exhaust Valve

No Specifications Symbol

Design

Value

(mm)

1 Diameter of valve face df 36.192

2 Thickness of the valve

disc t 3.0912

3 Diameter of valve stem ds 10.328

4 Maximum valve lift hlift 11.25

Table 4. Result Table of Heat Rate for Exhaust Valve

No Parameters Values Units

1 Total Thermal

Resistance, ƩRT’’ 155.505 K/W

2 Area of valve head, A1 1.05x 10-4 m2

3 Area of valve stem, A2 8.38 x 10-5 m2

4 Area of valve tip, Ai 1.03 x 10-3 m2

5 Area of valve stem tip,

Ao 8.38 x10-5 m2

6 Area of gas flow , A 7.11 x 10-4 m2

7 Heat power 8.5 W

4. ANALYSIS OF TEMPERATURE DISTRIBU-

TION FOR EXHAUST VALVE

The following figures show the tempertaure

distribution of exhaust valve at peak combustion

temperature 1616 K. In fact, the temperature on the tip

of the valve stem is higher than the ambient

temperature. Therefore, the temperature on the valve

stem tip can be seen 357.7 K in AISI 4340 low alloy

steel, 400.9 K in alloy steel and 342.4 K in AISI 4130

steel above ambient temperature according to

simulation results while running the engine. This results

are simulated in same parameters of exhaust valve with

different materials. The heat dissipation of these

materials are not very different but alloy steel is more

than others but less density. Therefore it’s size will be

larger than this designed parameters. Both materials of

AISI 4340 and AISI 4130 are same density but AISI

4340 is more heat power and thermal conductivity than

AISI 4130.



Fig 5. Temperature Distribution of Exhaust Valve

at Gas Temperature 1616 K (AISI 4340 Low Alloy

Steel)

Fig 6. Temperature Distribution of Exhaust Valve at

Gas Temperature 1616 K (Alloy Steel


Gas Temperature 1616 K (AISI 4130 Steel)


Gas Temperature 1500 K (AISI 4340 Low Alloy Steel)

According to the above review and simulation

results, the material AISI 4340 is more suitable than

AISI 4130 for this designed parameters.

The following table is described the results of

exhaust valve for AISI 4340 in peak combustion

temperature of 1616 K and average combustion

temperature of 1500 K.

Table 5. Result Table of Peak and Average Combustion

Temperature of Exhaust Valve for AISI 4340

Sr.No Name of Items

Peak Tempature

Average Temperatur

e 1 Material AISI4340 AISI4340

2 Melting

point 1705 K 1705 K

3

Combustion temperature 1616 K 1500 K

4 Valve stem tip temperature

298 K 298 K

357.7 K 331.5 K

5 Heat Power (Theorectical)

8.5 W 7.73 W

6 Heat Power (Simulation) 8.1 W 7.54 W

7 Deviation 4.7 % 2.4 %

5.CONCLUSIONS

The exhaust valve is one of the most critical parts

for the combustion system. It effects not only engine

performance but also volumetric efficiency. The

operating temperature is the most significant factor in

the performance of the exhaust valve. The physical

properties of the valve stem effects the temperature

distribution rate. Exhaust valve stem generally fails by

overheating. It is not considered as the coating effect

which can reduce the temperature from the valve. This

paper presents the heat dissipation process of the

exhaust valve used in the internal combustion engine. In

this paper, the material of exhaust valve is low alloy

steel (AISI 4340) and the length of valve is 130 mm.

The heat power generated from the exhaust valve is 8.5



W. The parameters of the designed valve is covered to

prevent the failure resulted from thermal stress because

the melting point temperature of alloy steel is 1432 ˚C

(1705 K). The exhaust valve stem can be performed to

reduce having valve stem tip temperature of 357.7 K

from peak combustion temperature of 1616 K.

ACKNOWLEDGMENT

The author would like to express his special thanks

to U Kyaw Lin Soe, Lecturer from Technological

University (Mandalay), Department of Mechanical

Engineering for his valuable advice and discussion.

Also author appreciate to U Kyaw Minn Khant,

Automobile Engineer from United Auto Car Service in

Mandalay for supporting some technical data.

REFERENCES [1] Danny Bender, Valve Guide Wear- What to Look

For, www.dannyengineportal.com, January 27,

(2016).

[2] Frank P. Incropera, David P. Dewitt,

Fundamentals of Heat and Mass Transfer, Seven

Edition.

[3] Prabhu. B. D., Valve and Valve Mechanisms, Unit

8 DP, (2015).

[4] Soni K.; Bhatt S. M.; Dayatar R., and Vyas K.,

Optimizing IC Engine Exhaust Valve Design

Using Finite Element Analysis, (2015).

[5] Taffmeisters, Exhaust Oversize Valve Seat for

Seat for Repairing Damaged Cylinder Head,

www.taffmeisters.co.uk, (2012).

[6] Kumar G.U. and Mamilla V. R., Failure Analysis

of Internal Combustion Engine Valves by Using

ANSYS, (2014).

[7] Pandey A. and Mandloi R. K., Effects of High

Temperature on the Microstructure of Automotive

Engine Valves, (2014).

[8] Sanoj. T. and Balamurugan S., Thermo

Mechanical Analysis on Exhaust Valve, (2014).

[9] A Text Book of Machine Design, “Internal

Combustion Engine Parts” Chapter 32.

[10] Hornik. A.; Jędrusik D., and K. Wilk, Unsteady

State Heat Flow in the Exhaust Valve in

Turbocharged Diesel Engine Covered by the

Layer of the Carbon Deposit, (2012).

[11] Snehal S.Gawale, Dr.S.N.Shelke, Diesel Engine

Exhaust Valve Design and Optimization,

www.iosrjournals.org.

[12] Rayner Joel, Basic Enfineering Thermodynamic,

Fifth Edition.

[13] www.Single-cylinder Diesel Engine Intake and

Exhaust Valve Enginebasiscs.com

[14] Heat and Mass Transfer Data Book,

Thermophysical Properties of Matter.

[15] Reid, R.C., J.M. Prausnitz and B.E. Poling, 1987.

“The Properties of Gases and Liquids”, McGraw-

Hill, New York, 4th edition.

http://www.dannyengineportal.com/

http://www.taffmeisters.co.uk/

http://www.iosrjournals.org/

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