ABSTRACT

by

John TauloDeputy Director (Research and Development)

Malawi Industrial Research and Technology Development CentreP.O.Box 357, Blantyre

E-mail: mirtdc@malawi.netWebsite: mirtdcmalawi.com

April 2005

This paper contains a technical report on a project to be carried out on

design and development of screw press. The project focuses on qualitative

analysis of existing screw presses and modifying them for production in

local workshops. The analysis will be based on both technical and financial

appraisals. Limitations therefore will include complexity of the turbine,

ease of fabrication and production costs among others.

The first screw press model was consuming relatively more energy and

needed high power source for its operation. The other problems were

unsmooth operation, oiling up, heavy foots, high oil content in cake fibre,

and comparatively high cost of maintenance. The model was found

inefficient for hard oil seeds such as sunflower, soybean and cotton seed.

The above enlisted shortfalls necessitated further research and

development.

In order to overcome the above stated problems, the technical

development work was geared towards further increasing the oil recovery

rate at a competitively low energy consumption, increasing the operation

efficiency and life of the expeller. Also achieving smooth operation both

on soft as well as hard oil seeds on expeller.

Depending on the defined limitations, the screw press is to process 100 to

120 kg oilseed per hour. This compares favourably well with TinyTech,

Lahore and Sundhara expellers . This report includes briefs on the designs

considered, engineering drawings and calculations of the final design

selected.

DESIGN BRIEF

The shortage of both edible as well as non-edible oils and fats for

industrial use has been continuing in Malawi for the past three decades.

As the local production of the vegetable oils is meagre compared to their

demand, so the shortage is met through imports from various sources.

Undoubtedly, the agriculture sector should endeavour to minimise our

dependance on imports but scientific and technological improvements

also have a major role to play.

One of the primary objectives of the present government is to provide

more job opportunities in the rural areas and to ameliorate the condition

of the rural population by raising their per capita income. The research

and development work on small scale oil expelling system is a significant

step taken by the Centre in this direction. This technology will not only

bring about uplift of the socioeconomic condition of the rural population,

but also reduce the gap between supply and demand of edible oils in

Malawi.

TABLE OF CONTENTS

Abstract

Design Brief

Introduction

Scope

Objective

Methodology

CHAPTER 1: INTRODUCTION

1.1 Screw Pressing

Continuous pressing by means of expellers (also known as screw press) is

a widely applied process for the extraction of oil from oil seeds and nuts.

It replaces the historical method for the batch wise extraction of oil by

mechanical or hydraulic pressing. The expeller consists of a screw (or

worm), rotating inside a cylindrical cage (barrel). The material to be

pressed is fed between the screw and barrel and propelled by the rotating

screw in a direction parallel to the axis. The configuration of the screw

and its shaft is such that the material is progressively compressed as it

moves on, towards the discharge end of the cylinder. The compression

effect can be achieved, for example by decreasing the clearance between

the screw shaft and the cage (progressive or step-wise increase of the

shaft diameter) or by reducing the length of the screw flight in the

direction of the axial movement. The gradually increasing pressure

releases the oil which flows out of the press through the slots provided on

the periphery of the barrel, while the press cake continues to move in the

direction of the shaft, towards a discharge gate installed at the other

extremity of the machine.

1.2 Statement of the Problem

The main problems of the old screw design were unsmooth operation,

oiling up, heavy oil in cake fairly high maintenance, heavy power

consumption and its unsuitability for hard oils seeds. Removal or

administering of these problems through research and development was

initiated with a view to: having the best screw configuration for a given oil

seed, consuming less power supply ,and upgrading its capabilities in

processing of hard oil seeds, (such as sun flower, soybean and cotton

seed).

1.3 Significance of the Project

One of the primary objectives of the present government is to provide

more job opportunities in the rural areas and to ameliorate the condition

of the rural population by raising their per capita income. The research

and development work on small scale oil expelling system is a significant

step taken by the Centre in this direction. This technology will not only

bring about uplift of the socioeconomic condition of the rural population,

but also reduce the gap between supply and demand of edible oils in

Malawi.

1.4 Aim of the Project

The aim of the project is to undertake the qualitative analysis of existing

screw presses, select , design and modify the selected expeller. The

selected expeller will then be fabricated for local manufacture. It aims at

using locally available resources to make the screw presses affordable to

the average Malawian.

1.5 Objectives

The main objective of the project will be to improve the village level oil

seed processing system. Specifically:

a) To continue technical developments in optimising the screw design

for screw press model (2)

b) To evaluate the performance of the newly developed screw press

during field trials

c) To conduct socioeconomic studies on the improved oil seed

processing technology.

CHAPTER 2: REVIEW OF LITERATURE

2.1 A brief History of Screw Presses

The seed crushing industry is one of the oldest industries in the world.

The Chinese were the first people to express oil seeds. As far as the year

3000BC, the Egyptians knew how to obtain oil using a press composed of

a sausage-shaped rush bag slug between vertical posts of a strong

wooden frame. In the 19th century, the ruins of Pompeii dated back to

79AD were excavated. A large pestle and mortar was found, a long pole

acted as a grinding pestle and hollowed trunk of a tree held the seeds. An

ass or an ox walked around the press, dragging the top end of the pole

and thus grinding the seeds in the hollowed tree trunk. Centuries later

other seeds such as linseed, rapeseed, cottonseed, groundnut, soy bean

and palm kernel which required greater pressure for oil expressing, were

available. This was done by a press employing a windlass and then by

using a water mill or windmill to apply the pressure.

The first mechanical press was successfully used back in 1906. The

manufacturers have come a long way since then with improved material

of construction, manufacturing methods, research and development and

have increased the efficiency of the screw press. As a result, various

types of improved expellers were developed to meet the requirement of

the processors.

The Malawi Industrial Research and Technology Development Centre

(MIRTDC) began to focus attention on village level oil seed processing in

the late nineties. The Centre adapted the ram press which produces 10

litres of oil per day from sunflower seeds. Subsequently, laboratory work

was undertaken to improve the efficiency of the machine.

The main modification introduced by MIRTDC was the use of perforated

cylinder instead of the slotted barrel. As a result, the capacity increased

to 12 litres per day from 10 litres per day as in the original press.

Similarly, a plate press (spindle press) was adapted which produces 20

litres per day.

Further work was continued on the development of the village expellers.

As a result of the above activities, a model was designed to serve as a

starting point from where a commercial prototype could be developed.

The model expeller did not prove successful in the processing of soft as

well as hard oil seeds. The expeller produced more foots than oil, an

average of 1.6 litres per 5 kg of seed was produced representing 26%

yield. Also power consumption increased from 5kW to 7.5kW. It was felt

to redesign the press to address the above problems. The village oil seed

technology as developed by MIRTDC could increase the production of oil

………………… percent from the already available conventional sources.

More over, this technology would be greatly helpful in the promotion of

non-conventional oil seed such as sunflower and soybean. Besides this, it

would provide self employment opportunities in rural areas of t he country

which would help improve the socioeconomic conditions of the villagers.

At the same time, the improved technology would greatly help in reducing

the gap between supply and demand in Malawi.

2.2 Operating principles

The screw press is designed to continuously remove oil from oil bearing

seeds. As such , it is fed continuously, and discharges oil and low oil

bearing solids continuously. This continuous separation of oil and solids is

effected by a pressure established by a screw turning within a confining

slotted cylinder or barrel. The screw pick up and force incoming oil

bearing material toward and through an adjustable opening at the

discharge end of the press.

This force creates a pressure on the material which causes the oil therein

to be released and pass from the expeller through the slots in the barrel.

Varying the size of the opening, or choke, for a given feed rate, within

limits, for a given choke setting, changes the amount of pressure on the

confined material and thus determines, to a great extent, the amount of

oil which will be removed or conversely, the residual oil content of the

discharge solids.

2.3 Vector Shaft Analysis Technique

The worm shaft is designed by adopting the Anderson Vector Shaft

Analysis Technique. The objective in the worm shaft design is to maintain

a steadily increasing pressure on a material as it moves from one worm to

another along the shaft. The pressure will compact the material and the

next worm will have to compensate for this compaction (loss of volume

due to higher density) just to maintain the pressure exerted by the

previous worm. Volume is also lost due to the removal of oil, which loss

also has to be compensated for. The idea is to subject the material to a

steadily increasing compaction, compensating for all losses in volume,

and do this as smoothly as possible from one worm to the next.

1) From the dimensions of all the worm parts and the rotational speed

of the shaft, the m3/min displacement of each worm along the shaft

is calculated;

2) From the production capacity of the material being pressed , the

kgs/min of material entering the shaft will be calculated ;and by

predicting the weight of oil and fines pressed out along the length of

the shaft, the kgs/min of material flowing across each worm shall

also be calculated;

3) Then density of the material as compaction progresses from

worm to worm shall be predicted, and computations on the m3/min

also made. Volume of material passing along the shaft. This will be

compared to the volumetric displacement of each worm and a

judgment will be made as to whether that worm will contribute to

compaction or allow a loss of compaction, and to what extent.

CHAPTER 3: METHODOLOGY

The methodology used in this project shall be as follows:

Brainstorming and literature review of oil extraction technologies. Screw

presses obtained in Malawi will be surveyed and performance data

compared. Among the expellers available, locally fabricated small size

expeller shall be selected for further improvement or adaptation.

Based on the qualitative data obtained, sketches will be made.

The design will then be evaluated using the evaluation matrix method.

The optimum screw press design will be selected using the set criterion.

Further analysis on this press type will then be carried out.

Theoretical design and performance parameters will then be developed.

Modifications on the design will be carried out and final design carefully

sketched. Detailed engineering drawings will then be produced and based

on this, the press will be fabricated.

CHAPTER 4:PRODUCT DESIGN SPECIFICATIONS

The final specifications cannot be concluded until

Designing, and manufacturing is done and tested;

Cost analysis is completed, and customer feed back after a trial of

10 MT seed has been crushed.

However, the basic requirements are as follows:

4.1 Functional Requirements Specifications

4.1.1 Frame

The frame should be from welded steel and provide maximum strength

and allow for simple installation of the press. The motor is to be mounted

on the press, either on the base plate behind the gear box or on the side

of the gear box.

4.1.2 Gear Box

The gear box should be of the worm box type, for high efficiency, it should

be fitted, as standard, with large bearings that will withstand the loads

generated by the pressing operation. It must run 24 hours/day, 7 days a

week for extended periods. It must b e best operated in a cool clean

environment and should be separated as far as possible for the hot dirty

environment of the pressing sections.

4.1.3 Cage Assembly

The cage should be split on the centre line with the two halves bolted

together. This design should allow for quick and easy access to the press

internals for cleaning or inspection.

When closed the cage lock firmly to the frame, to prevent it from

moving relative to the frame

The cage should form the drained barrel of the press. To form the

drained barrel the cage should be lined with lining bars separated

by spacers.

Provide a simple but effective clamping device system to hold the

lining bars in place.

4.1.4 Worm Assembly

The worm assembly should feature two compression zones to

maximise the efficiency of oil extraction without generating the high

pressures that absorb power and cause excessive wear of parts.

The worm shaft should be made up of loose sections, built up on a

keyed shaft. The parts should slide on to the shaft from the

discharge end, so that those parts that experience the greatest

wear and those most likely to need changing for different seed

types are first removed from the shaft.

The compression in each zone should be adjusted to suit the feed

material and the required duty, with simply interchangeable

pressure pieces.

4.1.5 Thrust Bearing Assembly

The thrust bearing carries the thrust loads generated by the press. These

are quite high , especially when full pressing a fibrous feed material. The

design should ensure that the bearings are located at the feed end of the

press and external to the main frame body to give the best working

environment(there is less pressure to force contaminants into the thrust

housing).

The design should allow separation of the thrust bearing from the gear

box oil and located between the working areas of the press and gear box

with a view to provide buffer zone to protect the more expensive gear

box internals from contamination.

4.1.6 Choke

The Choke is the discharge point of the solid residual from the press. The

cake at this point is under the highest pressure and it is the driest. This is

therefore potentially a high wear zone of the press. The design should

provide for a simple and effective choke system that needs a minimum

of adjustment. It should be long enough to form a good cake; have cake

cutters that can be easily be changed; and stand up to the high pressures

and the hot, dirty and steamy environment.

4.2 Outline Technical Specifications

4.2.1 Overall Press Requirements

The design offered must be field proven and manufactured using locally

available materials.

4.2.2 Capability

The press must be capable of performing the tasks as specified in the

Functional Requirements Specification under conditions and with

performance as defined in this Technical Requirement.

4.2.2.1 Availability

The press must be available 24 hours every day.

4.2.2.2 Environmental Conditions

Temperature : 25°C

Humidity : 20 – 80 % without condensation

Dust : to IP 65 specification

Noise : 85 dB(A)

4.2.2.3 Performance

Exact performance will depend on the type and quality of feed material

and the pretreatment used in the system . However, the press is expected

to achieve the following:

Throughput: 100-120 kg/hr (for feeds with 45% oil)

100-120 kg/hr (for feeds with 25% oil)

Oil in Cake: 8 -12 % (for feeds with approximately 45 %)

6 – 8 % (for feeds with approximately 25%)

Power requirements : 7.5 – 10 kW (main drive)

Energy: 75 Wh/kg seed processed

4.2.3 Reliability

The worm shaft with the bearings supported on the frame, is subjected

high torsional stresses. The design should allow the press to be used for

over 10 years.

4.2.4 Maintainability

Cleaning of the cage assembly, lubrication of the gearbox to be done

periodically for continuous running of the press. Rebuilding the worm

parts to be done by qualified technicians.

4.2.5 Size

The overall size of the press to be as follows:

Length: 1500 mm

Frame width: 500 mm

Gear box width: 500 mm

Nett weight : 200 kg

Cage dimensions: bore- 100 mm

No of fields – 3

Lining type-hard faced, landed bars

4.2.6 Product Cost

The final product has been projected to cost between K150,000 and

K400,000.

4.2.7 Quantity

Production quantities must be 10 units per annum.

4.2.8 Manufacture

Barrel bars, worm parts will be produced by sand casting method, or any

other cheap but suitable method. These will then be assembled. Shafts

will be produced by turning on the lathes.

4.2.9 Materials

Materials should be locally available. The product will mainly comprise of

metal which should be easy to form, fabricate and weld.

4.2.10 Standards

The product will conform to the standard specified by the BS1303

4.2.11 Product Life Span

The product will remain in production for up to 5 years.

4.2.12 Safety

To enhance operators safety, all moving parts shall be guarded;

The mechanism shall have good stability over the entire range of

operation induced vibrations.

4.2.13 Testing

The prototype will be tested for fatigue ,and torsional shear stresses.

Thereafter ,testing will be done for performance only.

CHAPTER 5: CONCEPTUAL DESIGNS

Concept 1: China Expeller

Concept 2: Dong- Kwang

Concept 3: Keller- P0015

Concept 4: Sundhara- Presses

Concept 5: Taeby-Press

Concept 6: Tiny Tec-Expeller

TABLE 1: CONCEPT EVALUATION MATRIX

CONCEPTCRITERIA

1 2 3 4 5 6

R 2 3 4 5 6 1 R 3 4 5 6 1 2 R 4 5 6 1 2 3 R 5 6 1 2 3 4 R 6 1 2 3 4 5 R

EASE OF OPERATION E - - + - - - E S - - + + + E - - - + S + E S + - S S + E - + S - + + E

EASE OF MAINTENANCE F

S S - S - SF

S S S S + -F

+ S S + - SF

- + + - - SF

S - + S S -F

OVERALL COST

E - - - - - + E S + S - + S E + - - + + S E - S + + S S E + + - - + - E

SAFETY R S S S S S S R S + S + S S R S S S - - R S S S S S S R S S S + S S R

PERFORMANCE E + + + + + - E S S S + - - E + - - - s + E S + - - S + E + - S S + - E

EASE OF MANUFACTURE

N - - + - + + N - + - - + S N + - - + S S N - + + - - + N + + + + + S N

NUMBER OF PARTS

C - - + - - C - S S S - - C - - S S - S + - + - - - S C S S + - + - C

E E E E E ETOTAL NUMBER OF `` +``

9 7 8 13 6 14

TOTAL NUMBER OF`` S``

8 16 10 13 14 11

TOTAL NUMBER OF ``-``

17 7 14 9 10 10

+ = Better than, S = same as , - = worse than

Note : the shaded concept was chosen for further development

The following factors will influence the choice of expeller design type:

1. Cost

The target user group of the press will be rural Malawians, most of these

live below the poverty line (i.e. earning less than USD 40 per annum).

The cost of the end product will therefore be kept as low as possible. The

design chosen will be one whose maintenance and operating cost are

relatively lower.

2. Technical Aspects

The worm assembly design

3. Ergonomics

4. Legal Aspects

CHAPTER 5: DESIGN CALCULATIONS

NOMENCLATURE

Density of oil seed (kg/m3)

Pitch (m)

D Hub diameter (m)

R1 Inner radius (m)

R2 Out radius (m)

N Shaft speed (rev/min)

Volume displacement

Q Production capacity

Pitchline velocity of gears

Angular velocity of gears

Angular velocity of gears

m Module for gears

Velocity ratio of gears

Tooth fillet radius of gears

n Speed of screw

Theoretical screw volume

Filling ratio of screw

Mass flow rate

P Power

Worm efficiency

f Displacement factor of screw

Allowable bending stress of shaft

Allowable torsional stress

T Torque

Tangential gear tooth force

Separating gear tooth force

Summation of vertical forces acting on shaft

Summation of turning moments acting on shaft

RA Reaction force acting on shaft at position A

RB Reaction force acting on shaft at position B

BM Bending moments acting on shaft

Equivalent torque acting on the shaft

d Diameter of shaft

C Circular pitch of gears

Safety factor of gears

Special factor of gears

Stress concentration factor

Minimum allowable tooth width

Slip factor of motor

Motor power

Shaft power

5.2.1 Determination of theoretical screw volume per pitch

We know that the theoretical screw volume is given by

= area x pitch

= (d2 – d2 s)

4Substituting into the formula

= (0.0962 – 0.0682 ) 0.065

4 = 2.344 x 10-4 m3

5.2.2 Determination of mass flow rate

Mass flowrate is dependent on product density , theoretical screw volume per pitch, speed of screw and filling.

i.e. = n

From table --- on page ---, = 45 %

= 2.344 x 10-4 x 40 x 640 x 0.45

= 2.701 kg/s

5.2.3 Determination of drive power

This is dependent on flow rate, gravity , displacement length, resistance.

i.e. P( , , L, , f i , )

Neglecting gravitational effects,

Drive Power = x L x

From tables,--- f I = 1.9 , = 0.9

Power = 2.7 x 0.8 x

= 4.56 k W

5.2.4 Electric Motor Power

5.2.4.1 Determination of efficiency

Adopting JIS on data for maximum and minimum efficiency as shown in appendix ---, the minimum efficiency of 79 % is chosen.

5.2.4.2 Worm Shaft power

Ps = = =

= 5.77 k WThe worm shaft will transmit 5.77 kW of power.

5.2.4.3 Motor Power

= ( 1 + )

From table 10.3 on appendix…, a value of = 0.25 is chosen. So we have

P = 5.77 (1 + 0.25) = 7.2 kWThe next available motor power is 7.5 kW and this is the one adopted for the design.

5.2.4.4 Actual motor speed

= (1 – )

The actual motor speed is given by where is the slip factor. Using table 10.4 on appendix…., is determined as = 6%. since we have

= 1440 (1- 0.06) = 1354

5.3 Design of worm shaft

In an expeller the design of the screw has basic importance as it determines the

performance efficiency. Literature survey reveals that there is scanty information

describing the effect a specific configuration work appears to have been done by

manufacturing industries using empirical approach . A more rigorous scientific approach

is thus required to predict more accurately the results of a particular worm configuration.

5.3.1 Screw Configuration

The screw will consist of 7 worm sections including reverse worm in the sixth position.

Table 1 : Screw Configuration – Option 1

Number of worm sections 1 2 3 4 5 6 7

Length of worm (cm) 12.7 10.16 7.62 5.08 4.46 3.81 3.81

Pitch of worm (cm) 12.7 10.16 7.62 5.08 3.81 3.2 3.2

Screw hub diameter (cm) 6.8 6.8 6.8 6.8 6.8 6.8 6.8

Spacers (starting from 2nd worm. 1.5 cm

Table 2: Configuration – option 2

Number of worm sections 1 2 3 4 5 6 7

length of worm(cm) 16.5 10.16 7.62 6.35 5.71 5.08 4.44

Pitch of worm(cm) 12.7 10.16 7.62 6.35 5.71 5.08 4.44

Screw hub diameter(cm) 6.8 6.8 6.8 6.8 6.8 6.8 6.8

Compression ratio: 1 :3.5, reverse 6th worm

Theoretical compression ratio 1:20

Table 3: Configuration - option 3

Number of worm

sections

1 2 3 4 5 6 7 8

length of worm(cm) 15.54 6.35 6.35 6.35 6.35 11.43 6.35 12.7

Pitch of worm(cm) 15.54 6.35 6.35 6.35 6.35 11.43 6.35 12.7

Screw hub diameter(cm) 6.8 6.8 6.8 6.8 6.8 6.8 6.8

5.1 Capacity of worm shaft

5.1.1 Determination of capacity of worm shaft

Displacement of each worm along shaft

70 kg mass translates to 70

640 m3 .

At 40 rev/min the displacement is 70 m3 /min

640 x 60

Each worm displaces 70 m3 /min

640 x 60 x 40

= 4.56 x 10 –5 m3/min

But distance moved by material for one worm at revolution = pitch

And volume VL = Vo - VI

= 2L (R1 – R2)

= 2 X 0.065 ( 0.04 – 0.03)

= 4.08 x 10-3 m3/min

Actual production capacity Q (kg/hr)

= 4.08 x 10-3 x 640 x 60 x 40

= 62.7 kg/hr

1.1 Determination of mass flow rate of each worm

From production capacity of 62.7kg/hr, the amount of material

Entering the shaft = 1.05kg/min. Assuming oil content of 40% and

residual content of 10%,

Weight of oil pressed along the length of the shaft/min = 0.428

kg/min and weight of cake plus fines = 0.622 kg/min.

Mass flowrate of material flowing across each worm

Each worm displaces 1.05 m3/min = 4.10 x 10 –5 m 3/min

640 x 40

3.3 Apparent Density

Apparent density = weight of material passing across a worm

volumetric displacement of that worm

= apparent bulk density (kg/m3)

or by computing the compression ratio = bulk density of 640

(kg/m3) as it enters the worm shaft, and is expected to return to

reach compensated bulk density of 1280 (kg/m3) as it flows the

last worm of the shaft. If the shaft has 7 worm segments, we

predicted that X % of the compaction would occur at any given

worm segment, say 70% by worm segment #3.

3.5 Gear box Design

Velocity ratio = 1/6

Driven gear : module 6, 96 teeth to rotate at 40 rpm

Angular velocity rv = n2 = 40 = 240 rpm n1 1/6therefore number of teeth on the driverrv = N1 , N1 = rv N2

N2

= 1/6 x 96 = 16 teeth

but pitchline velocity = pitch circle radius x angular velocity

vp = r2w2 and m = d2/N2

thus r2 = d2/2 = m N2 = 6 x 96 = 288 mm 2 2

w2 = n2 x 2 = 40r x 2 rad x 1 min = 4.2 rad/s 60 min 1 r 60s

and vp = r2w2 = 288 x 4.2 = 1209.6 mm/s

since vp is also r1w1 and r1 = 1/6 x 288 = 48 mm

w1 = w2 = 4.2 = 25.2 rad/s rv 1/6

and vp = r1w1 = 48 x 25.2 = 1209.6 mm/s

An investigation for combined bending and torsion under fatigue loading

PULLEYPINION GEAR

Pinion details : 16 teeth, 6 mm module, 20 pressure angle

Tooth fillet ro = 3.4 mm

Required safety factor (on fatigue limit) = 5

Material 220 MO 7 (EN 8) steel

Tensile strength = 620N/mm2

Bending fatigue limit = 0.4 x 620 = 248 N/mm2

Shear strength = 370 N/mm2

Torsional fatigue limit = 0.4 x 370 = 148 N/mm2

Bending stress concentration factor = 1.2

Allowable bending stress = 248/1.2 x 5 = 41.33 N/mm2

Torsional stress concentration factor = 1.5

Allowable torsional stress = 148/1.5 x 5 = 19.73 N/mm2

PINION SHAFT

Therefore allowable combined stress = ( ) 2 + 2

2

= (41.33)2 + 19.732

2

= 28.57 N/mm2

Speed of pinion = 240 rpm 25.2 rad/s

Power transmitted = 5 KW

Therefore torque T = P/

= 5000 = 198.4 Nm 198400 Nmm

25.2

Pitch circle diameter of pinion = 16 x 6 = 96 mm

Therefore pitch circle radius = 48 mm

Tangential gear tooth force Ft = T = 198400 = 4133.3 N

r 48

Separating gear tooth force Fs = 4133.3 tan 20 = 1504.4 N

Resultant tooth force = ( 4133.32+ 1504.42)

= 4398.6 N

Shaft load for a 600mm diameter pulley

Shaft load = torque = 198.4 = 661.3 N

Pulley radius 0.3

0.14 0.14 0.09 4398.6N 661.3N

RA RB

For equilibrium , Fv = 0, M = 0

RA + RB = 5059.9 N

Taking moments about A ,(positive clockwise)

(4398.6 x 0.14) + ( 661.3 x 0.37) – 0.28 RB = 0

therefore RB = 3073.2 N , RA = 1986.7 N.

0.14 0.14 0.09 4398.6N 661.3N

1986.7 3073.2 RA RB

1986.7

661.3

2411.9

-

59.5

0 0

Taking +ve clockwise

BM at A = 0

BM at B = 1986.7 (0.14) = 278.1 Nm

BM at C = -4398.6 (0.14) + 1986.7(0.28) = - 59.5 Nm

BM at D = -4398.6 (0.23) + 1986.7 (0.37) + 3073.2 (0.09) = 0

Therefore maximum bending moment = 278.1 Nm 278100 Nmm

Equivalent torque = M2 + 2

= ( 278100)2 + (198400) 2

= 341617 N mm

SHEAR FORCE DIAGRAM

BENDING MOMENTDIAGRAM

Te = qJ r

341617 = 28.57d4/32 d/2

d = 3 1/2 x 341617 x 3228.57

= 39.34 mm

Minimum allowable diameter = 40 mmGear Capacity

The pinion was the weaker of the two mating gears, and was more

likely to fail in strength than wear.

Using the Lewis Formula

Ft = YbC , and

Y = 0.154 – 0.912

N

= 0.154 – 0.912 = 0.097

16

Therefore circular pitch C = x module

= 6 x 3.142 = 18.85

Allowable stress = tensile or compressive strength

= SF x SP x Kt

where SF = safety factor

SP = speed factor

Kt = stress concentration factor

Pitch line velocity V = 1209.6 mm/s

SF = 5

Ratio = ro/C = 1/18.85 = 0.053

Stress concentration factor Kt = 1.6

Therefore , SP = 3000 + 1209.6 = 1.28

3300

allowable stress = 620

5 x 1.28 x 1.6

= 60.55 N/mm2

but tangential tooth load Ft = 4133.3

therefore, minimum allowable tooth width b = Ft

Y x C x

= 4133.3

0.097x18.85x60.55

= 37.34 mm

= 40 mm

therefore tooth thickness = m

2

= x 6/2 =9.42 mm

= 10 mm

3.7 Drive design

A drive was required from a 1440rev/min direct on line start

electric motor to a gearbox which had to run at 240 r/min for 8

hours a day at approximately 600 mm centers

Motor shaft was 35 mm diameter and gearbox 50mm dia.

Speed ratio = 1440/240 = 6 : 1

Service factor (from table 1) = 1.6, duty factor = 1.2

Design power = 5 x 1.6 x 1.2 = 9.6 KW

Belt section : 9.6 KW at 1440 rpm may be transmitted by SPA or

SPB. SPB was adopted.

Minimum pulley diameter from catalogue Table1 , 9.6 KW at

1440 rpm gives minimum pulley diameter = 80. 100mm diameter

pulley was selected.

Large pulley diameter = 100 x 6 = 600mm . It was accepted.

Belt length and center distance diameter of large + diameter

of pulley

small pulley = 700mm

Correction factor = 0.95

Basic power per belt from power rating table 140 mm and 1440

rpm gave 5.73 KW

Speed ratio power increment = 1.21

Corrected power per belt (basic power + increment) x factor

= (5.73 +1.21) x 0.95 = 6.59 KW

Number of belts = Dsign power = 9.6/6.59 = 1.46

Power per belt

Therefore use 2 SPB wedge belts.

Bore size

From dimension tables, a 100mm x 2 SPB has a maximum metric

bore of 60mm which is greater than the 35mm diameter motor

shaft. The 600 mm x 2 SPB also had a maximum metric bore of

75mm and was suitable for the 50mm gear box shaft.

Therefore , the drive specification adopted was:

Motor pulley - 100 mm x 2 SPB

Taper lock Bush - 2517 x 35 mm

Gear box pulley - 600 mm x 2 SPB

Taper lock Bush - 3020 x 50

Fenner wedge belts - 2 x 16 N (SPB)

Belt length = 3550mm

Centre distance = 746 mm

APPENDIX

(kW)

0.4

0.75

1.5

0.4

2.2

3.7 0.4 ~ 0.25

5.5

7.5

11

15

18.5

22

30

37

0.25 ~ 0.15

45 0.15 ~ 0.10

Table __: Data for typical allowances made for motor efficiency.

Table ____: Data for percent slip of different types of electric motors

kW

(%)

2 P 4 P 6 P

E E E

0.2

0.4

0.75

1.5

2.2

3.7

9.5

8.0

7.0

6.5

6.0

5.5

10.0

8.5

7.5

7.0

6.5

6.0

10.0

8.5

7.5

7.0

6.5

6.0

10.5

9.0

8.0

7.5

7.0

6.5

9.5

8.0

7.5

6.5

6.0

10.0

8.5

8.0

7.0

6.5

5.5

7.5

11

15

(19)

22

30

37

5.5

5.5

5.0

5.0

5.0

4.5

4.5

4.5

6.0

6.0

5.5

5.5

5.5

5.5

5.5

5.0

5.0

5.0

5.0

5.0

6.0

6.0

6.0

5.5

5.5

5.5

5.5

5.5

5.0

5.0

5.0

5.0

6.0

6.0

6.0