Design and Development of Screw Press

by John Taulo Deputy Director (Research and Development) Malawi Industrial Research and Technology Development Centre P.O.Box 357, Blantyre E-mail: [email protected] Website: mirtdcmalawi.com

April 2005

ABSTRACT 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: 1.1

INTRODUCTION

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: 2.1

REVIEW OF LITERATURE

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 m 3/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 4.1.1

Functional Requirements Specifications 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

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

:

and

with

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 : Energy: 4.2.3

7.5 – 10 kW (main drive)

75 Wh/kg seed processed 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

CONCEPT CRITERIA

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

S

S

-

S

-

S

S

S

S

S

+

-

+

S

S

+

-

S

-

+

+

-

-

S

S

-

+

S

S

-

F

F

F

F

OVERALL COST SAFETY

E

-

-

-

-

-

+

E

S

+

S

-

+

S

E

+

-

-

+

+

R

S

S

S

S

S

S

R

S

+

S

+

S

S

R

S

S

S

-

-

PERFORMANCE

E

+

+

+

+

+

-

E

S

S

S

+

-

-

E

+

-

-

-

s

EASE OF MANUFACTURE NUMBER OF PARTS

N

-

-

+

-

+

+

N

-

+

-

-

+

S

N

+

-

-

+

C

-

-

+

-

-

C

-

S

S

S

-

-

C

-

-

S

S

TOTAL NUMBER OF `` +`` TOTAL NUMBER OF`` S`` TOTAL NUMBER OF ``-``

E

-

S

+

+

S

S

E

+

+

-

-

+

-

E

R

S

S

S

S

S

S

R

S

S

S

+

S

S

R

+

E

S

+

-

-

S

+

E

+

-

S

S

+

-

E

S

S

N

-

+

+

-

-

+

N

+

+

+

+

+

S

N

-

S

+

-

+

-

-

-

S

C

S

S

+

-

+

-

C

7

8

13

6

14

8

16

10

13

14

11

17

7

14

9

10

10

S = same as ,

E

F

E 9

+ = Better than,

E

S

F

-

= worse than

E

E

E

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

ρ

Density of oil seed (kg/m3)

S

Pitch (m)

D

Hub diameter (m)

R1

Inner radius (m)

R2

Out radius (m)

N

Shaft speed (rev/min)

V

Volume displacement

d

Q

Production capacity

V

pg

Pitchline velocity of gears

ω2 g ω1g

Angular velocity of gears

m

Module for gears

r r

v o

n

Angular velocity of gears

Velocity ratio of gears Tooth fillet radius of gears Speed of screw

Vs

ϕ •

Theoretical screw volume Filling ratio of screw

m

Mass flow rate

P

Power

η

Worm efficiency

f

Displacement factor of screw

δ

Allowable bending stress of shaft

τ

Allowable torsional stress

T

Torque

F F

Tangential gear tooth force

t

Separating gear tooth force

s

F ∑

Summation of vertical forces acting on shaft

v

∑M 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

T

e

Equivalent torque acting on the shaft

d

Diameter of shaft

C

Circular pitch of gears Safety factor of gears

SF

Special factor of gears

SP

K

t

Stress concentration factor

b

Minimum allowable tooth width

s

Slip factor of motor

P P

m s

Motor power Shaft power

Determination of theoretical screw volume per pitch

5.2.1

We know that the theoretical screw volume is given by Vs

= area x pitch

Vs

=

π

(d2 – d2 s) S

4 Substituting into the formula Vs

=

π

(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.

•

m =V

s

ρϕ

n

From table --- on page ---,

ϕ = 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(

m

, Vs , L,

ρ ,f η i,

)

Neglecting gravitational effects,

∴

•

Drive Power =

m x Lx

From tables,--- f I = 1.9 ,

∴Power

η

= 2.7 x 0.8 x 1.9

= 4.56 k W

5.2.4 Electric Motor Power

=

0.9

0 .9

fi

η

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 P

4.56

η = 0.79 =

Ps =

= 5.77 k W The worm shaft will transmit 5.77 kW of power.

5.2.4.3

Motor Power

P

m

=

P

s

(1+

δ

)

From table 10.3 on appendix…, a value of δ = 0.25 is chosen. So we have P = 5.77 (1 + 0.25) = 7.2 kW The next available motor power is 7.5 kW and this is the one adopted for the design.

5.2.4.4

Actual motor speed

n =

N (1 –

s)

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

n

= 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 Length of worm (cm) 12.7 Pitch of worm (cm) 12.7 Screw hub diameter (cm) 6.8 Spacers (starting from 2nd worm. 1.5 cm

Table 2:

2 10.16 10.16 6.8

3 7.62 7.62 6.8

4 5.08 5.08 6.8

5 4.46 3.81 6.8

6 3.81 3.2 6.8

7 3.81 3.2 6.8

Configuration – option 2

Number of worm sections 1 2 3 length of worm(cm) 16.5 10.16 7.62 Pitch of worm(cm) 12.7 10.16 7.62 Screw hub diameter(cm) 6.8 6.8 6.8 th Compression ratio: 1 :3.5, reverse 6 worm

4 6.35 6.35 6.8

5 5.71 5.71 6.8

6 5.08 5.08 6.8

7 4.44 4.44 6.8

Theoretical compression ratio 1:20

Table 3: Number

Configuration - option 3 of

worm 1

sections length of worm(cm) Pitch of worm(cm) Screw hub diameter(cm) 5.1 5.1.1

15.54 15.54 6.8

2

3

4

5

6

7

8

6.35 6.35 6.8

6.35 6.35 6.8

6.35 6.35 6.8

6.35 6.35 6.8

11.43 11.43 6.8

6.35 6.35

12.7 12.7 6.8

Capacity of worm shaft Determination of capacity of worm shaft

Displacement of each worm along shaft 70 kg mass translates to

70 640 m3

.

m3 /min

At 40 rev/min the displacement is 70 640 x 60 Each worm displaces

m3 /min

70 640 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 =

2π L (R1 – R2)

= 2 π 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 n1

= 40 1/6

=

240 rpm

therefore number of teeth on the driver r v = N1 , N1 = r v 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 w2

=

n2

=

d2/2 = m N2 2 =

x 2π

60 and vp = r2w2

=

6 x 96 2

40r x 2π min

= 288 x 4.2

=

rad x 1 min = 4.2 rad/s 1r 60s

= 1209.6 mm/s

since vp is also r1w1 and r1 = 1/6 x 288 w1

=

w2 rv

and vp

=

=

4.2

288 mm

=

48 mm

= 25.2 rad/s

1/6

r1w1 =

48 x 25.2

= 1209.6 mm/s

An investigation for combined bending and torsion under fatigue loading PULLEY

PINION GEAR

PINION SHAFT

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 δ = 248/1.2 x 5 = 41.33 N/mm2

Torsional stress concentration factor = 1.5 Allowable torsional stress τ = 148/1.5 x 5 = 19.73 N/mm2 Therefore allowable combined stress =

√[( δ )2

+

τ

2

]

2 =

√[(41.33)2 + 19.732

] 2 28.57 N/mm2

= Speed of pinion = 240 rpm Power transmitted

25.2 rad/s

= 5 KW

Therefore torque T = P/ω =

= 198.4 Nm → 198400 Nmm

5000 25.2

Pitch circle diameter of pinion Therefore pitch circle radius

= 16 x 6 = 96 mm =

48 mm

Tangential gear tooth force Ft =

T

= 198400 r

Separating gear tooth force Fs Resultant tooth force

=

=

4133.3 N

48

4133.3 tan 20° =

= √ ( 4133.32+ 1504.42)

1504.4 N

=

4398.6 N

Shaft load for a 600mm diameter pulley Shaft load = torque

=

198.4

Pulley radius

0.14

0.14

=

661.3 N

0.3

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 0.14

,

RA = 1986.7 N. 0.14

0.09

4398.6N 1986.7

661.3N 3073.2

RA

RB

1986.7

661.3

SHEAR FORCE DIAGRAM

2411.9

59.5 0

0

BENDING MOMENT DIAGRAM

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 = = = Te J

+

τ

2

]

√[( 278100)2 + (198400) 2 341617 N mm

= q r

341617 π d4/32 d

√[M2

=

28.57 d/2

=

3

=

39.34 mm

√ [1/2 x 341617 x 32] 28.57 π

Minimum allowable diameter = 40 mm Gear 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 = YbCδ , and

Y = 0.154 – 0.912 N

= 0.154 – 0.912

=

0.097

16 π x module

Therefore circular pitch C = = Allowable stress δ

= =

where

6 x 3.142 = 18.85

tensile or compressive strength SF x SP x Kt

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

=

Stress concentration factor

0.053 Kt

=

1.6

Therefore , SP = 3000 + 1209.6

=

3300 allowable stress

=

620 5 x 1.28 x 1.6

=

60.55 N/mm2

but tangential tooth load Ft = 4133.3

1.28

therefore, minimum allowable tooth width b =

Ft YxCxδ

=

4133.3 0.097x18.85x60.55

=

37.34 mm =

π

therefore tooth thickness =

40 mm

m 2

= =

3.7

π x 6/2

=9.42 mm

10 mm

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 Design power = 5 x 1.6 x 1.2

=

= 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 1440 rpm gives minimum pulley diameter

=

80.

at

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 Correction factor

=

=

700mm

0.95

Basic power per belt →from power rating table 140 mm and 1440 rpm gave 5.73 KW Speed ratio power increment Corrected power per belt

→(basic power + increment) x factor =

Number of belts

=

= 1.21

(5.73 +1.21) x 0.95 =

Dsign power

=

6.59 KW

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

P

m

(kW)

δ

0.4 0.75 1.5

0.4

2.2 3.7 5.5

0.4

~ 0.25

7.5 11 15

0.25 ~ 0.15

18.5 22 30 37 45

Table __:

0.15

~ 0.10

Data for typical allowances made for motor efficiency.

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

s (%) 2P

kW

4P E

6P E

E

0.2

9.5

10.0

10.0

10.5

0.4

8.0

8.5

8.5

9.0

9.5

10.0

0.75

7.0

7.5

7.5

8.0

8.0

8.5

1.5

6.5

7.0

7.0

7.5

7.5

8.0

2.2

6.0

6.5

6.5

7.0

6.5

7.0

3.7

5.5

6.0

6.0

6.5

6.0

6.5

5.5

5.5

6.0

5.5

6.0

5.5

6.0

7.5

5.5

6.0

5.5

6.0

5.5

6.0

11

5.0

5.5

5.5

6.0

5.5

6.0

15

5.0

5.5

5.0

5.5

5.5

(19)

5.0

5.0

5.0

22

4.5

5.0

5.0

30

4.5

5.0

5.0

37

4.5

5.0

5.0