escooter project

Side view of Scoet

i

Executive

Summary

Every day we waste hours commuting to and from destinations. A remarkable feat of British design and engineering, Scoet

proposes an economically viable and environmentally friendly solution to this problem. It is a single person, efficient electric

scooter that will reduce the 10-minute journey time from South Kensington

station to Imperial College London to 4 minutes. Scoet boasts a robust deck structure with

sleek gear housing. With speeds of up to 24kmh-1 our design not only ensures speed but also long term usage and high safety- a minimum safety factor of 2 for our critical components. Scoet aims to provide the user with a unique and

enjoyable experience.

After creating some preliminary designs the group set out to design three

separate shafts in order to facilitate our 2-step transmission

system. The mixture of gears and pulleys provides a reliable and quiet transmission system. Our product design specification (PDS) outlined the

requirements for our scooter from which we then proceeded to calculate

stress, torque, fatigue, and deflection. The earliest design had a problem

concerning the ground clearance. After recalculating these dimensions and

readjusting the position of the housing to the rear of the scooter, the result

was propagated into a Computer Aided Design (CAD) model. In order to

assist our manufacturing process several components were further

simplified. In construction of the final design, some changes were made to

improve stability, reliability and effectiveness.

Reproduction for a large target market has been made possible by

the simplicity of the design and the reasonable part costs. The

total cost of raw materials for the

scooter is 90.75, this allows for reasonable profit margins whilst remaining competitive in the market. Preliminary

testing proved that the scooter design is robust, reliable and

safe for a 100 kg user to use on urban

terrain. Further design propositions include using lighter materials and a greater gear ratio to increase torque.

Isometric view of Scoet

Engineering

Drawings

Contents

1

Introduction

3

Concept

Layout

4

Product

Design Specification

7

Manufacturing Considerations

8

Project

Planning

9-10

Appendices

8

Discussion,

Conclusion and

References

5-6

Detailed

Component Design

and Finite Element

Analysis

1 -2

Engineering

Design

Analysis

1

The aim of the e-scooter was to modify a push scooter by creating an attachment that would include a kickboard and motorized rear wheel. The attachment would then be attached to the front wheels and steering assembly of the easily available Micro Monster push scooter. This provides an economically viable and modern approach to the solution to the problem proposed. The scooter will ensure faster and more efficient transport on the pavement.

How can we reduce the time and effort required for daily commute to Imperial?

This initial question provoked our responses to create an initial design plan which was summarised by the Gantt chart. This assisted us in accurately dividing our time out in our design stage of the process. We followed this process throughout the project in order to divide our manufacturing stage as well.

How can we ensure safety of our design?

Safety and reliability were key considerations in designing the e-scooter. The user could not be in contact with any uninsulated circuitry and the overall transmission- any high speed or hot components. Moreover, the motor could not overheat in the transmission casing. Similar to push off scooters, a mechanical brake had to be able to fully stop the e-scooter safely within a set short distance. In terms of portability, the scooter had to be light and small enough to carry on the tube easily on a daily basis during rush hours. The prototype had to be designed to withstand 100 hours of use by a 100 kg person to ensure durability and robustness of the product.

What were the constraints on our design?

In terms of design restrictions, the e-scooter had to include a plastic 120 mm diameter rear wheel, and must have a rotating wheel shaft. The power transmission had to include a 1kW effective power brushless DC motor, a rechargeable 22V Lithium battery and an electronic speed control system. Easily obtainable materials and parts were chosen to make the e-scooter to ensure that the e-scooter parts meet BSI standards and that the e-scooter can be easily reproducible for future manufacture. For practicality reasons and to ensure the design was not overly complicated, manufacture methods used were limited to those that mechanical engineering students could do in the Student Teaching Workshop (STW).

The final e-scooter prototype had to be a reproducible, portable, efficient, reliable and safe vehicle for college students who wanted a faster, easier alternative of transport.

Engineering

Design

Analysis

After identifying the main aims of the scooter,

further research on pre-existing e-scooters and

relevant drive transmissions in the market was

conducted for guidance and inspiration.

We started by brainstorming individually, then as a

team, any possible drive transmission ideas. It was

important to keep ideas varied and open at this

stage following the double diamond design process.

Introduction

In this current discover phase, we use secondary

research and creative thinking to brainstorm a

wide variety of ideas.

This is a divergent phase of thinking.

London has a population of approximately 9 million. The commuters that are part of this population put a daily stress on the public transport system that simply cannot handle the rapidly increasing number of both commuters and tourists. As a result a lot of the inner zone inhabitants (Zones 1, 2 and 3) have found using personal transport a faster, more efficient and cheaper method of travel. There are almost 20,000 cycling accidents every single year in London which has caused a great fear of the left hook on street bends. Although there have been several implementations to counter these accidents it seems that the pavement is rapidly becoming the only safe place on the rush hour streets of London.

This collage of initial ideas stems from a brainstorm session in which members of the team had to come up with as

many transmission ideas and PDS criterion as possible. Many ideas were repeated and some ideas were completely

unrealistic, however this did not matter as at this stage as no idea was to be rejected straight away. Rather, they were

meant to be developed upon and considered openly. Each member awarded (tick) marks to their favourite ideas. It was

upon these ideas that we were to further develop our design on. This process was integral to forming a core design idea

that was approved by everyone in the group. During the brainstorm it was made sure that everyone was comfortable

working on a specific design before we proceeded in any additional calculations. This made sure that everyone was fully

confident in what we were working on.

Outcome of Initial Brainstorm Session

2

Description Advantages -Greater reduction in speed whilst remaining relatively

compact. -More efficient and durable than belt drives, so better choice for high power transmission. -Unlike belt drives there is no propensity for the gears to slip hence making them more reliable

Disadvantages

-Steel spur gears would increase the weight of the scooter significantly. -More costly than using pulleys and belts. -Need to ensure high precision in shaft alignment for gears to mesh properly, this creates manufacturing difficulties due to the high loads the shafts experience.

Description Advantages -More efficient and compact than using a belt drive

transmission. -Reliability on par with gears as it can operate at high temperatures and will not deteriorate due to use in humid, sunny or dirty conditions.

Disadvantages

-Expensive, the shortest chain available would cost a minimum of 50 compared to the average cost of 10 for a pulley belt -Much heavier than the plastic pulley belt which decreases the portability of the e-scooter on the tube. -Requires frequent lubrication which would be impractical for daily use by a college student. -Manufacturing a chain drive would require tighter tolerances due to precise alignment needed. -The overall possible maximum transmission of a chain drive is lower than that of a gear or pulley belt drive.

Description Advantages -Much greater reduction in weight compared to gears.

-Less precision needed in manufacture, allows for speedier production. -Cheaper than spur gears.

Disadvantages

-More suited for low power transmission systems. -The shafts centre distances have to be a lot further apart than gears in order to achieve the angle of lap needed to prevent slipping. -Belt can wear and stretch over time and so is less durable and is more restricted in terms of operating temperatures. -Large amount of noise produced since timing belts compress air when rotating. Noise pollution makes usage rather ostentatious.

Aspect Weighting Score (1-5)

Weight 0.15 1

Compactness 0.10 4

Efficiency 0.15 4

Ease of manufacture 0.20 2

Reliability (in all weather) 0.05 5

Durability 0.10 4

Transmission ratio 0.15 2

Cost 0.10 1

Total 2.6


Weight 0.15 5

Compactness 0.10 3

Efficiency 0.15 3



Durability 0.10 2


Cost 0.10 5

Total 3.65


Weight 0.15 5

Compactness 0.10 2

Efficiency 0.15 5



Durability 0.10 5


Cost 0.10 3

Total 3.5

In this define phase, we analyse our research and

ideas to clarify our definition of the problem and

propose a rough solution.

This is a convergent phase of thinking.

Chain drive

transmission

2 step spur gear drive

Toothed pulley drive

However, whilst checking the design, the one-step transmission idea

was rejected mainly because only 6 teeth out of 10 teeth on the small

pulley would be in contact with the belt at any one time. Accounting

for the high speed of the motor shaft, this appeared inefficient and

impractical as a design since the belt would most likely slip off the

pulley. Another problem with the use of a 10 tooth pulley was its

maximum bore diameter of 8 mm. This would limit the design in using

a 2 x 2 key and keyway slot which would require higher skill and

precision in manufacture. A bigger pulley could not be used as this

would mean a maximum of 3 mm ground clearance between the

pulley belt and the ground.

Decision: We chose to develop the one-step toothed pulley belt drive

further as it had the greatest weighted score. A tensioner was

designed and 10 and 60 toothed pulleys were chosen to achieve

an appropriate total gear ratio. A very small pulley on the motor

shaft would allow the shaft and its components to hang from

beneath the board, creating a very compact design. A cut-in for

the large pulley and wheel would allow the rear shaft to hang

directly from beneath the board. However, in order to compensate

for the weakened board due to the cut out, a metal-plywood-

metal sandwich solution was decided upon, whereby two 1mm

thick sheets of steel would sandwich the rear half of the kickboard.

Having produced a range of ideas of transmission and features, we were then able to group the best ideas together for more technical analysis.

Engineering

Design

Analysis

In this define phase, we analyse our research and

ideas to clarify our definition of the problem and

propose a rough solution.

This is a convergent phase of thinking.

Initial one step and tensioner ideas

Concept

Layout

3

Shaft Positioning The shafts are in a triangular arrangement as shown above to minimise the horizontal and

vertical size of the assembly. The Vertical size was limited by how far down we could bring

the motor shaft before it interacted with the belt, and the horizontal size is limited by the

fact that our belt needed to be a certain length to guarantee we had adequate meshing.

Assembly Positioning The positioning of our assembly to be completely contained at the rear of the scooter is a result of multiple design iterations. Our initial design was located underneath the board and although it was cleaner in terms of aesthetics, the size of the gears and pulleys resulted in the board having to be at a very steep angle and high off the ground. Both would make the scooter not very ergonomic. By having all the shafts at the rear, we are able to ensure there is plenty of ground clearance underneath the board and the board is almost completely parallel with the ground.

Transmission Choices The first step of the transmission was done using spur gear reduction design due to the advantageous efficiency in high power and speed transmissions. To obtain the overall gear ratio, it was decided that the first step would achieve a ratio of 1:2. A greater ratio was not used as a larger gear would decrease ground clearance, make the transmission less compact and make the transmission significantly heavier. A 20 tooth spur gear was chosen to provide suitable face width for the high speeds of the motor shaft. By having the gears mesh between the motor shaft and the idler shaft, this would minimize the chances of gear misalignment since the main manufacture complication in the 2 step gear idea was having all 3 shafts parallel to each other.

The second step of the transmission used a pulley belt drive transmission to obtain a ratio of 1:3. To ensure suitable angle of lap and to remove the need of a tensioner, a 20 tooth pulley was used. By using a pulley belt drive in the second stage, misalignment issues caused by manufacturing and dynamic loading would be resolved.

Potential Improvements The housing of our assembly is comprised of five individual components. Two side plates

and two mounting blocks for the board, and one top plate. Ensuring all these components

line up precisely is the greatest challenge of our chosen design. One improvement that

could have been made would be to use the CNC machine and redesign our housing so it

could be created from two identical thicker side plates that could be directly mounted to

the board. This method would have resulted in a stronger cleaner construction as we would

have not had to use sheet metal, and may have been easier to line up the axis, however it

would have also added significant weight.

To spread out the stress in the connection between the plates and the board, intermediate aluminium blocks were used. This provided more surface area for the connection between the boards and the plates. Through bolting the aluminium blocks to the side plate horizontally rather than vertically this would decrease the shear force in the bolts when the kickboard deflects, since the stresses would be spread more evenly between the bolts, rather than having the bolt closest to the rear wheel carry most of the load.

Following the old one step design, the original plan was to have the entire transmission underneath the kick board, but after calculating the required incline angle of around 20 between the board and flat ground, it was decided that the transmission module would be moved to the back of the scooter. Although this lengthened the transmission, it increased ground clearance between the board and the ground. With the transmission to the back of the e-scooter, side plates would be required to hold at least the idler and motor shafts.

In order to decrease the bending moment caused by the user on the sheet metal, the rear wheel was placed right behind the rear edge of the kick board. This allowed the weight of the motor and idler shafts to further reduce the user-induced bending moment. The positioning of the shafts relative to each other allowed each components to move without interference, unless desired, and provided suitable ground clearance.

Our brake started in spring form to allow for metal

deflection but we decided to remove this to reduce

design complexity. However without this, even

allowing for extra thickness to avoid plastic

deformation, there would be too much force on the

boards bolts. So we loosened them to encourage a

pivoting motion to reduce this stress.

Bottom view of Scoet transmission

Final brake design

Brake idea 2

Brake idea 3

There was an issue with how attaching the rear wheel shaft to the bottom of the board would still create an incline angle between the ground and the board. To resolve this problem and to reduce the overall weight of the board, the idea of a cut-out for the wheel was removed, hence removing the need for the steel-plywood-steel sandwich. It was agreed upon that the side plates would hold all three shafts. This would also make manufacture easier, as CNC manufacture of the side plates would ensure a higher degree of parallelism between all three shafts.

Triangular positioning of shafts

2 step hanging assembly

The wheel coupling was created in

order to transmit the rotational

motion of the shaft to the wheel.

We could not manufacture a

keyway directly into the wheel due

to the fact that it is made out of

plastic and so would not be able to

withstand the high torques

needed. To combat this we

screwed an aluminium hub into the

wheel, which we could then put a

keyway in. The thickness of the hub

is limited by the minimum key size

necessary to transmit the torque.

In the develop phase we explore solutions to any problems in development in an as open ended manner as possible.

We then converge to our final solution in the

delivery phase.

Brake idea 1

Product

Design

Specification

4

Aspect Objective Criteria Test Plan

Performance Maximum Speed 24 kmh-1 On a 3 incline surface (tarmac) with driver weighing 67.5 kg On 50 m stretch of road in front of Eastside

Acceleration 0 to 24 kmh-1 in 20 m Accelerate to 24 kmh-1 in 20 m on 3 incline (tarmac) with driver weighing 67.5 kg, including an initial kick start

Calculation after speed trial

Durability No failure for 100 hours of service On continuous bumps of 0.02 m in size with 100 kg load on scooter Shakedown on a 100 m bumpy mew (cobblestones preferably)

Braking Distance (achieved by mechanical brake only)

Braking 5 m from 24 kmh-1 Deceleration from 24 kmh-1 in 5 m on tarmac with driver weighing 67.5 kg using mechanical brake only

Trial on stretch of road in front of Eastside

Maximum Mass 100 kg At 5 kmh-1 up and down a 0.1 m tall speed hump On speed hump within/near Imperial College with a mass of 100 kg on scooter

Operation in low temperatures -5C Fully operational at -5C, can fulfil acceleration and braking criteria Trial on stretch of road in front of Eastside during December.

Operation in rainy weather Performance unaffected by water Fully operational at -5C, can fulfil acceleration and braking criteria On 50 m stretch of road in front of Eastside on wet ground with water splashed on rear wheel + adapted parts of scooter during trial

Corrosion Resistance 12 months Must not have visible signs of corrosion 1 week after getting it wet and allowing it dry naturally indoors

Allowing it naturally dry after rainy weather test plan and checking for visible signs of corrosion 1 week after test

Package/ Dimensions

Ground clearance (between hanging components and board)

10 mm Minimum of 10 mm ground clearance when on a smooth, flat hard surface without person standing on the scooter

Measure using a calliper

Width of scooter 200 mm Width of widest part of scooter (rear wheel shaft) Measure using a steel rule Thickness of kick board 15 mm Maximum thickness of kick board Measure using a calliper

Length of kick board 400 mm Maximum length of wooden kick board Measure using a steel rule

Total mass of added adaptations 6 kg Includes board, rear module, battery, ESC system Electronic balance

Longest dimension of scooter 800 mm Maximum length of scooter Measure using a steel rule

Contact brake area 8000 mm Minimum area on the brake which the foot can step on (Total area of brake - area of brake parallel to board used to attach brake to wooden board)

Measure using a steel rule

Resources Cost 150 maximum Total cost of parts that require purchase from the RS catalogue (not including battery, ESC, permissible stock materials and motor) should be a maximum of 100 pounds

Calculate total price using RS catalogue

Manufacturing methods Can be manufactured using STW All manufacturing methods for parts must be doable in STW and all tolerances can be met using STW equipment and student operation of machines (except CNC)

Checking manufacturing methods and tolerances with machines available for use in STW before consolidating design

CNC part Part manufactured via CNC mill, lathe and cutter One part can be machined within the set tolerances using CNC machine tools available in STW- project coordinator must approve of part design

Check with project coordinator and workshop technicians that part is suitable for CNC can be made to correct tolerances

Spur gears, toothed pulleys, bearings

Standard Catalogue sizes Parts must be from the RS catalogue only Check catalogue for standard available sizes before consolidating design

Bearing housings, shafts, rear module housing, spacers, wheel coupling

Easily available material Part of permissible stock materials (Aluminium 6082 T6, Mild Steel EN1A)

Check permissible stock materials list to ensure that it can be made using the sizes of aluminium/ steel available

Bolts, screws, nuts, washers, keys Easily available material that fits the BS standard

Must comply with BS standard sizes in the list of permissible stock materials

Check permissible stock materials list to ensure that the size of bolt/screw thread/nut/washer/keys can be provided

Safety Protection against moving parts in motorized adaptation

Prevents user from getting injured by touching high temperature or high speed moving parts

Foot of user cannot directly touch motor, gears, pulleys, wheel during operation by 67.5 kg person

Trial on 50 m stretch of road in front of Eastside

Effective cooling of motor Prevents overheating of motor (leading to failure of scooter + smoke)

Temperature of motor must not rise more than 40C after 15 minutes of use

Leave motor in rear module running for 15 minutes

Protection against electrical components, cables and circuitry

Prevents user from getting electrocuted/ shocked and tripping over long cables

User should not get shocked by touching any part of the scooter when scooter is in use; user should not be able to trip on cables

Ammeter for checking electro-shock; Trial on 50 m stretch of road in front of Eastside to check whether cables will cause tripping

Deflection of kick board Prevents kick board from failing and causing injury

Kickboard must not have a deflection greater than 0.4% of the length of the board whilst 100 kg mass is placed on scooter

Calliper measurement difference between board and ground with and without 100 kg mass on scooter

Strength of shafts Prevents shafts from large deflections- causing misalignment in gears, bearings and prevents shafts from failing

Minimum safety factor of 2 Calculations of Safety factor via stresses in shafts

Reliability of mechanical brake Brake must remain intact after use and always be able to stop the scooter

Brake must be able to fully stop scooter in 5 metres from 24 kmh-1 90% of the time with a driver weighing 67.5 kg and not plastically deform

Test braking the scooter in 5 m from maximum speed in front of East side 20 times

Strength of side plates Prevent side plates from fracturing, bending or failing to perform to meet requirements

Minimum safety factor of 4 Calculations of safety factor via maximum stresses in the side plates

Edge/ Corner finishes All edges and corners must be smoothened/ rounded

User should not get a wound when using/ transporting scooter Visual inspection on quality of finish and design

To address all the design considerations that the e-scooter would have to fulfil, a product design specification was created for the chosen design concept.

Detailed

Component

Analysis

5

To obtain a transmission ratio of 1:2, gears with 20 and 40 teeth were chosen. The minimum

number of teeth allowable is 18 for a pressure angle of 20, however gears with 20 and 40

teeth were chosen as a suitable face width and module were more readily available. By

achieving a 1:2 ratio in the first step instead of a 1:3 ratio, the gear on the idler shaft would be

smaller and lighter, since a 60 tooth aluminium pulley was much lighter than a 60 tooth steel

gear. Although heavier, 080M40 was chosen as the material over POM and brass, as this would

provide a suitable safety factor to the gears without having the gears be needlessly large. A

module of 2 was confirmed after calculating what the minimum face width would be and

checking what was available and in stock on the catalogue. Details on the calculations can be

found in Appendix B.

Gears Selection

The minimum transmission ratio was determined to be 5.8 to provide sufficient torque for

acceleration. An overall transmission ratio of 1:6 was chosen to simplify gear and pulley choice. This

was then achieved by an initial 1:2 step down, followed by a 1:3 step down. The transmission was

determined to require 830W of power which, even accounting for the small power losses due to

transmission efficiencies of 98% per step-down, was well below the rated motor power of 1000W. This

meant that the motor could sufficiently provide the required torque to accelerate even without a push-

start. However, these calculations were done assuming no power losses through heat and assuming

no dynamic loads. Through testing, the actual e-scooter was predicted to much slower than the

maximum speed defined by in the PDS, but this was mainly due to the intrinsic limitations of the motor.

Details on the calculations can be found in Appendix A.

Transmission Ratio

Bearing Selection Deep groove ball bearings are the most

readily available and widely used type of

bearing in relatively light to normal load

applications. Hence they were chosen for use

in the scooter project. To decide on which

bearings were suitable for the application,

bearing calculations were performed, but

bearings were mainly chosen based on what

was available in the catalogue for set outer

and inner diameters. Since bearings on the

same shaft were made identical, the safety

factor of the bearings range from 4 to 15. The

safety factor was a lot higher for the rear

wheel shaft since the rear wheel would be

under higher loads and forces from the

collisions of the rear wheel with objects, such

as steps and pebbles, which would then be

directly transmitted to the shaft and thus the

bearings. Details on bearing calculations can

be found in Appendix D.

The tolerance of the inner diameter of the

bearing was found using the bearing loads

calculated. Using SKF as a reference, the

ratio of bearing load to was used to

determine the fit. SKF recommended a k5 fit

for normal loads (0.05

0.1) with a

shaft diameter between 17-100mm.

Therefore, as the bearings were within this

range, this fit was used.

Wooden Board A balance between comfort, compactness, safety and sturdiness in the kickboard had to be made. Minimising the size of

the board would make the scooter more portable, less heavy and would decrease deflections in the kickboard, making it

safer. On the other hand, it would give the rider less space on the board to place his feet on, thus making it less

comfortable. By making the wooden board 400 mm, this gave space for an average male foot to stand on the board, extra

space for the front part of the second foot as well as for bolted joints. For this board length, a safety factor of 3.8 was

calculated and was deemed to be reasonable considering that the load had been assumed to be the worst case scenario

of a point mass, although dynamic load factors were not included in the calculations. Calculations can be found in

Appendix E.

Pulley Selection A pulley belt transmission was chosen due to reasons listed in the engineering design analysis section, including easy installation, low maintenance and high reliability associated with the design. Despite the fact that belts are limited by

their power transmission capacity and speed ratio, the constraints specified by the PDS were within the limits for the

belts speed. A timing belt was chosen to prevent slip, thus allow better angular synchronisation between the driving and

driven shaft, ensuring a constant speed ratio. Flanges on the small pulley were necessary to reduce the propensity for slip

as the teeth had to always remain meshed in position throughout the arc of contact. The main disadvantage associated

with timing belts would be the large amount of noise generated as air is compressed between the teeth and the pulley.

This was viewed as less important as the pulley drive was used in the second slower step of the transmission, hence

reducing loss in efficiency. The belt material chosen was polyurethane due to its availability.

The ratio of 1:3 was best described by a 20:60 teeth pulley choice. From this ratio, the selection of pulleys, shown in

Appendix G, was optimised by choosing a large belt width whilst minimizing the difference in pulley diameter to reduce

the propensity for the belt slip. A salient factor when analysing the belt length was to calculate the teeth in mesh with

the pulleys. If this number were to be lower than 6 then the teeth would not mesh throughout the arc of contact and

therefore the mechanism would not work. This was an important consideration for the small pulley, and a key reason as

to why a smaller pulley was not chosen. Equations used in the calculations can be found in Appendix G, which found the

number of teeth in mesh to be 8.

Bearing Housings The bearing housings were originally designed so

three bolts would carry all of the load on the

bearings, but due to fatigue stress

considerations, the design was adapted so the

housings would sit inside the side plates. This

would improve durability and concentricity.

There are three parts to the pair of bearing

housings per shaft, there is one bearing housing

that constrains the outer race of the bearing and

one that does not constrain the outer race per

shaft. The fully-constrained bearing housing

consists of two parts- a main housing and a lid

as shown in Figure 1, and the unconstrained

bearing housing is shown in Figure 2 and allows

the bearing to move with the shaft in the

horizontal direction.

Figure 1: Fully constrained bearing

housing: Main housing (blue), 'lid' (green),

plate (pink)

Figure 2: Unconstrained bearing housing: Main housing (brown), plate (grey)

Detailed

Component

Analysis

6

Shafts Detailed calculations on the three shafts were performed to ensure that appropriate shaft diameter and safety factors were used. Greater detail can be found in Appendix C.

Motor Shaft Despite the unnecessarily large safety factor of 6.7, the

shaft diameter was not reduced as this was the diameter of

the pilot bore in the gear. Moreover, having big shoulders in

the shaft would increase stress concentrations and increase

the chances of not manufacturing a concentric motor shaft

that was parallel to the idler shaft. To reduce deflections in

the cantilevered beam to ensure parallelism with the idler

shaft, it was made as short as possible. Detailed analysis can

be found in Appendix C.1.

Idler Shaft Similarly, the idler shaft could not have a smaller maximum diameter

due to the pilot bore in the gear being 15 mm. However, the safety

factor of the shaft was a more reasonable 2.2. The miniscule

maximum deflection of 0.0175 mm would ensure parallelism between

the idler and motor shafts, ensuring gear mesh. The idler shaft was

made without shoulders and relied on the use of spacers to axially

constrain the parts. This reduced stress concentrations and made it

simpler to manufacture and adhere to the strict concentricity

tolerances. More details can be found in Appendix C.2.

Rear Wheel Shaft A maximum shaft diameter of 20 mm was required for the

wheel to fit, and a calculated safety factor of 2.6 would ensure

safety in case the person stepped on the wheel directly and in

cases of sudden impact force as well as dynamic loading. Steps

in the shaft were created to minimize the weight of the shaft,

and by having a graduated step-down this reduced stress

concentrations in the shaft. The step down was necessary to

ensure that the ends of the shaft were M10 threads in order to

use stock nuts and washers. By having steps in the shaft this

also allowed smaller sized bearings to be used to minimize the

size and weight of the bearings and their housings. In order to

ensure the durability of the rear wheel shaft that would suffer

the most from impact forces, an Abaqus analysis was

performed to check the calculated maximum predicted stress-

Figure 3. The calculated value was within a 3% margin of the

Von Mises stress obtained using Abaqus. Detailed calculations

can be found in Appendix C.3.

Figure 3: Finite element analysis performed on the rear wheel shaft using Abaqus

Brake The brake was designed as a pivot structure to prevent

excessive stress on the bolts in the board when stepping on the

brake, as illustrated in Figure 4. The bolts were not bolted in

tightly to allow it to act as a pivot. The exerted force required

to stop the scooter was calculated to be 143 N, meaning that

the weight on the back leg to provide the required deceleration

would be approximately 14.6 kg. A human weighing 67.5 kg can

comfortably exert this weight on the brake by pressing down

using one leg. Details on calculations can be found in Appendix

H.

To ensure that 2 mm thick aluminium could be used for the

brake yet remain in the elastic region after use, an Abaqus

analysis of the brake was performed (Figure 4), assuming that

the entire weight of the person (67.5 kg) acted on the top part

of the brake alone. A small elastic deflection of 3.91 mm and a

maximum stress of 63 MPa was determined, giving the brake

a reasonable safety factor of 2.9.

Figure 4: Finite element analysis of the brake using Abaqus

Side Plates The side plates were designed to connect the board to the three shafts as well as shaped to protect the user from moving parts in the transmission.

Since the plates were a critical part of the e-scooter, a greater safety factor was required for this part.

To check that the thickness of 3 mm was suitable, finite element analysis was performed on the side plates to ensure that the stresses in the plate

were within the fatigue limit. Forces used in the analysis can be found in Appendix F. The Von Mises stress- used for ductile materials was determined

to be 38.29 MPa, as demonstrated in Figure 5. Steel was chosen over aluminium as steel has a fatigue limit greater than aluminium. The fatigue limit

of mild steel - 180 MPa was approximated to be half the tensile strength (High Peak Steels Ltd., 2016), and would give the plates a safety factor of 4.7.

This meant that it was safe enough, accounting for sudden impact forces, that thicker plates would not be necessary.

The stress concentrations were found to be at the edge of the bearing housing and

the bolt holes for the bearing housings. These were in regions mainly determined by

the direction the forces on the bearings were acting on. The bolt closest to the wheel,

connecting the plate to the aluminium block had a higher stress concentration than

the other bolts as it would be taking a significant proportion of force from the

persons weight. By placing the rest of the gear transmission behind the rear wheel,

and having the wheel close to the board, a moment countering the couple generated

by the persons weight was induced thus reducing the force on this bolt.

The side plates were shaped so a top plate could be attached to shield the user

from accidentally placing his/her foot on the motor shafts spur gear. The side plate

was also designed so minimal material was used to reduce the weight of the e-

scooter.

Figure 5: Finite element analysis of one side plate using Abaqus

Manufacturing

Considerations

7

All our bearings used the k5 tolerance and so they needed to be manufactured with a precision of between +0.001mm and +0.009mm of the nominal diameter. The lathes that were available to manufacture allowed diameters to 0.005 mm. This meant that it was very difficult to manufacture to the required precision. In the case of the bearings because they were constrained axially and had radial loads, the inner race would not move even if the manufacturing of the component is slightly incorrect.

The gears and the pulleys were all held in place by keys such that a transition fit was not necessary. This meant that any deviations from the required tolerances (g6) were not a problem as long as the fit was a clearance fit.

Another unavoidable problem with manufacturing was that the tools used may have been worn out and old. This could cause deviations in the nominal values displayed by a machines digital read out as a part is being manufactured.

The gears that were being used had to mesh properly for the scooter to work as planned and to be efficient. Misalignments of the two shafts holding these gears needed to be reduced so that the gears could mesh properly. Because of this the side plates that support the bearing houses, which in turn support the shafts were manufactured using CNC laser cutting. This allowed the holes to be drilled in the most precise way, thus reducing the misalignment of the two shafts.

Because of the time constraints on the project it was important that the parts

were kept as simple as possible. This allowed the scooter to be manufactured in the allowed time and reduced manufacturing complications that would arise from complex parts, especially as the team didn't have extensive experience in manufacturing.

Brake and Top Sheet Bearing Housings Side Plates Aluminium Supports Wooden Board

Idler Shaft Motor Shaft Rear Wheel Shaft Spacers

To manufacture the components, appropriate tolerances needed to be chosen. Also an understanding of the limitations of the machines was important. The shafts were critical components that needed manufacturing to specific tolerances. The shafts needed to be concentric all the way along which meant that it was best to turn the whole part in one attempt on the lathe without removing it and returning it to the lathe, reducing the concentricity of each end. To achieve this, the length of material required was not to be longer than the specified length of the shaft and was then cut after all the turning had been done. The bearings, gears and pulleys that were fixed on the shafts also require specific tolerances.

Exploded view to highlight main components for manufacture

Special care had to be

taken in manufacture of

this component as only

one plank was supplied.

Sandpaper smoothed it

enough for aesthetic and

safety purposes. This

component required the

use of a pillar drill for its

holes and a band saw to

cut to size.

This component required

the use of a metal cutting

band saw to cut to size and

a milling machine for its

holes.

We manufactured using

CNC laser cutting. Curved

corners increased user

safety.

These two parts were

made using a sheet metal

bending machine, a

guillotine to cut to size and

a pillar drill for the holes.


the use of a lathe to cut

features to size and a

milling machine to create

the bolt holes.

The most important

dimension for these

components were their

lengths as they were crucial

for alignment. This

component required the

use of a lathe and a milling

machine. Combining

spacers of same inner/

outer diameters creates a

section of pipe that can be

bored in one go to create

several spacers.


the use of a lathe, milling

machine for the keyways

and a die to create the

external threads.

To attach the shaft to the

motor provided, holes

had to be milled. Studding

would connect these two

components. This

component requires the

use of a dye, lathe and

milling machine for the

keyways


the use of a lathe, milling

machine for the keyways

and a die to create the

external threads.

Project

planning

8

Aurisicchio, M. (2016). ME1 Machine Elements Notes. Department of Mechanical Engineering. London: Imperial College

London.

CES EduPack, c. (2016). Granta Design Limited.

Department of Mechanical Engineering, I. C. (2015). Mechanical Engineering Data and Formulae. London: Shell.

High Peak Steels Ltd. (2016). EN1A Steel Properties- 230M07. Retrieved December 02, 2016, from

http://www.highpeaksteels.com/en1a-steel-properties/

Kadiric, A. (2016). Introduction to Rolling Element Bearings and Mechanical Transmissions. London: Department of Mechanical

Engineering, Imperial College London.

To ensure coordination between team members and to maximise

productivity, a Gantt chart was created for the design and

manufacturing processes. This made sure that deadlines (coloured

in black) were met. Scheduled workshop and meetings (coloured

in orange) were designed as days to start new tasks. Buffer time

was not shown in the Gantt chart, although it was planned so

there was time just in case parts or sections had to be redone. This

report was written throughout the manufacturing phase of the

scooter report and often helped us spot overlooked problems and

issues with the design. Since there were only a certain number of

workshop sessions available to manufacture parts, it was vital that

the deadlines for each scooter part was met.

With writing the report, designing the components and manufacturing the scooter, members

were assigned tasks according to their strengths to maximise efficiency. This allowed us to

meet most of the internal deadlines, and created a lot of time for checking over each others

tasks and parts. Different people were assigned to make, check and approve each part to

guarantee that no errors had been overlooked. Towards the end of the manufacturing process

and after verifying sub-assemblies fitted together comfortably there was a shift of emphasis

towards the report. The Gantt chart allowed a structured approach to be followed in the long

term which allowed us to make certain milestones before their external deadlines. Also due to

the flexibility of our work process we could edit and adjust the chart so that it matched the

pace of the group in reality.

Discussion

One of the main limitations of the project arose from the manufacturing techniques. A lot of the work required exclusively the lathe and only a few components required the milling equipment or drills. We overestimated the amount of time we would need on the lathe which meant that our manufacturing process fell behind. Several components such as the spacers and bearing housings could have been simplified by adding grooves and circlips on shafts. Another limitation was using a mill to mill holes in circular components. This meant that there was a conversion between the CAD model which was using a radial co-ordinate system and a Cartesian co-ordinate system. This lead to minor inaccuracies which could have easily been rectified by using a CNC mill on these components. On a large manufacturing scale this would be realistic since the holes being milled were relatively simple.

Conclusion

Over the entire design and manufacturing process several iterations were made to the design in order to accurately fulfil all the details of the product design specification without over-complicating the manufacturing process.

The project brief stipulated that only one design of a part could be created by Computer Numerical Control (CNC) machine. We used this to create the most critical component for alignment: the side plates. By creating two identical parts using the laser cutter it would insure the accuracy when all the non-computationally generated components were added in the assembly. The rest of the manufacturing plan was designed around the three shafts and how all components were going to be constrained on them. Once the critical components had been manufactured then other components could be made and checked by fitting sub-assemblies together.

This allowed for parts to be remade if they didnt fit the specification and allowed problems with components to be determined early on in the manufacturing stage.

Having completed the assembly of our scooter the resultant transmission accurately illustrates our initial product design specification. From the assembly the structure seemed to hold steadily on its own and there were no loose-fitting components which is a testament to the protean design of the structure. The design overcame several difficulties that the team encountered: the clearance being too small, pulley size being too small and gears not generating enough torque. The CAD model specification which was based on calculations by both human and computerised methods was confirmed to be a virtuous design by the physical model that was produced from its drawings.

The evolution of Scoet correlated to the growth of the project team. After the manufacturing stage the increased integration between the team allowed more insightful thoughts about the design and further integration in writing the report. It is from this development that we are proud to produce an excellent product.

References

In future production, the design process could be ameliorated by the using the critical components as the defining structures for all calculations. During the initial design process a lot of time was spent designing a tensioner which eventually was removed from the entire structure since the newest pulley calculations did not require one for the chosen centre distance. In general, the main improvement would be to calculate identify the most reasonable reduction step. The initial design was rather optimistic with attempting to do 1:6 in a single step reduction.

In order to improve the overall design of the scooter lighter materials should be used to make the gear housing as the resultant structure in this

prototype, although strong, was rather heavy for its purpose. Another improvement to the design would be to increase the gear ratio so that the

resultant torque produced by the gears would be higher. This would mean that the acceleration of the scooter would be greater. The scooter is

designed for travelling a rather short distance therefore having a higher acceleration is somewhat critical since there are several traffic lights and

crossings on the journey there will be a lot of stopping and starting.

Appendices

9

Appendix A: Transmission

Ratio To calculate the required

transmission to provide sufficient

torque and speed to the e-scooter,

the total resistive force was

analysed. The force mainly consisted

of gradient, rolling and aerodynamic

resistance. Equation A.1 was used to

obtain the driving force at constant

maximum speed ().

In this evaluation, a maximum combined mass (m) of 110 kg was used

to calculate the driving force () required when travelling at 6.67

ms-1 up an incline () of 3.The rolling friction () between the ground

and the wheels was estimated to be 0.018, the maximum value within

the typical range (Kadiric, 2016). The density of air () for 15C was

used (1.2 kgm-3). The area of the front of the scooter () was estimated

to be 0.5 m2. The coefficient of drag () was estimated to be 0.9

assuming the scooter and person having a long cylindrical profile.

= + + 0.52 (A.1)

was calculated to be 87.9 N assuming constant velocity. To account

for acceleration, Equation A.2 was used. However, a mass of 67.5 kg

was used as the PDS specified the capability to accelerate to 6.67 ms-1

up an incline () of 3 within 20 metres. The required acceleration ()

was calculated to be 1.11 ms-2 using the equations of motion.

= + + + 0.52 (A.2)

was determined to be 124 N assuming constant acceleration. Since

accelerating requires a higher driving force, the torque () required

from the transmission was calculated using this value in Equation A.3.

The radius of the rear wheel (d) was measured to be 0.06 m.

= (A.3) A torque of 7.5 Nm was needed for the transmission and the power for

the wheel () was calculated using Equation A.4. For a speed of 6.67

ms-1, the angular velocity () was 111 radians per second.

= (A.4) The transmission required 830 W of power. To obtain the step-down

transmission ratio determined by the maximum velocity, Equation A.5

was applied.

=

(A.5)

The minimum ratio required to provide sufficient torque was

determined to be 5.8, using the motors specifications to find .

Appendix B: Gear calculations Gear calculations were done for a range of modules to find the suitable

gear size for 20 and 40 teeth gears. Firstly, the pitch diameter (d) was

found by multiplying the module (m) to the number of teeth in the

respective gears. The pitch line velocity (V) was calculated by Equation B.1 (Aurisicchio, 2016); n was the speed in rpm and was obtained by

multiplying the motors stated rpm per volt (280rpm/V) by the voltage

supplied (22V).

=

2

2

60 (B.1)

To account for the amount of impact during gear meshing, Equation

B.2 was used to obtain the dynamic factor () (Aurisicchio, 2016).

=6.1

6.1 + (B.2)

The transmitted load through gear meshing () was then calculated by

Equation B.3; this force is tangential to the pitch line. 100% power

transmission efficiency was assumed for the 1kW motor.

=Power

(B.3)

To determine a suitable face width for the gears, the Lewis form factors

(Y) had to be taken into account as seen in Equation B.4. The gears under

consideration had a module equal to the addendum, and thus the

relevant Lewis form factors were obtained through linearly

interpolation (Aurisicchio, 2016). The permissible bending stress of the

gear material, p, was found to be 117 MPa (Aurisicchio, 2016).

F =

v p (B.4)

Table B.1 shows the figures used and obtained for the chosen set of

gears. Finally, from the minimum face width required, a module of 2

was chosen due to its availability.

Table B.1: Calculated values for gears

Variable 20 Tooth Gear 40 Tooth Gear

d (mm) 40 80 n (rpm) 6160 3080 V (ms-1) 12.9 12.9

0.321 0.321 (N) 77.5 77.5

Y 0.30769 0.38117 F (mm) 3.35 2.71

Appendix C: Shaft Calculations In order to calculate the minimum shaft diameter required under

loading, stress and deflection calculations were performed on each

shaft. Before calculating the deflections, the forces induced by the

presence of gear mesh and pulley tension were obtained. Using the

transmitted load through gear mesh obtained previously (77.5 N), the

horizontal (W) and vertical (W) force components were

found after determining an angle of 40 between the horizontal and the

direction of transmitted load.

Using Equation B.1 and Equation B.3, the net force (1- 2) acting on

each pulley was found to be 194.83 N. These equations are valid for

pulleys as well as gears. To find the two tensions in the pulley belt- 1,

2, Equation C.1 was used ignoring the error induced by the differences

between a flat belt and a timing belt pulley. 100% efficiency in power

transmission was assumed, and an approximate coefficient of friction

() of 0.6 was used between the aluminium pulley and the polyurethane

belt. The angle of lap () on the idler shaft was 2.646 rad.

12

= (C.1)

Although 2 was determined through the model to only act in the

horizontal direction, 1 was split to its respective horizontal and

vertical components. Table C.1 shows the horizontal and vertical

forces arising from the pulleys and gears.

Table C.1: Calculated forces due to gear mesh and pulley belt tension

Parameter Value

(N) 59.38 (N) 49.83 (N) 274.57 (N) 150.69

(N) 118.37

Appendix C.1: Motor Shaft In calculating the maximum stresses in the shaft, axial stresses were

assumed to be negligible. Assuming that the gear and shafts weight

was negligible compared to the transmitted load, and simplifying the

connection between the motor shaft and the motor to a built in

support, a free body diagram of the shaft (Figure C.1) for the vertical

forces on the shaft was created. Msupport represented the moment at

the encastre support. The radius () was assumed to be a constant 6

mm all along the 63.9 mm long shaft.

Figure C.1: Free body diagram of the motor shaft

The second moment of area () for the motor shafts circular cross

section was calculated to be 1.02 X 10-9 m4 using Equation C.2

(Department of Mechanical Engineering, 2015).

=1

4 4

(C.2)

Using Figure C.1, the force along the beam was then calculated and

represented using Macaulays brackets. By integrating this equation,

the moment (M) relationship was obtained. By using the relationship

shown in Equation C.3 (Department of Mechanical Engineering,

2015), integrating twice and applying boundary conditions, the

deflection () of the beam was determined. The value used for mild

steels Youngs Modulus () was 207 GPa (Department of

Mechanical Engineering, 2015).

= 2

2

(C.3) (C.3)

Figure C.2: Vertical shear force, bending moment and deflection diagrams and equations for motor shaft

The respective equations and vertical shear force, bending moment

and deflection diagrams obtained are shown in Figure C.2. These also

show the maximum values of each parameter. The maximum normal

stress () experienced in the shaft was calculated via Equation C.4.

=

(C.4)

The above steps were repeated to find the shear force, bending

moment, deflection and maximum normal stress in the horizontal

direction. The values obtained for the motor shaft were written into

Table C.2.

Table C.2: Motor shaft- Vertical and horizontal values for parameters

Parameter Maximum vertical value (absolute)

Maximum horizontal value

(absolute) Shear force (N) 49.8 at x=0 59.3 at x=0

Bending moment (Nm)

1.19 at x=0 1.42 at x=0

Deflection (mm) 0.00378 at x=0.0639 0.00450 at x=0.0639 Normal stress

(MPa) 7.02 8.37

To calculate the maximum stress, it was necessary to find the shear

stress in the shaft. Using Equation A.4, the torque () experienced by

the motor shaft was calculated to be 1.55 Nm. Equation C.5 was used

to find the polar second moment of area () of the shaft- 2.03 X 10-9

m4.

=(2)4

32

(C.5)

Subsequently, Equation C.6 was applied to find the motor shafts

maximum shear stress () (Department of Mechanical Engineering,

2015).

=

(C.6)

The shear stress was calculated to be 4.57 MPa, and through

referencing Mohrs Circle, the actual combined maximum stress () in

the shaft was found using Equation C.7. represented the normal

stress in the x direction, and likewise for the y direction. Since in

this instance acts in the x-y plane, it was renamed .

= ( +

2)

2

+ 2

(C.7)

The value for was found to be 8.95 MPa, and through dividing the

fatigue strength of steel (180MPa) by (High Peak Steels Ltd., 2016),

a safety factor of 20 was obtained. By using the fatigue limit, this

would ensure that the scooter is durable. To take into account the

stress concentrations at the rounded corners of the keyways, the

stress concentration factor of a circle (3) was multiplied to , resulting

in a maximum predicted stress of 26.9 MPa, giving a safety factor of

6.7.

Appendix C.2: Idler Shaft The same procedure addressed above was performed for idler shaft,

with similar assumptions as those used with the exception that the

shaft was modelled as a simply supported beam. The pulley force

and the gear mesh transmitted force were both modelled as point

forces. Bending moment, shear force, deflection diagrams were

made for the 189 mm long idler shaft assuming a constant diameter

of 15 mm.

10

The results of all the calculations were written up into Table C.3 and

C.4.

Table C.3: Idler shaft- Vertical and horizontal values for parameters



(absolute) Shear force (N) 127 at x=0.028 285 at x=0.028

Bending moment (Nm)

5.61 at x=0.072 12.55 at x=0.072


(MPa) 16.9 37.9

Table C.1: Idler shaft- Calculated values for parameters required to obtain the safety factor

Parameter Value (absolute) (m4) 2.49 10

9

(m4) 4.97 109 Torque (Nm) 3.10

(MPa) 4.68 (MPa) 27.8

Safety Factor 6.5 As with the motor shaft, taking into account the presence of keyways,

the maximum predicted stress was determined to be 83.4 MPa, giving

an acceptable safety factor of 2.2.

Appendix C.3: Rear wheel shaft The same procedure was performed again with the same assumptions

as the idler shaft, except that the point loads would be from the pulley

belt tension and the weight of the person. The rear wheel shaft had a

length of 190.7 mm and for calculation purposes was assumed to have

a constant diameter of 20 mm. A static vertical load of 60g was

estimated assuming that half of the load on the scooter (person and

extra mass) would act on the rear wheel. The results of all of the calculations are shown in Table C.5 and C.6.

Table C.5: Rear wheel shaft- Vertical and horizontal values for parameters



(absolute) Shear force (N) 405 at x=0.026 280 at x=0.026

Bending moment (Nm)

23.6 at x=0.072 11.2 at x=0.066


(MPa) 30.05 14.3

Table C.6: Rear wheel shaft- Calculated values for parameters required to obtain the safety factor

Parameter Value (absolute) (m4) 7.85 10

9

(m4) 1.57 108 Torque (Nm) 9.31

(MPa) 5.93 (MPa) 23.0

Safety Factor 7.8 To take into account the presence of keyways, using the same

procedure as with the motor shaft, a maximum predicted stress of 69

MPa was obtained, giving an acceptable safety factor of 2.6.

Appendix D: Bearing Selection To choose the appropriate bearings, the required (dynamic load

rating) was calculated using Equation D.1 from the load on the bearing

(P), the life of the bearing (L million revolutions) and k, the bearing

constant (3 for deep groove bearings).

= (

106)

1

(D.1)

From the forces calculated in Appendix C and those obtained through

shear force diagrams (not shown) and applying static equilibrium,

forces on the bearings were found. Horizontal and vertical components

of the forces were combined to a total bearing load vector; these can

be found in Table D.1.

Using Equation D.2, the required life of the bearings () can be

calculated from the PDS requirement of a minimum of 100 hours of use

and the angular velocity of the shaft ().

=100 602

2

(D.2)

The life of the idler shaft was calculated to be 1.85 x 107 revolutions,

and the life of the rear shaft was calculated to be 6.16 x 106 revolutions.

The safety factors were also calculated for each bearing; bearings 1 and

2 were on the idler shaft, and the others on the rear wheel shaft.

Table D.1: Safety factor and loading of bearings

Bearing 1 Bearing 2 Bearing 3 Bearing 4 Bearing load (N)

310.22 190.03 492.62 352.71

required (kN)

0.820 0.502 0.903 0.647

of bearing

3.44 3.44 9.95 9.95

S.F. 4.19 6.85 11.02 15.39

Appendix E: Wooden Board To calculate the deflection on the board, a method similar to that used

for shaft deflections was applied. To simplify calculations, these

assumptions were made:

The vertical force acting on the wheels acted at the ends of

the board

The weight of the board was negligible compared to the 100

kg person on the board

The person on the board was modelled as a point mass acting

on the middle of the board

The Youngs modulus of the board was assumed to be 9.0

10 GPa (value for softwood)

By assuming the persons weight as a point mass, this would provide a

safe overestimate of the stresses in the board. Since the wooden board

had a rectangular cross sectional area, =1

123 was used to

calculate the second moment of area (Department of Mechanical

Engineering, 2015). The thickness of the board () was given to be

0.015 m and the width of board () was given to be 0.14 m. was

found to be3.94 108 m4. Using a simply supported model, the

bending moment, and deflection diagrams (similar to those obtained

for the shafts) for the 0.40 m long kick board were created.

Calculations similar to those performed for the shafts were done for

the kickboard, but only vertical loads were taken into account and

shear stress was ignored. In calculating the safety factors, since wood

does not have a fatigue limit, the yield strength of a wooden board

was used. The value was given to be 70 MPa (CES EduPack, 2016). The

results of the calculations are displayed in Table E.1.

Table E.1: Calculated parameters for the kickboard

Parameter Maximum value (absolute) Shear force (N) 491 at x=0 and x=0.2

Bending moment (Nm) 98.1 at x=0.2 Deflection (mm) 3.69 at x=0.2

Normal stress (MPa) 18.7 Safety Factor 3.8

Appendix F: Side Plates The free body diagram in Figure F.1 shows the forces included in FEA

of the plate. in the diagram is a force vector representing the force

exerted on the bolts by the aluminium block. Forces used originated

from the maximum bearing forces calculated previously, the values

for the static load were doubled to give a safe load estimate which

would account for dynamic loading and impact forces. The forces on

the motor shaft were ignored as they were relatively small.

Figure F.1: Free body diagram of side plate

Appendix G: Pulley selection Details on the selected pulleys are shown in Table G.1.

Table G.1: Selected pulleys and their associated dimensions

RS Catalogue Number 286-5663 286-5720

Material Aluminium Aluminium Belt width /mm 10 10 Pitch /mm 5 5 Teeth 20 60 Maximum bore dia. /mm 18 76 Outside diameter /mm 32 95.65 Hub diameter /mm 23 65 Having decided on the pulley choices, and through fixing a centre

distance (), the belt length () was calculated using Equation G.1.

In the equation, the pitch is represented by and the number of

teeth in the large and small pulleys was represented by and

respectively. Since there were specific belt sizes provided by the

catalogue, it was more appropriate to set an approximate centre

distance then adjust until it fitted. The minimum distance between

the centres would have to be 63.5mm due to the size of the pulleys

themselves. A centre distance of 123.4mm gave a belt length of

455mm which would perfectly fit the pulleys. The specifications for

the selected pulley belt are displayed in Table G.2.

2( + ) + 2 +

1

4[( )

]

2

(G.1)

Table G.2: Specification for selected pulley belt

RS Catalogue Number 474-5549 Belt teeth 91 Belt length /mm 455 Power rating /kW 5 Exact Centre Distance /mm 123.4

Using Equations G.2 and G.3, based on our chosen belt, the

teeth in mesh between the small pulley and the belt was

calculated. In the calculations, was the lap angle and

was the number of teeth in the belt. Our final pulley selection

allowed for a minimum of 8 teeth in mesh which was

adequate.

(1

) ( ) =

( )

( ) (G.2)

Number of teeth in mesh =

(G.3)

Appendix H: Brake

Using Figure H.1, the force required to stop the vehicle using the

brake was calculated. To decelerate the scooter 12kmh-1

(3.33 ms-1) to rest in 5 m whilst in use by a mass of 67.5 kg (), it

had to be verified that our brake could reliably produce the torque

required to achieve this. The deceleration () was calculated to be -

1.11 ms-2 using the equations of motion. The torque () required to

stop the wheel of radius () was calculated using Equation H.1 to be

-4.50 Nm.

= Force = (H.1) This calculation assumed R and P would act at a perpendicular

distance x from the bolts and for F, the tangential friction force to

act at a perpendicular distance y from the bolts. By taking moments

about the bolts and rearranging the expression, Equation H.2 was

formed, where is the coefficient of friction between the brake and

the wheel.

=

(H.2)

Using the first part of Equation H.2, with = , and rearranging to find P, Equation H.3 was formed.

=

(H.3)

P was calculated to be 143 N.

Figure H.1: Free Body Diagram of Brake

THIRD ANGLE PROJECTIONA

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3544 48

3236 49

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PART 37 REMOVED FOR CLARITYPART 31,33 NOT LABELED DUE TO VIEW OBSTRUCTION

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DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED

OMAR IMRANANTHONY DE SOUZAANTHONY DE SOUZA


Steel EN1A

0REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


4 x 4.50 THRU

50

24

A

A

3

ALL CHAMFERS 0.50 X 45

SECTION A-ASCALE 2 : 1


ANGULAR 1

IDLER HOUSING A

ES003DWG No.


TITLE:


MATERIAL:

DRAWN 06/12/16

DATENAME

CHECKEDAPPROVED

Y.Y.


Aluminium 6082 T6

REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


4 x 5.50 THRU

60

30

R23

3

ALL CHAMFERS 1 X 45


ANGULAR 1 Rear Bearing Housing A

ES-004DWG No.


TITLE:


MATERIAL:

29/11/16

DRAWN 03/11/16

DATENAME

CHECKEDAPPROVED

Yining YangY.Y.


Aluminium 6082 T6

REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


4 x 5.50 THRU

60

R23

B

B

30

+ 0.

021

0

26

36

- -0.0

070.

020

3

9

7

ALL CHAMFERS 1 X 45

SECTION B-BSCALE 2 : 1


ANGULAR 1 REAR BEARING HOUSING B

ES-005DWG No.


TITLE:


MATERIAL:

06/12/16

DRAWN 03/11/16

DATENAME

CHECKEDAPPROVED

Yining YangY.Y.


ALUMINIUM 6082 T6

REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


15.5

45.7

259.9

312.9

372.1

4 x 9 THRU ALL

8 THRU

140

3 x 6.60 THRU ALL

CSK 12.60 x 90 3

5 40

100

12.

5

127

.50

70

70 105

15

400


ANGULAR 1

BOARD

ES-006DWG No.


TITLE:


MATERIAL:

12/12/2016

DRAWN 03/11/16

DATENAME

CHECKEDAPPROVED

Yining YangWill Hey


BIRCH PLYWOOD

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


17.

3 + 0

.1 0

5 -00.03

15 H7

+ 0.0180


ANGULAR 1

40 TOOTH SPUR GEAR

ES-007DWG No.


TITLE:


MATERIAL:

10/12/201603/11/2016

DRAWN 03/11/2016

DATENAME

CHECKEDAPPROVED

LAI SZE WAIOMAR IMRANANTHONY DE SOUZA


STEEL EN8 (080M40)

0REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


50

4 x 4.50 THRU

R21

A

A

24

- -0.0

070.

021

28

+ 0.

021

0

7

3 10

ALL CHAMF

ERS 0.50 X 4

5



ANGULAR 1

IDLER HOUSING B

ES008DWG No.


TITLE:


MATERIAL:

10/10/2016

DRAWN 06/12/16

DATENAME

CHECKEDAPPROVED

Y.Y.SZE LAI


ALUMINIUM 6082 T6

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


25 1

2.50

4 x 9 THRU ALL

20 50 50 50 20

190

25 1

2.50

97 53

2 x 9 THRU ALL

25

25


ANGULAR 1

MOUNTING BLOCK

ES-009DWG No.


TITLE:


MATERIAL:

10/12/20163/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED

ANTHONY DE SOUZAOMAR IMRANANTHONY DE SOUZA


Aluminium 6082 T6

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


2

120

180

22

DOWN 90 R 10

UP 90 R 10

87

136

86

R10

3 x 9 THRU


ANGULAR 1

BRAKE

ES-010DWG No.


TITLE:


MATERIAL:

12/12/201605.12.16

DRAWN 05.11.16

DATENAME

CHECKEDAPPROVED

OMAR IMRANWILLIAM HEYANTHONY DE SOUZA


Aluminium 8021 T6

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


6 +-0.0150.015

12.

8 + 0

.3 0

20 H7

+ 0.0210

A

A



ANGULAR 1

60 TOOTH PULLEY

ES-011DWG No.


TITLE:


MATERIAL:

10/12/20164/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED



Aluminium 6082 T6

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


5 +-0.0150.015

10

+ 0.1

00

15 H7

+ 0.0180

A

ASECTION A-ASCALE 2 : 1


ANGULAR 1

20 TOOTH PULLEY

ES-012DWG No.


TITLE:


MATERIAL:

10/12/20163/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED



Aluminium 6082 T6

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


4 x 6.60 THRU

60

R23

A

A

30

+ 0.

021

0

36

- -0.0

090.

025

3

12

All Chamfers 1 X 45



ANGULAR 1 Rear Bearing Housing C

ES-013DWG No.


TITLE:


MATERIAL:

06/12/16

DRAWN 03/11/16

DATENAME

CHECKEDAPPROVED

Yining YangY.Y.


Aluminium 6082 T6

REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


4 x 4.50 THRU

50

R21

A

A

10

34

- -0.0

090.

025

28

+ 0.

021

0

3 All Chamfers 0.50 X 45



ANGULAR 1

IDLER HOUSING C

ES014DWG No.


TITLE:


MATERIAL:

DRAWN 06/12/16

DATENAME

CHECKEDAPPROVED

Y.Y.


Aluminium 6082 T6

REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


2 X 5.50 4 X 6.60

24

55.50

90

21

21

105

4.1

0

10

10 4

.10

UP 90 R 1.5

UP 90 R 1.5

1

175

D

E

F

C

1 2 3 4

B

A

321 5

C

D

4 6 7 8

A

B


ANGULAR 1

TOP SHEET

ES-015DWG No.


TITLE:


MATERIAL:

10/12/201610/12/2016

DRAWN 04/11/2016

DATENAME

CHECKEDAPPROVED

WILLIAM HEYANTHONY DE SOUZAANTHONY DE SOUZA


STEEL EN1A

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES

E


26.5 15 H7

+ 0.0180

21


ANGULAR 1

26.5MM IDLER SPACER

ES-016DWG No.


TITLE:


MATERIAL:

10/12/201605.12.16

DRAWN 05.11.16

DATENAME

CHECKEDAPPROVED

LAI SZE WAIH KANEANTHONY DE SOUZA


ALUMINIUM 6082 T6

2REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


35.5 15 H7

+ 0.0180

21


ANGULAR 1

35.5MM IDLER SPACER

ES-017DWG No.


TITLE:


MATERIAL:

10/12/201605.12.16

DRAWN 05.11.16

DATENAME

CHECKEDAPPROVED



ALUMINIUM 6082 T6

2REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


8 15 H7

+ 0.0180

21


ANGULAR 1

8MM IDLER SPACER

ES-018DWG No.


TITLE:


MATERIAL:

10/12/201605.12.16

DRAWN 05.11.16

DATENAME

CHECKEDAPPROVED



ALUMINIUM 6082 T6

2REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


23.0 15 H7

+ 0.0180

21


ANGULAR 1

23MM IDLER SPACER

ES-019DWG No.


TITLE:


MATERIAL:

10/12/201605.12.16

DRAWN 05.11.16

DATENAME

CHECKEDAPPROVED



ALUMINIUM 6082 T6

2REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


5 12 H7

+ 0.0180

16


ANGULAR 1

5MM MOTOR SPACER

ES-020DWG No.


TITLE:


MATERIAL:

12/12/201605.12.16

DRAWN 05.11.16

DATENAME

CHECKEDAPPROVED



ALUMINIUM 6082 T6

2REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


367.00

10.

00

94.91 136.91

6.60 THRU x2

5 THRU x4

7 THRU x3

9 THRU x4

15.00

52.19 50.00

81.19

25.00 75.00

125.00 175.00 226.16 272.16

249.16

24.

00

32.

04

10.

00

15.

94

44.

94

73.

94

90.

44

9.0

4

110.19

110

.00

8 THRU

34 THRU

26 THRU

111

.44

69.

44

21.85

42.85

63.85

4.50 THRU


ANGULAR 1

SIDE PLATE

ES-021DWG No.


TITLE:


MATERIAL:

12/12/20164/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED

CALEB GODDARDANTHONY DE SOUZAANTHONY DE SOUZA


MILD STEEL EN1A

0REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


4 +-0.0150.015

12 H7

+ 0.0180

7.8

0 + 0

.1 0

A

A

20.00 10.00



ANGULAR 1

20 TOOTH SPUR GEAR

ES-022DWG No.


TITLE:


MATERIAL:

4/11/20164/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED



ALUMINIUM 6082 T6

1REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


5 -0 0

.03

16.0


5 -0 0

.03


ANGULAR 1

5x5x16 KEY

ES-024DWG No.


TITLE:


MATERIAL:

10/12/20164/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED

LAI SZE WAIOMAR IMRANANTHONY DE SOUZA


MILD STEEL EN1A

0REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


25.0

5 -0 0

.03


5 -0 0

.03


ANGULAR 1

5x5x25 KEY

ES-025DWG No.


TITLE:


MATERIAL:

10/12/20164/11/2016

DRAWN 3/11/2016

DATENAME

CHECKEDAPPROVED

MICHELLE LAIOMAR IMRANANTHONY DE SOUZA


MILD STEEL EN1A

0REVISION


X = 0.5X.X = 0.1X.XX = 0.02

TOLERANCES


22 13

R2.50

199

61.5

20

30 + 0.100

179

R2.50

15

- -0.0

060.

017

15

+ -0.

004

0.00

4 G

G

H

IDENTICAL KEYWAY CROSS SECTION FOR BOTH KEYWAYS

BOTH ENDS IDENTICAL

.005 A

A 5

-00.03

12.

5 -0 0

.10

SECTION G-G

M10

R1

ALL CHAMFERS 0.50 x 45

DETAIL HSCALE 2 : 1


ANGULAR 1

IDLER SHAFT

ES-026DWG No.


TITLE:


MATERIAL:

10/12/201606/12/2016

DRAWN 03/11/2016

DATENAME

CHECKEDAPPROVED

HUGH KANELAI SZE WAIANTHONY DE S

escooter project

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