American Society of Mechanical Engineers Human …web.cecs.pdx.edu/~far/me492/PDS document/Y2008... · [Type text] [Type text] [Type text] American Society of Mechanical Engineers

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American Society of Mechanical Engineers

Human Powered Vehicle Team

ME 493 Final Report - Year 2008

6/4/2008

Design Team:

Ben Bolen

Erik Chamberlain

Kenneth Lou

Levi Patton

Bryan Voytilla

Academic Advisor:

Derek Tretheway

Industry Advisor:

Faryar Etesami

Sponsor:

American Society of Mechanical Engineers

i

Executive Summary

The Portland State Human Power Vehicle design team designed and constructed a fully faired

aluminum bicycle for use by the Portland State Human Powered Vehicle Race Team in the

ASME Western Regional HPV Challenge named Vike Bike. Vike Bike is a recumbent, rear wheel

drive bicycle covered by a full fairing.

Frame geometry was established based on previous years HPV design team research into

optimal power angles for the human body while riding a bicycle. The configuration is a rear

wheel drive, recumbent bicycle with front wheel steering, a fixed seat, and an adjustable boom.

Power is delivered to the rear wheel via a custom made chain ring attached to the adjustable

boom; the chain is routed underneath the frame by way of three idler gears. An integrated roll

bar and a five point seat belt provide safety protection for the rider. 6061-T6 aluminum of

varying size and wall thickness was used to construct the frame.

The fairing was designed using Nation Advisory Committee for Aeronautics shapes optimized

for pressure recovery. The fairing was constructed with three layers of S2 fiberglass and either

a Baltek Mat or balsa wood core.

Materials testing and computer analysis were performed on the frame and fairing to optimize

their design for criteria such as weight or strength. Explanations of these analyses are included.

The Portland State HPV race team continued its tradition by finishing third overall in the

Western Region HPV Challenge for the third year in a row. The race served as a testing ground

for vehicle design and survivability; lessons were learned in systems performance as well as the

overall quality of design which can be implemented by future Portland State HPV design teams.

Table of Contents

Executive Summary .......................................................................................................................... i

Introduction and Background information ..................................................................................... 2

Mission Statement .......................................................................................................................... 2

Main design requirements .............................................................................................................. 2

Top Level Design alternatives ......................................................................................................... 3

Design Concepts .......................................................................................................................... 3

Frame........................................................................................................................................... 4

Fairing .......................................................................................................................................... 4

Mechanical Systems .................................................................................................................... 5

Final Design and Evaluations .......................................................................................................... 6

Frame........................................................................................................................................... 6

Fairing .......................................................................................................................................... 9

Mechanical Systems .................................................................................................................. 13

Future Design Considerations ....................................................................................................... 15

Frame......................................................................................................................................... 15

Fairing ........................................................................................................................................ 15

Mechanical Systems .................................................................................................................. 16

Conclusion ..................................................................................................................................... 18

Appendix A: Safety ........................................................................................................................ 19

Appendix B: Maintenance ............................................................................................................. 20

Appendix C: Initial Design Concepts ............................................................................................. 21

Appendix D: Product Design Specification.................................................................................... 26

Appendix E: Frame Analysis .......................................................................................................... 30

Appendix F: Biomechanical Testing .............................................................................................. 46

Appendix G: Braking Analysis ........................................................................................................ 48

Appendix H: Top Speed Analysis ................................................................................................... 49

Appendix I: Turning Analysis ......................................................................................................... 53

Appendix J: Strain Gauge Testing.................................................................................................. 57

Appendix K: Fairing Aerodynamic Analysis ................................................................................... 60

Appendix L: Fairing Material Analysis ........................................................................................... 64

Appendix R: Vehicle Stability ........................................................................................................ 67

Appendix M: Frame Design Drawing ............................................................................................ 69

Appendix N: Fairing Design Drawings ........................................................................................... 81

Appendix O: Vike Bike Dimensions ............................................................................................... 85

Appendix P: Glossary of Human Powered Vehicle Terminology .................................................. 86

Appendix Q: Bill of Materials ........................................................................................................ 87

References .................................................................................................................................... 93

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Introduction and Background information

Because of increasing energy prices and growing concern over vehicle pollution the American

Society of Mechanical Engineers (ASME) created the Human Powered Vehicle (HPV) Challenge

to encourage development in human powered technology. The goal of the HPV Challenge is

that someday a HPV will be designed that is practical enough for everyday uses such as going to

the store or commuting to work.

The ASME HPV Challenge is a competition in which engineering students from around the

country design, construct, and race an HPV. An HPV can take many forms and varying rider

positions, such as upright, recumbent, or prone and can have any number of wheels. The

competition consists of three separate events: a 100m sprint race, a 65km grand prix style

endurance race, and a judging process for the vehicle’s design, safety, and formal presentation.

For the third consecutive year the Portland State University (PSU) chapter of ASME competed in

the HPV Challenge. Having placed third the two previous years the team sought to improve

upon those finishes and win the overall competition. A design team of five mechanical

engineering seniors was formed to design and construct an HPV for use in this competition.

Mission Statement

The goal of this project was to develop an innovative, light-weight, and aerodynamic HPV for

the Portland State HPV Race Team to win the overall ASME Western Region HPV Competition

on April 18th, 2008, and to complete it within time and budgetary constraints.

Main design requirements

During the initial portion of the design phase external and internal searches were conducted

into HPV design. The goals given in Table 1 were the preliminary design goals for the project.

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Metrics were developed to quantify the targets. The outcome of the goal is given in the right

hand column of the table.

Table 1: Initial design goals

Evaluation of Design Goals

Metric Target Produced Target Met?

Top Speed >45 mph 39 mph No

Stopping Distance Less than or equal to 20ft from 15 mph 10 ft Yes

Frame Factor of Safety > 1.5 Competition research Yes

Fairing Strength Greater than or equal to 2007 PSU HPV fairing Competition research No

Crash Recovery Time < 15 seconds 11 seconds Yes

Turning Radius < 25 ft Competition research Yes

High Speed Stability Does not wobble uncontrollably above 20 Mph

Does not wobble uncontrollably above 20 Mph Yes

Stright line aerodynamic efficiency Cd less than or equal to 0.14 0.04 Yes

Partical fairing removal for rider entry and exit <60 seconds 20 seconds Yes

Durration of HPV life in service Greater than April 2008 Still functioning after April 2008 Yes

Visual Appeal 30 Points 7 points No

Use of industry standard bike tools 100% 100% Yes

Total vehicle weight less than 50-lb Total weight 45-lb Yes

Rider Visibility Horizontal > 150 degrees Vertical > 60 degrees

Horizontal > 150 degrees Vertical > 60 degrees Yes

Fairing Strength Modulus of elasticity >= Vike Trike II < Vike Trike II No

Rider Comfort Low fatigue on rider Low fatigue on rider Yes

Rider Exchange time < 60 seconds 30 seconds Yes

Overall HPVC Finish First Place Third No

Under Budget $11,004.75 $12,725.41 No

Top Level Design alternatives

Design Concepts

During the design process initial conceptual sketches were made of potential HPV

configurations. These design ideas consisted of various vehicle, drive train, and fairing

configurations as well as different rider positions. These initial design concepts can be seen in

Appendix C.

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Frame

A top-level conceptual design required a decision on whether the vehicle would be configured

as a bicycle or a tricycle. The performance requirements of top speed, stability, weight

reduction, crash recovery, and straight-line aerodynamic efficiency are affected by the design

decision. Differences in amounts and types of materials used can affect the cost requirement.

Bicycle and tricycle configurations were evaluated using pros and cons (see Table 2).

Table 2: List of pros and cons for bicycle and tricycle configurations. Bike Trike

Pros Cons Pros Cons

Less components means lower weight and cost

Low-speed instability Low-speed stability High-speed instability

Small frontal area Tolerancing (front and back wheel linearity)

Team has experience working on tricycle HPVs

Scrub of the wheels while turning

Low weight Rider change out requires pit crew assistance

Uncomplicated chain routing Steering design and use

Statistically proven to do well in a competition.

Crash recovery is complicated in a fully enclosed fairing

Rolling resistance is increased with third tire

Drive train has to be routed to avoid the front tire

Aerodynamic resistance is increased with hole in the fairing for the third wheel

The cons to the tricycle configuration are inherent to the tricycle design. Cons to a bicycle

configuration can be overcome with engineering. The pro tricycle experience can also be used

to help the team with building a bike. The small frontal area and low weight of a bicycle would

be difficult (or impossible) to transfer to a tricycle setup. Therefore, the decision was made to

go with the bicycle configuration.

Fairing

Several concepts were developed during initial design for the fairing, a partial frontal fairing and

a full fairing. In general, the configuration of a fairing is determined by the end use of the

vehicle. If the vehicle is intended to be used at high speeds then a full fairing is beneficial

because of the reduced drag forces, however if the vehicle spends the majority of its time at

low speeds a partial fairing is preferred due to its low weight. Previous years experience

identified high speed efficiency as critical in both the sprint and endurance competitions.

Taking these and other attributes into consideration, a table that addresses the pros and cons

of each configuration was created.

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Table 3: A comparison of three possible fairing configurations Symmetrical Full Fairing Non-Symmetrical Full Fairing Partial Fairing

Pros Cons Pros Cons Pros Cons

Moderate manufacturing cost

High weight when compared to the partial fairing

Best reduction of drag force

High manufacturing cost

No manufacturing cost

Low reduction of drag force

Moderate manufacturing time

Poor crash recovery High manufacturing time

No manufacturing time

High materials cost High materials cost Low materials cost

Reduction in drag force is nearly as good as the non-symmetrical fairing

Good visibility must be designed

High weight when compared to the partial fairing

Low weight

Ventilation must be designed

Poor crash recovery Best crash recovery

Good visibility must be designed

Good rider visibility is inherent

Ventilation must be designed

Good rider ventilation is inherent

Though the partial fairing has the most pros, they do not compensate for its deficiency in drag

force reduction. The non-symmetrical fairing potentially has the highest reduction in drag force

but the time and cost associated with such a design is prohibitive for the 2008 PSU HPV team.

Given these constraints the symmetrical full fairing is the chosen design.

The symmetrical design reduces manufacturing time and cost by requiring fewer positive and

negative molds to be produced. Only a single negative mold is needed to produce both the top

and bottom half of the fairing since the intended shapes are identical. Keeping this

manufacturing process in mind, the shape of the fairing can be designed conforming to the set

of design requirements outlined in Table 1:

Reduced drag force when compared to previous PSU fairings

Complete coverage of the bike frame and rider (full fairing)

Accommodations for rider ergonomics

Mechanical Systems

The configuration of mechanical systems on the HPV was sought to optimize ease of installation

as well as overall performance and durability. Three mechanical sub systems on the HPV that

were examined for alternative designs are brakes, drive train, and steering.

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Several types of Front Wheel Drive (FWD) drive trains where examined, including: a swinging

boom, a u-joint, and a twisting chain. Each type of FWD uses a crank and gear set on the front

of the vehicle and a cassette on the front wheel of the vehicle. A swinging boom uses a boom

that is fixed to the fork of the vehicle, so that the boom turns when the front wheel turns. For a

u-joint drivetrain, power is transmitted from a boom (which is fixed to the frame) to the front

wheel. The twisting chain setup uses a fixed boom with a turning wheel, where the chain derails

slightly when the vehicle is turning. Table 4 is a list of pros and cons between choosing either a

FWD or RWD system.

Table 4: Drive Train Configuration Pros and Cons FWD RWD

Pros Cons Pros Cons

Small chain line, high

efficiency

Limited wheel size and

gearing constraints

Ability to use large wheel while

keeping center of gravity low

Long chain line, low efficiency

Easy chain routing Chain twisting will cause pre-

mature ware of chain.

Conventional configurations

and proper use of parts

Complicated chain routing to

rear wheel

Complicated configuration

provides expensive parts

Use of commercially available

parts

Final Design and Evaluations

Frame

The Vike Bike frame design is based off of the similar 3D space frame that was used for the Vike

Trike II with a few exceptions. The Vike Bike used the same body configuration angel (BCA) as

the Vike Trike II to achieve maximum possible rider force output while maintaining low rider

fatigue. The experiment performed by Reiser, of Colorado State University (2001), indicates

that a BCA between 130° and 140° is most efficient (see Figure 1).

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Figure 1: Rider position study for maximum recumbent power input (Reiser, 2001)

Additional literature review (Landwer and Too 2003; Too 1990; Too 1991; Too 1994; Too 1995)

was used to set rider configuration angles and frame sizes. The geometry for the 2008 Vike Bike

was modified slightly during the analysis to simplify manufacturing, provide stress relief, and

add stiffness. Figure 2 shows the current Vike Bike riding configuration.

Figure 2: Vike Bike frame configuration geometry

The existing design of the Vike Trike II was proved to be effective and efficient. This provided

the PSU HPV design team with a conceptual model to base the design of the Vike Bike from.

The greatest change in design of the frame was to configure the Vike Trike II (Figure 3) into the

configuration of a recumbent Bike while maintaining low rider position of 11.0” and long wheel

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base for recumbent bicycles of 54.95” for stability. Using the adjustable boom rather than an

adjustable seat proved to be the most efficient way to accommodate the difference in rider

height for the 2008 Vike Bike race team.

Figure 3: 2007 PSU Vike Trike II frame, Isometric view

The Vike Bike is manufactured of aluminum 6061-T6, providing the strength and light weight

desired by the PSU HPV Vike Bike design team. Figure 4 presents the 2008 PSU Vike Bike.

Similar to the Vike Trike II the Vike Bike’s roll bar is welded directly to the main tube to provide

stiffness. The frame height of the Vike Bike was raised to 8 ½ ” from the Vike Trike II’s 5 1/32”,

this was needed to fit a smaller 451 recumbent bicycle wheel and still maintain a short overall

rider height of 11” (see Figure 3). A glossary of Vike Bike components can be found in Appendix

A.

9

Figure 4: 2008 PSU Vike Bike frame, Isometric view

Fairing

Using experience gained from previous years of HPV competition, a full fairing was chosen for

both the sprint and endurance events. Though previous years have used a partial fairing for the

endurance event, it was evident that a fully faired vehicle could increase overall velocity and

provide better protection to the rider during a crash. With the symmetrical full fairing design

chosen, the planar shape of the fairing was the next design consideration. The 2007 HPV team

used a planar shape provided by the National Advisory Committee for Aeronautics (NACA). The

shape was chosen for its pressure recovery attributes as well as incorporating the geometry of

the Vike Trike I and II. Selection of the planar geometry for the 2008 Vike Bike was done by

using a 6-series NACA shape that would contribute to a reduced drag force while closely fitting

the shape of the 2008 Vike Bike. A comparison between the 2007 and 2008 planar shapes is

presented in Table 5.

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Table 5: A comparison of the Planar shapes for the 2006 and 2008 HPV fairings PSU 2007 Planar Shape PSU 2008 Planar Shape

Pros Cons Pros Cons

Good pressure recovery Frontal area is larger than

necessary for the 2008 bike

Better pressure recovery due

to increased laminar flow

Space for pedals and feet will

be reduced.

Plenty of room for pedals and

feet

Smaller frontal area

One of the reasons for increased pressure recovery of the 2008 fairing is because its maximum

width occurs further from its leading edge. This is readily apparent when comparing Fig. 5 and

Fig. 6. Note that both fairings have a nose to tail length of 106”.

Figure 5: PSU 2007 Planar shape

Figure 6: PSU 2008 planar view

The leading 2 ft of the side view contour (see Figure 7) is determined by a NACA 4-series shape,

after that the geometry is governed by rider dimensions, the rollbar and toe box. The toe box

highlighted in pink in Figures 6 and 7 contribute to the fairings unique 3-dimentional shape.

The selected planar shape can only be used by keeping the faring walls nearly vertical from the

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nose to the beginning of the toe box. This results in the distinctive ridges that begin at the nose

of the fairing then blend into the rest of the body.

Figure 7: PSU 2008 side view

Rider visibility was designed concurrently with fairing shape since these features are

intertwined. Both window cutouts and a bubble canopy design were possible solutions for

rider visibility. Examples of these designs can be seen in Figure 8 and Table 5 lists pros and cons

of a bubble canopy versus window cutouts.

Figure 8: Bubble canopy (left), Window Cutouts (right)

Table 6: Pros and cons list of fairing shapes

Bubble Canopy Window Cutouts

Pros Cons Pros Cons

Full visibility with no

obstructions

Often times increases the

coefficient of drag

Reduced drag Visibility obstructions

Requires a single head

location for optimum visibility

Location of riders head inside

the fairing can be variable

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Tomai (1999, 195) discusses that the addition of a bubble canopy to a streamlined shape can

increase the coefficient of drag anywhere from 0.045 to 0.15. Beyond the aerodynamic

concerns, the bubble canopies advantage of an unrestricted view is overshadowed by the

necessity to engineer an adjustable seat to maintain a constant head location near the canopy.

Hence a fairing with window cutouts was chosen due to its streamlined shape and innate

accommodation of rider ergonomics. The window locations shown in Figure 8 (right) were

chosen to maximize rider visibility, fairing strength, and aesthetics this allows the rider to see

158o out of 180o which is an increase over previous PSU HPV fairings (see Figure 9).

Figure 9: Rider visibility provided by window cutouts

During the design of the fairing it was found that a completely symmetrical design was not

feasible due to bike and rider geometry. A simple solution was found that results in the nearly

symmetrical geometry, shown in Fig. 7, while still using a single mold system. The flat bottom of

the fairing is achieved through a simple and reversible modification of the mold. Figure 10

shows how this modification is done; inserting a “restrictor plate” changes the inner dimensions

of the mold.

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Figure 10: View of mold and plate

The plate is made of ¾” fiberboard that is painted and waxed so the surface matches that of the

mold. This simple modification requires little in manufacturing time and cost while retaining

the benefits of a single negative mold system.

A major deficiency of previous PSU HPV fairings has been the lack of a fairing latching system.

This year a system was designed to facilitate quick entrance and exit of the rider and keep the

fairing halves together in the event of a roll over. The Vike Bike latching system (see Figure 11)

consists of two wedges that, when mated at the fairing seam, restrict all horizontal movement.

Vertical movement is restricted by a Velcro strap. Eight of these latching points will be located

on the fairing. However, the four furthest mounts from the rider will not include Velcro.

Figure 11: Model of the fairing latching system

Mechanical Systems

Braking on the HPV is provided by a mechanical disk in the front and a cantilever pinch brake in

the rear. A mechanical disk brake was selected for the front wheel for its superior braking

Restrictor Plate

location →

Restrictor Plate

↓

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power and heat dissipation. A cantilever rear brake was required based on the use of a full

carbon wheel in the rear which necessitates a disk brake cannot be used.

External research was performed to determine the optimal body configuration angles (see

Appendix C). Rear wheel drive was selected because it provides increased turning ability and

high speed stability. Biomechanical testing was performed to find the power output available

from each rider (see Appendix F). A gear ratio analysis was performed with the biomechanical

data to find the required gear ratio to reach the target metric of 45mph (see Appendix F). The

final drive ratio required necessitated the manufacturing of 65 tooth and 75 tooth custom gears

because these sizes are not available commercially (Drawings available in Appendix M). The

donated gears were overbuilt, and holes were drilled into the gears to cut weight while keeping

manufacturing time down. FEA was used to optimize the amount of weight that could be drilled

out of the gears. A commercially available 10-speed Sram Force cassette and derailleur was

selected because of range of gears and light-weight (due to carbon fiber construction). A

thumb-shifter was mounted near the handlebars for ease of shifting. The final drive train

configuration consists of a custom made front sprocket on the front of the vehicle driving a

chain that is routed under the rider via three idlers to a ten speed cassette on the rear wheel.

The Vike Bike reached a top speed of 44mph under a high-cross wind situation. This value was

still close to the 45.28 mph top-speed calculated through the gearing analysis. The Sram Force

derailleur is a short-throw derailleur which performs well through a fixed range of gears and

chain length but was found to be very sensitive to changing boom length or crashes. After the

HPVC races, rider surveys were performed to compare the 65 and 75 tooth chain ring to the 53

tooth chain ring on the Vike Trike II. The 65 tooth chain ring was said to have provided a good

range of gear ratios in the endurance race and female sprint race, and the 75 tooth chain ring

supplied the highest PSU HPV speed in the mail sprint race.

The steering system for the HPV was designed based on the HPVC required 25 foot turning

radius as well as the goal of high speed stability. Two options for bicycle steering configurations

are front wheel or rear wheel steering. Front wheel steering was selected for its high speed

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stability and ease of use. A detailed analysis of the HPV steering system can be seen in

Appendix H.

Future Design Considerations

Frame

The Vike Bike was designed with a higher accuracy of forces exerted by the rider to the frame

compared to the Vike Trike II by using the biomechanical testing and incorporating the results

from the ASME roll-bar criteria. Previous years have also obtained dynamic rider data while

riding the bike is in a stationary trainer and collecting strain gauge data to apply to the frame

design for accurate modeling. David Van Dyke, the ASME Senior section Chair for the 2008-

2009 year, offered to donate a mobile dynamic strain gauge device that would allow the design

team to record dynamic road data for the future design of the vehicle. Acquiring these forces

will allow the future design teams to design and fabricate a lighter and stronger vehicle.

Other future design possibilities include a composite frame. Composite frames prove to be

stronger and lighter than a conventional metal vehicle frame. Extensive testing involving

materials selection is required for the development of a composite frame and is seen as a

drawback. Other testing drawbacks for using a composite frame include testing the adhesive

capabilities of bonding a composite to a metallic object.

Fairing

Although the balsa wood composite used for the 2008 Vike Bike fairing is stronger, lighter, and

more ductile than the 2007 Vike Trike II Baltek Mat composite, it was extremely difficult to

shape and form the balsa wood to the contours of the fairing mold. This resulted in voids and

gaps between the balsa wood sheets creating many weak points throughout the fairing

structure. Future designs should consider using materials that are easier to form into mold

contours.

In addition to material complications, the vacuum bagging layup process also resulted in weak

points throughout the fairing. Sufficient vacuum could not be produced due to leaks in the

16

equipment and mold. As a result the material layers did not bond together leaving large voids in

the fairing. Future designs should consider using a vacuum bagging processes that can pull a

higher vacuum or leak proofing the mold and vacuum equipment.

Widow shaping and installation was a problem area for this year’s fairing. Acrylic was the

material used and is not optimal for this application. Another material should be sought for

future windows. Acrylic is a relatively brittle material that cannot withstand the impact of a

vehicle crash. During the competition weekend three windows were shattered and the sharp

edges could have done physical harm to the vehicle driver.

The acrylic windows are difficult to form without the aid of a dedicated thermoforming

machine. The windows must be cut and bent to shape in order to correctly fit the fairing and

blend neatly with the surface. The method used to shaped the windows for this year and

previous years has been to warm a single acrylic window with multiple heat guns then bend it

to shape. The heat guns do not uniformly heat the acrylic so localized warping occurs as the

windows are bent into place.

An alternative window would be to use thin sheets of lexan or mylar. These sheets a can easily

be cut to shape with scissors and then taped to the fairing. These windows are thin and will

most likely be destroyed in a crash. However, they are cheap (less than a dollar a window) and

are easily manufactured. Used as an expendable part, these would decrease manufacturing

time, cost and the end weight of the fairing.

Mechanical Systems

During the safety inspection, it was noted by the judges that the safety harness was not of

automotive quality because the use of plastic strap adjustments. In the future, the safety

harnesses should be made of high quality metal components and high strength straps.

The HPV crashed on the final lap of the endurance race, and an object was hit that forced

plastically deformed the 65 tooth chain ring. If a stress-strain curve is obtained for the type of

steel used in the chain ring, then FEA can be used to reverse engineer and find the amount of

force from a crash. This determined force can be useful when designing next year’s HPV. Using

17

an intermittent shaft with a gear reduction can provide the same gear ratios as a 65 or 75 tooth

chain ring. This could eliminate the need for a larger gear, lighten up the main chain ring, and

allow the use of commercial available gears (where commercially available gears can be

replaced easier and faster than custom made gears).

Due to the problems encountered in using a 10-speed road bike derailleur and cassette, it is

recommended that a future design should incorporate a 7-speed mountain bike derailleur and

cassette. Seven speeds should be used because they can absorb more damage before it starts

to effect shifting, and it is expected that future HPVs will crash during testing and racing

situations. Mountain bike derailleurs have a larger throw than road bike derailleurs, which

could help alleviate shifting problems on the Vike Bike.

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Conclusion

This year marked another successful year for the Portland State HPV design team. The team

researched, designed, and constructed a novel HPV that was used by the Portland State HPV

race team to finish third place in the ASME Western Region Challenge. While not all design

goals were met, the majority of high priority goals were either met or nearly met. Innovations

were made in frame and faring design, as well as construction methods for the frame and

fairing.

Future design concepts were developed during the tenure of this project. A more effective

fairing attachment system would prove invaluable during the sprint and endurance

competitions. Also lessons have been learned with fairing construction as to core material

selection and vacuum bagging processes. Low speed stability also appeared as an issue during

the endurance race.

The design team recommends a fairing attachment system that uses magnets to retain the top

section of the fairing as well as only having as much of the top fairing removable as necessary

to maximize fairing stiffness. Large scale mock ups of fairing construction should be done

before fairing construction, which include curved surfaces to ensure that the core material and

manufacturing process are suitable for the fairing construction. Research should be done into

recumbent bicycle design in effort to find improvements on the bicycles design for low speed

stability.

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Appendix A: Safety

The following safety guidelines must be followed for safe operation of the HPV. Failure to

comply with guidelines may result in serious injury or death.

Always wear well fitting approved CPSC or Snell helmet while operating HPV.

Do not wear loose fitting garments while ridding the HPV, as well as making sure that

pant legs are tight ensuring they will not be pulled into the drive train.

Keep all body parts away from HPV drive train while in motion.

Do not touch the ground until the HPV has slowed to a stop.

Safety belt must be worn while the HPV is in operation.

Perform all regular maintenance on the HPV to ensure proper operation.

Do not ride the HPV at speeds above the rider’s ability.

Follow all traffic laws while operating the HPV on public roads.

Do not operate the HPV under the influence of alcohol, controlled substances, or

medication that way cause downiness.

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Appendix B: Maintenance

Regular maintenance of the HPV is required to keep the HPV in proper running condition. The

following is a schedule for preventative maintenance. In the event repairs are needed between

regular maintenance intervals they should be performed to keep the vehicle running safely.

Prior to every ride check

That the tires are properly inflated

The operation of brakes and brake cables

The tightness of the crank set After every ride check

Wheels for leaks or sharp debris picked up by tires during the ride

The trueness of wheels

Clean the bikes drive train and other mechanical parts if they have become dirty Every two weeks

Lubricate the chain Once a month

Clean the entire bike including the drive train.

Check the drive train for wear; looking for chain link tightness and repairing as necessary.

Lubricate all brake cables and derailers.

Check for looseness in the handlebar bolts, boom clamping mechanism, crank bolts, deraler bots, and the brake mounting bolts.

Replace tires as needed. Every Three months

Inspect the frame for cracking or signs of wear. Every six months

Inspect and adjust as needed headset, hubs, pedals, and bottom bracket bearings. Annually

Completely disassemble the bike replacing all bearings, breaks, and shift cables.

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Appendix C: Initial Design Concepts

In the initial design stage of the Vike Bike I each team member brainstormed for their ideal

HPV. Five concepts were developed varying from a two wheeled upright bicycle to a virtual

HPV.

Concept 1: Fully Faired Recumbent Bicycle

One integrated concept is to use a short wheelbase, recumbent, bike, with rear wheel drive, an

adjustable boom, a fixed seat, and a fully enclosed torpedo-shaped fairing with landing gear

that will be used in the endurance and sprint races (see Figure 1, below).

Figure C1: Sketch of Concept 1

Concept 2: Partial Front Fairing Upright Bicycle

This concept is based off of the common upright diamond framed bicycle sold in many bicycle

shops. Concept 2 uses front wheel steering and rear wheel drive, with the rider seated in the

upright position in an adjustable seat which compensates for different rider sizes (see Figure 2,

below). To satisfy the ASME competition rules the concept will utilize a partial fairing for both

endurances and sprint races.

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Figure C2: Sketch of Concept 2

Concept 3: Fully Faired Recumbent Tadpole Tricycle

Another concept would be a continuation of the Vike Trike II HPV with improved fairing design,

high speed stability, and drive train efficiency (see Figure 3, below). It should have three wheels

and be rear wheel drive. The seat should be fixed and the boom should be adjustable. The

focus will be on refining existing sub-systems rather than innovating new solutions.

Figure C3: Sketch of concept 3

Concept 4: Reverse Driving Fully Faired Tadpole Tricycle

The fourth concept integrates a fully-faired, recumbent, tadpole trike with RWD (see Figure C4

and Figure C5, below). Visibility is achieved via an integrated video system. The rider is

positioned with their head at the front of the vehicle to reduce chain routing and frontal area.

The riding position is not prone, merely backwards. Accommodation for different size riders is

done with an adjustable seat.

23

Figure C4: Top view of concept 4

Figure C5: Side view of concept 4

Concept 5: Fully Faired Recumbent Bicycle with Bubble Canopy

Concept 5 integrates Long wheelbase, two sided FWD, pedals turn with front wheel, recumbent

bicycle, large gears, adjustable boom, low center of gravity, carbon frame, fully faired for sprint

and endurance race (see Figure C6 and Figure C7, below). The steering being attached to the

pedals will help aid in the spring allowing the rider to pull up on the handlebars as they push

down on the pedals simulating a sprint on an upright bicycle.

Figure C6: Drive Train for concept 5

← Direction of travel

Video camera ↓

← Video glasses

24

Figure C7: Side view of concept 5

Top-level final design evaluation and Selection

The concepts listed above were evaluated in a concept scoring matrix (see Table 2, below),

using the Vike Trike II as a datum. Scoring is from a 1 - 5 range, with scores meaning:

1 – Very inferior

2 – Inferior

3 – Acceptable

4 – Superior

5 – Much superior

Table C1: Concept scoring matrix Datum Concept 1 Concept 2 Concept 3 Concept 4 Concept 5

Performance

Top speed 3 5 1 4 4 5

Frame Strength 3 3 1 3 3 3

Crash recovery 3 2 5 3 1 2

High-speed stability 3 5 5 4 2 2

Straight line

aerodynamic

efficiency

3 5 1 3 2 2

Partial fairing

removal for rider

entry and exit

3 4 5 3 3 4

Cost

25

Stay under budget 3 4 5 3 1 1

Life in Service

HPV to last through construction,

testing, and

competition

3 2 1 3 3 3

Rider Safety

Visibility 3 5 5 4 2 4

Fairing Strength 3 4 1 3 4 4

Total Score 30 39 30 33 25 30

Concept 1 is nine points above the datum, and six points above the nearest competitor. Each of

the ten main design requirements is given equal weight in this scoring matrix, so the minimum

possible score is ten points and a maximum possible score is 50 points.

Dividing the total score by the number of requirements gives an average score range of one

through five. This is convenient because the same scoring range listed above can be used.

Therefore, Concept 1 is very close to a superior rating. Likewise, Concepts 2, 3, and 5 are close

to an acceptable rating and Concept 4 is half way between poor and average.

The concept scoring matrix clearly indicates that team will be manufacturing Concept 1: the

short wheelbase, recumbent, bike, with rear wheel drive, an adjustable boom, a fixed seat, and

a fully enclosed torpedo-shaped fairing with landing gear to be used in the endurance and

sprint races. With a commitment to working on Concept 1, each of the product sub-functions

requires detailed design decisions beyond the top-level concepts.

26

Appendix D: Product Design Specification

Table D1: High priority product design specifications Table

Criteria

Requirements Primary

Customer

Measurement Metric Target Target Basis Verification Method

Performance Top speed PSU-HPV Team Top speed in male and

female sprint races

mph >45 mph Competition

research

Vehicle Time trial research

Braking ASME HPVC

Judges

Stopping distance at

15mph

Feet =< 20ft Competition rules Vehicle testing

Strength ASME HPVC

Judges

Frame safety factor Non-dimensional >1.5 Competition

research

Design analysis

Strength PSU-HPV Team Fairing strength Flexure modulus Greater than or

equal to 2007 PSU

HPV fairing

Competition

research

Materials testing

Crash recovery PSU-HPV Team Time to self-upright Seconds < 15s Competition

research

Vehicle testing

Turning radius ASME HPVC

Judges

Turning ability Radius in feet < 25ft HPVC rules Vehicle testing

High-speed stability PSU-HPV Team Vehicle does not wobble

uncontrollably at straight

line speeds > 20mph

Steering axis

rotation in

degrees

< 5deg Competition

research

Vehicle testing

27

Criteria Requirements Primary

Customer


Performance Straight line

aerodynamic efficiency

PSU-HPV Team Coefficient of drag Non-dimensional <=.14 Frontal area is an

improvement upon

Vike Trike II fairing

Theoretical verification with

CFD and achieved with wind

tunnel testing

Partial fairing removal

for rider entry and exit

PSU-HPV Team Rider change out time Seconds <60s Improve upon Vike

Trike II fairing

Time trial testing

Documentation Fulfill ME 492/493

Class Requirements

PSU-HPV Team Score on capstone

related reports

Grade A ME 492/493 class

syllabus

Inspection of class grade

Life In Service HPV needs to last

through construction,

testing, and HPV

Challenge.

PSU-HPV Team Life of bike Months > April 2008 Bike must last until

HPV Challenge is over

Inspection

Table D2: Medium priority product design specifications Criteria Requirements Primary

Customer


Aesthetics Visual appeal ASME HPVC

Judges

Frame appearance and

competition design

presentation

Points, subject to

judges

interpretation

30 points Competition rules Review points awarded at

competition

28

Criteria Requirements Primary

Customer


Performance Maintenance PSU-HPV Team Industry standard bike

tools

Common bike

tool sizes,

percent

= 100% Competition

research

Vehicle Time trial research

Maintenance PSU-HPV Team Ease of access # of parts to

remove to get to

desired part

<= 1 part Direct comparison

to standard

recumbent bikes

Solid modeling, vehicle

testing

Light weight PSU-HPV Team Vehicle assembly lbs < 50 lbs Improve upon

Vike Trike II fairing

and frame

Measurement with scale

Cost Stay under budget PSU-HPV Team Stay under budget with

material and fabrication

cost

Dollars > Budget Competition

research

Expenditure accounting

Safety Rider safety PSU-HPV Team Visibility Degrees of

vertical and

horizontal view

Horizontal > 150

degrees Vertical >

60 degrees

Rider preference Measurement

Fairing Strength Modulus of

elasticity

>= Vike Trike II Experience of

previous fairings’

adequate strength

Vehicle testing

29

Table D3: Low priority product design specifications Criteria Requirements Primary

Customer


Ergonomics Rider comfort PSU-HPV Team Comfortable

temperature

Deg F > 65 deg Competition research Vehicle testing

Ventilation Energy out, Watts Energy out =

Energy in

Competition research Vehicle testing

30

Appendix E: Frame Analysis

Summary Section

FEA was performed with I-DEAS 5 to analyze stress, strain, displacement, modal frequency, and

weight in the Vike Bike frame. Nine FEA optimizations were performed to find geometry and

commercially available tube sizes that would give the highest stiffness and lowest weight while

making sure to stay above the required safety factor of 1.5.

Loading conditions were modeled assuming worst case scenario loading and a simplified beam

mesh (see Figure E1), as described by Furniss and others (2007). The Vike Bike was modeled to

show compliance with ASME roll bar standards, no yield or fracture in a top loading of 485lbf or

a side loading of 260lbf (Interpretation of Rollover/side Protection Rules for 2008 HPVC

Competition, 2008). A materials selection analysis determined that Aluminum 6061 would be

used instead of 4140 steel or Titanium. The final optimization results are presented in Table E1

for stress (psi), factor of safety against material failure and yielding. Table E2 shows the final

optimization results for modal frequency (Hz) and weight (lbf). Material displacement is shown

in Table E3 and indicates that the largest rider will not come into contact with the ground

during a rollover condition.

Table E1: Stresses and factors of safety in different loading conditions

Final

Optimization

Top Load

Side Load – Frame Side Load –

Fork

Maximum Von

Mises stress (psi)

7330 18500 9810 Fatigue Factor of saftey

(ASME/Gerber)

Yield strength

(psi)

35000 35000 35000 Main tube Boom

Factor of safety

against yield

4.8 1.9 3.6 4.1/3.9 1.8/1.7

31

Table E2: Frame modal frequencies and weight

Final

Optimization

Mode 1 (Hz) 20

Mode 2 (Hz) 46

Mode 3 (Hz) 62

Mode 4 (Hz) 72

Mode 5 (Hz) 100

Frame

Volume (in^3)

133

Frame Weight

(lbf)

13

Table E3: Displacement and strain due to top and side loading conditions. x indicates not applicable

measurements.

Top Load Side load

Maximum elastic deformation (in) 2.31E-04 7.20E-04

Critical location displacement X (in) -1.64E-01 x

Critical location displacement Y (in) -6.50E-02 x

Critical location displacement Z (in) x x

Beginning clearance 0.83 0.83

Ending clearance 7.65E-01 0.83

Formulation Section

Given: Model of the Vike Bike I and II, reported by 2007 PSU HPV Team in their ASME Design

report, show rider weight of 100lbf at the seat tube bends, 1680in*lbf moment about the z axis,

and a 1320 lbf moment about the y axis (see Figure E1, below). Vike Trike I and II displacement

results are listed in Table E4. The ASME “Interpretation of Rollover/side Protection Rules for

2008 HPVC Competitions” states a top load of 485 lbf and a side load of 260 lbf (see Figure E2),

where the loads are applied independently.

32

Figure E1: FEM with Boundary Conditions for the Vike Trike II.

Table E4: Comparison of Stress and Displacement for Vike Trikes I and II.

Figure E2: Loading requirements determined by ASME

33

Find:

a) Model 2008 Vike Bike design and determine the maximum stress and displacement

using 6061 Aluminum. Show validity of model through convergence. Compare results to

Vike Trike II.

b) Find maximum stress and displacement using the same model as in part (a), using

Titanium and 4140 steel. Compare results.

c) Perform a modal analysis of the three models in part (b). Compare results.

d) Select frame material.

e) Using ASME top and side loads, show that the roll bar is acceptable and that there is

no permanent deformation or fracture on either the roll bar or the vehicle frame. Find

factor of safety against fatigue failure.

Assumptions:

The fork, frame, and boom are all one piece.

Dropouts, wheels, headset, cranks, and bottom bracket are not modeled because they

are design to withstand bicycle loading.

ASME top and side loads are not considered cyclical and are not used in a fatigue

analysis.

Rider weight is a point load in the seat tube bends.

Solution:

a)

Geometry –

The geometry was created with as a 3-D wireframe sketch in Solid Works and imported into I-

deas 5. Beam cross sections were keyed in for the seat tubes, roll bar, head bar, boom, main

tube, forks, and fork bends. Where the tube and main tube come together and where the head

tube and steer tube come together were both modeled as beams with the outer diameters of

the bigger beam and the inner diameter of the smaller beams.

Mesh-

34

Meshes were created using the beam mesh option in I-deas. Four meshes using 6061 Aluminum

for the body and generic isotropic steel were created. Different element sizes were used to

show the solutions convergence. Meshes 1 through 4 had 113, 161, 257, and 283 elements

respectively

Boundary Conditions -

To compare stress and displacement to Vike Trike II, similar loading and restraints were

employed to those shown in Figure E1, above. Two point loads of 100lbs each were applied to

the seat tube bends. The rear dropouts were fixed from translation, and the front dropouts

were fixed from translation in the y and z axis. A boundary conditions set was then created. This

was all accomplished under the Boundary Conditions task-icon, heat transfer mode. All of the

constraints and forces included in the boundary conditions set can be seen in Figure E3.

Figure E3: Vike Bike I frame and boundary conditions for dynamic loading.

Results -

A solution set was created for the models with their boundary condition sets, and the models

were solved under the Model Solution task-icon. Table 5E5, below, shows the results for

35

maximum Von Mises stress, x displacement, y displacement, z displacement, and strain energy

for each of the meshes. The stress and displacement results converged almost immediately

(See Figures E4 and E5). The 283 element mesh was optimized to reduce the strain energy

encountered from the 257 element mesh (see Figure E6).

Table E5: Maximum Von Mises stress , displacement, and strain energy for each model.

Elements

Maximum Von Mises

Stress (psi) X (in) Y (in) Z (in)

Strain Energy

(in*lbf)

113 2.36E+04 8.12E-02 9.29E-02 1.40E-01 3.25E+00

161 2.36E+04 8.12E-02 9.29E-02 1.47E-01 3.87E+00

257 2.36E+04 8.12E-02 9.31E-02 1.40E-01 1.10E+00

283 2.36E+04 8.12E-02 9.31E-02 1.40E-01 6.51E-01

Figure E4: Convergence of Von Mises Stress

0.00E+00

5.00E+03

1.00E+04

1.50E+04

2.00E+04

2.50E+04

0 50 100 150 200 250 300

Maximum Von Mises Stress

(psi)

Number of Elements

36

Figure E5: Convergence of Displacement

Figure E6: Convergence of Strain Energy

0.00E+00

2.00E-02

4.00E-02

6.00E-02

8.00E-02

1.00E-01

1.20E-01

1.40E-01

1.60E-01

0 50 100 150 200 250 300

Displacement (in)

Number of elements

X (in)

Y (in)

Z (in)

0.00E+00

5.00E-01

1.00E+00

1.50E+00

2.00E+00

2.50E+00

3.00E+00

3.50E+00

4.00E+00

4.50E+00

0 100 200 300

Strain Energy (in*lbf)

Number of elements

Strain Energy (in*lbf)

37

b)

Mesh-

Meshes were created using the 283 element beam mesh used in part (a). Meshes 1 through 3

were made with 4140 steel, 6061 aluminum, and typical titanium bodies, while maintaining the

fork as generic isotropic steels. Tube sizes were kept the same for each mesh.


Boundary conditions were the same as in part (a).

Results -


were solved under the Model Solution task-icon. Table E6 shows the results for maximum Von

Mises stress, x displacement, y displacement, z displacement, and strain energy for each of the

meshes. The Vike Trike II stress and displacement results are reproduced for ease of

comparison. Titanium versus aluminum evaluation is included as well and shows a titanium

displacement of 80% that of Aluminum. Figure E6 shows the location of the maximum Von

Mises stress in the frame and the fork, for each model. Figure E7 shows the location of the

maximum displacement for each model.

Table E6: Tabulated results of stress and displacement in each body.

4140 Steel Al 6061 Titanium (Typical) Vike Trike II Ti vs Al (%)

Maximum Von Mises

Stress (psi) 2.34E+04 2.52E+04 2.46E+04 8.29E+03 97.62

Max. Disp. X (in) 5.50E-02 8.12E-02 5.53E-02 7.96E-02 68.10

Max. Disp. Y (in) 4.32E-02 9.31E-02 6.55E-02 2.06E-01 70.35

Max. Disp. Z (in) 7.50E-02 1.40E-01 1.10E-01 1.57E-01 78.57

Displacement

Magnitude (in) 8.45E-02 1.53E-01 1.22E-01 2.71E-01 79.74

Strain Energy (in*lbf) 6.67E-01 6.51E-01 1.24E+00

38

Figure E7: Maximum Von Mises Stress locations for (clockwise from top left)

4140 steel, 6061 aluminum, and typical titanium.

39

Figure 8H: Maximum displacement location for (clockwise from top

left) 4140 steel, 6061 aluminum, and typical titanium

c)

Mesh-

Meshes were the same as the meshes used in part (b)


Restraints were the same as those used in parts (a) and (b). Boundary condition sets were made

under normal linear dynamics, Lanzcos method.

Results -


were solved under the Model Solution task-icon.

40

Table E7 contains the results for the first five modes in each model. The bike weight is also

included for each type of material. It can be seen that 4140 is too heavy to be used effectively

in the human powered vehicle. Aluminum 6061 provides the lightest bike weight by far, with

similar modes to those found in 4140 and Titanium. Titanium offers 80% of the total deflection

than that of Aluminum with similar tube sizes.

Table E7: Lanzco’s method

Modal analysis 4140 Steel Al 6061 Titanium (Typical)

Mode 1 (Hz) 9.85E+00 1.12E+01 1.03E+01

Mode 2 (Hz) 3.13E+01 2.91E+01 3.09E+01

Mode 3 (Hz) 3.85E+01 3.76E+01 3.78E+01

Mode 4 (Hz) 7.39E+01 7.23E+01 7.21E+01

Mode 5 (Hz) 8.18E+01 9.03E+01 9.34E+01

Total Weight (lbf) 49.5 17.25 29

Densities from

eFunda

d)

The results (see Table E8) indicate that an aluminum frame with the same tube sizes is 56-60%

the weight of a titanium frame, and 35% of the weight of the steel frame. Comparison of the

Vike Bike (using aluminum) and the Vike Trike II shows the Bike with 56% of the total

displacement than that of the Vike Trike. The Bike has 50% more stress which lowers the factor

of safety from 4.46 to 3.06. With the light weight of aluminum, adequate factor of safety, low

cost, and ease of manufacture, it is recommended that Aluminum 6061 be used for the Vike

Bike I frame. Table E8 shows a comparison of the Vike Bike against the Vike Trike II.

Table E8: Comparison of Vike Bike and Vike Trike II

Al 6061 Bike Vike Trike II Al. Bike vs. Trike (%)

Maximum Von Mises Stress (psi) 1.21E+04 8.29E+03 145.96

Max. Disp. X (in) 8.12E-02 7.96E-02 102.01

Max. Disp. Y (in) 9.31E-02 2.06E-01 45.19

Max. Disp. Z (in) 1.40E-01 1.57E-01 89.17

Displacement Magnitude (in) 1.53E-01 2.71E-01 56.47

Factory of Safety 3.06E+00 4.46E+00 68.51

41

e)

Geometry –

The geometry was changed slightly to accommodate ease of manufacturing and new fork

models.

Mesh -

Meshes were kept 283 +/- 20 elements.


For fatigue, the dynamic loading listed above was broken into mean and alternating loading

conditions. A top loading and a side loading was created as per ASME rollbar standards. ASME

guidelines specify that the fork and dropouts must be free from translation. This restraint set

was used for all models. A modal boundary condition set was created using Lanzco’s method.

Results –

A sample (Final Optimization) calculation for fatigue in the Boom is provided:

Fatigue analysis - Boom

*All tables and equations in the fatigue analysis section refer to Mechanical Engineering Design, 7th ed., by Shigley

ksi Ultimate strength in 6061-T6 aluminium

Yield strength in 6061-T6 aluminum

Endurance strength in 6061-T6 aluminium

Marin Factors

Machined parameters from table 7-4

Sut 38

Sy 35ksi

Seprime 0.504Sut

a 2.70

42

Equation 7-18

Surface factor ka

Boom diameter

Equivalent diameter, equation 7-23

Equation 7-19, for diameter .11<=d<=2in

Size factor kb

Loading factor for bending, equation 7-25

Temperature factor at 20 deg C, table 7-6 and equation 7-27

95% reliability factor from Table 7-7

Equation 7-17

ksi

Modified endurance factor

Load line

Factor of safety against first cycle failure

b 0.265

ka a Sutb

ka 1.03

d 1.75in

de0.370d

in

kb 0.879de0.107

kb 0.921

kc 1

kd 1

ke 0.868

Se ka kb kc kd ke Seprime

Se 15.763

Se 15.682ksi

m 9.79ksi

a 7.65ksi

a

m

0.781

ny

Sy

a m

ny 2.007

43

Fatigue in the main tube was analyzed with a similar method but with slightly different Marin

factors to due to the different tube sizes.

The results for each optimization, including the original and modified geometry, are included

below. Table E9 shows the maximum stresses and factors of safety for the specified loading

conditions. Table E10 shows the modal frequencies and weight for each optimization model.

The results in Table E11 show that the maximum elastic strain and displacement are within the

ASME specified requirements and do not endanger the rider.

Table E9: Maximum stresses and factors of safety for specified loading conditions First Approximation Top Load -

frame

Side load - frame Side load -

fork

Dynamic -

frame

Dynamic - fork

Maximum Von Mises stress (psi) 4.42E+03 1.74E+03 4.28E+04 1.21E+04

Yield strength (psi) 3.50E+04 3.50E+04 1.02E+05 3.50E+04

Factor of safety against yield 7.92 20.1 2.4 2.9

Optimization 1

Maximum Von Mises stress (psi) 4.43E+03 1.96E+04 4.02E+04 9.26E+03 1.85E+04

Yield strength (psi) 3.50E+04 3.50E+04 1.02E+05 3.50E+04 1.02E+05

Factor of safety against yield 7.9 1.8 2.5 3.8 5.5

Optimization 2

ASME-Elliptic Failure Criteria, Table 7-11

Factor of safety against fatigue failure

Gerber Failure Criteria, Table 7-10

Factor of safety against fatigue failure

nf1

a

Se

2m

Sy

2

1

2

nf 1.778

nf1

2

Sut ksi

m

2

a

Se

1 12 m Se

Sut a ksi

2

1

2

nf 1.67

44

Maximum Von Mises stress (psi) 6.18E+03 2.08E+04 4.15E+04 9.26E+03 1.86E+04

Yield strength (psi) 3.50E+04 3.50E+04 1.02E+05 3.50E+04 1.02E+05

Factor of safety against yield 5.7 1.7 2.5 3.8 5.8

Top Load -

frame

Side load - frame Side load -

fork

Dynamic -

frame

Dynamic - fork

Optimization 3

Maximum Von Mises stress (psi) 5.63E+03 1.79E+04 3.87E+04 2.06E+04

Yield strength (psi) 3.50E+04 3.50E+04 1.02E+05 3.50E+04

Factor of safety against yield 6.2 2.0 2.6 1.7

Optimization 4

Maximum Von Mises stress (psi) 1.05E+04 1.79E+04 3.87E+04 Fatigue Factor of saftey

Yield strength (psi) 3.50E+04 3.50E+04 1.02E+05 Main tube boom

Factor of safety against yield 3.3 2.0 2.6 2.54/2.4 3.4/3.2

Optimization 5




Optimization 6




Optimization 7




Final Optimization

Maximum Von Mises stress (psi) 7330 18500 9810 Fatigue Factor of saftey

(ASME/Gerber)

Yield strength (psi) 35000 35000 35000 Main tube Boom


45

Table E10: Modes and weight for each model Origin

al

Optimizatio

n 1

Optimizati

on 2

Optimizati

on 3

Optimizati

on 4

New

Geometry

Optimizati

on 6

Optimizati

on 7

Final

Optimizati

on

Mode 1 (Hz) 11 17 20 21 21 21 20 19 20

Mode 2 (Hz) 29 33 33 38 36 36 35 32 46

Mode 3 (Hz) 38 36 38 40 39 39 40 49 62

Mode 4 (Hz) 72 69 70 82 80 75 70 69 72

Mode 5 (Hz) 90 93 96 115 111 112 111 100 100

Frame

Volume (in^3)

178 160 128 128 138 137 128 131 133

Frame Weight

(lbf)

17 16 12 12 13 13 12 13 13

Table E11: Displacement and strain due to top and side loading conditions. x indicates not applicable

measurements.

Top Load Side load

Maximum elastic deformation (in) 2.31E-04 7.20E-04

Critical location displacement X (in) -1.64E-01 x

Critical location displacement Y (in) -6.50E-02 x

Critical location displacement Z (in) x x

Beginning clearance 0.83 0.83

Ending clearance 7.65E-01 0.83

Analysis references:

Shigley, J.E.. 2004. Mechanical Engineering Design. 7th ed. New York: Mcraw-Hill

American Society of Mechanical Engineers. 2008. Interpretation of Rollover/side Protection

Rules for 2008 HPVC Competitions. http://files.asme.org/asmeorg/Events/Contests/

HPV/13615.pdf (10 March 2008).

46

Appendix F: Biomechanical Testing

Summary Section

For the finite element model of the frame it is important to apply accurate forces insuring that

the frame is not overbuilt. This helps cut the overall weight of the Vike Bike I. A top speed

analysis will also be more accurate knowing the actual forces and power the riders are able to

subject the bike to. These forces are measured in the x, y, and z directions using a PCP

Piezoelectric force transducer model U206A203 that is attached to the Vike Trike II pedal as

seen in Figure F1. A Model Shop low speed laser tachometer was also used to measure the

cadence for each rider.

Figure F1: PCB Force Transducer bicycle pedal adapter

The testing was performed on the Vike Trike II with the use of LABView and a NI-9233 data

acquisition device. The PDS states that the Vike Bike I will reach a top speed greater than 45

miles per hour. With this testing the Vike Bike I’s top speed can me more accurately calculated.

The PDS also lists a light overall weight of the Vike Bike I. The biomechanical testing will help

the design team design the frame so that it is light and strong enough to withstand the loading

each rider will produce while riding.

The PCB force transducer’s output signal is a voltage in each direction x, y, and z. A conversion

factor is determined through testing and calibration from PCB allowing the conversion from mV

to Newton or pound force. The signal output for the laser tachometer is also voltage that can

be tabulated using the virtual instrument. The data can then be graphed comparing the

resultant force for x, y, and z to the rider cadence. Below a table was constructed showing the

rider maximum force and cadence.

47

Table F1: Rider force output, cadence

Rider Max Force (Lbf) Average Force (Lbf) Cadence (rpm) Max Power (Hp) Average Power (Hp)

Ben 104.50 40.08 98.40 1.12 0.43

Bryan 105.78 30.12 90.00 1.04 0.30

Chantelle 96.36 42.39 72.00 0.76 0.33

Erik 136.10 52.37 88.00 1.31 0.50

Kenneth 163.47 36.16 78.00 1.39 0.31

levi 82.40 25.05 90.00 0.81 0.25

Formulation Section

Given: Figure F2

Figure F2: Rider resultant force and cadence output for 5 seconds.

Find:

-Moments about center of bottom bracket.

Assumptions:

-Ignore losses in pedal connection and bottom bracket.

Solution:

Analysis Reference: Bolen, Ben and Bryan Voytilla. 2008. ME 411 Final Lab Report

Force 104.5lbf

crank_radius 175mm

Mbb Force 2 crank_radius

Mbb 119.997ft lbf

48

Appendix G: Braking Analysis

The Vike Bike’s breaking performance is described by the distance required to come to a

complete stop at a given vehicle velocity. This distance is defined in Equation G1 (Wilson, 2004):

𝑆 =𝑉2

20(𝐶𝐴+𝐶𝑅) (Eq. G1)

where S is the distance required to stop in meters, V is the vehicle velocity in meters per

second, CA is the coefficient of adhesion for the material the vehicle is breaking on, and CR is the

coefficient of rolling resistance for the vehicle and is s function of vehicle weight. Figure G1

demonstrates that the Vike Bike is able to achieve a stopping distance of 8.33 feet while

traveling at a velocity of 15 mph, given concrete with a CA of 0.8 and a CR of 0.035.

Figure G1: log- log scale plot of stopping distance vs. vehicle velocity for the 2008 Vike Bike

10, 3.70

15, 8.3320, 14.82

25, 23.1530, 33.34

40, 59.2750, 92.61

1

10

100

1000

10 100

Sto

p D

ista

nce

(Fe

et)

Velocity (MPH)

Concrete (Dry)

Concrete (Wet)

Gravel, Rolled

Sand, Loose

Ice

49

Appendix H: Top Speed Analysis

Summary Section

Accurately demonstrate that the desired top speed of 45 mph can be achieved given the Vike

Bike I’s drive train. The Vike Bike I’s top speed analysis incorporates the biomechanical rider

force output data. These realistic forces can be applied to energy and gearing calculation to

determine the top speed that can be achieved given the drive train of the Vike Bike I. The Vike

Bike I drive train is based on the conventional drive train of a bicycle with a few exceptions in

gear selection. Two custom bicycle chain rings 65 tooth and 75 tooth were fabricated and

donated for use in the spring competition. The PDS states that a top speed of greater than 45

mph will be achieved during the HPVC sprint competition.

Given the tabulated results used in Appendix F: Biomechanical testing the forces are used to

calculate the overall achievable speed for the Vike Bike. Maximum speed: 45.28 Miles/hour

While calculating the top speed all forces exerted on to the Vike Bike I are included in the

overall attainability equation.

Formulation Section

Given: The current drive train setup on the PSU Vike Bike I human powered vehicle in Figure H1.

Air density, frontal area, coefficient of drag, mass of Vike Bike I is 1.266 kg/m3 , 0.535m2 , 0.14,

and 28 Kg respectively.

Figure H1: PSU Vike Bike I drive train.

Find: What is the theoretically maximum achievable speed?

50

Assumptions: Speed is achieved at sea-level

Solution:

Top Speed:

Frictional Force:

Tire_Diameter 700mm

Tire_Circumferance Tire_Diameter

Crank_Length 175mm

Cadence90

min

efficiency 0.90

ncrank 75

nwheel 11

Wheel_speed Cadencencrank

nwheel

efficiency

Wheel_speed 9.2051

s

Max_Speed Tire_Circumferance Wheel_speed

Max_Speed 45.28mile

hr

fkinetic_friction 0.0021

Frictional_Force Normal_Forcefkinetic_friction

Frictional_Force 2.585N

51

Air Drag:

Applied Force:

The attainability equation relates force balance of the mechanical force produced by the rider

and the friction and aerodynamic drag force applied to the Vike Bike I. A positive value

indicated a larger rider mechanical force than the drag force due to aerodynamics and friction

thus, proving a speed achievable.

Air_density 1.226kg

m3

Frontal_area 0.535m2

Cd 0.14

Air_drag1

2Cd Air_density Frontal_area Max_Speed

2

Air_drag 18.812N

Max_Force 104.5lbf

Max_Force Crank_Length

81.347J

Wheel_torque

ncrank

nwheel

efficiency

Wheel_torque 10.738J

Applied_ForceWheel_torque

Tire_Diameter

2

Applied_Force 30.679N

52

This top speed formulation was tested at the ASME competition where the Vike Bike race team

was able to reach an unofficial speed of 44 Mph in the sprint race with an allowable run up of

600 meters.

Attainability Applied_Force Air_drag Frictional_Force( )

Attainability 9.282 N

53

Appendix I: Turning Analysis

Summary Section

ASME rules state that a vehicle must be able to turn within a 25ft radius. A geometric analysis

was performed to see how many degrees the fork of Vike Bike I must turn to fit within this

circle. Through solid modeling, it was found that the maximum amount that the steering tube

could rotate was 31.88 degrees. Since the steering tube and fork are connected directly, this is

also the maximum amount that the fork is able to turn. With this information, the minimum

achievable turn radius (in feet) was calculated. The results of the analysis listed below:

a) Turning angle required to achieve 25ft turn radius – 10.6 deg

b) Turning radius with a 31.88 degree turning angle – 8.7 ft

Each situation was modeled as a worst-case scenario. After starting the initial turn, the angle

required to continue the turn will decrease. Leaning into the turn also decreases the turning

angle required. Accounting for continuing a turn and leaning into account, the amount of

turning angle and the turning radius will decrease. Therefore, the design meets and exceeds

ASME standards.

Formulation Section

Given: Figure I1 shows Vike Bike I having a wheelbase of 54.95 inches. The maximum angle that

the steer tube can turn without hitting the fairing is 31.88 degrees, as shown in Figure I2.

Figure I1: Top view of the Vike Bike. The wheelbase measurement is

important in the view plane (i.e. Delta X) only.

54

Figure I2: When the steer tube has reached its maximum swing, the angle

between the steer tube and the main tube is 31.88 deg, measured from the

top view.

Find:

a) The angle that the fork must turn to achieve the ASME required turn radius of at least 25ft.

b) The turn radius when the steer tube is at 31.88 degrees.

Assumption:

1. Leaning when riding improves the turn radius, and not leaning corresponds to lower

speed biking. Performing the analysis without taking leaning into effect will be

considered as a worst case scenario.

2. Starting the turn requires more of a turning angle than continuing the turn. Therefore,

only the initial angle will be considered as a worst case scenario.

Solution:

a) Figure E3 provides a sketch of the system at the start of a low speed turn.

55

Figure I3: Sketch of the wheelbase (solid line) and the 25 foot circle with

radius (dashed line). All dimensions are in inches.

The front wheel is coincident with the curve (see fig I3). To turn the Vike Bike in the circle, the

front wheel must be turned at an angle from the frame until it is tangent with the circle. A right

triangle is created from fig F3 and is shown.

Figure 4I: Sketch of tangent line (solid line) and right triangle. Hyp is the

hypotenuse and the radius of the circle, b is the wheelbase, and a is the

desired angle.

By geometric association, it can be seen that the desired angle is the same angle in the right

triangle (fig’s 4I & E4). Angle a is found using equation I1

(Eq. I1)

Substituting values for b and hyp into eqn I1 yields the desired angle.

a asin54.95

300

a asinb

hyp

a 10.6deg

56

Desired angle for the ASME prescribed minimum turn radius of 25ft.

b)

A sketch of the system with the given information is provided in figure I5.

Figure I5: Sketch of the wheelbase and turn circle when the fork is turned to

its maximum angle.

Solving equation I1 for the hypotenuse gives:

(Eq. I2)

Substituting the values of a and b into equation I2 yields:

Converting the turn radius to feet yields:

For the turn radius when the steering tube is at an angle of 31.88 degrees.

hypb

sin a( )

hyp54.95in

sin 31.88( )

hyp 104.044in

hyp 8.7ft

57

Appendix J: Strain Gauge Testing

The ASME roll-bar guidelines for 2008 state that each team is granted points for demonstrating

that their roll-bar is able to withstand a vertical load of 485-lb 8° off of vertical and a side load

of 260-lb at shoulder height (Figure J1). Each team is able to demonstrate verification by

physical testing, computational analysis, and safety. Failure to demonstrate the loading

requirements in any category will result in a deduction of points for that team.

Figure J1: ASME Roll-bar analysis loading criteria top load of 485-lb and side load of 260-lb

Strain gauges are placed in the areas’ of interest as pointed out in the frame finite element

analysis section and also seen in Figures J2 and J3. The vertical physical loading is

demonstrated in figure J4 with the Vike Bike restrained at the front and rear axils and tilted 8°

off of horizontal. A load of 492-lb was then hung from the middle of the roll-bar.

Figure J2: Strain gauge placement for top loading criteria (485-lb)

58

Figure J3: Strain gauge placement for side loading criteria (260-lb)

Figure J4: Strain gauge testing for the top loading roll-bar criteria (492-lb)

The side load of 265-lb was applied to the side of the roll-bar while the front and rear axles are

restrained (figure J5).

59

Figure J5: Strain gauge testing for the side loading roll-bar criteria (260-lb)

The final results are tabulated in table J1 and compared to the theoretical results from the finite

element analysis. Completing this demonstration of loading the bike resulted in zero points

deducted for the analysis, testing, and safety category of the roll-bar strength verification thus,

proving that the Vike Bike will be able to withstand a roll-over.

Table J1: Strain gauge results. Side Load Top Load

Strain Gauge Result (maximum elastic strain)

Finite Element Analysis (maximum elastic strain)

Strain Gauge Result (maximum elastic strain)

Finite Element Analysis (maximum elastic strain)

0.00079 0.00072 0.00031 0.000231

60

Appendix K: Fairing Aerodynamic Analysis

Summary

The goal is to verify that the drag forces have been reduced when compared to the Vike trike

‘06/07 fairing. By manipulation of the shape of the fairing, specifically in the planar view as

shown in figure 1I, the coefficient of drag and the frontal area will be reduced when compared

to the Vike Trike II.

Frontal Area= 864 in2

Coefficient of drag (Cd) = 0.06

The frontal area and Cd of the 2008 fairing have been verified as being reduced compared to

the 2007 Fairing.

Formulation

Given:

Two models, shown in Figures K1 and K2, are available for analysis. A solid model of the Vike

Bike’s external surface created in SolidWorks 2007 is can be used for theoretical analysis while

a physical model has been constructed and can be used to collect experimental data.

Figure K1: 3D solid model of the Vike Bike fairing

61

Figure K2: Completed fairing and bike assembled

Find:

a) The frontal area

b) The coefficient of drag

Assumptions:

-External flow is normal to the frontal area of the fairing.

Solution:

a) The frontal area of the ‘06/07 fairing is found using the SolidWorks 2007 tool “Section

Properties” . The frontal areas of both fairings can be seen in the Figures below.

Figure K3: 2006/07 Fairing frontal area Figure K4: 2008 Fairing frontal area

886in2 864in2

62

The 2008 faring has a frontal area of 864in2 while the 2006/07 fairing has a larger surface area

of 886in2. This gives a 2.5% reduction in frontal area.

b) The coefficient of drag was found using two methods, theoretical and experimental. The

theoretical Cd value was found using CFD while the experimental value was found using a wind-

tunnel test facility provided by Freightliner. Both methods are presented below.

An idealized 3D model of the fairing was imported into the CFD program. The body seams and

wheel cutouts were omitted in order to simplify the meshing procedure. Pressure and velocity

plots for a 15mph external flow are shown in FigureK5. Table K1 shows the results of the CFD at

15 and 30mph.

Figure K5: Pressure plot at 15mph (left), velocity plot at 15mph (right)

Table K1: Results from CFD

Vehicle velocity Coefficient of Drag Net drag force

15mph .084 0.14 lbf

30mph .072 0.48 lbf

63

The completed fairing was taken to the Freightliner wind-tunnel where it was also tested under parallel wind conditions. The wind-tunnel results are shown in Table K2 below.


Vehicle velocity (mph) Coefficient of Drag

39.21 .0507

39.36 .0539

The 2007 fairing Cd values are shown in the Table K3 below.


Vehicle velocity (mph) Coefficient of Drag

22.4 0.11

44.7 0.12

The values presented in the tables above are related by Equation K1.

𝐹 =1

2𝜌𝐶𝑑𝐴𝑉

2 Where: F = Drag force (Eq. K1)

𝜌 = fluid density Cd = Coefficient of drag A = Frontal Area V = Fluid or vehicle

velocity

Assuming similar fluid conditions. the aerodynamic characteristics of the two fairings can be compared using only the coefficient of drag and frontal area. This is shown in equations K2 and K3.

CdA (2007) = (0.11)(886 in2) = 97.46 in2 (Eq. K2)

CdA (2008) = (0.06)(864 in2) = 51.84 in2 (Eq. K3)

Finding the percent reduction of drag force.

97.46−51.84

97.46= 47% reduction of drag force (Eq. K4)

64

Appendix L: Fairing Material Analysis

A fairing material was sought that will be stronger, more ductile, and lighter than the Vike Trike II fairing. A stronger and more ductile fairing is desired to protect the rider and absorb more energy during an impact, while an overall lighter fairing reduces rolling resistance and inertial forces. Seven samples were created to test fiber orientations of 0°- 45° and 0°- 90°, balsa wood and BaltekMat core, vacuum bagging and air drying, and two to three fiberglass layers (see Table L1). The samples were compared against the Vike Trike II fairing since it was tested during the 2007 ASME HPVC West competition and remained intact after multiple crashes. Each sample used S2 fiberglass and polyester resin with a balsa wood or BaltekMat core between the fiberglass layers. Due to budgetary constraints, carbon fiber was not considered as a potential solution because of its high cost relative to S2 fiberglass. To create a statistical average, five samples of each configuration were tested. Table L1: Sample configuration guide

Sample Layer Orientation Core Layup Method Layers

A (Vike Trike II) 0°-90° BaltekMat Air 3

B 0°-45° BaltekMat Vacuum 3

C 0°-90° BaltekMat Vacuum 2

D 0°-90° BaltekMat Vacuum 3

E 0°-90° Balsa (3/32") Vacuum 3

F 0°-45° Balsa (3/32") Vacuum 3

G 0°-45° Balsa (1/8") Vacuum 3

Strips of each sample were tested with a three point bend test (see Figure L1) according to ASTM standard D 790-97 (1998).

Figure L1: Three point bend test

The test data for each sample was used to find the modulus of elasticity in bending, weight per area, yield strength, and strain at yield. Modulus of elasticity in bending EH (psi) is calculated using EquationL1:

65

𝐸𝐻 =𝐿3𝑚

4𝑏𝑑3

(Eq.L1)

Where L is the support span (in), b is the width of the sample (in), d is the thickness (in), and m is the slope of the tangent to the initial straight line portion of the load deflection curve (lbf/in). The stress of the outer fibers at midspan, S (psi) is determined to be:

𝑆 =3𝑃𝐿

2𝑏𝑑2(1 + 6

𝐷

𝐿 2

− 4 𝑑

𝐿

𝐷

𝐿 )

(Eq.L2)

P is the load at a given point along the load-deflection curve (lbf), D is the deflection of the center of the center of the beam. The strain in the outer fibers, R (in/in) is given as:

𝑅 =6𝐷𝑑

𝐿2

(Eq.L3)

Equations L1 through L3 were used to reduce the raw data obtained during the experiment (see Figs. L2 and L3). Figure 19 shows that only sample E has approximately the same bending modulus as sample A, while having a greater strain at yield. In Fig. L3, only sample E has consistently higher yield strength, while having a lower weight per area than sample A. Thus sample E is the only configuration that fulfills the requirement of being stronger, more ductile, and lighter than sample A, the Vike Trike II fairing. Table L2 is a direct comparison of the average values of sample E to sample A with percent differences. Sources of error are due to manufacturing variability, sample preparation, and measurement error.

700.00

800.00

900.00

1000.00

1100.00

1200.00

1300.00

1400.00

0.006 0.008 0.010 0.012 0.014 0.016 0.018

Be

nd

ing

Mo

du

lus

(ksi

)

Strain at Yield (in/in)

Sample A

Sample B

Sample C

Sample D

Sample E

Sample F

Sample G

66

Figure L2: Bending modulus vs. strain at yield

Figure L3: Yield strength vs. Weight per area

Table L2: Comparison of average Vike Trike II and Vike Bike fairing properties

Weight per Area (lbf/ft^2) Strain at Yield (in/in) Yield Strength (ksi) Bending Modulus (ksi)

Sample A (2007 Vike Trike II) 0.50 +/- 0.02 0.011 +/- 0.001 10.8 +/- 0.8 1050 +/- 70

Sample E (2008 Vike Bike) 0.31 +/- 0.025 0.014 +/- 0.001 16 +/- 2.8 1100 +/- 100

% Difference from A -38 30 48 4.8

6.00

8.00

10.00

12.00

14.00

16.00

18.00

20.00

22.00

0.250 0.300 0.350 0.400 0.450 0.500 0.550

Yie

ld S

tre

ngt

h (

ksi)

Weight per area (lbf/ft^2)

Sample A

Sample B

Sample C

Sample D

Sample E

Sample F

Sample G

67

Appendix R: Vehicle Stability

Stability

Recumbent bicycle stability depends on the wheel base and the wheel trail Figure 2R. A large

trail (more than 3.5”) and long wheelbase (in excesses of 60”) leads to better high speed

stability. A short wheel base (less than 60”) and a smaller trail (less than 3.5) allows for better

handling at low speeds. The Vike Bike was designed with a wheel base of 54.95” and a trail of

2.56”, as a compromise between high speed stability and low speed stability. The stability of

the Vike Bike was tested and verified in the 2008 ASME HPV Competition. During the

competition the PSU HPV race team was able to test the Vike Bikes high speed stability in the

sprint race. This was successful without any major crashes that were caused by the stability of

the bike. The PSU HPV race team did however find out that extra practice was needed to learn

how to ride a recumbent bicycle as opposed to a standard upright bicycle due to the relative

positions of the center of gravities of each.

Figure 2R: Vike Trike II Front wheel configuration

68

Figure 1R: Vike Bike head tube angle design configuration

During the endurance race the low speed stability was tested by the PSU HPV race team and

found prior to competition that the rotation of the handlebars was inadequate and the side

read windows would need to be removed for this event. Upon further investigation the race

team found the fairing hard to maneuver due to rider unfamiliarity and decided to use only the

bottom half of the fairing. The bottom fairing would still provide the needed 10% frontal area

covering required for the race and also made the Vike Bike much more comfortable for the race

team to maneuver.

69

Appendix M: Frame Design Drawing

70

71

72

73

74

75

76

77

78

79

80

81

Appendix N: Fairing Design Drawings

The following drawings are derived from .dxf files that were given to OMSI. OMSI used the .dxf files to

write a machine code that was imported to a CNC router. This router cut the cross sections shown

below out of medium density fiberboard (MDF) used for the fabrication of the fairing plug.

82

83

84

85

Appendix O: Vike Bike Dimensions

This appendix illustrates the major Vike Bike dimensions and angles for the chassis and fairing.

86

Appendix P: Glossary of Human Powered Vehicle Terminology

The Bicycle industry has its own terminology for variety of components. Proper understanding of these

terms is critical to understanding the analysis and design in this report. Figure P1 illustrates the location

of Vike Bike components discussed in this report.

Figure P1: Vike Bike parts and locations

1. Rollbar 2. Rear Cassette 3. Chain Stay 4. Chain Ring 5. Chain Pulley 6. Main Tube 7. Seat Stay 8. Seat Tube 9. Seat Cross Tubes 10. Head Tube

87

Appendix Q: Bill of Materials

Table Q1 HPV Overall Components Cost List Frame Parts List

Component Quantity Size Brand Color Model Source Retail Price Total Price Contact

Front Hub 1 32 hole Chris King Green Chris King $100.00 $0.00

Head Sets 1 1 1/8" Chris King Green Chris King $130.00 $0.00

Rear Hub 1 135 Chris King Green Chris King $150.00 $0.00

Chain 4 10 speed SRAM SRAM Chain PC-1050 10spd Revolver $22.25 $89.00 Jake Furniss

IRD Chain Link 10sp Shimano 3 10 speed Shimano IRD Chain Link 10sp Shimano Revolver $10.95 $32.85 Jake Furniss

SRAM Power Link 2000 9sp Gold 2 SRAM SRAM Power Link 2000 9sp Gold Revolver $4.95 $9.90 Jake Furniss

Inline Tube 700 x 18/20C PV / 60mm 4 Inline Inline Tube 700 x 18/20C PV / 60mm Revolver $4.95 $19.80 Jake Furniss

Cyclone Tube 20x1 1/8 PV 3 Cyclone Cyclone Tube 20x1 1/8 PV Revolver $4.95 $14.85 Jake Furniss

Maxxis Tire 700c X 23mm 2 Maxxis Tire 700c X 23mm Revolver $30.00 $60.00 Jake Furniss

Front Tire 2 451c Schwalbe Black Stelvio HS 350 Revolver $30.70 $61.40 Jake Furniss

Rear Tire 2 700c Schwalbe Black Stelvio HS 351 Revolver $28.44 $56.88 Jake Furniss

Front Rim 1 451c Velocity Black AeroHeat 451mm 32 Hole Black WO/MSW

Revolver $37.00 $37.00 Jake Furniss

Spokes 50 14G DT Silver Champion Revolver $2.00 $100.00 Jake Furniss

Nipples 50 14G DT Silver Champion Revolver $0.25 $12.50 Jake Furniss

SRAM Derailleur Force Rear 1 Sram SRAM Derailleur Force Rear Revolver $91.71 $91.71 Jake Furniss

Sugino double chainring bolt box of 5 1 Sugino Sugino double chainring bolt box of 5 Revolver $9.75 $9.75 Jake Furniss

Sugino Chainring Bolt Single Ring Set / 5 1 Sugino Sugino Chainring Bolt Single Ring Set / 5 Revolver $12.50 $12.50 Jake Furniss

Cranks 1 175mm SRAM Carbon SRAM Crankset Force GXP 175mm 39-53 WO/BB Revolver $231.15 $231.15 Jake Furniss

Shim Cleat SM-SH51 SPD Pair 7 Shimano Shim Cleat SM-SH51 SPD Pair Revolver $10.44 $73.08 Jake Furniss

Bottom Bracket 1 Sram Revolver $20.13 $20.13 Jake Furniss

Aheadset Starnut 1" 1 Aheadset Starnut 1" Revolver $2.95 $2.95 Jake Furniss

* REDLINE 18" Flight U6 Micro Mini Fork 1 1 Redline Black * REDLINE 18" Flight U6 Micro Mini Fork 1" Threadless Black

Revolver $35.34 $35.34 Jake Furniss

Cassette 1 SRAM OG-1070 10spd 11-26T Revolver $55.66 $55.66 Jake Furniss

Front Brake 1 6" Avid Avid Disc Brake BB7 Road Front 160mm Rotor Revolver $13.79 $13.79 Jake Furniss

SRAM Shift / Brake Lever Force Pair 1 Sram SRAM Shift / Brake Lever Force Pair Revolver $341.55 $341.55 Jake Furniss

Surly Brake Hanger Stainless 1 Surley Surly Brake Hanger Stainless Revolver $10.95 $10.95 Jake Furniss

Velox Rim Tape Cloth 17mm Single 2 Velox Velox Rim Tape Cloth 17mm Single Revolver $3.95 $7.90 Jake Furniss

Rear Rim 1 700c Velocity Black AeroHeat Revolver $37.00 $37.00 Jake Furniss

Jagwire Brake Cable Teflon Tandem 1.6mm X 2750mm ATB/RD 1 Jagwire Brake Cable Teflon Tandem 1.6mm X 2750mm ATB/RD $7.95 $7.95 Jake Furniss

JAG Tandem Der Cable Slick Gal 1.1 x 3000 1 JAG JAG Tandem Der Cable Slick Gal 1.1 x 3000 Revolver $3.95 $3.95 Jake Furniss

Rear Cantilever Brake 1 Avid Cantilever Brake Shorty Rear Revolver $24.15 $24.15 Jake Furniss

http://www.whiteind.com/

http://www.chrisking.com/

http://www.sram.com/en/avid/index.php

http://schwalbetires.com/node/201/ok




http://www.velocityusa.com/

http://www.velocityusa.com/default.asp?contentID=618

http://www.velocityusa.com/default.asp?contentID=618

http://www.revolverbikes.com/

http://www.dtswiss.com/

http://www.dtswiss.com/index.asp?fuseaction=spokes.bikedetail&id=8

http://www.dtswiss.com/

http://www.dtswiss.com/index.asp?fuseaction=spokes.bikedetail&id=8




88

Delta Axlerodz Bolt On Skewer 1 Delta Axlerodz Bolt On Skewer Revolver $17.95 $17.95 Jake Furniss

DT Champion 2.0 (14ga) 294mm single 64 1 Revolver $0.40 $0.40 Jake Furniss

Rear Wheel 1 650c Veloforma Black Veloforma $999.00 $0.00 Mark Duff

SRAM 10 speed TT shifter 1 Sram Carbon SRAM 10 speed TT shifter Veloforma $119.00 $119.00 Mark Duff

Cane Creek TT Bar end Brakes 1 Cane Creek

Black Cane Creek TT Bar end Brakes Veloforma $35.00 $0.00 Mark Duff

Idler Volae Over/Under... 1 64mm Terracycle Black Titanium 15% Discount Terracycle $133.45 $133.45 Robert Johnson

Idler Trice Dual Ti 2 Terracycle Titanium 15% Discount Terracycle $136.85 $273.70 Robert Johnson

Idler Mount 2" 2 Terracycle Universal Idler mount for 2" tube Terracycle $32.20 $64.40 Robert Johnson

GlideFlex Zero F 1 1" tube Terracycle Recumbent Steering mount for 1" tube Terracycle $92.65 $92.65 Robert Johnson

Handel Bar Mast 1 Terracycle Recumbent Bicycle Handlebar Mast Terracycle $75.65 $75.65 Robert Johnson

Handel Bars 1 Terracycle Handlebars Terracycle $29.75 $29.75 Robert Johnson

Fork 1 20" with D-Break Bachetta Coventry cycle Works $40.00 $0.00 Sherman

Chain Ring 2 75/65 Custom Premier Gear $0.00 $0.00 Ed Smith

6061 AL BB Relived Shell 1 41mm OD x 68mm wide Nova cycles $8.75 $8.75

AL7005 Road Dropout Adj 1 Al 7005 vertical Drop Nova cycles $19.15 $19.15

Easton head Tube AL 6061 1" 1 34.9 x 3.1 x 200mm Nova cycles $8.40 $8.40

6061 U-Brake Boss 1 6061 U-Brake Boss with center Miter Nova cycles $6.47 $6.47

Shipping cost 1 Nova cycles $40.00 $40.00

Total $2,363.41

Ironclad

Shirts and gloves $1,339.60 $669.80

Total $669.80

Tools

Angle Finder 4" 1 Tube Notching AC Winks $12.99 $12.99

1.5" 6 fluted end mill 1 Tube Notching McMaster Carr $71.87 $71.87

2" 8 fluted end mill 1 Tube Notching McMaster Carr $92.00 $92.00

Total $176.86

Hardware

Cap Screws 8 3" (3/8"-16) Clamps AC Winks $1.20 $9.60

All Thread 1 Bio-Mechanical Testing AC Winks $17.99 $17.99

Total $27.59

Raw Materials

6061 Aluminum Tubing 1 Vary Frame Materials Tube Service $170.00 $170.00

6061 Aluminum Tubing 1 Second Order Tube Service $188.00 $188.00

Aluminum Plate 1 2'x2'x0.75" Fixture Plate Alaskan Copper $238.48 $238.48

Total $596.48

Fairing Parts List

Tools

89

Vacuum Bagging 1 FiberGlast $705.55 $705.55

Extra Vacuum Bagging 1 FiberGlast $94.50 $94.50

1/4 Drill Bit 1 Home depot $7.98 $7.98

Lite Weight 1 Baxter Auto Parts $19.99 $19.99

Lite Weight 1 Baxter Auto Parts $19.99 $19.99

Liquid Wax 1 Schucks Auto Supply $11.49 $11.49

Wet/Dry 400 1 Schucks Auto Supply $3.99 $3.99

Wet/Dry 800 1 Schucks Auto Supply $4.49 $4.49

Primer and Stuff (Kenneth) 1 Foster Auto Parts $35.53 $35.53

Canopy (Acrylic / Poly) 1 $150.00 $150.00

Total $1,053.51

Raw Materials

S2 Glass 5.60oz 70 yd TAP $14.00 $980.00

E Glass 1.5oz Mat 8 yd TAP $3.70 $29.60

5 Gal Structural Resin 1 Can TAP $210.00 $210.00

5 Gal Structural Resin 1 Can TAP $210.00 $210.00

Whitle Gel Coat 1 Gallon TAP $74.95 $74.95

White Pigment 8oz 3 Can TAP $15.25 $45.75

Green Pigment 8 oz 3 Can TAP $15.25 $45.75

PVA Mold Release 1 Gallon TAP $23.40 $23.40

Acetone 1 Gallon TAP $16.00 $16.00

CAB-O-SIL 1/8 LB 2 Can TAP $9.50 $19.00

Box Gloves 1 Box TAP $8.50 $8.50

2" Brushes 24 TAP $1.05 $25.20

Formula Five Can 1 Can TAP $11.25 $11.25

Gel Coat Gauge 1 TAP $8.00 $8.00

1 LB Clay 1 TAP $3.60 $3.60

10 oz MEKP 1 Can TAP $8.50 $8.50

BaltekMat 3 yd TAP $1.50 $4.50

Velcro 3 yd TAP $1.50 $4.50

1 qt Cups 6 TAP $0.65 $3.90

Tack Cloth 5 TAP $1.40 $7.00

Surfacing Resin 1 Gallon TAP $51.00 $51.00

E Glass 0.75oz Mat 10 yd TAP $1.90 $19.00

Micro Spheres 4oz 3 Can TAP $8.45 $25.35

Box Large Gloves 2 Box TAP $8.50 $17.00

5 qt Buckets 5 Each TAP $2.15 $10.75

1 qt Buckets 6 Each TAP $1.25 $7.50

90

Latex Gloves 1 Each Walmart $4.32 $4.32

Duck Tape 1 Each Walmart $2.97 $2.97

Paint Brush 4 Each Walmart $0.88 $3.52

Syringe 1 Each TAP $3.10 $3.10

3/32" Balsa 12 Not Final Hobby Town USA $10.00 $120.00



Large Gloves 3 Box Rite Aid $7.99 $23.97

Medium gloves 3 Box Rite Aid $7.99 $23.97

Mask 1 Each Rite Aid $3.99 $3.99

MDF & (2x4) 1 Levi Home depot $185.00 $185.00 Total $2,320.84

91

Table Q2 HPV Overall Components Cost List Frame

Components $2,363.41

Tools $176.86

Materials $596.48

Hardware $27.59

Frame Cost $3,164.34

Fairing

Tools $1,053.51

Materials $2,320.84

Fairing Cost $3,374.35

Competition

Rooms $2,061.46

Rental $850.00

Gas Wednesday Truck/Van $810.46

Gas Firday Car $33.00

Trailer $100.00

Registration $350.00

Food $924.00

Ironclad $669.80

Screen Printing $388.00

Travel Cost $6,186.72

Total Cost $12,725.41

Fund Raising

ASME Budget 2009 $1,720.66

ASME Budget 2008 $7,404.75

Senior ASME $1,000.00

Cummins NW $1,000.00

Cummins Foundation $1,000.00

Tap Plastics $500.00

Competition $100.00

Student Fee's

Total Raised $12,725.41

Balance $0.00

92

Thank you to all of our sponsors.

93

References

American Society of Mechanical Engineers. 2008. Interpretation of Rollover/side Protection Rules for 2008 HPVC Competitions. http://www.asme.org (1 March 2008).

American Society of Mechanical Engineers. 2008. Rules for the 2008 Human Powered Vehicle Challenge.

http://www.asme.org (15 March 2008). American Society for Testing and Materials. 1998. Annual Book of ASTM Standards. Designation D 790-

97 Standard Test Methods for Flexural Properties of Unreinforced and Reinforced Plastics and Electrical Insulating Materials: 145-152.

Callister, Jr. W.D. 2003. Materials science and engineering an introduction. 6th ed. New York: John Wiley

and Sons. Furniss, J., Kappa, B., Smith, M., Stenkamp, J., and N. Tavan. 2007. The Vike Trike II ASME West Coast

HPV Challenge. Portland, OR: PSU Mechanical Engineering Department. Guitterez, A. 2006. High-Strength Stainless Steels. Advanced Materials and Processes (June 2006). Landwer,G. E and D. Too. 2003. Factors affecting performance in human-powered vehicles: a

biomechanical model. Human Power 54:14-16. Reiser, P. 2001. Anaerobic Cycling Power Output With Variations in Recumbent Body

Configuration, Colorado State University Mechanical Engineering Dept. Tomai, G. 1999. The Leading Edge, Aerodynamic Design of Ultra-streamlined Land Vehicles. Cambridge,

MA: Robert Bentley. Too, D. 1990. The effect of body configuration on cycling performance. In E. Kreighbaum& McNeill

(eds.),Biomechanics in Sports VI (pp. 51-58). Montana State University,Bozeman, Montana ------. 1991. The effect of hip position/configuration on anaerobic power and capacity in cycling.

International Journal of Sports Biomechanics, 7(4), 359-370 ------. 1994. The effect of body orientation on power production in cycling. The Research Quarterly for

Exercise and Sport, 65, 308-315 ------. 1996. Comparison of joint angle and power production during upright and recumbent cycle

ergometry. In J.A. Hoffer, A. Chapman, J.J. Eng, A. Hodgson, T.E. Milner, & D. Sanderson (eds.) Proceedings of the Ninth Biennial Conference and Symposia of the Canadian Society for Biomechanics (pp. 184-185). Simon Fraser University, Burnaby, British Columbia, Canada.

Wilson, D.G. Bicycling Science. Third Edition. The MIT Press. Cambridge, Ma.

2004.

http://www.asme.org/

http://www.asme.org/

American Society of Mechanical Engineers Human …web.cecs.pdx.edu/~far/me492/PDS document/Y2008... · [Type text] [Type text] [Type text] American Society of Mechanical Engineers

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