asme hpv 2011 report_california state university northridge

2011 ASME Human Powered Vehicle Challenge

California State University, Northridge

Design Report

Vehicle Number: Pending

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Table of Contents

Section Page

Abstract.3

Introduction..4

Design Description5

Analysis.12

Testing...22

Safety.27

Cost Analysis28

References.....30

Attachment #1Form 6 and 3-View Drawings........31

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Abstract

A Human Powered Vehicle (HPV) is an efficient, highly engineered vehicle that runs on

human muscle power. It can have everyday applicationsfrom commuting to work, to carrying

goods to market. The American Society of Mechanical Engineers International Human Powered

Vehicle Challenge (HPVC) provides an opportunity for students to demonstrate the application

of sound engineering design principles in the development of sustainable and practical

transportation alternatives. The mission statement of the California State University, Northridge

(CSUN) Human Powered Vehicle team is as follows: To design and manufacture a Human

Powered Vehicle that is practical, high performing, and furthers the art of HPV design. The

design goals were safety, speed, weight reduction, comfort, and ease of use.

The CSUN Human Powered Vehicle team has designed a tadpole recumbent vehicle.

The vehicle configuration is a fully faired recumbent tricycle, with two wheels in the front that

will provide steering and one single wheel in the rear that will provide forward motion. The

vehicle has been designed to be lightweight and very stable.

The primary material that was used to fabricate the body of the HPV was carbon fiber.

Compared to traditional materials such as, Aluminum 6061 and 302 Stainless Steel, carbon fiber

is lighter, more rigid and can have a yield strength that is comparable to steel when fabricated

correctly. Due to its high strength to weight ratio, it was an ideal choice of material to

manufacture a vehicle that will be powered by a human. Carbon fiber, which is stiff in both

tensile and compressive directions, due to its fiber weave orientation, acts as the reinforcement to

a structure when coated with epoxy resin.

Aerodynamic considerations were critical in designing the vehicle. To reduce the

aerodynamic drag force on the HPV, the team designed a full fairing with an integrated visor and

disc wheel covers. Two different configurations will be used for the sprint event and the

endurance event. In the sprint event, a low drag coefficient is of paramount interest; therefore,

the visor will be used in the vehicle configuration completely enclosing the rider and bicycle.

During the endurance event, the vehicle and rider are required to perform on a course with other

riders and sloping terrain for an extended period of time. Accordingly, heat removal, weight,

and visibility become prime considerations. The visor will be removed for this event allowing

forced convective heat transfer, visibility all around the rider, and a lowered system mass.

Multiple software packages were used to design and analyze the vehicle. The SolidWorks

CAD and Simulation package was the primary design and analysis tool used. It was used to

create a solid model of the entire vehicle assembly and perform Finite Element Analysis (FEA)

and Computational Fluid Dynamics (CFD) analysis. In addition to the SolidWorks Simulation

package, NEiWorks was used to analyze the structural belly-pan, which was complemented with

physical testing.

Various physical tests were performed to ensure the structural strength and safety of the

vehicle and to verify FEA. The tests that were performed were top and side load safety tests of

the roll bar, tensile testing of various layers of carbon fiber, abrasive testing and an adhesive

strength test to decide on the best adhesive to bond cured carbon to cured carbon.

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Introduction

A Human Powered Vehicle (HPV) is an efficient, highly engineered vehicle that runs on

human muscle power. It can have everyday applicationsfrom commuting to work, to carrying

goods to market. Furthermore, human powered transportation is often the only type available in

underdeveloped or inaccessible parts of the world. Well designed vehicles can be a valuable

form of sustainable transportation. By increasing mechanical advantage, human powered

vehicles afford the rider the benefits of increased range and shorter travel times.

The American Society of Mechanical Engineers International Human Powered Vehicle

Challenge (HPVC) provides an opportunity for students to demonstrate the application of sound

engineering design principles in the development of sustainable and practical transportation

alternatives. The 2010-2011 California State University, Northridge (CSUN) Human Powered

Vehicle team has designed a tadpole recumbent vehicle. The vehicle configuration is a fully

faired recumbent tricycle, with two wheels in the front that will provide steering and one single

wheel in the rear that will provide forward motion. The vehicle has been designed to be

lightweight as well as very stable and can be seen in Figures 1 and 2. Multiple software

packages as well as physical testing were used to design and analyze the vehicle. The

SolidWorks CAD and Simulation package was the primary design and analysis tool used. It was

used to create a solid model of the entire vehicle assembly and perform Finite Element Analysis

(FEA) and Computational Fluid Dynamics (CFD) analysis. In addition to the SolidWorks

Simulation package, NEiWorks was used to analyze the structural belly-pan, which was

complemented with physical testing.

Figure 1: Vehicle CAD Model with Fairing Figure 2: Vehicle CAD Model without Fairing

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Design Description

The mission statement of the 2010-2011 CSUN HPV team is as follows: To design and

manufacture a Human Powered Vehicle that is practical, high performing, and furthers the art of

HPV design. The design goals were safety, speed, weight reduction, comfort, and ease of use.

These criteria were used when deciding between recumbent and upright vehicle designs. Table I

displays the decision matrix created to evaluate the design concepts. The design goals were

assigned weights and the design alternatives were graded on a scale from 0 to 10, 10 being

excellent and 0 being failure. The recumbent 3 wheel design came up with the highest total

score, 7.3 out of a possible 10. The 3 wheel tadpole design allows for greater stability for the

rider. This design scores high marks for safety, comfort, and ease of use. This makes the vehicle

a viable choice for personal transportation.

Table I. Overall Design Concept Decision Matrix

Design Goals

Safety Speed Weight Comfort

Ease

of Use

TOTAL Design

Alternatives

Weighting Factors

0.25 0.30 0.20 0.15 0.10 1

Upright 6 7 9 4 9 6.9

Recumbent

(2 Wheel) 7 8 6 8 6 7.15

Recumbent

(3 Wheel) 10 5 5 10 8 7.3

The primary material that was used to fabricate the body of the HPV was carbon fiber.

Compared to traditional materials such as, Aluminum 6061 and 302 Stainless Steel, carbon fiber

is lighter, more rigid and can have a yield strength that is comparable to steel when fabricated

correctly. Due to its high strength to weight ratio, it was an ideal choice of material to

manufacture a vehicle that will be powered by a human. Carbon fiber, which is stiff in both

tensile and compressive directions, due to its fiber weave orientation, acts as the reinforcement to

a structure when coated with epoxy resin. Some of the general properties for each material

considered are listed in Table II.

Table II. Material Property Comparison

6061 Aluminum

Density

(lb/in3) Tensile Strength (psi) Yield Strength (psi) Brinell Hardness

0.0975 45000 40000 95

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302 Stainless Steel

Density

(lb/in3) Tensile Strength (psi) Yield Strength (psi) Brinell Hardness

0.29 90000 40000 150

Plain Weave Carbon Fiber

Density

(lb/in3)

Tensile Failure Stress

(psi)

Compress Failure Stress

(psi)

Tensile Modulus

(psi)

0.0636 511983 222000 33358679

There were three main structural components of the vehicle that were fabricated: the

belly pan, the frame, and the roll bar. Both the frame and roll bar were constructed by machining

out a mold from pink extruded foam as shown below in Figure 3. The molds purpose was to

create the geometry needed for each part and remained within the structure of the carbon fiber.

Figure 3: Pink Foam Mold of Frame Figure 4: Belly Pan Vacuum Bag Process

The fibers in the carbon fiber weave are oriented a specific way to allow for maximum

strength to occur in the direction of the fibers. Structural strength of the carbon became

maximized by following this pattern of orientation. Curing techniques for the carbon were the

most intricate part of the fabricating process. Both the roll bar and frame were cured using a

vacuum bagging technique. Figure 4 shows the belly pan vacuum bag process.

A structural belly pan was fabricated to provide a flat mounting platform to attach

components. The strength and ease of manufacturing made this design ideal. The belly pan was

constructed by sandwiching an aramid honeycomb core between 8 layers of carbon fiber

composites. This was one more layer than FEA recommended in order to maximize the

structures overall strength. The honeycomb core increased the compressive strength of the belly

pan while adding only minimal weight.

The dynamic components of the vehicle were broken down into two categories: standard

bicycle parts, and original manufactured components. The parts that were produced for this

project were the steering system, front wheel mounts, rear dropout brackets, and drive train path.

The decision matrix that assisted in determining the chosen designs for each category is shown in

Table III. Great emphasis was placed on manufacturability and how well parts would perform in

competition settings. The design goals were assigned weights and the design alternatives were

graded on a scale from 0 to 10, 10 being excellent and 0 being failure.

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Table III. Manufactured Parts Decision Matrix

Design Goals

Weight

Ease of

Manufacture Aesthetics Cost Performance

TOTAL

Design Alternatives

Weighting Factors

0.25 0.30 0.10 0.10 0.25 1

Steering

Wheel with Cable 6 2 9 6 6 5.1

Wheel with Shaft 3 3 9 3 8 4.85

Steering Levers 6 9 6 9 9 7.95

Cross-Pivot Levers 3 9 6 9 6 6.45

Wheel Mounting

Dual Headset 6 9 8 3 6 6.8

Hub and Spindle 9 6 9 9 6 7.35

Chain Path

Through Frame

Axle 3 9 9 9 6 6.75

Off-set axle 6 6 9 6 6 6.3

One Chain 6 9 3 9 3 6.15

When it came to sizing all of the manufactured parts, there were several parameters

considered. All of the parts had to have a minimum factor of safety of 2. Also, many of the parts

had to be designed to be compatible with other standard bicycle parts. To simplify the assembly

process, all of the bolts that assemble the manufactured parts are all the same. This allows team

members to use the same tools on all of the components of the vehicle. The primary material that

was used was 6061-T6 Aluminum. 6061- T6 Aluminum was chosen for its strength, low weight,

price, and availability.

While the axles were able to be made on a lathe, many of the custom parts had to be

made on mill or CNC machine. For some of the more simple parts, a Bridgeport 2 axis Accurate

Control mill was used. The complex parts that required more detail were cut on a Haas VF2, 3

axis CNC mill. The CNC parts were programmed using Mastercam which converts SolidWorks

part files into machine code. Most, if not all of these more complicated parts which would

normally take hours to machine could be produced in less than an hour. Overall, the total

machining time is extremely low considering the amount of machined parts being produced.

The steering system consists of steering levers attached to a hub and spindle setup. The

levers are able to rotate opposite each other and attach underneath the riders seat on an axle that

is mounted through the frame. The levers will then attach to a push-pull mechanism that serves

as the steering linkage. The movement of the push-pull mechanism is channeled to the spindle

through a bell-crank located at the center of the vehicle. The purpose of the bell-crank is to

convert the linear motion of the steering levers into rotational movement of the front wheels. The

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handlebars are made with 7/8 inch OD aluminum tubing; each side is bent to two 45 degree

angles, the first near the linkage system which connects to the steering rods located on the

underside of the HPV. The tubes are bent twice at 45 degrees instead of a single 90 degree bend

so that the tubes will not collapse; this will allow the aluminum steering arms to be much

stronger than if the tubes were to be bent with a single 90 degree bend. The steering handlebar

assembly is pictured in Figure 5.

Figure 5: Steering Handlebar Assembly

The front wheel hub and spindle design has a built in 10 of caster and 4 of camber in

order to give the vehicle excellent handling. Figure 6 shows the front wheel mounts. The edge of

the frame is used as a brace to prevent any displacement of the front wheels from the hub. The

spindle is a C-shaped design that fits around the hub similar to most vehicles. The camber and

caster work together to change the angle of the wheel when the vehicle is turned. These wheel

mounts were designed and built in-house, providing the ability for a custom design not

available in stores.

The rear dropout brackets shown in Figure 7 are another unique feature of the vehicle.

Most bicycles have the rear dropouts permanently fixed to the frame. This vehicle utilizes a two

piece rear dropout design. An aluminum plug with two tapped holes on both sides of the frame

remains permanently fixed. The actual dropout is a separate piece that is custom built and bolts

onto the permanent plugs.

Figure 6: Front Wheel Mounts Figure 7: Rear Dropout

For the drive chain path, a two-chain system was the best choice to meet the design

criteria. The two chains are connected with a transfer hub that mounts to an axle attached to the

frame (Figure 8). This design takes advantage of the carbon fiber frame by eliminating the need

for a separate structure that would be bulky and heavy. It also allows the chain to travel very

close to the frame. The transfer hub consists of a cassette mount designed for a regular rear

bicycle wheel. Instead of mounting the full 13 gear cassette, there is a 15 tooth, and two 20 tooth

gears that act as a power transfer between the front pedals and the rear wheel. The 15 tooth gear

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is intended for use in the sprint event. The 20-20 gear combination will provide a lower gear

ratio for use during the endurance event. A 10 speed cassette is used at the rear wheel to provide

a range of overall gear ratios for the most efficient power transmission from the rider to the rear

wheel.

Figure 8: Transfer Hub

The parts that were purchased were all parts common to bicycles that are readily

available from stores and would be impractical to manufacture. It was determined that the

purchase of these parts would make the project more time efficient and economical. Parts that

were designated for purchase include: the rear wheel, 10 speed cassette, tires, tubes, transfer hub,

and the front wheels. The front and rear wheels were custom made and laced by hand.

Air resistance can approach 90% of the total retarding force on a bicycle. Consequently,

aerodynamic considerations were critical in designing the vehicle, particularly for the sprint

event. The power required to overcome aerodynamic drag increases with velocity cubed,

limiting a riders top speed. Three principal ways of decreasing wind resistance applied to the

design of the aerodynamic include: decreasing the frontal area, streamlining bicycle components,

and smoothing the surfaces of the fairing, roll bar, and rider.

To reduce the aerodynamic drag force on the HPV, the team designed a full fairing with

an integrated visor and disc wheel covers. Design alternatives were evaluated by means of a

design matrix given below in Table IV. The design goals were assigned weights and the design

alternatives were graded on a scale from 0 to 10, 10 being excellent and 0 being failure.

Table IV. Aerodynamics Design Concept Decision Matrix

Design Goals

Frontal

Area

Low Drag

Coefficient

Heat

Dissipation Rigidity/Weight Cost

Ease of

Manufacture

TOTAL

Design Alternatives

Weighting Factors

0.35 0.20 0.20 0.15 0.05 0.05 1

Exposed rider, roll bar, &

enclosed rear wheel 10 6 10 10 10 10 8.7

Enclosed front wheels/Exposed

rider & roll bar 6 4 10 6 8 8 6.2

Enclosed rider & Roll

bar/Exposed disc wheels, event

specific 10 10 10 8 6 6 9

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With a score of 9 out of a possible 10, the best choice was determined to be an event

specific design. The sprint configuration shown below in Figure 9 (left) completely encloses the

rider and major vehicle components. This configuration will lead to an optimal drag coefficient.

The endurance configuration shown in Figure 9 (right) will not incorporate the canopy, leading

to a lowered system weight, increased ventilation and visibility.

Figure 9: Sprint configuration (left) and Endurance configuration without canopy (right)

For correctly streamlined fairing geometries, pressure drag is minimized. In order to

minimize boundary layer separation over the surface, a long profile of 103 inches was modeled

using the NACA 0015 profile generated in Excel and imported as guide curves within the

SolidWorks environment (Figure 10). Using curvature combs, the computed airfoil data was

modified to meet rider geometry and vehicle dimensions. Particularly, the widest point moved

back towards the rear of the vehicle.

To facilitate the fabrication of the fairing skin, visor and stiffeners, a male plug was

fabricated out of a 2ft x 4 ft x 8 ft solid block of 1 lb density expanded polystyrene foam (EPS).

An initial rough cut, shown in Figure 11, formed by using a hot wire 3-axis CNC machine, was

followed by hand carving to produce the streamlined shape of the fairing. Final symmetry was

checked using cross sectional area templates. Areas of concern, in particular the pedal box,

shoulder, and head clearance were checked for conformance. Surface filling followed by surface

sealing stages were then applied to the foam sculpture. A joint compound was applied to the

entire surface to fill in all gaps and surface voids then sanded as shown in Figure 12. Two coats

of a water based primer (StyroPrime) followed by several coats (3-4) of a liquid plastic

(StyroSpray) supplied by Industrial Polymers were applied to the foam surface. This barrier

allowed for the application of a polyprimer (PLC). The entire surface was then sanded, cleaned,

and sprayed with an epoxy resin. A mold release was then applied (5 layers) in preparation for a

wet layup.

Figure 10: Surface geometries governing

frontal area reduction

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Figure 11: Rough cut foam block shown with inch routed carving templates (left), hand

carving to final shape (right)

Figure 12: Filling in all gaps and surface voids (left), surface sealing stage (right)

The wet layup consisted primarily of an inner carbon fiber layer and an outer hybrid

carbon/Kevlar layer. Once the initial shell had cured, additional hybrid fiber reinforcement

layers were bonded into the tub and canopy components at specified locations, near the rider,

interfacing with the road during rollover. A framing rib structure layup consisting of two layers

of carbon fiber and Lantor Soric XF foam was fabricated. The layup components were chosen to

increase stiffness and reduce overall weight, due to lower resin retention, when compared with

equivalent all fiber lay-ups. Following alignment of the tub and canopy skins, the ribbing

structure was bonded to the inner surfaces of each.

A female fiberglass tool was fabricated to drape form lexan for a canopy visor. The tool

was build with a shoulder to allow the lexan to be formed with a corresponding shoulder to

ensure a flush mate between the visor and canopy skin when bonded. Aluminum sheet metal

brackets were fabricated and bonded to the inner surface of the tub at specified locations to

mount the tub to the frame.

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Analysis

The rollover/side protection component of the design (Roll bar) consists of a single,

continuous carbon fiber feature which sweeps around the riders seat and head (Figure 13). The

roll bar is rigidly and directly attached to the main body of the vehicle (frame-belly pan) on the

mounting faces. The integrated design of the belly pan, frame and roll bar has provided a single

solid and strong base structure for mounting the rest of the components such as wheels, crank,

seat, etc.

Since the strength of the roll bar is of paramount importance in any roll over scenario,

extensive finite element analyses (FEA) have been

conducted to ensure the integrity of the structure in these

situations. Using NEiWorks composite capabilities and

through an iterative design optimization process, different

carbon fiber layups have been examined and optimized to

achieve minimum weight, minimum material waste and

maximum strength based on ASME requirements.

First, in a series of static analyses the roll bar has

been subjected to 800 lb top (12 degrees from vertical) and

500 lb side loads with the objective of finding the optimal

layup as stated above. With the roll bar held fixed on the

mounting faces, the experimentally obtained material

properties, the Hill composite failure criterion and the

allowable bond shear stress of 1000 psi (determined by

numerous testing samples) were used to define the boundary

conditions and computational constraints of the analyses.

Figure 14 presents the optimized carbon fiber layup obtained with the static analyses.

This optimized layup (left in Figure 14) is an 8 layers carbon fiber sandwich consisting of uni

and plain weave fibers with a horizontal wrap direction (right in Figure 14) following the shown

stack up layup.

Figure 14: The optimized carbon fiber layup of the roll bar (left) along with the wrap direction (right)

Mounting faces

Figure 13: Integrated components of the vehicle

including the roll bar, frame, and belly pan

Wrap direction

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The displacement and composite status of the optimized model for the two load scenarios

are shown in Figures 15 and 16. The maximum displacements obtained with the optimal layup

were 0.96 inch and 0.87

inch for the side and top

loads, respectively.

Furthermore, the composite

failure indexes reported in

both cases indicated the

healthy status of the

structure.

The static analyses

have then been extended to

dynamic (impact) analyses

for further examinations in

order to account for a more

realistic roll over scenario.

Since it is expected that in the case of a roll over the roll bar would initially come in

contact with the ground on its sides, a side collision has been considered for the impact analysis.

To simulate this side impact in the software (Figure 18) the roll bar has been carefully shot with

initial velocity of 25 miles/hr in a straight line trajectory from a distance of 1 ft to hit a plane

which is simulating the ground.

Upon examining Figure 17 and Figure 18 it can be seen that the obtained optimal layup

has brought about not only a healthy

structure (failure indices less than one) but

also an acceptable (ASME rule) side

deflection in case of the real impact. The

magnitude of this deflection at the roll bars

shoulder (denoted by node 628) was 0.93

inch.

Figure 18: The roll bar side impact composite max failure

index results

Figure 17: The Roll bar side impact displacement results of node 628

Figure 15: 800 lb top load displacement

results

Figure 16: 500 lb side load displacement

results

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The frame-belly pan is an assembly of two components which are manufactured

separately but are bound together to form the base skeleton of the vehicle. The frame itself is

holding the crank housing on the front while supporting the back wheels directly on its rear

handles. The belly pan on the other hand serves as a mounting structure for the front wheels

while acting as a stress reliever for the frame. The geometry (Figure 19) has been designed as

narrow as possible to decrease the surface area and reduce excessive flexing in X and Y

directions. Following a roll bar-like approach, an efficient carbon fiber layup was developed

through a system of static and dynamic analyses

while taking into account the stress-generated

failures that may occur in the critical segments of

the two components.

Following this procedure the analyses were

resolved into crank housing, structure flexing and

wheel supports.

The crank housing, which involves the

frame only (Figure 20), has been manufactured by

inserting the crank insert (an Aluminum tube) into

the pre-cut opening in front of the frame which has

then been wrapped and positioned with multiple

layers of carbon fiber. This analysis was concerned

with determining the sufficient amount carbon fiber

layers required to secure the crank in place and

avoid any failure.

The failures may occur due to the coupled

loads that are exerted during the pedaling.

Following the EN 14766 (European Mountain Bicycles

Committee) recommendations, the crank housing should

withstand a 400 lb couple load (Figure 20). The frame

boundaries were fixed sufficiently far from the crank

housing to provide a valid solution in the region of interest.

After several iterations, the analyses have

determined that 8 layers of plain carbon fiber with a layup

schedule of (0/+45/0/0/0/+45/0/0) would be capable of

capturing stress disturbances in the crank housing. The

simulation results of the segmental body of the frame

(Figure 20) showing total displacement and composite

failure indexes are presented in Figures 21 and 22. They were determined by implementing the

optimized layup stated above.

From Figure 21, it is seen that the total displacement reported is 0.005 inch illustrating the

strength and rigidity of the crank housing. As for the composite failure index result, it also

indicates a healthy status for the crank housing.

Figure 19: Frame-belly pan assembly with critical segments

Figure 20: The crank housing

Crank Housing

Seat Location

Back Wheel Supports

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Once the optimal layup for the crank housing was determined, the same layup was

applied to the entire frame to quantify the amount of flexing present in the X and Y directions.

To avoid boundary condition related inaccuracies, the belly pan was rigidly connected to the

frame, constituting a single continues model. Consequently the analysis has also aimed to define

an optimal layup for the belly pan to help the frame in increasing the overall vehicle rigidity. The

structural flexing can occur either due to the static loading of the riders weight or the dynamic

loading of a rough surface contact (such as a bump), along with the cyclic pedaling loads at the

cranks.

To simulate the effect of these three loads simultaneously while reducing the

computational resources, a procedure presented in Figures 23a-23e has been followed to

efficiently replace the complexity of a dynamic analysis with a conservative static analysis. To

initiate this approximation, a separate static analysis (including only the riders weight) was

performed. Next using the resulting displacements, an equivalent stiffness (Keq) has been

calculated for the frame and belly pan combined. At this point the frame-belly pan assembly has

been mathematically modeled by a single mass at the center of gravity attached to the equivalent

spring (Figure 23a). Using this simplified model, an arbitrary load (representing the surface

bump) has been approximated with a conservative static load (P0) as shown in Figure 23b. The

value of this conservative static load was determined based on a scenario where the vehicle

experiences a 3 inch high bump. It can be seen from Figure 23c that the spring will be displaced

by less than 3 inches at the top of the bump. Assuming the worst case scenario of the full 3

inches displacement of the spring, P0 was determined (using the Keq and displacement) to be

1.5xWrider where Wrider is the average riders weight. Now that a single known static equivalent

load is determined to approximate the dynamic loads of the road the frequency spectra chart

(Figure 23d) was used to take into account the impulse (short period) nature of this load. It can

be seen from the chart (shown in red) that this conversion can be done by multiplying the P0 by

the magnification factor of two. Collectively it could be said that: P0dyn = 3x Wrider = 700 lb.

Applying the obtained force at the CG and accompanying that with the 400 lb pedaling force (as

discussed in the crank housing analysis) a full loading scenario of the structure was achieved. In

addition the structure was fixed at the wheel locations (front and back) as shown in Figure 23e.

Figure 22: The composite failure index result of the crank

housing

Figure 21: The displacement results of the crank housing

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The displacement and composite status results obtained from these analyses showed that

the same layup schedule introduced in the crank housing could be extended to the entire body of

the frame. Furthermore it was determined that the optimal layup for the belly pan includes a 0.75

inch Nomex honey comb core which is sandwiched between 7 layers of plain weave carbon

fiber. Figures 24 and 25 summarize the belly pan and frame final layup schedule along with their

zero angle (wrapping) directions.

Figures 23a through 23e show the process of defining the dynamic loading of the belly pan and frame

Figure 24: The layup schedule (left) and the wrap direction (right) of the belly pan

Figure 25: The layup schedule (left) and the wrap direction (right) of the frame

Figure 23e

Figure 23a

Figure 23b

Figure 23c Figure 23d

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The simulation predictions of the

deflection of the frame-belly pan assembly in Y (up

and down) and X (side to side) directions

performed with the optimized layups are presented

in Figure 26 and Figure 27. From the displacement

results in Y and X (0.09 inch and 0.10 inch,

respectively) it is seen that the profile flexing of the

structure is nearly negligible.

To further investigate the results obtained

for the frame optimized layup, a separate

simulation was employed on the end segment of

the frame whose rear supports are holding the back

wheel.

Unlike the previous analysis (see Figure 23e)

where the rear supports were held fixed, this time

they were loaded directly by two vertical forces

(same on each support) whose magnitudes were

determined by taking into account the reaction

forces produced by P0dyn (see Figure 23e).

It has been determined that approximately

2/3 of this load (400 lb) is transferred to the rear

supports (Figure 28). By changing the boundary

conditions to include the roll bar in the analysis, and

fixing the structure far from the rear supports, each

component was assigned with its previously

obtained optimal layup.

The first round of analysis indicated a small

de-lamination at the area near the end of the right

support based on the reported composite failure

indices (Figure 29).

Figure 29: Shows the small de-lamination at the right handle

Figure 26: The displacement results of the structure in Y

direction

Figure 27: The displacement results of the structure in X direction

Figure 28: Shows the

rear support FEA model

along with 200 lb loads on

each support

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To solve the problem the original frame layup was locally improved via adding two extra

layers of carbon fiber. The subsequent analysis presented in Figure 30 confirmed the effect of the

imposed correction.

Figure 30: Shows the healthy status of the handles after addition of extra layers

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Using SolidWorks flow simulation to perform

flow analysis on the solid model of the fairing, the

design team optimized the fairing geometry. The flow

simulations have aimed at predicting the fairing (and

eventually the full model) drag coefficient, defined as:

Where F is the fairing drag force, A is the

fairing frontal area; is the air density and V is the

vehicle velocity.

To set up these simulations, an incoming

uniform air flow of 40 mile/hr, temperature of 59F,

and turbulence intensity of 1% have been specified

at the boundaries of the computational domain.

This computational domain have further been

equipped with robust mesh refinement and

extended end tail (Figure 31) to increase the

accuracy of the results.

Through an iterative design optimization

process the fairing geometries have been

improved at each step with the objective of

achieving a streamlined (low pressure drag)

shape. Figure 32 and 33 present the 6th

and 8th

(final) iterations along with their flow patterns and

pressure fields. The rise- drop- rise of the pressure

fields at the nose-shoulder -tail of the models are

in full accordance with the theoretical

expectations of a streamlined shape. However the

steady stream of the flow over and past the final

design has formed a more moderate nose-tail

pressure gradient resulting in a lower drag

coefficient of 0.06. This pressure gradient is

shown in detail with a pressure profile graph

extending from the tip to the back of the fairing

(Figure 34). It is seen that the nose-tail pressure

drop governing the fluid flow for the final fairing

becomes nearly recovered (0.014 psi), reducing

the pressure drag to a great extend. In fact an

approximate 57% improvement in drag

coefficient from iteration 6 to the final iteration 8

was realized as tabulated below in Table V.

Figure 32: 6th iteration along with its flow trajectory and pressure field

Figure 31: The computational extended region and mesh refinement

Figure 34: The pressure gradient profile of final fairing

Figure 33: 8th iteration (final) along with its flow trajectory and pressure field

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Table V. Aerodynamic Drag Data

Configuration Drag Coefficient -Cd Drag Force (lb) Frontal Area-AF (in2)

Analysis of Fairing ONLY

Iteration 6-

Sprint

Configuration 0.094 2.24 841.35

Iteration 7-

Sprint


Iteration 8-

Sprint


Analysis of Full Model

Iteration 8-

Sprint

Configuration

(Full Model) 0.073 1.44 1035.9

Iteration 8-

Endurance

Configuration

(Full Model) 0.209 5.64 948.36

Finally to achieve the true magnitude of the drag coefficient for the final fairing, a full

model simulation was considered by including all the exposed components of the vehicle.

Results for the full model analysis are tabulated in Table V. From sprint configuration in Figure

35 (left), it can be seen that the frontal area has increased to 1035.9 in2 and the drag coefficient

has also increased to 0.073. These values are still smaller when compared to the other shapes

considered. Figure 35 (right) shows the endurance configuration which has a drag coefficient of

0.209 and a frontal area of 948.36 in2.

Figure 35: Flow trajectory and pressure field for sprint configuration (left) and endurance configuration (right)

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To substantiate the effectiveness of the fairing design, a rough evaluation of the expected

performance at constant velocity, no winds, level ground, a conservative rolling resistance

coefficient of 0.008, overall drag coefficient of 0.073, negligible transmission loss, and typical

system mass of 88 kg (based on expected vehicle mass of 27 kg and rider mass of 61 kg) was

input to the Power equation given below.

: projected frontal area, (m)

: aerodynamic drag coefficient : rolling resistance, typically .003 < < .006 at 10 m/s G: grade, percent divided by 100 (zero for this case)

M: total system mass (kg)

P: cyclist power, (W)

t: time, (s)

V: velocity, (m/s)

: wind velocity, head or tail winds (zero for this case), (m/s)

: acceleration of the vehicle, (zero for this case)

W: weight of the system, cyclist and machine, (N)

: air density, (kg/m) at standard conditions

Based on the idealized model, it can be noted that for a stipulated top speed goal of 40

mph, the power input required is 376 Watts using the fully enclosed sprint fairing configuration.

Under the same parameter values, using the endurance fairing configuration, not including pit

stops, the team can expect an average velocity of 25 mph corresponding to 62 miles of distance

covered for the 2.5 hour duration of the endurance event based on the measured average rider

power input of 206 watts.

To maximize performance during the endurance event, research has shown that the rider

must release three units of heat for every unit of power input to the pedals. With an expected

power input of 200 watts, 600 watts of heat must be removed to avoid rider discomfort and

significant reduction in efficiency. Due to the riders head being exposed to air flow, it was

calculated that 100 watts may be removed by convective heat transfer. The calculated radiative

heat loss through the fairing was 100 watts. To facilitate heat loss due to sweat evaporation and

natural convection over the remainder of the riders body, a NACA duct was fabricated and

installed in front of the visor.

Refer to page 28 for a detailed cost analysis of the vehicle.

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Testing

Various physical tests were performed to ensure the structural strength and safety of the

vehicle and to verify FEA. The tests that were performed were top and side load safety tests of

the roll bar, tensile testing of various layers of carbon fiber, abrasive testing and an adhesive

strength test to decide on the best adhesive to bond cured carbon to cured carbon.

The roll bar test was performed to ensure that the impact of a possible crash would not

endanger the safety of the rider. The structural strength of the roll bar, specified by ASME,

should have a top load requirement that can with withstand a 600 lb load at an angle of 12

degrees with respect to the vertical axis and a deflection of less than 2.0 inches. The test setup,

shown in Figure 36, mounted the roll bar to a metal frame with heavy duty straps. The top of the

roll bar was pulled at an angle downward of 12 degrees from the vertical by a hoist. The hoist

was attached to a scale at the other end, which digitally read the tensile force in pounds. At a

force of 670 lb with a minimal and non-measurable deflection, the top load test passed the test

requirement. When compared to FEA, the deflection was less than the predicted value of 0.87

inches at an 800 lb load.

Figure 36: Top load roll bar test setup

The side load test, shown in Figure 37 and Figure 38, was setup by applying heavy duty

straps and clamps to the bottom of the roll bar to restrain it from movement. A second set of

straps were tied to the side of the roll bar, at shoulder height, and attached to a hoist. The other

end of the hoist was attached to a scale which read the tensile force as the roll bar was pulled

horizontally to the side. The ASME rules specify that the roll bar should withstand a minimum

force of 300 lb to the side of the roll bar at shoulder height with a deflection less than 1.5 inches.

The side load test passed its specification requirement with a load of 306 lb and a deflection of

1.1 inches. However, this deflection was slightly higher than the FEA prediction of 0.96 inches.

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Figure 37: Side Roll Bar Setup Figure 38: Side Roll Bar Result

Tensile tests were performed on various samples of carbon fiber, each with a specific

number of layers. This testing was conducted to verify the material strength properties of the

carbon fiber used in FEA. Each sample was layered at a zero and forty-five degree fiber

orientation to mimic the same process performed when fabricating each component of the

vehicle. The test setup consisted of creating carbon samples with a known neck length and

clamped into a tensile machine fixture as shown in Figure 39 and Figure 40. The results and data

for each sample are outlined in Table VI below.

Figure 39: Tensile Test Setup Figure 40: Tensile Samples after Test

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Table VI. Tensile Testing Final Results

Tensile Testing

# of Carbon

Layers

Gauge Length

(in)

Peak Load

(lb)

Width

(in)

Thickness

(in)

Area

(in2)

Stress

(psi)

1 1.5 257.11 1 0.012 0.012 21425.83

3 2.75 676.04 1 0.035 0.035 19315.43

4 3 1604.84 1.25 0.048 0.060 26747.33

6 5.75 2499.61 1.25 0.053 0.066 37729.96

8 6 3499.45 1.25 0.070 0.088 39993.71

It was shown that as the number of layers of carbon increase, the peak stress prior to

failure increases as expected. At 8 layers, as shown in Table VI, there is sufficient strength to

handle an ultimate yield stress of 39,993.71 psi. The stress values found matched closely with the

material properties used for the carbon fiber for FEA.

Abrasion testing was performed to experimentally determine the best possible fairing

material to resist abrasion in case of a vehicle roll over. Two procedures were performed in

order to experimentally determine the best possible material. In the first procedure, a special

fixture, shown in Figure 41, was built and four test samples were attached to the apparatus shown

in Figure 42.

Figure 41: Abrasion Fixture Figure 42: Abrasion Test Setup

The test samples were subjected to abrasion by dragging the test sample for 100 ft from

rest against asphalt, in order to determine the change in mass after abrasion. The material with

the lowest change in mass after this abrasive process was considered best. The second procedure

was called a skid test, consisting of replicating an initial velocity of 20 mph, with the fixture

shown in Figure 41. The samples were released with this initial velocity and the distance they

traveled before coming to a complete stop was measured. The material with the lowest stopping

distance was deemed as being best because the sliding distance after a rollover would be

minimized.

Table VII shows the results from the abrasion test. It can be seen that the carbon/Kevlar

samples had the lowest change in mass of 0.0519 g, 0.0634 g, and 0.0486 g, respectively, before

and after the experiment.

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Table VII. Abrasion Test Results--Dragged for 100 ft from rest at 20 mph

Test

# Material/Configuration

masso (g)

massf

(g)

m

(g)

l

(in)

w

(in)

Area

(in2)

1 Plain weave Carbon Fiber 0o

40.9651 40.8523 0.1128 2.5 2.5 6.25

2 Plain weave Carbon Fiber 0o

41.6812 41.5924 0.0888 2.5 2.5 6.25

3 Plain weave Carbon Fiber 45o 39.4177 39.3147 0.103 2.5 2.5 6.25

4 Plain weave Carbon Fiber 45o 41.2423 41.1422 0.1001 2.5 2.5 6.25

5 Kevlar plain weave #549 39.986 39.9057 0.0803 2.5 2.5 6.25

6 Kevlar plain weave #549 40.4099 40.3357 0.0742 2.5 2.5 6.25

7 Kevlar plain weave #2469 (smaller fibers) 40.3965 40.326 0.0705 2.5 2.5 6.25

8 Kevlar plain weave #2469 (smaller fibers) 41.6584 41.544 0.1144 2.5 2.5 6.25

9 Kevlar Twill weave 40.1187 40.0434 0.0753 2.5 2.5 6.25

10 Kevlar Twill weave 40.5889 40.5183 0.0706 2.5 2.5 6.25

11 Carbon/Kevlar hybrid (Kevlar direction of travel) 40.5306 40.4469 0.0837 2.5 2.5 6.25

12 Carbon/Kevlar hybrid (Kevlar direction of travel) 42.0082 41.9596 0.0486 2.5 2.5 6.25

13 Carbon/Kevlar hybrid (Carbon in direction of travel) 40.3251 40.2732 0.0519 2.5 2.5 6.25

14 Carbon/Kevlar hybrid (Carbon in direction of travel) 42.3793 42.3159 0.0634 2.5 2.5 6.25

Table VIII shows the results from the skid test. It was shown that the plain weave carbon

orientated at 45 degrees stopped with the shortest distance of 15.25 ft. However, the

carbon/Kevlar samples also performed well on this test. Based on the results of both tests, the

carbon/Kevlar was chosen for the fairing.

Table VIII. Skid Test Results--Initial velocity from 20 mph to a complete stop

Material/Configuration Stopping Distance (ft)

Plain weave Carbon Fiber 0o

18.917

Plain weave Carbon Fiber 45o 15.250

Kevlar plain weave #549 17.250

Kevlar plain weave #2469 (smaller fibers) 16.208

Kevlar Twill weave 17.875

Carbon/Kevlar hybrid (Kevlar in direction of travel) 16.167

Carbon/Kevlar hybrid (Carbon in direction of travel) 17.667

Adhesive testing was performed to experimentally determine the best adhesive to use for

bonding cured carbon to cured carbon. This was performed during fabrication when bonding the

roll bar to the frame and the frame to the belly pan. The ideal material needed to be able to

withstand maximum force without experiencing shear failure. A total of nine test samples,

shown in Figure 43, Figure 44, and Figure 45, of carbon fiber were created and bonded with

three different types of adhesives: epoxy resin, Hysol, and 3M DP-420.

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Figure 43: Epoxy Sample Figure 44: Hysol Sample Figure 45: DP-420 Sample

A tensile testing machine was used to accurately determine the peak load prior to shear

failure. The ultimate stress before failure for each sample was determined by using the peak load

and dividing it by the measured bond cure area. A total of three samples were used to find the

average peak load per adhesive. The average peak stress for DP-420, epoxy resin, and Hysol

were calculated to be 1852 psi, 1376 psi, and 2583 psi, respectively. These results are tabulated

in Table IX. Ultimately, Hysol was chosen due to its high stress tolerance.

Table IX. Adhesive Test Results Sample Run Length (in) Width (in) Area (in

2) Peak Load (lb) Peak Stress (psi)

DP-420 1 0.562 0.598 0.336 436 1297

DP-420 2 0.573 0.563 0.323 652.3 2022

DP-421 3 0.530 0.649 0.344 770 2238

Average 619.333 1852

Std Dev 169.275 493

Epoxy 1 0.507 0.592 0.300 544 1812

Epoxy 2 0.534 0.576 0.308 310 1008

Epoxy 3 0.528 0.580 0.306 463.78 1307

Average 439.26 1376

Std Dev 118.911 407

Hysol 1 0.708 0.594 0.421 1006 2392

Hysol 2 0.665 0.579 0.385 1108.475 2879

Hysol 3 0.678 0.582 0.395 978.006 2478

Average 1057.188 2583

Std Dev 72.531 260

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Safety

A recumbent vehicle design provides many safety benefits to the rider. Greater safety is

possible because of the near impossibility of taking a header over the front wheel or of

catching a foot or pedal on the ground when cornering. There is far greater comfort in an almost

complete absence of pain or trauma in the riders hands and wrists, or back and neck, or crotch.

This design also allows better visibility forward and to the side for the rider compared with that

for a diamond-framed bicycle with dropped handlebars. For safety purposes, the vehicles

canopy was designed to give the rider at least 180 degrees of visibility during the sprint event.

During the endurance event, the canopy will be removed, thus giving the rider an even greater

range of visibility.

As the case with any vehicle, certain precautions need to be taken to ensure that all of the

riders of the recumbent vehicle and other competitors are safe. First and foremost is having a

seatbelt since the vehicle requires the rider to lie in a reclining position. The design incorporates

a simple lap belt that will wrap around the frame and secure the ride comfortably within the seat.

Brakes are required on any vehicle to prevent any unwanted contact with surrounding

vehicles or obstructions. The vehicle incorporates two different brake systems. The front brakes

are cable driven and are to be used by the rider while riding. The rear brake will be used as a

parking brake that will assist the rider while getting in and out of the vehicle. Recumbent

vehicles can be difficult to get in and out of and this will make driver changes safer, quicker, and

more efficient.

All nuts and bolts will be checked prior to each event in order to maintain their security.

Safety wire and Loctite are two of the most common fastener security techniques and will be

utilized to ensure that no mechanical parts become loose. Since many of the vehicle parts will be

custom made out of aluminum, it is important that there are no sharp edges or loose chips. Each

machined part has been cleaned up using files and deburring tools. Similarly, all open end tubes

are capped using either plastic plugs or Delrin plugs. Certain areas with a high potential for

pinching, such as the chain and gears, must be evaluated for safety. Most of the pinch points are

behind or below the rider so they are of minimal concern. The key area is the chain that runs

down between the riders legs. A chain guard is incorporated to prevent any leg contact.

of 33

Cost Analysis

This section provides a summary of cost analysis for this project. In order to properly

illustrate the overall cost of a single quantity production of a human powered vehicle, each and

every component of the vehicle must be properly tabulated to have a detailed cost analysis. Mass

production costs for this vehicle will also be included in this section since this method is more

cost efficient. The tables below provide the division of labor cost, material cost, and overhead

equipment cost. The cost for the utilities and production facility was not included in this analysis

since this varies with location.

Table X. Labor Cost Estimate

Labor Quantity

Hourly

Rate

Hours per

Week

Cost per

Week

Cost per

Month

Design Engineer 3 $28.00 40 $3,360.00 $13,440.00

Project Supervisor 1 $36.00 40 $1,440.00 $5,760.00

Carbon Fiber Team 3 $15.00 40 $1,800.00 $7,200.00

Machinist 1 $20.00 40 $800.00 $3,200.00

Welder 1 $18.00 40 $720.00 $2,880.00

Assembly Team 3 $15.00 40 $1,800.00 $7,200.00

Total Cost $9,920.00 $39,680.00

Table XI. Material Cost Estimate

Individual Vehicle Cost 10 Vehicles Cost

Materials Quantity Unit Cost Total Quantity Total

Carbon Fiber (1/32'') 14 $65.00 $910.00 140 $9,100.00

G-10 1 $85.00 $85.00 10 $850.00

Fairing Mold Foam 1 $350.00 $350.00 1 $350.00

Aluminum 3 $700.00 $2,100.00 30 $21,000.00

Composite Supplies 1 $120.00 $120.00 10 $1,200.00

Epoxy Resin 1 $95.00 $95.00 10 $950.00

Adhesive 1 $80.00 $80.00 10 $800.00

Nomex Honeycomb 1 $650.00 $650.00 10 $6,500.00

Subsystem Components 1 $1,200.00 $1,200.00 10 $12,000.00

Total Cost $5,590.00 $52,750.00

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Table XII. Overhead Equipment Cost

Manufacturing Equipment Quantity Cost Total

TL-2 CNC Lathe 1 $27,000.00 $27,000.00

Vacuum Bagging Pump 1 $470.00 $470.00

Computer/Software 3 $1,000.00 $3,000.00

Welding Machine 1 $3,200.00 $3,200.00

CNC Milling Machine 1 $22,000.00 $22,000.00

Tools 1 $1,500.00 $1,500.00

Total Cost $57,170.00

Table XIII. Cost Analysis Summary

Cost of Single Vehicle 6 Year Production-10 Vehicles Per Month

Materials $5,590.00 $4,024,800.00

Labor $9,920.00 $7,142,400.00

Overhead Equipment $57,170.00 $57,170.00

Total Cost $72,680.00 $11,224,370.00

Total Cost Per

Vehicle $72,680.00 $15,589.40

As Table XIII shows, the cost per vehicle is greatly reduced as the number of vehicles

being produced increases. This is a direct result of the high cost of overhead equipment. For a

production run of 6 years, assuming 10 vehicles a month are produced, the total cost per vehicle

is $15,589.40. However, if only one vehicle is produced, the cost is $72,680.00. It should be

noted that the vehicle presented at competition does not cost nearly this much. All overhead

equipment was provided and the students were not paid for their labor. Therefore, the only costs

incurred were the material costs.

of 33

References

"ASME - Human Powered Vehicle Challenge (HPVC)." ASME - Home. Web.

.

CyclesPublished Standards. European Committee for Standardization. 2005. Web.

.

Kyle, M. W. "Aerodynamics of Human-Powered Vehicles." Proceedings of the Institution of

Mechanical Engineer. Part A, Journal of Power and Energy (2004): 141-54. Print.

Ryan, R. "Fluid Dynamics." Dec. 2011. Lecture.

Salary.com - Salary Information, Job Search, Education Opportunities and Career Advice.

Web. .

Shields, D. "Sculptor/Propmaker Teaching Session." Personal interview. 31 Dec. 2011.

Wilson, David Gordon, Jim Papadopoulos, and Frank Rowland. Whitt. Bicycling Science.

Cambridge, MA: MIT, 2004. Print

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2011 Human Powered Vehicle Challenge West Sponsored by ASME and Montana State University

Form 6: Vehicle Description

Due April 11, 2011

(Dimensions in inches, pounds)

Competition Location: Montana State University

School name: California State University, Northridge

Vehicle name: Bicycle Engineering @ Northridge (BE@N) Vehicle number : Unknown

Vehicle type Unrestricted Speed__X_____

Vehicle configuration

Upright Semi-recumbent X

Prone Other (specify)

Frame material Carbon fiber composite

Fairing material(s) Carbon fiber with Kevlar fabric sandwiched

Number of wheels 3

Vehicle Dimensions

Length 103.08 in Width 37.56 in

Height 53.28 in Wheelbase 37.56 in

Weight Distribution Front 60% Rear 40% Total 100%

Wheel Size Front 20 in Rear 27.5 in

Frontal area Sprint Configuration: 1035.9 in2 and Endurance Configuration: 948.36 in2

Steering Front _X Rear

Braking Front Rear Both __X

Estimated Cd 0.073

Vehicle history (e.g., has it competed before? where? when?)

This vehicle is a clean sheet design and has not competed in any event.

of 33

Attachment 1Figure 1: 3-View Drawing Without Fairing (dimensions shown in feet)

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Attachment 1Figure 2: 3-View Drawing With Fairing (dimensions shown in feet)

asme hpv 2011 report_california state university northridge

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