proceedings of the asme 2015 international mechanical

GRADUATION DESIGN PROJECT: DESIGNING AND BUILDING A PIN-ON-DISK TRIBOMETER

Elspeth Ochs1 Mechanical Engineering Department

Rochester Institute of Technology, Rochester, NY, USA

508-360-3699, [email protected]

Patricia Iglesias Victoria Mechanical Engineering Department

Rochester Institute of Technology, Rochester, NY, USA

ABSTRACT

The purpose of this paper is to describe an independent

study project required before graduation with a master of

engineering degree in the department of mechanical

engineering at Rochester Institute of Technology. The goal of

this graduation project was to design and build a Pin-on-Disk

Tribometer. The tribometer will be used by the department for

future student research projects. This paper includes the details

of the design process, such as adaptations from existing

tribometers, all required calculations, and descriptions of the

machining and assembly. New design concepts presented

demonstrate simplicity of construction and use. The required

calculations include shear and bending moments in the arm,

bearing calculations, lead screw load analysis, counterweight

balancing, torque calculations for selecting the correct motor,

and strain calculations for determining the correct strain gage

set-up. The paper details part selection, pricing, machining, and

assembly. Mistakes and alternate design considerations are

also discussed.1

NOMENCLATURE

The nomenclature for this paper includes:

C dynamic load capacity (bearings) [N]

Cl critical load (lead screw) [N]

Cs critical speed (lead screw) [RPM]

D diameter of scratch (pin) [m]

E Young’s modulus [Pa]

FA force at point A/pin (arm) [N]

FB force at point B/bearing pivot (arm) [N]

FC force at point C/counterweight (arm) [N]

FN normal force (pin) [N]

FR reacting force (pin) [N]

1 Elspeth now works full-time at UTC Aerospace Systems.

I second moment of inertia (arm) [m4]

J Polar moment of inertia [kg-cm2]

L length between supports (lead screw)[m]

L10 bearing life with 90% reliability (bearings) [cycles]

M moment around neutral axis [N-m]

M‹x› third order singularity equation (arm) [N-m]

N fixity (lead screw) [--]

P applied load (arm) [N]

R rotating speed (motor) [RPM]

Tfric torque due to friction (motor) [Nm]

Trod torque due to rotation (motor) [Nm]

V‹x› second order singularity equation (arm) [N]

Varm arm volume (arm) [m3]

VCW counterweight volume (arm) [m3]

Vpin pin volume (arm) [m3]

Vrod1,Vrod2 volume of threaded rods holding applied

weight and counterweight (arm) [m3]

W load [kg]

b base width (arm)[m]

c centroid (arm) [m]

d root diameter (lead screw) [m]

e ball bearing constant (bearings)[--]

g gravity [m/s2]

h height (arm) [m]

l1, l2 length of arm, length of threaded rod (arm) [m]

m1, m2 mass of arm, mass of threaded rod (arm) [m]

q‹x› first order singularity equation (arm) [N/m]

t time (motor) [s]

x distance from front end of arm (arm) [m]

xarm distance from arm pivot (arm) [m]

xCW distance from counterweight (arm) [m]

ε strain (strain gage) [--]

μk coefficient of kinetic friction [--]

ρ density [kg/m3]

Proceedings of the ASME 2015 International Mechanical Engineering Congress and Exposition IMECE2015

November 13-19, 2015, Houston, Texas

IMECE2015-51148

1 Copyright © 2015 by ASME

INTRODUCTION

Every year, much of the energy that the world consumes is

wasted through friction and wear in mechanical and

electromechanical systems. It has been estimated that

approximately 11% of the total energy annually consumed in

the U.S. in the four major areas of transportation,

turbomachinery, power generation and industrial processes can

be saved through new developments in lubrication and

tribology [1]. Friction is responsible for a major loss of useful

mechanical energy and wear is a major reason for replacing

equipment. Thus, a better understanding and utilization of the

principles of tribology is particularly important for conservation

of energy and materials in engineering design [2].

The Pin-on-Disk tribometer [3] is the most common type

of friction and wear testing machine. Its basic design consists of

an arm with a pin (rolling ball) attached to the end of an arm

with the pin in direct contact with a test specimen secured on a

rotating disk. A normal load is applied at the end of the arm

directly onto the pin, which causes the rolling ball inside the

pin to scratch the surface of the disk. The radius of scratch and

the sliding speed can be adjusted. The friction force is

measured, in this case using a strain gage attached to the arm.

The amount of wear can be calculated by measuring the wear

width of both specimens [3], from the change of cross-

sectional area measured with a profilometer [4] and weighting

both specimens before and after the test.

The task at hand was to design and build a custom version

for Dr. Patricia Iglesias Victoria, a faculty of the mechanical

engineering department and tribologist. The rig will be used by

Dr. Iglesias and future graduate and other research students

working with her.

DESCRIPTION OF THE APPARATUS

In this design, an arm holds a pin and, by means of a

counterweight, balances so the pin is just touching the surface

of the tested material. The pin holds a 1.5 mm diameter ball.

The specimen is clamped down onto a plate below the pin.

Before the test begins, a known weight (FN) is added to the top

of the arm directly above the pin, this is the normal force. As

the test begins, the specimen rotates at a constant predesignated

speed and the pin wears the specimen surface in a circle. The

friction coefficient between pin and specimen can be measured

from strain in the arm and converted to force (FR). Using these

two known forces, the coefficient of friction can be calculated

using Eq. (1).

(eq. 1)

The radius of the wear track can be adjusted with a lead

screw. A swivel allows the arm to rotate out, and can lock at set

positions to keep the arm stationary both for testing and

changing the pin. The motor and strain gage recordings are

controlled with LabVIEW software. The final assembly is

shown in Fig. 1.

Figure 1. The pin-on-disk tribometer with labeled parts.

Creo

The entire design was modeled with Creo as the first step

in the design process. Once the three-dimensional model was

created, two-dimensional drawings were made for references

during the machining process. The Creo model is included in

Fig. 2. All the designs discussed in the paper were first

developed in Creo. The three-dimensional model was the

primary tool developing the design process.

Figure 2. The pin-on-disk tribometer modeled in Creo.

CALCULATIONS AND DESIGN

The requirements of the tribometer were to test the

coefficient of friction between a specimen and the pin. The rig

1 Arm

assembly

2 Counter

weight

3 Applied

weight

4 Pin

5 Specimen

6 Disk

7 Plunger

assembly

8 Motor

(below

table)

9 Lead

screw

assembly

10 Mount

11 Swivel

12 Table


needed to test coefficients between .01 and .99 for both

lubricated and dry tests. The table holding the specimen

needed to spin at a constant speed up to 500 RPM. Test length

must range from a few minutes to 15 hours or possibly more.

Specimens and the pin must be replaceable, preferably easily,

and the pin height needed to be adjusted. The rig required a

way to adjust between the scratch radius. Components needed

to be strong enough to perform properly, with addition of the

strain gage arm designed to be weak enough to measure strain.

Finally, the rig must be safe to use with proper protection from

possible harm and spray from lubricated tests. All of these

requirements are proven met within the calculations and design

section.

Force and moment analysis in the arm

In order to decide which material would be best for the

arm, a force and moment analysis was completed for both steel

and aluminum. For this calculation, the arm was assumed to be

one piece with the same cross section as the current arm front

and rear sections). The weight of the rod, threaded rod, counter

weight, and applied weight along with their displacements were

considered to find the bending and shear diagrams. The weight

was treated as an even load distribution along the arm. The free

body diagram is included in Fig. 3. The singularity equations

are as follows in Eq. (2)-(4) below.

Figure 3. Free Body Diagram of force and moment analysis.

(eq. 2)

(eq. 3)

(eq. 4)

The pivot was located at 0.21m, the end of the arm and

beginning of the threaded rod at 0.25m, and the threaded rod

ended with the counter weight at 0.3m.

The factor of safety was also found for each case. Because

the factor of safety was so great (632 with aluminum),

aluminum was determined to be the appropriate arm material.

The exact weight of the counterweight was optimized, as well

as the size. The counterweight is made of brass, as its high

density minimizes the size needed. Only one weight was

required, which could easily be assembled to one threaded rod

extended from the middle of the arm (alternate designs have

multiple counterweights [4]).

For the middle of the arm, a ductile material was

necessary to easily measure the strain. Additionally, aluminum

is much easier/faster to machine. The corresponding shear and

moment plots are shown in Fig. 4.

Figure 4. Shear and moment plots for Aluminum.

Counterweight

In order for the pin to sit on the surface of the specimen,

the arm needed a counterweight on the other side to balance it.

Brass was chosen as the counterweight material because of its

large density. Because the volume of each part was known

(from Creo software) as well as the densities of the material,

the volume of the counterweight could be calculated using the

following moment sum:

(eq. 5)

Gravity can be canceled out from each term to simplify the

above equation. The volume of the arm on the right of the

pivot was subtracted out with its same volume on the left side

of the pivot. Assuming a conservative distance for the

counterweight to be 9 cm from the pivot, the volume was

calculated as just below 50,000 mm3. A rod with an outer

diameter of 21 mm was available.A 6 mm diameter was

machined out of the counterweight for threading it onto the rod.

The length of the counterweight was machined to 38 mm, a

reasonable length. The corresponding mass of the

counterweight is .42kg.

-4 -3 -2 -1 0 1 2 3 4 5

0 0.05 0.1 0.15 0.2 0.25 0.3

Sh

ear

Forc

e (N

)

Arm Length (m)

Shear Force for Aluminum

-0.40

-0.30

-0.20

-0.10

0.00

0 0.05 0.1 0.15 0.2 0.25 0.3

Ben

din

g M

om

ent

(N-m

)

Arm Length (m)

Bending Moment for Aluminum


Lead screw

In order to allow multiple tests on the same sample, the

arm needed to be able to adjust linearly to change the diameter

of wear track. A simple solution for the linear movement was

to incorporate a lead screw assembly below the arm. An

example from the supplier is included in Fig. 5.

Figure 5. Lead screw assembly excluding shafts from

MiSUMi, the supplier [5].

The lead screw included two shafts to hold the weight of

the assembly, it is important that the screw did not take on any

extra weight. Each end of the screw has a pillow block bearing,

one fixed and one supported. The lead screw length needed to

include difference in radius of scratch, thickness of the mount,

and addition space on each side to keep the lead nut from

bottoming out.

The critical speed of the screw is determined in Eq. 6 to

decide whether it will meet the assembly needs:

(eq. 6)

N is 1.47 (one end fixed, one end supported) in this case, and d

is provided by the supplier.

Obviously the lead screw will not exceed its critical speed, it is

plenty thick and it is much too short for it to be unstable.

The critical load of the lead screw was calculated as such:

(eq. 7)

N is 2.00 for one end fixed, one end supported. Because N is

provided for English units and the supplier dimensions are

English, the load was first calculated in pounds and converted

back to metric.

Again, the critical scenario is not close to being reached, so this

lead screw more than satisfies the requirements.

At the end of the assembly is: a handle to rotate the screw,

a position indicator, which tells the location of the pin in

millimeters from the center of the specimen, and a lock to hold

the arm in place so it does not move during a test. The handle

hangs off the edge of the table to eliminate the possibility of

hand or arm interference. This is shown in the Fig. 6.

Figure 6. The lead screw assembly incorporated on the

tribometer.

Swivel design

A design requirement was to make the pin height

adjustable and completely removable in order to replace the

ball within. To do this, the arm would need to move enough

distance from the testing surface for easy accessibility. A

common approach with pin-on-disk tribometers is to allow the

arm to bend up at a very high angle, possibly up to 45º [4].

This method did not seem desirable for the sake of the

counterweight interfering with the lead screw assembly beneath

it. This could be avoided by raising the height of the arm above

the lead screw a considerable distance, installing two lead

screws instead of one and the weight would drop down between

them (a common solution), or resolving a way for the

counterweight to move entirely out of the way, such as to the

side or completely off. None of these results sounded desirable.

It was best if the arm didn’t tilt upwards to replace the pin, so

an alternate to this was to rotate the arm out to the side.

Originally, the idea was to rotate just the arm above the lead

screw, but a locking swivel that was the right size for that

application was not easy to find. Slightly larger locking

swivels, however, were available, but would not fit easily

between the lead screw and arm. The design of the swivel

location needed to be reevaluated. The idea came to put the

swivel below the entire lead screw assembly. Because the

tribometer is small in general, the locking swivels were large

enough in size to fit the lead screw assembly above it. The arm

did not need to swing up at a steep angle in order to replace the


pin, which made for an easy replacement. The swivel and the

subassembly it supported is pictured in Fig. 7.

Figure 7. Side view of the swivel.

Once the swivel was purchased, the lock and rotation

worked sufficiently for a commercial product, but there was

some considerable play in the locked position considering the

design needed to minimize any possible source of vibration. In

order to tighten up the locking mechanism, the swivel was

partially disassembled, the locking pin hole was bored out

larger for a boss to be press fit in, and a new pin was made to

be long enough to hold in the new boss. A section of the swivel

plate was also machined out in order for the handle to rise up

higher for the longer locking pin.

Motor selection

A motor needed to be selected to meet the torque

requirements at the speed of 500 RPM (8.3 RPS). The motor

speed would kept at a constant throughout a test. A stepper

motor would easily fit the requirements, and Aneheim

Automation provides stepper motors. Stepper motors are less

expensive than servo, and the precision of a servo motor was

not necessary because only the speed being output needed to be

known.

In order to calculate which motor would satisfy the

conditions, the torque generated by the assembly upon start up

needed to be calculated. The torque caused by all the rotating

elements acts similarly to a flywheel and can be found using

Eq. (8)

(eq. 8)

The torque caused by the friction from the pin is calculated as:

(eq. 9)

The max condition values were used for calculating the

torque, so assuming the maximum size of the sample (causes

larger moment of inertia), largest wear track radius, largest

coefficient of friction, and maximum applied load. The

moments of inertia were calculated assuming circular sections

except for the plungers, which were point loads. It was

assumed the motor accelerated from naught to 500 RPM in one

second. The following torques were found:

(eq. 10)

The torque is then converted to English units for referrencing

the supplier given chart.

Figure 8 shows that motor 23MD106 will be appropiate

considering the RPS of 8.3 s and max torque of 62.5 oz in. The

motor is connected to the data acquisition (DAQ) system,

which acts as a counter for the motor, and the DAQ is

connected to a computer. A LabVIEW Virtual Instrument (VI)

code was written to control the motor. An external power

supply was also required separately to power the motor, which

had 24 V and 2.7 A voltage and amperage requirements,

respectively.

Figure 8. Torque curve for motor selections available.

Taken from Anaheim Automation website [6].

Strain gage calculation

During testing, the resulting friction force needed to be

measured, this is done by means of a strain gauge mounted on

the arm. Three kinds of strain could act on the arm: axial,

torsional, and bending. As a frictional force is applied to the tip

of the pin, no axial strain acts on the arm, and both bending and

torsion strain act on arm. Torsional forces cancel from the full

bridge setup of strain gage, this will be explained further on.


The only remaining strain is caused by bending, the equation

for pure bending strain is noted below:

(eq. 11)

The Young’s Modulus will be the same with each experiment,

and the load will vary from 8.8N (1kg, μ=.9) to .05N (500g,

μ=.01) as extremes, depending on the mass applied to the top of

the arm and the range of estimated coefficient of friction. The

strain value calculated needs to fall in a measurable range of 10-

3 and 10

-6 strain, so the length, width, and height of the section

are altered to maximize the strain values. The dimensions

selected were a length of 76mm, width of 10mm, and thickness

of 6mm.

Typically, a section of the arm is machined down to alter

the length, width, and thickness parameters to get the strain to a

measurable value. The same concept is applied here, but

instead of machining down the surface, a smaller section of the

arm is detachable. Making this section removable benefits the

design, for example if the smaller section is bent, it can be

replaced with a new piece, and the detailed machining work in

the rest of the arm would not need to be redone. On top of this,

the option to remove the arm is available for replacing the

strain gage. This section is pictured in Fig. 9.

Figure 9. Strain gage bonded to surface and lead wires

soldered.

The configuration of the purchased strain gage for the

tribometer is shown in Fig 10. Mounting a strain gage on both

sides of the arm will result in a full bridge set-up, which

maximizes accuracy of the strain measurement.

Figure 10. Strain gage purchased.

The strain gage acts as a resistor in a Wheatstone bridge

(Fig. 10), a circuit which can measure unknown resistance.

Figure 11. Wheatstone bridge circuitry.

In a full bridge, each resistor is replaced with a strain gage.

Gages experiencing positive strain are connected directly with

gages experiencing negative strain (Fig. 12).

Figure 12. Wheatstone bridge with respect to strain layout.

Because each resistor is being replaced with a strain gage in

a full bridge set-up, effects such as change in ambient

temperature and extraneous loads cancel out, as long as each

strain gage is centered and parallel to the beam as well as the

strain gages on opposite sides are symmetrically loaded.

Because both the strain gage and the arm are fastened in the

center of the arm, torsional effects are also canceled out.

Arm mid-section design

Because the middle section of the arm was designed to be

weaker to increase the axial bending (and strain), it was

important to check if it could hold the weight of the arm and

pin. One possible rudimentary calculation involves making

sure the stress in the arm did not exceed the yield stress, which

calculates stress as:

(eq. 12)

Each is calculated in Equations (13)-(15).

(eq. 13)

(eq. 14)

(eq. 15)

Stress is solved as:

The yield limit of aluminum is 200 MPa. The dimensions of

the arm are satisfactory.


Because the mid-section of the arm was designed to be

weaker, it is also designed to be removable. This is because if

it ever needed to be replaced, the detailed machining work at

both ends would not need to be redone. If the mid-section ever

needed to be replaced, the length of the exposed section should

be fastened to be exactly 76.2 mm so that the pin is centered

properly on the disk. A rubber stop is included just below the

arm towards the back so that the counterweight hanging off the

back would never smash into the position indicator directly

below it.

Plunger design

Many current tribometer designs mount the test specimen

by either bolting down the specimen at its center [X] or by

clamping down the specimen to a plate to a specific bolt pattern

[X]. Both designs have advantages. The first allowing the

specimen to be any diameter (assuming it is big enough to fit a

bolt and leave enough room to test), but takes time to machine

each test sample. The second does not need any adjustments

made to the sample, but takes time to unbolt and bolt the

clamps for each test that mounts a different size specimen.

Although this method allows the samples to be a range of

diameters, they still have to be specific.

The goal was to design something that was both quick to

set-up and allowed for any size diameter within a defined range

(.5 mm to 23.5 mm scratch radius was requested). Machining

into the part was not an option, as the specimen size is small

and would leave little room for testing. The solution was

simply to clamp down the specimen with an adjustable threaded

fixture (Fig. 13). This fixture was a stationary piece with a

threaded hole. A long screw was added with a nut to hold it in

place, and a rubber tip to hold the specimen down. Because the

plate will never spin faster than 500 RPM and the pin does not

exert much lateral force on the specimen, tightening the

specimen down with hand tightness is enough to keep the

specimen from moving during tests, The nut can be tightened

with a wrench for a more comfortable clamp.

The specimen also has a range of heights (10 mm to 50

mm). This can be accounted for with the pin height.

The platform was machined with a ring pattern to allow the

specimen to be centered visibly.

Figure 13. Specimen mounted and tightened with plungers.

Shaft

Simply put, the rotation of the motor transfers into the

mounting plate. In order for rotational output from the motor to

properly transfer into the plate, specific design considerations

needed to be accounted for in order to minimize slop and

wobble within the subassembly. The motor shaft and hardened

shaft are connected by means of a flexible coupling (Fig. 14).

The shaft connects to the plate by a machined part, it was press

fit and secured with set screws to keep from slipping. The part

is flanged and bolted into the plate. A flanged ball bearing is

bolted to the table, which takes care of a great amount wobble

in the shaft. In between this bearing and the machined part is a

flat bearing; this small piece is important because it allows the

weight of the plate to be held by the table through the bearing it

sits on. Without it, the motor would not be able to function

with the weight of the plate assembly pressing down on it.

Lastly, spacers are surrounding the bolts holding the motor in

order to further prevent wobble.

Figure 14. Shaft assembly and counterparts.

Bearing calculations

The bearing life of the flat bearing can be calculated as

such:

(eq. 16)

C is provided by the supplier. If the load capacity was the same

as the load applied to the bearing, the bearing would last 106

revolutions. The value e is 3 for ball bearings, so is the same

value for each bearing in the tribometer. Assuming that the

bearing runs at 400 RPM (on average) and is used for 10 hours

a day, the flat bearing will fail in 362.5 days (according to that

calculation). In actuality, the tribometer will go without being

used for many days at a time throughout the year. This bearing

coupling


is easy to replace, only one screw on the coupling needs to be

loosened, and the entire plate assembly lifts out, and the

bearing simply sits in place.

The flanged linear ball bearing would see the same number

of rotations, but not a huge amount of force. There may be

some force from the motor jerking, but not enough to notice

visibly, as it runs smoothly. Very little force would be exerted

from the weight tilting because it is restrained from tilting

much, a fraction of the 23N. The dynamic load capacity of the

bearing is 320 N, so the bearing will easily last a long time.

The remaining bearings barely rotate or rotate slowly.

These bearings include the two ball bearings in the arm at the

pivot, the two linear ball bearings sliding on each shaft on

either side of the lead screw, and the two ball bearings in the

pillow blocks. Because they all do not move much, as long as

the load on the bearings does not exceed the load capacity, it is

safe to assume these bearings would last plenty of time to serve

the assembly. Table 1 shows the loads and load capacities for

each remaining bearing.

Table 1. Bearing load capacities and loads

BEARING

LOCATION

LOAD

CAPACITY [N]

LOAD

[N]

Arm pivot 3247 5.40

Linear shafts 262 30.3

Pillow blocks 1300 10.0

Normal force

Once the arm is balanced, a known weight is applied to the

top of the arm directly above the pin. The masses range from

100g to 1000g in 100g increments. The normal force is simply

the mass times gravity:

(eq. 17)

Although this may seem trivial, it is important to simplify the

design whenever possible.

Ground surfaces

Figure 15. Table during grounding process.

All surfaces being built upon each other needed to be

precision ground. All the ground parts include the table, both

the plate and platform which mount the specimen, as well as

the arm mount and plate the arm assembles too. The ground

surfaces (Fig. 15) are flat within a tolerance of .13 mm.

Case

A case was also designed for both safety reasons and to

prevent splatter for lubricated tests. Because of the handle

hanging off the back of the assembly, it made more sense to

only enclose the rotating section instead of the entire assembly.

The case was designed to not interfere with the arm when

stationary or rotated out, and needed to be removed for easy

access to sample set-up. Instead of making one side removable,

all four sides were attached together for rigidity, and a

detachable piece allowed space for the arm to rotate out. The

piece is attached with hand screws and a bracket. The sides

were fixed together with an acrylic chemical that melts the

plastic together. The case overhangs the edges on three sides

and the forth side is supported by the table; on the opposite

edge, the case is held up/kept from sliding out with two

brackets. The case is made of plexiglas (Fig. 16).

Figure 16. Case assembled and mounted on tribometer.

Detachable piece removed for arm to rotate out.

Because acrylic cracks easily when drilled through, it was

risky to drill right into the sides, holes were drilled in a separate

piece and attached to the handles, then glued to the case. The

pieces with holes were taped during the drilling process to keep

them from cracking, and machined slowly with soapy water as

lubricant.

ASSEMBLY

The entire machining and assembly process was completed

over the final three months of the project. The first parts and

materials arrived right before this period, on schedule for

machining.


Machined parts

Most components in the assembly were machined, either

entirely or partially. Aluminum was used whenever something

stronger was not necessary to simplify the machining. All of

the machined parts are included in Table 2. Figure 17 includes

a picture of the fully machined front end of the arm, showing

that individual parts have much machined detail.

Table 2. Fully and partially machined parts

FULLY MACHINED PARTICALLY MACHINED

Arm front Threaded rods (2)

Arm back Shaft

Arm mid section Swivel

Arm mid section mounting

brackets (2)

Handle

Counterweight Lead screw

Mounting block Case

Plate

Disk

Platform

Plunger holders (4)

Shaft to disk connection

Table

Legs (4)

Figure 17. Front end of arm detached after completed

machining.

Assembly order

The advantage of doing an independent project over an

industry product is the order of assembly does not have to be

entirely defined before the parts have arrived or even before the

machining process is complete. This advantage allowed many

parts to be marked with respect to a location of another part’s or

subassembly’s geometry. This lessened the tolerance on many

machined details, which is important for precision work.

The order of assembly is as follows in Table 3.

Table 3. Assembly order summarized

ASSEMBLY ORDER

Table including legs

Swivel to table

Plate with mounting block and lead screw subassembly to

swivel

Lock and handle to lead screw

Arm subassembly to mount

Threaded rods, counterweight, applied weight, and pin to arm

Motor and linear bearing to table

Disk subassembly, flat bearing unattached to motor/linear

bearing, coupling fastened last

Case, sitting on table/brackets fastened to table

The mounting block provides a good example of a part that

needed precision machining and benefited from partial

assembly. The bearing holes were bored out for a press fit

installation. Their locations were determined from the shafts

already assembled to the plate. The press fit requirement

allowed almost no location and diametral tolerance. This

scenario is similar anywhere something was press fit, such shaft

and other bearing transitions.

The arm was able to unattach easily. It was helpful during

the assembly process for machining updates and strain gage

installment.

Another feature that benefited from machining during the

assembly process was the shaft hole in the table. This was

machined after the arm was installed, marked by the location of

the pin. The four motor mounting holes were machined with

respect to the shaft. This step is captured in Fig. 18.

Figure 18. Tribometer partially assembled during machining

process.

The final mechanical assembly was the case. As previously

mentioned, the case was assembled with an acrylic paste, and

the handles attached first by a separate piece then pasted onto

prevent cracking the main plexiglas.

The strain gage and motor were then wired to the DAQ,

which connected to the computer.

CONCLUSIONS

As expected, many issues arouse during the machining and

assembly process, due to either unforeseen circumstances or

inexperience. Many design improvements were made along the


way. The design already is a streamlined version of existing

pin-on-disk tribometers, but it could be streamlined even

further, in terms of both function and physical appeal. For

example, making the plate, platform, and shaft transition pieces

lighter would have greatly reduced the power the motor needed.

In terms of overall aesthetics, it would have been more visually

pleasing if the bolts were threaded into the plate instead of

attached with nuts. The idea was to minimize tolerance, but

threading into material has a much nicer appearance. The final

difference would be to know many of the design change

descisions beforehand as to keep shipping costs down.

Overall, the tribometer was a success, both for providing

equipment to the department and as a learning experience.

ACKNOWLEDGMENTS

The authors would like to acknowledge the financial

support of the Mechanical Engineering Department at the

Rochester Institute of Technology. Special thanks to the RIT

machine shop for allowing use of their machines and answering

questions throughout the production process. Additional thanks

to professors, peers, and staff willing to provide feedback and

answer questions throughout the entire design, build, and

writing process. About the author

The project was completed at the Rochester Institute of

Technology while Elspeth was working towards her Master of

Engineering Degree with a focus in Mechanics and Design.

Elspeth now works full-time at UTC Aerospace Systems in

Windsor Locks, CT as a Mechanical Design Engineer in the

Space Systems group. Her current work supports the design for

the Orion Spacecraft.

REFERENCES [1] Bronshteyn, L. A., and Kreiner, J. A. H., 1999, "Energy

Efficiency of Industrial Oils," 42, pp. 771-776.

[2] Czichos, H., 1983, "Tribology: Scope and Future Directions

of Friction and Wear Research," Journal of Metals, 35(9), pp.

18-20.

[3] Standard test method for wear testing with a pin-on-disk

apparatus, ASTM G99-05.

[4] Iglesias, P., Bermudez, M. D., Moscoso, W., Rao, B. C.,

Shankar, M. R., and Chandrasekar, S., 2007, "Friction and Wear

of Nanostructured Metals Created by Large Strain Extrusion

Machining," Wear, 263(1-6 SPEC. ISS.), pp. 636-642.

[5] MiSUMi-Product Specifications. “Lead Screws –For

Support Units-” June 2014.

[6] Anaheim Automation. “23MD - Stepper Motors with

Integrated Drivers.” 2011. June 2014.

[7] CSM Instruments. “CSM TRIBOMETERS--Nano & Micro

range for Tribological studies.” pp. 2. June 2014

[8] Nanovea. “Tribometers.” pp. 1-4. June 2014.


proceedings of the asme 2015 international mechanical

Documents