Design and Fabrication of a Miniature Tensile Load frame for a

Scanning Electron Microscope

Senior Design Team 04-004

Critical Design Report

May 14, 2004

Robert Rinefierd – Team Manager, Mechanical Engineer Evan Kastner – Lead Engineer, Mechanical Engineer

Nicholas Currier – Mechanical Engineer Blaine Stuart – Mechanical Engineer

Kennedy Mogwai – Industrial Engineer Evan Brunner – Computer Engineer

Executive Summary

The tensile load frame is part of a series of projects funded by the Mechanical

Engineering Department to construct specialized lab equipment. The purpose of this

project was to design and construct a load frame for the scanning electron microscope

(SEM) in the CIMS Materials Science Lab. Mechanical Engineering faculty and students

will use the machine for research of metallographic structures during tensile testing.

Most tensile load frames will not fit within a small vacuum chamber, which makes this

design unique. In addition, most microscopes will not allow for dynamic analysis of

bonds, welds, or other areas of interest during a tensile test.

In addition to fitting inside the narrow packaging envelope in the vacuum

chamber, the load frame must operate safely in a vacuum environment. Many

mechanical constraints of the chamber limit the usable space to an envelope of 10 inches

long, 9 inches wide, and 3 inches high. All components must be vacuum-rated to avoid

contamination and outgassing. Also limiting the design is a maximum cost of $7,500.

Several concepts were evaluated, including a design using two power screws and

a motor inside the vacuum chamber, a hand driven design with a removable crank, a

design using a single driving screw and an internal motor, and a two screw design with a

motor mounted externally to the chamber. The two-screw design with an internal motor

was the recommendation from feasibility assessments and the design was developed.

The test samples will be standard ASTM cylindrical geometry with threaded ends.

Loads will range from 200 lb in compression to 2000 lb in tension. After calculating the

torque required to raise the tension to 2000 lb from a static condition, motors, gearboxes,

and drivetrain components were researched to create a mechanism to apply the necessary

torque to the power screws. The design utilized a mounting point on the existing SEM

position fixture, necessitating a cantilever design, where the two power screws support

the free crosshead.

The driving source is a stepper motor using a controller with load feedback and a

position estimate. Load and position data will be recorded in LabView and a user

interface will be designed to run the load frame from a laptop computer. The design was

analyzed with finite element modeling for stress and deflection before production began.

A functional prototype was built without any major problems.

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Table of Contents

1 Recognize and Quantify the Need .............................................................................. 1

1.1 Project Mission Statement........................................................................................................... 1

1.2 Product Description .................................................................................................................... 1

1.3 Scope Limitations ........................................................................................................................ 2

1.4 Stakeholders ................................................................................................................................ 2

1.5 Key Business Goals ..................................................................................................................... 2

1.6 Financial Analysis ....................................................................................................................... 3

1.7 Preliminary Market ..................................................................................................................... 3

1.8 Secondary Markets ...................................................................................................................... 3

1.9 Order Qualifiers .......................................................................................................................... 3

1.10 Order Winners............................................................................................................................. 4

1.11 Innovation Opportunities............................................................................................................. 4

1.12 Background Research.................................................................................................................. 4 1.12.1 Describe the Need.............................................................................................................. 4 1.12.2 Categorize the Need........................................................................................................... 4 1.12.3 Constraints ......................................................................................................................... 5 1.12.4 Existing Products ............................................................................................................... 5

2 Concept Development................................................................................................. 7

2.1 Overview of Tensile Testing Equipment ...................................................................................... 7

2.2 Integrating the Product with the Current Fixture ....................................................................... 9

2.3 Concept Design Proposals ........................................................................................................ 11

2.4 Concept 1 – Internal Motor with Two Driving Screws.............................................................. 14 2.4.1 Concept Overview ................................................................................................................ 14 2.4.2 Design Features .................................................................................................................... 14 2.4.3 Preliminary Bill of Materials ................................................................................................ 15

2.5 Concept 2 – Internal Motor with Single Driving Screw ............................................................ 17 2.5.1 Concept Overview ................................................................................................................ 17 2.5.2 Design Features .................................................................................................................... 17

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2.5.3 Preliminary Bill of Materials ................................................................................................ 18

2.6 Concept 3 – Manually Driven Load Frame............................................................................... 21 2.6.1 Concept Overview ................................................................................................................ 21 2.6.2 Design Features .................................................................................................................... 21 2.6.3 Preliminary Bill of Materials ................................................................................................ 22

2.7 Concept 4 - External Motor with Two Driving Screws ............................................................. 24 2.7.1 Concept Overview ................................................................................................................ 24 2.7.2 Design Features .................................................................................................................... 24 2.7.3 Preliminary Bill of Materials ................................................................................................ 25

3 Feasibility Assessment.............................................................................................. 26

3.1 Introduction............................................................................................................................... 26

3.2 Evaluation of design concepts ................................................................................................... 26

3.3 Pugh Evaluation ........................................................................................................................ 27

3.4 Weighted Concept Evaluation ................................................................................................... 27

3.5 Results ....................................................................................................................................... 29

3.6 Conclusion................................................................................................................................. 32

4 Objectives and Specifications ................................................................................... 33

4.1 Design Objectives...................................................................................................................... 33

4.2 Performance Specifications....................................................................................................... 33

4.3 Design Specifications/Implementation ...................................................................................... 34

4.4 Evaluation Criteria.................................................................................................................... 34

4.5 Safety Standards ........................................................................................................................ 35

5 Analysis and Synthesis ............................................................................................. 36

5.1 Design Structure Matrix (DSM) ................................................................................................ 36 5.1.1 Introduction .......................................................................................................................... 36 5.1.2 Problem Statement................................................................................................................ 36 5.1.3 Results analysis..................................................................................................................... 36

5.2 Motor and Gearbox Selection ................................................................................................... 37

5.3 Drivetrain Design...................................................................................................................... 39

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5.4 Gripping Mechanism................................................................................................................. 41 5.4.1 Specifications of the Grips.................................................................................................... 41 5.4.2 Purchased or Machined?....................................................................................................... 42 5.4.3 Final Design of Grips ........................................................................................................... 42

5.5 Base and Frame......................................................................................................................... 43

5.6 Control and Display .................................................................................................................. 44 5.6.1 Beginning Estimates ............................................................................................................. 44 5.6.2 Control response constraints................................................................................................. 45 5.6.3 Sample Response Characterization....................................................................................... 45 5.6.4 Hardware Design Fundamentals........................................................................................... 47 5.6.5 Control Software................................................................................................................... 49

5.7 Vacuum Interface ...................................................................................................................... 51

5.8 Stress Calculations for Critical Components ............................................................................ 51

6 Preliminary Design ................................................................................................... 53

6.1 Part Numbers for Pro Engineer files......................................................................................... 53

7 Engineering Models .................................................................................................. 54

7.1 Part and Assembly Modeling..................................................................................................... 54 7.1.1 Modeling with Pro/Engineer................................................................................................. 54 7.1.2 Finite Element Analysis with Pro/Mechanica and I-Deas .................................................... 55

7.2 Testing the Product for Reliability and Quality ........................................................................ 61 7.2.1 Integrated Test Plan for SEM Load Frame ........................................................................... 61 7.2.2 Individual Test Plans ............................................................................................................ 62 7.2.3 Component Processing ......................................................................................................... 63 7.2.4 Subassembly Testing ............................................................................................................ 63 7.2.5 Deployment Testing ............................................................................................................. 64

8 Final Design .............................................................................................................. 65

8.1 Changes from Preliminary Design ............................................................................................ 65 8.1.1 Gripping................................................................................................................................ 65 8.1.2 Shaft and Bearing Setup ....................................................................................................... 66 8.1.3 Free End and ACME Nuts.................................................................................................... 67 8.1.4 Custom Gear Design............................................................................................................. 67

8.2 Cost Analysis ............................................................................................................................. 68

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8.3 Design for Manufacture ............................................................................................................ 69 8.3.1 Introduction .......................................................................................................................... 69 8.3.2 Design Phase ........................................................................................................................ 70 8.3.3 Design Optimization............................................................................................................. 71 8.3.4 Results .................................................................................................................................. 71 8.3.5 Suggestions for Redesign ..................................................................................................... 72

8.4 Included Parts ........................................................................................................................... 72

9 Production Planning.................................................................................................. 75

9.1 Material Considerations............................................................................................................ 75

9.2 Tooling Design and Machine Setup........................................................................................... 76

9.3 Manufacturing Process Sheets .................................................................................................. 77

10 Pilot Production .................................................................................................... 79

10.1 Manufacturing Difficulties ........................................................................................................ 79

10.2 Assembly Design........................................................................................................................ 80

10.3 Manufacturing Pictures............................................................................................................. 80

10.4 Recommendations for Improvement .......................................................................................... 83

References......................................................................................................................... 85

Appendix........................................................................................................................... 86

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1 Recognize and Quantify the Need 1.1 Project Mission Statement

The mission of this senior design team is to design and construct a tensile load

frame for the scanning electron microscope (SEM) in the CIMS Materials Science Lab.

The tensile stage must be lightweight, affordable, and easy to carry.

1.2 Product Description A scanning electron microscope (SEM) allows for high-powered magnification of

the surface or near-surface structure of specimens. Images are produced when a beam of

electrons is reflected off of the sample surface. A detector monitors the radiation and the

electrons scattered by the surface. Scattered energy and electrons form a surface profile,

which is mapped to a cathode ray tube and the image is formed. Apertures and magnets

act to focus the beam in much the same way that a lens would. However, the image is

not controlled by lenses, amplified only by changing the size of the raster, the area

scanned by the electron beam. Scanning electron microscopes are capable of much

higher levels of magnification than ordinary light microscopes, reaching magnifications

of 25,000X. In comparison, a typical light microscope may reach a maximum

magnification of 2000X. The most important characteristic of a scanning electron

microscope is that it has a large depth of field, allowing the image to stay in focus across

a rough surface. This is why the machine is ideal for examining fracture surfaces.

Currently, samples may only be evaluated without a load in the CIMS scanning

electron microscope. The surface analysis can only be performed on a sample before or

after testing. A load frame would allow the surface behavior of a sample to be monitored

under tensile loads. Distortion and elongation can be monitored dynamically on a

microscopic level while a tensile load is being applied. Applications for this project

would be to monitor local stress fields in welds, bond sites, or high stress areas. The

tensile stage is to be designed specifically for the SEM used in the CIMS Materials

Science Lab. Though opportunities may exist for incorporating product features into

future designs in industry, the product will be a single unit with no plans for mass

production.

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1.3 Scope Limitations The tensile stage must be designed and manufactured within 22 weeks of the

winter and spring quarters 2003-04. Sponsor Dr. Elizabeth DeBartolo will own the

tensile stage upon completion and it will be stored in the Mechanical Engineering

Department’s facilities. The National Center for Remanufacturing and Resource

Recovery (NCR³), owner of the SEM in the CIMS Materials Science Lab, has no current

interest in the product and will not be using it. Therefore, the product will need to be

transported easily to the CIMS building from the Engineering building.

The microscope has a position fixture, which is adjustable along the X, Y, and Z-

axis, tilt, and rotation. A load frame would need to be positioned to scan the desired area,

necessitating a positioning mechanism. Due to the logistics of replacing the existing

setup and the expense of building an additional module, the design will use as much of

the existing fixture as possible. It is not feasible to construct a new position fixture and

incorporate it with the load frame and is therefore beyond the scope of the project.

1.4 Stakeholders Major stakeholders in the project are Dr. DeBartolo and faculty members of the

Mechanical Engineering Department who may be performing materials research.

Additional stakeholders would include any thesis students or undergraduate students who

wish to perform advanced materials research with a scanning electron microscope.

1.5 Key Business Goals This project is part of a series of projects funded by the Mechanical Engineering

Department, the goal of which is the design and construction of lab equipment for use by

mechanical engineering students and faculty. With financial constraints limiting the

ability to purchase new equipment, some highly specialized equipment that may not be

widely used will be designed and constructed with funding from the Mechanical

Engineering Department. Ultimately, the goal is to design for the production of one load

frame. If the design is innovative and successful, it may be presented at an ASTM

conference in the spring with other student designs. A possibility exists to market the

design to various manufacturers at such conferences, but this is not a requirement.

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1.6 Financial Analysis Dr. Hensel and Dr. DeBartolo, the faculty coordinator and faculty mentor,

respectively, approved a budget of $7,500 for design and construction of the load frame.

In addition to the purchased components, raw material cost is the only other major

expense. Use of the College of Engineering machine shop is free and manufacturing time

is not a monetary expense. Also, the necessary software packages (Pro Engineer and

ANSYS) are already available in the Mechanical Engineering labs, so no software costs

exist.

1.7 Preliminary Market As the project is funded by the Department of Mechanical Engineering, the

preliminary markets for this product are the students and faculty of the department who

may be performing metallographic research with the SEM in the CIMS Materials Science

Lab. The market will be very selective, as all users shall obtain permission from the lab

supervisor in CIMS as well as a faculty advisor before using the machine.

1.8 Secondary Markets The secondary markets for this product are students of other disciplines of

engineering and possibly their faculty. In addition, manufacturers of mechanical testing

equipment may be approached with the results of this project. Though they currently

have no interest in the product, engineers in the CIMS NCR³ Materials Science Lab may

find applications for it if a successful product is developed. This module may lead to

new NCR³ research projects involving the SEM.

1.9 Order Qualifiers The product shall be modular, easy to install, and lightweight. It should be simple

to manufacture, consisting of components that can be machined and assembled during

spring quarter. It shall have at least one working set of grips to accommodate either flat

or cylindrical specimens. Most importantly, it will have a position and load display,

allowing the samples to reach the correct tensile loads. The machine will meet the

sponsor’s minimum load capacity of 1000 lb.

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1.10 Order Winners The frame shall fit inside the SEM chamber; this distinguishes it from the

standard laboratory load frames. Most importantly, the load frame shall not interfere

with the functionality of the SEM, allowing for dynamic analysis of a sample in tension.

Tension can be adjusted while the chamber is at vacuum pressure, so multiple images

may be created at varying tensile loads. It will accommodate loads ranging from 2000 lb

in tension to 200 lb in compression.

1.11 Innovation Opportunities Though this is a single product not intended for major production or design

improvements, some innovation opportunities exist. The possibility exists for using

interchangeable grips. The use of threaded grips, v-grips, and flat grips would

accommodate almost any test sample that fits within the SEM chamber and load frame.

Cylindrical and rectangular samples could be analyzed. Design ideas may also be

marketed at an ASTM conference in the spring. Future senior design projects could

adapt the product for high temperature testing and testing microscale structures. Other

projects might involve the design of a tensile fatigue load frame roughly based on the

tensile frame developed in the project.

1.12 Background Research 1.12.1 Describe the Need

A description of the customer’s needs was prepared using sketches and various

written ideas. The manager’s ideas were compared with the customer’s needs. In the

span of three initial meetings, the customer provided feedback and contributed additional

ideas towards the Needs Assessment.

1.12.2 Categorize the Need

Category 4. New Problem, No Process or Product.

The SEM currently analyzes samples with no loading. Though reasonable

observations may be made with no applied load, much can be learned about the material

by its dynamic behavior under various tensile loads. RIT does not own a product capable

of applying tensile or compressive loads within the microscope’s vacuum chamber.

Relatively recent technology led to the development of other SEM tensile stages and only

a few designs currently exist. Though a load frame might be available for purchase, the

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cost would be prohibitive. As a result, the project will allow for the design and

development of a cheaper customized product to meet the research needs of the customer.

1.12.3 Constraints

The major constraints of the design process are the size limitation, material cost,

time for design and development, and time for manufacture. The fixture must fit within

the vacuum chamber and any electrical or mechanical connection must be properly

sealed. Port doors are replaceable with custom doors or windows and the seal must be

strong enough to withstand the vacuum pressure. The load frame must stay within an

envelope 3” high, 9” wide, and 10” deep. Components may extend outside the chamber

if the protruding parts are properly sealed. Most importantly, the components should be

free of any oils and materials should be safe in a vacuum environment. Special vacuum-

safe lubricants are available.

As the fixture is intended for the private non-commercial use of students and

faculty, it will not have any applicable industry standards or regulatory agencies to

govern the design, installation, or service of the product. However, it will meet the

approval of the SEM owner before installation. The product is a single unit of

production, customized for the microscope and position fixture in the CIMS Materials

Science Lab and will not be sold. Market prices are not important, but cost must be kept

within the allotted budget. The design is therefore not under as many constraints as a

marketed product, and it is a new problem. Though no product exists for this

microscope, designs may be based on similar existing products.

1.12.4 Existing Products

Two existing products were identified which provided some ideas for the design.

The design created by Lehigh University (Figure 1.1) was not patented, but a detailed

photograph allowed for analysis of the drivetrain setup and provided the team with an

idea with which to begin the design process [1]. It was the basis for the idea of a two-

screw motor-driven design. The design employs two power screws with spur gears on

the ends of each screw. A position sensor and load sensor are included. A motor and

gearbox combination drives the system. A series of gear reductions occur with the help

of a worm/worm gear pair and some spur gears.

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The other product was a US patent [2] and was not as helpful in creating the

design. In both cases, the budgets seemed to be considerably higher than that of this

team, so only the basic layout and some machine functions could be adopted into this

design. However, this discovery confirmed that such a fixture could be designed and

built.

Dr. David Davidson of the Southwest Research Institute in San Antonio is one of

the few individuals to design and build such a load frame. He has designed machines for

fatigue testing applications within a scanning electron microscope [3]. His designs were

too expensive and too complex to integrate into the development of this project.

Individuals such as Dr. Davidson usually design these miniature load frames instead of

large companies because each machine is a custom design.

Figure 1.1 - Lehigh University concept

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2 Concept Development Initial concepts were designed after a period of research, capturing a wide variety

of ideas from all team members. Beginning with background research on existing tensile

testing equipment, the design team reviewed existing full-scale load frames. Full-scale

machines provided the team with ideas for the initial concept designs. Brainstorming

sessions were performed for the more serious design concerns, including the integration

of the load frame into the existing scanning electron microscope and determining a basic

setup for the electronic control system. A series of team proposals was made for various

combinations of modules and for a general machine layout. The top four proposals were

developed with detailed concept sketches, a short report, a preliminary bill of materials,

and a preliminary cost analysis. Team members reviewed each concept proposal and

contributed ideas towards the robust development of each design in preparation for the

feasibility assessment.

2.1 Overview of Tensile Testing Equipment Uniaxial tensile testing is one of the most frequently performed mechanical tests.

On a full-scale machine, this type of test generally involves gripping a specimen at both

ends and subjecting it to an increasing axial load until it fractures. Collection of load and

elongation data during the test allows the operator to determine several characteristics

about the mechanical behavior of the material, such as strength and stiffness. Tensile

testing equipment consists of several types of devices that apply a controlled tensile load

to test specimens. The equipment is capable of varying the rate of load and accurately

measuring the forces, strains, and elongations applied to the specimen.

Equipment has evolved from purely mechanical or electromechanical machines to

sophisticated instruments that employ advanced electronics and microcomputers [4]. The

current technology of tensile testing equipment primarily consist of the force application,

moving crosshead, gripping mechanism, fixed base, control and display panels (Figure

2.1).

The load frame for this project was designed to fit and function within the

chamber of a scanning electron microscope (SEM), allowing for microstructure analysis

during tensile tests. Full-scale load frames would not fit within a microscope chamber.

This machine will allow analysis of microstructures during the tensile tests. The load

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frame is scaled-down, but will still include some of the technological advances of a

modern tensile load frame.

Figure2.1 – Diagram of Tensile Testing Equipment

Tensile Testing

Equipment

ElectromechanicalServo hydraulic

Driving

screwsCrosshead

Gripping

Method

Fixed

Base

High-Torque

Motors

Control &

Display

2 Screws 4 ScrewsScrew

Action

Wedge

Type

Button

head

Pneumatic

Action

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2.2 Integrating the Product with the Current Fixture Some of the most critical and difficult design decisions involved the incorporation

of the load frame within the existing microscope chamber. The scope of the project does

not involve designing a position fixture with the load frame, and as a result, the load

frame must be designed to fit the constraints of the existing position fixture. Early in the

concept development process, a brainstorming session identified most of the potential

problems that must be overcome to integrate the load frame into the existing SEM. The

twenty problems are listed in Table 2.1. Each group member placed four votes for the

problems that they deemed the most important to overcome. The important issues

identified were: maintaining a vacuum seal for wires and/or components that may pass

through the chamber, the size and location constraints of the chamber, the ability to

control the applied force and position, and the portability of the module. Though not

identified as an initial concern, cost grew in importance as research progressed on grips,

motors, and other purchased components. Also growing in importance was the vacuum

rating of all components.

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Table 2.1 Brainstorming results for development problems. Problems that must be overcome to integrate load stage with SEM Votes Rank

Maintaining vacuum seal / sealing methods 5 1

Size and location constraints of the chamber 5 1

Ease of assembly 0 NR

Grounding of sample (electrically) 2 5

Ease of installation 0 NR

Removing part of existing fixture / disconnecting wires 0 NR

Don’t interfere with electron gun 0 NR

No electrical interference 0 NR

Meets approval of machine owner (Mike H.) 1 6

Displaying load 0 NR

Portable 3 3

Lightweight 0 NR

Control of applied force and position 3 3

Control of grips 0 NR

Wire interface through door 0 NR

Cleanliness / no oil or impurities 0 NR

Securing load frame 1 6

Mechanical losses in motor 0 NR

Cost limit ~ $7,500 0 NR

Vacuum rating of components 0 NR

Distribution of load evenly in samples 0 NR

Since the team is composed of four mechanical engineers, one industrial engineer

and one computer engineer, incorporating an electrical control system was a major

challenge for the team. In order to further analyze the interaction of the tension control

system, an empathy session was performed. One student performed the role of the power

screws, while another performed the role of the motor. A third student performed the

role of the grips, a fourth acted as the control/display module, and a fifth performed the

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role of the test sample. An operator (not portrayed) places the sample inside of the grips

and tightens them manually. Once the sample is sufficiently tight, the operator turns a

dial on the control box to apply the tension. The control box sends a signal to the motor

to rotate at a specified torque based on the desired tension. The motor needs a drivetrain

(not portrayed) to interact with the screws. As the screws turn, the sample gradually

stretches. A load cell (not portrayed) is needed to transmit force data and a position

sensor (not portrayed) is needed for transmitting position data to the controller. When the

sample reaches its specified tension, the control box needs to cut the power on the motor

and lock position at the desired tension. The screws are self-locking and hold the sample

at the desired tension. Simply cutting the power to the motor is a sufficient solution.

Several issues presented themselves during the empathy session. One major issue

was supplying power to the motor and controls. Also, if the control box needs to interact

with the motor in addition to load and position sensors, the module must be customized

and have some programming functionality. The module may have to interpret signals

and convert them to a desired output. In order to prevent sample slipping, the grips

should be threaded, knurled, or grooved. The screws should not elongate or deflect in

bending. Fine pitch screws would provide the best position control and minimize

backlash. Most importantly, the motor’s torque output should be sufficiently high to

avoid stalling, a problem with potentially disastrous results when a load is applied.

2.3 Concept Design Proposals The initial concept design proposals involved various load frames based around

the same concept with different features. Four design choices were made for each design,

as the machine already had several constraints and would not allow radically different

concepts. Grip types were varied, including screw-driven, self-locking wedge grips, and

motor driven. Motors were varied between stepper motors and torque motors. The

support structure was either sliding or cantilever. The cantilever would have a fixed grip

and a moving grip that is supported only by the driving screws. A sliding support would

have a plate with runners to support the moving grip. Screws, gears, and belts were

considered as drivetrain options. Each group member had a maximum of four votes, but

did not have to use all of them.

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Table 2.2 - Initial Concept Proposals Grip Type Motor Type Support Drivetrain Votes Rank

1 Screw grips Torque motor Sliding base Screws/gears 5 1

2 Screw grips Stepper motor Sliding base Screws/gears 1

3 Self locking Torque motor Sliding base Screws/gears 3 2

4 Motor grips Torque motor Cantilever

base

Gears/shafts 0

5 Motor grips Torque motor Cantilever

base

Screws/gears 1

6 Motor grips Torque motor Sliding base Screws/gears 2 3

7 Screw grips Torque motor Cantilever

base

Screws/gears 2 3

8 Screw grips Torque motor Sliding base Gears/shafts 1

9 Motor grips Torque motor Cantilever

base

Gears/shafts 1

10 Self locking Torque motor Sliding base Belts 0

11 Self locking Stepper motor Sliding base Gears/shafts 1

12 Motor grips Stepper motor Sliding base Gears/shafts 0

13 Motor grips Stepper motor Sliding base Screws/gears 0

14 Screw grips Torque motor Sliding base Belts 0

15 Screw grips Torque motor Sliding base Hydraulics 0

The initial brainstorming session led to several design choices. Belts and

hydraulics would not be considered as options for the drivetrain. Hydraulics would not

be feasible because of the vacuum pressure and possible leakage of hydraulic fluid. In

addition, the price of a hydraulic system would be prohibitive. Due to the small size of

the system, belts and pulleys would not achieve the necessary speed reduction. A good

gearbox should be sufficient to achieve the necessary speed reduction from the motor to

the screws. The top four concepts had many similar features. Distinguishing between a

design decision and a completely different concept design was difficult. Purchased

components, such as the motor, gearbox, grips, and load cell dramatically increase cost.

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Choices must be finalized after various concept proposals incorporate the components in

different configurations. As a result, the group proposed a new set of designs. The set of

concepts was more general and focused on the creation of a generic load frame varying

the style and location of the drivetrain.

Table 2.3 - Final Concept Development Proposals Concept Description Votes Rank

1 2 driving screws with internal motor and gearbox 5 1

2 1 driving screw with internal motor and gearbox 5 1

3 2 driving screws with mechanical hand crank/gearbox 5 1

4 2 driving screws with an external motor/gearbox 4 1

5 1 drive screw with an external motor/gearbox 0 NR

6 4 drive screws with an internal motor/gearbox 0 NR

7 4 drive screws with an external motor/gearbox 0 NR

8 2 drive screws with motor and no gearbox 1 5

Concepts 1-4 were chosen for development and analysis in further detail.

Concept development studies focused on a concept overview, design features, a

preliminary bill of materials and some initial assembly sketches to work out the basic

logistics of the design and obtain a cost estimate.

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2.4 Concept 1 – Internal Motor with Two Driving Screws 2.4.1 Concept Overview

Concept 1 is an internal motor driven, two-screw, cantilever load frame contained

within the vacuum chamber of the SEM. The main features of this concept are screw-

tightened grips, a gearbox to screw drivetrain powered by a stepper motor, and a

cantilever mounting setup. The specific type of motor to be used and the method through

which it is controlled was a topic of discussion through the preliminary design stages. As

long as the grips are light and the power screws are of sufficient size, deflection shouldn’t

be an issue with this design. All fabricated parts within the SEM will be machined from

AISI stainless steel (303 or 304). If the screws bend too much, loading would not be pure

axial tension and load readings would not be accurate. The weight of the support on the

fixed end should be fairly small, as the gearbox weighs less than 2 pounds. Grips could

be purchased, but another cost-reduction option is a custom set of grips to fit the exact

specifications of the chamber. As a result, weight and size are minimized.

2.4.2 Design Features

On the fixed end, the gearbox and one grip will be fastened to an L-shaped

support. The support will be stiff, strong, and corrosion resistant. The two power screws

will support the free end and drive it by applying tension or adjusting position before

gripping. As stated above, stainless steel was the material of choice for the screws and

both grip supports because of its strength and corrosion resistance in the vacuum

environment. All machined surfaces must be free of any cutting oils and contaminants

before installation to avoid contamination of the vacuum chamber.

Turning a screw on the jaw modules will loosen and tighten grips. Ideally, grips

will be purchased from Tinius Olsen, MTT, Instron, or another leading competitor. The

cost of a vacuum rated motor and gearbox, vacuum-sealed electrical feed-through, not to

mention the load cell, leave very little capital for grips, and fabricating a set that will

work to perfection in a laboratory is cause for concern. Tinius Olsen has set a current

standard with a set of grips that meet the design requirements, although a slightly larger

than preferred, for $2000. Instron may be able to provide a set of smaller remanufactured

grips for considerably less money. The grips will have a wedge mechanism, moving the

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jaws closer together as a screw is tightened. The fixed grip will incorporate a load cell to

measure tensile force in the sample.

A small display box will incorporate a live display for tension and position.

Position display will be absolute and of mostly cosmetic benefit, however, the extremes

of travel will be entered into the control logic to keep the free end within its specified

limits. Wires will run through an existing port in the microscope chamber and connect to

a position sensor on the motor and a load cell in the fixed grip.

A feasible design could be created for $6,000-8,000, but several decisions must

first be made. Grips could be machined, but the issue would become incorporating the

load cell in the grips. The aim of this design is to provide a cost-effective solution to the

expressed wishes of the customer without major compromise, respecting that cost-

effective may still exceed the original budget allowance.

2.4.3 Preliminary Bill of Materials

Qty Description Purchased or

Machined

Price

(estimated)

1

1

Vacuum rated motor

Vacuum-rated gearbox

Purchased

Purchased

$1316

$2789

1 Additional Gearing Worm/ Idler/ Mount to Screw Purchased $100

2 Power Screws Machined $225

1 Load cell Purchased $500

1 Live display for load/position Both $100

1 Fixed end, Free end – grip support (material cost) Machined $95

1 Base support (material cost) Machined $37

1 Set (2) of wedge acting grips Purchased $2000

1 Set of assorted wires for electronic controls Purchased $50

50 Assorted sizes of cap screws Purchased $50

1

1

Vacuum Sealed Electrical Interface

Control module for motor

Purchased

Both

$500

$100

TOTAL ESTIMATED COST (Without Shipping) $7872

15

Figure 2.2 - Concept 1 sketch – top view (upper) and front view (lower)

16

2.5 Concept 2 – Internal Motor with Single Driving Screw 2.5.1 Concept Overview

Concept 2 is similar to Concept 1, except that it has a single driving screw. This

concept will consist of six modules: a high-torque motor, driving screws, a mounting

frame, grips, vacuum electrical interface, and display and control for load and position.

The gripping module will employ a set of wedge-type grips. One grip will be fixed to the

mounting frame. The free end’s grips will connect to the motor drive shaft through a

plate guided by a keyway in the mounting frame to transform rotational motion into

linear motion. The size of the motor and position of the grips results in a difficult

challenge within the space restrictions of the chamber. With the restrictions in mind, the

plan is to purchase the following modules: motor, grips and drive screws. The other

modules will be fabricated to fit the purchased components.

2.5.2 Design Features

MODULE

Action and Feature

Gripping

Purchase

Method of fastening

o Self locking

Power Purchase

o Torque motor

Mounting/Frame Fabricate

Fixed base

1 Sliding Grip

Control/Display Fabricate/Purchase

External

o Load

o Position

o Start/Stop

17

Drive Train Purchase

Drive Screws

o 1

Interface Vacuum

Clearance

Based on test specimens that will be approximately 3 inches long with a gage

length of 1 inch, a set of Action Wedge grips will be suitable. Their cost is not known,

but they come in various sizes and load ratings. The grips will have a load capacity of

1,000 to 60,000 lb, with flat or V-style inserts, movable grip body, stationary inserts, and

a hand wheel activated sure-grip unit configured for quick attachment and removal. The

body will be made of high-strength aluminum or a steel alloy with anodizing or an oxide

finish.

For the force application module, the concept will utilize a frameless limited

angle torque motor. The details of the motor are shown in the supporting documents

attached. The cost of the motors varies with size and rating.

A Control/Display module box will incorporate a live display of the force and

position of the grips. A closed loop servo drive system will be considered (funds

permitting), that will control the crosshead (free-end grip) motion. Because of the high

forces involved, the drive train will be machined to close tolerances to eliminate

backlash, friction and tear. All components enclosed in the vacuum chamber will be

vacuum-safe and free of interference.

2.5.3 Preliminary Bill of Materials

Qty Description Purchased

or Machined

Price

(estimated)

1 Torque Motor Purchased $2000

1 Motor Shaft Purchased $50

1 Lead Screw Purchased $100

18

1 Transfer plate Machined $100

1 Bearing and pre load nut Machined $100

1 Set (2) of Wedge grips Purchase $2500

1 Optical Encoder Purchased $250

1 Mounting Frame Machine $131

20 Cap screws and washers (Assorted sizes) Purchased $20

1 Interface circuitry Purchased $50

1 Control/Display Panel and Accessories Both $400

1 Vacuum interface Purchased $500

TOTAL ESTIMATED COST (without shipping) $6201

19

Figure 2.3 - Concept sketch - single screw internally driven

20

2.6 Concept 3 – Manually Driven Load Frame 2.6.1 Concept Overview

Concept 3 is a mechanically driven load frame with a drivetrain inside the

microscope chamber. The main features of this concept are screw-tightened grips, a

gearbox to screw drivetrain powered by a hand crank, and a cantilever mounting setup.

The potential benefit of this design is a lighter and cheaper frame, as no plates would be

used for sliding guides. Friction would be reduced without runners or a support plate. As

long as the grips are light and the screws are stiff and strong, bending deflection

shouldn’t be an issue with this design. However, if the screws bend too much, loading

would not be pure axial tension and force readings would not be accurate. As a result the

load must be kept uniaxial. The weight of the support on the fixed end should be fairly

small, as the gearbox weighs less than 2 pounds. Grips could be purchased, but another

cost-reduction option is a custom set of grips to fit the exact specifications of the

chamber. As a result, weight and size are minimized.

2.6.2 Design Features

Similar to other potential concept designs, some of the major components will be

purchased. A generic “black box” version will represent the purchased components until

decisions are finalized in the feasibility assessment. Further analysis will be done on the

gearbox and the other “black box” components after they are selected for the system.

On the fixed end, the gearbox and one grip will be fastened to an L-shaped

support. The support will be stiff, strong, and corrosion resistant to avoid contamination

of the vacuum chamber. The two power screws will support the weight of the free end

and drive it while applying tension or adjusting position before gripping. AISI 304

Stainless steel will be the material of choice for both grip supports because of its

corrosion resistance. All machined surfaces will be free of any cutting oils and

contaminants before installation to avoid contamination of the vacuum chamber.

Turning a screw on the jaw modules will loosen and tighten grips. Ideally, grips

will be purchased from Tinius Olsen, MTT, Instron, or another leading competitor.

Should the motor be eliminated from the design, grips can be purchased without

adversely affecting the budget. Tentative prices are around $2,000 for a set of Tinius-

Olsen grips that is ideal for this fixture. The grips will have a wedge mechanism, moving

21

the jaws closer together as a screw is tightened. The fixed grip will incorporate a load

cell to measure tensile force in the sample. Money saved from eliminating the motor can

be spent on the load and position sensors or possibly a vacuum seal through the chamber

door for an external crank.

A small display box will incorporate a live display for tension and position.

Functions within the display will allow both load and position to be zeroed. Wires will

run through an existing port in the microscope chamber and connect to a position sensor

on the grip and a load cell in the fixed grip.

Cost is the major factor in this design. A feasible design could be created for

$4,000-5,000, but several decisions must first be made. Grips could be machined, but the

issue would become incorporating the load cell in the grips. The potential shortcoming

with this conceptual design is the 15-20 minutes necessary to depressurize the chamber

after each load adjustment. A possible solution is to feed a flexible vacuum-sealed

driveshaft through the chamber wall, allowing tension to be adjusted dynamically. If the

cost limit is not exceeded, this design should be acceptable.

2.6.3 Preliminary Bill of Materials

Qty Description Purchased

or Machined

Price

(estimated)

1 Vacuum-rated gearbox Purchased $1500

1 Stainless steel for crank mechanism Machined $40

2 Power Screws Both $250

1 Load cell Purchased $500

1 Live display for load/position Both $100

1 Fixed end, Free end – grip support Machined $100

1 Base support Machined $50

1 Set (2) of wedge acting grips Purchased $2000

1 Set of assorted wires for electronic controls Purchased $50

50 Cap screws and washers (assorted sizes) Purchased $50

1 Control module for motor Both $100

TOTAL ESTIMATED COST (without shipping) $4740

22

Figure 2.4 - Concept 3 sketch (top)

Figure 2.5 - Concept 3 sketch (side)

Figure 2.6 - Concept 3 gearbox and crank interaction

23

2.7 Concept 4 - External Motor with Two Driving Screws 2.7.1 Concept Overview

The main feature of Concept 4 is that the motor supplying the tension is mounted

on the outside of the machine. The potential benefits of this design are more space within

the chamber for the rest of the load frame and the avoidance of problems with the motor

overheating. In a vacuum, air cooled motors will overheat if they are operated for too

long. The main challenge for this design is maintaining a good driving torque as the load

frame moves into focus under the electron gun.

2.7.2 Design Features

Most of the major components will be purchased. Depending on the price and

size constraints, sizes and locations may change somewhat. As a result, a generic black

box will represent the purchased components until decisions are finalized in the

feasibility assessment. Further analysis will be done on the motor and the other “black

box” components after they are selected for the system.

A vacuum interface for the motor and shaft will be mounted on the door. Because

of tight tolerances, this product must be purchased. A potential solution costs $2625.00,

which is expensive, but may eliminate some other costs in the process. The shaft in the

chamber is rated to 150 ounce-inches torque. The shaft will be connected to a gearbox

that will provide a torque sufficient enough to apply the desired tensile force to the

sample. The motor will power the driving screws, the means of applying tension.

A control box will incorporate a live display for tension and position. The control

box on this motor may aid in controlling the system. Buttons on the live display will

allow both load and position to be zeroed. An additional wire port will be added for the

load cell. The wire pass-through costs about $80.

This design, however, is practically impossible with the existing machine, due to

the support plate for the existing SEM position fixture. The plate enables tilt control and

partially obstructs the port that is targeted for the external motor interface.

24

2.7.3 Preliminary Bill of Materials

Qty Description Purchased

or Machined

Price

(estimated)

1 Motor & Pass-Through (vacuum) Purchased $2750

2 Power Screws Both $250

1 Load cell & Wire Pass-Through Purchased $600

2 Grip supports Machined $200

1 Set (2) of wedge acting grips (Tinius Olsen) Purchased $2000

1 Set of assorted wires for electronic controls Purchased $50

50 Cap screws and washers (assorted sizes) Purchased $50

1 Gearbox Purchased $800

TOTAL ESTIMATED COST $6750

Figure 2.7 - Concept 4 - similar to other concept designs, except for the external motor.

25

3 Feasibility Assessment 3.1 Introduction

Evaluation is a critical part of the design process. Original designs are evaluated

at the concept stage and after the details of a machine have been finalized. Different

evaluation methods are required for different activities. Methods used for assessing

detailed designs are generally inappropriate for the evaluation of design concepts because

the specific details of the purchased equipment are not yet available at the concept stage.

In many cases, proprietary equipment will be selected and incorporated into a larger

design scheme. Therefore methods for comparing equipment are needed. Proprietary

equipment can be evaluated in great detail, but to carry out very detailed assessment of

large machines would be time consuming and, in most cases, inappropriate. In some

cases, only a limited amount of detailed information is available for the proprietary

equipment. Therefore, the equipment needs to be evaluated with respect to the particular

case under consideration. Qualitative and quantitative methods may be used according to

the requirements of individual cases.

In this case, a range of evaluation methods and principles were used to

qualitatively and quantitatively assess the four concepts. The attributes to be included in

this evaluation are resource, economical, schedule, and technical feasibilities. Methods

used for this comparison were Pugh’s method and a weighted evaluation.

3.2 Evaluation of design concepts For this project a weighted evaluation is used side by side with a Pugh evaluation.

In each case the resource feasibility is broken down into sufficient skills, sufficient

equipment, sufficient number of people, and availability of purchased components.

Components of the schedule feasibility include the chances of meeting the intermediate

mileposts, chances of meeting the PDR requirements, and chances of meeting the CDR

requirements. Economical feasibility will be scaled by the percentage of total required

funds allocated. Technical feasibility will be based on distinguished levels as depicted in

the evaluation worksheets.

26

3.3 Pugh Evaluation Pugh’s evaluation was a qualitative evaluation in which design concepts were

compared to a reference design concept. This method is similar to the ‘paired

comparison’ method of evaluation used in creative problem solving. The reference

concept was a derivative of the generic tensile testing configuration, scaled down to size

to fit in the SEM chamber. An evaluation matrix was constructed, (Table 3.1), consisting

of the four concepts, which are compared against the assessment criteria. The reference

concept was chosen as the datum. Each concept was then compared with the datum with

respect to each assessment criterion, or attribute, independently. If a concept was deemed

better than the datum with respect to a certain attribute then a plus sign was inserted into

the matrix for that attribute. If a concept was deemed to be worse than the datum with

respect to the attribute, a minus sign was entered into the matrix. If it was the same as the

datum, or if no judgment whatsoever can be made, then a zero was inserted. Thus the

pluses, minuses, and zeroes for each concept were totaled to complete the matrix. The

matrix highlights the strength and weakness of concepts. The objective is to eliminate

week concepts and to identify those strong concepts that are suitable for further design

work.

3.4 Weighted Concept Evaluation Also known as the systematic quantitative method, the weighted concept

evaluation was performed in six steps: defining the assessment criteria, setting the value

judgment, determining relative importance of criteria, predicting performance, converting

the performance to score values, and computing the overall value.

The assessment criteria resemble that of the Pugh evaluation except that they

estimate relative importance of attributes. Weights were developed through a comparison

of row attributes and column attributes and tallying the row, column, and overall totals.

The relative weights or importance of each attribute were normalized by dividing each of

the attribute total values by the sum of the total values.

Horizontal arrows signified that the row attributes were more important than the

column attributes. Vertical arrows signified that column attributes were more important

than the row attributes and a diagonal arrow signifies no significant difference between

row and column attributes.

27

Value judgment was set by defining a range of performance from an upper value

of perfectly acceptable performance to a lower limit that defines the threshold of

complete unacceptability. A score of 5 was given to the perfectly acceptable concept that

was much better than the reference concept, 4 points were awarded for a concept that was

better than reference concept, 3 implies same as reference concept, a value of 2 is given

for a worse design than the reference concept, and a value of 1 is given when a concept is

much worse than the reference concept.

The overall performance value is computed by the summation of the product of

the weight value with attribute score for each concept. The design concept with the

greatest overall value is the preferred choice. However, if concepts are very close in

score, the designers may not choose the best concept from the feasibility assessment, as

they would be practically equal.

A weighted evaluation was done for both the relative importance of the modules

and for the resource, economic, schedule, and technical feasibility criteria. The weighted

evaluations for the modules were shown in Tables 3.2 and 3.3. The weighted evaluations

for the various design concepts were shown in Tables 3.4 and 3.5.

28

3.5 Results

Table 3.1 – Pugh evaluation of concept proposals

Pugh Evaluation Worksheet

Baseline Concept Internal motor/gearbox -2 screws

ATTRIBUTE CONCEPT 1 CONCEPT 2 CONCEPT 3

Resource Feasibility

R1 Sufficient skills 0 - +

R2 Sufficient equipment 0 0 +

R3 Sufficient number of people 0 - +

R4 Availability of purchased components 0 - +

Economical Feasibility

E1 % of total required funds we have 0 - +

Schedule Feasibility

S1 Chances of meeting the intermediate mileposts 0 - +

S2 Chances of meeting the PDR requirements 0 - -

S3 Chances of meeting the CDR requirements 0 - -

Technical Feasibility**

T1 feasibility Level (L0, L1, L2, L3, L4) 0 0 0

Scores 0 7 4

Number of Pluses 0 0 6

Number of zeros 9 2 1

Number of negatives 0 7 2

Conclusions:

**Level L0: Are we trying to break the laws of science?

Level L1: Are fundamentally new inventions required?

Level L2: Has a similar technology been used before (by anybody)?

Level L3: Has the technology been demonstrated by our team?

Level L4: Has the customer tested the technology?

29

Table 3.2 - Relative weights for module importance to overall design

Table 3.3 - Weighted importance of modules for concept proposals

Table 3.4 - Relative weights for weighted concept evaluation

30

Table 3.5 - Relative weights for concept evaluation

Weighted Evaluation Worksheet

Baseline Concept Internal motor/gearbox -2 screws

ATTRIBUTE RELATIVE WEIGHT#

CONCEPT 1

CONCEPT 2

CONCEPT 3

Reource Feasibility

R1 Sufficient skills 0.083 3 2 4

R2 Sufficient equipment 0.097 3 3 4

R3 Sufficient number of people 0.014 3 2 4

R4 Availability of purchased components 0.153 3 2 4

Economical Feasibility

E1 % of total required funds we have 0.194 3 1 5

Schedule Feasibility

S1 Chances of meeting the intermidiate mileposts 0.028 3 1 4

S2 Chances of meeting the PDR requirements 0.167 3 1 2

S3 Chances of meeting the CDR requirements 0.194 3 1 2

Technical Feasibility**

T1 Feasibility Level (L0, L1, L2, L3, L4) 0.069 3 3 3

RAW SCORE 2.997 1.581 3.399

NORMALIZED SCORE 0.999 0.527 1.133

**Level L0: Are we trying to break the laws of science?

Level L1: Are fundamentally new inventions required?

Level L2: Has a similar technology been used before (by anybody)?

Level L3: Has the technology been demostrated by our team?

Level L4: Has the customer tested the technology?

31

3.6 Conclusion Based on the results of the two evaluation methods, two concepts will be

considered for future development. The main focus, however, will be devoted to the 2-

screw internal motor concept, which yielded the greatest overall performance value of 3.

This concept will consist of six modules. The concept will use a machined set of

threaded grips, an internally mounted stepper motor, a twin-screw drivetrain, external

mounted control and display panel, a fixed mounting frame, and a vacuum-friendly

interface for wire feedthroughs. There is a similar working model at Lehigh University.

The other concept, with the second largest overall performance value of 2.931, is

concept 3, which differs from the latter by employing manual power instead of the torque

motor. The concept will be pre-loaded by a hand crank prior to closing the SEM

Chamber. This concept will also serve as a backup plan, depending on the economical

feasibility of stepper motors and other purchased components for the preferred concept.

32

4 Objectives and Specifications 4.1 Design Objectives

The design objectives for the S.E.M. Load Frame, the key purposes and goals for

this project, must be met while taking into effect the constraints of the vacuum chamber.

The load frame must fit within the scanning electron microscope’s vacuum chamber. As

the load frame must work inside the vacuum chamber of the electron microscope

compartment all components of the frame must be designed to work safely within the

vacuum environment. While inside the vacuum chamber, the load frame must not

interfere with the electron gun or the detector. Such interference would compromise

image quality and possibly damage the machine if components come into contact with

each other. At least one set of working grips will be included in the load frame’s design

and be able to handle at least 2000 lbs of tension. These main grips will hold cylindrical

specimens with threaded ends. Samples will be easy to install, since the work needed to

install a sample will be minimal.

As the load frame will be the property of Dr. DeBartolo, it will have to be

modular and easy to install and remove from the load frame. In addition, the module will

be stored in the mechanical engineering facilities, but used at the CIMS building. As a

result it must be lightweight and easy to carry between buildings.

4.2 Performance Specifications The load frame must meet several performance criteria. In order to meet the

customer’s minimum needs, the machine must support a load of 2000 lb in tension and

200 lb in compression. For safety and logistical reasons, the machine must have the

capability of both position and load control, with a switch that allows the operator to

select either option. Position control is needed for adjusting the movable grip and

securing the test specimen. Load control will drive the sample during the tensile test, but

in the event of a fracture, position control may be used to halt motion of the grip.

Since the customer desires a display for load and specimen deflection an

automated control system will incorporate live displays for load and position. The

control system will remain outside of the chamber and will have a power source separate

from the SEM.

33

4.3 Design Specifications/Implementation Due to the design of the existing SEM position fixture, the load frame will be

customized to that specific machine. The position fixture contains several stepper motors

that control x, y, and z position, tilt, and rotation. Budget and time constraints prohibit

designing a new position fixture to accommodate the load frame. As a result, the rotation

module will be removed to make room for the load frame and provide a mounting

location for the load frame without removing the position fixture. To allow for easy

mounting, the design must be modular, consisting of several components that easily

assemble.

The operation of the machine involves a series of mechanical devices to achieve

appropriate sample tension. The drive train consists of a stepper motor, which couples to

a gearbox, driving a series of spur gears, which in turn drive a pair of ACME thread

power screws. One grip will remain fixed while the two ACME screws drive the other

grip. Custom designed grips will accommodate threaded cylindrical samples with a 1”

gage length. The motor control is located outside of the SEM vacuum chamber, so the

wires must pass through a vacuum-tight feed-through in the chamber door.

4.4 Evaluation Criteria Before installation of the product, it must meet a set of evaluation criteria. All

part designs will be verified by modeling in Pro Engineer and critical parts will be

evaluated for stress and displacement in ANSYS. Fatigue calculations will ensure that

the power screws and any fasteners will hold for the lifetime of the module (105 cycles

should be sufficient). In addition, torque calculations for the power screws will ensure

that the motor will meet its performance standards and avoid stalling. After verifying the

performance capability through simulation and development testing, the load frame will

be tested outside of the vacuum chamber by loading a sample specimen to maximum

tensile and compressive loads.

Other important criteria for the machine’s performance that must be met are as

follows: If a sample fails in the vacuum chamber, there will be safety constraints to

ensure that the sample will not damage the microscope. The drive train must handle the

specified loads of 2000 lb tension and 200 lb compression. Samples will be tightly

secured and should not loosen during operation. The module must fit in the chamber and

34

be easy to install. The surfaces of all parts must be corrosion resistant and free of any

impurities and oils to avoid contamination of the vacuum chamber. The automatic

controls must function while the chamber is sealed and provide a live load and position

feedback within reasonable accuracy.

4.5 Safety Standards Though no ASTM standard exists for vacuum tensile tests, several safety

precautions must be implemented in the design of the load frame. Vacuum environments

can be damaging to an air-cooled motor if it is operated for a long time, so motor use

must be limited. A regular motor may also experience outgassing in a vacuum, but a

vacuum rated motor would ensure no outgassing. Parts must be secured to the position

fixture inside the vacuum chamber. Due to the vacuum pressure and sensitivity of the

equipment, the system may take a few minutes to pump back to atmospheric pressure

before the chamber can be opened. Therefore it would be impossible to stop the pump

and immediately retrieve a loose part without inflicting any damage.

The effects of the electron beam may result in two problems, grounding and

overheating. As electron beams are fired at the surface, the load frame may become

charged during operation. It is important to note that a grounded connection must be

established with the fixture. To ground the load frame, an existing grounding clip from

the rotation module will be attached to the base of the tensile frame’s support structure.

As long as the motor is kept from overheating, heat will not be a problem for the fixture.

The electron beam will not cause a significant temperature change. The operators of the

machine stated that plastics and other insulators might have problems with melting

because they absorb the energy of the beam, but metallic structures will not experience a

considerable temperature change.

The existing position fixture was not designed to support heavy loads, and weight

must be minimized. Using the analysis features in Pro Engineer, weight can be

monitored for individual parts and for the entire assembly. The load frame’s weight must

not hinder the functionality of the existing position fixture. Most importantly, it must not

damage the position fixture.

35

5 Analysis and Synthesis 5.1 Design Structure Matrix (DSM)

5.1.1 Introduction

A software program run within Microsoft Excel, Design Structure Matrix,

analyzed the dependencies of various design decisions. A DSM model of the design

process was used to quantify a process configuration and lead to a logical order of design

decisions. Cost, duration or schedule, and variances in both are largely a function of the

number of iterations required in the process execution and the scope, or impact, of those

iterations. Since iterations may or may not occur (depending on a variety of variables),

this model treats iterations stochastically, with a probability of occurrence depending on

the particular package of information triggering rework.

This model characterized the design process as being composed of activities that

depend on each other for information. A change in that information results in a ripple of

change in the design. Thus, an alteration to one activity can cause a chain reaction

through supposedly finished and in-progress activities. Reworking is a function of the

probability of a change in inputs and the impact of the change in inputs. The model also

assumes independent activities can work concurrently.

5.1.2 Problem Statement

The goal of this procedure was to map the whole picture of systems and parameter

relationship to aid in understanding the implications of changing any parameter to the

others. It was limited to the scope of the project in two respects; first it is used after a

number of parameters have been established and second, it was a new technology that

needed to be tried and tested in the project due to high dependencies of parameters

associated with this project.

5.1.3 Results analysis

A total of nineteen system attributes were mapped out and modeled in the DSM

software. After three iterations it was realized that the dependencies are of high nature

and no further candidates were available to complete the bottom half as expected. It was

noted, however, that only sample size, load cell and applied loads could be addressed

independently with minimal impact of rework on other parameters. As a result, further

testing of the DSM was halted, as no iteration would produce the required results. In that

36

case the remaining attributes with feed forward dependencies will be treated one at a time

by assuming desired parameters and adjusting the rest of the others accordingly.

Figure 5.1 - DSM Matrix for the tensile load frame

5.2 Motor and Gearbox Selection The selection process for the motor and gearbox required the resolution of the

following issues: vacuum compatibility, torque requirement, size constraints, power

requirements, and cost. Vacuum compatibility greatly limited the field. Contact with 25

different motor manufacturers – makers of low force linear actuators excluded – yielded

referrals to three companies: Danaher Motion, Empire Magnetics, and Bay Side Motion.

Danaher was unable to assist the team as they did not make vacuum rated equipment, but

offered Empire Magnetics as a referral. While Bay Side is able to provide gearboxes

suitable to the laboratory vacuum, they do not manufacture motors with the capability or

gearboxes beyond a 100:1 ratio in the size range required. Also, the minimum pressure

of Bay Side’s vacuum rating was not available. Cost served to eliminate the servomotor

as an option. With servomotors, the necessary onboard components are expensive to

dissipate large quantities of heat from the motor in a vacuum environment. As a result, a

stepper motor was chosen.

37

Size and cost constraints make the selection of a NEMA 17 frame stepper motor

coupled with a 700:1 gearbox necessary. Potential solution included a complicated bevel

and worm gear setup or larger motors, which with a 100:1 gearbox (Section 5.3) is next

to impossible to fit, without even considering how to handle axial thrust control of the

gears. Empire Magnetics makes a 700:1 three-stage gearbox in both NEMA23 and

NEMA17 sizes. The NEMA17 can only transmit a maximum of 600 lb·in continuously;

loading over 600 lb·in will severely shorten gearbox life and it could fail at any time.

The NEMA23 gearbox can transmit 1040 lb-in before its life is threatened. Team chose

to adapt a NEMA17 motor to a NEMA23 gearbox, saving $400 compared to purchasing

the larger motor.

There are other advantages to using a NEMA17 motor instead of a NEMA23,

which include safety and simplicity. With a maximum motor torque output of 2 lb·in and

a gearbox efficiency of 0.72, it is impossible for the NEMA17 motor to damage the

NEMA23 gearbox. The maximum torque applied by the gearbox is 1008, based on the

motor limitations using Equation 5.1, where Tmax is the maximum torque, Ng is the ratio

of the gearbox (700:1), and Tm is the torque applied by the motor. Also the slimmer size

of the NEMA17 motor allows for positioning such that a 2.5:1 cylindrical gear ratio can

be used while keeping the center lines of the motor, gearing, and power screws in the

same plane, saving material cost and simplifying the machining setup.

ggmg NTT ⋅⋅= η (5.1)

Use of a three-stage continuous 700:1 gearbox was a costly, but worthwhile

decision. With a 100:1 ratio, the number of independent gears outside of the gearbox

could number at least 6, and require the use of bevel or worm gearing to provide

additional reduction. The efficiency losses and axial thrust created with these gearing

combinations have not even been fully calculated, because the team unable to find a

combination to even meet the space constraints and preserve the safety of gear shafts.

Each complexity in the power train introduces additional possibilities for error, additional

complexities of manufacture, and additional tolerances. A 700:1 gearbox requires the

precision alignment of three holes and precision thickness of the base housing. These

parts were easily fabricated in the machine shop.

38

5.3 Drivetrain Design The fundamental problem for this design was to determine how to apply a

maximum tensile load of 1000 lb to the test sample. Though 1000 lb was the maximum

required load, the system was designed to handle the ideal load of 2000 lb. Most

mechanically driven tensile testers use some type of power screw, usually ACME power

screws. ACME power screws were chosen for this design because of their self-locking

capability. When the motor is turned off, the screws hold their position, avoiding any

unnecessary loads on the motor. In order to provide the axial tension in the test sample,

the screws are in compression to balance the forces. Provided that the two screws are the

same size and properly aligned, they will share the load equally. Thus, a maximum

compressive load of 1000 lb will be applied to each power screw when the sample is

loaded to 2000 lb tension. A torque must be transmitted to the power screw to obtain the

axial load in the sample, so a system of three cylindrical gears was devised to couple the

power screws with the motor and gearbox, obtaining the necessary speed reduction.

The torque to apply the tensile load determined what gear reduction was

necessary in the drivetrain and gearbox. The maximum required torque to turn the

screws is 200-250 in·lb, depending on lubrication and bearing resistance, as determined

by Equation 5.2 if the sample load was 2000 lb, the maximum load that the screw will

experience during operation. W is the load on each screw (1000 lb), dm is the pitch

diameter of the screw, the coefficient of friction is f for lubricated friction (.15), L is the

lead (0.1 inches), and αn (10°) is the lead angle. Due to the lack of suitable governing

equations for roller bearing friction, and because of the low shaft speed, lubricated

coulomb friction was used to provide a conservative estimate of bearing friction, a

function of shaft speed. The second term of the equation is the friction calculation for a

collar, treating the bearing like a collar with a very low friction coefficient.

2coscos

2,cc

nm

nmmLR

dWLfd

LdfdWT ⋅⋅+

⋅⋅⋅±⋅⋅

⋅⋅

=µ

απαπ

m (5.2)

Frictional losses in the drivetrain and efficiency in the gearbox (72%) necessitate

a gear reduction ratio of 250:1 to achieve the necessary torque in the power screws. With

a 100:1 gearbox and a 2.5:1 spur gear ratio, the factor of safety would be 1 while

operating the motor at 1 in·lb. Using the 700:1 gearbox instead of the 100:1 gearbox, an

39

additional 2.5:1 reduction by spur gears is sufficient to reach the maximum torque and

include a factor of safety of 5. Though this is an excessive level of safety, no

intermediate gear reductions were available. The spur gears on the power screws (2.5

inches) are connected to the gearbox by a 1 inch pinion gear. For the spur gears, they

will survive the maximum load. The maximum stress per tooth is around 90 ksi when the

machine is operating at maximum capacity. After computing gear stresses, shear stresses

were computed for the shafts using Equation 5.3. The stress on the output shaft can be

found using equation 5.3.

3

16dT

⋅⋅

=π

τ (5.3)

The major problem with the drive train was connecting the spur gears to the

power screws. Machining the component from a single piece was not even a possibility,

considering the expense of cutting an ACME thread and the time required to machine a

2.5 inch diameter shaft (needed for the spur gears) down to a 1 inch power thread. A

contact with years of machine shop experience estimated that the cost would be $4,000-

5,000 for the two power screws. The best solution was to purchase an ACME threaded

rod, spur gears, and appropriate bearings, washers, and fasteners. The method of joining

these parts came under much scrutiny. A square broach, a pin, and a key were considered

as methods of fastening. The pin would be the simplest design. Cutting the thread off of

the end of the ACME rod and drilling a matching hole in the collar, the components

would be inserted into each other and a hole will be drilled for a dowel pin. Keys are

good at transmitting torque, but are weak in axial loading. A broach is functional, but the

square hole would be costly to machine with wire EDM (electrical discharge machining).

The pin was able to transmit the torque from the gear to the collar and from the collar to

the shaft. The stresses in the pins were 45 ksi and 34 ksi for the gear/collar and

collar/thread, respectively. However, to achieve a safe pin, the diameter was ¼ inch with

a 1 inch power screw. Collar length would increase to allow more thickness to account

for the hole.

The final design was to select a larger ACME shaft (1-10) to allow material for

stepping down the end of the shaft to accommodate bearings and gears. Drilling through

the center axis of the gear, the gear fits on the 0.500 diameter at the end of the shaft, held

40

in place by the diameter step, washers and a 10-32 screw. The 5/8 inch diameter shaft

section passes through a clearance hole in the fixed end and is held by a radial load

bearing inside the fixed end. Stress in the power screw is not a problem (Section 5.8), as

a vacuum safe lubricant was used to reduce friction. PTFE thickened Krytox vacuum

lubricant was viscous enough to remain on the components inside the vacuum and

provide the necessary protection. Bearing stress in the thread is about 22 ksi, not

lubricated, and about 7 ksi lubricated, which are both below the yield strength of 30 ksi

for AISI 316 stainless steel.

5.4 Gripping Mechanism 5.4.1 Specifications of the Grips

One of the most crucial components to provide stability and safety in the SEM

Tensile Stage is a set of grips to secure the metallic specimen. A machine worth several

hundred thousand dollars is potentially damaged if a sample is released while under load.

Without a firm grip, the tensile test results would be null and void. It was in the best

interests of the team to properly analyze the gripping problem and design a viable

solution.

When designing the grips, a major decision was whether the grips should be

designed for predetermined specimen geometry or if the specimen should be designed

based on the grips. This was easily determined since a standard ASTM specimen exists

for a variety of sizes and samples could easily be acquired or manufactured [4]. As a

result, the grips were designed first with the specimen geometry adjusted to fit within the

grips.

Grip specifications were determined by the environment and stress conditions for

the operation of the machine. As with the design of other components within the SEM

tensile stage, it was important to safely operate in a vacuum pressure of 10-7 torr, and this

component was no exception. Also, it is very important that the grips do not contain any

grease or oil to avoid contamination of the SEM chamber. Since the design goals for the

tensile stage are to load a specimen to a maximum of 2000 lbs tension and 200 lbs

compression, it would be best to design the grips to be able to withstand a tensile of force

of 3000 lbs, a factor of safety of 1.5. Most importantly, the grips must also be small

enough to fit within the narrow packaging envelope of the tensile stage.

41

5.4.2 Purchased or Machined?

With a number of purchased components already in the design, the team faced

mounting costs. The original idea was to purchase a set of small, flat, grips. Initially, it

was decided that these grips should be purchased instead of designed and machined

because of time constraints and machining issues. It became evident, unfortunately, that

the specific type of grips that were needed would be too expensive and consume too

much of the budget (Table 5.1).

Table 5.1 - Grips available for purchase Manufacturer and Type Load Capacity (lb) Cost

MTT – Flat Wedge Action Grips 5000 $4,800

Tinius Olsen – Flat Wedge Action Grips 2250 $1,910

Instron – Remanufactured Wedge Action Grips 5000 No response

With potential prices that consumed a large portion of the budget, several cheap

alternatives were considered. Some of which included self-locking grips and a custom

design that for a flat dog bone specimen, but the flat designs didn’t allow for the samples

to be examined across both width and thickness. Therefore, the team decided to design

and manufacture a set of grips for a cylindrical sample with male threaded ends. This

design was met with approval from the team because it solved three major problems. The

grips would be very easy to manufacture. A 1-inch diameter cylinder of stainless steel

could be drilled and tapped with the sample’s thread size. Given the machine’s small

height envelope of 3 inches, the size of the grips is ideal. The cost to manufacture these

grips would be much less than $100, freeing up funds for other design features.

5.4.3 Final Design of Grips

The grip is simple to manufacture and easily replacable. Created from a small,

cylindrical piece with a threaded through hole on the center axis, the free end grip is 1

inch in diameter and about 1.5 inches long. The hole will be tapped with a 3/8-24 thread,

the same size of thread used on the Futek load cell that is designed to handle up to 3000

lbs. The fixed end will use the load cell as a grip. The sample threads into the female

end of the load cell and the male end will of the load cell is threaded into the fixed end of

42

the frame. The material choice is AISI 303 stainless steel because of its corrosion

resistance and machinability.

As with any design, the decisions were verified through calculations. To just trust

the thread specifications for the Futek load cell is not enough to ensure that the threads

are safe. It is advisable to check stresses to make sure the threads can withstand the force

applied to the grips. This is the only calculation that is of any concern to the grips

because the weakest part of the grips is going to be the threads. The following equation

gives the stress applied to one thread, given an applied load:

tp

ddP

i

⋅−⋅⋅

=)(

422π

σ (5.4)

P is equal to the force applied to the threads, which in this case is 2000 lbs, and

d²-di² is the tensile area of the 3/8-24 threads. The variable t/p is the number of threads in

which the load is being applied. For this case the length of the male end of the specimen

fitting into the grip is designed to be 9/16 inches. At 24 threads per inch, a total of 13.5

threads are in contact. Therefore, the average stress per thread is around 6.75 ksi, which

is safely below the yield strength of the material (30 ksi) of the grips.

5.5 Base and Frame The base and supporting frame were designed around the existing SEM position

fixture and the purchased load frame components. The base of the load frame must serve

three purposes: to act as a platform to support the load frame, to act as an apparatus to

hold the SEM’s longitudinal position resolver so that the sample can be positioned under

the microscope, and to act as a means to connect the test specimens to the SEM’s

grounding path.

To support the load frame, the base will be composed of two pieces. For

corrosion resistance and strength, the base is machined from AISI 303 stainless steel bar

stock. A pattern of holes in the center of the plate matches that of the SEM position

fixture base plate. A set of three holes is used for mounting the fixed end to the base.

The base plate must hold the position resolver, a function that was previously done by a

module that is removed for the installation of the load frame. The position resolver

mounts on the side of the base in the same location where it would normally be located

on the removed rotation module. With the position resolver mounted to the base plate,

43

the existing position control for the stage will function normally. In regards to providing

a path to ground, the plate must be composed of sufficiently conductive steel to avoid a

buildup of charge in the load frame. It is an important facet, however, because without a

path to ground the electron gun would charge up the specimen quickly and a discharge

could damage the equipment inside the SEM. However, the SEM is equipped with a

warning to alert the operator when the system is not properly grounded.

The fixed end plate is the critical component to manufacture, controlling

alignment and position of various components. On the plate, mounting holes are

precision machined to allow for the placement of the gearbox/motor assembly, base plate,

power screws, and the necessary bearings to align the power screws. Composed of 1.25-

inch plate stock, deflection should be minimal, despite the number of mounting holes for

the motor, base, and load cell. Stress and deflection were verified in ANSYS (Chapter

7). As this component was manufactured accurately, assembly and alignment issues were

be negligible.

Acting as the crosshead, the free end travels along the power screws. Supported

by the screws, the free end crosshead is also composed of 1.25-inch thick AISI 303

stainless steel for manufacturability and corrosion resistance. The two holes to

accommodate the power screws are threaded with a 1-5 2G ACME tap. An additional

mounting hole locates the bolt that secures the threaded grip.

5.6 Control and Display 5.6.1 Beginning Estimates

To generate the

necessary torque at

reasonable speeds, it

has been decided to run

the stepper motor with

an applied voltage of 24

volts. At this potential

the motor should be

able to reach a

maximum speed of

44

Figure 5.2 - Torque vs. speed curves for the Empire Magnetics U17-2 vacuum rated stepper motor. This graph demonstrates that at 24V a maximum speed of nearly 20 rotations per second can be maintainedand still generate the necessary torque (15 oz inches).

nearly twenty rotations per second (RPS) and still generate the desired torque of about

15oz inches (Fig 5.2). With a complete gear ratio of 1750:1, and ten threads per inch on

the ACME screws, this translates to free-end movement at a rate of about an inch every

14.6 minutes (875 sec/rotation). The motor requires 200 steps/rotation, when each of four coils is

activated in turn. To generate the ‘parallel’ curves seen above, these four coils are wired

in parallel to function as two coils. This generates a slower drop off in the torque speed

curves and changes the resolution per step to 3.6 degrees. Thus there will be only 100

steps per rotation for our application. At a rate of 20 rotations/second requires the controller to

have the ability to drive the motor with a 2 kHz signal to reach the maximum allowable

motor speed. However, due of the characteristics of the controller device, this upper limit

will never be achieved using this setup.

5.6.2 Control response constraints

To reduce expenditure, a method of control was found that could be implemented

for no monetary cost. This choice, however, comes with cost in performance. A

Measurement Computing device called the miniLAB 1008 was received from the

Mechanical Engineering department for no cost. This device has no ability to do

processing independent of the host computer. It is connected to the host via a USB 1.0

connection. Thus its abilities are all limited fist by the bandwidth of the USB 1.0

connection and second by the software response time of the host computer. The

miniLAB 1008 has two analog output channels that have a collective output rate of 100 samples/second. This means that one channel can be used to run 100 samples/second, or both

channels can be operated simultaneously at a reduced rate of 50 samples/second. The result is

a maximum driving frequency of 100Hz, a maximum speed of one rotation/second. With the

gearing this means movement of the free end at a rate of 57µinches/second. Although this rate

of movement is seemingly very slow (one inch takes just under 5 hours to traverse) the

rate of change in load is much more substantial.

5.6.3 Sample Response Characterization

Through relatively simple calculation, using the relation of stress and strain

allowed by the modulus of elasticity, the applicable rate of change for the load on the

sample can be revealed. The stress on the sample is equivalent to the force applied over

45

the sample’s cross-sectional area (Eq 5.5). Strain is the change in length of the sample

proportional to the original length of the sample before force was applied (6).

AF

=σ (5) llS ∆

= (6) SE *=σ (7)

A stress and strain are related through the modulus of elasticity (7). This formula can

then be manipulated, and the rate of change of length substituted for ∆l to estimate a rate

of change in the load applied to the sample. Thus: l

EAlF ∗∗∆=∆

Assuming the

5.3, the rates of cha

assuming that the p

output frequency of lbs/sec. This, as we

Table 5.2 – Assumed worst case sample characteristics

Sample characteristics

Length 1.25 (inches)

Radius 0.125 (inches)

Modulus Elasticity 9900000 (PSI)

Table 5.3 - Rate of sample loading for various motor speed characteristics Loading Characteristics

Motor Speed Rate of Load Change Time to run full 2200 lb Range

(RPS) (lbs/sec) (sec) (min)

0.0625 1.39 1584.5 26.4

0.125 2.78 792.2 13.2

0.25 5.55 396.1 6.6

0.5 11.11 198.1 3.3

1 22.22 99.0 1.7

10 222.16 9.9 0.2

20 444.31 5.0 0.1

characteristics of the cylindrical sample to be as are state in Table

nge for various rotational speeds are shown in Table 5.4. Thus

arallel connection reduces the step size by half, and a maximum

100Hz the device can achieve 1 RPS and a loading rate of ±22.2

ll as data for other operational speeds is shown in table 5.4.

46

Pessimistically the team can assume 2 – 5 minutes to move the entire valid load range of

the device at the highest motor speed possible with the miniLAB 1008TM.

5.6.4 Hardware Design Fundamentals

Figure 5.3 - The MOSFET H-Bridge used to drive the stepper motor coils. A single motor coil is simply modeled here as a 10 Ω resistor (R1). This model was developed to determine if this configuration could source the necessary power to the motor coil. Additions to this circuit in implementation include capacitances to quite oscillations caused by the coil’s inductive nature and level shift conditioning of the miniLAB 1008TM output. Here the control voltages are represented by V3 and V4, sinusoidal voltage supplies.

Figure 5.4 - Simulation of MOSFET H-bridge seen in figure n+1 demonstrating proper application of power to the motor coil for voltages sufficient to saturate the MOSFETs. The 10Ω resistor modeling the motor is subject to 1.81Amps at 18 volts, values satisfactorily close to our desired values of 2 Amps and 24 Volts.

47

This analysis asserts that a 24 volt driving voltage would be feasible and produce

a reasonable rate of control. Thus a 24 volt 50 Watt single ended power supply was

acquired. The fact that the power supply is single ended limits the precision of control

without exponentially increasing the complexity of the driving mechanism. Thus a

configuration of transistors commonly referred to as an ‘H-Bridge’ was implemented to

provide the required bi-polar drive. The H-bridge was implemented using metal oxide

field effect transistors (MOSFET) to provide maximum efficiency at high currents. A

model of the circuit was assembled to prove the functionality and feasibility of the

design. The model is shown in Figure 5.4 and the results in Figure 5.5. The probe colors

match the plot of the specific variable measured at that point. All plots are voltages with

the exception of the blue line which represents the flow of the current through the motor

coil. The output of the miniLAB 1008TM can be fed through a level shift and directly into

the H-Bridge which will in turn feed a motor coil, thus control of the motor is

accomplished.

A 10 volt excitation voltage must be fed to the load cell so that output can be read.

This will be generated from the 24 volt supply and fed into the load cell. Noise on the 24

volt power supply is expected to be a major issue while the motor is running due to the

inductive nature of the coils. All attempts possible will be made to quiet the effect of the

motor on the power supply. Feeding the sensor from the noisy 24 volt power supply is

feasible because sampling of the sensor signals cannot be achieved simultaneously with

the movement of the motor due to limitations of the miniLAB 1008TM. Thus to read

sensor signals the motor will be stopped and measurements will be taken in a fraction of a

second. This will allow the settling of the supply voltage and add minimal delay to the

achievement of the user desired load.

Filtering of the sensor signals will require only a low pass stage to filter out

excessive high frequency noise as the signals to be measured are expected to change

relatively slowly (temperature and load). Any additional or more complex filtering or

numerical analysis can be achieved through LabVIEW functions, or virtual instruments

(VIs).

48

Figure 5.5 - A visual summary of the hardware layout and connections.

5.6.5 Control Software

With all aspects of the hardware interface laid out as can be seen in Figure 5.6, all

that remains is the software to interface the system to the user. The sponsor of the project

Figure 5.6 - This figure shows the general layout of the LabVIEW program. A series of VIs interprets the users commands and outputs them to the miniLAB 1008tm, which in turn dictates the signals to the tensile stage and interprets inputs from the stage

preferred that control be implemented through National Instruments LabVIEW version

6.1. The miniLAB 1008TM came with “Universal Libraries for LabVIEW” that offer

49

functions, or VIs, for interacting with the miniLAB device. The software that is provided

to control the tensile stage is built up from these universal libraries and as such, can either

be used as is by the end user, or merely as a set of guidelines for defining their own

functionality for the machine.

There are two main types of VIs provided as a result of this project. First there

are VIs that prompt, direct, and interpret

the user’s commands into signals to be

output to the device. The second type of

VI simply outputs the generated signals

to the device, or receives data from the

t

p

i

t

m

d

o

a

Figure 5.7 - An example GUI demonstrating the range of control devices, and their possible applications in the case of the tensile load stage. The GUI interface will be easily end-user definable.

tensile load frame. The VIs that send

and receive signals to and from the

device are built most directly on the

Universal Libraries for LabVIEW that

came with the miniLAB1008TM. The

subVIs, as they are called, that were

developed for this project in this

situation mainly serve to simplify the

interaction with the miniLAB 1008TM

device, and make programming in this

environment more directly oriented with

he resulting actions of the tensile stage. This is intended to simplify the development

rocess for an individual who may not fully understand the functions of all the devices

nvolved, but knows how they want the tensile stage to respond to their commands.

The other type of VI developed for this project is oriented more towards defining

he user interface and dictating a sequence of actions to the user to set up and use the

achine. These VIs would include signal generation and user prompts. A layout of a

eveloped function can be seen in Figure 5.7.

Using these VIs, the user can take inputs and outputs and generate whatever plots

r GUI devices they want or need for their application of the tensile stage. Figure 5.8 is

n example interface that shows several of the control input/outputs available for GUI

50

development in LabVIEW. The tensile stage software will come with a default GUI, but

the end user will always have the option of easily designing custom interfaces or

functions.

5.7 Vacuum Interface In the design of an electromechanical system to in a vacuum, an important

consideration is the transfer power and data signals through the chamber wall. One ideal

port is available in the SEM door for transmitting necessary signals. Motor signals,

temperature feedback, and the load cell data travel through the door to the control module

in the laptop computer. Major companies manufacture or sell these feed-throughs

included Nor-Cal Products, MDC Vacuum Products, ISI, and PTL vacuum.

The port available in the SEM has a diameter of 1.75 inches. After viewing

several schematics, it was found that a 2.75 inch flange was required. At 10-14 wires

need to pass through the chamber wall, 4-8 for the stepper motor, 2 for the thermocouple,

and 4 for the load cell). The first option was from MDC, which consisted of a twenty pin

feed-through. Model number MDC647055, it protrudes into the chamber with 3.14 inch

pins. An advantage of this feed-through is the possibility of upgrading of the load frame

with the inclusion of a position encoder. This is a feature that would not have been

accommodated by other components under consideration, which included a ten pin

feedthrough and a 9 pin D-type connector.

The wires are removable from the feed through to allow for easy installation and

removal. Using crimpers, the wires for each component are attacted to a ceramic disk for

spacing and insulation. The disk and crimper assembly attaches to the pins on the feed

through, which is coupled to the airside plug. Wires from the airside plug are routed into

the control box.

5.8 Stress Calculations for Critical Components Though fatigue life and yield strength were initial concerns, the design is safe for

operation. The excessive factor of safety in some components allowed everything to be

safe in fatigue, as fatigue life for stainless steel is generally equal to or slightly greater

than the yield strength, especially for low cycle fatigue [9]. Calculations shown below

were the most critical in the design process.

51

Table 5.4 - Selected stress calculations Part Section of Part Stress Allowable Stress FS

------------ ---------------------------- KSI KSI ------------

Free End Vertical Section Through ACME

CTR Line 4 30 7.50

Free End Vertical Section Through CTR 10.61 30 2.83

Free End Bolt Cross Section at Thread 22.36 30 1.34

Free End Bolt Individual Thread 6.94 30 4.32

Cylindrical

Fastener Cross Section at Thread 2.96 30 10.14

Compression

Collar Cross Section 0.3 30 100.00

Cast Iron Nut Cross Section 4 43.5 10.88

ACME Shaft Thread 4.5 30 6.67

ACME Shaft 1"-10 to 5/8" Step 3 30 10.00

ACME Shaft Radial Load Bearing (Maximum

Bending Moment) 15.77 30 1.90

ACME Shaft 5/8" to 1/2" Step 17.35 30 1.73

ACME Shaft Keyway 15.64 30 1.92

Needle Roller

Thrust Bearing Spec: 2170lbf @ 11,000 RPM

1100lbf

@0.7RPM 2170lbf

MINIMUM

1.97

Needle Roller

Radial Bearing Spec: 2330lbf @ 16,000 RPM 266lbf @ 0.7RPM 2330lbf

MINIMUM

8.76

Fixed End Vertical Section Through CTR 8 30 3.75

Gearbox 1008lbf-in MAX Production;

1040lbf-in MAX Allowable 220lbf-in Design 1040lbf-in MAX

MAXIMUM

4.73

Pinion Gear 60.4KSI Sim, T=1040lbf-in@

0.86RPM, large Face Wd 64.86 180 2.78

Spur Gear All Gears 8620 Case Hard (Must

be for shock load) 25.95 180 6.94

Base/ Fixed End

Connection

Screws Designed to Withstand

Maximum Possible Moment Due

to Externally Applied 100lbf

Force

N/A N/A N/A

52

6 Preliminary Design Several changes were made to the preliminary design and are discussed in detail

in Chapter 8. The initial cost estimate for the preliminary design was $7,500. At this

stage of the design, the necessary components and materials were purchased from

suppliers.

6.1 Part Numbers for Pro Engineer files -1st digit =1 if component, =2 if assembly

-2nd digit =0 if purchased, =1 if manufactured

-3rd digit = module number (1=motor/gearbox, 2=drivetrain, 3=grips, 4=base/frame,

5=control, 6=vacuum interface, 7=miscellaneous

- 4th and 5th digits=component number within module (01, 02, etc.)

Figure 6.1 Preliminary Design

53

7 Engineering Models 7.1 Part and Assembly Modeling

7.1.1 Modeling with Pro/Engineer

In addition to modeling and assembling the components in three dimensions

before manufacture, computer aided modeling offered several other advantages. In

transferring the models to detail drawings, Pro/Engineer allowed for the analysis of

fabrication and assembly issues before any of the components and materials were

purchased. Manufacturing issues were considered as the components were created and

assembly issues were considered as the components were assembled. Critical dimensions

and tolerances were resolved with the use of the analysis features, which allow for

checking fits through measuring distances, weight analysis, global clearance, and global

interference. Pro/E served a final check to review all of the dimensions before the parts

were manufactured. In addition, the assembly drawings with regular and exploded views

were created from the models and a bill of materials was automatically generated. All of

the mechanical parts and some of the electrical components were modeled in

Pro/Engineer (Figure 7.1). A great deal of time, money, and effort was saved through the

use of Pro/E to analyze the design, essentially allowing the team to create a prototype

before the prototype was built.

Figure 7.1 – Components such as the fixed end were modeled in Pro/Engineer

54

Figure 7.2 – Component fits were tested with a model of the SEM position fixture

7.1.2 Finite Element Analysis with Pro/Mechanica and I-Deas

In order to verify the safety of the design and accuracy of stress calculations,

several of the fabricated parts were analyzed with finite elements. Though most of the

components were purchased and were rated to specific loads, the fabricated components

needed to be evaluated before proceeding with the design. The high stress components

were tested, which included the gears, power screws, and the free end. The gears were

analyzed in Pro/Mechanica, but the power screws and free end were analyzed with I-

Deas. Though Mechanica worked most consistently with the existing models, some team

members had more experience with I-Deas. Finite element analysis is never the sole

method of design, as garbage in is equal to garbage out. Models were carefully designed

and constrained, as boundary conditions have a drastic impact on stress magnitudes and

locations. Calculations were done first to estimate the stresses and were a used for

verification of the finite element methods.

Gears, which were case hardened, were not expected to fail in the current design.

The pinion gear and 2.5 inch gears were modeled as 3D solid elements in Pro/Mechanica

structural analysis. The bore surface was constrained in all degrees of freedom and the

gears were loaded with a tangential uniform force of 440 lb to the surface of one tooth,

55

the equivalent contact force from the maximum gearbox torque. The number of elements

was rather large due to the complicated geometry from the several gear teeth. For each

model, displacement, Von Mises stress, and shear stress were checked.

As expected, the pinion gear was the weaker gear, due to its smaller size and

smaller hole to diameter ratio. Its maximum Von Mises stress was 97.3 ksi, which is a

factor of safety of 1.85 with respect to the yield strength. Shear stress was the greatest at

the root of the tooth and was 54.5 ksi. In comparison, the calculated value was 64.86 ksi.

Displacement was also relatively small, as the maximum displacement was .000973

inches at the tip of the tooth. Stress and displacement contours for the pinion are shown

in Figure 7.3. A shear stress contour plot is shown in figure 7.4

Figure 7.3 - Von Mises stress and displacement in the pinion.

Figure 7.4 - Shear stress in pinion and gear.

56

As expected, the spur gear did not approach its failure criteria. Due to a larger

size and a smaller ratio of bore diameter to overall diameter, stresses were smaller. The

maximum Von Mises stress was 49.7 ksi, compared to a yield strength of 180 ksi. The

factor of safety is 3.6 for the spur gears. The maximum shear stress was 27.5 ksi,

compared to a calculated value of 25.95 ksi. Displacement was less than half of the

maximum displacement in the pinion, as it was only 0.000437 inches at the tip of the gear

tooth. As with the pinion, the numbers were in the same general range and the two

components were deemed safe. Stress and displacement contour plots are shown in

Figure 7.5 and a shear stress contour plot is shown in Figure 7.4.

Figure 7.5 - Von Mises stress and displacement in the gear

When modeling the power screw in I-deas, the shaft was created to almost the

same geometry except that the threads were only modeled for a total length of 0.750

inches. The length of the shaft in which the ACME threads are be engaged was all that

was necessary to model. The surface of the threads facing away from the other end of the

shaft was constrained in every direction to simulate the conditions of the shaft before

enough torque is applied to overcome the self-locking properties of the ACME threads

when the shaft is applying 1000 lbs of force to the free end. The shaft is also constrained

in the x and z directions along the surface of the shaft that is inserted inside the needle

roller bearing, since those are the directions in which the bearing restrains the shaft. The

rest of the boundary conditions were all applied as distributed loads along certain

surfaces of the shaft.

57

The main axial load of 1000 lbs was applied as a distributed load along the flat

surface of the shaft created by the first diameter step. The resulting pressure applied to

that surface was 3036 psi. Since the torsion from the gear is to be transmitted through the

key, the resultant distributed load was applied to the keyway. A pressure of 21.308 ksi

was applied at one of the sides of the keyway in order to approximate the torque being

applied to the shaft. A pressure of 5334 psi was also added to the bottom of the keyway,

in order to simulate the force being applied to the end of the shaft by the downward force

component being applied to the gear teeth.

For meshing, tetrahedral solid elements were used. An element size of 0.125

inches was set and the mesh was created, resulting in somewhere around 60,000 separate

nodes. When the FEA analysis solution was run and the data post-processed the

following two Von Mises stress distributions, below, developed.

Figure 7.6 - Gear end of lead screw

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Figure 7.7 - ACME end of lead screw

Figure 7.6 shows a max stress of 34.6-ksi around the edge of the keyway, which

can most likely be attributed to the local stress concentrations of the pressures that are

applied there. Because of these local concentrations, the stress calculated in that region

may not be 100% accurate. Most likely the stress in that region is around 27.2-ksi, which

is marked in orange. One area of interest in this figure is the fillet edge on the diameter

step. It is here that a stress concentration is expected to occur, and the Von Mises stress

curve supports that theory. The area of interest is surrounded by a shade of green but the

round is noticeably yellow, meaning that most likely the stress there is about 27.7-ksi.

The Von Mises stress plot in Figure 7.7 supports preliminary analytical

calculations. The ACME threads were not a point of failure for the power shaft. The

stress plot shows a max stress of around 5.5 ksi, which is much lower than the max stress

to be achieved at the other end of the shaft.

The material used for the power screw was 304 Stainless Steel. This grade of

stainless steel has a tensile strength of 73.2 ksi and yield strength of 31.2 ksi.

Considering that the above boundary conditions that were set were at an absolute worst-

59

case scenario it appears that the loads in the power screw will not approach the yield

strength. Also, since the load frame will see at most 100,000 cycles, fatigue failure is not

an issue. Theoretically this shaft should also be able to withstand 106 cycles, since the

fatigue life for stainless steel is nearly infinite for loads below the yield strength [9]. It

appears that the lead screw should withstand the operating conditions that it was designed

for.

The finite element analysis of the free end was done using the I-DEAS software

package. Symmetry could have been used to cut computation time but the part was

simple enough that modeling in quarter-symmetry was not necessary. The initial design

for this part required a one thousandths press fit of iron nuts into the holes for power

screws. This press fit would create a pressure of 2000 psi in those holes. The 2000 lb

load was applied over the counterbore for the grip and the bases of the counterbores were

constrained not to move in the Z (tensile) direction. After meshing and running the

solution a maximum stress of 13.5 ksi was obtained, which translates to a factor of safety

of 2.2. The locations of max stress were along the outside, where the wall thickness from

the press fit is thinnest and along the setscrew holes in the direction of applied force.

For the compressive case, 200 lb was applied to the circular area in contact with

the grip collar. The maximum stress was found to be 15.1 ksi producing a factor of

safety of 2. The stress concentration in this case is greater than for the other case, despite

of the drop in load by a factor of ten because there is much less area supporting the load.

While machining the ACME nuts and free end, the press fit was removed from

the design. The same boundary conditions and loads were applied as before, with the

exception of the press fit. The maximum stress was found to be 8.34 ksi producing a

factor of safety of 3.6. The stress concentration is on the edge where the iron nut would

contact the free end. As expected, the stress is higher along the edge of the hole with the

most bending stress (Figure 7.8).

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Figure 7.8 - Von Mises stress in the free end

Finally, an analysis was run on the “Compressive” stress state without the press

fit. After solving, it was found that the stress was largely negligible for the entire piece.

Analysis showed that the part in its final form has a factor of safety of 3.6 and that the

removal of the press fit from the design was a good engineering decision from the

perspective of stress in the free end.

7.2 Testing the Product for Reliability and Quality 7.2.1 Integrated Test Plan for SEM Load Frame

The SEM tensile load frame involved the development of six modules. Each

needed to be tested for technical design specification and customer requirements as well

as safety requirements associated with incorporating the system in the SEM platform.

The scope of testing was limited by the project duration, budget allocation and technical

competency of the design team. It should be noted that this project involved the design

and commissioning of the tensile testing system and will differ from commercial projects

by virtue of the constraints governing the size and user requirement, making it a unique

case never attempted anywhere.

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The focus of the test plan will take into account the developmental process, from

the needs assessment to production and deployment. It should be noted, however, that

there are two primary approaches to this kind of testing; one is to test during development

and second during operation. It is possible to run the two procedures concurrently for

well-funded complex projects or programs involving sophisticated subsystems and

involving high-risk catastrophic failures. Since the scope of this project terminates with

deployment, it is reasonable to integrate a test plan to cover the development of the

system, which will have a test planning team consisting of the customer and the owner of

the SEM machine (NCR³).

Figure 7.9 - Planning flow for Integrated Test [ref]

Test Planning Team Individual Test

Planners (CIMS)

Project Office

Test Performers

Test

Policies

and

Ground

Rules

Summary

Specifications

Test procedures Test reports

Test Planning

Forms

Integrated

Test Plan

Management

Summary and

Status Vol.1

Special Studies

Test integration

Project direction

Contract negotiation

7.2.2 Individual Test Plans

The Development Test and Evaluation for this project will be categorized into

three parts. The first part covers individual components or parts. The second covers

subassemblies, and the third covers deployment testing. The first two categories were

incorporated into the concept feasibility assessment phase. The team members on their

respective assigned modules of the design will conduct testing. Most of this part of

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testing will be completed after the preliminary design review and will verify that

components or subsystems satisfy the customer needs and technical specifications.

7.2.3 Component Processing

Many parts were outsourced from vendors and upon delivery they were routed to

assembly or manufacture, depending on whether they were stock material or finished

components. For stock material, the first check was to verify the material specifications,

as stated in the order form. A manufacturing sheet has been developed for every

fabricated part to state all manufacturing requirements, including the module, part

number, sequence of machine operations and related checks as well as engineering

drawings. Otherwise, if parts were found to be defective at this stage, they were sent

back to the supplier for replacement.

During manufacture, some of the parts possessed features that required

specialized operations and multiple set-ups. For each part routed for manufacturing, the

detailed processing requirements depended on the type of operation required and the

availability of facilities and machines. Most of the parts required normal machine

operations like milling, turning, boring, etc, which generated scrap material. In light of

this undesirable situation, the stock was ordered to minimize scrap and reduce processing

time. Tolerances and dimensions were recorded and compared with the requirements and

further action was taken depending on whether the specifications were satisfied or not.

Finally, the finished part advanced to assembly. Those that did not meet specifications

were routed for reworking or were scrapped.

7.2.4 Subassembly Testing

Once parts met and satisfied the requirements, they were routed for subassembly

integration. During this phase, individual components were joined to others with special

techniques that have been outlined in the feasibility assessment. The operations varied

depending on each subassembly requirement. It was decided to verify interactions and

tolerances at this stage to eliminate any undesirable conditions that may arise before

integrating into respective modules. Each subassembly will have its own set of

instructions, as per the specifications depicted in the feasibility assessment phase. For

components that did not require manufacturing processing, but need to be integrated into

63

subassemblies, re-working of parts that have been routed through manufacturing

compensated any discrepancies encountered.

7.2.5 Deployment Testing

When all components and subassemblies satisfactorily met their requirements,

subassemblies were incorporated into their respective modules. During this phase, all

stakeholders were invited for a complete system test. This test was conducted with the

system outside the operating chamber. Recorded data included maximum force and

displacement to verify customer specifications. Safety features were tested at this phase

and the software was calibrated. The module was later verified in the chamber once the

functionality was tested. Documentation made available to the customer included

assembly instructions and an operation manual.

Figure 7.10 - Material flow for SEM load frame manufacturing.

Receiving Manufacturing Assembly Deployment Test

• Lathe • Mill • Drill • Work

Table

• Torque • Tolerances • Vacuum • Function

• Supplier Info.

• Part number

• Specification

• Routing

Parts

Stock

• Control o Power o Safety

Stop • Display

o Force o Distance o Motor

Temp

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8 Final Design

8.1 Changes from Preliminary Design 8.1.1 Gripping

The setup for gripping evolved into a simpler, more ergonomic design.

Previously, three cylindrical grips were employed with one grip on either end of the load

cell on the free end and one on the fixed end. However, with the load cell on the free

end, there was no feasible way to grip and ungrip the samples. The problem was solved

rather easily, as the male-ended load cell was replaced with a male/female-ended load

cell. The size was perfect to fit in place of the cylindrical grip on the fixed end and

eliminated an additional grip that had been on the other side of the load cell. The new

setup is very easy to use and the samples can be installed in three basic steps. The

sample is threaded into the load cell, the grip is threaded onto the sample, and the end

bolt is passed through a protective collar and threaded into the grip. The collar was once

a two-piece cylindrical design that was removable to leave room for unscrewing the grip

without removing the bolt. The intention of the collar was to hold the grip in place,

provide a bearing surface during compression tests, and to provide a method of securing

the end bolt if the sample fractured. With the current design, there was no need to use

65

the two-piece design and it was replaced with a cylindrical component that resembled the

grip.

Figure 8.1 - Final design of gripping module

Grip

End bolt

Collar

Load cell

Figure 8.2 - Preliminary design of gripping module

Grips

Collar

8.1.2 Shaft and Bearing Setup

Though the preliminary design of the shafts was feasible, the team decided to

reduce the necessary torque with the redesign of the lead screws. Addition of bearings,

washers, and vacuum-safe lubrication reduced the necessary torque considerably. The

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design employs larger power screws (1 inch instead of ¾ inches) to allow for stepping

down the shaft twice. The first section is held by needle roller bearings inside the fixed

end and supported on either end by needle roller thrust bearings (Figure 8.3). Everything

on the shaft is held together with a 5/8-18 nut. A keyway on the end of the shaft allows

power transmission from the gear and a #10-32 cap screw secures the gear to avoid any

axial motion.

Figure 8.3 - Assembly of bearings, washers, and gears to power screws

8.1.3 Free End and ACME Nuts

In order to provide a threaded surface for the screws to interface, several designs

were considered. Originally, a nut and flange design (see Figure 8.1) was considered, in

addition to ACME tapped holes. Due to the price of taps, which far exceeded $150 each,

the ACME nut concept was chosen. Using machinable gray iron cylindrical ACME nuts,

the components were easily adjusted to the 1.250-inch diameter and 0.750 inch length.

Holes in the free end provided a light interference fit and two ¼-20 setscrews were also

employed to secure the ACME nuts.

8.1.4 Custom Gear Design

In order to keep the gears strong and avoid any deformation or destruction, they

were custom designed. With the help of a contact at the Gleason Works, the gears were

designed and manufactured free of cost. Materials available in purchased gears were

mostly plastic, brass, and steel. In order to increase the yield strength, case hardened

8620 steel was chosen for the material. The gears have a pitch of 16 and a contact angle

67

of 20°. Manufactured with pregrind hobs, the gears reached a tolerance class of AGMA

11 before heat treatment an AGMA 9 after heat treatment. They will withstand a

maximum stress of 180 ksi and maximum torque beyond the limit of the gearbox. They

mate with the small diameter on the end of the power screws and interface with a key to

transfer the torque.

8.2 Cost Analysis The team was limited to a budget of $7,500 for the design and fabrication of the

SEM tensile load frame. Standard components such as fasteners and washers, which

usually can be obtained for free in the shop, were purchased because the black oxide

finish would not be safe in vacuum conditions. Several washers and fasteners were

composed of stainless steel, plastic or titanium. Manufactured components were

fabricated from 303 stainless steel rod and bar stock and 304 stainless steel power screws.

Most of the budget was spent on the vacuum rated motor and gearbox from Empire

Magnetics. Space constraints and environmental conditions limited the options, resulting

in a higher price tag than initially expected ($5500). The gears were obtained for free

from the Gleason Works and several of the electrical components were also obtained for

free from the Computer Engineering Department. As a result, the total cost of $6614.08

stayed well below the allotted budget of $7,500. An item and supplier cost list is given in

Table 8.1.

Table 8.1 - Cost List Part No. Component Qty Unit Cost Total Cost

10101 Stepper Motor 1 $991.00 $991.00

10102 Gearbox 1 $2,789.00 $2,789.00

10201 ACME Screw (RH) 1 $132.00 $132.00

10202 ACME Screw (LH) 1 $174.89 $174.89

10203 ACME Nut (RH) 1 $35.67 $35.67

10204 Spur Gear 2 $0.00 $0.00

10205 Pinion 1 $0.00 $0.00

10206 Needle Bearings 4 $2.13 $8.52

10207 Needle Bearings 2 $4.64 $9.28

10208 ACME Nut (LH) 1 $51.70 $51.70

10302 Cap Screws 1 $5.82 $5.82

68

10307 Load Cell 1 $900.00 $900.00

10501 Control Box 1 $0.00 $0.00

10503 Stepper Drive 1 $0.00 $0.00

10504 Power Supply 1 $15.00 $15.00

10505 Serial Cable 1 $16.00 $16.00

10601 Electrical Interface 1 $413.25 $413.25

10602 Air Side Connector 1 $47.50 $47.50

10603 Ceramic Spacers 2 $10.45 $20.90

10604 Wire Crimpers 4 $57.00 $228.00

10701 Cap Screws 1 $5.58 $5.58

10701 Cap Screws 1 $4.27 $4.27

10702 PTFE Washer 1 $5.87 $5.87

10703 Bearing Washers 10 $0.75 $7.50

10704-5 Washers 1 $6.67 $6.67

10706 Nuts 1 $7.90 $7.90

10707 Washers 1 $5.10 $5.10

10708 Washers 1 $9.48 $9.48

10709 Washers 1 $5.01 $5.01

10710 Key Stock 1 $1.89 $1.89

11301, 3 Stainless Steel 1 $16.20 $16.20

11306 Specimens 1 $12.31 $12.31

11401, 2 Stainless Steel 1 $82.32 $82.32

11403 Stainless Steel 1 $36.98 $36.98

N/A Motor Adaptator 1 $325.00 $325.00

N/A Vacuum Lubricant 1 $146.02 $146.02

N/A PTFE Tape 1 $2.02 $2.02

N/A PTFE Tape 1 $4.06 $4.06

Subtotal $6,540.76

Shipping $73.32

Total $6,614.08

8.3 Design for Manufacture 8.3.1 Introduction

Product Design for Manufacture and Assembly (DFMA) presents an overview of

the factors influencing product design and the manufacturing cycle which include

69

component design and analysis, design for manufacturability, function, design of manual

assembly, and manufacturing processes.

DFMA software is a combination of two complimentary tools, Design for

Assembly (DFA) and Design for Manufacture (DFM). DFA software allows for

systematic analysis of a product design with the goal of reducing manufacture and

assembly costs, improving quality and speeding time to market. DFMA includes DFA

and the part cost estimation software DFM. Consolidating parts into elegant and

multifunctional designs, DFA reduces the complexity of a product. DFM software then

allows the engineer to quickly judge the cost of producing the new design and to compare

it with the cost of producing the original design.

The early and accurate cost estimate provided by DFMA software allowed the

team to proactively budget resources while the product was still under development.

During the development stages of a new product, cost and cost drivers deserve careful

consideration. The goal was to improve the product without increasing the costs.

However, this project tackles a new case. Usually, designers have no way of accurately

quantifying whether the specific innovation they are contemplating will increase or

reduce overall product cost. The Design for Manufacture and Assembly suite of software

gave the team tools to use during the product development cycle to analyze and

understand the cost effects of design decisions.

8.3.2 Design Phase

Control of part count is paramount to maintaining cost targets and overall

functionality. DFA software tools helped simplify the product by focusing the attention

of design team on part count and part count reduction. Product simplification was

achieved through the application of the software’s industry-tested minimum part count

criteria. The analysis allowed the determination of the theoretical minimum number of

parts that must be in the design for the product to function as required. Identifying and

eliminating unnecessary parts also eliminated unnecessary manufacturing and assembly

costs. Suppliers were a rich source of feedback during product simplification,

particularly if one of the options is to consolidate multiple parts into one part with

multiple features.

70

The potential for cost reduction continued when the team selected the optimal

materials and manufacturing processes for each part in the design. Using DFM software

tools, a thorough understanding of the primary cost drivers associated with manufacturing

the product was established (a benchmark for what the product "should cost."). Central

to the “should-costing” approach is accumulating real information about manufacturing

costs and noting where specific costs arise in your design. Large costs in product

development are associated with design manufacturability, so sharing “should-costing”

information with suppliers can make your collaboration more fruitful.

8.3.3 Design Optimization

The models in DFM Concurrent Costing software guide the user through an

assessment of alternative processes and materials and provide cost information for the bill

of material. Costs update automatically as you determine tolerances, surface finishes,

and other part details. Gradually, as effective shape-forming processes and considering

how to modify part features to lower cost, the product becomes more optimized.

8.3.4 Results

Achieving higher DFA indices, a measure of assembly efficiency, depends on the

following: part integration, ease of assembly, part inserted from top, part is self aligning,

part does not need to be aligned, one hand assembly, no tool required, assembly in single

linear motion and part secured by insertion.

The total number of parts was 61. The theoretical minimum number of parts

calculated by the software is 15 with a design efficiency index of 6.32%. There are 27

different parts. The total product cost is $7246.48 with total manufacturing tool cost of

$1800.00. These figures are based on a 10-year production plan for 100 products. The

manufacturing cost for each part fabricated was estimated to be $100.00.

A base part was chosen for other parts or subassemblies to join. The structure is

divided into subassemblies, which are pre assembled. There are three subassemblies: the

motor, free end and the fixed end assemblies. The fixed end has two subassemblies,

namely the right and left shaft assemblies.

71

8.3.5 Suggestions for Redesign

Discussed below is the list of suggestions from the DFMA software and the

team’s replies. Many of the changes suggested by DFMA were not applicable to the

design for reasons such as machining cost and loss of function and flexibility.

Incorporating integral fastening elements into functional parts or changing the

securing methods will eliminate the need for separate fastening elements such as screws,

nuts and keys. It was necessary for our design to have the above to support disassembly,

repair and replacement parts or service.

Combining connected parts or attempting to rearrange the structure of the product

could eliminate the parts whose function is solely to make connections. The Key was

identified as one of these parts but could not be substituted; it was incorporated to

maximize torque transmission.

Reduce the number of parts in the assembly by combining with others or

eliminating parts. The parts are mainly washers and bearings, which aid in alignment and

could not be replaced. However, combining a part with another may eliminate parts such

as fasteners.

Consider redesign of the individual assembly parts that require a grasping tool

such as the key. Changing the key would require another method of connection and

hence will be not changed.

Consider redesign of the individual assembly parts that require severe insertion

difficulties such as the bearings. This assembly set up was required for this fit and will

not be changed at this time.

Only three parts were identified to pose ergonomics difficulties for the assembly

worker. The bearings require significant pressure to insert and the motor pinion gear and

drive gears require a cover as the have sharp edges.

8.4 Included Parts Drawings for all of the manufactured parts and the major purchased components

are included in Appendix 1. The parts modeled in the assemblies are listed below. All

but the vacuum components and fasteners have a drawing. Some part numbers were used

on parts that are no longer in the design. Hence, some numbers were skipped.

72

20001 – SEM load frame (2 sheets)

20002 – SEM chamber with load frame (2 sheets)

20003 – Electrical feedthrough assembly

10101 – Motor

10102 – Gearbox

10201 – Right handed ACME power screw

10202 – Left handed ACME power screw

10203 – Right handed ACME nut

10204 – 2.5” spur gear

10205 – 1” spur gear

10206 – Needle-roller thrust bearing

10207 – Needle-roller thrust bearing

10208 – Left handed ACME nut

10307 – Load cell

10601 – Vacuum rated 20 pin feedthrough

10602 – Air side plug

10603 – Ceramic spacer

10604 – Wire crimpers

10700 – Cap screw - 10-32 x .5

10701 – Cap screws – 10-32 x 2

10702 – Compressible PTFE washer

10703 – Washer

10704 – Spherical washer

10705 – Spherical washer

10706 – Nut

10707 – Clamp washer

10708 – Shim washer

10709 – Shim washer

10710 – Key

10711 – Set screw – ¼-20 x 3/8

10712 – Cap screw – 3/8-24 x 2.5

73

11301 – Cylindrical grip

11303 – Compression collar

11306 – Specimen

11401 – Fixed end

11402 – Free end

11403 – Base

74

9 Production Planning 9.1 Material Considerations

The choice of material for any component within a design is crucial. Many

different aspects have to be considered in order to choose what best works for the

problem at hand. The material has to have sufficient properties in order to effectively

perform in its operating environment, i.e. tensile strength. Also, the material must be

able to be formed or machined into its specified component with certain degrees of

accuracy. Lastly, cost must be taken into consideration. Budgets cannot be exceeded and

if the material costs too much then it will be of no use. For the SEM Tensile Load Frame

all of these aspects were taken into consideration when choosing what materials to use

when making all the different components within the design.

One of the metals most widely used within the load frame was austenitic stainless

steel, specifically AISI 303 stainless steel. Stainless steel was used to manufacture the

free and fixed ends of the load frame, the grip, and the grip collar. The power screw

shafts (both left and right handed) and many of the fasteners were composed of AISI 304

(also known as 18-8) stainless. For all of these components it was found that stainless

steel had sufficient strength and stiffness in order to sustain the stresses applied to the

load frame. It also resists corrosion, which is very important so as to reduce the

possibility of contaminating the electron microscope chamber. In addition, it does not

outgas, to decompose or evaporate, at the vacuum pressure in which it will be placed, and

the cost of the material was low enough to fit into the allocated project budget.

The base of the load frame was constructed out of ½ inch thick aluminum. It was

decided to use aluminum for the base instead of steel for a number of reasons. Firstly,

aluminum is much easier to machine than stainless steel, therefore cutting down

production time. Secondly, strength and stiffness for the base was not nearly as much of

an issue as with other components because its only function is to fasten the load frame

onto the base of the microscope chamber. Thirdly, the aluminum was readily available

and added no cost to the project.

The gears were made out of 8620 case hardened steel. This steel was chosen

because of its high yield strength (180 ksi). Because of the torques acting on the gears,

the potential shock loads would be too much for the gears to handle if stainless steel were

75

to be used. Therefore gear teeth stripping would have occurred. The case hardened steel

can easily handle the applied stresses as well as well as not outgas in the vacuum

environment. It is not as resistant to corrosion as stainless steel, however, so care must

be taken so as not to allow any rust to form on the gears. Because of the difficulty to

machine 8620 steel and the overall difficulty in making gears, The Gleason Works

manufactured the gears.

For the 1-10 ACME threaded nuts in the free end, 3-inch diameter nuts were

purchased from McMaster-Carr for custom machining. The two metals in which

McMaster Carr offered these parts were in bronze and iron. Tin has a low vapor

pressure, and brass was potentially a problem with outgassing. Though brass was assured

by the Microelectronic Engineering Department to be safe, the iron inserts were chosen.

Some slight rusting could be a problem so care will be taken with handling and cleaning.

Figure 9.1 - Removing a large amount of material from the fixed end

9.2 Tooling Design and Machine Setup All of the work done on the free end, fixed end and base of the load frame was

done on a milling machine. All processes done on the milling machine were accurate to a

thousandth of an inch, with the aid of digital coordinate readouts for the x, y, and z-axes.

Cutting speeds used for machining the fixed and free ends were much lower than what

was used for the base. This was because the base was made out of aluminum and

material could be removed much faster than it could be for the stainless steel fixed end

and free end. The flats for the ACME nuts were also machined on the mill.

The power screws, grip, grip collar, and ACME thread inserts were all machined

on a lathe. The speeds they were turned at were between 150-200 rpm. They could not

be turned at much higher speeds without causing damage to either the parts and/or the

76

carbide tool bits. The accuracy achieved on the lathe was, like the milling machine, to

within one thousandth of an inch, but was instead achieved though the dials used to adjust

the location of the cutting tool. Due to the decreased cutting speed, the feed rates were

fairly slow. Also, the amount of material removed in one pass was kept small, about

0.030 inches for the threaded nuts and other cylindrical components, because the metals

were difficult to machine.

In the case of the power screws, the external threads and 5/8-shaft section were

sanded to fit cleaner in the fixed end of the load frame. Emery paper was used instead of

a cutting tool because the change in diameter needed to only be about one thousandth of

an inch, which was better suited to be removed by emery cloth. Great care also had to be

taken when cutting threads within the power screw shafts, because of the strength of the

stainless steel. It is easy enough to break a tap in aluminum but in steel the threat

becomes even greater.

9.3 Manufacturing Process Sheets Process planning involves the transformation of part or product definition data

into detailed production instructions. Production data may include materials, geometry,

and topology, tolerances, and demand volume. According to Alting and Zhang [ref]

(1989) this process can be summarized in ten steps. The steps include: interpretation of

product design data, selection of machining process, selection of machine tools,

determination of fixtures and datum surfaces, sequencing of operations, selection of

inspection devices, determination of production tolerances, determination of the proper

cutting conditions, calculation of standard times and costs, and generation of process

sheets and numerical control code.

The plan depicts detailed manufacturing, assembly, test, and service instruction in

a normal industry arrangement. However, for this particular project the plan was slightly

modified to focus on operations, machine selection, engineering drawings, supplier name,

part number and total manufacturing time and part material description, machine tools

and verification. This information was then consolidated into a manufacturing route

sheet for each part that required indoor fabrication. For ready to assemble parts, the sheet

will have drawings, and other relevant information without the machine code, machinist

assignment and processing times.

77

As part of the team quality standards, the manufacturing route sheets provide the

basis for continuation of manufacturing information and the relevant manufacturing

sequence used to fabricate the parts. This will also serve as record for improvement and

as a reference for all fabricated parts. This included the tooling changes that resulted

from changing operations during the manufacture of parts. It was observed that some

parts required special attention during manufacture due to the limitation of the tool

availability and material constrains in manufacturing and tool selection. Therefore, every

sheet was updated accordingly to the operations used in the machine shop. The

manufacturing route sheet thus serves as a record of all activities involved in operations

as well as a quality control reference during fabrication.

Figure 9.2 - Manufacturing process sheet for the free end

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10 Pilot Production 10.1 Manufacturing Difficulties

Despite the relatively fast production, the project encountered some

manufacturing difficulties along the route to success. Though AISI 303 is considered a

free machining alloy, it is still stainless steel. No special cutting tools were required, but

cutting speeds were drastically reduced. As a result feed rates were also reduced.

Despite the slow cutting speeds, only ten parts were machined. Fabricated parts included

two ACME shafts, two ACME nuts, the base, fixed end, free end, grip, collar, and a

sample.

The fixed end, which required a large cutout for the motor mount, was the first

component to create machining difficulties. Using a standard end mill, only .030 -.060

passes were made and with an inch of material to remove, something needed to be done.

The cutter was glowing cherry red from the heat. A roughing cutter of 1.9 inches in

diameter was used to clear out the remainder of the material. With a slow steady feed,

the cutter machined the motor mount section in a total of 5 passes. Large holes for the

power screws also created problems. In order to machine the 5/8 diameter holes, a pilot

hole was first drilled with a 3/8 bit. Counterbores and bearing bores were machined with

the use of end mills, rather than use flat bottom drills, since counterbore tooling was not

available in the specified size.

The free end was the only other source of trouble, due to the massive amount of

material that was removed. Most of the lessons learned from the fixed end machining

were applied to the free end and it was machined without any problems. As with the

fixed end, the excess material was removed as the component was trimmed to the proper

length and width. Clearance holes for the ACME taps were first piloted with several

sizes of drills to keep material removal minimal and ensure a clean cut.

The ACME nuts, though fabricated from gray iron, caused some manufacturing

issues. The 1.375 inch nuts were machined from 3 inch cylindrical nuts. Though gray

iron was machinable, material removal was kept to a maximum of .030 inches in

diameter for each feed, due to excessive noise. The parts took roughly an hour and a half

per part to machine, once the setup issues were resolved.

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10.2 Assembly Design The only major problem facing the design team was the assembly of components

with several degrees of freedom. In order to ensure proper alignment and a consistent

location of several key components, the team incorporated several assembly guides. The

ACME nuts were ground flat on the top and bottom surfaces to be secured by the

setscrews. Once the nuts were secure within the counterbored surface in the free end, the

screws were aligned so that the free end was parallel to the fixed end. The free end/shaft

subassembly was fitted with the appropriate bearings and washers and was added to the

fixed end, which already had the bearings. The additional nuts and washers were added

before everything was secured with a jam nut. Screws were adjusted so that they were

aligned with a parallel block and the surface of the free end. Gears were aligned and the

shafts were marked with locations for keyways. After cutting the keyways and fitting the

keys, the load frame was assembled without any problems. To ensure that the position

and alignment of components is kept consistent, a center punch was used to mark the

upper end of the ACME nuts and the free end. Gear bores were matched to a specific

power screw and the gear for the right-handed power screw was marked on its inner

surface with a center punch. If components are replaced, such as the ACME nuts or the

power screws, the machine should be assembled and realigned before the flats on the nut

or keyways on the shaft are machined.

10.3 Manufacturing Pictures The following pictures show the load frame in various stages of manufacture.

Figure 10.1 - The load frame was carefully assembled to check fits before cleaning.

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Figure 10.2 - Machining the iron nut

Figure 10.3 - Drilling holes for ACME nuts in free end

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Figure 10.4 - Gears were tested with a makeshift motor shaft.

Figure 10.5 - Test samples were machined to a 1-inch gage length

Figure 10.6 - Finished load frame awaits a motor.

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10.4 Recommendations for Improvement Though the design is successful in meeting its primary goals, some improvement

can be made on weight, position accuracy, and the control system. The weight of the

module was not given much consideration during the design, as other constraints such as

strength, stiffness, and corrosion resistance were dominant. High strength aluminum may

be a suitable replacement, but the team avoided the use of aluminum in load bearing

components because of the superior stiffness and strength of stainless steel. In a test

application, deflection should be kept to a minimum in the test apparatus and focused

primarily in the sample. With aluminum samples in a stainless steel load frame, the

machine should never yield or fail before the samples.

For position accuracy, a position encoder could be added to obtain a stiffness

estimate. Due to cost constraints, the team eliminated the encoder. One issue with

position accuracy is the compliance of the machine. A compliance curve would have to

be created by testing an extremely stiff material and recording the displacement of the

system. The displacement of the system will be subtracted from the travel in the free end

to obtain the actual extension in the sample. Even with a compliance curve, there is no

substitute for an extensometer. Position resolution is only accurate if an extensometer is

used, and neither can be used in destructive testing inside the vacuum chamber without

opening the chamber to remove the extensometer.

The control system can’t be optimized until the motor arrives, which will be in

early June. Internal paperwork delayed the purchase of the motor and gearbox due to

their high cost and the team was not able to optimize the control code before the project

deadline.

While the prototype is an embodiment of sound mechanical design, there are

some minor flaws to account for. Most notable among these flaws is a slight wobble of

the cantilevered free-end, only noticeable without a sample in the grips, as it is advanced

or retracted. The undesirable movement is very slight and not likely of sufficient

magnitude to affect the stresses in the sample or frame significantly. However, it is

unintended and, therefore, its’ existence must be explained. The sources of wobble and

slight oscillatory misalignment are beyond the control of the team for reasons deriving

from budget and project scope. Listed in the order of probable greatest effect, they

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include: the grade of thread tolerance used, the interface between the ACME nuts and the

free-end, and the limitation of available machining accuracy. All three are functions of

cost, although the second is, also, to a large degree, a problem of adequate space. Cost

was a problem until the gears were obtained for free, but the team was not aware of the

gear price until after the threads were ordered.

A 6C ACME thread class, and the associated lead accuracy and straightness, were

it available to our budget, would eliminate any wobble perceptible to the naked eye. Of

course it would introduce the problem of hole placements accurate beyond the machining

capability of the RIT machine shop. Only 2G ACME, general purpose, thread and

corresponding nuts, were within the scope of the budget, even before considering other

implications of an increased tolerance.

Standard nut and flange combinations are not possible in this instance due to inner

and outer flange diameter constraints eliminating all possible choices from the field. To

tap the ACME threads would require the purchase of both left-handed and right-handed

taps. The choice of a less conventional thread pitch only serves to further increase this

cost. Only the best attempt at a precision fit of a gray iron nut, turned to a proper outer

diameter, is feasible.

A tensile load frame is a precision piece of equipment. As the scale decreases, so

does the increment of precision. At best, machining accuracy is limited to 0.001in. Also,

ordered parts have their own tolerance ranges and errors which must be provided for.

The quality of the finished prototype is unmistakably good, especially when its’ status as

a first generation prototype is considered. Again, the deviations in the product do not

seem to have a noticeable impact on any testing conducted.

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References [1] Vinci, R.P., 2000, "Mini-Tester Capabilities,"

http://www.lehigh.edu/%7Erpv2/minitensile_tester.pdf

[2] In Situ Tensile Testing Machine and Sample for a Scanning Electron Microscope,

Chiron et al., 1997. US Patent number 5,606,168.

[3] Nagy, Campbell, and Davidson, 1983, “High-temperature, cyclic loading stage for the

scanning electron microscope,” Rev. Sci. Instruments, Vol 55 no. 5, May 1984, pp. 778-

782.

[4] ASTM E8-01, Annual Book of ASTM Standards, Vol. 03.01

[5] Fundamentals of Machine Component Design, Juvinall, Marshek, 3rd edition, 2000.

[6] Integrated Product Testing & Evaluation. H. Gilmore, H.Schwartz. New York 1986.

[7] Testing to Verify Design & Manufacturing Readiness, Mc Graw – Hill, Inc 1993.

[8] Defense Systems Management College, Test & Evaluation Management Guide, Fort

Belvoir, VA: DSMC, March 1988, p.2-8

[9] Stainless Steels – Chromium and Nickel, Technical Data, Allegheny Ludlum

Corporation, Pittsburgh, PA, 2003.

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Appendix Appendix 1 – Drawings for assemblies and parts.

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