THE GEORGE W. WOODRUFF SCHOOL OF Georg Tech MECHANICAL ENGINEERING

Georgia Institute of Technology

Atlanta, Georgia 30332-040q

Mechanical

ME 4182

Design Engineering

NASA/UNIVERSITY

ADVANCED DESIGN PROGRAM

LUNAR DEEP DRILL APPARATUS

June 1989

Jill Harvey, Editor

Professor: Mr. J. W. Brazell

AJIEqualEducationand EmploymentOpportunib' Institution A Fnitof the [rniversibSystemof (_rgia
1.0 EXECUTIVE SUMMARY ....................................................................................

2.0 PROBLEM STATEMENT ..................................................................................

2.1 PERFORMANCE OB JECTIVES .............................................................

2.2 CONSTRAINTS ........................................................................................

3.0 GENERAL ARRANGEMENT ..............................................................................

4.0 DEEP DRILL APPARATUS SYSTEMS DESIGN ........................................

4.1 BIT CUTTING LOADING INTO STEPPED AUGER ..................................

4.1.1 CUTTING TEETH ..................................................................................

0

2

2

3

5

9

9

9

4.1.2 SHOVEL APPARATUS ....................................................................... 2 0

4.1.3 INTERMEDIATE HOLDING DEVICE ............................................. 2 1

4.1.4 STRUCTURAL DESIGN ....................................................................... 2 2

4.1.5 MATERIAL SELECTION .................................................................... 2 4

4.1.6 BIT LIFE ............................................................................................... 2 5

4.1.7 STEPPED AUGER ................................................................................. 2 5

4.1.8 BIT AND ROD INTERFACE ............................................................... 2 6

4.2 RODS, AUGER, AND STEPPED AUGER ........................................................ 2 7

4.2.1 INTRODUCTION ................................................................................... 2 7

4.2.2 DESIGN CONSIDERATIONS .............................................................. 2 7

4.2.3 MATERIAL SELECTION .................................................................... 2 8

4.2.4 INTERFACE DESIGN ............................................................................ 2 9

4.2.5 ROD DESIGN .......................................................................................... 3 1

4.2.6 MOTOR INTERFACE DESIGN ........................................................... 3 1

4.2.7 VERTICALLY RECIPROCATING AUGER DESCRIPTION ......... 3 3

4.2.8 VERTICALLY RECIPROCATING AUGER DESIGN ..................... 3 3

4.2.9 VERTICALLY RECIPROCATING AUGER INTERFACE

DESIGN ............................................................................................................... 3 5

4.2.10 AUGER DESCRIPTION ..................................................................... 3 6

4.2.11 DESIGN OF VERTICALLY RECIPROCATING AUGER

EMPTYING DEVICE ....................................................................................... 3 8

4.2.12 COMPONENT FABRICATION ........................................................ 3 8

4.3 FOOTPLATE .......................................................................................................... 4 0

4.3.1 SYSTEM OPERATION ......................................................................... 4 0

4.3.2 BASE ........................................................................................................ 4 1

4.3.2.2 CASING ................................................................................... 4 1

4.3.2.3 FINS ........................................................................................ 4 1

4.3.2.4 ADAPTOR ............................................................................. 4 2

4.3.2.5 SIGNALLING ........................................................................ 4 2

4.3.3 SPRING ................................................ . .................................................. 4 2

4.3.3.1 GRIPPERS ............................................................................... 4 3

4.3.3.2 SPRING SUPPORT ............................................................... 4 3

4.3.3.3 SPRING STOPS ..................................................................... 4 3

4.3.4 LIFTERS ................................................................................................. 4 4

4.3.5 MOTOR .................................................................................................... 4 4
4.3.6 POWER STORAGE................................................................................ 45

4.3.7 MATERIALS ......................................................................................... 45

4.3.8 ANALYSIS ............................................................................................. 46

4.3.9 MASS ANALYSIS ............................................................................... 46

4.3.10 REGOLITH PROTECTION AND REMOVAL ............................... 46

4.3.11 FAILURE ANALYSIS ...................................................................... 47

4.4 POWER SUPPLY, ROTARY DRIVE, AND HEAT

MANAGEMENT ...................................................................................................... 48

4.4.1 POWER SUPPLY .................................................................................. 49

4.4.2. MOTOR ................................................................................................. 53

4.4.3 TEMPERATURE MANAGEMENT .................................................. 55

4.4.4 HEAT RADIATION ............................................................................ 64

4.4.4.1 ALTERNATE EQUATION IN SI UNITS .......................67

4.4.5. FRICTION CLUTCH ........................................................................... 67

4.5 STRUCTURE AND MECHANICAL INTERFACE DESIGN FOR

SKITTER DRILLING IMPLEMENT ........................................................................ 69

4.5.1 INTRODUCTION ................................................................................... 69

4.5.2 CONSTRAINTS ..................................................................................... 69

4.5.3 PERFORMANCE CRITERIA ............................................................... 70

4.5.4 GEOMETRY ............................................................................................ 71

4.5.5 CONSTRUCTION AND ASSEMBLY ................................................. 72

4.5.6 CONNECTIONS TO ATTACHMENTS .............................................. 74

4.5.7 FASTENER CONSIDERATIONS ........................................................ 75

4.5.8 DESCRIPTION OF LOADS ................................................................ 75

4.5.9 STRUCTURAL ANALYSIS ................................................................ 77

4.5.9.1 CASE ONE ............................ _.................................................. 78

4.5.9.2 CASE TWO ............................................................................. 79

4.5.10 MATERIAL SELECTION AND TREATMENTS .........................80

4.5.10.1 ALUMINUM ALLOYS ..................................................... 81

4.5.10.2 TITANIUM ALLOYS ........................................................82

4.5.10.3 ALUMINUM - LITHIUM ALLOYS .............................82

4.5.10.4 METAL MATRIX COMPOSITES ................................... 83

4.5.10.5 MICROMECHANICAL PROPERTIES............................85

4.5.10.6 MACROMECHANICAL PROPERTIES ..........................89

4.5.10.7 TREATMENT ...................................................................... 92

4.5.10.8 ADHESIVE ANALYSIS ....................................................94

4.5.11 MECHANICAL INTERFACE DESIGN AND

OPERATION ...................................................................................................... 96

4.6 ROD CHANGER AND STORAGE RACKS ....................................................... 99

4.6.1 GRIPPING MECHANISM .................................................................. 99

4.6.2 MATERIALS ......................................................................................... 108

4.6.3 RACK AND PASSIVE GRIPPER SYSTEM .....................................113

4.6.4 MECHANICAL ARM AND ROTATING BAR ...............................115

2
4.6.5 MOTORS AND SPEEDREDUCERS.................................................... 117

4.5.6 CONCLUSIONS ...................................................................................... 119

4.7 MACROCORE RETRIEVAL AND STORAGE SYSTEM ..............................121

4.7.1 EXPLANATION OF CORE REMOVAL PROCESS........................121

4.7.2 GENERAL OVERVIEW OF EQUIPMENT .......................................121

4.7.3 INTRODUCTION .................................................................................. 123

4.7.4 CORE SAMPLE .................................................................................... 123

4.7.5 CONSTRAINTS ................................. ................................................. 123

4.7.6 DESIGN CONSIDERATIONS ............................................................. 124

4.7.7 DESIGN .................................................................................................. 125

4.7.8 RETRIEVER SHELL ............................................................................ 126

4.7.9 TOP CONNECTION .............................................................................. 126

4.7.10 PIPE CUTTER .................................................................................... 126

4.7.11 SPRING ............................................................................................... 128

4.7.12 RESET GEAR...................................................................................... 128

4.7.13 PIN ....................................................................................................... 128

4.7.14 GRIPPERS .......................................................................................... 130

4.7.15 PROTECTIVE GEAR COVER .......................................................... 130

4.7.16 CORE STORAGE DEVICE ................................................................ 13 0

4.8 CONTROLS FOR ENTIRE DRILLING OPERATION ................................... 132

4.8.1 TASK DEFINITION ............................................................................ 132

4.8.2 DRILLING OPERATIONS ............................................................. 134

4.8.2.1 OPERATIONAL FLOWCHARTS .................................... 134

4.8.2.2 OPERATIONAL POSITIONS .......................................... 136

4.8.3 HUMAN FACTORS DESIGN .........................................................137

4.8.3.1 WORKSTATION DESIGN .............................................. 137

4.8.3.2 MENU SCREENS/ANIMATION ..................................138

4.8.3.3 CONSOLE DESIGN ........................................................... 138

4.8.4 MENU DRIVEN OPERATIONAL DISPLAYS ..........................140

4.8.4.1 EARTHBOUND CONTROL ................................................ 140

4.9 VERTICAL ACCELERATOR ............................................................................. 143

4.9.1 INTRODUCTION .................................................................................. 143

4.9.2 DETAILED OPERATION DESCRIPTION ........................................144

4.9.3 SPECIFICATIONS ................................................................................ 145

4.9.4 CONTROL OPERATION SEQUENCE................................................. 146

4.9.5 MATH MODEL FOR THE VERTICAL ACCELERATOR ..............146

4.9.6 CAM-FOLLOWER ANALYSIS ......................................................... 152

4.9.7 COMPONENT ANALYSIS .................................................................. 157

4.9.8 DESIGN DEVELOPMENT ................................................................... 159

4.10 MICO-CORING DRILL .................................................................................... 162

4.10.1 CONTROLS REQUIREMENTS ........................................................ 162

4.10.1.1 MICROPROCESSOR CONTROL SYSTEM.....................163

4.10.1.2 MICROPROCESSOR SELECTION ...................................165
4.10.1.3 PIEZOELECTRIC ACCELEROMETER ............................166

4.10.1.4 ROTATIONAL SPEED CONTROL .................................166

4.10.1.5 DRILL BIT TEMPERATURE CONTROL .....................167

4.10.1.6 LIMIT SWITCHES ...........................................................168

4.10.2 DRILL PROTECTING DEVICE ..................................................... 169

4.10.3 BRACING MECHANISM ...............................................................170

4.10.3.1 CAMS ................................................................................... 170

4.10.3.2 BEARINGS .......................................................................... 171

4.10.3.3 BRACING PADS................................................................ 172

4.10.3.4 BRACING PAD RETURN ASSEMBLY .........................172

4.10.3.5 SELECTION OF MATERIALS ....................................... 172

4.10.4 ROTATIONAL MOVEMENT OF DRILL BIT ............................173

4.10.4.1 HEAT TRANSFER ANALYSIS ......................................173

4.10.4.2 TORQUE REQUIREMENTS FOR ROTATION OF

DRILL BIT ............................................................................................ 174

4.10.4.3 ROTATIONAL MOTOR SELECTION ...........................174

4.10.5 DRILL TRANSLATION MECHANISM ...................................... 176

4.10.6 MICRO-CORING DRILL BIT DESIGN ........................................ 177

4.10.6.1 OVERALL DESCRIPTION .............................................. 177

4.10.6.2 DIAMOND SELECTION AND MOUNTING ................177

4.10.6.3 BIT CLAMPING MECHANISM ....................................178

4.10.7 LUBRICATION AND BEARINGS .................................................180

4.10.8 DUST JACKET ................................................................................... 181

4.10.9 FAILURE SCENARIOS .................................................................... 182

5.0 CONCLUSIONS .................................................................................................... 192

6.0 RECOMMENDATIONS ....................................................................................... 194

6.1 BIT CUTTING LOADING INTO STEPPED AUGER .......................194

6.2 RODS, AUGER, AND STEPPEDAUGER ............................................ 194

6.3 FOOTPLATE ............................................................................................. 195

6.4 POWER SOURCE, ROTARY DRIVE, AND HEAT

MANAGEMENT ............................................................................................... 195

6.5 STRUCTURE AND MECHANICAL INTERFACE WITH

WALKER ............................................................................................................ 195

6.6 ROD CHANGER AND STORAGE RACKS ........................................... 197

6.7 MACRO-CORE RETRIEVAL TOOLS AND STORAGE....................197

6.8 CONTROLS FOR OPERATION AND CONSOLE................................197

6.9 VERTICAL ACCELERATOR FOR STEPPED AUGER .....................198

6.10 MICRO-CORING DRILL ..................................................................... 198

6.11 OVERALL RECOMMENDATIONS ................................................... 198

7.0. REFERENCES....................................................................................................... 200

7.1. ACKNOWLEDGEMENTS ...................................................................... 200

7.2. BIBLIOGRAPHY ................................................................................ 201

4
7.3.

7.2.1. BIT CUTTING LOADING INTO STEPPED

AUGUR ................................................................................................... 2 0 1

7.2.2. RODS, AUGER, AND STEPPED AUGUR ....................... 2 0 1

7.2.3. LUNAR FOOTPLATE ........................................................ 2 0 2

7.2.4.POWER SOURCE, ROTARY DRIVE, AND

TEMPERATURE MANAGEMENT ................................................... 2 0 3

7.2.5. STRUCTURE AND MECHANICAL INTERFACE

WITH WALKER .................................................................................. 2 0 3

7.2.6. ROD CHANGERS AND STORAGE RACKS ..................... 2 0 5

7.2.7. MACRO-CORE RETRIEVAL TOOLS AND

STORAGE ............................................................................................... 2 0 7

7.2.8. CONTROLS FOR OPERATION AND CONSOLE ........... 2 0 9

7.2.9. VERTICAL ACCELERATOR FOR STEPPED

AUGUR ................................................................................................... 2 0 9

7.2.10. MICRO-CORING DRILL ................................................ 2 0 9

SOFTWARE .......................................................................................... 2 1 2

APPENDIX A ............................................................................ 2 1 3

APPENDIX B ............................................................................ 2 1 4

APPENDIX C ............................................................................. 2 1 5

APPENDIX D ............................................................................ 2 1 6

5
FOREWORD

This report represents the combined effort of 49 students from the

Georgia Institute of Technology's School of Mechanical Engineering Senior

Design class. Although this report describes a conceptual design, it serves

two purposes. First it serves to teach the design process. Second, it

provides NASA with a general conceptual design of a Lunar Drilling

Apparatus capable of many drilling and sampling tasks.

The students of ME 4182 were organized into nine groups of five and

one group of four. Each group was given a part of the Lunar Drilling

Apparatus to design as best they could in a period of eight weeks. Each

group met with Professor Brazell once a week for advisement and

accountability. Four weeks into the project, each group made an oral

presentation of their progress and built a quarter-scale model. This report

served as a source of communication between the groups. Students were

required to provide weekly written progress reports with graphics to

serve as communication between the groups.

Students were encouraged to use any resources available to them,

including, but not limited to vendors, the Georgia Tech library, handbooks,

imagination, and similar terrestrial designs. Databases, NASA documents,

and long distance phone service were made available as well.

The Lunar Drilling Apparatus (LDA) is a part of a larger design for a

previously designed three-legged lunar rover named "SKITTER". SKITTER

should be able operate from a lunar base and be commanded from the

earth. It will go from the base to given locations to drill holes for samples

and other scientific and engineering purposes.

The design itself is conceptual and should be checked for engineering

credibility. Experimental testing and detailed analysis should be

completed before this design could be considered functional. Students'

time was limited by the overall time frame of the project and other class

requirements.
CONTRIBUTORS

Agin, Arthur Harlan

Avery, Robert Eugene

Black, Angela Michelle

Blair, Dwight Nelson

Blare, Thomas Franz Alfred

Bourcier, Mark Allen

Bowden, Allen Kenneth

Bridges, Timothy Scott

Bruzzano, Marco Anthony

Collins, Robert Jackson

Cooper, Jeffrey Allan

Donahue, David Philip

Dyer, John Francis

Fortson, Reginald Damell

Gargour, Farid Gilbert

Halferty, John Peter

Harper, Ronald James

Headley, Frank Anthony, Jr.

Hurst, Mark Edward

Hutto, James Douglass

Jennings, Scott Patrick

Kennedy, John Douglas

Kinney, Jody Tilmon

K.lima, Gerald Gregory, Jr.

Kuster, James Robert

Marsh, William Hill

Martin, James Stephen

McBrayer, Michael Sean

McDaniels, Robert Dale

McGinn, Patrick Laun

Miller, Scott Edward

Moy, Jennifer Ann

Murphy, Christopher Browning

Powell, Barry Edward

Ridlehoover, Kathryn Elizabeth

Rosario-Cuevas, Edward

Rush, Phillip Joseph

Saxon, Gregory Bruce

Seckinger, Cheryl Ann

Sheiner, Jeffrey Stuart

Sherrard, Roger Scott

Sobel, Bradley Paul

Stelzenmuller, Gregory Brennan

Stewart, Chilton Taylor

Waldee, Timothy L

Wielbruda, Richard Frank

Williams, Margaret Crecelius

Willingham, James Thomas, Jr.

Wolfe, Charles Vincent

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4.2

4.8

4.2

4.5

4.10

4.3

4.5

4.5

4.10

4.4

4.6

4.7

4.6

4.10

4.1

4.6

4.8

4.7

4.6

4.7

4.3

4.4

4.8

4.4

4.1

4.10

4.2

4.7

4.9

4.9

4.8

4.7

4.1

4.10

4.6

4.2

4.1

4.9

4.3

4.3

4.9

4.4

4.9

4.3

4.5

4.5

4.4

4.1

4.8
LIST OF FIGURES

FIGURE 3.0.1 :

FIGURE 3.0.2:

FIGURE 4.2.1.1 :

FIGURE 4.2.4.1:

FIGURE 4.2.4.2:

FIGURE 4.2.4.3:

FIGURE 4.2.4.4:

FIGURE 4.2.4.5:

FIGURE 4.2.5.1 :

FIGURE 4.2.6.1 :

FIGURE 4.2.8.1 :

FIGURE 4.2.8.2:

FIGURE 4.2.8.3:

FIGURE 4.2.9.1 :

FIGURE 4.2.9.2:

Side View of Drilling Implement

Top View of Drilling Platform

Role of PTS and CRS in Lunar Drilling

Front View of Male Interface and Back View of

Female Interface

Side Views of Male and Female Interfaces

Top View of Male Interface and Bottom View of

Female Interface

Rod Interface Coupled

Coupling Collar

Drilling Rod

Magnetic Motor Interface

Stepped Auger Segment

Cutaway View of Stepped Auger (VRA)

Overview of a Single VRA Winding

Upper VRA Interface

Lower VRA Interface

FIGURE 4.2.10.1: Auger Segment

FIGURE 4.2.10.2: Auger Helix

FIGURE 4.2.10.3: Lower Auger Interface Cutaway

FIGURE 4.2.10.4: Upper Auger Interface Cutaway

FIGURE 4.3.1.1:

FIGURE 4.3.1.2:

Side View of Footplate with Spring Open

Elevation View of Spring Closed on Drill String

FIGURE 4.4.a:

FIGURE 4.4.b:

FIGURE 4.4.c:

FIGURE 4.4.d:

FIGURE 4.4.1.1:

FIGURE 4.4.1.2:

FIGURE 4.4.2.1:

Design Process for Power, Motor, and Temperature

Management

Overall Arrangement Top View

Overall Arrangement Bottom View

Overall Arrangement Plan View

Power System Arrangement

Tank Cross Section

Curvo-Synchronous Motor Load Characteristic
FIGURE 4.4.2.2:

FIGURE 4.4.2.3:

FIGURE 4.4.2.4:

FIGURE 4.4.2.5:

FIGURE 4.4.2.6:

FIGURE 4.4.2.7:

FIGURE 4.4.3.1:

FIGURE 4.4.3.2:

FIGURE 4.4.3.2:

FIGURE 4.4.3.3

FIGURE 4.4.3.4:

FIGURE 4.4.4.1:

FIGURE 4.4.5.1:

FIGURE 4.4.5.2:

FIGURE 4.4.5.3:

Curvo-Synchronous Motor Current Characteristic

CSM Schematic

Motor Mount and Flange

Bracket Detail

Thrust Bearing

Thrust Bearing Detail

Heat Pipe Concept

Heat Pipe Arrangement

Heat Pipe Detail

Heat Pipe Specifications

Typical Wrapped Wick Screen

Assumption of Constant Temperature

Friction Clutch

Shaft Interface

Friction Clutch Specifications

FIGURE 4.5.5.1:

FIGURE 4.5.6.1 :

FIGURE 4.5.6.2:

FIGURE 4.5.8.1:

Typical Sleeve Detail

General Layout

Bottom Panel Attachment Locations

Shear Induced by Flexural Bending Due to Rod

Weight

FIGURE 4.5.10.1: Fiber/Matrix Interface

FIGURE 4.5.10.2: Strength at Temperature for Candidate Composite

Materials

FIGURE 4.5.10.3: Modulus at Temperature for Candidate Composite

Materials

FIGURE 4.5.10.4: Composite Strength vs. Temperature

FIGURE 4.5.10.5: SCS/Aluminum Tension-Tension Fatigue

FIGURE 4.5.11.1: Mechanical Interface Latch

FIGURE 4.6.1.1:

FIGURE 4.6.3.1:

FIGURE 4.6.3.2:

Robot Arm Work Cell

Spring Clip for Rod Mounting

Rod Rack and Arm Side View

FIGURE 4.7.1:

FIGURE 4.7.2:

Exploded View of Cutting Mechanism

Horizontal View of Retriever and Cutter
FIGURE 4.7.3 Side View of Retriever and Cutter

FIGURE 4.7.10.1: Positioning of Cutters

FIGURE 4.8.2.1:

FIGURE 4.8.2.2:

FIGURE 4.8.2.3:

FIGURE 4.8.2.4:

FIGURE 4.8.2.5:

FIGURE 4.8.2.6:

FIGURE 4.8.2.7:

FIGURE 4.8.2.8:

FIGURE 4.8.2.9:

FIGURE 4.8.2.10:

FIGURE 4.8.2.11:

FIGURE 4.8.2.12:

FIGURE 4.8.3.1:

FIGURE 4.8.3.2:

FIGURE 4.8.3.3:

FIGURE 4.8.4.1:

FIGURE 4.8.4.2:

FIGURE 4.8.4.3:

FIGURE 4.8.4.4:

FIGURE 4.8.4.5:

FIGURE 4.8.4.6:

FIGURE 4.8.4.7:

FIGURE 4.8.4.8:

FIGURE 4.8.4.9:

FIGURE 4.8.4.10:

FIGURE 4.8.4.11 :

FIGURE 4.8.4.12:

FIGURE 4.8.4.13:

Basic Operations

Systems Check

Footplate Routine

Drill Segment Routine

Macro-coring Routine

Micro-coring Routine

Vertical Accelerator Routine

Add Rod Subroutine

Connect Rod Subroutine

Remove Rod Subroutine

Disconnect Rod Subroutine

SKITTER Drilling

NASA-STD-3000

NASA-STD-3000

Artists Conception

Log On

Operations Positions

Figure 9.2.4.2.2-1

Figure 9.2.4.2.2-2

of Operator and

System Access/Preoperation Check

Preoperation Checklist

Return Control to Main Console

Activate Drilling Operation

Set Footplate

Drill Segment

Drilling

Macro-coring

Micro-coring Decision

Micro-coring

Drilling Complete

Revert Control to Main Console

Console

Decision

FIGURE 4.9.1:

FIGURE 4.9.2:

FIGURE 4.9.3:

FIGURE 4.9.4:

Vertical Accelerator (General

Vertical Accelerator (Internal

Vertical Accelerator (Detailed

Detailed Specifications (Clutch

4

Arrangement)

View)

View)

and Shaft Details)
FIGURE 4.9.5:

FIGURE 4.9.6:

FIGURE 4.9.8:

FIGURE 4.9.9:

FIGURE 4.10.1 :

FIGURE 4.10.2a:

FIGURE 4.10.2b:

FIGURE 4.10.3:

FIGURE 4.10.4:

FIGURE 4.10.5:

FIGURE 4.10.6:

FIGURE 4.10.7:

FIGURE 4.10.8:

FIGURE 4.10.9:

Detailed Specifications (Housing Details)

Straight Line Approximation

Particle Directions

Cam Profile

Cam Displacement Diagram (Dwell=30 degrees)

Cam Displacement Diagram (Dwell=45 degrees)

Cam Displacement Diagram (Dwell=60 degrees)

Cam Profile

Micro-coring Drill Rod Chamber

Drilling Operation

Drill Bit Translation Mechanism

5
LIST OF TABLES

TABLE 4.1.1:

TABLE 4.9.1:

Power Requirement Chart

Design Development

6
LIST OF ABBREVIATIONS

a

A

AISI

A1

amp

BTU

C (deg)

cm

CRS

CSM

CW

D,d

deg

e

E

F

F (deg)

FH

FT

ft

g

g

G

h

Hp

I

in

J

kg

kW

L,1

coefficient of thermal expansion

area

American Institute of Steel and Iron

aluminum

ampere

British Thermal Unit

Celsius

counter clockwise

centimeter

Cutting Removal System

Curvo-Synchronous Motor

clockwise

diameter

degree(s)

ductility

Young's modulus

force

Fahrenheit

fence height

number of fences per turn

foot

gram

acceleration due to gravity

Giga (10 9)

height

horsepower

moment of inertia

inch

angular moment of inertia

kilogram

kilowatt

length

7
lb

LDA

LDDA

m

M

_t

max

MCR

MIRS

mm

MMC

MMI

n

N

NASA

NiCad

P

7t

PLC

PTS

q

9

R,r

RAM

RIM

wavelength

pound

Lunar Drill Apparatus

Lunar Deep Drill Apparatus

meter

Mega (106)

viscosity

maximum

Macro Core Retrieval

Motor Interface Rod Section

millimeter

Metal Matrix Composite

Magnetic Motor Interface

efficiency

Newton

National Aeronautics and Space Administration

Nickel Cadmium

pitch

pi (3.14159...)

Programmable Logic Controller

Power Transfer System

heat transfer

density

radius

Random Access Memory

Rod Interface Mechanism

RPM,rpm revolutions per minute

Cry

Crut

SAE

see

SKITTER

Sn

STD

T

yield strength

ultimate tensile strength

Society of Automotive Engineers

second

Spatial Kinematic Inertial Translatory Tripod Extremity

Robot

tin

standard

temperature

8
Ti

V

V

VRA

W

W

CO

wt

titanium

velocity

volts

Vertically Reciprocating Auger

width

Watt

radial frequency

weight

9
1.0 EXECUTIVE SUMMARY

The scope of this project is to design a self contained, mobile

drilling and coring system to operate on the Lunar surface and be

controlled remotely from earth. The system will use SKITTER as its

foundation and produce Lunar core samples two meters long and

fifty millimeters in diameter.

The drill bit used for this is composed of 30 per carat diamonds

in a sintered tungsten carbide matrix. The bit assembly is 2 m long.

Because high temperatures will accelerate bit wear and 70% of the

heat generated from drilling will be held in the cuttings, a stepped

auger will be used to bring the chips away from the bit and up out of

the hole. The auger will be accelerated vertically upward by a camfollower mechanism which will convert the rotational motion of the

rotary drive into a vertically reciprocating motion.

To drill up to 50 m depths, the bit assembly will be attached to

a drill string made from 2 m rods which will be carried in racks on

SKITTER. The rods will be retrieved by a mechanical arm system

which will place them into the drill string. This mechanical arm

system allows two degrees of freedom: rotational and rotary

movement. The drive will be detached during rod addition of

removal, so the footplate will be used to support the drill string at

these times. This device is stored in the drive via an adaptor and is

placed onto the Lunar surface at the drill site first. Its support

mechanism consists of an aluminum lithium spring locking device

similar to an automotive speed washer. The spring is deflected by a

battery powered motor to open it during drilling.

Rotary power for drilling will be supplied by a Curvo -

Synchronous motor. This motor has only one moving part, can

operate at temperatures up to 350 F , and is 92% efficient. The motor

will be powered by a hydrogen oxygen fuel cell which has an output

10
of 2 to 12 kilowatts. Heat generated by the motor will be transferred

to a thermal radiator oriented towards deep space through a system

of three heat pipes.

SKITTER is to support this system through a hexagonal shaped

structure which will contain the drill motor and the power supply.

This structure will be made from silicon carbide aluminum and it will

attach to SKITTER with three actuator latches.

A micro-coring drill will be used to remove a preliminary

sample 5 mm in diameter and 20 mm long from the side of the core.

This drill utilizes one motor for rotational movement and one for

vertical translation. Finally, after all drilling, the core will be

removed in 2 m sections by cutting the sample with the macro core

retriever. The samples will be cataloged and placed in a protective

storage cell.

This whole system is to be controlled from Earth. This is carried

out by a continuously monitoring PLC onboard the drill rig. A touch

screen control console allows the operator on Earth to monitor the

progress of the operation and intervene if necessary.

11
2.0 PROBLEM STATEMENT,

In future exploration of the lunar surface, drilled holes and

core samples will be required for various applications in the initial

phases of outpost habitation. Drilling holes for the emplacement of

scientific instruments, setting of structural anchors, and the

acquisition of geological samples for analysis are but a few of the

potential uses of a drilling apparatus. The drilling will be conducted

by a device attached to a three-legged walker called SKITTER (Spatial

Kinematic Inertial Translatory Tripod Extremity Robot).

The lunar environment is different from the terrestrial

environment, therefore terrestrial methods of drilling may not be

directly applied. Characteristics of the lunar environment which

need to be considered include lunar dust, reduced gravity, lack of

atmosphere, micrometeorite bombardment, extreme temperature

gradients, radiation,and the 14-day diurnal cycle.

2.1 Performance Objectives.

The drill will be remotely controlled and will be attached to the

bottom of SKITTER. The system must be firmly anchored and braced

to avoid movement during the drilling operation. It should be light

enough so that SKITTER can easily transport it to the desired drill

site and position it. The device should also be detachable so that

SKITTER can disengage and walk away from it if necessary.

A downward force must be used to effectively drill into the

lunar surface. A minimum downward force of 200 pounds is

required with a rotational speed of 300 RPM. The drill bit must be

able to drill a hole into sandy or solid basalt and must be able to

complete a 50 meter hole without being replaced. About 80 percent

of the heat created by the bit during drilling will be retained in the

12
cuttings. To prevent the bit from overheating, these cuttings must be

removed from the bottom of the hole periodically.

The drilling implement will include enough rods to drill a hole

up to 50 meters deep. The diameter of the hole may range from 50

millimeters to 100 millimeters. Drilling angles of up to 20 degrees

from vertical must be accommodated. Core samples must be

removed and stored periodically during drilling. The removal

process must occur with minimal damage and disturbance to the

samples. Cataloged storage will facilitate later testing and analysis.

The storage system must be able to hold all the core samples which

are removed from a 50 meter hole.

A power supply must be able to allow the drilling of a hole up

to 50 meters deep without recharging. Effective heat dissipation

from the apparatus is necessary so that drilling will be possible

anytime during the 14-day lunar day or night. Control of SKITTER

and the drilling apparatus will be coordinated from Earth. After the

successful completion of drilling, the device will disengage from the

drill hole and return to a base so that core samples can be deposited

and fresh batteries retrieved. The absence of human operators

makes it necessary for the design to be fully maintained robotically.

This will be best accomplished by having the least number of moving

parts and striving for the least complex design possible.

2.2 Constraints.

Size and weight limitations due to transportation have not been

specified. Taking into account the estimated $20,000/lb price tag for

lunar freight, the design should be as light, as possible.

The drilling device will be attached to the triangular base plate

on the bottom of SKITTER which is equilateral with each side

measuring 1.5 meters. SKITTER's clearance of one meter, minimum,

and four meters, maximum, must be included in the design. When

13
retracted, the apparatus should be as non-intrusive as possible to

allow clearance of the legs for "walking".

Temperature extremes and vacuum environment prevent the

use of fluids for cooling of the drill bit and removal of cuttings. Heat

management must be accomplished by the use of conduction and

radiation. Cuttings will be removed from the drill hole by use of a

vertical accelerator and stepped auger. Another characteristic of the

lunar surface which will affect the bit is the constant temperature

(minus 50 degrees C) which exists at a depth of 80 centimeters or

more

Lunar dust is abundant and tends to adhere to all exposed

surfaces. This can cause problems for moving parts. It may also

reduce the effectiveness of radiation panels. The dust will also

adhere to the stored core samples, so a micro-coring device must be

designed to extract small core samples from the drill hole and protect

them from the lunar dust.

Since the moon has no protective atmosphere, solar radiation

and micrometeorite bombardment must be considered in the

selection of materials. Material strength is important due to the high

drilling forces that will be necessary to penetrate the basaltic lunar

crust which has properties close to that of granite.

Reduced gravity on the moon must also be considered in the

design of the drill. Since a lightweight apparatus is desired and the

moon's gravity is only one-sixth of that on Earth, some type of

downward force must be provided for drilling.

A specific control problem arising from the remote operation of

the drill is the 2.6 second communications delay between Earth and

the moon. Visual monitoring of the system may not be practical, and

lack of audible transmission will also make control more difficult.

14
3.0 GENERAL ARRANGEMENT

THE SKITTER (Spatial Kinematic Inertial Translatory

Tripod Extremity Robot) has been developed as a tool in lunar

exploration and in the initial phase of outpost habitation.

Projected duties planned for the robot are drilling, mining, site

preparation and benefication of lunar soil prior to pilot-plant

oxygen extraction and other soil processing techniques.

The drilling implement is one of a series of

interchangeable implements whose individual use will be

intermittent. The drill's uses will include obtaining soil

samples for scientific analysis, mineral locations, foundation

analysis and hole preparation for scientific instrument

placement, structural anchors, excavation explosives and

mining methods. The drill is designed to take advantage of the

lunar environment while operating within the constraints

imposed by that environment.

Figures 3.1 and 3.2 at the end of this section show the

general arrangement of the drilling implement as described

below.

The drilling implement connects to SKITTER via the

structural platform and interfacing latches. Three latches

mounted in the underside of SKITTER are designed so that the

structure can be engaged and disengaged from controllers on

Earth. The space-frame platform supports all parts of the

drilling implement including the power system, robot arm and

rod holders, rods and associated devices, and control box.

The drill motor and power system (consisting of pressure

vessels and fuel cell) sit inside the platform. The motor shaft

15
protrudes through the underside of the structure where it

connects to the drill string. Heat generated from the motor is

transferred away by use of heat pipes. A rejection panel (not

shown) is mounted on the outside of the structure. This is the

primary power system for the main drilling operation.

A rotating plate, supporting a robot arm, is mounted

underneath the structure. The rotating plate allows the twodegree-of-freedom arm to reach up and grab the rods from the

rod holders and connect them to the motor shaft and the sunk

drill string.

A footplate will hold the drill string while the robot arm

connects another rod or device. The rod holders extend out

from the structure in between SKITTER'S legs and use clips to

attach the rods. Approximately thirty rods and associated

devices can be held at a time. The 25 rods that make up the

drill-string are hollow and contain a helically-stepped auger

which serves to collect the broken outer rock for removal by

the vertical accelerator. The rods have specially designed

connections for attachment to the motor shaft, periphery

devices, drill bit, and each other. The full set of rods and

devices can drill to a depth of 50 meters.

Several associated devices will fit in the rod racks and

are integral to the drilling operation. The vertical accelerator

shakes the rods to force the broken rock up the auger and out

of the rods. It is manufactured in the shape of a rod and is

oscillated vertically by a rotary driven cam. Thus, it attaches

and operates identically to the drill string. The macro-core

retrieval device is implemented after the vertical accelerator

clears the small chips away. Its purpose is to break off and

remove the cylindrical rock section left in the center of the bit.

The micro-core retrieval device is implemented after the hole

is drilled and cleared. Its purpose is to remove core samples

from different shaft depths for analysis. Manufactured in the

16
shape of the rods, the device contains its own power source,

drive train, drill, retrieval and storage devices.

The drill bit has a diamond cutting face consisting of 54

facial diamonds. The maximum downward drill force required

is 250 pounds at 300 revolutions per minute. This force

required still leaves ample weight on SKITTER for good traction

between the lunar surface and SKITTER's legs.

A preliminary controls scenario has been designed for the

entire drilling procedure. This scenario is for Earth-based

control and considers the three second time lag between Earth

and the moon, the desired sequence of operation, and the

actuators needed to trigger the individual operations. A

carefully detailed sequence of operation is presented in the

controls section of this report.

The basic sequence of the drilling operation begins with

the SKITTER squatting down and engaging the structure,

supporting and containing all drilling devices, via the latches.

The drilling process initiates with the robot arm placing the

footplate on the lunar surface. The arm then obtains the drill

bit rod from the rod holders and attaches it to the motor shaft.

Drilling occurs as SKITTER slowly squats down to a specified

level. The footplate then clamps the drill bit rod so the rod can

be disconnected from the motor shaft. Once this is done,

SKITTER stands up, the robot arm retrieves the vertical

accelerator apparatus from the rod rack and connects it to the

motor and the sunk drill string. The broken rock is shaken out.

The robot arm and footplate then work to place the vertical

accelerator and drill bit rod back in the rod racks. The macrocore retrieval device is then used to clear the drilled shaft of

the unbroken rock. The drill bit rod is then placed back in the

cleared shaft and held by the footplate. An additional drill

string rod is attached to the shaft and drill bit rod. SKITTER

slowly squats down, drilling to the next level. This process is

- 17
repeated until the desired drill depth is reached. At each rod

change or addition, the footplate is used to hold the drill string

already in the shaft so new rods can be put in place. After the

drilling is completed, the micro-core retrieval device is used,

along with rods to extend to desired depth, to obtain off-hole

samples.

18
Rod

Rod Rack

Robot

Arm

\

Motor

/

BOTTOM OF SKITTER J

t

_1_ &,_ &,J_4..----- mechanical int erf ace

pressure vessels

pipes and rejection

panel not shown

platform/structure

-Rotating Plate

Basic rod, Vertical accelerator device, or

Macro and Micro core retrieval devices,

all have the same basic shape and attaching

method

footplate, carried by

robot arm when

in transit

drill bit

I

Figure 3.0.1 - side view of

drilling implement
ROD

ROD CLIP DEVICE

ROD RACK AND BRACKET

ACTUALLY HOLDS 10 RODS

(ONE ROD RACK ON

EACH LONG SIDE)

POWER PLANT

SKITTER'S LEG AREA

MOTOR

control box

PRESSURE VESSELS

Figure 3,0.2 - top view of drilling platform
4.0 DEEP DRILL APPARATUS SYSTEMS DESIGN

4.1 BIT CUTTING LOADING INTO STEPPED AUGER

4.1.1 Cutting Teeth

In this design, diamonds were selected as the cutting teeth due

to their extreme hardness. Diamonds were also chosen as the cutting

teeth because they require less thrust than other cutting materials.

The teeth will consist of Type IA diamonds, 30 per carat, having a

dodecahedron shape. Past research has shown that cutting teeth

having these characteristics provide the best results. The diamonds

will be bonded to the bit by a powder matrix material infiltrate

tested by Kennametal Corporation in Latrobe, Pennsylvania [Ref

4.1.1] The test results from this corporation showed that Type 4

powder metal would be the best at holding the diamonds in place.

There will be a total of 54 facial diamonds of which 12 of them

will be inside or outside cutting diamonds. An additional 12

diamonds will be used to ream the 100mm hole to 104ram. This

larger hole will allow space for the bit to turn without rubbing the

side of the hole. These diamonds will be placed on the drill bit as

shown in figure 4.1.1.1 in Appendix D. The diamonds are mounted

on small ridges with a negative rake angle of 7.5 degrees to insure

smooth particle flow around the teeth and into the shovel apparatus.

These ridges are at 15 degree angles to push the particles to the

outside of the bit and to prevent them from obstructing the teeth.

For this bit design, the maximum downward force and bit

speed will be 250 pounds and 300 revolutions per minute,

respectively. The reason for using only half of the available

downward force SKITTER can provide, is to insure that SKITTER is

able to have traction on the surface. The bit speed was selected to

keep the heat generated to a minimum. From calculations in

19
Appendix A, assuming steady state and continuous drilling, the

temperature of the cutting teeth would only reach 1092 degrees F.

However, this design will not be continuously drilling which should

lead to a bit temperature of approximately 800 degrees F to 900

degrees F. This temperature is well below the critical temperature of

952 degrees F, at which there is a build up of carbon on the

diamonds. This is called graphitization or oxidation. When this

happens, the cutting ability of the diamond is nullified.

With a maximum cutting depth of 0.2mm for each diamond, the

expected drilling rate will be approximately 61ram per minute

calculated from the following formula:

CUTTING RATE = RPM * CUTTING DEPTH

The teeth will produce a maximum torque of 41.689 N-m as

calculated in Appendix A. From this torque and the additional

torque from the bit dragging on the side of the hole, a set of power

requirements was calculated. These calculations show that the worst

case power requirements of 9.81 kW (13.2 Hp) will be sufficient to

do the cutting. However, in normal drilling, the entire drill string

should not rub the side of the hole. In this case, the power

requirement is 2.8 kW (2.9 Hp).

This drill bit will not be operating continuously for the entire

50m of drilling. The bit will only drill 8mm before the intermediate

holding device is filled. After filling the holding device, the drill bit

will be raised from the bottom of the hole and vibrated up and down

to advance the cut particles up the stepped auger. During this cycle,

the cutting teeth will have time to radiate approximately 50 degrees

F which will help increase the life of the bit.

2O
4.1.2 Shovel Apparatus

The purpose of the shovel apparatus shown in drawing 4.1.2.1

is to scoop the loose particles from the bottom of the bit to the

holding device inside the bit. There will be two shovel mechanisms

to feed both storage devices at the foot of the double helical auger.

The shovel will be made of tungsten carbide instead of diamonds.

The reason for not putting diamonds on the shovel is because its

primary use is to scoop up the loose particles, rather than to actually

do any actual cutting. The front of the shovel is designed to come to

a point in order to remove even the smallest particles cut.

The shovel path has a rectangular cross sectional area of 100

mm 2 that will provide a steady flow of loose particles to the holding

chamber. In order to maximize the size of the holding chamber, the

shovels will be placed at an angle of 40 degrees as shown in drawing

4.1.2.1 in appendix d. These shovels will span the entire width of the

drill face and the reaming teeth in order to remove all of the loose

particles. Only 19.9 Newtons will be placed on these shovels, which

is low enough to guarantee long life. Also the shovels are designed to

last as long as the cutting teeth last. As the cutting teeth wear down

the shovels will wear down and remain functional.

4.1.3 Intermediate Holding Device

The holding device is an intermediary step between the

removal of the cuttings from the teeth and shaking the cuttings up

the stepped auger (Drawing 4.1.3.1). The shovel will dump some of

these cuttings directly onto the stepped auger, which will begin at

the edge of the device, and the remainder into the holding device.

Thus its main purpose will be to prevent the cuttings from falling

back down the shovel and onto the teeth.

21
The need for this device arises due to the fact that about 80%

of the heat to which the teeth will be subjected is contained within

the cuttings themselves. The "hot" cuttings must be kept out of

contact with the teeth as much as possible in order to avoid thermal

failure. Grinding of the teeth over cuttings that have fallen back into

the hole is also avoided.

Due to uncertainty as to how much of the cuttings will travel

directly onto the auger and how much of them will need to be

contained in the holding device, it has been designed as large as

possible, restricted only by the size of the drill bit and the positions

of the holes.

The device will be made from ferritic steel with a high

chromium content; which was chosen to match expansion factors

with interfacing parts.

Calculations show that if the intermediate holding device was

required to hold all of the cuttings, it would be full after drilling

approximately five millimeters. This was arrived at with an

expansion factor of three from solid rock to cuttings.

The possibility of insulating this device from the bottom of the

drill bit and the teeth has been considered. There will be a vacuum

gap between the bottom of the holding compartment and the upper

surface of the section of the bit holding the teeth; which will prevent

heat flow from the cuttings back down to the teeth.

4.1.4 Structural Design

The entire drill bit is 2 meters in length. The overall

dimensions for the bit are given in Drawing 4.1.4.1. The upper

section of the bit from the connection to the upper shafts down to the

splined connection is referred to as the upper bit. The lower 2.5 cm

22
section containing the cutting teeth and shovel mechanism make up

the lower bit. These two sections are separate and removable from

one another. The combined mass of the two sections is 16.4 kg. This

separation is beneficial because it allows for the replacement of the

cutting teeth without disposing of the entire upper bit. The spatial

separation between the lower bit and the holding device minimizes

the heat transfer from the cuttings back down to the teeth. The life

of the cutting teeth is extended by decreasing this downward heat

flux.

The primary responsibility of the spline connection between

the two sections [Drawing 4.1.4.2] is the transference of the torsional

load to the cutting teeth. The downward force on the bit is not

transferred via the two retaining screws in the splined section.

Rather, the end of the spline on the upper bit rests directly on the

lower bit transferring this force. The length of the spline connection

between the retaining screws and the flange on the lower bit was

minimized the shear force on the retaining screws resulting from

slight differences in the thermal expansion between the two sections.

By removing the two screws from the inside shaft the lower bit can

be removed and replaced after it has sustained sufficient wear.

The shovel mechanism described in Section 4.1.2 is two

separate pieces -- one embedded in the lower bit, the other

extending from the front of the holding device on the upper bit. The

scoop which actually lifts the cuttings from the bottom of the hole

extends up into the lower bit to a vertical height of 15 mm. This

results in a hole through the lower bit for removal of the cuttings.

The scoop extension from the upper bit mates with this hole

when the two bit sections are connected. The assembled structure

results in a channel from the cutting teeth to the holding mechanism.

To help seal the channel connection the upper bit scoop extension fits

into a small vertical protrusion from the lower bit. This will restrict

the leakage of cuttings from the channel between the two bit

sections.

- 23
The choice of overall bit dimensions was somewhat arbitrary.

It was decided that a core sample of 50 mm in diameter was

desirable. Thus, a maximum of 25 mm in radial length remained for

containment of the auger. The decision to use the maximum outside

diameter of 100 mm was based on allowing for optimum efficiency

of the stepped auger. However, the major drawback to using such a

large diameter is the high amount of work required to drill as

compared to drilling a small diameter hole. Due to the restrictions of

the project an optimum outside diameter for the drill bit was not

analytically determined.

4.1.5 Material Selection

The extreme operating conditions necessitate close examination

of the material requirements of the core drill bit. In the lower bit,

the material must be wear resistant and capable of retaining

strength at temperatures approaching 500°C. The upper bit is

responsible for transfer of torsional load from the upper connecting

shafts to the lower bit.

Tungsten is highly recommended for drilling in rock as hard as

basalt. Thus, a straight tungsten carbide with 10% cobalt is most

suitable for the lower bit. This sintered carbide will provide

sufficient high temperature strength (Rockwell A > 80) at

temperatures of 530°C. The addition of 10% Co will improve the

impact properties which are of some importance in the design. Other

beneficial effects of using the tungsten carbide are its relatively high

thermal conductivity (150 - 180 W/m.°C) and its low coefficient of

thermal expansion (5.9 micro inches/inch).

The upper bit is not exposed to the extremes in temperature as

in the lower section. Therefore, the use of tungsten carbide is

unnecessary as well as impractical from a fabrication standpoint. A

24
material with a low coefficient of thermal expansion is necessary in

the upper bit because the lower section also will not change

significantly with temperature. AISI-SAE Hl l is recommended as the

material in the upper bit. This hot work tool steel provides a low

coefficient of thermal expansion and high temperature strength

which is of importance in the splined connection. Its low cobalt and

molybdenum content give this steel desirable impact properties as

well. The wear resistance of Hl l is qualitatively fair which is

sufficient for the operating conditions in the upper bit.

4.1.6 Bit Life

The life of the drill bit has been approximated using

experimental data from reference one. Comparing the parameters of

the bit in this reference to those under which the bit being designed

is operating, an estimation of the life can be made. The drill bit in

reference one was tested in basalt in the quarries at Dresser,

Wisconsin, simulating the worst conditions of lunar drilling. From

these tests the bit was found to have a life of ten feet (roughly three

meters), when drilling continuously at 1,000 RPM with up to 400 lb.

downward thrust. The speed at which these tests were conducted

and the fact that the drilling was continuous, are the two major

factors determining the bit life. The bit designed in this report will

not be running continuously, allowing for a cool-off period. It will

also not exceed 300 RPM in operation. From the above it can be

concluded that the life of the drill bit in this paper will be at least

three times the life of the bit in reference one, probably even several

times that number. This corresponds to a drilling depth of at least 30

ft. (about 10 meters), if drilling were continuous. Building and

testing the bit would provide more accurate figures.

4.1.7 Stepped Auger

- 25
The stepped auger is double helical and will operate by

jumping the cuttings up over a series of steps. After reaching the top

of this two meter section, they will travel onto a similar auger in the

next section. For details on the augers see section 4.2.8.

4.1.8 Bit and Rod Interface

The bit and rod interface is the mechanism that will allow the

travel of cuttings between two meter auger sections. It may effect as

much as a 70% reduction in the cutting's flow. It will be a variation

on the collar interface design (section 4.2.4), and is detailed in section

4.2.9.

26
4.2 RODS. AUGER. AND STEPPED AUGER

4.2.1 Introduction

A system for lunar deep drilling requires a method for power

transference from the surface to the bit and a method for the

transport of cuttings form the bottom of the hole back to the surface.

Conventional earthbound technology is insufficient for these

purposes because component mass, interfacing requirements, and the

need for a substantial fluid reservoir to flush cuttings from the hole

do not suit the moon environment.

Several constraints in addition to those imposed by the moon

environment must be accommodated by the power transfer system

(PTS) and the cutting removal system (CRS) (See Fig.4.2.1.1). Some

constraints are imposed on the PTS and CRS by the other systems

which are to be designed to support lunar drilling operations. These

systems require that a 100 mm diameter hole be drilled in two

meter increments, that the interfacing of drill train components be

performed by robots remotely operated from the earth, and that a

variety of arrangements of in-hole components (rods, augers, and

stepped augers) be possible. Specific constraints are the power and

compressive force needed for bit operation (which have been set at

50 ft-lb torque, 150 rpm, and 200 lb), the number of degrees of

freedom of the rod changer which must perform the interfacing

operation (which has been set at two), and the load carrying ability

of the rod changer and footplate clamp (which is not expected to

exceed 200 lb).

4.2.2 Design Considerations

Aside from effectiveness and durability in their designated

tasks, the PTS and CRS were designed with the goal of optimizing

specific properties. Foremost among these properties was the

- 27
minimization of system mass and thus feasibility for transport to the

moon. Secondary considerations were the desire to minimize energy

and articulation requirements for component interfacing, the desire

for overall design simplicity with a minimum of moving parts and

potential failure modes.

Work has been done in the past at the Georgia Institute of

Technology on the design of a stepped auger, or vertically

reciprocating auger (VRA), for the removal of granular material

during lunar drilling. Because of the availability of information on

this design and its potential for effectiveness, it was decided to use a

VRA as the primary component for the CRS. The other component

which was required for this system was a device to remove loose

material from the uppermost portion of the lunar surface as a

prerequisite to the insertion of the bit for drilling hard rock. For this

purpose a simple helical auger similar to those used in soft soil on

the earth was chosen.

For the PTS a system of rods and interfaces was chosen as the

most practical method for the mechanical transfer of the compressive

and torsional forces needed for bit operation and the tensional and

compressive forces required to drive the VRA. Since the number of

CRS components in the hole at any given time must be determined

by the geological conditions in the drilling region, it was necessary to

design all interfaces to correspond with those of the VRA and the

auger. If hard rock is encountered near the surface, there may be

only one CRS component in the hole. Therefore, all in-hole CRS

components not only functioned as PTS components in their

prescribed tasks, but also could be used to operate as rods in the PTS

if necessary.

4.2.3 Material Selection

Because the strength requirements for each component were

similar and because the extreme operating temperature range made

it preferable for all components to have the same coefficient of

28
thermal expansion to prevent interface seizures, all the components

of the PTS and CRS should be constructed of the same material. Since

none of the physical dimensions for any component were dictated by

its function alone (with the exception of the outside diameters of the

auger and VRA), it was decided to minimize weight by the selection

of a material with a high strength to weight ratio and the reduction

of component dimensions wherever stress analysis proved this to be

feasible. For this purpose titanium alloys were considered and a Ti5 A1-2.5 Sn alloy was chosen. Since this alloy is classified as a

refractory metal, there should be no significant loss of properties at

operating temperatures up to 800 degrees Fahrenheit in air and up

to a somewhat higher temperature in space where oxygen

embrittlement will not be a significant problem. Data contained in

NASA SP-5012 indicates that there is consistency in material

properties of Ti-5 A1-2.5 Sn down to -300 degrees Fahrenheit. This

means components constructed from the alloy should be functional at

any low temperature conditions which might me encountered on the

moon. Some properties of this alloy are listed below:

Elastic modulus ........................................ 128 x 104 MPa

Yield Strength........................................... 827 MPa

% Elongation ...............................................10%

Fracture Toughness (at -300 F) ......... 0.8

Density .......................................................... 4500 kg/m3

Power Transfer System Component Designs

4.2.4 Interface Design

The interface design which was decided upon incorporates a

solid inner cylindrical cutout in the form of a dovetail connected to

an outer cylinder which constitutes the body of the component which

contains the interface (rod, auger, or VRA). The interface to be

oriented toward the surface end of the hole has a tapered male

dovetail design (See Fig.4.2.4.1) and the lower interface has the

female counterpart to this configuration (See Fig.4.2.4.4). At the

29
upper interface there is a bistable collar which surrounds the

component and slides on a diametrically opposed set of C-shaped

grooves (See Fig.4.2.4.5) such that in its lower position it surrounds

the upper end of the component but sits clear of the interface. In its

upper position the collar surrounds a portion of the component, the

entire interface, and a portion of the second component to which the

first will be coupled. This design allows for a connection operation as

follows:

(1) The footplate clasps the lower rod by the collar with the

collar in its lower stable position.

(2) The rod changer brings a new rod in from the side such that

the dovetail is fitted into place.

(3) The new rod is rotated, depressed, and rotated again to

bring the collar to its upper stable position covering the

interface.

(4) The motor is connected to the upper rod (see the section

4.2.6 on the motor interface for a detailed description of this

operation).

(5) The footplate releases the collar and drilling is resumed.

The same set of operations is performed in reverse for the

disconnection of the rod coupling.

The nature of the rod interface design can be seen in Figures

4.2.2.1 through 4.2.2.4.

A stress analysis was performed on this design.

failure which were considered:

The cases of

• shearing through the weakest horizontal and vertical planes

in either end

• gross plastic deformation due to axial forces

• plastic deformation of the collar due to interface flexure

30
The case of collar deformation due to radial forces from the

footplate was not considered. In view of the current footplate design

this type of failure is a probable failure mode and warrants further

analysis.

4.2.5 Rod Design

The design of connecting rods for the drilling operation was

dictated primarily by the interface requirements at either end. The

rod itself is simply a titanium tube which is externally identical to

the VRA and the auger (See Fig.4.2.5.1). The outer diameter of the

rods was chosen to be the maximum which the 100 mm hole size and

interface collar requirements would permit. This allowed for a

maximum resistance to failure in torsion and column buckling with a

minimum of rod weight. For this purpose the outer rod diameter

was chosen to be 98 mm, and the inner diameter was chosen to 96

mm. A stress analysis was performed on the rod cross section. The

failure modes which were considered were:

• column buckling

• shear yielding in torsion

• compressive yielding

• tension yielding

• fatigue due to tensile forces exerted by the vertical

accelerator

4.2.6 Motor Interface Design

The motor interface will utilize two permanent magnets coated

with a protective layer of paper. The contact force of these magnets

would be approximately 45 Newtons. The surface of the magnets

will be textured to prevent rotational slipping through a surface gear

arrangement. This contour will work better than a flat arrangement

since it would be held by the material instead of the frictional

relationship.

31
There will be a female section of the Magnet Motor Interface

(MMI) permanently attached to the shaft below the vertical

accelerator. It will have an additional sheath for alignment and

shielding purposes. This sheath will produce a recessed appearance.

A special rod section will be required having a male MMI at

the top and an RIM at the bottom. It will also be equipped with an

auger for use during the first two meters of soil. This assumes that

the hard rock will not be encountered until two meters of loose

material has been removed. This Motor Interface Rod Section (MIRS)

will always be the top rod section on a rod string.

The MIRS will be completely out of the hole at the start of a

two meter drilling period. At the end of the two meters, the MIRS

will be fully inserted in the hole. The diameter of the MMI will be

larger in diameter than the hole to prevent the rod string from

accidentally falling into the hole.

The connection process is as follows:

(1) Rod string is lowered to maximum depth

(2) Footplate grips coupler

(3) Rod changer grips MIRS

(4) RIM is engaged

(5) SKITTER lowers onto the magnet

(6) Drilling begins

The disconnection process is as follows:

(1) SKITTER raises MIRS until the footplate can grip the coupler

(2) Footplate grips the coupler

(3) SKITTER raises up with 45 Newtons

(4) RIM is disconnected

(5) MIRS is returned to rack

32
There are several problems with this interface design. Initial

component insertion is difficult, and adding or removing components

from the rod string requires that two operations be performed per

component. Furthermore, no suitable magnetic material could be

found to implement this design.

4.2.7 Vertically Reciprocating Auger Description

The VRA or stepped auger is similar in appearance to an

ordinary helical auger. The VRA, however, has a series of steps

comprised of vertically oriented fences along its its windings (See

Fig.4.2.8.2). These fences create small compartments that retain dirt.

In operation the VRA is unlike conventional augers. It is not rotated

but instead reciprocates along its axis. In its initial upward

movement the soil particles retained on the steps obtain a significant

upward velocity. When the VRA is then accelerated downward, the

particles are deflected from the upper helical surface in a direction

consistent with their progression up the windings. The particles are

then prevented from sliding back to their original position by the

subsequent fences of the VRA.

4.2.8 Vertically Reciprocating Auger Design

The design of the VRA began with a careful study of the

optimization program developed by an ME 4182 group in the spring

of 1988. The parameter for which this program was designed to

optimize was VRA Volumetric throughput. For the purposes of this

design a high throughput was desirable, but additional consideration

had to be given to the volume retention of the VRA. The VRA was

the most practical in-hole storage device for cuttings as well as the

mechanism for transport of these cuttings. However, the VRA is

incapable of moving cuttings beyond its uppermost step for deposit

into a more appropriate storage device. Thus the geometric

constraints of the hole preclude efficient storage of cuttings outside

of the VRA.

33
The VRA throughput-optimization program was modified to

calculate retained volume in addition to throughput. The input mode

to the program was altered to vary only those parameters which

current design constraints permitted. The number of helices (N) was

permanently set at two to facilitate the transport of cutting from the

bit to the first stage of the VRA and through the VRA interfaces. The

number of fences (FT) was set at six since this had proven to yield a

substantial throughput with a minimal number of particle-fence

collisions in the previous VRA study. Calculations of the retained

volume were performed based on the assumption that the pitch

angle did not vary with the radius. Although the pitch actually

varies in inverse proportion to the radius, this assumption was a

reasonable engineering approximation since the majority of cuttings

are retained at the outer edge of the VRA helix and in the VRA core.

From the modified optimization program it was determined

that the most effective VRA configuration was one with a pitch of

about 4/5 of the VRA outer radius. Although the throughput of this

configuration was much lower than might have been obtained with a

steeper pitch, and the energy loss through particle interaction was

relatively high, this design provided the advantages of high VRA

volume retention and low energy input per stroke requirements.

Based on this design and an estimated three-to-one expansion of cut

rock, it was determined that once the hole is sufficiently deep for the

insertion of nine VRA segments, all of the cuttings produced by two

meters of drilling may be stored in the VRA string.

The final VRA configuration is as follows:

Inner diameter ...................................... 52 mm

Outer diameter ....................................... 96 mm

Pitch ............................................................ 81 mm

Fence height ............................................ 18 mm (45%)

Number of fences per winding ....... 6

Total height ............................................. 2 m

Volume throughput ............................. 48.6 cm3/stroke

34
Volume retention ................................. 5,350 cm3/segment

A VRA was designed within these parameters. The interface

for this component had to be modified from the basic interface

design to allow for the passage of cuttings through the interface.

A stress analysis was performed on the VRA design. The

principle load bearing portion of this structure was considered to be

the outer casing. Therefore, the analysis for VRA failures resulting

from its use as a PTS component were separated from those resulting

from its use in the CRS. The failure cases which were considered

were:

• gross plastic deformation from tensile loading

• plastic deformation due to torsional shear

• column buckling

These analyses were identical to those performed on the rod

cross section. The analysis of failures resulting from tile interface

modification was not performed. The only modes of failure which

this could effect were compressive yielding, where the inner core of

the VRA would provide additional support not available for the rod

interface, and shearing, where the cross section carrying the torsional

load was not significantly reduced.

4.2.9 Vertically Reciprocating Auger Interface Design

The VRA interface was modified from the basic interface

design by the removal of the transverse plate. Torsion was

transferred from the core to the sheath through the last two half

windings of the VRA helix which were reinforced for this purpose

(See item B in Fig.4.2.9.1). The steps of the helix near the interface

were cut flush with the termination of the sheath so that these would

not interfere with coupling and decoupling operations (See item D in

Fig.4.2.9.1). Ports were added to the steps just below the interface so

35
that particles could drop from these steps into the VRA core for

storage (See item A in Fig.4.2.9.1).

When the VRA is in operation, cuttings will be transported to

the interface by a normal VRA progression. Upon reaching the

interface, the particles will be thrown past the lower helical surface

off of which they had been rebounding and will instead be deflected

by the lower surface of a winding in the next VRA segment. A

percentage of these particles will then be picked up in one of the

first steps of the upper VRA segment and begin their progression up

this segment on the following stroke. In this interface the coupling

collar functions to prevent material from dropping off the sides back

into the hole. A net reduction in throughput will occur at the

interfaces due to the relatively large distance which particles must

traverse in order to reach the upper winding and due to the reduced

fence height on the final steps of the lower helix. A conservative

estimate of this reduction was calculated by the input of a variety of

similar geometries into the VRA design program. This estimate was

that as much as a 70% reduction may be expected.

Clearly the VRA interface will result in a substantial waste of

energy since particles will be sliding back to steps in the lower VRA

segment which they have already traversed. Most of this energy loss

will occur at the interface between the lowest VRA segment and the

bit segment since all other VRA segments will have the same

throughput reduction at both ends. This effect might be reduced by

the selection of a steeper pitched VRA to be incorporated in the bit

segment. This VRA design would have to have reduced fence heights

to match the throughput of the interface and a sufficiently steep

pitch to maximize particle contact with the windings of the upper

VRA without requiring a greater input of energy than the vertical

accelerator can provide or than the upper VRA can make efficient

use.

- 36
4.2.10 Auger Description

The auger functions to remove loose material from the hole at

the start of drilling operations. It is a piece of technology borrowed

directly from conventional earthbound drilling in soft soil. The

function of the auger in lunar drilling is essentially identical to its

function on the earth with lower gravity and soil adhesion being the

only relevant environmental differences. The auger moves material

upward as a result of its rotation and downward progression just as a

screw does in soft wood or soap. Two types of auger segment were

needed for this purpose. The first was a 2 m auger with a cutting

edge at the lower end and an interface above. This segment was to

be used to begin the hole. The second was a 2 m auger with two

interfaces (See Fig.4.2.10.1). This type was to be inserted behind the

first segment. Although is is unlikely the material would be

displaced from the bottom of the hole beyond the first auger

segment, the upper auger segments could function to catch soil

dislodged from the sides of the hole and transport it upward.

All auger interfaces were required to mesh with the drive

motor and the connecting rods. The rod-auger connection will be

necessary if the depth of loose material exceeds the combined length

of the auger segments available. Since the auger interfaces were

designed to match the rod interfaces, the augers may also function as

rods. This connection provides an overall weight savings in the PTS

by reducing the number of rods necessary for the drilling operation.

The parameter of auger design which was of primary concern

was pitch. An auger with an extremely steep pitch would require an

impractically large driving torque. An auger with too low a pitch

angle would not provide a suitable throughput and would run the

risk of clogging if it were to encounter large particles. Since the

output torque of the system's driving motor was known to have a

maximum of 50 ft-lb of torque and the depth of loose material was a

variable dependant on the selected drilling location, it was decided to

select a single helix auger with a pitch which was comparable to its

37
diameter. This configuration has proved effective in earthbound

drilling operations. This auger provided the additional advantage of

having a high retention of material within the auger windings (See

Fig.4.2.10.2). Thus, if the depth of loose material is greater than the

ability of the auger to transport material, the auger may be drawn

directly upward after a stall condition is reached; emptied by

agitation, inclination, or reverse rotation; and then reinserted for

further drilling.

4.2.11 Design of Vertically Reciprocating Auger Emptying

Device

The design of the VRA emptying device can only considered

from a conceptual standpoint since there is not sufficient capability

in either the current rod changer or footplate designs to empty the

VRA. The recommendation for the design is as follows:

(1) The stepped auger is detached from the motor.

(2) The auger is swung upward until it hangs at a sufficiently

steep angle for soil to slide along the helix winding.

(3) The auger may be either spun or shaken.

(4) The cuttings will fall into the 2-wheeled dump truck or an

intermediate storage device and be carted away.

4.2.12 Component Fabrication

The fabrication of the components for the CRS and the PTS was

not considered in depth. The titanium alloy specified may be welded

using an inert gas technique and can be shaped by most standard

machining operations. Titanium alloys similar to the one specified

are commercially available in the forms of bar and sheet. Where

applicable in stress calculations, the presence of weldments with

penetration depths comparable to the thickness of the smaller

member was assumed. To construct the specified components, cores,

sheaths, and collars may be reduced from bar stock by lathing.

Fences and helix windings may be cut and drawn from sheets. The

38
fabrication of the interface geometry will be difficult, but a series of

lathing and milling operations should be capable of producing it. All

connections may then be welded. These weldments should be

inspected for defects by X-ray or ultrasound techniques.

39
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4.3 FOOTpLATE

4.3.1 System Operation

The footplate provides a stable platform for supporting the

drill string during rod string addition or removal. For the following

description of operation procedure, it will be assumed that the

footplate will be stored in the ground with an adaptor rod attached.

(This adaptor rod provides the interface between the drill drive and

the footplate.) Initially, the adaptor rod with footplate will be

attached to the drill drive. SKITTER will squat down on the adaptor

rod connected to the footplate to drive the serrated casing as far

down into the ground as possible. SKITTER will then screw the

footplate counterclockwise into the ground. The serrated casing eases

the torque requirements while the base fins on the bottom of the

footplate provide torsional stability. Once the footplate has been

completely driven into the ground, SKITTER will turn the drill drive

1/2 turn clockwise to release the adaptor from the footplate. The

rod changer places the adaptor into the storage rack, retrieves a drill

string and connects it to the auger. Drilling may then begin.

As each drill string is driven its length into the ground, a

signal is sent to unlock the drive, causing the spring to clamp closed

around the drill string. It is the tension in the spring which causes

the clamping to occur; no motor power is required in this motion.

Once the next string has been added, another signal causes the lifters,

operated by a battery-powered motor, to rotate the spring open. The

procedure for removal of drill strings is similar. Once drilling

operations have been completed, all drill strings and auger sections

are removed and the adaptor is reconnected to the footplate by the

rod changer. The drill drive will engage with the adaptor and

unscrew the footplate from the ground. SKITTER can then walk to

the next anticipated drill site and the drill drive can screw the

footplate into the ground again. SKITTER can leave the footplate in

this position for storage until the new drilling operation is set to

begin.

40
4.3.2 Base

The base houses all working mechanisms of the footplate.

Made of aluminum-lithium, it is 550 mm in diameter and 855 mm in

height and is welded together. Additional pieces are welded onto the

base unless otherwise specified. See drawing 4.3.2.1 for footplate

plan and elevation views.

4.3.2.2 Casing

The casing prevents loosely-packed regolith near the surface

from collapsing into the newly drilled hole. Originally, variable

depth casing which could be cut to size was to be used. After further

investigation, however, it was determined that only the regolith 200

mm from the surface was loose enough to pose a problem, so a fixed

casing was designed. This casing, 200 mm in length and 120 mm in

diameter, is made from aluminum-lithium, like the base, to minimize

thermal expansion differences and is permanently attached to the

footplate. The bottom of the casing is serrated with four equallyspaced teeth to ease the torque requirements needed while rotating

the footplate into the regolith. These four teeth have bottoms to

prevent shear failure under torsion. Finally, the exterior of the

casing has a double helix thread to further ease the torque

requirements necessary to screw the footplate into the ground.

4.3.2.3 Fins

The four fins on the bottom of the footplate are designed to

prevent the footplate from being forced out of the ground while

drilling operations are in progress. The four fins, equally spaced and

triangular in shape, and welded to the footplate base at their top

and join the casing at their bottom. (Approximately 10 cm from the

bottom of the casing.) These fins provide the torsional stability and

are cut out to reduce their total weight..

4.3.2.4 Adaptor

41
The interface with the drive will be similar to the interface of

the drill string. This adaptor will have a special connection to the

footplate that will be able to withstand the amount of torque needed

to drive the footplate into the ground. The adaptor will be made of

aluminum-lithium and will be of the same diameter as the rods (98

mm) so that it may be stored in the rod rack. The interface to be

attached to the footplate will be of the pin-groove type. The

footplate will contain the grooved (female) section while the adaptor

contains the male end, a pin which runs through the diameter of the

adaptor rod. The pin will enter the groo_(ed section, rotate through a

slot, and lock. This mechanism will allow for the footplate to be

suspended from the drive without allowing the footplate to fall.

Drawing 4.3.2.2 shows the design of the adaptor.

4.3.2.5 Signalling

A flexible antenna, 500 mm in length, will be mounted to the

outer perimeter of the footplate. This antenna can withstand

bending motion and will be able to function even while covered with

regolith. Accompanying the antenna will be a transmitter box with

two relays. The relays, manufactured by Opto, will be solid state,

single-pole, single-throw. One relay will turn the motor on and off;

while the other relay will enable and disable the locking mechanism

in the motor/gear box. A radio frequency signal from the controls

group will operate the relays.

4.3.3 SPRING

The spring is the device that grips the drill string or auger

while other drill strings are being added or removed. The spring is

actually composed of a section of aluminum-lithium which is

deflected upward and outward by lifters. It is shaped like a leaf

spring with a hole cut in the center and slits cut along both sides of

the hole and is similar in operation to a speed washer. (A drawing

of this can be seen in the appendix as Figure 4.3.) The spring is 1.5

mm thick, 200 mm wide, and 500 mm long. In the closed position,

42
the two sections of the spring are flat and separated by a distance of

100 mm so the drill string can be tightly gripped. Figure 4.3.2 shows

an elevation view of the spring in the closed position holding a drill

string. In the fully open position, the two sections of the spring are

separated by a distance of 120 mm allowing a 1 cm leeway on either

side of the opening to account for stray drill string motion. Figure

4.3.1 shows a side view of the footplate with the spring in the open

position. The upper section of the spring is reinforced by aluminumlithium ribs to help prevent buckling and fatigue failure.

4.3.3.1 Grippers

Grooved aluminum-lithium pads on the ends of the spring

grasp grooved sections on the drill string couplings to increase

gripping capability. These pads are rounded to follow the contour of

the strings and the auger and are 25 mm in length so that the auger

can be held in a stable position. The grippers are welded to the

center tongue of the spring.

4.3.3.2 Spring Support

In addition to spring stops, the spring will be supported by

ribbing which runs the whole length of the deflecting flanges. Two

equally-spaced lengthwise ribbings will reinforce the flanges and

will help prevent the spring from overflexing.

4.3.3.3 Spring Stops

The base contains aluminum-lithium stops underneath the

spring to provide additional support to the spring material and to

help prevent the spring from overdeflecting. These stops,

rectangular in cross-section and 20 mm in width, are placed at the

lip of the spring opening and are welded to the base structure. They

are curved in shape to closely fit to the spring opening. The stops

are placed 5 mm from the grippers creating a cantilever

arrangement.

43
4.3.4 Lifters

The lifters are the mechanisms which cause the spring to

deflect open and, therefore, release the drill string. The deflection

mechanism for the spring will be similar in operation to a camshaft

in an internal combustion engine. The lifter lobes are actually arms

connected to a shaft with a roller at the end which is in contact with

the spring. These lobes will act on each side of the drilling hole.

There will be a dry sealed roller bearing at the end of the lifter. The

roller bearing reduces the amount of torque required to turn the

lifters because the sliding motion is replaced by rolling motion. The

two lifters will be connected by a solid shaft which runs just outside

the radius of the drilling hole. The shaft will be mounted in brass

bushings at each end. One end of the shaft will be connected to a

motor with a sealed internal gear reduction mechanism. The force

required to lift the spring was calculated to be 41.95 N and the lifter

must raise the spring 38.3 mm up at the center.

4.3.5 Motor

A motor will be used to turn the shaft that lifts the spring to

release the rod string. The motor will be a Micro Mo Electronics, Inc.

Gearhead Series 34PG. Within the motor's sealed casing is an

internal gear reduction set which reduces the torque and size

requirements of the motor itself. There will be a sealed output shaft

from the gearbox that is connected to the shaft. The operating

temperature range of the motor is from -55 C to 125 C, so a

molybdenum blanket will be used to insulate the motor. Also, the

footplate will be covered in insulating regolith soon after drilling

operations begin, further protecting it from extremes in temperature.

The motor and gear set, weighing a total of 451 grams, can

provide a maximum torque of 3.5 N-m at the output shaft. The gear

ratio is 1080:1 and the motor assembly is 87.33 mm long. The motor

44
will be connected to the base by mounting threads which are on the

front of the gear reducer.

4.3.6 Power Storage

It is assumed that the power required to operate the

transmitter and relays is relatively negligible compared to the power

required to operate the motor. To power the motor, four WhittakerYardney #LRO5 Silvercels were chosen. These will be connected in

series to provide a total of 0.5 Amp-hrs. At 6.0 V (1.5 V each), these

cells are suitable for long life and low-rate use. A recharge will not

be required in 50 holes. The total mass of the four cells is 0.0817 kg

and the cells will occupy a space of 1505 mm 2 at a maximum height

of 39.62 mm (including terminals). The batteries will also require an

insulating molybdenum blanket to protect them from temperature

extremes.

4.3.7 Materials

Aluminum-lithium alloys are best known for their low density

and corrosion resistance. Electrical conductivity, ease of fabrication,

and appearance are also attractive features. Aluminum-lithium

alloys are part of a class of alloys known as high modulus alloys,

isotropic metallic materials having a specific modulus higher than

that of conventional aluminum, titanium, or iron base alloys. The

weight saving potential of aluminum-lithium is now widely

recognized and significant improvements in production capabilities

and mechanical properties of alloys produced on a commercial scale

have been achieved in the last two years.

In a solution treated condition, some aluminum-lithium alloys

have low strength but high ductility which allows considerable cold

forming. Strength can be increased substantially by aging, especially

if the material has been cold worked before aging. A1-Li castings

with good surface quality can be made under controlled conditions.

45
The particular aluminum-lithium alloy chosen was aged

Weldalite, an alloy produced by Martin-Marietta. This alloy has

suitable properties for the footplate design and can be welded to

satisfaction, a problem with some of the aluminum-lithium alloys.

See Table 2 for the range of properties of this alloy.

4.3.8 Analysis

See following pages for analysis.

4.3.9 Mass Analysis

See Table 1 for weight calculations.;

4.3.10 Regolith Protection and Removal

The drilling of a 50 m deep hole will involve removing 0.29 m 3

of regolith and depositing it upon the footplate. This amount of

material would weigh about 465 kg and posses a lunar weight of 760

N. This volume and mass of cuttings could severely impair the

operation of the footplate. The dirt could all pile on top of the spring

mechanism either crushing it or rendering the motor useless under

the weight. These cuttings may also find their way beneath the

spring causing it to remain partially open and reducing the

effectiveness of the grippers clamping ability. To avoid this problem

a cylindrical bellows will be utilized between the base plate and the

spring to keep the cuttings from getting underneath the spring. The

bellows will be made of a high nickel alloy such as Mech metal with

properties maintained in temperature ranges of -350 C to 250 C.

Furthermore, a structural cover will be used to support the cuttings

over the spring device and keeping the footplate operational.

4.3.11 Failure Analysis

Unfortunately, there are few parts on the footplate that can fail

and still allow the footplate to continue operating properly. No

46
redundant systems have been incorporated into the design. Should

any part of the motor or gearbox fail drilling operations would come

to an abrupt halt. If the shaft, a lifter, a roller, a relay, the

transmitter, or the spring fail, the result will be the same. If a fin

breaks off of the bottom of the base there would only be a slight

balance problem, but drilling can continue. If the casing should

break in some way the extent of the setback will depend on the

severity of the damage, and the footplate might again be slightly

unbalanced. If the adaptor breaks drilling can not continue past the

hole that the footplate is currently in. Should a cell in the battery go

bad the voltage output will be reduced by 1.5 volts, and more power

will be drawn out of the battery If a cell goes bad early enough in

the drilling operations, a dead battery might bring a premature

ending to the operation of the footplate, and, therefore the drilling.

47
FIGURE 4.3.1

SIDEVIEW OF FOOTPLATE WITH

SPRING OPEN

i
FIGURE 4.3.2

ELEVATION VIEW OF SPRING

CLOSED ON DRRL STRING

S

__ 5pc; _"3
4.4 POWER SUPPLY, ROTARY DRIVE, AND HEAT

MANAGEMENT

This section addresses the power supply, motor, and heat

management requirements of the lunar deep drill. Figure 4.4.a is an

outline of the design process used for these elements. Power will be

supplied by a hydrogen-oxygen fuel cell. This system, which has

been used on the Space Shuttle Orbiters, has an output range of 2 to

12 kilowatts. This will adequately supply the maximum of 10

kilowatts necessary for the worst case scenario; however the fuel

requirements were designed for a normal operating point of 4

kilowatts of power output since the drill motor was designed to

operate at 2.81 horsepower and 300 RPM. This design point may be

too low for successful operation of the drill due to a lack of

communication between design groups regarding power

requirements near the end of the project. The on-board hydrogen

and oxygen supplies will be contained in external cryogenic tanks

which will be re-filled between drilling operations. Excess heat

generated by the power supply will be conducted away by two heat

pipes and dissipated to deep space by two radiator panels located on

the periphery of the implement.

The drive mechanism for the drill will be the newly-developed

Curvo-Synchronous Motor (CSM), which has only one moving part,

operates at temperatures up to 350 degrees F, and has an efficiency

of 91.3%. Heat generated by the motor will be transferred to a

thermal radiator by a single heat pipe. The heat pipe will also act as

a thermal diode to protect the motor if the radiator panel should

overheat. The radiator panels will be oriented toward deep space by

the positioning of SKITTER at the drilling site. Finally, a torquelimiting friction disc clutch has been incorporated into the design to

protect the motor in the event of drill bit lock-up. This clutch will

"run free" at an overload torque of 80 ft-lbs and will automatically

re-engage when the overload has been removed. Two different 3-

48
dimensional views of the components from this section are shown in

Figures 4.4.b and 4.4.c, and a plan view is included in Figure 4.4.d.

4.4.1 Power Supply

The design of the power supply system for the lunar deep drill

apparatus (LDDA) involved fulfilling many of the general design

criteria for the implement as a whole such as minimum weight,

manageable volume, durability, and ability to operate in the lunar

environment. The basic concept of the power supply system as a

whole was that the power supply on-board the implement would

serve as a temporary source which would be recharged or replaced

by an external source, such as a solar array or some other outside

source. The specific requirements of the on-board power supply

were that it be compact enough to fit into the implement, be light

enough to be portable, and that it be able to supply enough power to

drill at least one complete 50 meter hole without needing to be

recharged or replaced. A hydrogen-oxygen fuel cell power plant was

the device chosen to provide the necessary power to the LDDA. Its

relative light weight, durability, low maintenance requirements, and

operating characteristics made it the obvious choice. The only other

feasible alternative was the use of rechargeable batteries, which

ultimately proved to be far in excess of the weight limitations of the

project.

By summing all of the power requirements of the various

groups, the capacity of the power supply system was set at 100

kilowatt-hours. The fuel cell power plant design chosen was one that

has already seen extensive use on the Space Shuttle Orbiters. This

system, manufactured by United Technologies, is capable of

delivering power from a minimum of two kilowatts to a peak output

of twelve kilowatts, and has physical dimensions and weight that are

suitable for use in such an application. The total power output

requirement during the entire drilling operation occurs during the

49
actual drilling of the hole when the motor will require four kilowatts

of power. The length, width, and height of the power plant are 35 cm

x 38 cm x 101 cm, and it weighs 91 kilograms. These dimensions

and weight apply only to the power plant; the fuel for the power

plant will be contained in external cryogenic tanks, and the byproduct water will be collected in its own tank.

The fuel cell power plant that will be used essentially involves

a reaction between hydrogen and oxygen which can be characterized

as reverse hydrolysis wherein the reactants are combined to produce

electrical energy and water as a by-product. In addition to

producing the desired power for the drilling operation, the power

plant will produce roughly 1500 watts of excess heat that must be

dissipated. This heat will be conveyed through heat pipes to radiator

panels for dissipation. The power plant and the necessary tanks are

shown schematically in Figure 4.4.1.1 inside the skeletal structure of

the implement. The power plant consists of three basic parts: the fuel

cell stacks, the control and electrical components, and the necessary

plumbing to convey the fuel and water to and from the plant,

respectively. The heart of the power plant is a pair of two parallel

stacks of fuel cells in which the hydrogen-oxygen reactions occur. At

the operating point of four kilowatts, the cells are capable of

producing 31.5 volts. This particular power plant is designed to

operate as an unregulated direct current source and any necessary

power conditioning must be done on the output of the power plant

before it is used. Recent developments in power semiconductors

have made this requirement possible as the technology available to

date allows for the conversion from a D. C. source to an alternating

output if such output is mandated. However, for the purposes of

operation of the LDDA, only an increase in the output voltage of the

power supply is required.

The capacity of the power supply depends solely on the

quantity of fuel available on-board the LDDA at any given time.

Therefore, by adjusting the reactant tank sizes, a similar adjustment

of the capacity of the power system results. At an operating point of

5O
four kilowatts of output, the power supply will consume 0.087

pounds of hydrogen per kilowatt-hour, 0.69 pounds of oxygen per

kilowatt-hour, and will produce a maximum of 3 pounds of water per

hour of operation. For the complete drilling of one hole, the total fuel

requirements will be 17.4 pounds of hydrogen, 69 pounds of oxygen,

and the amount of by-product water will be 75 pounds of water.

Using this data, the necessary tank capacities were determined to be

0.5585 cubic meters for the hydrogen tank, 0.0275 cubic meters for

the oxygen tank, and 0.034 cubic meters for the by-product water

tank.

Using the given tank capacities and the space limitations of the

structure, cryogenic tanks for the reactants were designed as well as

a holding tank for the by-product water. Since little information was

available on cryogenic tank design, a rigorous design for the reactant

tanks was not possible; however, some recommendations on the tank

design are in order. First, a selection of materials is necessary for the

actual pressure vessel. Probably the leading choice is a Kevlar

composite or due to its high strength, and its light weight. The

insulating material should be a multi-layered reflective coating such

as mylar with a metallic coating such as aluminum. One way to gain

the best insulating characteristics for such a material would be to

fabricate corrugated sheets and alternate the orientation of

successive sheets. In this manner, the highly desirable insulating

properties of a vacuum can be used to an advantage. Finally, the

layout of the pressure vessel should be designed so as to minimize

any stress concentrations that might occur because of improper

geometry. Bearing these requirements and considerations in mind,

an effort was made to calculate a rough estimate of the size required

for each of the reactant tanks. In order to minimize both space

limitations and possible stress concentrations within the tank, a

cylindrical tank geometry with hemispherical ends was chosen. It

was assumed that three inches radially would allow enough space for

both the pressure vessel and the mylar insulation. In addition, it

was decided that when the tanks were filled to desired capacity, that

an additional five percent of tank capacity should be left to allow for

51
venting of reactants in gaseous form for emergency temperature

maintenance if necessary. Using these requirements, the sizes of the

reactant tanks were calculated as follows. Since the size of the tanks

and hence the structure that contains them should be minimized, a

minimum diameter was sought which would allow this minimization

of space within the structure. The dimensions of the base of the

structure were predetermined by the structures group (See Section

5). Using these dimensions, it was seen that a maximum length for

the tanks was 1.23 meters or slightly more if the tanks were going to

fit between the motor and one side of the structure. Using this

length as a constraint for the larger tank size, a determination of the

diameter was made using the following relation between the tank

volume and the size of the tank which was characterized by the

length of the cylindrical section and the radius of the cylinder and

hemispheres.

V = (rc)(R2)(L) + (4/3)(rtR3)

After successive iterations, the final hydrogen

determined to be

L = 31.24 inches

R = 8.86 inches.

tank dimensions were

Once the radius of the cylindrical section was set, the length of the

oxygen tank was determined by simply rearranging the equation to

solve for the length. The resulting dimensions of the oxygen tank are

therefore

L = 14.40 inches

R = 8.86 inches

Finally, the minimum volume for the water tank which was given

earlier calls for an tank size which will include the necessary

insulation and heating capability to maintain the waste water in a

liquid state. It is recommended that the water reservoir be insulated

in the same fashion as the reactant tanks and that heating of the

tank to avoid freezing of the water be accomplished by resistive

heaters built into the tank. In this manner, temperature

52
management can be closely monitored and adjusted by the on-board

PLC ( See Section 8). Based on this recommendation a cubic tank 45

cm on a side should be large enough to accommodate both the

required capacity as well as the insulation and heat controls.

4.4.2. Motor

When choosing a motor to operate the drill string, several

aspects had to be considered. First, the efficiency of the motor was

crucial. Low efficiency means that much excess heat will need to be

dissipated. Due to the negligible lunar atmosphere, this heat cannot

be convected away. With only conduction and radiation left, a highly

inefficient motor would pose serious problems. In addition to

thermal considerations, physical limitations also had to be taken into

account. The motor fits onto the supporting structure attached to the

bottom of SKITTER. Weight is minimized so the LDDA can easily be

picked up and transported by the walker. The power was

determined by finding the torque necessary for successful drilling.

The normal operating torque, assuming that there is no friction on

the sides of the drill string, is :

T = (f) (W) (r)

where: T is in ft-lbs

W is in lbs

r is in (ft)

Therefore ...

T = (1.5) (200 lbs) (1.96/12 ft) = 49.21 ft-lbs

and at the normal operating speed of 300 RPM...

HP = (49.21 ft-lbs) (300 RPM)/(5252)

HP = 2.81 horsepower

53
Considering these parameters, the choice was the recentlydeveloped Curvo-Synchronous Motor (CSM). This motor was

developed by Dr. Kent Davey from the Electrical Engineering

Department at the Georgia Institute of Technology. The CSM has

many desirable characteristics. The motor has only one moving part

so that inertia and friction losses are small. The innovative winding

scheme is such that the flux path between the magnets is minimized.

See Figure 4.4.2.3 for a schematic of the CSM. Due to the radial

positioning of the magnets, a large number of thin magnets can be

used thus providing a thin disc-like shape. Efficiencies of 95% to 98%

are not uncommon and the motor runs at temperatures up to 350

degrees F. Conventional permanent magnet motors were rejected for

the same reasons that the CSM was chosen. They usually have a low

efficiency of 30% to 50%, weigh a lot, take up more room, and are less

tolerant of heat.

The particular motor that will be used is a 5 horsepower CSM.

The 5 horsepower will accommodate some of the extreme values

seen in Appendix A yet provide smooth operation at the normal

operating speed and torque of 300 RPM and 49.21 ft-lbs,

respectively. A torque limiter set at 83.3 ft-lbs will prevent motor

overload, see 4.4.5, Friction Clutch, for details. Because the CSM is a

direct drive system, a thrust bearing is included in the the design to

transfer applied loads to the frame of the motor. A diagram of this

thrust bearing is shown in Figures 4.4.2.6 and 4.4.2.7. The CSM's load

characteristics appear in Figure 4.4.2.1 and its current characteristics

appear in Figure 4.4.2.2. See calculations in Appendix A for the

details of these two graphs. The motor has a diameter of 18 inches, a

thickness of 2 inches, and weighs approximately 40 pounds according

to Dave Ross, a CSM specialist from the VA Hospital in Decatur,

Georgia. The efficiency of this motor was found by multiplying the

winding resistance, R, by the square of the current, I. In this case,

933 volts and 4 amps are required. So ...

Heat loss = (22.135 ohms)(4 amps^2) = 354 Watts

54
The efficiency is then ...

n = (3730 Watts)/(3730 + 354 Watts) = .9133 = 91.33 %

where:

3730 Watts is input power

354 Watts is power loss

To dissipate the 354 Watts of excess heat, we employ a "sheath" of a

highly conductive aluminum alloy which covers the sides motor. A

heat pipe is then used to transfer this heat to a radiator panel. This

motor mount and flange can be seen in Figures 4.4.2.4 and 4.4.2.5

and also provide for the mounting onto the supporting structure.

The materials used for the "sheath" were an Aluminum alloy, 1199,

and a Titanium Beta alloy. The justification for this choice is simple.

First, no heat should be radiated to SKITTER; it should all be

conducted to the heat pipe for proper disposal. With this in mind, a

material with a low thermal conductivity and emmisivity value was

needed for material "B" and a highly conductive material, Aluminum,

was needed for material "A", refer to Figure 4.4.2.4. Secondly, the

materials must have similar coefficients of thermal expansion to

permit welding. The two materials that satisfied these criteria are

AA 1199 O and Ti Beat Alloy. The Aluminum alloy has a linear

coefficient of thermal expansion of 13.3/deg F x 106 and a thermal

conductivity 1540 BTU/ft2(h)(in.)(deg F) while the Titanium alloy has

10.1/deg F x 106 and 119 BTU/ft2(h)(in.)(deg F). This arrangement

will inhibit heat radiation towards SKITTER's under-belly, with the

help of a coating on the Titanium alloy, and promote conduction to

the heat pipe. Dissipation of heat is crucial because the CSM is not

operable at temperatures exceeding 350 degrees F. For further

details, see sections 4.4.3, Heat Pipes, and 4.4.4, Heat Radiation.

55
4.4.3 Temperature Management

Three heat pipes will provide proper temperature control of

the CSM and the fuel cell power plant.

Heat Pipe Concept

The heat pipe is a device with an effective thermal

conductivity of as much as 500 times that of copper. This property

makes it especially suited for applications in which a large amount of

heat must be transferred from one area to another. Figure 4.4.3.1

illustrates the basic concept for a standard heat pipe. The three basic

elements are the working fluid, the porous capillary wick, and the

outer pipe structure. The wick is saturated with the liquid phase of

the working fluid and the remaining volume of the tube contains the

vapor phase. Heat applied at the evaporator vaporizes the working

fluid in that section. The resulting difference in pressure drives the

vapor from the evaporator through an adiabatic section to the

condenser where it condenses thus releasing heat and energy. The

liquid enters the wick surface where a capillary pressure is

developed. This capillary pressure pumps the condensed liquid back

to the evaporator where the process can begin again. This passive

heat exchange process can continue indefinitely if the heat pipe is

properly designed.

Some of the advantages of utilizing heat pipes to regulate the

temperatures of the motor and the power supply are:

a. Heat pipes are able to transfer high heat fluxes with

small temperature gradients. A heat pipe is therefore

very nearly isothermal and close to the saturated vapor

temperature at all times.

bo Heat pipes do not require any external pumping power.

Fluid movement is due solely to capillary forces rather

56
than gravitational forces in returning the condensed

liquid to the evaporator.

C. The heat pipe can act as a thermal diode (i.e. operation in

one direction only). This allows for the heat pipe to "shut

off" if the heat sink or condenser side becomes hotter

than the heat source side. This can be achieved by an

adequate design of the wick configuration such that the

liquid capillary pumping exists in one direction only.

do

e°

The heat pipe can be manufactured into many common

shapes (i.e. rectangular and circular) and can be bent or

flexed to accommodate unusual geometries. The

flexibility feature will be used in this design.

The absence of gravity allows it to be more efficient

lending it to be advantageous in space applications.

A limit to its heat

disadvantage of a heat pipe.

transport capabilities is the

These limitations are due to:

largest

a. capillary pumping ability,

b. choking of the vapor (i.e. sonic limitation),

Co tearing of liquid off the liquid-vapor interface by vapor

flowing at high velocity (i.e. entrainment limitation), and

d° the disruption of the liquid flow by nucleate boiling in

the wick (i.e. boiling limitation).

These limitations will determine the requirements for the

optimal design of the three basic elements of the heat pipe; the

working fluid, the wick, and the pipe structure.

57
Design of the Heat Pipe

The design procedure used is outlined

INITIAL CONSTRAINTS _

FLUID, PIPE, WICK [

SELECTION

PIPE DIAMETER DETERMINED

SO THAT SONIC LIMIT NOT EXCEEDED

WICK DETAILS SELECTED SO THAT

CAPILLARY LIMIT NOT EXCEEDED

MECHANICAL DESIGN TO DETERMINE ,

CONTAINER DETAILS

CHECK ENTRAINMENT AND BOILING

LIMITS TO ENSURE PROPER OPERATION

OF HEAT PIPE

MPLETED HEAT PIPE

below:

58
The geometry of the heat pipe will be the initial constraint for

its design. A circular cross section with flat circular end caps is the

chosen geometry for this design. It is the most common and easiest

to manufacture. The first heat pipe to be designed will be the one

which removes the excess heat developed by the motor. Then by

simply following the same steps, the two heat pipes for removal of

waste heat from the power supply will be accomplished. Figure

4.4.3.2 illustrates the space requirements for the heat pipe for the

motor. The heat pipe will be wrapped around the motor so that

temperature management over the entire surface of the CSM will be

properly controlled. The peak power output of the CSM's power

cycle will not exceed 5 horse power. With an efficiency of 91.33 %,

346 Watts must be adequately dissipated. The temperature of the

CSM will be operated at 212F to ensure safe operation. We have

assumed that the heat will be generated uniformly by the CSM so

that the required dissipation of heat is shown below:

(346 Watts) x factor of safety of 2 = 692 Watts

The heat pipe will be 173 inches long so it may work in coordination

with the radiator located 59 inches away in the rod rack. The

radiator will be used as a heat sink which will be discussed in more

detail in the next section. The evaporator and condenser section will

have a length of 61 inches and 56 inches, respectively. Because of

the width of the radiator, the overall diameter of the heat pipe

cannot exceed 1.5 inches. Three U brackets with corresponding pop

rivets will be used to connect the heat pipe to the sheath of

aluminum about the CSM. Similarly, two such brackets will hold the

heat pipe to the radiator panel.

For the successful operation of a heat pipe, its working fluid

must have a melting point temperature below and a critical point

temperature above the pipe operating temperature. For a pipe

59
operating at 212F, water, methanol and ammonia are suitable

working fluids. The capillary limitation equation will be used to

determine the most suitable liquid. The capillary limitation

represents the required pressure balance at the liquid side of the

liquid-vapor interface in the wick. An expression for capillary

limitation on the heat transport factor (QL)c,max is represented by:

(QL)c,max = 2( ) (_)k (2nrvtw)

k/re

2Hrvtw

-liquid transport property

- wick property

- wick cross sectional area

The larger value of the heat transport factor will thereby give a

larger upper capillary limit. The heat transport factor is directly

proportional to the the liquid transport factor. Water has a larger

liquid transport factor and has higher conductance properties than

methanol or ammonia. Therefore water has been chosen as the

working fluid. According to Professor Gene Colwell at the Georgia

Institute of Technology, freezing of the water during the lunar night

will not pose a problem to the heat pipes performance. The heat

pipe will act as a heat sink when the temperature falls below 32F.

As the temperature rises due to the operation of the CSM, normal

operation of the heat pipe will resume.

Selection of the pipe and wick material depend largely on their

compatibility with the working fluids. Two major concerns of

incompatibility are corrosion and the generation of noncondensable

gas. Solid particles from corrosion may be transported with the

flowing liquid to the evaporator region leading to an increased

resistance to fluid flow. The generation of gas will have the effect of

inactivating a portion of the condenser. Copper, nickel, and titanium

are compatible materials for a water pipe. Copper is chosen because

of its superior conductance properties to nickel and titanium and to

its relative low cost.

60
The size of the pipe diameter necessary for a given application

should be determined so that the vapor velocity is not excessive. At

high Mach numbers, the compressibility of the flow of vapor may

cause choking to occur. Therefore, the heat pipe will be designed so

that its maximum Mach number does not exceed 0.2. Incompressible

and laminar flow can now be assumed and the sonic limitation is

immediately satisfied. Under these conditions the required vapor

core diameter is determined by:

(20 Qmax)

(np&._ 7vRv%)

The vapor core diameter is:

dv = 0.10 inches

with a sonic limitation of:

Qs,rnax = 4.27 x 10 6 Watts.

The capillary limitation is the most common of the four

limitations. During normal operation of the heat pipe, vapor and

liquid flow past each other separated only by the wick. There exists

a pressure gradient called the capillary pressure along this liquidvapor interface. The capillary limitation on the heat transport factor

was determined by:

(QL)c,max =

Peru

FI - Fv

and represented by Figure 4.4.3.4, Figure 4.4.3.5, and Figure 4.4.3.6

of Appendix A. This relationship yielded the following wick

dimensions:

61
Mesh number (N)

Wick area (Aw)

Wick diameter (d)

= 250 inch -1

= 0.9818 inch 2

= 0.0016 inches

with a capillary limitation of:

Qc,max = 700 Watts.

The container tube dimension can now be determined.

maximum hoop stress subject to internal pressure is given by:

The

P(do 2 + di 2)

fmax = (do2 _ di2)

di = 0.1.12 inches

do = 0.1.25 inches

tl = 0.13 inches

A factor of safety of 2 was used in the calculation. For the most

common circular heat pipes, a flat circular end cap is used. Assuming

a thickness of less than 10% of the diameter, the wall thickness of the

end cap can be determined by:

Pdo

fmax - 4t2

t2= 0.06 inches

Again, a factor of safety of 2 was used. The remaining space

between the vapor core and the inside of the heat pipe wall will be

left for the wick and liquid flow.

The basic parameters of the heat pipe have now been designed.

The entrainment and boiling limitations must be checked to ensure

62
the proper operation of the heat pipe. The entrainment limit ensures

that the wick will not dry out. Since the vapor and liquid move in

opposite directions, a shear force exists at the liquid-vapor interface.

If the vapor velocity is high enough, the liquid may be torn from the

surface of the wick causing the wick to dry out. The entrainment on

the axial heat flux is determined by:

_J _rv Qe,max = )_Av 2rh s

and represented by Figure 4.4.3.7 of Appendix A. The value of Q/Av

for the designed heat pipe is 1.78 x 105 Btu/hr-in 2. The entrainment

limit, Qe,max/Av, of 2 xl05 Btu/hr-in 2 will not be exceeded.

The boiling limit is associated with the formation of vapor

bubbles in the wick structure at the evaporator section. This

formation can cause hot spots and obstruct the circulation of the

liquid. The boiling limitation can be represented by:

Qb,max -"

2(5

21-I LekeTv-rn

(? pv

and represented by Figure 4.4.3.8 of Appendix A. A boiling

limitation of 1.5 x 10 4 Btu/ft-hr is above the operating value of Q/Le

for this pipe at 464 Btu/ft-hr.

Figure 4.4.3.3 illustrates the final design specifications of the

heat pipe for the motor. In designing the heat pipes to convey waste

heat from the fuel cell power supply, approximately 1500 watts

must be dissipated at the normal operating point for the supply.

Like the heat dissipation scheme for the motor, the excess heat will

be radiated to deep space with the use of radiator panels. Two

panels identical to the one used for the motor's excess heat will be

63
used; therefore, each heat pipe which feeds a radiator must carry

750 watts of heat.

In the design of the heat pipes for the temperature control of

the power supply, an assumption has been made that the heat

generated by the power supply is localized in a region near the

center of the power supply since no actual heat sink location

information is currently accessible. Furthermore, an assumption that

a 10 inch evaporator length for the heat pipes is possible. Since the

radiator panels for the power supply will be identical to the one for

the motor, the condenser length will also be the same. The total

length of the heat pipes was determined to be 96.5 inches: 56.5

inches for the condenser section, 10 inches for the evaporator

section, and 30 inches for the adiabatic section. Using the same

series of calculations and the same material selection as for the heat

pipe for the motor, the final dimensions of the heat pipes for the

power supply were determined to be as follows:

di = 1.45 inches

do = 1.30 inches

tl = 0.15 inches

t2 = 0.065 inches

4.4.4 Heat Radiation

If we had a perfect motor there would be no excess heat

generated. Unfortunately there are no perfect motors. Our motor

design is very efficient, however we still will have up to 350 watts of

waste heat to remove from the drill assembly. The power source

generates another 1500 watts of waste heat. This gives us a total of

1850 watts or 6310 Btu/hr or waste heat.

The lack of atmosphere in the lunar environment models a

virtual vacuum. The presence of the vacuum precludes energy loss

64
from any surface of a solid by conduction or convection. Cooling is

only available in the form of thermal radiation from the surface.

One radiator will be connected to the motor and two radiators

will be connected to the power sorce by heat pipes that will run

along the length of the radiators as was described in section 4.4.3.

This is shown in figure 4.4.3.2. This radiator will be welded to the

rod rack along one side. Heat dissipation by radiation is controlled

by the equation:

q = Aeo (Ts4 - Tsur 4)

where:

q

A

IJ

£

Ts

= Heat dissipated in Btu/hr

= Area of the radiator

= .1714 x 10 -8 Btu/sq ft hr o R4

= Emissivity

= Temperature of radiator

Tsu r = Temperature of sink

The emissivity of the radiator depends on the material on the

surface. The radiator can be coated with an enamel to give it the

greatest possible emissivity and the lowest possible absorptivity.

The low absorptivity will prevent the radiator from absorbing heat.

Because of the heat pipe in our design there should not be a problem

of the radiator absorbing heat. If the radiator becomes hotter than

the motor the heat pipes will simply shut off. The heat pipes will

only work one way. Enamel coatings that are currently in use will

give the radiator a emissivity of 0.9 and an absorptivity of 0.3. (I&D)

By coating the radiator with a very smooth enamel the affects of

lunar dust will be minimized. The radiator should be highly polished

after the enamel is sprayed on.

The maximum heat rejection will occur when the motor is at its

maximum operating temperature of 212F. In order to keep the

65
radiator at a uniform temperature across its surface a material with

a high thermal conductivity should be used for the internal part of

the radiator. Aluminum is an excellent choice because it is light as

well as highly conductive. Keeping the radiator thin will also help

the radiator remain at a uniform temperature. Due to the size

constraints of the positioning of the radiator in the rod rack, the

radiator will be 1.5 inches in width.

The sink temperature will be the temperature of deep space

which is -400F. Care will have to be taken so that the radiator is

pointed toward deep space. This may be accomplished by the proper

positioning of SKITTER so that the radiator will face towards deep

space at all times during the normal drilling cycle.

Using the above equation to determine the required area of the

radiator panel yields:

6310- Area x 0.9 x 0.1714x10 -s x (7104-604 )

Area =16 ft 2

With three radiators that would mean that each radiator would need

a surface area of 5.36 ft 2- This would be the ideal radiator size.

Lunar dust and changing sink temperatures are just a few of the

places where error could occur. For this reason a factor of safety of

at least 2 should be included. This would result in an area of:

Area = 2 x 5.36 ft 2 = 772 square inches

The radiators will be able to radiate off of both sides. This

gives us another factor of safety of two, but introduces additional

error. As shown in figure 4.4.4.1 the radiator is 50.45 inches in

height and 15.3 inches in length. The volume of each aluminum

radiator would be:

Volume = length x width x height

66
V = ( 50.45 )( 15.3 )( 1.5 ) = 1157.8 inch 3

The weight of each aluminum radiator can then be determined as:

mass = density x volume

m=pxv

m = ( 0.098 lb/inch 3 ) x ( 1157.8 inch3 )

m = 113.5 lb

The height along which the heat will be applied to the radiator is 3.7

times the length. Therefore, the above calculations assumed an

isothermal temperature gradient along the radiator as is shown in

Figure 4.4.4.1.

4.4.4.1 Alternate Equation in SI Units

In si units the value of cr is 4.876x10 -8 kcal/sq m hr o K 4. 346

watts is equal to 297.8 kcal/hr Changing our values to si units we

get the following equation:

297.8 = A*.9*4.876x10-8*(3944-334)

Solving this equation we find the area to be .28 square meters.

With the factor of safety of two this gives us a radiator with an area

of .56 square meters. The radiator is 1.44 meters wide. This means

that the radiator is .32 meters long.

.32 meters is approximately 12.6 inches the error in this term

is due to very inaccurate conversion factors. This section is to help

readers more familiar with si than English units. The design should

be based on the English calculations.

67
4.4.5. Friction Clutch

In the event of a bit lock-up, a shut-off safety device was

required to prevent damage to the drill motor, the supporting

structure, and possibly SKITTER. This torque limiter can be used in

this case and should be set at a shut-off torque which is high enough

to allow for effective drilling, yet low enough to protect the drilling

device. This will act as a backup to the control shut-off which senses

a maximum current value as an indicator for a bit lock-up.

A disc friction clutch is an effective tool in torque overload

protection. This clutch will transmit torque up to a predescribed

value, after which it will "run free". It resets itself automatically

once the overload is removed and speed reduced. Friction forces and

a normal force provided by a spring are responsible for the

transmission of power. See Figure 4.4.5.1 for a schematic of a

standard friction clutch. The general formula governing the size of

the clutch is as follows:

T = 0.5 (i) (f) (Fa) (Din)

where:

T = transmitted torque

i = number of pairs of contact surfaces

f = applicable coefficient of friction

Fa = axial engaging force

Dm = mean diameter

Higher torques mean larger diameters unless more than one pair of

contact surfaces is employed. For this reason, multiple plate friction

clutches are popular. Since there are torque limiters on the market

today which suit our needs, design for such a device is not necessary.

The Helland R&D Company offers a wide array of friction clutches

known as Torq-Tenders. The model compatible with our needs is the

68
TT4X Standard Torq-Tender. The torque available is 1000 in-lbs

(83.3 ft-lbs) which will well protect our system at an assumed

overload torque of 960 in-lbs (80 ft-lbs). See Figure 4.4.5.3 for

specifications. The bore size chosen was 1.75". Although

incompatible with our drill string diameter, it was the largest

diameter offered by the TT4X. To correct this, a variable diameter

shaft interface was designed to accommodate the torque limiter and

the upper-most end of the drill string, see Figure 4.4.5.2 and

calculations.

The advantages of such a system are many. In addition to

protecting our Lunar Deep Drill Apparatus, it requires no additional

power, generates no heat, and is unobstructive. The Torq-Tender

also makes up for improper alignment. It accepts up to 1.5 degrees

misalignment and up to .015" of parallel misalignment. The system

requires no maintenance and resets itself automatically. A torque

limiter such as this makes for a good addition to our LDDA.

69
onstraints

Power Requirement

Available Motors

I

Dissipation of Heat h

Motor

Temperature Management

Radiator Panels to aid

in dissipation of heat

)

Figure 4.4.a Design Process for Power, Motor, and Temperature Management
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4.5 STRUCTURE AND MECHANICAL INTERFACE DESIGN FOR

SKITTER DRILLING IMPLEMENT

4.5.1 Introduction

This section describes the design of the implement base

structure and mechanical interface of the soil drilling mechanism

which will be attached to the bottom surface of SKITTER. The

implement base structure will be referred to as either the structural

interface or platform throughout this section. This structural

interface will accommodate the rotary drill drive, rod rack and

changer arm to handle drill string rod elements, fuel cell power

source, control module, antennae, and heat dissipation devices.

4.5.2 Constraints

Certain conditions found on the moon warrant special

consideration during design to eliminate potential performance

failures. Some of the possible problem conditions are lunar dust,

radioactivity and extreme temperature ranges. A slightly smaller

problem encountered on the lunar Surface is micrometeorite

bombardment.

The moon is covered with a layer of very fine dust which acts

almost magnetic in nature due to the lack of gravity and atmosphere.

This dust tends to cause friction and lubrication problems in moving

parts and bearings. This played a large role in the mechanical

interface design which is detailed in Mechanical Interface Design and

Operation section.

7O
The moon, as with most planets lacking an atmosphere, is

subjected to intense solar winds, gamma rays and ultra-violet

radiation. This radiation can cause weakening, embrittlement and

creep in most metals without some kind of protection. Solar wind

consists of hydrogen and helium nuclei and is the dominant source of

the lunar atmosphere. Solar radiation is not the cause of the most

harmful effects. Galactic cosmic radiation contains gamma and

ultraviolet rays which are most harmful to the materials. The

protection used on the structural interface is described in Material

Selection and Treatment section.

The temperature extremes on the moon range from -130 deg. C

to 158 deg. C. This is very detrimental in the areas where different

materials are mated together. Tolerances must be closely monitored

to avoid seizing of moving parts.

The bombardment by micrometeorites on the structural

interface surfaces could cause microscopic scratches which could lead

to crack propagation over an extended period of time. The polymer

sealant used for radiation protection is also very effective in

minimizing the microscratches from the micrometeorite

bombardment.

4.5.3 Performance Criteria

There are three major performance criteria which are required

of the structure during operation: it must provide a stable work

platform; it must have mechanical simplicity to allow remote

operation without human help; and it should have commonality of

parts for standardization of repairs and compatibility of other

implement devices designed in the future.

In order to be a stable work platform, the structure was

designed to take all possible forces generated during operation

71
without significant deflections occurring which may disrupt or

interfere with the drilling process. Deflections cannot be allowed to

introduce bending in the drill string or cause vibration which may

damage attachments which are fragile such as the heat dissipation

panels or control module. The fact that it should not catastrophically

fail is quite obvious since this would abruptly end the entire

operation.

Mechanical simplicity is required to allow for the remote

operation from the earth ground control base. The easiest way to

insure simplicity is to use the smallest number of moving parts. In

the design of the mechanical interface, which is the connection

between the structural interface and SKITTER, the general idea was

to use basic mechanical cams and springs which require no

lubrication or maintenance. The few components which required

separation from the dusty environment were enclosed in a selfcontained cavity.

Since there will be other other implements designed for use by

SKITTER, a universal latch design was chosen which could be easily

used for all of these tools. It would have been short-sighted to

design the mechanical interface as a unique connection which could

not be adapted to all of the other implements.

4.5.4 Geometry

The geometry design evolved as a result of the SKITTER

configuration and the outside surface area required by the rod

holder attachments. The structure is a six-sided figure of

symmetrical shape with three of the alternating sides longer than the

other three. Figures 4.5.4.1 through 4.5.4.3 display the geometrical

dimensions and angles of the structural interface.

72
The three sides which are shorter are located under the legs of

SKITTER and are in effect "dead" space. Any objects placed in these

areas would interfere with the motion The other three sides are the

attachment points of the rod racks. These attachments required as

much length as possible which resulted in the unequal side lengths.

The height of the structure was determined by the power plant and

accompanying fuel storage cylinders which supply electricity to the

drill drive. A height of .5 meters was required mainly for the

hydrogen, oxygen and waste water storage cylinder capacity

requirements.

The top panel of the structure has a large opening in order to

allow changing of damaged equipment such as the drill drive motor

or control module. The bottom platform has more structural

members to support the equipment and to provide a point of

attachment for the drive motor. The side panels consist of cross

framing to resist the torsional stress applied by the drilling

operation. The side panels also contain vertical members to resist

the upward force of the vertical acceleration used to reject the drill

cuttings from the hole.

4.5.5 Construction and Assembly

The structure was analyzed using FEM (Finite Element Method)

to find the highest stressed member. The rod diameter which was

chosen to withstand buckling, stress and fatigue loading over the

design life of the implement, is 25.4 ram. All the members are

specified to be the same diameter to avoid a large number of

different rod extrusion diameters and sleeve configurations. A later

study to minimize individual diameters is recommended if the

weight constraint is critical enough to warrant such a diverse

number of rod diameters and sleeve configurations. The existing

mass resulting from these dimensions is 71.1 kg. which is 697.5

earth pounds and 116.3 moon pounds.

73
The members are to be extruded using a matrix of 6061

aluminum with continuous beta silicon carbide fibers. This

composite is described in the Materials Selection and Treatments

section.

The sleeve connections are to be cast in parts with the number

of parts depending on the geometry of the joint. The sleeves are to

be cast using the same material as the structural members are

extruded from. Joints joining members which are all in the same

plane only require two halves, while three or four parts will be

needed for three-dimensional joints. The joint construction will

follow the following sequence:

1)

2)

3)

4)

One part or half of the sleeve is placed down on a flat

surface

All structural members are glued into the sleeve part or

half

The other half or parts of the sleeve are glued over the

members

The sleeve halves or parts are bolted together through

flanges provided on the perimeter of the sleeve

The sleeve recess which receives the structural members is

sinusoidally machined at the edges to lessen the stress concentration

applied at the end of the sleeve. A typical sleeve arrangement is

shown in Figure 4.5.5.1. The insertion length of the member into the

sleeve is 25 mm, which is the member diameter.

4.5.6 Connections to Attachments

One of the important duties of the structural interface is to do

just that, interface. We had to be sure that the structure would not

only withstand the moon environment and rigors of drilling, but that

it could also accommodate the various apparatus for the operation.

74
This meant that we had to keep our design flexible until the final

designs of the other groups were established. Figure 4.5.6.1 shows

the different components that the platform must hold and the

general location of each. Figure 4.5.6.2 shows the corresponding

location of the fasteners and mounts for the components. The

assumption was made that the platform and its attachments make up

a self-contained, permanent apparatus. That is, there won't be a

need for removeable fasteners between members and attachments.

The power generation equipment was placed inside the

platform because it was large, awkward and sensitive. The power

design group specified the dimensions of the plant and the three

pressure vessels, as well as the brackets they will be mounted in.

Since the vessels and power plant aren't stressed laterally, these

fasteners won't need any critical design. They merely need to hold

the vessels plant in place. Because large forces are transferred

through the motor it was placed in the center of the structure so as

not to create adverse moments on the structure due to the large

forces transmitted through it. The power design group specified the

sizes and geometry of the fasteners needed to secure the motor.

Heat pipes used to transfer motor heat to the rejection panel are

small and are easily placed inside. The control box is also placed

inside the interface since it needs to be protected from unnecessary

contact with moving parts.

The rods will be placed on racks around the perimeter of the

platform. The interface will hold three racks consisting of 10 rods

each. Each rack lies on the long sides of the interface in between the

legs as seen in Figure 4.5.6.1.

The heat rejection panel is small( area = .016 sq. meters)

compared to the racks and will fit easily between any of the rod

racks. It will be attached similar to the rod racks.

According to the rotary disc design group, the disc will be

connected to a bearing which will in turn be connected to the

75
structure. The bearing geometry and fastener locations haven't been

specified. Consequently, we are designing under the assumption that

a U-bolt arrangement will be adequate.

4.5.7 Fastener Considerations

As stated above, each of the objects attached to the platform,

with the possible exception of the motor, are permanently affixed.

We have specified locations and beam sizes and the other groups are

designing their fasteners accordingly. We have informed them to use

U-bolts. These type of "wrap around" fasteners are preferred

because drilling holes into the beams will introduce stress

concentrations.

4.5.8 Description of Loads

The rotary motor produces and transfers the largest forces on

the platform. As mentioned above, the motor will be attached with

three fasteners on the 'floor' of the structural interface. The torque

and drilling force will be transferred at these fastener locations.

The motor design group used a torque of 68 N-m (50 ft.lbs.) as

a maximum for drilling into the basalt of the moon. However, the

drill bit group didn't confirm this figure and didn't specify a different

one. Accordingly, we decided to use the maximum load torque of

the motor itself, 271 N.m (200 ft.lbs). Thus, if the drill string jams

and forces the motor up to its limiting torque, the structure will

remain intact. Dividing this torque by the effective radius of the

fasteners (.23 meters), we obtain a force of 395 N (89 lbs.) acting

tangent to lines intersecting the fasteners and the center of the

motor.

76
The vertical drilling force will also be transferred through the

motor connections. Estimates by the drill-bit design group indicated

a 890 N (200 lbs) force was necessary to successfully drill into the

moon basalt. For our analysis, an assumed drill force of 300 N (75

lbs) will act vertically at each of the three motor fasteners.

The rod holders and associated rods create flexural bending at

the sides of the platform. Figure 4.5.8.1 shows how the shear forces

will subject the top and bottom members to flexural bending. The

maximum weight of a rod holder set is 22.1 lbs (98.1 Newtons = 10

rods/set X 9.81 Newtons/rod). The rod holders and the length of the

rods constitute the moment arms. The rod holder design group

specified the geometry shown in Figure 4.5.8.1. The rod holders

extend the rods' center of mass approximately 2 meters. From these

values and the geometry of the figure, we calculated the reaction

magnitudes at the structure interface/brace attachments to be 214 N

(48 lbs.) for the lower fastener and 200 N (45 lbs) for the upper

fastener for each brace.

The motor design group also needs to house a mini-power plant

inside the structure. This plant will consist of a main generating

facility, three pressure vessels (one fuel tank, one oxygen tank and

one waste water tank). The main gener_ting facility weighs 149 N,

(33 lbs), and is located along one of the structural interface's three

longer sides, as shown Figure 4.5.6.1. A resulting moment of 80.5

Nom is produced on the structure. The total weight of the three

pressure vessels is approximately 65 N, (14.5 lbs), and they will be

located along the remaining two long sides. The vessels produce a

resulting moment of 15.3 N°m. The downward forces will be

transmitted through the fasteners described in the previous section.

The vertical accelerator will create a vibratory force translated

by the drill string, through the motor, and into the platform at the

motor connections. The vertical accelerator design group indicated

that maximum force of 533 N (120 lbs) will be generated as the rock

is accelerated up the auger. Since this force occurs at a different

77
time and in a different direction than the drilling load, we created a

separate analysis for it. Thus, we have two cases for the stress

analysis.

The robot arm, rotary disc, motor weight, heat rejection panel

and control box all amount to very little in the moon environment.

Together, they constitute less than 5% of the drill, accelerator, and

rod loads. Consequently, we neglected them in our analysis.

4.5.9 Structural Analysis

The integrity of the structures' design was evaluated by

utilizing SDRC I-DEAS finite element software package. Preparation

of the analysis included setting parameters, defining the boundary

conditions, loads, and making some assumptions. Parameters for the

physical and material properties were derived in the Material

Selection and Treatment section. Two load cases where analyzed due

to the different loading conditions.

The boundary conditions of the structure consisted of the

latch/structure connection. It was assumed that when the structure

is connected to Skitter, the latch positions would remain fixed.

Therefore as the loads were applied to the structure, reaction forces

occurred at these locations that could be used to design the

mechanical interface.

4.5.9.1 Case One

The first case included the motor, rod and power plant loads.

From this analysis, the maximum principle tensile and compressive

stresses were determined, 545 kPa and 597 kPa respectively. These

values where then compared to the corresponding maximum stress

values for Silicon Carbide Aluminum, 458 MPa and 49.25 MPa. The

78
dramatic difference indicated that failure due

compressive yielding was highly unlikely.

to tensile or

Of equal concern was the possibility of failure due to buckling.

A critical slenderness ratio of 9.1l was determined in order to

confirm the consideration of buckling over yielding. When compared

to the slenderness ratio for Silicon Carbide Aluminum, 193.7, it was

determined that if failure would occur it would indeed occur by

buckling, and the Euler Column formula was used to determine the

critical load. A value of 27,591 N was calculated for the critical load

using the following equation:

pA2E

Pcr = A(1/k)^ 2

Pcr - 27,591 N

where A = 5.067e-4

! = 1.23 m

k= d/4 = .0254 m/4

E= 207e9 Pa

This value was then compared to the maximum load of 276 N,

which was determined from the analysis as occurring on a member

of length 1.23 m. Here again, by a considerable difference, the

analysis indicates that failure due to buckling is improbable.

Another concern is the possibility of fast fracture. At the

maximum tensile stress of 545 kPa, the critical crack length was

calculated using the following equation:

Kc = s('_pa) where Kc = 91 MN/mA3/2

a = critical crack

length

solving for a, the number is on the order of a hundred meters. This

indicates that fast fracture is not going to be a factor at this stress

level.

79
Another consideration is the strength of the adhesive used to

join the structures' members. Evaluating the maximum shear stress

equation below:

T, max = _- x coth _-

h = (G) (1^2) / ((E) (tl) (t2))

tl = adherend thickness

t3 = glue thickness

It was found that the shear strength of the adhesive is 5.77

MPa. From the FEM analysis, the maximum shear stress component

was determined to be .53 MPa. This indicates that the weak link in

the overall design is at the joints where a factor of safety was

determined to be approximately l l.

4.5.9.2 Case Two

A second load case was run in order to analyze the vertical

accelerator operation. The second case consisted of the loads

resulting from the vertical accelerator, rods and the power plant.

Because of the vertical accelerators' cyclic operation, fatigue loading

was a prime concern. In order to determine whether the existing

fatigue was either high or low cycle fatigue, the maximum stress for

the cyclic loading was calculated. This value was found to be 1.05

MPa. When compared to the yield strength of the aluminum matrix,

169 MPa, it was obvious the loading is in the high cycle fatigue

regime. Therefore, any scratches, nicks or dents introduced onto the

structure can reduce the fatigue life by orders of magnitude. In

order to determine the number of cycles to failure at this stress

level, Basquin's law will be evaluated.

In the steady state regime the crack growth rate is described by:

80
da

dN= A(AK)m where AK is the cyclic stress intensity

A and m are material constants

Solving for the number of cycles to failure, Nf :

1

Nf=ArAs_^4p^2(1/a0_, ,, 1/ac)

Since ac is much larger than a0, it will be ignored. Performing the

calculation with A = 4 e-13 and m = 4, the number of cycles to failure

are 18e12.

Considering the maximum number of cycles in one day, the

number of days in a year and the service life of the structure, the

number of cycles expected is approximately 100,000. Compared to

the number of cycle s to failure, the material has an infinite life at

this stress level.

4.5.10 Material Selection and Treatments

The space environment is such a hostile work space that

demands on materials are rigorous. Ordinary structural materials

such as steels and irons are inadequate when subjected to the harsh

conditions in deep space. Materials used in this environment must

be designed according to the following criteria:

1. High strength to density and stiffness to density ratios.

2. Long useful life at the operating temperatures.

3. High creep resistance.

4. Resistance to thermal and mechanical fatigue.

5. Low thermal expansion coefficients.

6. High fracture toughness.

7. Resists outgassing.

81
Although this list appears to exclude all known materials, potential

does exist for the following space-age materials"

1. Aluminum alloys

2. Titanium alloys

3. Aluminum-Lithium alloys

4. Metal matrix composites

Aluminum and titanium were considered because of the extensive

use of these alloys in previous space applications. AluminumLithium alloys and metal matrix composites are new materials that

are specifically designed for space applications. These promising

materials are on the forefront of the space age materials technology.

In order to choose the proper material to build the structure, each

candidate's mechanical properties were evaluated and based on the

seven criteria and a suitable material was chosen. In addition to the

mechanical properties, another important factor was evaluated,

fatigue strength. Since the structure will be subjected to cyclic

stresses, and since numerous scratches will be inflicted onto the

structure, the ability to withstand fatigue loading is paramount. The

following sections present an analysis of each candidate.

4.5.10.1 Aluminum Alloys

Aluminum alloys have been used in space applications for the

past 30 years. By possessing a low density, approximately 2.7

Mg/m^3, and a high modulus, 71 Gpa, these alloys are very attractive

materials. However, because most alloy elements have a very low

solubility in aluminum they ,'ire segregated to the dendrite cell

boundaries. The dendritic structure, side branches growing from

crystals in the metal, usually are formed from second phase iron

particles. The iron particles are large and add only small increments

to strength, however they do reduce the fatigue strength of the alloy.

The fatigue strength can be improved with certain alloys, however

aluminum alloys with good fatigue properties typically have poor

82
thermal or cryogenic properties. Since the temperature on the moon

will cycle between -130 degrees C. and 160 degrees C., the structure

will be subjected to thermal as well as mechanical cycles. Over the

period of ten years, it must be able to resist fatigue failure.

4.5.10.2 Titanium Alloys

Titanium alloys, unlike their aluminum counterparts, have good

fatigue properties. The alloy Ti-6AI-4V exhibits the best all around

characteristics. This material is especially promising at very low

temperatures. Since the temperature on the moon will drop to near -

130 degrees C, a material must withstand the rigors of the extreme

cold. The unnotched strength, fatigue crack propagation rate and

fatigue strength are surprisingly maintained at temperatures as low

as -253 degrees C. However when notched specimens were tested the

results indicated that the fatigue strength decreased dramatically in

this alloy.

Another problem with titanium alloys is their density.

Titanium alloys usually have a density of 4.5 Mg/m^3 or a density

that is twice that of most aluminum alloys. With the cost of space

flight proportional to the weight of the payload, any savings in

weight will reduce the cost of the planned mission.

4.5.10.3 Aluminum Lithium Alloys

In recent years, a number of aerospace metals firms have been

placing a great deal of research into tile area of aluminum alloyed

with lithium. A major reason is because each weight percent of

lithium, up to 4 wt.%, added to aluminum results in a reduction of

density of 3 percent and an increase in elastic modulus of 6 percent.

The potential for weight savings and an increase in stiffness have

been the primary reasons for the heightened expectations of this

alloy.

83
Although this alloy has received a lot of attention one major

drawback is that, it does not possess good structural properties. In

particular, the fatigue properties of aluminum-lithium are not

promising. This is a result of the impurities contained in lithium

which tend to be cites of microcrack initiation. Current research at

the Alcoa Research Center is centered on the removal of the

impurities, however a time frame for a high fracture tough alloy has

not been set.

4.5.10.4 Metal Matrix Composites

Materials composed of metals reinforced with fibers have

demonstrated unparalleled mechanical properties. When compared

to metals, metal matrix composites have the following advantages:

1. Major weight savings due to higher specific mechanical

properties

2. Dimensional stability for space structures

3. Strength retention at higher operating temperatures

4. Improved fatigue characteristics

With respect to organic matrix composites, MMC offer;

1. Higher operating temperatures

2. Does not outgas or degrade in space

3. Radiation survivability, laser, nuclear, UV, etc.

4. Improved joining characteristics

5. Better transverse properties

The main reason for the introduction of fibers into the metal

instead of traditional work hardening methods is that precipitation

or dispersion hardening result in dramatic increases in the yield

strength with little influence in the modulus. However, when fibers

84
are placed into the metal-matrix they increase the stiffness and

because of there light weight an increased stiffness to density ratio is

achieved. A high stiffness to density ratio is sought for structural

purposes involving compression or flexural loading, which are two of

the prime loading mechanisms expected in the structure. The

following mathematical analysis demonstrates the effect of a high

stiffness to density ratio on material selection.

Consider a round bar loaded in axial tension:

1 = length d = diameter P = applied force

p = density m = mass A = area

total deformation: d = Fl/AE

m = Volume x p = (p)(d^2)(I)(p)/4

A = (p)(d)(1)

m = (A)(d)(p)/4

A = 4(m)/(dp)

d = Fldp/(4mE)

therefore mass = Fldp/(4dE)

= F/d x ld/4 x p/E

From the final equation it can be seen that for a given rigidity

F/del a minimum mass is achieved for maximum E/p (stiffness to

density ratio). In other words, it makes sense to use high modulus

fibers that are lighter that the metallic matrix for reinforcement in

structural applications.

Although it has been established that metal matrix composites

are the materials of choice for structural applications, an important

question arises, which MMC should be selected? From research and

conversations with various vendors aluminum 6061 reinforced with

47 percent continuous beta silicon carbide fibers will be used to

design the structure. In order to have made a selection, a number of

areas were examined: These include micomechanical properties such

as transverse stresses between fibers and matrix, composite thermal

85
expansion coefficient and macronlechanical properties such

strength, toughness and fatigue.

as

4.5.10.5 Micromechanical Properties

When a metallic composite is used with fibers a problem that

can arise is that of poisson's ratio mismatch. This imbalance is

caused by having a metal with an extremely different modulus

compared to the fibers. An imbalance of this sort can have a

dramatic impact on the overall properties of the composite.

Specifically, the transverse stresses and the pressure that results at

the fiber-matrix interface can cause enormous residual stresses in

the composite. The residual stresses can cause the ductile metal to

deform plastically near the interface which results in a weaker

composite. The following mathematical analysis demonstrates how

poisson's ratio can influence the composite.

HOOKES LAW: (cylindrical coordinates, see Figure 4.5.10.1)

ler 00I Isr 0 OI

I 0 eqO I =(l+n)/E x I0 sq OI -n x

I0 0 ezl I0 0 szl

(sr+sq+sz)/E

l1 00I

x I0 1 OI

10 0 11

equilibrium condition: d(sr)/dr = (sr:sq)/r = 0

plane strain condition in terms of displacements:

er = d(ur)/dr eq = ur/r

ur = radial displacement ez = const

so/k= (l-n) x eq + n x (er+ez)

sr/k = (l-n) x er + v x (eo+ez)

k = E/((l+n) x (1-2n)

Substituting the plane strain equations into the above two equations:

86
so/k = n x d(ur)/dr + (l-n) x ur/r + n x ez

sr/k = (l-n) x d(ur)/dr + n x ur/r + n x ez

Substituting the above two equations into the equilibrium condition:

d^2(ur)/dr^2 + (l/r) * d(ur)/dr ur/r^2 = 0

The solution to the differential equation takes the form of:

ur = Cr + B/r

The constants can be found by applying known boundary conditions

at the interface.

By denoting the matrix as component 2 and the fibers as 1 and

letting sr,2 = -p. An expression for the radial pressure at the

interface can be derived.

p = (n,2 - n,1) x 2(ez) x V,2 / (V,1/kp,2 + V,2/kp,1 +1/G2+,2)

V = volume fraction

kp = E/((2(l+v) x (1-2v)

The stress components at the interface are:

Fibers: sr,1 = qo,1 = -p

sz,1 =E,1 xez- 2 x (n,1) x p

Matrix: sr,2 = function of p

sq,2 = function of p

sz,2 = E,2 x ez + function of p

From the above analysis it can be seen that when the poisson's

ratio difference (n2 nl) goes to zero the radial pressure goes to zero

and takes with it most of the stress components.

87
Of the metal matrix composites examined, the aluminum

composite that was chosen possesses a very low poissons ratio

difference. Aluminum 6061 has a n of .36 and the fibers have a

ratio of .25, the difference being .11.

As discussed in the previous section the dissimilar poissons

ratios can be attributed to the different modulus of elasticity for each

component. In the case of Silicon Carbide Aluminum the moduli are

of the same order of magnitude, which indicates why the poisson's

ratios are similar. In order to calculate the modulus of the

composite, the theory of elasticity should be used instead of the rule

of mixtures. Generally, the often used rule of mixtures will only give

an order of magnitude approximation of the desired value. The only

draw back with there use is that the should be solved numerically or

with the aid of complex variable techniques. For equally precise

calculations without the complicated theory of elasticity equations,

K.K. Chawla in COMPOSITE MATERIALS suggest using the the Halpin -

Tsai Equations. These equations take the form of:

p/pm = (1 + xnVf) / (1-nVf)

n = (pf/pm 1) / (pf/pm + x)

Values for x:

E longitudinal 2(l/d)

E transverse .5

G longitudinal 1.0

G transverse .5

K 0

Where p and pm are the desired quantities and the matrix value

respectively.

Using the above values the composite modulus in the

longitudinal direction is 207 Gpa and in tile transverse direction it is

159 Mpa.

88
Another micromechanical area is the thermal expansion

coefficient. Just like a large difference in poisson's ratio can create

residual stresses so too can dissimilar a's. When a choice is made to

combine a metal and its reinforcing fiber careful thought is used so

as to preserve a thermodynamic equilibrium when the composite is

formed.

From the mechanical data available, the aluminum composite

discused is comprised of components with similar a's. Aluminum

6061 has an a of 20 exp(-6)/K. The fibers have an a of 5.7 exp(-6)/K.

Because the matrix and fibers have am order of magnitude

similarity, the residual stresses in Silicon Carbide aluminum will be

minimized.

An often used method of calculating the thermal expansion

coefficient of the composite is the rather simple rule of mixtures

formula. However, the use of the formula will overestimate the

expansion because it does not take into account the restriction that

the stiff fibers impose on the matrix. The fibers by restricting the

movement of the matrix produce a composite that has a cte much

less than a weighted average formula can predict. If the assumption

is that the poisson's ratio differences are slight, as was shown earlier,

then R.A. Schapery has demonstrated in the book COMPOSITE

MATTER the validity of the following formulas:

a longitudinal:

a transverse:

VfEfaf + Vm Em am/(EfVf+Em Vm)

af (Vf)^.5 + am(1-(Vf)^.5)( 1 + K)

K = VfnmEf/(EfVf+ Em Vm)

using the following values:

Vf = .4 Vm = .6 Ef = 430 Gpa

Em =71 Gpa am = 20exp (-6)/K

af = 5.7 exp (-6) / K

nm = .36

89
The thermal expansion coefficient of the composite in the

longitudinal direction is 7.94 exp(-6) / K. The thermal expansion

coefficient in the transverse direction is 12.1 exp(-6)/K. Notice that

the composite will tend to expand more in the tr_znsverse direction.

This is caused by the continuous fibers restraining tile matrix more

in their principal direction, longitudinal.

4.5.10.6 Macromechanical Properties

The macromechanical properties that will be of most concern

are, compressive and tensile strengths, toughness and fatigue. The

compressive strength of a unidirectional composite cannot not be

predicted by examining both the compressive strength of the matrix

and then the fiber because the main failure mode in compressive

loading is buckling of columns. More specifically, fiber composites

fail in compression by periodic buckling of the fibers. In order to

determine the compressive strength of the composite, two buckling

modes were examined. The first, in phase buckling, is characterized

by the fibers buckling with the same alignments. In such a case the

composite strength in compression is proportional to the matrix

shear modulus, Gin. Whereby:

s,c = Gm/Vm

For an isotropic matrix, of which aluminum 6061 is;

Gm = Em/(2)(l+n,m)

Therefore;

s,c = Em/(2)(l+n,m)(Vm)

= 71(Gpa) / (2)(1+.36)(.53)

= 49.25 Gpa

9O
The other failure mode is out of phase buckling. This buckling

mode involves transverse compressive and tensile strains. Here the

compressive strength is given by the geometric mean of the fiber

and matrix Young moduli:

s,c = 2/3 (Vf) ((Vf)( Em)(Ef) / (Vm) ^ .5 )

= 51.56 Gpa

The particular mode of buckling is primarily dependant upon the

volume fraction of fibers. In phase buckling is noticed at higher

volume fractions because the fibers exert more influence on each

other and tend to couple in phase. Since Silicon Carbide Aluminum

uses 47 percent fibers, the upper bound on compressive strength is

49.25 Gpa.

The tensile strength can be similarly calculated with the

following results:

Tensile Strength (0 degrees) 1458 MPa

(90 degrees) 86 MPa

High fracture toughness is one of the most attractive

characteristics of composites. Because of the crack blunting effect of

the fibers, these materials can resist crack propagation both statically

and dynamically. In the particular case of Silicon Carbide Aluminum,

the continuous fibers act as roadblocks to any crack that is

propagating in the longitudinal direction. The fracture toughness of

Silicon Carbide Aluminum is compared with Silicon Carbide particles

and unreinforced aluminum in the following table:

MATERIAL

NOTCHED TENSILE STRENGTH

% OF UNOTCHED STRENGTH

APPARENT K

(KSI SQRT(in))

SCS2/AL 100 7 3

FP/AL 6 5 3 6

SiCp/AL 6 5 2 4 - 2 8

91
As illustrated in the table, Silicon Carbide Aluminum has a far

superior fracture toughness compared to unreinforced aluminum and

Silicon Carbide Aluminum with whiskers, a rival material.

Fatigue is an area of composites that has received a great deal

of attention. Fatigue failure mechanisms in composites are quite

different than in metals. Fatigue failure in metals occurs as a result

of initiation and growth of a crack. Fiber composites, on the other

hand, can withstand a variety of subcritical damage (matrix cracking,

fiber fracture, fiber/matrix decohesion, and ply cracking). The

biggest difference in fatigue failures is that in metals the critical

factor is crack length while in composites it is crack density.

The various forms of damage mentioned above result in a

reduction in the load carrying ability, which in turn causes a

reduction in stiffness and strength. Fatigue, in composites, is

characterized by two distinct stages. In the first stage, noninteractive cracks appear in isolated piles. In the second stage, the

damage becomes localized in areas of crack interaction.

Just as the micro and macromechanical properties of Silicon

Carbide Aluminum were evaluated, so were the elevated and

cryogenic temperature properties. As mentioned in the opening of

this section, the lunar environment is severe. In fact, the lunar

temperature will cycle between 160 degrees C and -130 degrees C

within one lunar day. Typical mechanical properties evaluated at

room temperature may become meaningless at the range of

temperatures expected on the moon.

Figure 4.5.10.2 illustrates that the strength of Silicon Carbide

Aluminum, in the fiber direction, is far superior to other reinforced

aluminum composites or even unreinforced aluminum 6061 at

elevated temperatures. Figure 4.5.10.3 is an another example of the

excellent high temperature capabilities of this composite. From the

figure, it can be determined that the composite retains it stiffness up

to approximately 300 degrees C.

92
Another area of prime concern with the use of materials on the

moon is thermal fatigue. In metal matrix composites, thermal fatigue

can have devastating effects on the material. More specifically, if the

fiber and matrix have vastly different thermal expansion

coefficients, the interface can be weakened causing the fibers to

pullout of the matrix. The end result can be catastrophic failure of

the structure. However, as discussed earlier, aluminum and the

silicon carbide fibers have similar cte's which results in a composite

that is not easily susceptible to thermal fatigue.

Some of the cryogenic properties of Silicon Carbide Aluminum

are illustrated in Figure 4.5.10.4. The plots of composite strength

versus temperature indicate that the material is not expected to

loose any strength or modulus at subzero temperatures. Although

the data was obtained down to -50 degrees C, these values should

not change radically at the expected moon low temperatures. In fact

the only problem that might arise is with the fatigue properties of

the composite. Since the matrix of aluminum alloy 6061 becomes

less ductile at these temperatures, the possibility of matrix failure

could arise. Examining Figure 4.5.10.5, the fatigue properties are

evaluated at high and low temperatures. From this data, it can be

determined that the fatigue data corresponds favorably at these

temperatures.

4.5.10.7 Treatment

An area that has not received any attention as of yet is the

performance of the material when exposed to high levels of radiation

for an extended period of time. Since the moon does not have an

atmosphere that can filter the immense amount of radiation flux, any

material used on the moon must withstand radiation from all

sources. These include solar wind, solar flares, cosmic rays and

ultraviolet radiation. Silicon Carbide Aluminum has a superior

resistance to radiation degradation, however after a period of time

93
the radiation bombardment can attack the fiber-matrix interface

causing a catastrophic failure. In order to prolong the life of the

structure, a number of alternatives have been examined. One of

these is to coat the entire structure with a radiation absorbing

polymer. This polymer would act as a shield to protect the composite.

One serious flaw with this approach is that the polymer would have

to a near identical thermal expansion coefficient as the composite.

If the polymer did not expand and contract as the composite does,

the coating would loose its adhesiveness to the composite and peel

off.

The other approach considered is of applying a sealant to the

composite. The sealant would be similar to sealants used on wood to

prevent water damage. By using a sealant the thermal expansion

coefficients would not play as significant of a role because the

composite would tend to allow the sealant to penetrate the first few

microscopic layers. The sealant that has been used quite extensively

is polydimethylsiloxane (PDMS). This thermoset polymer has as its

backbone the high strength Silicon-Oxygen combination and methyl

groups (CH3) as the side groups. This sealant is characterized by the

following;

1. Glass transition temperature of 150 K.

2. Maximum operating temperature of 250 degrees C.

3. Thermal stability due to high bond energy of the

backbone.

4. Good adhesive bonding with composites in the thermal

cycling environment.

From the list it is evident that PDMS possesses excellent lunar

properties. One particular advantage is that the glass transition

temperature is so low that it will retain its' elastomeric properties at

very low temperatures. In order to make the sealant provide UV as

well as other forms of radiation protection, a UV absorber will be

added to PDMS. The UV absorber, 2-hydroxybenzophene, will make

94
the sealant opaque thereby

penetrating into the composite.

preventing the radiation from

One additional problem must be addressed at this time. The

point has already been made that the composite does not outgass,

however PDMS does. In order to prevent the outgassing of the

polymer research has shown that the following measures can be

taken:

1. Curing the polymer at temperature well above the conventional

curing levels.

2. Vacuum purification of the polymer.

3. Solvent extraction.

4.5.10.8 Adhesive Analysis

In order to build a rigid structure, the individual members

must be connected in some permanent manner. Bolts and rivets can

be used with composites, but it is often necessary to have load

spreading inserts bonded into the structure, in order to reduce the

stress concentrations. One particularly attractive bonding method is

the use of adhesives. Adhesive bonding allows for a more gentle

diffusion of the load into the structure, thus reducing the localized

stresses encountered with the use of rivets and bolts.

The stress analysis for an adhesive bond is very complex. In fact,

most of the methods used to determine the exact stress and strains

on a joint are solved by numerical methods. Although this may

make the calculations too complex, the tubular lap joint stresses and

strains can be approximated by the ordinary double lap joint.

Utilizing the double lap joint, the resulting stresses will be rough

approximations to the actual values.

Since the adhesive joint will be subjected to high temperatures,

an elasto-plastic analysis will be used to determine the stresses. The

analysis was completed in detail in Structural Adhesive Joints in

95
Engineering by Robert D. Adams.

condensed version of his analysis.

The following section is a

Adhesive joints are typically subjected to three forms of

stresses; shear stresses, bending moments, and tearing stresses. The

analysis has shown that the maximum stress is a shear stress and

occurs at the end of the joint. The stress is approximated by the

following equation:

T, max = sqrt(h/2) x coth(sqrt(h/2))

h = (G) (1^2) / ((E) (tl) (t2))

tl = adherend thickness

t3 = glue thickness

The tearing stresses or bond peel stresses cannot be ignored

because their magnitude is less than the shear stress. The bond peel

stress is much higher at the end of the joint and decreases to a

minimum near the midsection of the joint. These stresses are caused

by a high stress concentration at the edge of the adhesive joint. In

order to reduce the stress concentration and to provide for a

smoother transition of stresses into the otherwise unstressed regions

of the joint-adherend interface, the overlap will be machined into a

sinusoidal pattern to a distance of roughly one half of the overlap

length. This pattern will enable a greater amount of shear energy to

be transferred uniformly between the adherend and the adhesive.

In order to determine the glue thickness and the overlap length, a

set of graphs for adhesives are employed. These graphs indicate for

a certain glue thickness the optimum strength and for a overlap

length the optimum strength. However, for a first order

approximation the overlap length can be set to the diameter of the

adherend. The glue thickness can be found from the graph of a

typical adhesive to be approximately .15 mm.

The choice for an adhesive is similar to the choice for a material, it

must be capable of withstanding the rigors of outer space. In the

previous section it was mentioned that siloxanes are well suited to

96
the lunar environment and after much research, it was discovered

that a form of PDMS can also be used as a space adhesive. This

material, methacrylate-siloxane with 25 percent silica reinforcing

filler(Whacker HDK 2000), possesses excellent physical properties.

The properties were evaluated at both high and low temperatures

and are summarized in the table below.

Tensile Strength

Modulus

Tear Strength

PROPERTIES

2.86 MPa

.48 MPa

8.9 KPa

4.5.11 Mechanical Interface Design and Operation

The mechanical interface refers to the latch system which connects

the drill implement to SKITTER. A very important aspect of the

design was to conceive a connection arrangement which has the

following properties:

1)

2)

3)

4)

5)

Simplicity

Self-aligning movement

Self-closing operation

Small number of moving parts

Dependability

The latch design was kept simple in order to insure remote

operation and robot control. Only one command is needed from

ground control during the entire drilling process. This command is

sent when the drilling task is complete and the implement is to be

detached from SKITTER.

The closing of the latch was designed to be completed without

any mechanical actuation required other than the force of SKITTER

97
lowering itself onto the implement. An involute cup with flat sides

with an attachment rod in the center is the target of the latch claw.

Both the claw and receiving cup are made with involute profiles. An

involute curve is generated by tracing a point on a cord as the cord is

unwrapped from a cylinder. This allows a smooth closing action

between the mating surfaces. As the claw enters the cup, it is forced

to close because the involute shape gets narrower towards the

bottom and forces the two claw halves towards each other. The

involute surfaces cause the claw to close until it reaches a point

where the vertex of the claw hits the attachment rod. At this point, a

lateral force is required to fully push the claws together. This is

accomplished by placing one cam on each of the claw halves at a

point where the attachment rod will contact the cams and create a

moment about the claw pivot pin. This moment completely closes

the claw and allows a spring-loaded lock pin contained on one half of

the claw to extend into a mating hole on the other half of the claw

which will keep the claw locked until implement detachment is

desired. When drilling is complete, an electrical actuator is activated

which pulls the lock pin out of the mating hole. The two halves of

the claw are automatically pulled apart by two extension springs

which cause the normal position of the claw to be open. Once the

claws are open, SKITTER raises off of the implement. The latch

mechanism remains open until SKITTER lowers onto another

implement which operates on the same attachment basis. Figure

4.5.11.1 shows a typical latch with each part labeled.

The number of moving parts was kept to three. The two claw

halves and the lock pin are the only moving parts, except for three

springs. The lock pin-actuator mechanism can be separated from the

environment by a dust cover to prevent interference. This design

should insure a dependable mechanism which is not dependent on

automated control.

98
Io

TYR SLEEVE DETAIL

...... ii iii i ........

NOT TO SCALE

FIGURE 4-.5.5. t
General Layout

ROD

ROD CUP DEVICE

ROD RACK AND BRACKET

ACTUALLY HOLDS 10 RODS

(ONE ROD RACK ON

EACH LONG SIDE)

POWER PLANT

SKITTER'S LEG AREA

MOTOR

control box

PRESSURE VESSELS

Figure 4.5.6. I
r.

i__

_7

I--

ll.i.l

_.J

._I

_D

I_L m_llmI t.--i

Z

__ o,

_E
" Shear Induced

_ Due To

By

Rod

Flexural

Weight

Bending

shear in Skitter's members

9 degress

Rod Rack

22.1 degrees

98.1 Newtons

(Weight of 10 Rods)

Figure 4.5.8.1
Fiber/Matrix Interface

I

!

t

I

t

!

FIBER

i

!

E, V Gt,

!

I

I

I

I

J

,_ ";

....x

MATRIX

E,v,(*

Figure 4.5.10.1
'.4"1

['J

u.._
LLILU

y_0

J

,¢o

GO /I /

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iOF pOOR QUALITY
4.6 ROD CHANGER AND STORAGE RACKS

4.6.1 Gripping Mechanism

Since the end effector is responsible for the handling of the

rods and other drilling implements, it's design and employment is

vital to the success of the robotic system. These are the

characteristics needed to define a gripping system"

1. Characteristics of the arm and especially the maximum

transport load capacity of the arm.

2. Characteristics of the object, including

• weight

• range of specifications

• nature of material (roughness)

• geometry, dimensions and structure

• contact forces

• forces acting on object

• environmental conditions

3. Maintenance

The gripping mechanism for the robot arm will be a bilateral

gripper. This type of gripper was chosen over a unilateral design

because of the gripping needed to grasp a rod. A magnetic unilateral

gripper would not function as well as the bilateral. Unilateral

grippers use only one surface to grip while bilateral grippers have

two finger-like jaws. The multilateral gripper was determined to be

too complex and unnecessary for our application. With our bilateral

gripper, the degrees of freedom of the arm are reduced to three.

This keeps the overall design of the arm simple and easier to

maintain.

Considering various powering means, electric power use is light

in weight and simple in maintenance for robotic arms. Electric power

99
is simpler to maintain versus an oil pressure system because it does

not use any fluids that have to be sealed. All plastic or rubber seals

would have to be covered to prevent them from being damaged by

radiation. While the electrical method is inferior to the oil pressure

method in output power, the electrical method is more suited to our

design needs. Reliability in terms of maintenance is more important

in our application versus output power. A large amount of output

power is not needed since the heaviest object to lift is 16.67 lbs. on

the moon. The fact that electric power is light in weight and

simplistic was the determining factors in choosing it for gripper

power requirements.

The gripper will be made of aluminum. For discussion on the

choice of aluminum, please refer to the materials section, Section

4.6.2. The gripper has five main components. These include the

motor, leadscrew, leadscrew nut, main rectangular bar, pinned bar,

and the jaw part of the main rectangular bar. Please see Drawing

4.6.1.1 for the gripper design.

The jaw section of the gripper mechanism will be 81 mm tall.

This will enable the jaws to grasp the 100 mm auger over two helix

sections because of the double helix design, offering flexibility in

grasping the auger. The gripper can grasp the auger on any part of

its length. The main rectangular bar will connect to the jaw section

at the middle of its 81 rnm height. This will reduce the moment the

object causes on the connection between the gripper and the robotic

arm. The jaw section is 170 mm of the main rectangular bar length.

The gripper has to be designed to grasp both 100 mm augers

and 50 mm rods. A V-type gripper jaw was chosen for its ability to

accommodate both the 100 and 50 mm diameter digging implements.

This design will enable the jaws to grip both size objects on the

centerline of the gripper, considered to be a key design parameter.

Therefore, the robot arm will only have to be programmed to go to

one position for each rod storage and retrieval location.

100
The gripper design of a screw-driver has been chosen. Screwdriven grippers are operated by turning the screw uniformly

through a motor. Our design is operated by turning the screw in a

right handed direction to force the threaded block to shift to the left,

thereby tightening the jaws through the action of the toggle links.

This gripper was designed as a non-overrunning device. If the motor

loses power the gripper will not open and drop the object it was

holding.

The leadscrew, 5 mm in radius, is made out of tempered steel.

It is 110 mm in length, with a pitch of 50 threads per inch. Acme

screws with metal nuts range in efficiency from 35% to 55%. A

median value of 40% was chosen for our application. The steel screw

will be lubricated to reduce the coefficient of static friction. The lead

screw nut will have pins on it for the pinned rectangular bar

connection. The lubricated leadscrew will be covered by a tubular

cover that will be connected to the nut. The cover will slide with the

nut. It will have ball bearings on its ends to help it slide and offer

protection from the atmosphere. We realize that the cover needs to

protect the lubricated leadscrew, although we have not engineered

the cover because of time constraints.

The gripper will have a rectangular bar, 53.41 mm in length,

that is pinned at both ends. This rod will be connected to the

leadscrew nut. The bar rests at 86.3 degrees from the horizontal

when the gripper is closed. The bar then moves to the right with the

nut when the leadscrew turns. The bar rests at 90 degrees from the

horizontal when the gripper is fully open.

Each gripper main rectangular bar will be pinned respectively

at the left end. This is the pivot point upon where the gripper

system rotates. The gripper mechanism will be welded to the robotic

arm at a downward angle of 16.5 degrees. The angle enables the

gripper horizontal access to the vertical rods, held by the footplate,

while the robot arm is off center. By free body diagram, the force

101
required by the gripper is equal to the product of the coefficient of

static friction and the weight of the heaviest object to lift.

F=gW

A total factor of safety was chosen by the product of the factor

of safety for uncertainties in regard to the load and the factor of

safety for uncertainties in the strength of the gripper.

n = nl ns

Fu

nl= F

S

ns -.- --

Crp

Where:

Fu is the limiting value of F

S = yield strength

C_p = permissible stress

The max load of the gripper was then computed using the

factors of safety.

Fu=nlF

An assumption was made in the calculation of the maximum

bending moment. The main rectangular arm of the gripper that has

a bend of 159.68 degrees is approximated as a straight rectangular

arm.

The maximum stress was then computed.

The section modulus for a rectangular section is:

102
I

Z-'-

c

bh 2

Z=--

6

The maximum stress is then reduced to:

6 M

gP-b h 2

The depth of the rectangle can then be computed by solving for b:

S 6 M 6Mns

ns b h 2 b = S h 2

Calculations for b:

nl = 2.0

ns= 2.0

n=4.0

= 0.4, aluminum on aluminum

F = (100/6)

0.4 ( 4.448 ) - 185.3 N

Fu = (2.0) ( 185.3 ) = 370.6 N

Mp = ( .210 ) ( 185.3 ) = 38.9 N-m

The yield strength of aluminum was approximated to be 49 kpsi.

b= ( 6 ) ( 38.9 ) ( 2.0 )

(.012)2 ( 49k ) ( 6890 )

b = 9.6 mm

The height of the rectangular arm is 12 mm, while the base has

been calculated to be 9.6 mm. Therefore the thickness of the

rectangular gripping arm needs to be 9.6 mm to achieve a total

safety factor of 4.0. All rectangular bars of the gripper will have

103
these dimensions. The gripper jaws will have at least 9.6 mm at the

smallest part of the base. Other areas of the gripper jaw base will be

wider.

The clearance for the gripper to move between two adjacent

rods of 100 mm is 168 mrn. The gripper may be holding a 100 mm

object. For the gripper jaw, 9.6 mm of material is required to

prevent yielding. Each jaw only requires 20 mm to open fully and

slip by the rod that was just stored. This leaves approximately 2 mm

of clearance between each side of the gripper jaw and the rods it

passes by. To achieve this 20 mm of opening for the gripper jaws,

3.45 mm of horizontal translation of the leadscrew nut is required.

The gripper will have to go to a preset position for the jaws to

grasp the rod properly. The gripper can not move forward toward

the rod and then be stopped by the rod. The rod would come in

contact with the lead screw. The controls for SKITTER will have to

preset the forward movement of the gripper to grasp the rods.

The torque needed to close the gripper with a force of 185.3 N

must be designated. F1 will be the 185.3 N force required by the

gripper. F2 will be the vertical component of the force on the pin at

100 mm point along the length of the gripper, as measured from the

left end. The relationship between the two forces is:

0.310

F2 = F1 (O_)

F2 is then calculated as:

0.310

F2 = 185.3 (_--0--_) = 574.5 N

The pin is at an angle of 3.7 degrees when the gripper is fully

closed. The component of F2 that is the force on the leadscrew, along

104
the pinned rectangular bar, will be F3.

as"

574.5 N

F3 = = 575.7 N

cos(3.7)

This force is then calculated

The force acting as a weight along the horizontal length of the

lead screw is a component of F3"

W= 575.7 sin ( 3.7 ) = 37.2 N

The torque required to drive the load W using a lead screw

with pitch, p, and efficiency, e, has the following components"

TTotal = TFriction + TAccclcration

F

TFriction = 2xpe

Where:

F = frictional force in ounces

p = pitch in threads/inch

e = lead screw efficiency

F = Its W for horizontal surfaces where Its is the coefficient of

static friction and W is the weight of the load.

1 JLoad CO

TAcc = 3 8 6 ( e + JLeadscrew + JMotor ) T

m = 2xpv

W

JLoad = (2xp)2

rcLgR 4

JLeadscrew = 2

Where"

105
J = inertia, oz-in 2

co = angular velocity, radians/sec

t = time, seconds

v = linear velocity, in/sec

L = length, inches

R = radius, inches

p = density, ounces/in3

The formula for the load inertia converts linear inertia into the

rotational equivalent as reflected the motor shaft by the lead screw.

The following assumptions where made in the calculations for the

required torque"

Its = 0.15, for lubricated aluminum on steel

e = 0.4, for acme lead screw with metal nut

p = 50 threads per inch, a fine pitch is needed

to move only 3.45 mm

L= ll0mm

R = 2.5 mm

9 = 1.54 oz-in 3

v = 4 in/sec

co = 100 milliseconds

JMotor = 0.08 oz-in 2

The torque calculations are as follows"

F=(0.15)(37.2N) =5.58N

1 lb 16oz

F= (5.58 N)(4.45 N)(1 l b ) = 20.1 oz

TFriction =

20.1 oz

2 rc ( 50 rev/in) ( 0.4 )

= .160 oz-in

106
JLoad --

4.45 N 1 6 oz

37.2 N ( 1 lb ) ( l lb .)

( 2rt (50))2

= 1.36 x 10 -3 oz-in 2

JLscrew =

1 in 1.54 oz .25 cm

( 11 cm ) (2.54) ( ) ( )4 cm in3 2.54 cm/in

JLeadscrew = 9.83 x 10 -4 oz-in 2

1 1.36 x 10 -3

Tacc - 386 ( 0.4 + 9.83 x 10 -4 + 0.08 oz-in 2)

2_ ( 50 rev/in ) 4 in oz-in-sec 2

0.1 sec 1 sec oz-in 2

TAcc = 2.681 oz-in

TTotal = 0.160 + 2.681 = 2.841 oz-in

The torque required by the gripper is 2.841 oz-in. The torque

of the motor will be controlled by regulation of the supply amperage.

A plot of current versus torque can be generated from experimental

procedures. An interpretation of the current pattern will enable the

correct current to be supplied for the needed torque.

Experimentation with current versus torque is needed to prevent too

much torque being supplied. This design helps to protect the gripper

system. If the pinned rectangular bar jammed, the current would

increase as the gripper would try to close. The system would be

programmed to decrease the current at a determined value. This

would prevent the gripper from continuing to close on the jammed

bar.

107
The gripper motor will be discussed in the motor section of the

paper, Section 4.6.5. The weight of the gripper motor and gripper

mechanism is 3.33 N.

4.6.2 Materials

Materials Selection

Many areas must be examined when choosing a material for

any engineering application. These areas include indemnification of

material requirements, analysis of the environmental capability of

the material and whether such a material has been used in a given

application successfully. In the case of space exploration, the

environmental capabilities of the materials becomes the most

important factor. Three factors that will affect material selection are:

Temperature - A chosen material must withstand the moon's

temperature gradient which ranges from -158 C to 130 C

during the two week lunar day. Creep and brittleness of

material can occur.

Vacuum Environment This type of environment causes loss

of material by volatilization of diffusion. Many materials

experience cold welding in this environment.

Radiation Damage can occur by formation of lattice defects

which cause loss of strength. Chain scissors and cross

linking of organic materials develop.

Weight is also an important factor.

high strength to weight ratio.

A material chosen should have a

In considering a material for the arm, gripper, rotating plate

and rack and passive gripper system, environment considerations

are of great importance. Plastics can be eliminated due to off gassing

108
and radiation induced decomposition. Both of these mechanisms

cause an increase in brittleness and a decrease in strength. Ceramics

cannot be used because they can not handle loads in tension. They

are also brittle and have extremely low coefficients of expansion

making ceramic -metal combination difficult to design. It necessary

that the mechanical are gripper, arm and rotating disk be of the

same material so as to avoid having to design for differing expansion

coefficients. High-strength alloys based on magnesium, aluminum,

titanium, beryllium and steel are used in aircraft application and

appear to be most appropriate for moon use. Of these alloys,

titanium and aluminum appear to demonstrate the best properties in

the moon environment.

Aluminum has several properties which are particularly

advantageous to moon use. First, the alloys have a high strength to

weight ratio. Second, they are highly reflective. Third, they are

good thermal conductors. Fourth they are corrosion resistant and

fifth, have a high melting point. The 2000 and 7000 series

aluminum alloys are high strength alloys used for structural

components. These alloys have been heat-treated and age-hardened

making them capable of withstanding a large temperature gradient

without deterioration of mechanical properties. The alloy 2219-T37

is particularly attractive. This alloy's major alloying element is

copper. Copper provides increased strength and improves elevated

temperature properties and machinability. It has been heat-treated

solution treated and cold worked to an almost full hard state.

Therefore it is strong but still retains some ductility. This alloy has

structural use at high temperatures up to 600 F. It has also been

used in cryogenic applications and aircraft parts. 2219-T37 is

available in sheets, tubing, bars, extrusions and forgings. The

mechanical properties are as follows:

ultimate tensile strength = 414 MPa

yield strength = 338 MPa

M_g

density = 2.99 m3

109
Young's modulus = 71 GPa

Brinell Hardness = 117

Endurance limit of infinite at 500,000,000 cycles

Titanium alloys have properties much like aluminum. The alloy, Ti6A1-4V, physical properties make it the best titanium alloy for the

moon. This alloy is used for cryogenic applications and has excellent

corrosion and creep resistance. Its mechanical properties are as

follows:

ultimate tensile strength = 1000 MPa

yield strength = 900 MPa

density = 4.4 m3

Young's modulus = 115 GPa

Titanium,s performance in a vacuum environment is slightly better

than than of aluminum. Aluminum cold welds at temperatures

above 300 C while titanium cold welds at temperatures above 500 C.

Considering all properties aluminum and titanium are almost equal

in mechanical performance. Therefore it necessary to compare the

cost, weight, weldability and commercial forms of the materials in

question. Aluminum is much cheaper, weighs less, is easily weldable

and is readily available is more commercial form than titanium.

Further, aluminum has been used successfully in the moon

environment previously. Considering these factors AI 2219-T37 is

the material of choice. The arm, mechanical gripper and rotating

plate will be made of this material.

The rack system will be constructed of the same composite

used in the structural interface. This decision was made because the

rack must directly attach to the structural interface. Thus, providing

good continuity between these two parts. The rack can be attached

with the same type of connection which is used to hold the members

of the structural interface together.

110
All passive grippers, pins, bolts and lead screws will be made

of tempered steel to ensure a long fatigue life. Tempered steel has

been used previous for equipment clips on the lunar rover.

The differences in materials between the rack and the passive

grippers leads to the question of how will the different thermal

expansion coefficients affect the design. Because metals generally

have expansion coefficients which are relatively similar, the design

should not be effected. However, this should be tested.

Although it is traditional to use metals for bearings, plastics

have been considered for their light weight and "self-lubrication"

features. The problems using plastics in space environments that

were previously mentions will no be in effect because the bearings

will be enclosed and not subjected to the damaging radiation.

Plastics can be made self-lubricating be adding graphite,

molybdenum disulfide, or silicone oil to the forming process. The

polymer that has the best characteristics for this application is

polyimide. Polyimide can take a maximum load of 10000 psi,

ft

operate continuously at 5000 rain at a critical temperature of 316 ° C.

It is one of the strongest polymers for this application having a

modulus of 0.45 x 106 psi Plastics have already proven themselves

superior to other engineering materials and have been used

successfully on the moon for interior structural purposes.

Lubrication

Several modes of lubrication are available that can be used on

the moon. Again the environment must be taken into consideration.

The regolith presents a special problem when dealing with

lubrication. The regolith has the consistency of flour and tends to

accumulate on all parts exposed. If an oil or other liquid system is

used it would soon become gummy due to the regolith and lose its

effectiveness. Liquid lubricants are also subject to evaporation,

become unstable due to radiation exposure and lose oxidation

lll
stability. A solid lubrication system is the other form that can be

explored. The advantages of solid lubrication are listed as follows:

1. Shearing takes place more easily when high loads are applied.

2. Can be used at low speeds because it prevents the stick-slip

action of asperities on mating surfaces.

3. Operate under extreme temperatures.

4. Can be used in dirty and abrasive environments.

5. Prevent fretting wear.

6. Simplify design.

7. No maintenance or unreliable parts.

8. Intermittent action with long periods of disuse.

9. No reaction with environment.

10. Stability under nuclear radiation.

From this list, it is apparent that solid lubrication is very appropriate

to moon use. The only disadvantage with this system is that it is not

renewable. Therefore the amount of lubricant film placed on the

parts exceed the required amount for safety purposes. Resin-bonded

molybdenum disulfide would be the lubricant of choice. It has a

wear life of 9,869,000 cycles at 35,000 psi. It also has the lowest

coefficient of friction at 0.036.

4.6.3 Rack and Passive Gripper System

Rack

The rack structure contains three sections, one between each of

the legs of SKITTER. It is made of the same composite that the

structural interface group has chosen. The reason for this selection is

for continuity between the two structures. Both structures will

thermally expand the same and they can both be fastened with the

same connectors and glue.

112
The rack's members are circular rods. The top rods are 17 mm

in diameter. These two rods are 1 meter apart and extend 868 mm

out from the top of the structural interface beneath SKITTER. These

rods are supported by two other rods that are 36 mm in diameter

and extended from the bottom of the structural interface. These rods

support the center of the top rods and are at an angle of 49 degrees

from the horizontal. These rods are 662 mm long. The actual storage

area has to be in an arc because the arm is on a rotating bar. This arc

enables the rod to be grasped in the same place every time by a

spring clip The arc has a radius of 1.51 meters and an arc length of

2.1 meters. The arc has spring clips connected to it that hold the rods.

These clips are at a 61.5 degree angle from the horizontal. The

placement of the clips at an angle is used to compensate for the angle

of the rods when the are placed in the rack by the rod changer arm.

There are 14 clips on each of the three rack section_. The rack has a

space in the center to allow the arm to move from behind the rack to

retrieve another rod or implement. Ten of the clips are for 50 mm

rods and the other four clips are for 1O0 mm implements. The clips

are placed symmetrically with respect to the support rods. There is

140 mm between the center of each of the clips which allows room

for the gripper to release the rod and move behind the rack.

When a rack section is completely loaded, the two support

structures must support 227 kg each. The diameters stated above

were calculated with a factor of safety of 2 and the actual calculations

are shown in Appendix A.6. The stress calculations for the arc were

done by approximating it as a straight beam. The top beams are

experiencing shear while the rods extending from the bottom of the

structure are in compression. The force to place a rod into the rack is

only 34 N for the 100 mm rods and half that for the 50 mm rods.

This force was insignificant in the calculations done.

In addition to force and stress calculations, deflection of each of

the beams were found. The greatest deflection was in the support

structure. It was found to be 5 mm. The arc was found to only

113
deflect 2 mm. The total weight of the a single section of the rack is

5.65 kg.

Passive Gripper

The passive gripper design is based on the principle of spring

clips. This idea was discovered after examining the methods used to

store tools on the lunar rover and in ordinary household situations.

To determine the holding force required by each clip a simple

friction calculation was used:

I.tN = F

p. will be set at 1 for simplicity. The inner surface of the clip will be

roughened during processing to meet this criterion. It is assumed

that the roughening will not cause significant stress concentrations in

the clip. To solve for F the limiting weight of 16.7 lb. was resolved

into x and y components using the 61.5 ° as shown in Drawing 4.6.3.1.

Fx = sin61.5(16.7) = 14.68 lb

Fy = cos61.5(16.7) = 7.97 lb

the requirement of the largest component must be satisfied,

therefore the holding force must meet or exceed 14.68 Ib per clip. A

clip made by De-stat-co fit the design criterion. This clip is made of

tempered steel which eliminates the calculation for fatigue. This

material has an infinite fatigue life up to 500,000,000 cycles. Fatigue

would be the main failure mode as the clips can be considered to be

cyclically loaded. The final dimensions of the clip are shown in

Figure 4.6.3.1.

114
4.6.4 Mechanical Arm and Rotating Bar

Mechanical Arm

In the design of the mechanical arm it was essential to simplify

the mechanisms involved to maximize reliability and minimize

weight. The final accepted design was that of a three degree of

freedom system including the following:

1. A rotating bar to select rack position and change side to grip

drilling implement if needed.

2. A jointed arm to move between the drill string and storage

racks.

3. A bilateral gripper to grasp the various drilling implements.

The jointed arm was designed to grasp the drilling implement at

a point 0.3 meters from the top. The joint will be powered by a

brushless DC motor and will be attached through a double reduction

worm gear from Boston Gear. This motor and speed reduction system

will be attached to the stable upper part of the arm and will be

covered by a small aluminum structure to protect it from the harsh

environment of the moon. The housing will guide the output shaft

with a ring of sealed bearings.

The upper arm will consist of two 0.5 meter long rectangular

cross section beams projecting directly down from the rotating beam.

This connection could be welded or, the rotating beam and upper arm

could be manufactured as one piece. The cross section of the two

rectangular portions of the arm will be 50 mm by 15 mm (See Figure

4.6.4.1).

The lower cylindrical section of the arm goes between the two

upper arm sections. They are connected by a 20 mm shaft that is

part of the lower arm. The shaft is also surrounded by sealed

115
bearings in the upper arm. The lower arm is 25 mm in diameter,

1.02 meters long, and the opposite end is attached to the mechanical

gripper. The total weight of the arm will be 6.1 N moon weight. From

the deflection calculations in Appendix A.2.6 the total deflection of

the arm and rotating bar is not more than 10 mm.

Rotating Bar

The purpose of the rotating blade is to eliminate one degree of

freedom from the mechanical gripper. The blade will consist of a

simple beam which will be attached to the underside of the SKITTER.

Power will be provided by a motor as specified in the motors section.

The blade will rotate on a ring of sealed ball bearings. A

counterbalance will be used to remove stresses from the bearings.

The calculations for the beam are based on the superposition of

a cantilever beam with an intermediate load and a cantilever beam

with an intermediate moment. Using formulas for maximum

resulting stress in terms of loading for the Handbook of Formulas of

Stress and Strain for the loading situations above and setting the

maximum resulting stress at the yield strength of the aluminum with

a factor of safety of 1.5, the dimensions of the beam were found to be:

height = 3.4 cm

thickness = 1.7 cm

length = 0.5 m

The end deflection was calculated using the superposition of the

two different cantilever beam loading situations with formulas from

the Handbook of Formula of Stress and Strain. The deflections were

found to be:

y(point load) = .0052 m

y(moment) = .0094 m

y(total) = .0146 m = 14.6 mm

116
The weight of the bar is 0.838 kg. The calculations for all values can

be found in Appendix A under A.6 Rod Changer and Rack, Rotating

bar calculations.

4.6.5 Motors and Speed Reducers

Brushless Electric Servomotors

Brushless servomotors were chosen for their high efficiency,

high power to weight ratios, and their durability. Also, motors with

brushes will produce significant arcs in a vacuum, an undesirable

side effect. To increase service life, vacuum sealed bearings will be

used. A controller box, containing power control electronics will be

necessary to run each of these three phase motors.

Three motors are used to actuate the rod changer system. One

is used to rotate the beam on which the arm is mounted, another

motor drives the elbow joint on the arm, and a third powers the

gripping mechanism.

Speed Reducers

Speed reducers were chosen based on required torque and

weight considerations. Worm gears were chosen for the arm's elbow

joint and the for the rotating blade, to eliminate backdrive. In the

case of a power failure the locking action of the worm gears will

insure that the rod changing system does not swing uncontrollably

under SKITTER. The gripping mechanism employs a power screw

driven by a harmonic speed reducer and a brushless motor. The

power screw locks the gripping mechanism if power is lost, so as not

to drop a rod or drill implement.

117
Rotating Beam: motor anti speed reducer

For the rotating beam, a relatively moderate torque, low output

rpm is necessary. An Ailing-Lander double reduction worm gear

was chosen that provides a 5000:1 gear ratio. The torque necessary

to rotate the beam is:

T = (load) x (arm) x (safety factor).

T = (96 N)(1.02 m)(1.5) = 147 N-m.

This means that an input horsepower of 0.0012 is required of the

drive motor. At a speed of 10,000 rpm, a Kollmorgen Co. model

A714 brushless motor can provide 1.2 oz-in of torque. The worm

gear weighs 67 N, and the motor weighs 15 N.

Arm elbow joint: motor anti speed reducer

A high torque at a low speed is necessary to swing the arm

from directly beneath SKITTER to the side mounted rod racks. This

move is to be accomplished in about 30 seconds. The maximum

torque necessary is calculated in Appendix A.2 at 140 N-m. A

double reduction worm gear such as Hub City's 1300 series, style URR provides a 3000:1 reduction in speed. A Moog model 303-002

brushless motor with an input horsepower of 0.0112 at 1750 rpm

provides the gear with the required power. Output from the gear is

at 0.583 rpm and 160 N-m. The worm gear weighs 98 N, and the

motor weighs 11 N. These elements can be seen in drawing 4.6.4.1.

Gripping Mechanism: motor anti speed reducer

A power screw is driven to open and close the gripping

mechanism. Torque is transferred to the screw from a brushless

motor thru a harmonic drive. A harmonic drive was chosen for its'

compactness, high efficiency, and low weight. Quincy Technology

manufactures a 2000:1 ratio unit, that is 75 mm in diameter, and 60

118
mm long. The torque required to drive the grippers' power screw is

2.841 oz-in. The harmonic drive requires 0.0098 hp to provide 3.0

oz-in. An Inland Co. brushless motor provides 0.02 hp at 6500 rpm;

a sufficient amount to power the gripper. The harmonic drive

weighs 12 N, and the motor weighs 6 N. The configuration and

approximate dimensions of these elements can be seen in drawing

4.6.4.1.

The weight of the three motors and three speed reducers total

to 209 N.

4.5.6 Conclusions

This system has

consideration in mind:

1.

2.

3.

4.

been design with the following

high reliability

simplicity, with few moving parts

low maintenance requirements

minimum degrees of freedom

5. light weight

6. functionality with other SKITTER components

7. withstand extreme temperature ranges

8. high strength to weight ratio materials

9. durability for long life

10. regolith contamination

The design was limited to three degrees of freedom. The arm

joint, gripper and rotating bar each represent one degree.

Several materials were used in the design so that each

application uses the material with the highest strength to weight

ratio possible. This design can be incorporated well with other

systems that need its use. For this design, the total weight was

as follows:

119
Three racks" 166.28 N

Motors and speed reducers: 209 N

Arm" 6.1 N

Mechanical gripper" 3.33 N

Total weight: 384.71 N

120
d

-tun©

t>©

(.o

U.J

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c_

LL

c_

CO

Q. 00

,'E
FIGURE 4.6.3.1 SPRING CLIP FOR ROD MOUNTING

7.75 cm

(3.25 cm)
aW

n-
4.7 MACROCORE RETRIEVAI_ AND STORAGE SYSTEM

4.7.1 Explanation of Core Removal Process

After the drill and stepped auger are removed from the drill

hole, a device will be sent down the drilled hole to extract the core.

The retrieval device will be attached to the drill string and slowly

lowered down into the drill hole. When the core retriever gets

within two meters of the core, the core retriever is lowered at a

slower pace. If the core is broken and lying against the wall of the

drill hole, the core retriever may need to be rotated to properly feed

the core into the retriever and allow the retriever to continue

lowering. Upon reaching the desired depth the device will

discontinue its descent and rotate a small increment around the

sample to release the cutters and grippers. The device will then be

rotated in the same direction around the core. After several

revolutions, the core will be severed. The grippers will be engaged,

and the core will be lifted out of the drill hole. Once on the surface of

the moon, the core will be released from the core retriever and put

into a storage cell. The core retriever will then be set aside so that

drilling can be resumed.

4.7.2 General Overview of Equipment

Once the core retriever is placed around the core sample, it is

necessary to cut the core. The sample can then be lifted to the lunar

surface. The method which is employed for this task is modeled

after a pipe cutter.

A pipe cutter employs a sharp circular disk to cut pipes. This

disk is attached to a screw such that as the screw is turned, the

pressure applied from the disk to the pipe is increased. The disk is

spun around the pipe and tightened at certain intervals to cut

through the pipe.

121
The same general principle is used in the lunar core sample

case. However, rather than cutting through the core sample, three

disks are used to propagate a crack across the core and severe it.

The propagation is caused mainly by the shape of the disks. The

disks will be thicker at its center than at its edges. Therefore, the

disks also act as a wedge. These wedges are the main driving force

behind the crack propagation.

The wedges propagate the crack because the material of the

core is rock which is very brittle and experiences little to no elastic

deformation. As the wedges are driven further into the core, the

rock simply separates.

Heat dissipation is not a major concern for the mechanism will

be moving slowly, not unlike a pipe cutter. However, since human

assistance is unavailable to tighten the cutter, small springs are

activated by rotating the top of the device and used to drive the

disks very slowly into the rock as the entire apparatus is spun

around the core sample.

After the core is separated, it will be gripped so that the core

can be lifted to the surface. It is important that the grippers apply a

uniformly distributed load across the cross section. Any type of

point loading would run the risk of damaging the specimen. After

gripping the core, the retriever containing the core sample will be

lifted to the surface of the moon. When the specimen is ready for

storage, the grippers and cutters will be reset using a gear. The gear

will be rotated and reset by pressing it against the top of the storage

device. The core sample will then slide out and be deposited into a

storage cell.

122
4.7.3 Introduction

After the drill bit and stepped auger are removed from the

hole, a core retrieval device must be sent down to bring the core to

the surface.

4.7.4 Core Sample

The core sample will be 50 millimeters in diameter and up to 2

meters in length. Maximum mass will be 46 kilograms. It must be

removed from a hole which has a 100 millimeter diameter and can

be up to 50 meters deep. The core should be brought up to the

surface and stored in a storage device that will protect it from the

lunar environment.

4.7.5 Constraints

Various environmental and equipment constraints must be

considered in the design of the macro-core retrieval and storage

system. Moving parts must be kept to a minimum due to abrasion

and wear caused by lunar dust and rock fragments. Temperature

extremes and cosmic radiation must be considered when deciding

upon materials. Reduced gravity and near vacuum conditions must

also be taken into account.

Since little human supervision or maintenance will be

available, the mechanism should be self automated and highly

reliable. A fully mechanically operated system would be ideal due to

power requirements. Transportation of the system puts restrictions

on the size and weight of the retriever.

Restrictions are also put on our design by other parts of the

drilling device. The connection device between the drill string and

the motor attached to SKITTER limits our pulling force to 200 pounds.

The motor provides a maximum torque of 50 ft-lbs.

123
4.7.6 Design Considerations

Several different ways of severing the core were explored. The

methods that were researched included tension, torsion, bending,

cutting, or a combination of these.

The properties of basalt, which is a lunar rock with high tensile

and torsional strength, were used to compute the forces needed to

break the core. Using pure tension to break a 50 millimeter core of

basalt would require a force greater than 54 Mega-Newtons, which is

much more than the 890 Newton (200 pound) restriction (Table

4.7.6.1). The pure torsional force required to severe the 50

millimeter core is also greater than 1353 Newton-meters (Table

4.7.6.2). Thus, pure torsion was not a feasible consideration in

breaking the core.

The force needed to cut the core by pure bending is

approximately 340 Newtons applied at the top of the core (Table

4.7.6.3). Although this method is realistic a sleeve would have to fit

tightly over the core to prevent it from breaking. For a softer rock,

clearance at the top of the sample may not be sufficient to break

core.

Cutting or scoring the bottom of the core could be used in

conjunction with one of the above methods to break the sample. The

core could be scored creating a stress concentration which

significantly reduces the force needed to break the sample (Table

4.7.6.4). The core could also be completely severed using a cutting

device stored in the bottom of the core retriever. The core could

then be held with a gripping device and raised to the surface.

The macro-core retriever uses a pipe cutting device to cut the

core. By spinning a rotating cutting wheel around the core, the core

124
will be severed. The cutting wheel is very thin at the initial contact

point, and gets larger towards the center of the wheel. The geometry

of the wheel causes an outward force at the cut, pushing the core

apart. The outward force, along with the compression force of the

cutter, causes the core to break with a relatively small cut. The core

breaks because of Crack propagation, caused by the concentrated

stresses at the point of contact of the cutter.

4.7.7 Design

The macro-core retrieval is a cylinder 2 meters high with a 100

millimeter diameter (Drawing 4.7.7.1). It is lowered into the hole by

the drill string, and is stopped 2 meters from the bottom of the hole.

At this time the retriever is lowered slowly over the core. The

bottom of the device is conically shaped in case the core is broken.

This will feed the broken core into tile inner tube. The retriever will

be spun counterclockwise if it gets hung up on a broken core.

When the retriever gets to within 3 centimeters of the bottom

of the hole, its descent is stopped. The cutters are then activated by

rotating the retriever 30 degrees clockwise. There are three cutters

in the device. Each is spaced 120 degrees apart at the bottom of the

core retriever. They are activated by a pin placed at the top of the

cutter. The rotation of the top of the retriever pulls a chain that

raises the pin (Drawing 4.7.7.2). The cutters are then pressed against

the core by a tension spring. The device is rotated counterclockwise

around the core while the spring presses the cutters into the sample.

When the cutting blade is compressed 5 millimeters into the core, the

core should be broken. A plate made to the contour of the core will

then make contact with the core.

The plate will grip the core, and the retriever will be raised out

of the hole by the drill string. When the core and retriever reach the

surface of the moon, the rod changer will grab the core retriever and

take it over to the core storage device. The storage device will

125
facilitate the resetting of tlae retriever. A reset gear is contained in

the helical section at the bottom of the retriever. It is cut such that it

is flat against the helical section when ttae cutters are not activated

and swings out when the cutters are activated (Drawing 4.7.7.3). The

storage device has a ridge on its top that will push the gear flat

against the helical section when the retriever is pressed against it.

This releases the core. The core will then slide out of the retriever

into the core storage device.

4.7.8 Retriever Shell

The shell of the retriever will be 2 meters long with an outer

diameter of 100 millimeters and an inner diameter of 55 millimeters.

It will consist of an inner and outer tube. The tubes will be 3

millimeters thick and will be connected by supporting bars placed

between them. The shell will be made out of a titanium alloy, Ti5A1-2.5Sn, which maintains its high strength under the temperature

extremes which will be encountered.

4.7.9 Top Connection

The top connection will allow a slight rotation, with a male and

female connection between the two parts of the shell. This

connection will be made of the same titanium alloy as the retriever

shell.

4.7.10 Pipe Cutter

To cut the core sample, a 10 millimeter disk cutter similar to a

pipe cutter is employed (Drawing 4.7.10.1). The core retrieval device

will use three disk cutters spaced 120degrees apart (Figure

4.7.10.1). These cutters will be spring loaded with a force of 222

Newtons. They will be rotated around the core by rotating the entire

retrieval unit within the hole.

126
This method of cutting the core sample is unusual. Therefore,

no information about removing a core sample in this manner could

be found. However, a similar type of cutting disk is used in some

tunneling operations. As a result, the design of the disk cutters is

based upon disks used to cut tunnels.

The disks within the core retrieval unit are rotated by the

friction developed between the cutting edge of the disk and the core

sample. Traditional steel and steel alloy disks have a tendency to

slip while in contact with the rock. Slipping increases the wear on

the cutting edge and thereby decreases its life. The use of tungsten

carbide helps to prevent slipping on a rock surface. Additionally,

tungsten carbide cutting edges have been shown to successfully cut

through extremely hard rocks which traditional cutting edge metals

do very poorly. For these reasons, tungsten carbide was chosen as

the material to make the disk cutter from.

Another important parameter of the cutting disk is the angle of

the cutting edge. Sharp edged disks are best for penetrating into

hard rocks quickly, but usually need frequent replacement. Disks

with blunt angles tend to penetrate rock more slowly. A cutting edge

angle anywhere from 55 to 60 degrees has been found to offer the

best compromise between effective penetration and wear. From

available data the expected cutting life of the disk cutter is

approximately 200 hours.

The principle by which this method of rock cutting works is

kerfing and compressive spalling. With the kerf principle, the sharp

edge of the disk cutter cuts a groove into the rock, and the wedge

shape of the disk imparts a shearing force causing the rock to break.

The sharp edge of the cutting disk also induces a very large stress

concentration where it makes contact with the rock. This stress

concentration can get as high as 5 times the unconfined compressive

strength of the rock. The force of this stress concentration breaks

through the rock by essentially crushing all of the rock in the path of

127
the cutting edge.

spalling.

This crushing of rock is referred to as compressive

The cutter disks are designed to penetrate the core sample to a

depth of 5 millimeters if necessary. Realistically, the cutters should

never have to penetrate anywhere near 5 millimeters. There are no

formulas on which to base this statement. Rather, this statement is

based on theoretical considerations from books and upon a solitary

test conducted by Mr. James Brazell. In this test, Mr. Brazell used an

ordinary pipe cutter to cut a granite core sample approximately 45

millimeters in diameter. With a very minimal amount of force, the

core sample was separated in just one revolution.

4.7.11 Spring

The spring will be a simple tension coil spring (Drawing

4.7.11.1). It will have 8.5 coils, an outer diameter of .0832 meters,

and have a wire diameter of .0019 meters. It will be grounded and

squared at both ends. The spring constant (K) will be 2.32

Newtons/meter and exert a force of 222.5 Newtons when fully

compressed. A 186.59 Newton force will be exerted at the end of the

cutters stroke. A material for the spring has not been found that will

give the characteristics needed on the moon.

4.7.12 Reset Gear

The reset gear, used to reset the cutters after the core has been

extracted from the hole, will be a normal gear with 18 teeth. The

pitch diameter is 2 centimeters, and the angle of incident is 20

degrees. The face width will be 2 millimeters allowing a 222.5

Newtons of tangential force. The gear will be cut so that it will be

flush with the conical section of the retriever when the cutters are

reset.

128
4.7.13 Pin

The pin will be cylindrically shaped with rounded ends. It will

be made of tungsten carbide with 10 percent cobalt to insure good

wear life.

4.7.14 Grippers

There will be three pads that will grip the core after it is

broken. The pads will be located directly above the three cutters

and are powered with the same spring that the cutters use. They are

shaped at an arc that will enable them to grip the core with the

entire pad (Drawing 4.7.14.1). The geometry of the pads prevent

point loading, which could cause harm to the sample. The pads will

be made out of tungsten carbide with 10 percent cobalt. This

material has a good wear life and can be used at the temperature

extremes that the retriever will encounter.

4.7.15 Protective Gear Cover

There will be a flexible material around the gear and spring to

protect them from dust and dirt. The material will be made of

polyester yarn of Dupont type 56 dacron and has a urethane coating

system by Mobay Chemical Company. It has a cyclical life of 2 million

cycles. This material is used for space suits.

4.7.16 Core Storage Device

The storage system is designed to facilitate the storage of up to

26 core samples from a single hole. Each core sample has an

individual unit-cell storage cylinder 100 millimeters in diameter and

slightly bevelled to 50 millimeters. These unit-cells, however, are

located in an overall rectangular storage system (Drawing 4.7.16.1).

Four alternating columns of 6 and 7 unit-cell cylinders respectively

129
are positioned at a 60 degree angle to enable the SKITTER arm to

swing upward, release the core, and allow gravity to assist in

depositing the sample. In addition, the inner two columns will be

recessed to allow for the SKITTER arm's sideways motion.

The entire system will be positioned in a location away from

the drilling area but within easy access of the SKITTER. The operator

will guide the SKITTER arm from the hole to the storage cylinder

group and position it over the correct, pre-determined unit-cell for

deposit. A marker band will be stretched across the entrance of each

unused cell to assist the operator in discerning where to place the

core sample. The marker will be torn and broken upon core deposit

into the unit-cell so that the operator will be better able to tell which

cells are full. Also, each unit-cell will be pre-labeled to account for

the core sample's position in the hole and the particular hole from

which it came. A circular ridge lining the bevelled area of each unitcell will facilitate the macro-core retriever resetting process. As the

core sample is deposited into the unit-cell, the ridge provides

resistance to the pressure applied by the macro-core retriever and

the gear is reset.

The storage system and unit-cells will be constructed of Ti-5

A1-2.5 Sn alloy. This material has an extremely low coefficient of

thermal expansion (9.4 micro-in./in/C), and all material properties

remain essentially the same at both high operating and low

environmental temperatures.

130
Positioning of Cutters

Drill Rod Interfac_

Activator Linkage

Cutter Assemblies

120 Q apart

SECTION A - A

TOP VIEW

Re set

/ \

I I

I I

SIDE VIEW OF RETRIEVER

Figure 4.7.10.1

A
4.8 CONTROLS FOR ENTIRE DRILLING OPERATION

4.8.1 Task Definition

The initial step of designing the controls for this operation was

to define the tasks that were to be performed to complete the

drilling operation. The entire operation was broken into five basic

routines that appear in the main flowchart. Each of the specific

activities of the drilling operation was then specifically determined

in flowchart form. The flowcharts and a discussion of the methods

needed to read them appear in section 4.8.2. Each routine is

uploaded to an onboard Programmable Logic Controller, PLC, on the

drilling apparatus as the routine is needed. When the specified task

has been completed that routine is then down loaded and the routine

for the next task is then uploaded. This process allows maximum

autonomy for the drilling apparatus while maintaining consistent

contact for the operator.

After the tasks to be performed were defined and organized, it

was determined where the operator would have input into the

drilling operation. This involved careful consideration of the current

design as well as the operation of other components of the drilling

operation and SKITTER in general. Since there is approximately a 2.6

second communications delay between the Earth and Moon and back,

full operator control is not possible without risking damage to the

equipment; also, the time required to complete the drilling procedure

would be significantly increased. Fully automating the process,

however, would lock out human input and possibly prove as

damaging as full operator control. In addition, the task of operating

the equipment would cause the operator to become distracted due to

lack of involvement. It was therefore decided that that a semiautomatic approach was more suited to the problem. Human input

would be placed at critical points in the routines where decisions on

131
options must be made as well as to allow tile operator to monitor the

operation and intervene when a problem develops.

Drilling Operations

Operational Flowcharts

In the course of designing the control methodology for the

drilling operation, it was found that tile most efficient way of

illustrating the steps in an operation was to construct a flowchart. In

this section, each significant operation has been reduced to a

flowchart that illustrates the complete sequence of events needed to

complete a task.

Initially, tile drilling operation was broken clown into its major

activities; these are referred to as routines. In each case these

routines will be uploaded and downloaded to the onboard PLC to

reduce the size of the required RAM memory areas. The five basic

routines and the overall operating sequence for the drilling operation

are illustrated in Fig. 4.8.2.1.1.

Once determined, the routines were written as the design

process was completed for the respective areas. Each routine is

designated in its own flowchart anti appears in Figures 4.8.2.1.2

through 4.8.2.1.11. Each flowchart specifies the action of the integral

operation and refers to tile operational subroutines.

The subroutines were developed as areas of operation that are

necessarily repeated frequently. Tile appropriate subroutines are

appended to each routine that utilizes them and uploaded and

downloaded as a set. Each subroutine appears as a separate figure in

this section and are organized to be near those routines using them

most often.

In reading the flowcharts, there area few standards of which

the reader must be made aware. Boxes represent specific system

132
commands that are acted upon directly by system hardware.

Diamonds are decisions that are integral to the routine and are a

answered affirmatively or negatively by logical iterative procedures

in the routines. Ovals represent subroutine calls. In each subroutine

call, the machine returns to the original routine at the same point it

exited to the subroutine and continues forward from there. The

unwritten understanding here is that SKITTER and movable

hardware is always left in the same standard position as it was found

upon entering a routine or subroutine. SKITTER is always in the

"normal" position at the beginning and end of an operational

segment, and the rod changer arm is always in the indexed position..

These positions are discussed in the next subsection.

133
4.8.2.2 Operational Positions

The specific areas that must have operational positions defined

for SKITTER is the actual drilling operations and the vertical

accelerator. These positions are defined in relative terms since

quantitative information was not available at the time of writing.

The specification of these positions is noted as a recommendation for

further study as it may prove a good troubleshooting method in

evaluating operations. These positions are illustrated in order of

discussion in Figures 4.8.2.2.1 a,b,c,d.

SKITTER begins the drilling operation in the "normal" position.

This position is the maximum working height that will be used in the

drilling operation and will serve as the standard indexed reference

position for SKITTER. At the "normal" position the rod changer arm

has clearance to operate while carrying a rod, auger, bit, and the

various devices used in other operations. When the desired

component has been placed in position over the hole or drill string,

SKITTER lowers into the "chucking" position at which the footplate

can grasp the component. In the "chucking" position the rod changer

arm has clearance to move only when not carrying a component.

Generally, the rod changer will then be indexed to its standard

position after releasing the component. SKITTER then lowers to the

"motor lock" position which allows the motor to engage the drill

string directly. The drill motor would then start and SKITTER would

lower through the drilling procedure ending finally in the "drilling"

position. At this point, components other than those in the hole must

be stowed or in their standard indexed position as this is the point of

least clearance between SKITTER and the lunar surface.

It must be noted that these positions are used as standards for

automatic operation. Operation of the routine requires that specific

positions be defined to ensure the smooth transition from routine to

subroutine. Specifying these positions by no means disallows the

134
operator from moving SKITTER or the various components manually.

In fact, this will certainly be the required action in an error situation

or when damage occurs.

4.8.3 Human Factors Design

In the design of the control system for the lunar deep drill,

considerable analysis of the interface between the operator and the

drill were made. The first part of the analysis dealt with dividing

the control tasks between operator and computer. The second part

was to design the optimal control console to be used from Earth.

4.8.3.1 Workstation Design

Several criteria were used to determine the optimal means of

human-to-hardware interface. The console had to be easy to use,

keep the operator's attention, allow the operator to monitor the

drill's systems and tile drilling operation's progress, allow simple but

complete control over the drill, and meet NASA specifications. After

careful consideration of the requirements and alternative solutions, it

was decided that a computerized system utilizing both a touchsensitive screen and a normal computer keyboard would be used.

On-screen menus would be the primary means of input from the

operator.

As well as the visual feedback provided by animation of the

menu screens, the operator will be able to monitor the operation

through audio signals. A computer generated "whirr" will imitate the

sound of the motor that the operator would hear if he were at the

drilling site. This system will be analogous to the driver of a manual

transmission automobile. Accompanying the motor simulation will

be an audio simulation of the grinding of the bit into the rock, giving

the operator a sense of how hard the drill motor is working and

when it is approaching a stall. This sound will correlate to the motor

torque much as the motor "whirr" will correlate to the motor rpm.

135
These two audio signals will enhance the operator's sense of the state

of the drilling operation.

4.8.3.2 Menu Screens/Animation

The touch-screen computer system will use a number of

different screens to allow access into the computer system, monitor

the operations, give the operator visual feedback, and allow the

operator to input decisions. The screens are multicolored with

simple animation gauges and animated sketches showing the

functions Skitter is carrying out. They are sequentially ordered

according to the decisions made by the operator. The operator must

touch the green "CONTINUE" area to advance to transmit the

instructions to Skitter and advance to the next screen. They were

designed in accordance with NASA-STD-3000 9.3.3.4.7. Touchsensitive areas in use are outlined to show the operator what his

choices are. On any given screen, there are few enough active touch

areas to space them adequately to prevent accidental or erroneous

activation. During normal operation, the keyboard is only needed to

enter identification/password/access code and to enter the target

drilling depth. Touch area activation is indicated by a color change in

the activated area, except in the case of the "CONTINUE" area which

will result in a new screen appearing. Examples of these screens can

be found in Figures 4.8.4.1.1 4.8.4.1.14, with explanations and

instructions following each screen layout.

4.8.3.3 Console Design

The primary criteria in the design of the control console was

NASA-STD-3000. The console will be textured matte grey, which is

in accordance with NASA-STD-3000 9.2.2.2.4. This color reduces

reflections from illumination sources in the immediate environment.

The touch-screen shall be positioned directly in front of the operator,

136
within arm's reach, and tilted back 10 degrees from vertical (NASASTD-3000 Figure 9.2.4.2.2-1). The edges of the screen will be near

the optimal +15 degree vertical and horizontal eye movement ranges

(NASA-STD-3000 Figure 9.2.4.2.2-2). This orientation of the screen

will allow the operator to easily see tile screen with a minimal range

of eye motion. Standard display controls (intensity, contrast, etc.)

will be below the screen. Speakers for the audio feedback will be

placed at the top of the console on either side of the screen. The

computer keyboard is retractable, since its use is limited, so as not to

inconvenience the operator. Armrests will be provided on each side

of the console, to help prevent operator fatigue while operating the

touch-screen. The Touch-Screen Disable toggle switch shall be placed

on the right side of the screen; this switch will disable the touchscreen capability, allowing the screen to be cleaned of smudges and

dirt without accidental actuation. The Emergency Stop button shall be

located on the right side of the screen, positioned to provide

immediate activation capability. The console will stand at a height

such that the operator's knees will be free from contact with the

console. Shelving for operation manuals, etc. will be available within

arm's reach of the operator to his/her left. The design developed

based on these criteria can be found in Figure 4.8.3.3.3-.4 as well as

an artists rendering in Figure 4.8.3.3.5.

137
4.8.4 Menu Driven Operational Displays

As discussed in the previous section, the method of direct

interaction between operator and apparatus will be through the

touch screen control console. Approximations of the actual screens

appear in order of operation in the following figures. These screens

illustrate, generally, what the operator will see while running the

apparatus through its various operations. The following are

comments on the screens themselves.

4.8.4.1 Earthbound Control

Initially control will transfer from the SKITTER control console

to the DRILLING OPERATIONS console when SKITTER seats the

drilling operations rig, thereby activating it through passive sensors

in the bomb latches securing the operation. The operator will move

to the DRILLING OPERATIONS console and go through a series of

entry requests shown in Fig. 4.8.4.1.1. Once into the system, Fig.

4.8.4.1.2, the operator will choose the pre-operation check so that all

the systems on the drilling operation can be verified to be ready. If

the operator has specific programming or editing that must be done

he would choose system access. At this point, notice that the screen

has a continue arrow at the bottom right and a stop figure at the

bottom left. These are utilized by the operator to proceed after a

decision has been made or return to the previous screen or system

access to manually resolve a problem.

Choosing the pre-operation checklist brings up the screen in

Fig. 4.8.4.1.3. As the system performs its checks the screen

highlights those areas being checked with a box and highlights those

areas that do check out with a green check mark. Those systems that

fail the check are indicated with a red "X" mark. If a system has

failed, the operator can strike the stop sign and go to system access

or accept the failure if trivial and continue. Assuming all systems

are operational, the operator will strike "go" and tile screen in Fig.

138
4.8.4.1.4 will appear. At this point the drilling operation has been

checked and is prepared to begin drilling. Control is then returned to

the SKITTER console so that the operator can move SKITTER to the

drilling site. Fig. 4.8.4.1.5 remains on the screen while SKITTER is

moving to the drilling site, at which point control will revert back to

the DRILLING OPERATION console and the operator will strike

continue.

At this point the drilling operation can begin and the DRILLING

OPERATION console has nominal control over SKITTER movement

vertically. The Set Footplate screen, Fig. 4.8.4.1.6, comes up and the

appropriate routine is uploaded. As the routine completes the task,

the animated figure of SKITTER and the footplate go through a rough

analogy of the motions that SKITTER is actually completing. The box

at the bottom of the screen will show in words the current stage of

the routine, similar to the corresponding flowchart points. The

vertical bar on the left side is a graphic representation of completion

of the routine and goes from top to bottom as the routine is

completed.

The Drill Segment screen, Fig. 4.8.4.1.7, asks the operator to

input the depth to which the hole will be drilled. This sets the

markers at the left, which will be maintained to give the operator

constant update as to the level of completion of the drilling

operation. Upon entering the desired depth, the operator will strike

continue and the Drilling screen, Fig. 4.8.4.1.8, will appear. The

appropriate routines are uploaded and downloaded, the bit is

positioned and drilling begins. Again the screen animation illustrates

the actions of SKITTER as the operation progresses. Throughout the

drilling process the screen is updated as to depth, motor rpm, motor

amperage, SKITTER position, and rod changer arm position, as well as

the ongoing depth update on the left.

Macro-coring animation is shown on the screen as in Fig.

4.8.4.1.9. Again the operation is animated and the operator can

monitor the flowchart through the bottom box. At the completion of

139
the macro-coring the operator is asked to decide whether microcoring is desired or not on the screen shown in Fig. 4.8.4.1.10. Here

the choice is made and is indicated by a lighted box on the screen. If

"Continue Drilling" is chosen, the screen reverts to that shown in

Fig.4.8.4.1.8 and the process is iterated, if "Micro-coring" is chosen,

the screen in Fig. 4.8.4.1.11 is revealed and the animated process is

monitored by the operator once he sets the clesired depth. Again the

decision screen in Fig. 4.8.4.1.10 appears and the operator selects the

next option. The drilling complete option is not available until the

desired depth has been reached so that the operation can only be

halted by the manual interdiction of the STOP button on the console

or through system access. Once the desired depth has been reached

and all coring is complete, the drilling complete option can be

selected and the screen in Fig. 4.8.4.1.12 appears. The footplate is

then retrieved and the screen in Fig. 4.8.4.1.13 is shown. At this

point the system can be accessed for other work or control is

reverted to the SKITTER console so another site can be selected.

These screens are not all inclusive of the detail that will be

necessary in operating SKITTER. A great deal more detail will be

required, but the direction chosen appears to be the most efficient

and interesting to the operator. Further activities are reported in the

Recommendations for consideration in design.

140
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4.9 VERTIt_AL ACCELERATOR

4.9.1 Introduction

In order to remove the hot particles from tile drilling chamber

and the bit, a vertical reciprocating stepped auger must be

implemented, as discussed earlier. These particles will collect in the

bit collection chamber until it is full. A vertically reciprocating device

can now be used to cause the particles in the collection chamber to

advance up the small steps of the stepped auger. Therefore, this

vertically reciprocating device will be able to induce a certain

amount of volumetric throughput of particles per stroke. This

volumetric throughput is mainly a function of the design parameters

of the stepped auger and the upward applied force of the

reciprocating device.

The final design for the vertical accelerator for the stepped

auger is one which uses a sinusoidally shaped cam and a cam

follower. This cam converts the rotational mechanical energy of the

rotary drive into vertically reciprocating mechanical energy. The

motion supplies the necessary acceleration to the drill string and the

stepped auger to cause the particles in auger to advance. Illustrations

of the accelerator are shown in the following figures: 4.9.1, 4.9.2,

4.9.3, 4.9.4 and 4.9.5.

The cam-follower idea was chosen over the others because of

its simplicity and reliability. This method does not need a secondary

motor which would add weight and complications to the overall

system. This idea also has only two degrees of freedom, the rotating

sinusoidal cam and the reciprocating cam chamber which attaches to

the drill string.

The accelerator runs for 20 seconds at 50 rpm for a period of

time, Trun. This time duration is only a function of the amount of soil

141
moved per stroke of the accelerator, through the stepped auger, and

the amount of soil contained in the lower soil collection chamber. Tru n

should be long enough to completely empty the bit collection

chamber (193,000 mm 3 of soil). Two accelerator strokes occur per

revolution so the stroke frequency is 1.67 strokes per second. For

the mathematical analysis of the advancement of the particles up the

auger and component designs, see subsequent sections.

4.9.2 Detailed Operation Description

When the bit collection chamber is full of moon regolith the

goes into auger mode. The rotary drive cuts off and the SKITTER

stands up 152 mm to disengage the jaw clutches (A & B). Refer to

figures 4.9.1-4.9.5. Now no torque is transmitted to the drill string,

but the telescoping shaft still transmits torque to the cam (H). The

rotary drive begins to turn the cam, and the cam housing (F) will also

turn until the rollers on the outside of the cam housing hit against

the vertical stops. As the rotary drive now continues to turn the

cam, the cam roller followers roll on the cam, accelerating the cam

housing up and down and in turn shaking the entire drill string. The

cam housing rollers roll up and clown the vertical stops keeping the

cam housing from rotating. When the bit collection device is empty,

the rotary drive stops, and the SKITTER goes back into drill mode by

squatting back down engaging the jaw clutches and releasing the cam

housing rollers from the vertical stops allowing it to rotate. This

process is then repeated every 13 mm of drilling.

142
4.9.3 Specifications

Parts List

(A)

(B)

(C)

(D)

(E)

(F)

(G)

(H)

Upper jaw clutch

Lower jaw clutch

Telescoping shaft upper shaft - outer splines

Telescoping shaft lower shaft inner splines

Cylindrical vertical accelerator housing

Cylindrical cam housing

Vertical stops

Sinusoidal shaped cam

4 Cam roller followers with needle bearings (25 mm dia.,

13 mm thick)

2 vertical stop roller followers (25 ram, 13 mm thick)

Operational Requirements

Stroke distance

Frequency

motor speed

Power

Distance SKITTER lifts to aug

56 mm

1.67 strokes/sec.

50 rpm

300 watts

152 mm

Material

The entire accelerator is made of titanium alloy : Ti-6A1-4V

at 204o C (400oF) yield strength = 4.88 MPa

at 24o C (75oF) yield strength = 5.74 MPa

Roller follower and bearings: McGill Camrols-case hardened

steel for smooth rolling and low wear.

Roller follower: Case hardened steel

Needle bearings: silver nitride ceramic; recommended by

Torrington.

143
Because no lubrication is necessary, they can be manufactured

to our specifications.

4.9.4 Control Operation Sequence

1). SKITTER squats engaging the jaw clutch which will transmit

torque to the drill string.

2). The rotary drive drills down thirteen millimeters into the

lunar surface, then drive stops.

3). SKITTER stands up 152 millimeters disengaging the clutch so

that the drill string will not turn.

4). The rotary drive begins to turn at 50 rpm which means two

strokes per revolution or 100 strikes per minute. The drive

should run for 20 seconds.

5). When the auging is completed, the motor stops and the

SKITTER squats back down engaging the jaw clutch and begins

drilling another thirteen millimeters.

4.9.5 Math Model for the Vertical Accelerator

The advancement of the particles through the stepped auger

requires that the particles be given an upward velocity great enough

such that if the stepped auger were abruptly stopped, the particles

would continue upward until they would impact with the underside

of the helix above them. This velocity of impact must be great

enough such that the particles would be "knocked" into the steps

further up the stepped auger helix. The maximum distance that the

particles could be knocked would be a perfectly symmetrical

rebound path off the underside of the helix. However, due to the

gravitational pull of the moon and the inelastic particle-particle and

144
particle-auger collisions, this overall advancement will be somewhat

less than a symmetric rebound. This non-ideal rebound path would

therefore be somewhat parabolic rather than a straight line as shown

on the following page (Figure 4.9.1).

Due to the complexity of the analysis, a few assumptions have

been made. The assumption that gravitational effects on the

particle's rebound path will be small when the applied force is large

was made. Also, the assumption that all particle-particle and

particle-auger collisions will be elastic due to the difficulty in

measuring these quantities without experimentation was made.

•ght Line Approximation

T

Parabolic

Profile

P

Figure 4.9.6

Therefore, the particle's advancement will only be a function of the

applied upward force, and its rebound path will be linear and

symmetric with a trajectory of impact when the upward force is

sufficient.

The particles must be given a certain velocity in a certain

distance, so they must be accelerated. The velocity required to be

145
sufficient to impact the helix adequately was determined from an

energy balance on the particle's path, letting V1 be the particle's

velocity when it leaves the auger and VIMP be the velocity of the

particle when it impacts the underside of the helix, or the roof.

Applying the concept that the particle's increase in potential energy

will be approximately equal to its decrease in kinetic energy, and

then solving for Vt yields the relationship shown below.

VI = [ gmoon ° p + ] 1/2

[4.9.1]

In the above equation, P represents the distance that the particle

must travel upward. This is the equivalent to the stepped auger pitch

for single-helix augers and one-half of the pitch for double-helix

augers.

The vertical reciprocating device must be capable of achieving

V1 in a certain distance, hi, which will be determined by the space

constraints of the overall design. Using the following relationships,

V1 = (h) [4.9.2]

a = I-_) [4.9.3]

the acceleration of a particle will be given by the following

relationship after the expression for V1 has been substituted into it.

"P) moon +

al = h

VIMp2) [4.9.4]

2h J

146
Therefore, the required force to produce this acceleration will be a

function of the total mass, mtot, of tile drill string. This includes the

masses of the drill bit, the stepped auger (empty and full), and the

drill rods which vary from one to twenty five lengths. This force will

also have to compensate for any frictional effects that the stepped

auger encounters on the sides of the hole. It will be assumed that the

diamond tip coated drill bit will produce rather smooth walls in the

hole. The hole diameter also allows for some clearance between the

stepped auger and the hole walls. Therefore, frictional effects will be

considered negligible. The required upward force is shown below.

Fup(min) = mtot

• [ gmoon " P (VIMP2_ h + _ ] [4.9.5]

In order to complete the model, it becomes necessary to derive

an expression for the required impact velocity, VIMP, tO provide the

assumed straight line rebound approximation. The analysis of the

dynamics of the stepped auger will be simplified somewhat as it will

be treated as if the helices of the auger were simply unwrapped.

Therefore, we are dealing with a ramp, rather than a helix This

assumption is not a poor one considering the fact the particles

furthest away from the center of the auger advance the least

anyway. Therefore, the outside radius is used as the basis for the

VIMP model. The particle-wall collisions, again, were assumed elastic,

therefore, the path of the particles up the auger steps in the helix

will not vary much from the path of the particles in the unwrapped

ramp, except for the absence of induced vorticity in the bulk particle

motion in the helix. The parameters involved in the stepped auger to

be used are the vertical pitch, P, the number of helix turns, N, the

number of fences per turn, FT, the fence height, FH, the outer radius

of the stepped auger, Ro, the angle of inclination of the helix, 0O, and

the advancement of the particles, A. The inclination angle is shown

below and given by the following relationship.

147
Therefore; 0 = tan-1 [4.9.6]

It is also necessary to analyze the vector dynamics of the

velocities involved in the paths of the particle, both before and after

the impact with the roof. This geometry is shown in the figure.

IVtMPI

Roe, o,_ ''X Vy

,,, \

'._p..-/'lVIMpI

s

Figure 4.9.8

In the above figure, the quantity c_ represents the angle (0), 13

represents (90 0), and X represents (90 - 20). Therefore, the

following relationships represent Vx and Vy.

Vx = IVIMpI cos (90 - 20)

Vy = IVIMpI sin (90 - 20)

[4.9.7]

[4.9.81

Using the following equations of motion for a particle in a

gravitational field;

x = Vx't

Y = - gmoon "

2 + .t

[4.9.101

[4.9.11]

148
The impact velocity relationship was then determined. Eq. [4.9.10]

was solved for t and substituted into Eq. [4.9.11]. Eqs. [4.9.7] and

[4.9.8] were then substituted in to yield the following relationship.

Y = (gmoon'). t 2 ) / IVIMr, I cos(90-20) x )+ x tan(90-20)[4.9.12]

In Eq. [4.9.12] the variable x represents the horizontal component of

the advancement shown in figure 4.9.1 and y represents the vertical

component. Therefore, Eq. [4.9.12] can be satisfied by the

x and y components of the position at which the rebound leaves the

particle, shown below.

x = A cos do [4.9.13]

y = A sin do - P [4.9.14]

Substituting Eqs. [4.9.13] and [4.9.14] into Eq. [4.9.12] and solving for

the impact velocity yields the following relationship.

VIMp = 2 (cos(90_ 2do))ZqiA-sin__ -g ° (A cos0) p_ (Acosdo)tan(90_Zdo)) 2 / [4.9.15]

The expression for the overall upward force required for the

reciprocator to accelerate the particles enough to advance up the

stepped auger is found by substituting Eq. [4.9.15] into Eq. [4.9.5].

The parameters which now need to be quantified are P, h,

m TOT, dO, and the advancement, A. If the rebound trajectory is

assumed to be a straight line, the maximum advancement would

149
result. From the geometry of the stepped auger, this would result in

an advancement shown below.

Amax = 2P sin[ tan-1 ] [4.9.16]

Amin = FHsin # +I_) [4.9.17]

Eq. [4.9.17] was used to check the stepped auger design to be sure

that the particles would make it to the next step and over the fences

given the calculated upward force. The parameters P, N, FH, FT, Ro,

and _ are all design parameters of the stepped auger. The vertical

accelerator for the stepped auger will therefore be designed to

optimally supply the auger with the required force. The only

parameter left to be determined is the vertical displacement of the

auger, h, which will be discussed in the next section.

The maximum upward force which can be tolerated must also

be considered. The vertical accelerator will pull on the drill string

with a force, Fup. When the stepped auger is abruptly stopped, the

force required to stop the drill string from continuing upward must

be considerably less than the weight of the drilling apparatus so that

SKITTER will not hop around.

4.9.6 Cam-Follower Analysis

In order for the rotational mechanical energy from the rotary

drive to be adequately converted into vertically reciprocating

mechanical energy for the drill string, a sinusoidal cam and follower

system will be used.

From the math model discussed earlier, the velocity that the

particles must have to impact on the underside of the helix and then

150
rebound with a straight-line approximation is approximately 427

mm/sec. This value was determined after calculating the necessary

impact velocity to be 368 mm to have the particles follow a straight

line approximation given the design of the stepped auger and all of

its parameters. Therefore, the cam must accelerate the drill string

and the stepped auger such that the final velocity of particles in the

stepped auger, when it is abruptly stopped, will be about 427

mm/sec.

The cam-follower design is basically a contoured disk mounted

to a shaft which extends through the top casing of the accelerator

which attaches to the motor. The cam itself has a "hill" profile which

the weight of the drill string, supported by the rollers, must travel

along. In the course of this motion, the follower must ride smoothly

and properly over the cam's hill profile. The rotary drive rotates the

cam with a certain angular velocity, and, in conjunction with the

heights of the hill profile anti the cam radius, Ream, will determine the

velocities and accelerations of the drill string, stepped auger, and in

turn, the particles will be determined.

Certain initial constraints had to be set when the motor speed,

cam profile, and cam size were being determined. To support the

weight of the drill string, two followers, 180 ° apart, are used so that

the reciprocating motion occurs smoothly without any horizontal

movement or moments caused by using only one roller follower.

Therefore, for the weight of the drill string to be reciprocated

smoothly, the entire hill profile of the vertical motion must occur in

180 ° of rotation in the cam. The cam virtually has the exact same

hill profile repeated twice in 360 ° of the cam. This is so that the

same heights of the hill profile will coincide with the follower

supports at the same time.

The hill profile can be be broken down into three different

periods of consideration: the rise section, the settling section, and the

dwelling section. This was when the next constraints came into play.

The rise section must incluce a fin_ll velocity of V1 = 427 mm/sec to

151
the drill string as discussed earlier. Also, the settling section must

allow the particles sufficient time to travel up off of the auger,

impact with the helix, and rebound into the next steps. Finally, the

falling section must allow the drill string, the stepped auger, and the

particles on the auger to lower to the starting position with an

acceleration equal to or less than that of the gravity of the moon.

This was done to minimize the amount of particle chaos in the

stepped auger. In addition to these time constraints, trise, tsettle, and

tfall, the size of the cam and the necessary vertical displacement, h,

were constrained by their magnitudes to keep the accelerator light

and easy to handle. A 30 ° maximum pressure angle was also

implemented.

The design of the settling time was first considered. It was

found that a settling time of 0.2 seconds was appropriate for the

particles to rebound and land in their final destinations. With this

time known, the trise and tfall must be split up from the remaining

time for 180o of rotation of the cam.

The value of tratl was considered next. It was found that this

time was a function of the height, h, that the drill string had been

displaced. The relationship is shown below:

[4.9.19]

Therefore, the determination of the value of h was necessary before

the natural necessary falling time for the drill string could

be determined and compared to the amount time for failing which

had been allowed.

In order to determine the height, h, and trise the initial

constraint that V1 must be at least 427 mm/sec was employed as

shown below:

152
trise = [4.9.20]

Many iterations were done to test all of the constraints of the final

cam design.

It was finally decided that a 113 mm radius cam can be used at

50 rpm with 55 mm of overall vertical displacement. This design

creates the necessary velocity required for particle advancement in

the auger and satisfies all of the mentioned constraints. The final

cam hill profile for 180 ° of rotation is shown below with the vertical

displacement as a function of time, induced by the motor speed, the

angle, and the actual horizontal displacement of the cam if the 180 °

of cam were unwrapped into a line instead of a semi-circle.

E

vE 55-

¢-

T

0.13 seconds 0.20 seconds 0.27 seconds

of rise of settle of fall

45 90 135 180 Degrees

0.15 sec. 0.30 sec. 0.45 sec. 0.60 sec. Time

89 mm 178 mm 267 mm 356 mm Displacement

The thickness of the edge of the cam changes slightly on the exterior

of the cam. This is necessary so that the smooth passage of the cam

through the roller supports on each side will not be impeded. During

the transitions from rise to settle or settle to fall, the cam will

become thicker or thinner respectively to compensate for the change

in cam angle. A modeled cantilever beam analysis could be

performed to determine the minimum thickness of the cam with the

153
given radius, but the maximum thickness will be specified as 20 mm.

This corresponds to a minimum thickness of about 17 ram.

4.9.7 Component Analysis

Bearing Analysis

The following is the analysis of the maximum load on the cam

bearings:

Weight of full drill string = 245 N (on moon)

Acceleration of drill string (maximum) = 3277 mm/s 2

Maximum force to accelerate drill string

F=ma= I 245 )1(m ) N . . 3277

mm

1646

Fmax = 487 N

Because the roller follower is being pushed at a maximum of a 30

deg. pressure angle it actually reacts to a larger force.

487 N

Fmax = cos(30 o) - 562 N

Maximum force on any one roller is then half of this total force.

Froller = 281 N

Load rating"

CR = Frollcr" _,nD))

Using an L10 life of 5000 hr and design rpm of 450:

LR = 3000 hr

nR = 500 rpm

154
LD = 5000 hr

nD = 450 rpm

a = 10/3

((5 000"_ (4 5 0")']3 / 10

CR = 281 N, _,_) _,5-'-0-O))

CR = 317.5 N

The roller followers have a much higher load rating than the

computed one estimated by Torrington; so the roller follower life

should last through about 450 fifty meter holes, but we oversized the

follower for about ten times that amount. Since there are extreme

operating temperatures on the moon going as low as -158 °C at night,

lubricants would just evaporate, therefore, the bearings used are

designed without the need for any lubrication.

Shaft Analysis:

Diametermin = 25 mm

T max = 28.6 N-m

Moment = 0

Sy = 827 MPa

n = ? (factor of safety)

Using a basic static failure analysis and solving for n (factory) gives,

(32___M___ a 0.8 66T))1/3 d= ( r¢ I.Sy + Sy

d3Sy7_

n= 0.866(32)T

(1)3(827 MPa)(3.14)

n= 0.866(32)(28.6)

n = 645 Factor of safety

155
The reason for this high factor of safety on the smallest diameter

shaft is because our design uses the same size rods as the rods in the

drill string, and the vertical accelerator does transmit the required

torque to the drill string during the drilling process. Thus, the shafts

designed in the vertical accelerator are consistent with the drill

string design and not designed based on factor of safety.

4.9.8 Design Development

Prior to the derivation of a satisfactory Vertical Accelerator,

very in-depth analyses of design options were accomplished.

Initially, various notions for oscillating the control rods and the

stepped auger were conceptualized and recorded. These various

methods of operation were divided up into specific categories and

developed as a Version 1 design. The categories were as follows:

Reciprocating Cam Mechanisms, Spring Actuated Devices, Gear

Powered Devices, Secondary Motor Configurations, and a

Miscellaneous section for designs which did not fit into one of the

preceding areas. Next, a full evaluation was performed on each

Version 1 design. The evaluation emphasized design effectiveness

with respect to initial constraints. A final recommendation for each

design was rendered. Unsatisfactory designs were discarded and the

remaining feasible designs were developed further into subsequent

versions. Following these versions, a more critical in-depth

evaluation was performed with a more significant emphasis placed

on initial constraints and a final design was brought forth (see table

4.9.1).

156
Table 4.9.1

DESIGN A DESIGN B DESIGN C DESIGN D DESIGN E

(CAMS) (SPRINGS) (GEARS) (2nd MOTOR) (OTHER)

Version 1 Version 1 Version 1 Version 1 Version 1

Version 2 Version 2 Stop Here Stop Here Version 2

Stop Here Version 3 Final

Design

Stop Here

Design E (Version II) will be the final configuration choice for the

report

Final design provides effective means of oscillating Auger to

raise particles. If the pitfalls in following designs (A through D)

could be overcome, these alternative designs could prove useful.

Their difficulties are as follows:

Design A (Cams) -

Cam designs incurred complexities regarding controlling and

secondary motors. Space constraints and cam alignment posed minor

obstacles as well.

Design B (Springs)

Although simple in design, spring options were exhausted due

to poor precision, minimal oscillations, and abrupt motion.

Design C (Gears) -

Gear Configurations posed significant space interference

problems. Secondly, the presence of lunar dust on the exposed gears

gives rise to a demand for lubrication. A more in-depth study on

these items could possibly solve discrepancies and lead to a

satisfactory gear device.

157
Design D (External Motors) -

External Motors incur additional controlling and power

requirements. Supplemental maintenance and high operational costs

greatly minimize the feasibility of an accelerator requiring this

added power.

For a more in-depth analysis of the design alternatives refer to

Appendix B9.

Present Design

The present design omits need for external motors and

controlling. This accelerator fits onto the rod changer table. It is

activated by the SKITTER raising up and down.

158
Fig. 4.9.1 - Vertical Accelerator

( General Arrangement )

Rotary Motor

Skitter

. "......

t

i

"o°..o

..... Vertical

Accelerator

Refer to Fig 4 9 2 Housing

Footplate

_.. Drill

_" String / /

Stepped /

Auger Drill

Bit _
Fig. 4.9.2 - Vertical Accelerator

( Internal View )

Rotary Drive

Interface

_'_'.':'Y,_.:_::_X¢_2.:::::_:':':':':"

Control Rod Table

Torque Limiter

Upper Roller Bearing

Upper Clutch Plate

Telescoping Shai_

Lower Clutch Plate

Cam Housing

Vertical Stop

Drill String Magnet

r Detailed View, Refer to Fig. 4.9.3

m
Fig. 4.9.3 - Vertical Accelerator

( Detailed _iew )

Interface k_

Control Rod Table _"

Torque Limiter

®

®Q

@©
Fig. 4.9.4- Detailed Specifications

(Clutch and Shaft Details)

_ Upper Clutch Plate _ Lower Clutch Plate

,-{- 102 mm -_-{ _,

_deg.

64 mm

m

76 mm (I.D.) Through Hole

© Inner Shaft _) Outer Shaft

45 de_i._._,/ 32 mm

ili

-_:--.::j

i ._i_i_ J

51mm

25 mm

45 deg.

I

{

{

32 mm

57 mm

64 mm

89 mm
Fig. 4.9.5 - Detailed Specifications

(Housing Details )

® Vertical Accelerator

Housing

© Cam Housing

-_ _-38 mm -_k 6 mm

165 mm

--_ _-25 mm

394 mm

_--19 mm

56 mm

6 mm thru hole

4 places

6 mm

m

165 mm

r

B

70 mm

25 mm

305 mm
4.10 MICRO-CORING DRILI,

4.10.1 Controls Requirements

Once the drill has been lowered to the required depth, an

acoustical signal will be sent through the rod. This signal will be

generated by a vibration induced in the rod by the impact from an

object This impact must produce a vibration with an amplitude

larger than that of the noise that will be associated with the process

of connecting and moving the rods. Experimental analysis of the

entire project will be necessary to determine the amplitude and

frequency of the vibration produced by the motion of the rods and

the strength of the impact needed to produce a vibration that will

have a greater amplitude than the above mentioned noise. The

results of the experiments should take into account the damping of

the vibration due to the length of the rods and the interface at which

the rods are connected.

The vibration will be sensed by a piezoelectric accelerometer

placed on the inner surface of the rod. At this point an impulse will

be sent to a motor that controls a camshaft. A set of two cams will

be used to rotate the bracing device (see 4.10.3) into position. This

180 ° rotation will also rotate two more cams that will lower the

sliding door used to protect the drill bit and the sample. This

camshaft rotation will trigger a limit switch that will initiate the

drilling process. The drill will translate 54 mm with a positive

inclination of 20 ° in order to remove the pristine sample. At the end

of the translation the sample will displace a plug and the bit will

clamp down on the sample. At this time two limit switches will be

stuck in order to reverse the translation and stop the rotation, one by

the motion of the plug, and the other by the drill bit.

The rotation will not be stopped immediately, but be allowed to

continue for about 5 seconds in order to Cut a grove into the sample.

159
cams will be rotated 180° , locking the bracing pads into

the lunar basalt surrounding the micro-coring drill rod.

Upon completion of the core retrieval process, the microcontroller will send a signal to the bracing motors to

rotate the cams to their original positions, allowing the

spring-loaded bracing pads to retract away from the

lunar rock surrounding the drill rod.

(3) The Speed of the Rotational Drive Motor- The

rotational speed of the drill will be monitored through

the use of photodiodes. The photodiodes will send a

digital signal which will be analyzed by the microcontroller and converted to an analog signal by a D/A

converter. This analog signal will be used to control the

duty cycle of the power source of the rotational drive

motor to maintain an angular velocity of 400 rpm (300

rpm if the bit becomes too hot) during the coring

process.

(4) The Temperature of the Drill Bit- The drill bit

temperature will be monitored by the use of

thermocouples attached to the bit. If the temperature of

the drill bit exceeds 290 ° K, the rotational speed of the

micro-coring drill will be reduced to 100 rpm (to allow

the bit to cool) until the temperature of the bit drops

below 280 ° K. If the temperature of the bit exceeds 300

o K, the rotational motor will be turned off until the bit

cools by 25 ° K. Through these control functions, the core

sample obtained will not be contaminated by excessive

heat.

161

PRECEDING PAGE BLANK NOT FILMED
4.10.1.2 Microprocessor Selection

The microprocessor-based control system for the micro-coring

drill will utilize a Motorola HCMOS MC68HCllA8 8-bit microcontroller. The hardware features of this device include:

• 8K Bytes of ROM memory

• 512 Bytes of EEPROM

• 256 Bytes of RAM

• 8 Channel, 8-Bit Analog-to-Digital Converter

° Enhanced 16-Bit Timer System

Assembly language programs to control the processes listed in

section 4.10.4 of this report will be written by Control Dynamics

Corporation in Huntsville, Alabama. This company has vast

experience in the field of controls and currently is working on

several NASA projects. Control Dynamics will interface with

Motorola for pattern generation of these programs on the chip.

Data from the photodiodes and LED's used to control the

rotational speed of the motor will be inputted to the micro-controller

through pin A1 (located on serial port A of the device) into a pulse

accumulator. By comparing the accumulation of pulses with a known

internal clock speed for the micro-controller, the rotational speed of

the motor can be determined.

Signals from the thermocouples located in the drill bit and from

the piezo-electric accelerometer will be processed by the analog to

digital converter for the micro-controller through pins on general

input port E. Communications to the device controlling the duty cycle

of the power source for the rotational drive motor will take place

general I/O port D. A schematic of the microprocessor-based control

system for the lunar micro-coring drill is shown in Figure 1.

162
4.10.1.3 Piezoelectric Accelerometer

The micro-coring operation will begin once a signal is sent

through the rods in the form of a vibration. The vibration will be

sensed by an accelerometer located near the drill bit, on the inner

surface of the rod. In an accelerometer a coil-in-magnet

arrangement is used between the mass and the base and this forms

an electrodynamic system. At frequencies above the natural

frequency of the mass on the spring, the voltage output from the coil

is almost directly proportional to the velocity of the base of the

transducer.

Piezoelectric accelerometers can measure accelerations in a

range of 2x10-6 g to 105g in a frequency range of 0.1Hz to 60kHz.

Bruel and Kjaer (UK) Ltd. is the manufacturer that produces the best

piezoelectric accelerometers. After experimentation the frequency

range of the vibration associated with the micro-coring operation will

be determined, and the accelerometer that is best designed for this

application may be selected. The sensitivity of commercially

available piezoelectric accelerometers ranges from 1.5 mV g-1 to 10

V g-1 and the weight of the accelerometer from 0.5 g to 500 g.

4.10.1.4 Rotational Speed Control

From previous calculations (see Appendix A.10.1b), the optimal

rotational speed for the micro-coring drill was determined to be 400

rpm. Due to the inherent nonhomogenous cross section of the lunar

rock (into which the coring will take place), it would be nearly

impossible to

power the drill at this velocity by a constant speed motor. Therefore,

a feedback control system will be used to maintain the drill speed at

a nearly constant value.

The light beam from a light emitting diode (LED) will be passed

through a slit cut into the drill bit (near the rotational drive motor).

163
This light signal will be be collected by a photo-diode which

transforms the optical signal to a series of digital pulses. These

pulses will then be read into pin A1 of the microprocessor, where it

will be fed to the pulse accumulator of the chip. By comparing the

rate of pulse accumulation with a known internal clock speed for the

micro-controller. By comparing the actual motor speed to the

desired motor speed, the microprocessor will adjust the duty cycle of

the nickel-cadmium batteries powering the rotational drive motor.

Triggering devices and opto-isolators will be needed to interface the

micro-controller to the motor and photo-diode, thus completing the

motor control circuit.

4.10.1.5 Drill Bit Temperature Control

Since the core sample cannot have a cross-sectional

temperature gradient of greater than 50°, it will be critical to closely

monitor the surface temperature of the drill bit. To accomplish this

purpose, thermocouples from the Omega Corporation will be placed

in the crown of the drill bit 0.010 inch from the bottom of a face

diamond.

Westinghouse Electric Company had success using this type of

set-up in simulated lunar drilling tests conducted in 1967. Analog

signals from the thermocouples will be fed through operational

amplifiers (to increase signal strength) into general input port E of the

microprocessor. The thermocouple input is then passed through the

micro-controller's A/D converter and is processed by the chip as a

digital signal.

The temperature from the thermocouples will be sampled every

second. If the average sampling temperature from the thermocouples

exceeds 290°C (corresponding to a temperature gradient of 40°C for

the core sample), the desired rotational speed for the drill bit will be

reduced from 400 rpm to 300 rpm. If the mean temperature of the

thermocouples exceeds 300 °C (corresponding to a temperature

164
gradient of 50°C for the core sample), the rotational power to the drill

will be cut off for a period of ten seconds. After this ten second

delay period, the temperature will be analyzed again. If the

temperature gradient of the sample is small enough at that point, drill

rotation will be continued.

4.10.1.6 Limit Switches

Limit switches are electromechanical devices which must be in

contact with the object sensed and usually have a digital output.

There are several possible actuators, such as rods, levers, or cams,

and the selection of the actuators depends on the object to be sensed

and its direction of approach to the limit switch. Because of the

application, limit switches are recommended over the more sensitive

but more delicate photoelectric sensors. Limit switches are low cost

devices designed for industrial applications and, because of their

dependability and resistance to wear, are indicated for the rugged

lunar environment. For our application, we will use limit switches

from Micro Switch Company located in Charlotte, North Carolina.

4.10.2 Drill Protecting Device

The drill rod is divided into different sections. The actual drill

is contained in the bottom section of the rod. The core sample

obtained from the drilling process must be kept safe from

contamination. This is accomplished by raising and lowering a

protective cylinder over the drill. When the drill rod is lowered to a

specific depth the drilling process will take place. The protective

cylinder will be lowered out of the way. When the drilling is

completed, and the sample is retrieved into the rod, the cylinder will

be raised over the drill. The movement of the cylinder is

accomplished by the use of the same camshaft that drives the

bracing system. The cylinder is attached to a set of connecting rods

which are driven by the camshaft. The protective cylinder needs to

be lowered far enough so that the drill bit can be translated outward.

165
The drill bit is 22 mm. in diameter, therefore the protective casing

must be lowered at least 22 ram. if not a few more, in order to

provide adequate room for the drill to translate out into the

surrounding rock.

* NOTE - The bracing system and protective cylinder will rely on the

same control system.

4.10.3 Bracing Mechanism

The micro-coring rod is designed to obtain a rock sample from

the lunar surface. The sample will be obtained by the use of a coring

drill. It is desired that the drill remain stationary while the drilling

operation is taking place. For this to happen the drilling rod must

remain relatively fixed. This is accomplished by a bracing system.

The bracing system will consist of a motor and camshaft

configuration. A set of bracing pads will be extended from the microcoring rod. These pads will be rectangular in shape and will be bent

to conform to the side of the main hole. The pads will be equipped

with an abrasive surface. The bracing pads are attached to a set of

connecting rods which are manipulated by a camshaft. The camshaft

will be rotated by a chain drive or a set of gears. The power will be

supplied by a D.C. motor.

4.10.3.1 Cams

The main objective of the camshaft is to provide translational

motion to the bracing pad rods. Since the camshaft will only rotate

once every drilling mission, the cam does not have to be designed

with a high degree of precision, therefore a simple harmonic rise and

return motion was chosen to design the cam lobes. A cam mechanism

with reciprocating roller follower was chosen for this design. The

motion of the bracing pads are described or detailed by a

displacement diagram. A displacement diagram relates a known

input motion (in this case the rotation angle) to the output motion,y.

The diagram is divided into three regions; the rise, where the motion

166
of the follower is away from the center of the cam, the dwell, where

the follower is a rest, and the return, where the motion of the

follower is back towards the center of the cam. The displacement

diagrams for simple harmonic rise and full-return motion are shown

in Figures 2a. and 2b. The equations for simple harmonic motion are:

(rise) y=_(1-cos L b)

(return) y= L_(1 +cos_b )

The micro-coring drill rod has a diameter of 90 mm. , therefore the

pads must be translated radially outward a distance of 5 mm. The

cam will rotate 180 degrees, at this point the bracing pads will be

fully extended. When the drilling operation is completed the cams

will rotate another 180 degrees to return the pads to there original

positions. When the rotation has been completed the drilling rod is

ready to be raised. The camshaft used for our bracing system was

designed to achieve a maximunl lift of 5 ram. When the bracing

pads are in there extended position a force , due to the drilling, will

be exerted on the cams. To stop the cam from slipping once the pads

are in place a dwell was placed in our cam design. Various dwell

angles were considered including 30, 45, and 60 degrees ( see

Figures 3, 4, and 5 for the displacement diagrams ). A dwell angle of

45 degrees was chosen after consulting with various individual

involved in this field. This means that for a cam rotation of plus or

minus 22.5 degrees from the maximum lift angle, the lift will remain

5 ram. This dwell should keep tile cam from slipping due to any

forces. The actual cam profile is given in Figure 6.

4.10.3.2 Bearings

167
The camshafts will be supported by a pair of roller bearings.

The life of an individual bearing is defined as the total number of

revolutions, or the number of hours at a given speed before failure

occurs. Since the camshafts in our rod only rotate once per drilling

operation, bearing life should not be a major problem. Therefore, a

set of bearings should be picked which are light weight. Due to the

lack of an atmosphere the bearings can not be lubricated using

traditional means. Instead, a dry lubrication system will have to be

employed.

4.10.3.3 Bracing Pads

The drill rod will be kept firmly in place, during the drilling

operation, by a set of four bracing pads. The pads will be positioned

flush against the drill rod's outer surface. When the drill rod is in the

correct position the pads will be extended outward. To insure a

stable base for the drilling operation, the pads must be contoured to

fit to the surrounding wall. The pads must also have an abrasive

surface to make sure that the drilling mechanism does not rotate. If

the drill rod rotates the drill bit could be misaligned and the drilling

operation would have to be cancelled. Various abrasive surfaces

would have to be tested in order to determine the optimum design.

4.10.3.4 Bracing Pad Return Assembly

The bracing pads are connected to the camshaft by the use of

connecting rods. The connecting rods or cam followers must remain

in contact with the cams at all times. Furthermore, the pads must

return to there original position before the drilling rod can be raised.

This is accomplished by a retaining spring. The springs will be placed

around the connecting rods, see Figure 7 A spring rate must be

chosen to insure that the cam follower remains in contact with the

cam. Testing on such a cam system would have to be done before the

correct spring rate can be determined.

4.10.3.5 Selection of Materials

168
The bracing pads will be constructed from high strength steel,

this is also the case for the connecting rods. The cam shaft will be

machined from a steel similar to the one used for the bracing pads.

Since none of the bracing operation entails a high degree of

reciprocation, fatigue life should not be a major concern. All

mechanisms should be constructed as light as possible due to the

high cost of space transportation.

4.10.4 Rotational Movement of Drill Bit

4.10.4.1 Heat Transfer Analysis

Since the core sample is to undergo a series of selenological

tests, including carbon dating, it cannot be contaminated with

excessive heat. Thus, the amount of power that can be used to rotate

the drill bit in the process of obtaining a micro-core sample is limited

by the amount of heat generated during the drilling process. The

ambient temperature of lunar basalt (into which the micro-coring

process will take place) is -23 ° C and the temperature of the lunar

selenological lab will be 27 ° C. Thus, we chose 50 ° C, the temperature

difference between these two locations to be the maximum

allowable temperature gradient across the sample during the microcoring process.

Through the use of Fourier's conduction law, the maximum

allowable temperature gradient for the sample was used to calculate

the maximum permissible amount of heat generation for the coring

process. This value was set as the upper limit for the rotary power

input into the micro-coring drill. Ideally, a two-dimensional,

transient heat transfer analysis of the region near the drill bit should

be carried out to relate the heat generation from the drilling process

to the temperature profile of the core sample. However, by making

several reasonable assumptions, we simplified the heat transfer

problem to a one-dimensional, steady state problem to meet the

169
time limitations of this project.

The assumptions and calculations performed to determine the

maximum allowable rotational power input for the micro-coring drill

are listed in Appendix A.10.1. From these calculations, we

determined the maximum permissible input power for the drill

(taking into account only thermal contamination) to be .242 HP.

Since the one-dimensional heat transfer calculation included several

assumptions, we chose a more conservative .12 HP as the maximum

allowable input power.

4.10.4.2 Torque Requirements for Rotation of Drill Bit

The torque the bit requires to cut the lunar rock was

determined by estimating the dry coefficient of friction between the

diamond tipped drill bit and the lunar rock. The coefficient of

friction was found from the experimentai data of the Westinghouse

lunar drill experiment discussed in the 1967 Industrial Diamond

Revolution Conference in which basalt was substituted for lunar rock.

The speed, thrust force, and power were recorded at various points

in the experiment and were used to determine a value of .64 for the

dry coefficient of friction at the rock-diamond interface. The

experiment was performed using a drill bit larger than the microcoring bit used in our design, but this size difference should not

affect the value obtained in the Westinghouse test.

The thrust force driving the micro-coring drill was set at 50

lbs. The thrust force multiplied by the coefficient of friction gives a

resultant force on the bit. A radius of 20 mm was used to calculate

the maximum torque on the bit of 12.6 in-lbs. This value of torque is

considered to be the maximum value since the the thrust force will

usually be less than 50 lbs.

4.10.4.3 Rotational Motor Selection

1 70
From material property considerations at the diamond bitbasalt interface, the coring drill will be required to deliver a

rotational torque of at least 12.6 in-lbs (see calculations in Appendix

A.10.1.2). Taking a factor of safety into account, the rotational drive

motor will be designed to deliver a minimum 14 in-lbs of torque

over a speed range of 300 to 500 rpm. When inputting its maximum

power at 500 rpm, the motor will still generate less than half the

limiting horsepower of .242 (to prevent thermal contamination of the

sample). To meet these design criteria, we will use an Electrocraft E400 servo-motor which has the following features:

Peak Torque

Thermal Resistance

Length

Weight

Maximum Oper. Temp.

15 in-lb

4.0/Watt

25.4 mm (1.0 in)

2 lb 2 oz

155°C

This motor is a high performance permanent magnet

servomotor which combines many of the characteristics of a "torque

motor" with those of a low inertia moving coil type. It has excellent

voltage regulation and low inherent ripple torque.

4.10.5 Drill Translation Mechanism

The translation mechanism for the micro-coring drill consists of a

small motor attached to a rack and pinion drive system. The pinion

measures one-fourth inch in diameter, and power necessary for

motor is 0.1 Hp.

By using a drilling time of two minutes, a rock drilling rate of 26

mm per minute was obtained. The torque on the motor shaft was

found to be 12.6 in-lbs by using a coefficient of friction of 0.64

between the rock and the diamond-tipped bit. In order to traverse

the drilling distance in the two minutes, the speed of the pinion and

thus the translation motor was found to be 1.3 rpm's. Operation of

171
small motors at speeds as low as 1 rpm is non-traditional, but can be

done. Electrocraft, a manufacturer of these type motors has assured

us that these specifications for a motor are feasible.

During the drilling process, the drill bit is guided by a bearing

sleeve around the bit. This sleeve is mounted to the side walls of the

rod chamber both at top and bottom. The rack used in translating

the drill bit is supported by a series of support members. These

members are anchored to the main support, a bulkhead mounted

horizontally within the rod chamber.

4.10.6 Micro-Coring Drill Bit Design

4.10.6.1 Overall Description

The drill bit design is one that resembles many coring bits used

on earth with a few modifications. The micro-coring bit is composed

of natural West African Boartz Diamond cutters and a 4140-stainless

steel body (see Drawing 4.10.6.1). There will be four longitudinal

slits cut in the bit, dividing the face of the bit into four distinct

sections. The four sections of the bit will be machined to provide an

inward clamping angle of two degrees (see Drawing 4.10.6.3) to aid

in breaking the micro-core sample. The longitudinal slits also direct

cuttings from the micro-coring process away from the 6 mm

diameter sample. Finally, the slits increase the flexibility of the

body of the drill bit, allowing a teflon plug to be inserted inside the

bit.

The inner diameter will be 10 mm for a majority of the length

of the bit (it will have a 6 mm I.D. for the 3 mm of length closest to

the diamonds. Located inside this diameter will be a 10 mm wide

Teflon plug which will aid in the bit clamping mechanism. The entire

bit will be mounted to a shaft which extends from the rotational

drive motor.

172
4.10.6.2 Diamond Selection and Mounting

The diamond type and configuration on the micro-coring bit

was determined by using the diamond bit recommendation chart

located in a book written by the Acker Drill Co. and from the results

from the Westinghouse simulated lunar drilling experiment. The

lunar rock (into which the coring operation will take place) was

classified as a medium hard, solid, abrasive material similar to

granite or basalt. The recommendations are as follows:

1. Diamond size:

2. Diamond pattern:

3. Diamond amount:

1/50 carat

Spiral Snow Plow Set

127 stones

There will be a total of 2.15 carats on the end of the bit. Since

cost was not a parameter, the quality of the diamonds will be high

commonly known as AAA diamonds. From the Westinghouse

experiment, tests results showed that the arrangement referred to

above effectively drill a core sample from basalt under high vacuum

conditions of lxl0 -6 ton-.

4.10.6.3 Bit Clamping Mechanism

The mechanism will operate as normal drill bit. The bit will

rotate and form a sample 6 mm in diameter. The significance of the

Teflon plug comes in to play at this point. The drill bit will be

translate inward and as the bit moves inward the sample moves

deeper into the bit, thus pushing the Teflon plug backward. At a

point 46 mm from the front of the bit, the I. D. will widen again to 10

ram. The plug will move into this area of greater diameter and

release the four sections of the bit. The whole mechanism will be

designed to stop rotating at this point and retract. The bit has firm

grip on the sample and the retraction will break the sample off in

173
tension. In order to insure that the sample breaks, the bit will rotate

a few addition times after the plug has released the sections and this

rotation will cut a groove in the sample which creates a flaw. Also,

the bit clamps will be pointed and this will increase the radial force.

The increase in radial force has been shown to lessen the force in

tension necessary to break the sample. The flaw and the

concentrated radial force will insure that there is a clean break and

this will provide a sample 46 mm long and 6 mm in diameter. The

sample received is larger in diameter than required and this

adjustment is necessary for two reasons. First, the outer surface of

the sample will be contaminated with fragments from the drill bit so

an additional .5 mm was added to account for this contamination.

Secondly, the sections of the bit may begin to clamp inward and

actually taper the sample so an additional .5 millimeter was added in

order to insure that at the smallest end the sample would still be 6

mm in diameter.

4.10.7 Lubrication and Bearings

The bearings used for the rotational and translation motors and

the guide bearing will be anti-friction bearings in which the retainer

rings are made of a solid composite lubricant consisting of silvermercury and teflon. This composite was used with great success in

the Westinghouse lunar drill experiment. The gear and sliding

mechanisms will be coated in a dry lubricant containing molybenum

disulfide. Molybenum disulfide is not adversely affected by the

vacuum conditions. The trade name for the dry lubricant that would

be used is Surf-Kote. It is a molybenum disulfide-resin mixture. The

lubricants will lessen the power requirements by reducing the

friction loss.

4.10.8 Dust Jacket

174
The movement of the rods against the main coring hole will

cause large amounts of dust and debris to fall downward towards the

micro-coring drill bit rod (located at the bottom of the drill string),

possibly jamming it. The best way to avoid this problem would be to

insert a rubber-like barrier to isolate the bit from a majority of the

dust. This barrier will attached to the outside of the micro-coring

rod just above the top of the sliding door mechanism used to expose

the bit to the lunar rock for drilling. The dust trap will be able to

conform itself to the shape of the hole and will have a concave

upward orientation as the drill string slides down the 100 mm

diameter coring hole. This contour will allow the trap to catch a large

amount of the dust falling from the upper regions of the main coring

hole while the sliding door mechanism is open and drilling is in

progress. The rest of the falling debris will pass around the outside

edges of the trap (which is wide enough to protect the drill bit from

this dust.)

Once the micro-coring operation is completed, the sliding door

mechanism closes (with the drill and micro-core sample inside the

micro-coring rod) and the drill string begins to move upwards. As

the micro-coring drill rod begins to ascend, the rubber part will take

on a concave downwards orientation, releasing the dust. To allow the

dust trap to change contour without cracking in the extremely cold

lunar environment, a low durometer rubber should be used for this

application. However, the material must be hard enough to support

the load of the debris falling onto the trap during the micro-coring

process.

4.10.9 Failure Scenarios

Since the micro-coring drill is located in the bottom rod of the

drill string (at a depth of up to 50 m from the surface of the moon),

failure of the micro-coring drill bit must be considered. To prevent

the whole micro-coring rod or even the whole drill string from being

175
lost if the bit becomes embedded into the lunar basalt, we developed

methods of overcoming the following failure modes:

(1) Drill Bit Becomes Embedded - Can Be Retracted- If

the microprocessor reads the actual drill speed to be 0

rpm for over ten minutes and the temperature gradient

of the sample does not exceed 30 ° , the direction of the

translational motor will be reversed in an attempt to pull

the drill bit out of the lunar rock. If the bit can be

retracted from the micro-coring hole, it will activate a

limit switch which will close the micro-coring rod sliding

door mechanism, retract the rod's bracing, and signal the

surface to raise the drill string. When the micro-coring

rod is brought to the lunar surface, the bit will then be

inspected and corrections will be made to prevent future

failures.

(2) Drill Becomes Imbedded Cannot be Extracted- In

the event that the drill bit cannot be extracted from the

lunar rock, our design ensures that only the bit and the

sliding door mechanism (as opposed to our whole rod or

the whole drill string) is lost. If reversing the direction

of the translational motor fails to free the drill bit from

the lunar rock, a distress signal will be sent to the

HCMOS micro-controller. This will instruct the controller

to retract the micro-coring bracing and send Skitter a

signal to lower the rod with a force of at least 100 lunar

lbs and raise it with a force of at least 200 lunar lbs. The

downwards force will break the bit off from the

rotational motor at the motor shaft (actually, only 30

lunar lbs. of force are needed to do this), and the

upwards force will break off the sliding door mechanism

from the rest of the micro-coring rod. The removal of

this door will allow the drill string to be raised even if

the micro-coring bit remains locked in the lunar basalt

(like a cantilever beam).

176
i

Motor To Rotational Controller ]

and LED's for

Rotational Motor

Control

To Thermocouples

and Acceleromemeter

Figure 4.10.1.

177
Figures 2a. and 2b.

----_. gz /t//y _----_ _ i'L

_0_5 Y '"_ \ \" "\'-. 1 I

y
Figures 4.10.2a. and 4.10.2b

5

,--, 4

E

E

'-" 3

2

0

Cam Displacement Diagram

(Dwell = 30 degrees)

............... 1........ ......... i............... I............... I............... !.... _,

0 30 60 90 120 150 180 210 240 270 300 330 36

angle [degrees]

Figure 4.10.3

178
5

E

E 4

._ 3

=_

2

Cam Displacement Diagram

(Dwell = 45 degrees)

....... _ ............ ! ....... i ............ _ i

3/ _ i I i _ [ \!

• I ' ' I ' I ' t • ' ' , ' ' t ' I • "_

0 30 60 90 120 150 180 2 0 240 270 300 330 360

angle [degrees]

Figure 4

179
,

i_====1

E 4

E

3

'N,=,,,

"- 2

5-

Cam Displacement Diagram

(Dwell = 60 degrees)

0

= i _ ! i

1 ! i i i

!)!:iiii! .....

!!!,

0 30 60 90 120 150 180 210 240 270 300 330 360

angle [degrees]

Figure 5

180
Cam Profile

90

120 .¢-.- 60

_°_ _°22°- _°- ° -

240 300

27O

Figure 6

181
Micro-Coring Drill Rod Chamber

r-- -'-I

Bracing

Mechanism

Drill Bit/

Motors

Bracing

Mechanism

F i g u r

182
Drilling Operation

Translation

Motor

Bulkhead

Lunar

Rock

Figure 4.10.8

183
Drill Bit Translation Mechanism

Lunar

Rock

debris guard

guide bearing

Figure 4.10.9

Lunar

Rock

184
5.0 CONCLUSIONS

The entire drilling system was designed with three objectives;

minimize weight, minimize moving parts, maximize reliability. The

cost of transporting the implement to the lunar surface is directly

related to weight. Therefore, when possible, light weight materials

were chosen for all design elements. The control sequence from the

earth incorporates an approximate 3 second time delay. To minimize

control problems from complex design, the minimum number of

moving parts was chosen. Also, maintenance on the lunar surface will

be very difficult therefore minimizing the moving parts will help to

increase the reliability of the machine.

The overall Deep Drill Design uses the SKITTER mobile platform

to transport and help actuate the drilling process. The drill consists of

a 100 mm coring bit, augers, rods, a footplate, rod changer, storage

racks, a support structure, a macro-core retrieval tool, and a microcoring apparatus. The control sequence described provides the

automation and control necessary to produce 25 , 2 m core samples.

This design is an initial attempt to produce a mining or

construction machine capable of drilling to such depths. The design

indicates that a better material transport system from the bit to the

top of the hole is needed. This would minimize the heat at the bit and

allow for longer drilling periods. Also, the bit life would be greatly

enhanced. The rods, which connect the bit to the rotary drive and

transmit the power, are tubular with a simple connector which is

easily handled by the rod changer. While this design is a good

concept, initial interfacing suggests that a better design for the

connection collar is needed. The auger and stepped augers are now

being used primarily as a means to store the cuttings away from the

bit. The auger interface is very similar to the rod interface. However,

it causes a significant reduction in throughput. Therefore, a closer

look at the interface is indicated. The driver for the stepped auger is a

185
cam which is located in one of the rods of the string. The cam

changes the rotary motion of the motor shaft into an upward

acceleration. The foot plate provides a very good working base for

the lunar drill. The fins on the base of the footplate aid in the

support of the hole along with the attached casing. The spring driven

rod holders allow the rod changer ample time and space to add and

remove rods to the drill string as needed. Storage racks are located

between the legs of the SKITTER and hold all the drill implements

necessary to complete one 50 meter hole. The power system is

fueled by a hydrogen-oxygen fuel cell located in the supporting

structure of the drill. The actual rotary drive for the drill is a curvosynchronus linear motor. This motor has an efficiency rating around

95%. The fuel cells also power several other small motors on the

implement. After the desired depth has been achieved, the microcore retrieval unit is inserted to retrieve smaller core samples from

the walls of the hole. The unit is housed inside a rod and controlled

with a microprocessor.

The Lunar Deep Drill Apparatus design needs more work in the

area of interfacing between the different systems. In the next

section, numerous recommendations are made to improve upon the

overall design.

186
6.0 RECOMMENDATIONS

The following is a list of suggestions to improve the functioning of

the entire design. Specific recommendations are given for each

section of the report. Overall improvements and observations follow

the section recommendations.

6.1

6.2

Bit cutting loading into stepped auger

* A cooling system could be designed to transfer heat from

the bit. This heat reduction would allow for a greater RPM

to be used during drilling.

* Wear testing of the bit portion of the drill should be

performed to determine and optimize cutter life, rake angle,

and bit orientation.

Rods, Auger, and stepped auger

* Throughput and volume retention testing must be

performed to evaluate the auger's capabilities. Pitch

variations should be used to optimize the design.

* Fracture testing of the device should be performed under

extreme lunar conditions.

* The AED should be redesigned to be compatible with the rod

changer. Presently, the AED will not work as designed.

* The rod to motor magnetic connection should be redesigned.

The design is not functional.

187
6.3 Footplate

* The design appears to be fully functional. A mechanical

model should be built to analyze the footplate performance.

* Strength calculations and testing should be performed on

the footplate.

* The storage procedure for the footplate should be reevaluated.

6.4 Power source, rotary drive, and Heat Management

* The power of the motor has to be increased. The current

power rating does not take into account non-ideal drilling

conditions such as drilling uneven surfaces. Also, the effect

of power dissipation due to friction from the drill string and

lunar core has not yet been considered.

* Fuel cell testing should be performed to evaluate

performance, especially at sub-zero temperatures. Possible

freezing of the water could present a problem.

Modifications will have to be made if freezing occurs.

* A refueling interface and a conversion location for the fuel

waste must be designed.

6.5 Structure and Mechanical Interface with walker

* A new structural design should be considered. The

construction of the present clesign does not appear to be

possible due to the fastening method which is used.
6.6 Rod Changer and Storage racks

* The disk which provides the rotational movement of the

arm does not provide a method to stop the rotation with any

degree of accuracy. A device to accomplish this task must

be designed:

* The model of the rod changer should be built, and a stress

analysis should be performed at extreme operating

temperature and conditions.

* The storage racks should be tested in conjunction with the

rods, drill, auger, and core retrieval devices to determine the

functionability. Also, the crushing capabilities that the racks

supply should be tested.

6.7 Macro-core Retrieval tools and Storage

* Extensive testing must be performed to optimize the shape

and size of the cutter. Performance should be rated on size,

life, and overall cutting ability.

* The cutter should be placed on the centerline with the

spring to prevent binding.

6.8 Controls for Operation and Console

* Components such as actuators need to be specified and

designed where needed in the project.

* The procedures for the following possible failures must be

addressed:

- Bit jamming during drilling
- Rod drops

- Bit breaks

- Rod breaks

- Drill string falls into the drill hole

- Actuator fails

- Power fluctuations

6.9 Vertical Accelerator for stepped auger

* The accelerator model should be built and tested under

lunar conditions (1/6th gravity). The design appears to be

functional.

6.10

6.11

* A method to keep the dirt from entering the device as the

rod moves up and down should be investigated.

* A bearing study for the cam should be performed.

Micro-coring Drill

* A heat transfer analysis should be performed on the bit.

* Detailed control and actuation for the micro-core needs to be

developed.

* An optimum drilling

determined.

angle should be experimentally

Overall Recommendations

* The interaction (interconnection) among the different

components of the lunar drilling device should be studied.

I _f-6
Once component modifications have been made, a model

should be built to study this interaction.

* A detached study of materials for each component should be

performed. Optimization of materials should include

strength-to-weight ratios, overall strength, reliability, and

life. Studies should consider the extreme temperature

conditions.
7.0. References

7.1. Acknowledgements

We would like to thank the following individuals for there help and

support in the completion of this project;

Mr. Bob Delauder of Brainerd-Killman, Inc.

Mr. Parks C. Stewart of Factory Automation Controls, Inc.

Mr. Anthony M. Aretakis of Whittaker-Yardney Power Systems

Mr. Gustav M. Geil of Micro Mo Electronics, Inc.

Mr. Dave Ross, VA Hospital Research

Dr. Kent Davey, EE Dept., Georgia Tech

Dr. Gene Colwell, ME Dept., Georgia Tech

Dr. Moussad Tavakoli, ME Dept., Georgia Tech

Mr. Dan Gibson, Good Tools

Dr. Robert Bachus, CE Dept., Georgia Tech

Dr. James S. Lai, CE Dept., Georgia Tech

Mr. Joe Chudecki of the Cerbec Corp.

Mr. Jim Kenedy of the Torrington Corp.

Dr. Johnson, ME Dept., Georgia Tech

Dr. Ferri, ME Dept., Georgia Tech

Dr. Black, ME Dept., Georgia Tech

Dr. Desai, ME Dept., Georgia Tech

Mr. Jaime Edwards

Dr. Mike Kelly of GTRI, Atlanta, Ga.

Mrs. Shannon Wolfe

Ack Electronics of Atlanta, Ga.

Franz Blam Racing of Atlanta, Ga.

Ms. Jill Harvey

and, of course, Mr. J.W. Brazell
7,2.

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Crouch, D.S., Lunar Rock Coring Device Design Study - Final Report,

NASA Document 66N15364.

Hickson, P.J., Modifications to the Apollo Lunar Surface Drill, NASA

Document 78N79023.
7.3. Software

Autocad 9.02

Word Perfect, 1.0

Word Perfect, 5.0

Catia CAD software

Cadam CAD software

Versacad CAD software

McDraw II, Claris

Harvard Graphics 1.01

Cricket Draw

Microsoft Word, 3.01
APPENDIX A

SAMPLE CALCULATIONS
1

A.1 BIT

Power Requirements:

Definitions"

Formulas:

Calculations"

F = Downward force on bit

Fd = Drag force placed on tooth

l.t = Drag coefficient

r = Radial position

T = Y_Fdri

P= Tto

HP = (Tn)/63000

F = 250/54 = 4.629 lbs.

Fd= _F = (1.0)(4.629) = 4.629 lbs. = 20.6 N

T = 20.6(6)[0.025 + 0.03125 + 0.0375 + 0.04375 +

0.05 + 0.0281 + 0.0344 + 0.041 + 0.047]

= 20.6 N (6 starts)(0.3375)

= 41.7 N-m or 368.98 lbf - in

Power = To.) to = 300 rpm = 31.14 rad/se c

= 1.37 kw

Horsepower = (Tn)/63000

= 1.757 Hp

Load = 375 lbs = 1668.075 N

(4 starts) = 11.25 lbs per tooth

T = 67.544 N-m = 597.81 lbf-in

Power =2.2Kw Hp =2.85 Hp

Load = 500 lbs = 2224.1 N

(4 starts) = 15 lbs per tooth

T = 90.058 N-m = 797.085 lbf-in

Power=2.8 Kw Hp= 3.8 Hp
_ 2

Power Requirements due to friction against side of the hole:

bit diameter = 100 mm

hole diameter = 104 mm

clearance = 2 mm (all around)

downward force = 250 lbs

RPM = 300

Length of rod = 2 m

At Bit leaning against wall:

(1) Find force (fx) at the bottom of the bit by

summing the moments about the end of the

bit.

fx = 0.5 lbs = 2.224 N

B. Entire bit touching the wall

(1) Find one constant force of 4448.2 N

(2) Torque = Force x Radius

= 4448.2 x 0.05

= 222.4 N-m

(3) Add torque due to teeth.

(4) Find power and Hp.

Ptot = Tt.o

= (222.4 + 90.058)(31.4)

= 9.816 Kw

Hp = (Tn)/63000

= (2765.58)(300)/63000

= 13.169 Hp
Intermediate Holding Device:

Due to size restrictions:

L = 60 mm • L = centerline length of holding device

Wall thickness (t) = 1 mm

W = ro-ri- 2t = 50- 25 - 2(1) = 23 mm

w

Volume of solid rock cut (Vs):

V s = AT d " AT = area of cutting teeth contact

each revolution

d = depth cut

AT = X(492 -262 ) = 5419 mm 2

With an expansion factor of 3 (see section 4.1.3):

Cs = 3Vs = 3ATd " Cs = volume of cuttings

Maximum capacity of the holding device (vmax):

vmax = WLt = (23)(60)(18) = 24840 mm 3

t = depth of device

The volume of cuttings cannot exceed this value.

Cs = 3ATd < 24840

Solving for d:

d = 24840/(3AT) = 8280/5419 = (1.5 mm)(2) = 3 mm

Although the above is very small, it does not determine the

drilling depth restrictions since much of the cuttings will jump

directly onto the auger.
TABLE 4.1.1

POWER REQUIREMENT CHART

LOAD

RPM (NEWTONS)

300 1112.1

300 1334.5

300 1556.9

300 1779.3

30O 2001.7

300 2224.1

POWER (KW) HORSEPOWER

9.8

9.9

11.6

13.3

14.9

16.6

13.2

13.4

15.6

17.8

20.0

22.3
** TEMPERATURE REQUIREMENTS **

At steady state for a cylinder straightened out as a wall '

q=KA(T1-T2)

L

where" L = Length = 2m

A = Cross Sectional Area

= g/4(.052 - .0482) + x/4(.0272 - .0252) = .0024m2

K = Thermal Conductivity K @ 5200 K = 19.8 Wim-K

T2 = 00 C = 273.150 K

Q = Kf* W * y = 1 * 1112.05 * (31.4 * .05) = 1746.804 Watts

Q' = 1746.804 * .2 = 349.361 Watts

Also, 20% of the heat will remain in the cutting teeth.

** MODEL **

d

%

m

lu

t 1 _ t _ t t

Iql._ c,,-

: I ,q"/q m

/

2

** MODEL ** (At steady state)

(1) There will be 12 diamonds so treat

the model as 12 study areas.

(2) Therefore, 12 areas will dissipate

29.113 Watts each.

_4_
w_

STEADY STATE HEAT"

I Iq7 ,.-, J!

2q.113 WA_S

29.113 = x

2 1.97

x = 28.676 Watts

28.676 Watts = 40 ( .0024 ) [T1 - 273.15 ]

1.97

T1 = 861.62OK = 1091.24o F

A = _ (.052 - .0252) = .006m2

29.113 Watts = 159 ( .006 ) [ T2 - 861.62 ]

.0254

T2 = 862.3950 K = 1092.640 F

Even though this temperature is above 952o F, the bit should still

work because when calculating T2, steady state was assumed. In order to

reach steady state in this application, constant drilling must be done. This

drill bit, however, will never be run constantly due to the shaking process

that is needed to advance the cuttings up the auger. This will then lead to

a temperature well below the calculated 1092.640 F.
STRESS ANALYSIS

All failing modes have same order of magnitude for similar

calculations. Shown are those we estimated to be critical. We intend

that resulting safety factors are large enough to account for

discrepancies due to unknown information.

Mode 1

Description:

Normal stress in plane of smallest area in male interface.

Assumptions:

Curved edges taken as straight squared edges for area calculation.

Calculations:

a

_x

m

B

D

D

total load

normal area

1,228.93 N

(10 ram)(50 ram) + 2[0.5(5 ram)(50 mm)l

= 1.64 MPa

Safety factor:

F,S, m

Oy

ff

[in.it x

827 MPa

1.64 MPa

= 504.71

w

Mode 2

Description:

Shear stress in plane of smallest area at interface of both male and

female.
Assumptions"

Curved edges taken as straight squared edges for area calculation.

Calculations"

tt'u_ x

total load

parallel area

1,228.93 N

2(25 ram)(20 ram)

= 1.23 MPa

Safety factor"

F°S, n

"l:y

_rt_x

33o.8 MPa

1.23 MPa

= 269.18

Mode 3

Description"

Shear stress in circular plane around male and female rod interface

that intersects rod.

Assumptions"

None.
Calculations:

1_ ttmx

total load

parallel area around interface

1,228.93 N

[2x(25 mm)][(9/48) + 9 ram]

571.59 KPa

Safety factor"

F,_,

B

B

1;y

"[;max

330.8 MPa

571.79 MPa

578.74

Mode 4

Description"

Torsion in rod.

Assumptions:

None.

Calculations:

IT_X

Tr_x

J

(67.79 N-m)(49 ram)

(x/32)[(98 ram) 4- (96 ram) 4]

= 4.63 MPa
w -

Safety

F°S, m

B

m

factor"

"l:m_ x

330.8 MPa

4.63 MPa

71.39

Mode 5

Description:

Bending all

Assumptions:

None.

Calculations:

around top of rod, below the interface due to axial load.

O

rt'l_x

Safety

F°S. B

D

m

m

M¢

I

( 1,228.93 N)[( I mm)/(2_,48 mm)(48 mm)1(6.67 mm)

= 15.66 MPa

factor:

Oy

II

n_x

827 MPa

15.66 MPa

52.83

(I/12)(I mm)(10 mm) 3
Mode 6

Description:

Bending on one side of female interface.

Assumptions:

The cross-sectional area at the base of one side of the female

interface was used instead of the one perpendicular to the moment

arm. This approach is more conservative since the used area is

smaller resulting in a higher critical bending stress. Also, this area

was approximated as a parallelogram to simplify calculations.

Calculations:

M¢

a --- m

max I

m

m

(1,228.93 N) sin(11.31'),

[(25 mm)/sin(78.69*) + (5 ram) tan(11.31')].

(4.33 ram)

2(1/36)(IO mm)( 15 rnm)3_(I/12)(30 ram)(15 mm) 3

= 2.68 MPa

Safety factor:

F.S,

CIy

CI

I'l_X

827 MPa

2.68 MPa

= 308.58

m

?,
Mode 7

Description:

Deflection of one side of the female interface due to tension allowing

male interface to slip out.

Assumptions:

The cross-sectional area at the base of one side of the female

interface was used instead of the one perpendicular to the moment

arm. This approach is more conservative since the used area is

smaller resulting in a higher critical bending stress. Also, this area

was approximated as a parallelogram to simplify calculations.

Calculations:

YtIl_ X

F L3

3El

u

m

(1,228.93 N) sin(l 1.3i°),

[(25 ram)/sin(78.69 °)+ (5 ram) tan( 11.3 I°)]3

3( 128,000 MPa).

2( 1/36)(I0 ram)(15 mm)3_ (I/12)(30 ram)(15 ram) 3

= 1.13 mm

Safety factor:

Common length of a side of male-lemale interface

Ytt_X

5 ITirn

1.13 mm

= 4.42
m

Mode 8

Description:

Eccentric loading of one rod.

Assumptions"

None.

Calculations:

p _ total load

A _ rod cross-sectional area

k = sqrt[(I/16)[(98 mm)Z+ (96 ram) z]

= 34.30 mm

e= l

c= i

p Oy

A i + (e.c/k) sec[( i/Ic)sqrt[P/(4 A E)]

P

[(49 ram) 2- (48 ram)z]

827 IvIPa /mm z

1 + (2.04)secl(2.33)(10)'3sqrt(Ipl)

Iterating,

P : 82,895 N
Safety factor:

m

P

total load

82,895 N

1,228.93 N

67.45

Mode 9

Description:

Buckling of one rod.

Assumptions:

None.

Calculations"

C=I

(L/k)cr = sqrt[(2C_2E)/oy]

55.27

(L/k) = 2 m/34.30mm

= 5831

(L/k) < (L/k)cr

C_ 2E I

Per - L2

= 112.211<N

_II_ Euler column
Safety factor:

Per

total load

112.21 KN

1,228.93 N

= 91.31
APPENDIX A.2

ROD WEIGHT

Rod Volume

Rod Mass

Earth Weight

Moon Weight

WEIGHT CALCULATIONS

= Hollow Cylinder + Top and Bottom Interface

Support + Interface + Collar

= (7_/4)(9 8 mrn2-9 6mm2) (2000mm)

+ 2(r_/4)(98mm2)(2/3 )(10mm)

+ (_/4)(50mm2)(25mm)

+ (r_/4)(100mm2-9 8mm2)(1 00mm)

= 7.903x105mm 3 = 7.903x10-4m3

= 9V = (4500 kg/m3)(7.903x10-4m 3) = 3.6 kg

= 7.8 lb = 1.75 N

= 1.3 lb = 0.292 N

w

AUGER WEIGHT

Auger Volume

Auger Mass

Earth Weight

Moon Weight

= Sleeves + Tube + Helix + Interface + Collar

= 3(7_/4)(98mm2-9 6mm2)(100mm)

+ (r_/4)(5 0mm2-46 mm2) (2000mm)

+ Oz/4)(50mm2)(25mm)

+ (_/4)(100mm2-98mm2)(100mm)

+ (2000ram/1 0mm)(g/4)(100mm2-50mm2)_

(cos 45)(lmm)

= 9.42x105mm 3 = 9.42x10-4m3

= 9V = (4500 kg/m3)(9.42xlO-4m 3) = 4.24 kg

= 9.3 lb = 2.09 N

= 1.6 lb = 0.360 N
VRA WEIGHT

VRA Volume

VRA Mass

Earth Weight

Moon Weight

= Outer Tube + Inner Tube + Interface

+ 4 1/2 Windings + Windings + Steps + Collar

= (rU4)(98mm2-96mm2) (2000mm)

+ (_/4)(52mm2-50mm2)(2000mm)

+ (_/4)(50mm2)(25mm)

+4(_/4)(96mm2-52mm2)(1/cos 39)(lmm)

+ 2((2000/81)-l)(r_/4)(96mm2- 52m m2)(0.2mm)

(1/cos 39) + (2000/81)(6)(18.225)(22)(0.2)

+ (_/4)(100mm2-98mm2)(100mm)

= 1.11xl06mm 3 = 0.00111m 3

= pV = (0.00111m3)(4500 kg/m 3) = 5.0 kg

= 11 lb-- 2.47 N

= 1.83 lb = 0.411 N

WEIGHT OF SOIL CUTTINGS

Volume = (Volume Drilled)(Expansion Factor)

= (_/4)(100ram2-5 0mm2)(2000mm)(3)

= 4.5x107mm 3 = 0.045m 3

Density of Cuttings = Density of MLS-1 Simulant

= 0.108 lb/in 3 = 3000 kg/m 3

Mass of Cuttings from 2m Worth of Drilling = (.045m3)(3000 kg/m3)

= 134 kg

Moon Weight = 49 lb = 11.0 N
TOTAL STRING WEIGHT (LOADED AND ASSUMING NO AUGERS ARE

USED AS RODS)

Mass = 9(VRA Mass) + 16(Rod Mass) + Soil Mass

= 9(5.0 kg) + 16(3.6 kg) + 134 kg

= 237 kg

Earth Weight = 520 lb = 117 N

Moon Weight - 86.75 lb = 19.49 N

Safety factor allows for 1.3 moon g's vertical acceleration

(0.22 earth g's)

TOTAL SYSTEM WEIGHT ON EARTH

16 Rods at 7.3 lb each = 1.75 N each = 28 N total

9 VRA's at 11 lb each = 2.47 N each = 22.23 N total

3 Augers at 9.3 lb each = 2.09 N each = 6.27 N total

Total System Weight = 56.50 N

m
A.2 Failure due to Static Loads Calculations

FAILURE OF A ROD

Mode #1 Description: Rod yields in tension

Assumptions"

Tensile Force = 900 N (200 lb) worst case

Calculations"

Cross Sectional Area, A = g.(98 mm 2- 96 mm 2) = 305 mm 2

4

=F= 900N = 3 N =2.95 x 106N

A 305 mm 2 mm 2 m 2

Safety Factor, n = _y_ = 827 x 106 N/m2_= 280

c_ 2.95 x 106 N/m 2

Mode #2 Description: Rod yields in shear due to tension

Assumptions"

Tmax = 50 Ib-ft = 67.8 N-m = 6.78 x 104 N-mm (motor torque)

"Cy = 0.5Cyy = 414 MPa

Calculations:

J =/._.(984-96 4 ) = 7.17 x 105 mm 4

32

r=rmax = 98 mm - 49 mm

2
L

_max = (6.78 x 104N-m)(49 mm) = 4.63 N = 4.63 x 106 N

(7.17x 105 mm 4) mm 2 m 2

n = _S_zv-- = 414 x 106 N/m2 = 89.3

"Cmax 4.63 x 106 N/m 2

Mode #3 Description: Rod buckles as a column under compressive

force

Assumptions:

Both column ends pinned at interfaces (end constraint factor=4)

Force is purely axial

Calculations:

1 = 2000 mm = 2 m

E = 127 x 104 MPa = 1.27 x 106 N/mm2 (for titanium)

I = J/2 = 3.59 x 105 mm 4

Example:

Pcr =.4.g-2(1.27 x 106 N/mm2)(3.59 x 10 5 mm4_ = 4.5 x 106 N

(2000 mm)2

Example:

n = Pcr = 4,5 x 106.._NN = 5000

P 900 N

FAILURE OF AUGER

Mode #1 Description: Component yields in tension or compression

Assumptions: All axial forces are resisted by auger core alone. Worst

case axial force = 900 N
Calculations"

Cross section area = A = (re /4 )((50ram) 2 - (46mm) 2) = 302

o=F/A = (900N/302mnl 2) = 3 N/mm2 =3.10 2 N/m

eg. n = Oy/O = 827.106/3.10 6 = 280

mm 2

Mode #2 Description: Component yields in shear due to torsion

Assumptions: All torsional loads resisted by core alone, maximum

torque = 50 ft-lb - 67800 N/mm 2 ,'_y = 0.5(Oy) =414 MPa

Calculations:

"Cma x = Tr/J

J : (_/32) ((50mm) 4 + (46mm)4) = 174000 mm 4

rma x = 50mm

_max = (67800 N mm)(50 ram)/(174000

eg.

mm 4) = 19.5 N=

= 1.95 * 107 N/m 2

n: "Cy/Xy : (414.10 6N/m2)/(19.5 .106 N/m 2) = 21

/.:
_ Stepped Auger Volume Retention Calculations

surface of soil arc length = 18.225 (sin(51)/sin(24)) = 34.8 mm = L

projection of L in XY plane = Lsin(75) = 41.88

angular subtense of soil arc = (41.8/(_.96)).360 = 50 degrees

Y projction of soil surface = Lcos(75)= 90 mm

volume of empty space above soil and below horizontal =

(1/2)(49.9/360)(7:/4)(962-522)(9) = 3190 mm

_ volume of soil section per step =

((1/2)(49.9/360)(7:/4)(962-522)(90)-3190=28,700 mm3

number of steps = (2000/81)(2)(6) = 296

_r

total retained volume in steps = (296)*(28,700) =8.5.106 mm3

volume of core = (7t/4)(502)(2000) = 3.93,106 mm3

total volume that can be held by a VRA segment =

(8.5.106+ 3.93,106)mm 3 = 0.0124 mm 3
A.3. Footplate

pLACEMENT OF FOOTPLATE

Calculation for torque needed to overcome friction regolith. Assumed

coefficient of friction of dirt to be C = 200 lb/ft 2. This simulates tightly

packed soil on Earth. Torque calculation is as follows:

T = Area x C x Radius

where,

A = surface area of casing

r = radius of casing

L = length of casing

c = coeff, of friction

Area=2pirL = pidL

d = 120 mm = 4.724 in = 0.394 ft

L = 0.67 ft

r = 0.197 ft

A = pi(0.394 ft)(0.67 ft) = 0.83 ft 2

T = (0.83 ft2)(200 lb/ft2)(0.197 ft} = 33 ft-lb

Maximum torque available from the drive is 200 ft-lb.
A.3. Footplate

I. STRUCTURE

MASS CALCULATIONS

Part of Structure

Base

Fins

Casing (est.)

Adapter (est.)

Interface

Roller, Shaft

Supports (6)

(est.)

Volume

(3.14/4)(.55)2(.005)

4(.5)(.21)(.215)(.005)

(3.14/4)(.1302 .1202)(.3)

[(. 1252 -.1552)/2] (.1)(.873)

(6)(.01)(.025)(.04)

= 1.19 x 10 -3 m 3

= 4.52 x 10 -4 m 3

= 5.89 x 10 -4m 3

= 1.05 x 10 -4 m3

= 6.0 x 10 -5 m3

Material Removal from Base

Square Sections (2)(.13)(.05)(.005)

Triangular Sections (2)(.5)(.225)(.13)(.005)

Total Volume of Structure

Density of Aluminum-Lithium = 2.6 x 103 kg/m 3

Mass = p Vtot = 5.67 kg

II. PARTS

=- 6.5 x 10 -5 m3

=- 1.4_i x 10 -4 rn3-

= 2.18 x 10-3 m 3

w

Part Volume

Rollers (2)

Shaft

Lifters (2) (est.)

Spring (est.)

( 2)( 3.14/4)( .02 )2(. 24 )

(3.14/4)(.015)2(.24)

(2)(3.14/4)(.015)2(.086)

(.5)(.2)(.0015)

Total Volume of AI-Li Parts

= 1.51 x 10 -4 m 3

= 5.4 x 10 -5 m 3

= 3.0 x 10 -5 m 3

= 1.5 x 1O:_ rn_-

= 3.85x 10 -4 m 3
Mass = pVtot,parts = 2.6 x 103 kg/m 3 x Vtot = 1.00 kg

III. ADDITIQNAL PARTS

Part

Gear Reducer

Motor

Batteries

Transmitter

Relays

&

Mass

0.39 kg

2.18 oz. = 0.063 kg

(4 cells)(0.72 oz.)

2.88 oz. = 0.08 kg

1.0 lbf = 4.54 N

= 0.46 kg

= 0.39 kg

= 0.06 kg

= 0.08 kg

= 0.46 k_,

Total Mass of Additional Parts = 0.99 kg

Overall Total Mass (Sum of Parts I, II, IlI) = 7.67 kg

The weight on Earth will be Mass times an acceleration factor of 9.81 m/s 2.

Earth Weight = 7.67 x 9.81 = 75.243 N.

The acceleration constant on the lunar surface is (9.81 m/s2)/6 ;

aL = 1.65 m/s 2,

therefore, Lunar Weight = (7.67)(1.635) = 12.54 N.
_ A.3. Footplate

SPRING DEFLECTION

CALCULATION

160mm

cos-o= 160 mm/ 170mm

-o= 19.75:'

need deflection of 19.75 _

displacement of

clamp lever

= (170 mm) sin-o-= 57.45 mm

220 mm

exaggerated for

detail

19.75'

220 mm

cos + = a/h

find radius

r = 220 mm /cos70.25 _

r = 651.05 mm

y = FI / 48EI

y = 0.0383 m

I = 0.440 m

E = 64 X I0 N/m

l=bh/12

b = 0.108 mm

h = 0.003 mm (spring thickness)

deflection in center of

spring =38.30 mm

651.05 mm-612.75 mm = 38.30 mm
A.3. Footplate

DEFLECTION OF LIFTER AND LIFTER LENGTH CALCULATIONS

F = (41.95 N) / 2 = 20.975 N

L = 0.084 m (ideal case, without bending)

Ymax = FL3 / 3EI; E = 64 X 109 N/m 2

I = pi d 2 /64

d = 15 mm = 0.015 m

Ymax = (20.975 N) (0.084 m)3 (64) / 3 (64 X 109 N/m 2) (0.015 m) 4 pi

Ymax = 26.0 X 10-_..mm

1.) ideal

o5

k

75" ,,,,_

o

HOLE MoTo R

38.3 mm = deflection of spring required to open spring during drilling

2.) with bending

F_ 2_ x, IO-¢°m

tan -1 (38.3 mm/ 75mm) = 27.05 ° - rotation required from the motor

and the lift¢r length = 75 mm I cos 27.05 ° = 84,21 mm, where the arm
length is specified as the length from the centerline of the motor to the

centerline of the roller.
A.3. Footplate

MASS ANALYSIS OF ADAPTOR

Adaptor diameter = .098 m

Material : AILi, 19 = 2.6 x 103 kg/m 3

Length = 1.75 m

The adaptor is an annular rod with inner radius of 44 mm and outer

radius of 49 mm. The choice of an annular device was made to limit

the weight requirements for transportation to the lunar surface.

Volume = n(.0492 -.0442)(1.75) = .003 m3

therefore,

Mass = pV = (2.6 x 103)(0.003) = 6.65 kg
w

A.3. Footplate

ANNULAR TUBE ADAPTOR

The following calculations are a force analysis to determine the strength of

the annular adaptor rod.

Sy = 6.67 x 108 Pa (Yield stress of Aluminum- Lithium)

1: = 1:max p/C = Tp/J (shear stress)

J = (_c 4 -_b4)/2

Tmax = 200 ft-lb

For positioning of footplate : T = 33 ft-lb

c = 49 mm = 1.93 in.

Case 1: Let b = 29 mm = 1.14 in.

J = _(1.934 - 1.144)/2 = 21.79 2.65

J = 19.14 in 4

Xmax = Tc/J = 200(12)(1.93)/19.14

= 242.01 lbf/in 2 = 1.67 x 106 Pa

so, Xmax < Sy

therefore, this design is fine but it can still be reduced.

Volume = [_(.0492 - .0292)/2](1.75) = 8.6 x 10 -3 m3

Mass = pV = (2.6 x 103)(8.6 x 10 -3 )

= 22.36 kg

Case 2: Let b = 34 mm = 1.34 in.

J = 16.73 in 4

"_max = 276.87 lbf/in 2

= 1.91 xl06pa

Zmax < Sy

Volume = 6.85 x 10 -3 m 3

Mass = 17.81 kg

Case 3: Let b = 44mm= 1.73 in

J = 7.72 in 4
Xmax= 600 lbf/in2

= 4.14 x 106 Pa

"_max < Sy

Volume = 2.56 x 10 -3 m3

Mass = 6.65 kg

Case 3 will be used due to its mass value and strength requirement.
m.3. Footplate

YIELD STRESS CALCULATIONS FOR LIFTER AND SHAFT

1.) Maximum torsion applied to the arm link at the motor hook-up is:

T = (41.95 N) / (0.075 m) = 3.15 Nm

= Tr/J; J = (pi)d 2 / 32 = polar area moment of inertia

= (3.15 Nm) (0.015/2 m) (32) / pi (0.015 m) 4

=4,7_5 X 106N/m 2

Since oy = 6.27 X 108 N/m 2, the link is not in danger of failing and

the factor of safety = 6.27 x 108 / 4.75 X 106 = 132

w

2.) Maximum force applied (pure bending) to each lifter is:

F = (41.95 N)/ 2 = 20.975 N

and

o-= Mc/I; I = J /2 = pi d2 / 64 = rectangular moment of inertia

c = (0.015 m) / 2

M = Fr, where r = 0.075 m.

o-= (20.975 N) (0.075 m) (0.015/2 m) (64) / pi (0.015 m) 4

o= 4.7_ X 1O6N/m 2

and c_ > o;, so the lifters are not in danger of failing.

The factor of safety is o- /o"= 132.

y
A.3. Footplate

OPEN SPRING FORCE AND YIELD STRESS CALCULATIONS

Sorina Force Calculation:

Spring Thickness = 1.5 mm = 0.0015 m

From previous page, Ymax = 38.3 mm, and Fma x = 48EIymax/L3.

E = 64 X 109 N/m 2 and I = bh3/12.

I = (0.108 m) (0.0015 m)3/12 = 3.0375 X 10 -11 m 4

Fma x = (48) (64 X109 N/m 2) (0.0383 m) (3.0375 X 10 -11 m4)/ (0.440 m) 3

= 41.95 N

Fmax, per arm = (41.95 N_/2 = 20.975 N. per arm

Tma x = (41.95 N) (0.075 m) = 3.15 Nm = 445.5 Qz-in

Yield Stress Calculation:

cr = Mc/I M -- (41.95 N/2) (0.220 m) = 4.615 Nm

c = (0.0015 m) /2 = 0.00075 m

I = 3.0375 X 10 -11 m 4

o-= (4,615 Nm) (0.00075 m) [ (3.0375 X 10-J!3-m4-2 = 1.140 X 101_-N/m2

and Cry = 6.27 X 108 N/m 2. % > o-so the spring is not in danger of failing.

The factor of safety is:

o-y/o-= 6,27 /1,140 = 5.5
A.3. Footplate

CLOSED . SPRING STRESS CALCULATIONS

From the rod group, the worst case maximum weight is 70 lb (318.2 N). We

will assume the system to be a simple cantilever:

CS

m ,,

) _///,/_ O. O0._m

i

o" = Mc/I

M = (159.1 N)(0.005 m) = 0.795 Nm

c = (0.0015 m)/2 = 0.00075 m

I = 3.0375 x 10 -11 m 4 (previous page)

(3" = (0.795 Nm)(0.00075 m)/(3.0375 x 10 -11 m 4)

cr = 0.196 x 108 N/m2

and O'yield = 6.27 x 108 N/m 2

Factor of Safety = (ry/Or = 32
A.3. Footplate

BALANCE CALCULATIONS

The unsymmetrical parts on the plate must not create a mass-moment

about the center. The only movable parts are the batteries and the

transmitter and relays box. Calculations for the combination that seems to

be the best of several are shown below •

J C

. Motor,Gear,Shaft,ShMt Supports

Rela , \ /

_ Lif%rs (32.03,13) /

-i __ShaftSupports' I i

(75,o) /

Gear (75,-137.5) /

i_ / Battery Pack

/ ( 1 15,- 150) Honor /

(75,- 1e 1.6)/

-j

The moments created by the non-moving parts :

Part Mass(k_)x Moment arm rel. to "0" (mm)

Gear 0.39 x 75i-(100+10+55/2)j -75i-137.5j

Motor 0.06 x 75i-(100+10+55+33/2)j =75i-181.6j

Shaft 0.11 x 75i+0j =75i+0j

Lifters 2(.04) x (75-85.95/2)i+0j =32.03i+0j

Supts. 2(.02) x 75i+ 0j =75i+0j

-- M0m.(kg-mm)

= 53.35i + 29.1j

= 11.44i + 4,73j

= 0i + 8.25j

= 0i + 2.56j

= 0i + 2.93j

Total Moment of Non-moving Parts - 64.79i + 47.56j kg-mm.
The moments created by the batteries and the transmitter and relays must

equal the total of (-64.79i 47.56j) kg-mm for the structure to balance.

The battery will be placed at location (115, -150) near the gear and motor.

Its moment is •

64.79i + 47.56j

12,2_i + 9,40i

77.05i + 56.96j kg-mm

or, -77.05i 56.96j kg-mm

= M x r= 0.46 kg k x r

r = (moment j /mass)i

r TO TRANSMITTER = (-123.02i + 166.41j) mm
A.3. Footplate

MOTOR POWER REOUIREMENTS

The motion of raising the spring will occur in one second.

traverse an arc of 29.240.

The lifter will

29.24o x 1 rev x 60 _e¢ = 4.87 rpm at gearbox output

1 sec 360 ° 1 min

Since the maximum torque required is 444.5 oz-in, and occurs at maximum

deflection, we will cut this figure in half to get an average, continuous

torque for the means of calculating input power requirements.

Tavg,cont = 444.5 oz-in/2 = 222.75 oz-in

With a reduction of 1,080 and a gearbox mechanical efficiency of 55%, the

motor output torque is:

Tmotor = 222.7_ 9z-in = 0.375 oz-in

(1080)(0.55)

According to the Micro Mo Catalog, torque in oz-in mutiplied by rpm and a

conversion factor of 0.00074 gives mechanical power in watts:

Pmech = (0.375 oz-in)(4.87 rpm)(1080)(0.00074)

Pmech = 1.46 W

And with a motor efficiency of 80%:

Pmech = 1.82 W

Since we are dealing in DC voltages:

I = P/V Supply Voltage is 6 VDC

I = 1.82 W/6 VDC = 0.303 Amps/raise motion
5000 cycles will be required at 1 second per cycle. This gives battery

power as:

(0.303 A)(5000)(1 s/cycle)(1 hr/3600 s) = 0.4213 Amp hrs

It is assumed that the power required to operate the transmitter and

relays is relatively negligable. Four Whittaker - Yardney #LR05 Silvercells

connected in series will provide a total of 0.5 Amp hrs. At 6.0 V (1.5 V

each), these cells are for long life and low rate use. A recharge will not be

required in 50 holes.

The total mass of four cells is 2.88 oz or 0.82 kg.

The four cells will occupy a space of 2.33 in2 (1505 mm2) on the plate.

Cell base width = 0.54 in

Cell base length = 1.08 in

Cell height (inc. terminals) - 1.56 in = 39.6 mm

(4 cells)(0.54 in)(1.08 in) = 2.33 in2 = 1505 mm2
A.4 POWER SOURCE

4.4.5. Friction Clutch.

These calculations were used to determine the required

diameter for the motor shaft. Also, see Figure 4.4.5.2.

From Mohr's Circle...

where:

Tmax = ( (0x/2)'2 + (Txy)"2 )'.5

Tmax = 16(M'2 + T'2)^.5 / pi(d)'3

M = moment = 0 (assumed)

T = torque = 83.3 ft-lbs (from Torq-Tender specifications)

Tmax = 16(83.3)/pi(d)'3

from Maximum Shear Stress Theory, Ssy = Sy/2

AISI 1045 Steel - Sy = 276 kpsi

with a factor of safety of 2

d = .42"

The d dictated by the maximum Torque-Tender shaft diameter to

accommodate the 3.94" interfacing with the drill string is 1.75". This

diameter is large enough. See Figure 4.4.5.2 for the shaft interface

and Figure 4.4.5.3 for the Torq-Tender specifications.
4.4.3 TEMPERATURE MANAGEMENT CALCULATIONS

Heat Pipe Wall Thickness Calculation

At 212F, the water vapor pressure is 180 psi and the Copper

Ultimate Tensile Strength is 18 kpsi. The maximum allowable stress

on the container walls is one-quarter of the ultimate tensile stress.

The maximum pressure stress is determined by:

do --

fmax = P (do 2 + di 2) / (do 2 - di 2)

do 2 = [(-P/frnax - 1)/(P/fmax - 1)] di 2

[(-180/1.8 x 103 -1)/80/(1.8 x 103-1)] (0.10) 2

do = 0.1.25 inches

tl = 0.1.25 - 0.1.12

tl = 0.13 inches

End Cap Calculation

For the most common circular heat pipes, a flat circular end cap is

used. Assuming a thickness of less thaT1 10%, the wall thickness of

the end cap can be determined by:

fmax = P do/ 4t2

fmax = 1/4 (UTS) = 1/4(18 kpsi) = 4.5 kpsi

4.5 kpsi = 2(180 psi)(0.1.25 inch)/4t2

t2= 0.06 inches

A factor of safety of 2 was used.

Diameter of Vapor Core Calculation

dv [(20 Qmax) / (1-Io, _-I'¢, RvTv)l/2)]l/2

dv =

20(4.3 E6) 1.33)(462)(373.15) (3.1415)(2.254 E6)(0.58)

dv = 0.10 inches

\

\
Properties of Water Vapor at 212F and Nomenclature

pv = vapor density (3.62 E-2 lbm/ft 3)

= latent heat of evaporization (754.3 E3 ft-lbf/lbm)

_'v = vapor specific heat ratio (1.33)

Rv = vapor gas constant (85.78 ft-lbf/lbm-F)

Tv = operating temperature

fmax = maximum hoop stress in the wall

P = inside pressure of heat pipe

do = tube outside diameter

di

dv

tl

t2

Aw

Av

Pl

_tz

Qmax = maxmimum dissipated heat per heat pipe

= tube inside diameter

= vapor core diameter

= cylindrical thickness

= end cap thickness

= wick cross sectional area

= vapor core cross sectional area

= surface tension coefficient (0.8521 lbf/ft)

= liquid density (59.97 lbm/ft 3)

= liquid viscosity (0.682 lbm/ft-hr)

Qe = entrainment limitation

Qb = boiling limitation

Qc = capillary limitation

rc = screen wick capillary radius

tw = screen wick thickness

Le = length of evaporater section

La = length of adiatbatic section

Lc = length of condenser section

N = wire screen mesh number

d = wire screen wire diameter

Fv = vapor frictional coefficient

F1 = liquid frictional coefficient

Pcrn = maximum capillary pressure
Capillary Limitation Calculation

The capillary limitation on the heat transport factor was

determined to be"

(QL)c,rnax =

Pcm

F Fv

where

Pcrn = maximum capillary pressure

F1 = liquid frictional coefficient

F,, = vapor frictional coefficient

Using Figure 4.4.3.4 below with a Mesh number of 250 inch t

at the operating temperature of 67(11,_ (_ 2F), the maximum capillary

pressure is:

P,'m --47 Ibf;q'

120

100 Fluid: W.ter --

Mesh numbs, N in "I

!6o "_ ",,,.

,_ .2oo_ _,,,,,_ _.

E 150 ___

•_ 100 _........_......__ _ ..-_._

50_

0

4,00 5OO 6OO

---=.==.......=,_ ,,.==,.._ ,.,..=_,=

700 80O 9OO 1000

Teml:mr|ture, T R

Figure 4.4.3.4 Maximum capillary pressure for screen wicks
Figure 4.4.3.5 shows the vapor frictional coefficient F,, at a vapor core

diameter of 0.10 inches to be:

Fv = 2 x 10"4 (lbf/ft 2) / (Btu-ft/hr)

Qc.max = (QL)c,max / [0.5(Lc) + La + 0.5(Le)]

700 = (QL)c,max/[(0.5)(O.1.55) + (t.50) + (0.5)(1.43)]

(QL)c.max = 2076 Watt-meters

Pcm

(QL)c,max = F L " Fv

Fl= 0.0022 (lbf/ft2_ / (Btu-ft/hr)

• 10"4

_- ", \ Fluid: l W,,,tl¢ l

! I. •

\ \

,,., \ k

..- \ \ , .,

> N "

\\ 1

1G"

0 0.2 0.4 0.6 0.8 .I .0

VNsar.core di_neter, dv In

1.2

Figure 4.4.3.5 Vapor frictional coefficient for round tubes
By use of Figure 4.4.3.6, with a mesh number of 250 inch-1 at

670R (212F), a wick diameter d of 0.0016 is obtained. The liquid

frictional coefficient can be determined:

FlAw =1.5 x 10 .5 Ibf (Btu-ft/hr)

Fi = 0.0022 (Ibf/ft -_t / (Btu-ft/hr)

A,,, = FIA,v/Fp

Aw = 1.5 x 1(I-5 / 0.0022

Aw = 0.0068 ft 2 =- 0.9818 inch 2

$

! \ , x \

| "-.,

I0

-_=_ ....,.._ o,0o_6--'_

I ,; J, o 6__

_ .L_ 50 0.0130

t J

700 800 900 tO00 1100 1700

Temoe,rttuce, T R

Figure 4.4.3.6 Liquid frictiontll c_wfficient for round tubes
Entrainment Limitation Calculation

The entrainment on the axial heat flux is determined by:

Qe,max = _.Av [(orv) / (2rh,s)] 112

Figure 4.4.3.7 below shows a plot of the above relationship and

is used to determine the value of Qc.ma,,/Av for the designed pipe.

The operating temperature and wire mesh number are used as

parameters. The value of Q/A,, for the designed heat pipe is 1.78 x

105 Btu/hr-in 2. At 670R (212F), a mesh number of 250 and a screen

wire diameter of 0.0016 inches will yield a Qe.max/Av of 2 x 105

Btu/hr-in 2

Av = F1 d,,2/4

Av= FI (0.075)2/4

Av= 0.0044 inch 2

Q/A_ = 746/().()()44

Q/Av = 1.7 x 1115 Btu/hr-in 2

2

%

£

u"

C

e"

ek_

10

a

$

?

$

s

4

10"

i

500 600 700 800 900 1000 11O0 1200

Temperature, T R

Figure 4.4.3.7 Entrainment limil on :txial heat flux density
Boiling Limitation Calculation

The boiling limitation can be represented by"

Qb.max = [(2FI LekeTv)(2o/rn)[ / l(2.0,, ln(ri/rv))]

Figure 4.4.3.8 below shows a plot of the above relationship to

determine Qb,max/Le for the designed pipe. Using the value of 0.896

for di/doas the controlling parameter for the plot shows a boiling

limitation of 1.5 x 104 Btu/ft-hr. Tile operating value of Q/Le for this

pipe is:

Q/L,: = 2362.4/(61/I 2)

Q/L.c=464.75 Blu/ft-hr

w

lO"

di/d"

7

6 1.11

s 1"

1.25 "-'-" _._,_

2'

-.=

_=_ et

1_ 71

ii 61

.J

4

d ,

g

'5 10_

1.50

H20/Cu Heat Pil_e,

\

1.75 _ \

'2.00 _ ',,.

\\\\ \

\N\ \ \

"i\\

l \\\

)

1(

400 500 600 700 800 900

Temperature, T R

Figure 4.4.3.8 Bc_ilin_..z,I.imitation

lOOO
&.5"

-Refer to figure 4.5.11.1"

;umming Moments about A"

I; Ma = 0 = Fb(.0826) - (142.5)(.1223)

F b = 210.9 N

Summing Forces about B"

_2 Fy = 0 = 210.9 - 142.5 - F a

F a = 68.5 N

'Moment of Inertia at smallest h"

bh^3 (.0095)(.0191)^3

b- 12 - 12 = 5"502e-9 m^4

_argest Moment:

--Mb = (68.5)(.0826) = 5.66 N m

_argest Tensile Stress:

(Mb)(C)

qb(flexural) - I -

Fb 210.9

_b(bearing) - Ab -

(5.66)(.0191)

5.502e-9 - 19.64 MPa = 2.85 kpsi

.00009 -

2.34 MPa

";u = 42 GPa >> 19.64 MPa ; therefore very good!!

This is obviously an over-design, but considering the fatigue loading

over the life of the latch and relatively small weight this is justified.
tENSION SPRING DESIGN:

Desired Spring Forces: 8 N when claw is open

5 N when claw is locked

Material: Music Wire (UNS G10850)

-Property Constants: m -- .146

A = 2170 MPa

G = 79.3 GPa

)esired Geometry: O.D. = 5 mm; d = .8 mm

D = O.D. - d = 4.2 mm

D

-3pring Index: C = --= 5.25 d

;hear Stress Mult. Factor:

.5

K s = 1 + "_" = 1.095

Fmax + Fmin

_,It. Force: Fa = 2 = 3.5 N

- Fmax - Fmi n

Mean Force: Fm - 2 = 1.5 N

CASE ONE: Static

8Fmax D

-.Max. Shear: Xmax = Ks _d 3 - 114.4 MPa

A

_Jlt. Strength: Sut - d m - 2.24 GPa

(ield Strength: Sy = .75Sut = 1.68 GPa

Jsing Dist.-Energy Theorem: Ssy = .577Sy = 970 MPa

Static Factor of Safety: i"1static = osy = 8.48

"Cmax

w_

CASE Two: Fatigue

For peened springs from pg. 436 Shigley, 5th edition.

I

Endurance Limit: Sse = KaKbKcSe ' = 465 MPa

Fa

Mt. Shear: Xa = Xmax lrmax

- 80.1 MPa

Zatigue Factor of Safety"

Sse

rl fatigue - "Ca

- 5.81
_.TQMPRE$$ION SPRING DESIGN"

Desired Spring Forces: 8 N when claw is open

5 N when claw is locked

-Material" Music Wire (UNS G10850)

_roperty Constants" m = .146

A - 2170 MPa

G = 79.3 GPa

Desired Geometry" O.D. = l0 ram; d = .9 mm

- AF

Spring Rate: k- AY - 600

M

Tree Length:

F1

Lf = L 1 +'k-'- = 38.3 mm

d4G

ND3- - 1.084e-5

8k

L2

"l'otal # Coils: N t < T = 25

"#Active Coils: N =Nt- Nd= 23 ; Nd -- # dead coils

ND3 3 3

-'Coil Diameter: D = (--"_-)" - 1.084e-5

--inside Diameter: I.D. = D - d = 6.9 mm

)utside Diameter: O.D. = D + d = 8.7 mm

Compressed Length" L s = N t d = 22.5 mm

Spring Index: C = D/d = 8.65

- .5

Shear- Stress Mult. Factor: K S = 1 +-C" = 1.06

Max. Force: Fma x = (L2 -Ls) k+F2 =9.5 N

- 8FmaxD

Max. Shear: Zmax = Ks red3 - 273 MPa

-'Ult. Strength:

A

Sut - d m - 2.2 GPa

Wield Strength: Sy = .75 Sut = 1.65 GPa
Using Dist.-Energy Theorem" Ssy = .577 Sy = 953 MPa

'_ > Ssy therefore it's O.K. to use the .9 mm music wire with both ends

- ground and squared.

_tatic Factor of Safety: q static = Ssv = 3.5

_max
A. 6 ROD CHANGER & STORAGE RACK

Calculations for the arm dimensions:

_w

Constants:

PAlum = 2.99 Mg

m3

E = 71 GPa

Ssy = 49,000 psi = 336 MPa

G = 26.2 GPa

with safety factor of 2.5: Ssy = 134.4 MPa

Lower arm

F

(cylindrical):

94.17 N + 10 N

MN

134.4

m2

= 0.775 ram2

Mc

I

I M 5.1 N m +18.06 N m + 116.7 N m

C S MN

134.4

11"12

= 1040.4 mm 3

I nd 3

- - 1040.4 mm 3

c 32

set

d = 21.97 mm

d = 25 mm

m m

nd 2

4

- 490.9 mm 2> 0.775 mm2

Vol. -

nld 2

4

- 500,691 mm3 W = pgV = 14.67 N = 2.44 N (moon)

Pin:

nd 4

d = 20 mm J32

- 15.71 x 10 -9 m 4

torsion

Tr (136.86 N m)(0.01 m) N

'_max='] ''= 15.71 x 10 -9 m 4 -89.03 x 10-6_-'_ "

deflection L = 20 mm

TL

Gj m

(136.86 N m)(0.02 m)

(26 2 x 10 9 _-_")(15 71 x 10 -9 m 4)

• m 2 •

= 0.007 °
Upper arm (two rectangular sections):

F 115.17 N

A=-= = 0.859 mme

c N

134.4 x 106 m2

I M 139.86 N m

-=--= = 1040.4 mm3

c _ N1

143.4 x 106 m2

set cross section 15 mm x 50 mm

A = 2 x 15 (50) = 1500 mm 2

Vol. = AL = 1500 mm2(500mm) = 750,000 ram3--7.5 x 10 -4 m3

W = pgV = 21.98 N = 3.66 N (moon)

Stress concentration:

d = 20 mm w = 50 mm h = 20 mm

A = (w - d)t = (50 mm - 20 mm) 15 mm = 450 mm a

F 120 N N

G0 A 450 mm 2 - 2.67 x 105 md 20 mm h 20 mm

- 0.4 - O.4

w-50 mm w-50 mm

d h

-- --on stress concentration graph. Look up w and w

Resulting Kt = 3.5

N N

_max = Kt_o = 3.5(2.67 x 105 m'-_'_)= 9.34 x 105 ,,m""T

Deflection:

FL 3

Upper arm:Xmax- 3EI-

(279.72 N)(0.5 m) 3

N

3(71 x 109 _-"5")(312,500 x 10 -12 m 4)

= 0.0005 m = 0.5 mm
FL

Ymax = EA

(115.17 N)(0.5 m)

(71 x 109m_)(0.0015 m 2)

=5.407 x 10 -7 m = 0.00054 mm

Lower arm:

FL3

Ymax (when horizontal) = _-_

(137.18 N)(1.02 m) 3

3(71 x 109 _--_)(312,500 x 10 -12 m 4)

m 2

= 0.0022 mm = 2.2 mm

u

FL

×max (when vertical) =

(104.17 N)(1.02 m)

(71 x 109_7)(0.00049 m 2)

= 3.054 x 10 .6 m = 0.00305 mm
ROTATING BAR CALCULATIONS

DIMENSIONS

W_ + Mc

- Smax = I

where:

Smax = maximum resultant stress = Ssy/1.5

a = distance from arm attachment to bearing attachment

W = the point load from the arm

M= the resulting moment from the arm

c= 1/2 the beam thickness

I = moment of inertia

bh 3

1-

12

-where"

w

therefore"

b=2c

h =2b

I= 16/12(c) 4

now"

!

where:

-substituting values"

c3_ 1.5 12(Wa + M)

Ssy 1 6

Ssy = 338 MPa

W = 115.17 N

M = 138.86 N

a = 0.425m

c 3 = 6.2x 10-7m

c = 8.5 x 10-2m

b = 1.7cm

h = 3.4cm
DEFLECTION:

FORCE D_ON:

_where:

- W 2 a 3

ymax = 6E--_I3a - L- )

E = 71 GPa

L = .5m

Note: all other values are the same as above

substituting:

Y max "-" (- 1/6)( 115.17/[(71 x 109)(6 x 108)] (3 (. 425 )2_.5_( .425 )3)

Ymax for force = 0.0052 m

MOMENT DEFLECTION:

Ymax = (Ma/EI)(L 1/2ta))

substituting:

-- Ymax = [(138.86)(.425)]/[(7 lx 109)(6x 10-8)11.5

Ymax for the moment = .0094 m

TOTAL DEFLECTION = 1.4cm

1/2(.425)]
STORAGE RACK

DIMEN$IQN$;

868mm

A "B C

D

500mm

-FORCE CALCULATIONS:

1 i 1

-- 5. N 45.5N

L

E F

10 rods x 43.6 N/rod = 436N

4 impl. x 76 N/impl. = 304N

Total = 740N

Each of the two support rods must withstand 370N.
B

4

i i

370N

To find the forces at A and B, sum of moments and forces

equations were used.

MA= B(.434m) sin(49)

B = 980.5N

370N(.868m) = 0

Fy = B sin(49) - Ay- 370 = 0

Ay = 370N

Fx = B cos(49) - Ax = 0

Ax = 643.3N

m

Rod BD is a two force member in compression. The force on this

rod is 980.5N. The forces in the members that counter the moment

-created by placing rods into the rack are are magnitudes smaller

(34N) than the forces of the implements themselves that they are

-negliable.

Stress Calculations;

The arc was approximated as a straight beam for the stress

calculations. The maximum bending stress will occur at the supports.

Mmax = 45.5(.28 + .42) + 76(.14)

Mmax=42.49 N-m
the shear stress is •

t_= 32(M)/rc(d) 3

= 32(42.49)/_(d)3

_ Set this shear stress equaton to the maximum shear stress

divided by the safety factor (2) and solve for the diameter (d).

S = Ssy/2

S = 316 MPa

316 = 32(42.49)/n(d) 3

d= 11 mm

_ The moment created by placing a rod into a spring clip is

14.28N-m is much less than the moment created by the weight of the

rods so the stress calculations have been neglected.

The maximum bending moment on the top beam of the support

-structure occurs at the center. And is calculated as follows

Mmax = 340(.434)

Mma x = 147.6 N-m

The stress was calculated using the same technique as above

and then d is solved for.

S = 32(147.6)/_(d) 3

Smax = 316 MPa

316 = 32(147.6)/_(d) 3 d = 16.8mm
- This beam is also in tension do to the angle of the support beam

below it so a tension calculation must be done. The force on the beam

-is divided by the area of the beam.

a = P/A

= 147.6/re(d)2

The ultimate tensile stress is then divided by the safety

-factor(2) and the result is the maximum allowable tensile stress.

Smax = 1461 MPa/2

Smax = 730.5 MPa

_ This value is then pluged in the equation above and the

diameter is solved for.

730.5 = 147.6/n(d) 2

d = lmm

-- The diameter is taken to be 16.8mm because it is the greatest.

the next diameter that is needed s the lower support beam. This

_beam is in compression and the greatest concern is with buckling.

The beam is assumed to be a Johnson column and this will be checked

later. The first calculation is for the constant 'b'.

b = (Sy/2n)2/CE

C end conditions constant

E - Young's Modulus

Sy - Yield strength

b = (1096/2n)2/(1.2)(206000) : 0.123
This is then pluged in to the Johnson equation

S = Sy b(1]k) 2

S = 1096 - .123( l/k) 2 k=d/4

The stress on this beam is then calculated by dividing the force

by the area of the beam.

S = 980/x(r) 2

S = 1248.4/d 2

Plug this into the equation above and solve for the diameter.

.001248.4/d 2 = 1096 (1.48/d 2)

d = 36.8mm

Deflection:

The curved beam was approximated to be a straight beam for

the deflection calculations. The deflection of the ends of the arc were

-calaculated as follows.

y = (1/12)(SL2)/(Ec)

S shear strength

L - length of the bar

E - Young's Modulus

c - distance to the center of the bar

y = (1/12)(316)(.42)2/(207000)(.011)

y = 2mm
u

The deflection at the cented of the beam is calculated using the

-same formula but the moment and the shear at the center must be

calculated first.

M = 45.5(.14+ .28 + .56 +.84 + .98) + 76(.42 + .7) - 370(.56)

M = 5.32 N-m

S = 32(5.32)/pi(.011) 3

S=40

y = (1/2)(316)(40)(1.12)2/(207000)(.011)

y = 1.8mm

The support structure is next

y = (1/12)(316)(.868)2/(20700)(.017)

y= 5.6mm

-Weight Calgulations:

Curved beam:

Top Beam:

Bottom Beam:

Volume

2.1 (_(.011 )2/4)

.868(_(.017)2/4)

.662(rc(.0368)2/4)

Weight = V(p)

- = (.00199)(2840)

= 5.65 kg (per section)

= .(1002 m 3

= 2(0.00019 m 3)

= 2(0.0007 m 3)

Total = .00199 m 3
A.7 MACROCORE RETRIEVAL

Mass of Core Sample

=-.==4)-

2m

50 ram

W

I

I

)

'91==.=== BasaLt

p=#

V = _ (r'2) n

m =pV

g = 9.8 m/s^2 (Earth)

W = m g (Earth)

W = m g /6 (Moon)

9 = 2.9 g/cm'3

V = 7t(5 cm)'2 (200 cm) = 15,708 cm'3

M = (2.9 g/cm'2)(15708 cm'3) = 45,553 g = 45.553 kg
A.7 MACROCORE RETRIEVAL

Mass of Retriever:

Material: Ti - 5AI - 2.5Sn

p -- 4.46 g/cm'3

Outer Shell:

V = x(r'2) L = _[(5 cm)'2 - (4.7 cm)'2](200cm)

- 1828.41 cm'3

M = p V = (4.46 glcm'3)(1828.41 cm'3)

= 8154,69 grams

Inner Shell:

V = ;t (r'2)L = x [(2.8 cm)'2 - (2.5 cm)'2](200cm)

= 999.03 cm'3

M = p V = (4,46 glcm'3)(999.03 cm'3)

= 4455.66 grams

Single Support Beam:

V = Lwh = (200 cm)(1.9 cm)(.2 cm) = 76 cmA3

M = p V = (4.46 g/cm'3)(76 cm'3)

= 338.96 grams

Single Mechanism (Model as Solid Block):

V = Lwh = (2.1 cm)(1.5 cm)(.2 cm) = 6.3 cm'3

M = p V = (4,46 glcm'3)(6.3 cm'3)

= 28.098 grams

Total Mass of Retriever:

= 8154.69 g + 4455.66 g + (3 x 338.96 g) + (3 x 28.098 g)

= 13,711.52 g

- 13.71 kg
A.7 MACROCORE RETRIEVAL

Simple Buckling of Outer Shell

P

Io = J/2 = 7rravg3X = _(4.85 cm)3(3cm) = 1075.21 cm3

Per = 7r2EIo/4L 2 = _2(7.586 x 107 Pa)(10.7521 x 10 -6 m4)/4(2 m)2

= 503.14 N
A.7 Macrocore Retrieval

Assumptions:

Cutting Force = 501bs

Cutting Stroke = 1.0cm

Width of spring = 1.0cm

Material is UNS461500

m=.167

A = 167kpsi

G=11.5 x 106psi

Ultimate and Shear Strength

Sut = A/d.167 Sy=.75(Sut) Ssy= .577Sy

= 169kpsi/d .167 Ssy = (.577)(.75)Sut

Ssy = 73.135kpsi/d.167

Diameter of wire(d) and spring(D)

Ssy= T = 8FD/_td 3 ( I +.5/C) C = D/d

D = 1.0/2.54 - d

C=(1.0/2.54 - d)/d

73.135/d.167 = [8(50)(l.6/2.54-d)/Ttd3] [l+.5(1.0/2.54-d)/d]

by iteration d = .07379 in. = .0019 m

D = 1.012.54 - d = 3.199 in. = .0813 m

Number of Coils in the Spring (Nt)

fully compressed N = Nt + Nc

Nt = (1.6/2.54)/d Squared and Grounded ends

Nt = 8.5 turns Nc = 2 turns

N = 6.5 turns
Minimum Forces during cutter stroke

K= d4G/8D3N = (.07379)4(11.5 x 106 )/8(.3199)(6.5)

K= 20.5 Ibslin. = 2.32 Nlm

force when gripper should touch

F= 501b. - K(1.0/2.54)

Fmin = 41.93 Ibs = 186.59 N

Conclusion

Diameter of coil (D)

Diameter of wire (d)

Number of coils (Nt)

Spring Constant (K)

Fmin = 41.93 Ibs =

Fmax = 50.0 Ibs =

,- 3.199 in. = .0813

= .0738 in. = .0019 m

= 8.5 turns

= 20.5 Ibs/in. = 2.32 N/m

186.59 N

222.5 N
A.7 Macrocore Retrieval

Reset Gear Desig_

Assumptions:

Tangential Force:

Pressure Angle:

Number of Teeth:

Pitch Diameter:

Wt -, 50 Ibs = 222.5 N

0 = 20°

N=18

d = 2 cm =.7874 in

Velocity Factor:

Assume V = I in/sec = 5 ft/min

Kv= 5O/(5O + V0.5)

= O.957

Y = 0.29327

(Hobbed Gears)

(From Shigley & Mitchell)

Face Width:

F = WtP/KvY(Sp = (50 * 22.86)/((0.957 * 0.29327) * (Sp)

3p = Face Widthmin = 0.412 in

5p = Face Widthmax = 0.687 in

(C_p: permissible bending stress = Sy/4

where 4 = safety factor)
A.7 Macrocore Retrieval

From Face Width Criterion:

SPmax = 9884.75 psi

SPmin = 5926.25 psi

Therefore, Sy range = 23705 psi to 39539 psi

= 163 MN/m'2 to 272 MN/mA2

Gear Properties:

N = 18 Teeth

P = 22.86

= 20 °

d = 0.7874 in = 2 cm
A.7 Macrocore Retrieval

_n..

70 Ibs

Contact Points:

5mm

2mm

3mm

20 Ibs

50 Ibs

I = _d4/64 = _(.3/2.54)4/64 = 9.5525 x I0 -6 in4

Mma x = 501bs x 5mm x (cm/10mm) x (in/2.54cm)

= 9.84in-lbs

(5 = Mc/I = (9.84in-lbs x (.15/2.54)in) / 6.552 X 10 -6 in 4

= 6.083 X 104 psi = 419 MN/m'2
A.7 Macrocore Retrieval

Breaking the Core Sample:

Using Tension:

Fmin = (St x A = (1kpsi)(3.043in 2) =

Fmax = C_t x A = (4kpsi)(3.043in 2) =

13.5kN

54.1kN

Using Torsion:

Tmin = (_ x _ x d3)/16 = (2kpsi)(_)(.05m)3/16 =

Tmax = (T x _ x d3)/16 = (8kpsi)(_)(.05m)3/16 =

338.2Nm

1352.9Nm

Using Bending:

Fmin = (5 x

Fmax = (5 x

I/L*c = (2kpsi)(3.07x10-4)/(2m)(.025m) = 84.6N

I/L*c = (8kpsi)(3.07x10-4)/(2m)(.025m) = 338N
A.7 Macrocore Retrieval

Table 4.7.6. I: FORCENEEDEDTO BREAK CORESAMPLE

USING TENSION: F = St *A

Core

Diameter

(mm)

5O

45

4O

35

30

25

20

15

I0

5

Area

(m'2)

1.96E-03

1.59E-03

1.26E-03

9.62E-04

Fmin

St = Ikpsi

(N)

13,528.52

10,958.10

8,658.25

6,628.97

Fmax

St = 4kpsi

(N)

54,114.06

43,832.39

34,633.00

26,515.89

7.07E-04

4.9 IE-04

3.14E-04

1.77E-04

7.85E-05

1.96E-05

4,870.27

3,382.13

2,164.56

1,217.57

541.14

135.29

19,481.06

13,528.52

8,658.25

4,870.27

2,164.56

541.14

Table 4.7.6.2: TORQUE NEEDED TO BREAK CORE SAMPLE

USING TORSION: T = Tau*Pi*d'3 / 16

Core

Diameter

(mm)

5O

45

40

35

d'3

(m'3)

1.25E-04

9.1 IE-05

6.40E-05

4.29E-05

Tmin

Tau = 2kpsi

(N-m)

338.21

246.56

173.16

116.01

Wmax

Tau = 8kpsi

(N-m)

1,352.85

986.23

692.66

464.03

30

25

20

15

2.70E-05

1.56E-05

8.00E-06

3.38E-06

73.O5

42.28

21.65

9.13

292.22

169.11

86.58

36.53
A.7 Macrocore Retrieval

Table 4.7.6.3: FORCE NEEDED TO BREAK CORE SAMPLE

USING BENDING: F = (Sb*l) / (L'c)

Core

Diameter

(mm)

50

45

40

35

30

25

Moment of

Inertia

(m'4)

3.07E-07

2.0 IE-07

1.26E-07

7.37E-08

3.98E-08

1.92E-08

Fmin

Sb= 2kpsi

(N)

84.5532

55.4754

34.6330

20.3012

10.9581

5.2846

Fmax

Sb = 8kpsi

(N)

338.2129

221.9015

138.5320

81.2049

43.8324

21.1383

20

15

I0

5

7.85E-09

2.49E-09

4.9 IE- I0

3.07E- I I

2.1646

0.6849

0.:353

0.0085

8.6582

2.7395

0.5411

0.0338

Page 2
A.7 Macrocore Retrieval

Failure K = Kc

Kc = 3 MN/ml.5 for Granite

K=a x(_xa)0.5

a = K / (x x a) 0.5 = (3 MN/m 2)/(rt x .001 m) = 53.52 MN/m 2

where a = .001m

A = 7t x r 2 = rt x (.025m) 2 = 1.96 x 10-3 m 2

where r = .025m

F=ax A = 54MN/m 2 x.00196m 2 = I05.8kN

where a = 54MN/m 2

Friction Force (Gripper)

F

P

P = F / _l.s = (I0 Ibs) / (0.I) = 100 Ibs = 445 N
A.7 Macrocore Retrieval

Material - Granite (Westerly)

Kc = 3 MN/mA(3/2)

K = Sigma * (pi*a)'(I/2)

Failure K = Kc

a

(mm)

i

2

3

4

5

Sigma

(MN/mA2)

53.52

37.85

3O.9O

26.76

23.94

A = pi*r'2

radius

(mm)

25

23

21

19

17

Area, A

(mm'2)

1,963.50

1,661.91

1,385.45

1,134.12

907.92

Area

(m'2)

1.96E-03

1.66E-03

1.39E-03

1.13E-03

9.08E-04

Page I
A.7 Macrocore Retrieval

Table 4.7.6.4: Force Calculation with Varying Stress Concentration

F = Sigma * Area

Force (MN)

Sigma

(MN/m'2)

54

38

3:

27

24

0.00196

0.10584

0.07448

0.06076

0.05292

0.04704

0.00166

0.08964

0.06308

0.05146

0.04482

O.O3984

Area (m'2

0.00139

0.07506

0.05282

0.04309

0.03753

0.03336

0.00::3

O.O6102

0.04294

0.03503

0.03051

0.02712

0.000908

0.049032

0.034504

0.028148

0.024516

0.02 1792

Page 2
A.7 Macrocore Retrieval

Friction Force (Gripper)

P = Coefficient of Friction * Force

P (Ibs)

Coef. of Fric.

0.I

0.2

0.3

0.4

0.5

Force (Ibs)

I0 Ii 13 15 17

I00.0 II0.0 130.0 150.0 170.0

50.0 55.0 65.0 75.0 85.0

33.3 36.7 43.3 50.0 56.7

25.0 27.5 32.5 37.5 42.5

20.0 22.0 26.0 30.0 34.0

P (N) Force (N)

Coef. of Fric.

0.I

0.2

0.3

0.4

0.5

44.5 48.95 57.85 66.75 75.65

445.0 489.5 578.5 667.5 756.5

222.5 244.8 289.3 333.8 378.3

148.3 163.2 192.8 222.5 252.2

I 11.3 122.4 144.6 166.9 189. I

89.0 97.9 i 15.7 133.5 151.3

* I pound = 4.45 Newtons

Page I
A.8 CONTROLS

There are no sample calculations available from the controls

group.
A.9 VERTICAL ACCELERATOR

Bearing Analysis

The following is the analysis of the maximum load on the cam

bearings:

Weight of full drill string = 245 N (on moon)

Acceleration of drill string (maximum) = 3277 mm/s 2

Maximum force to accelerate drill string

F=ma= 5N ' • 3277 m m

1646

Fmax = 487 N

Because the roller follower is being pushed at a maximum of a 30 deg.

pressure angle it actually reacts to a larger force.

487 N

Fmax - cos(30o) - 562 N

Maximum force on any one roller is then half of this total force.

Froller = 281 N

Load rating:

CR I_L-'_/nR _I/a.. : Froller * [_D))

Using an L10 life of 5000 hr and design rpm of 450:

LR = 3000 hr

nR = 500 rpm

Lo = 5000 hr

nD = 450 rpm

a = 10/3
CR = 281 N. ((3000) 00 _4_5-'_)) 5 0)?/10

CR = 317.5 N

The roller followers have a much higher load rating than the computed one

estimated by Torrington; so the roller follower life should last through

about 450 fifty meter holes, but we oversized the follower for about ten

times that amount. Since there are extreme operating temperatures on the

moon going as low as -158 oC at night, lubricants would just evaporate,

therefore, the bearings used are designed without the need for any

lubrication.

Shaft Analysis:

Diameterrnin = 25 mm

T max = 28.6 N-m

Moment = 0

S y = 827 MPa

n = ? (factor of safety)

Using a basic static failure analysis and solving for n (factory) gives,

(_a 0.8 6 6 T))I / 3 d= _ r_ _'_y + Sy

d3Sy_

n= 0.866(32)T

(1)3(827 MPa)(3.14)

n= 0.866(32)(28.6)

n = 645 Factor of safety
A.IO Microcore Retrieval

CAM DESIGN

Simple Harmonic Rise y = (1 - cos rt¢_t13) * L/2

Simple Harmonic Full Return y = (1 + cos r_/13) * L/2

13 = 150 L = 5 mm 0<=_<=360

Stations f y (mm)

0 0 0

1 30 .477

2 60 1.727

3 90 3.273

4 120 4.523

5 150 5.000

6 180 5.000

7 210 5.000

8 240 4.523

9 270 3.273

10 300 1.727

11 330 .477

12 360 0

y=2.5*(1-cos 1.2¢)

y =2.5 * (1 + cos 1.2_)

y (ram)

5

Lift vs. Rotational Angle

72 144 216 288 360
APPENDIX B

DESIGN ALTERNATIVES
B.1 BIT

There are no design alternatives available from the bit group.
B.2 Design Alternatives for Rod, Auger, anti Stepped Auger

B.2.1 Rod

A solid rod cross section was considered for strength and fracture

toughness but rejected because of weight considerations.

B.2.2 Rod Interface

A rod interface design which was identical for both top and bottom

ends of the rod was considered (see drawing in progress report). This

design was rejected because the collar required for the coupling could not

be coupled by the rod changer (see Fig.D.4.2.1).

B.2.3 Vertically Reciprocating Auger Design

Several pitch and fence height configurations were considered but

rejected to achieve a suitable balance between throughput and volume

retention.

B.2.4 Vertically Reciprocating Auger Interface

An interface with a continuous helix was considered l_ut rejected

because no suitable method for coupling could be arrived at given the

abilities of the rod changer. Ignoring the VRA interface by using only a

single VRA incorporated in the bit segment was considered but rejected

because the volume retention was much too low for 2 m of drilling.

B.2.5 Motor Interface

A motor interface which was identical to the rod interface but

utilized a collar which was actuated from above was considered. Although

rejected this design is probably more practical than the current MMI

design.
C I I

l.d

I-.4

O.0

,'L

LdHW

I-I H

I-.- Z0

n,

I.L I a ? I
B.3 FOOTPLATE

A different design utilizing a tool chuck (like those on power

drills) with rotary clock springs was also considered, but turned

down. The outer housing is fixed to the plate while the center of the

chuck rotates, opening and closing the chuck jaws on the drill string.

A series of clock springs, acting independent of each other would

provide opening and closing power for the chuck. Mechanical power

for energizing the springs would be provided by the drive at the

bottom of its stroke as it meshes with the rotating part of the chuck.

A spring is de-energized by the flip of a lever, but does not cause the

other springs to de-energize, hence the independence. The design

would be very complex and difficult to seal from the lunar

environment, but would have offered the advantage of not requiring

a battery and motor to provide mechanical power. The design of the

chuck, as it appears on the patent, is attached.
_ nited States Patent

rHetzen et id.

I i

[54l

75|

731

221

121l

631

[|91

TOOL CHUCK

Inventors: WIIImm IL _ Fayetteville.

N.C.; St_ J. nrlidt, Owings

MUls. Md.

Amiln_: Tim Meek m,d _

A_ Cemlmay, To,on.

Md.

Feed: _ 2, ]_3

Appl. No.: Xf_61$

F4mml US. APl_lttom Dim

Coatlmmliom of Set. No. 126.472. March 22. 1971.

52!

-_511

[S8)

US. Cl. ................................ 279/62; 279/I ME

lat. CLt .......................................... B23B 31/04

el _ ............. 279160. 61.62. 63.64.

2?9/65. I K. I ME

1561

|.162.164

-_JAIO.tM2

Itderwem Cltld

UNITF.D STATES PATENTS

8/1969 Willtsmam ............................. 279162

511974 IDerblrd_re ........................... 2"/9162

14

BO

I

ii l 3,949,998

1451 Apr. 13, 1976

FOREIGN PATENTS OR APPLICATIONS

435.216 5/1948 Italy ...................................... 279162

1.255.450 I 1/1967 Germany .............................. 279/62

Primary Examiner--Willie G. Abererombie

tLt_umt F.rami_r--Z. R. BiJimky

Attorney, Agent, or Firm--Joseph R. Slotnik; Edward

D. Murphy; Leonard Bloom

15"/] ABSTRACT

A tool chuck for drills and the like including a body

having gripping jaws slideably supported in tapered

ways formed in the body. A split nut mtatably sup.

ported on the body threadedly engages the jaws and is

opera>le to engage and release the jaws from a bit.

The nuL and body are interfitted _ rotatably support

and axially retain the nut on the body. A cup is fitted

on the nut to hold it together and on the body and, together with the interfitted body and nut, protects the

internal chuck parts from dirt and other foreign

material.

12

I

43

OF POOR QUALi'i'_,
B.4 POWER SUPPLY, ROTARY DRIVE. AND HEAT

MANAGEMENT

Design alternatives considered but discarded are listed below with a

brief reason for their rejection.

I) Rechargeable Batteries: Silver-zinc batteries were considered

for the application, but proved to be far too heavy for the

design,

2) On-board Solar Panels: The power requirements made this

alternative not feasible.

3) Gearing for the Motor: Gearing proved to be unnecessary since

a motor could be designed which would meet all of the design

requirements.
B.5 STRUCTURE AND MECHANICAL INTERFACE

There are no design alternatives available from the structure

and mechanical interface group.
B.6 ROD CHANGER AND STORAGE RACKS

The first mechanical gripper to be used as a proposal was a

gripper that had one jaw in a fixed position. The other parallel jaw

would move by means of a lead screw towards the fixed jaw. This

type of gripper could accommodate both size diameters of digging

implements, 50 mm and I00 mm. The problem this design created

was that these different shaped digging implements could not both

be held on the center line of the gripper. The robot arm will have a

fixed position to travel to for each individual clip. This position

would not be the correct position for the rod storage if the rod was

off center while in the grasp of the gripper. A control problem would

arise if for every different size rod, the robot arm position would

change. The design we chose held the different size rods on the

center line of the gripper.

Another gripper design had parallel jaws with two lead screws.

The gripper would be powered by a motor turning these righthanded and left-handed threaded lead screws. When the lead

screws turned the two female threaded jaws, the rotating motion

would be translated into linear motion. The two parallel jaws would

approach each other in a clamping motion. This design was not

chosen because of the fact that it has two lead screws. Our simplified

design only has one lead screw. The width of this gripper design is

wider than the chosen design. The smaller design in width will

enable the rack to hold more rods. This is because the gripper

requires clearance to move in between the rods.

Another gripper design used a fixed body and a rotating arm.

The gripper had a U shape body that was stationary. The U shape

had a radius of 50 mm to hold the I00 mm digging implements. The

rotating arm was driven by gears. This gripper needed a curved slot

cut into it for the 50 mm rods. This slot would position the smaller

rods so that they were on the center line of the gripper. This design

was not used because the chosen design was less complex. This

gripper design would also drop whatever it was carrying because the

gears would turn, reducing the force on the gripped object if there
was a loss of power to the motor. This design would not be a nonoverrunning system.

These gripper designs originally had sensors to detect the force

exerted by the gripper on the rods. A simpler way of measuring the

amperage required by the gripper to detect when the gripper was

closed with the proper force was used.

As stated in the materials section titanium, particularly the Ti4V-6AL alloy was considered as a structural material. It has similiar

properties to the aluminum alloy chosen. However it is heavier, and

costs quite a bit more than aluminum. The titanium alloys are

difficult to weld and are not readily available in many forms. The

properties of the aluminum are sufficient for our design.

Two clip designs were considered, both based on clips made by

Gibson Good Tools. The first design considered was a u-shape where

each side bends back on itself. This design was refected because of

excess material and the design used was easily approximated as a

cantilever beam, thus simplifying the calculations required. Also the

accepted design more easily accomodates for drilling implements of

differing sizes.
A. Water Jet

B. Thermal Shock

C. ElectricArc

D. Laser

E. Acid

F. Tension

G. Torsion

H. Bending

I. Impact

_.Tension _

F1. Chinese Handcuffs

F2. Locking Ring

F3. Screw-type Ring

F4. Pivoting Rings

F5. Post-holeDigger

F6: Cams

Bending __ Wedge

• Post-hole Digger

with Bending

J. Thermal Expansion "-/-Sc°ring

/

K. Scoring

Cutting

L. Cutting

M. Blasting

"L--K 1. Pipe-Cutter

Mechanism

f L1. Diamond Teeth

N. Sonic Vibrations

O. Melting

P. Crushing
B.6 DESIGN ALTERNATIVES

Gripper

The first mechanical gripper proposed was a gripper that had

one jaw in a fixed position. The other parallel jaw would move by

means of a lead screw towards the fixed jaw. This type of gripper

could accommodate both size diameters of digging implements, 50

mm and 100 mm. The problem this design created was that these

different shaped digging implements could not both be held on the

center line of the gripper. The robot arm will have a fixed position to

travel to for each individual clip. This position would not be the

correct position for the rod storage if the rod was off center while in

the grasp of the gripper. A control problem would arise if for every

different size rod, the robot arm position would change. The design

we chose held the different size rods on the center line of the

gripper.

Another gripper design had parallel jaws with two lead screws.

The gripper would be powered by a motor turning these righthanded and left-handed threaded lead screws. When the lead

screws turned the two female threaded jaws, the rotating motion

would be translated into linear motion. The two parallel jaws would

approach each other in a clamping motion. This design was not

chosen because of the fact that it has two lead screws. Our simplified

design only has one lead screw. The width of this gripper design is

wider than the chosen design. The smaller design in width will

enable the rack to hold more rods. This is because the gripper

requires clearance to move in between the rods.

Another gripper design used a fixed body and a rotating arm.

The gripper had a U shape body that was stationary. The U shape

had a radius of 50 mm to hold the 100 mm digging implements. The

rotating arm was driven by gears. This gripper needed a curved slot

cut into it for the 50 mm rods. This slot would position the smaller

rods so that they were on the center line of the gripper. This design

was not used because the chosen design was less complex. This

gripper design would also drop whatever it was carrying because the
gears would turn, reducing the force on the gripped object if there

was a loss of power to the motor. This design would not be a nonoverrunning system.

These gripper designs originally had sensors to detect the force

exerted by the gripper on the rods. A simpler way of detecting this

was to measure the amperage required by the gripper to close with

the proper force.

Materials

As stated in the materials section titanium, particularly the Ti4V-6AL alloy, was considered as a structural material. It has

properties similar to the aluminum alloy which was chosen for the

final design. The titanium alloy was not selected for many reasons.

It is heavier, and it costs quite a bit more than the aluminum. Also,

titanium alloys are difficult to weld and are not readily available in

many forms. The properties of the chosen aluminum alloy are more

than sufficient for our design criteria.

Clips

Two clip designs were considered, both based on clips made by

Gibson Good Tools. The first design considered was a U-shaped clip

where each side bends back on itself. This design was rejected

because of excess material and the design used was easily

approximated as a cantilever beam thus simplifying the calculations

required. Also the accepted design more easily accommodates for

drilling implements of differing sizes.

Rack

A first design for the rack was a free-standing rack. It held the

rods in a cylindrical holder. This design resembled a vertical

baseball bat holder. The holder was going to sit under SKITTER to

permit easy access for the robot arm. This design became unfeasible

when SKITTER would have to squat down. The cylindrical holder

was then going to be placed out from under SKITTER. If the holder

was placed away from SKITTER, the rod changer would need another
degree of freedom. Either the mechanical gripper would need a wrist

of the robot arm would need a third section.

Another design for the rack involved a revolving pulley system

between the legs of SKITTER. This design would allow twice as many

clips as the design of a fix rack. This design was decided to be too

complex. The spacing between the clips on the fixed rack was

reduced to account for all the required clips. Because of this, we

were able to use the fixed rack. Another problem with the revolving

rack was the parameter of the rack needing to be in an arc shape. A

revolving rack that traveled the path of an arc would be even more

complex.

Al'm

Many different types of arms were considered for this

application with degrees of freedom from 2 to 6. To improve

reliability and reduce complexity, the arm finally decided upon

contained 3 degrees of freedom. This allowed us to maximize

simplicity and yet be able to achieve all of our performance

objectives. For the actuation of the arm, electric, hydraulic, and

pneumatic driving system were examined.
B.7

I°

MACROCORE RETRIEVAL AND STORAGE

Methods of Breaking Core

A. Water J_e_la

DESCRIPTION: Water jet technology employs a narrow, high velocity

stream to cut through a variety of hard materials.

DISCUSSION: Impractical in a lunar environment due to hard

vacuum and lack of water supply.

B. Thermal Shock:

DESCRIPTION: Sufficient thermal stress in a localized region causes

crack propagation. Application of liquid nitrogen or

other supercooled liquids can accomplish this task.

DISCUSSION: Hard vacuum prevents application of any type of

liquid bath. Also, if this method could be employed,

the core could be destroyed.

C. Electric Arc;

DESCRIPTION: High energy path is produced between two electrodes

in an electric arc which can be used to vaporize or

melt rock.

DISCUSSION:

DESCRIPTION:

DISCUSSION:

Lack of atmosphere prevents formation of an arc.

Lasers can be used to cut through many materials

including rock.

This would be an effective method, however, there is a

high degree of complexity and relative lack of dirt

tolerance involved.
DESCRIPTION: Certain strong acids can eat through rock.

DISCUSSION:

F. Tension:

DESCRIPTION:

DISCUSSION:

G. Torsion:

DESCRIPTION:

Hard vacuum prevents the use of volatile substances.

Sufficient tension will break a mineralogical specimen.

Most rock cores on earth are broken by this method.

This is a good solution, but the force required in this

situation may approach 4000 pounds per square inch.

The skitter has the capacity of providing

approximately 200 pounds of upward force. This

limited force could be used in combination with some

type of jacking device mounted on the footplate. The

release of energy upon breakage could be potentially

catastrophic and decrease the overall reliability of the

systeml

Rock specimens can be broken by applying a torsional

force.

DISCUSSION:

ILJ_k_d_g_

DESCRIPTION:

DISCUSSION:

Essentially the same problems and advantages exist as

in tensional breakage.

Application of a bending moment can break a rock

specimen.

This is a good method for it requires much less force

than tensional or torsional breakage. However, due to

the size of the outer annular diameter, there is little

room to apply this force to the top of the core sample.
I. Impact;

DESCRIPTION: Sufficient impulse loading on a localized

causes crack propagation and breakage.

area can

DISCUSSION: Limited space makes this method impractical because

a reasonable amount of room would be required to

accelerate this impacted mass into the core. Also,

excessive damage of the core could occur.

J. Thermal Expansion:

DESCRIPTION: Water in a confined, localized area of a specimen can

an exert a large amount of pressure due to freezing,

thus causing crack propagation.

DISCUSSION: Hard vacuum prevents the use of water in this

application.

K. Scoring:

DESCRIPTION: Creating a local high stress area to assist in breaking

by another method (eg. breaking glass).

DISCUSSION: Good concept, but must be used in conjunction with

another method of removal.

DESCRIPTION: Grinding or cutting the specimen either partially

totally with some type of abrasive arm or wheel.

or

DISCUSS ION: This is a good solution since it does not require large

amounts of torque or lifting force. It would interact

well with another method of core removal.

Use of explosives will break rock.
DISCUSSION: Dangerous and catastrophic forces involved and lack of

oxygen for burning make this method impractical.

N. Sonic Vibrations:

DESCRIPTION: High frequency sound waves resonant a local area to

its natural frequency, thus weakening and breaking

material.

DISCUSSION: This method would not be feasible due to its

complexity and the varying types of strata

encountered in the lunar regalyth.

O. Melting:

DESCRIPTION: Application of high heat source could be

weaken and melt an area of the specimen.

used to

DISCUSSION: Lack of oxygen makes this solution untenable.

P. Crushing_

DESCRIPTION: Application of a large load to crush a localized area of

the specimen.

DISCUSSION: The necessity of applying incredibly large forces as

well as the possible destruction of the sample makes

this method undesirable.
MACROCORE RETRIEVAL AND STORAGE

II. Methods of Breaking Core using Tension

F1. Chinese Handcuffs;

DESCRIPTION: This method is a system of interlocking cables which come together

locking down on the core when tension is applied to either end of the

cylindrical cable network.

DISCUSSION:

 g_Ring 

DESCRIPTION:

DISCUSSION:

F3. Screw Ringz

DESCRIPTION:

DISCUSSION:

This is a good method for the network will hold the core intact.

However, the interlocking cable network is statically indeterminate and

hard to manage.

The bottom of the core retrieval device flairs out into several flanges.

By forcing a ring down onto the flanged region, its inner diameter

decreases locking onto the core for subsequent breakage in tension and

removal.

This method would work, but several problems become instantly

obvious. First, the ring sliding over the outer body of the retriever must

overcome possible vacuum welding conditions. Second, this design

requires a veritable plethora of moving parts. Finally, dirt could become

lodged between the ring and the core retrieval body and result in wear

problems.

This is an extension of the locking ring concept in which the locking

ring moves down over the flangesby means of screw threads.

The mechanism of this system would provide an excellent steady

increase of pressure on the core for gripping purposes. However, dirt

tolerance creates difficulties in using any type of threaded device in the

lunar environment.
-- MACROCORE RETRIEVAL ALTERNATIVE DESIGNS

Figure B.7.1 CHINESE HANDCUFFS

Core

Interlocking Cables

Shell

Core

Locking Ring

Figure B.7.2 TENSION RING

2

\

Confining Stops

Shell
F4. Pivotin_gs:

DESCRIPTION: A pair of rings,hinged on an annular axis,is rotatedsuch that each

member ispulled in an opposing direction. Points of high stressare

createdatthecontactarea affectingeventualbreakage forremoval.

DISCUSSION: This mechanism will work, but it is not capable of optimum core

retention.

F5. Post-holeDigger."

DESCRIPTION: The name of this design belies its description. Two half cylinders

rounded atone end can descend to the core and tightenaround it. The

core can then be broken and removed by a tensileorbending action.

DISCUSSION: This method iswell-adaptedto the removal of broken cores. Problems

ariseinthe areasofgrippingand removing of core thatisintact.

DESCRIPTION: The core retrievalbody isa cylindertwo meters in length lined on tile

insidewith springdriven cam-shaped elements. These elements allow

the retrieverbody to slidedownward over the core, but force into the

core with any upward movement. Ifsufficientupward tension can be

appliedto the core in tandem with thisgripping action,the core can be

broken and removed.

DISCUSSION: This would be an effectivemethod of gripping the core in the hole.

Problems arisein the area of loading stressessince the individu_ cams

will contactthe core and cause largepoint loads possibly crushing the

core.
- PIVOTING RINGS

Core Retrieval Device

Inner Casing

Core

Inner Tube

Core

Figure B.7.3

Shell

Pivoting Rings

Figure B.7.4 TENSIONING ANNULAR REDUCER

Power Screw

Shell

Wedges
_._ FigureB.7.7 MODIFIED WEDGED GRIPPER

OuterShell

InnerShell

Core

t*

p

fieval Device

Core

Figure B.7.8 CAMS

Shell

Cams
III.

H1. Wedge."

DESCRIPTION:

Methods of Breaking Core using Bending

The cylindricalretrieverbody would be equipped with a wedging device

at itstop. When the retrieverisfullyin place around the core, the

wedge is forced down between the top of the core and the retriever

body producing a bending moment on the core itself.This moment

would break the core.

DISCUSSION: This isnot a good way to break the core,since itwill likelyconsistof

varying stratawith a vast range of yield strengths.Therefor the core

would break at the weakest point and not necessarilythe bottom. Itis

desiredto retrievethe fullcore so thisisnot a good solution.

H2.

DESCRIPTION:

Post-holeDigg_:_i_A_hm_

This design isa post-holedigger concept combined with a method of

applying a bending moment tothe core once itissecurelygripped. The

bending moment breaks the core and the retrieverisretractedfrom the

hole.

DISCUSSION: Although this is a good design, there are problems in dealing with

possiblebroken cores as well as with the mechanics of gripping the

core. The mechanism isvery long and the bottom sectionsof a broken

core could dislodgeand fallback intothe hole.

IV. Methods of Breaking Core using Scoring

KI. Pipe Cutter Mechanism:

DESCRIPTION: A disk of hardened metal with a sharp wedge-shaped edge is forced into

one side of the core while a rollingwheel is applied to produced an

opposing contactforce. The disk and wheel rotatein tandem around

the core. This actionproduces a relatedstressfieldwhich follows the

contactpointforce of the rotatingdisk. Itis this stressfieldand the

stressesinduced by the wedge-shaped disk that causes the cylindrical

core sectiontosplitapart.
Core

Posthole Digger Apparatus

Power Screw

Core

Posthole Digger Apparatus

Figure B.7.5 Post-hole Digger

"\\\"\"_ilil

•x\\\\\\\ _, :::::

\\\\\\\N2 :::::

x\'N\'NN\\_

,\\\\ xNNa,_

_\x\-_x\\. iiii

_\\\\\\\_ i[ii

,-.-.-oo,\.., i2!!

I

.\\\\\\

i! x,\,,,

ii ...... N\\\\ \

N\\\\\

°

I

|

Power Screw

_,.,..._'_-\\\M:.:,:- _ ,,_::::::: \_,

.,::i:i:i i:i:i!:!:!:i:i:!:i

Shell

Locking Arms

Figure B.7.6 WEDGE

...... .,. • _\,,,,_

i£1i;;!iii::ili N

i:i;!:iiiii?iii::_

i!iiiii:ili:iii_i_

:i111!2!ii_i[5!ii_

i_i)iiiiiiii:_

}iiiiiiiiii2;ii

@x2_ _ ..... _- k,_

Re

!!:iiiiiiii!_i_ 'i

ge

Shell

Pivot
DISCUSSION: This is an excellent method because little penetration of the disk is

required to actually break the core. The heterogeneous and brittle nature

of a typical regalythic core assists in the fast fracture mechanics of this

method. In addition, this method is relatively quick for the same

reasons mentioned above. It must be used in conjunction with a device

that will grip the core in a removal process.

V. Methods of Breaking Core using Cutting

Li. Di,_nond Teeth Cuttin_

DESCRIPTION: Rocker arms inside of a gripping device are imbedded with diamond

teeth. At the proper location on the core, the arms could be

mechanically activatedto contact the core under force of a torsional

spring. The drillstringisthen rotatedby the skitterand the arms cut

intothe core untilitisbroken and ready forremoval.

DISCUSSION: This method creates a heat management problem in the cutting arms.

These arms would be relatively small bodies of metal incapable of

dissipating much heat. Also, there are possible problems with core chip

removal and abrasion of the arms.
B.8 Controls

The alternative designs discussed for control of the drilling

operation consisted of both a more traditional design as well as a

more futuristic design.

The traditional design that was discussed consisted of the

"typical" NASA space control panel common to the Apollo and earlier

space missions. The panel contained analog gauges and solid state

switching. Data entry and routine commands were executed through

the use of a keyboard or specifically designated push-button

controls. This system was not chosen because of its size,

maintenence difficulty, and lack of concern for the comfort of the

operator.

The futuristic design was not chosen due to the unknown

reliability and expense of its leading-edge technology. This design

would, however, result in the ultimate in operator comfort and

control. The work site on the moon would be monitored by video

cameras at three or more multi-occular positions. The views would

be assimilated by a computer and a three dimensional projection

would appear on a global screen about the operator. This would give

the operator the sensation of witnessing the procedure in person.

The benefits for the operator of such asystem are obvious, but

incorporation of the technology would not be feasible at this time.
B.9 Vertical Accelerator

Design Alternatives

DESIGN A

(CAMS)

DESIGN B

(SPRINGS)

DESIGN C

(GEARS)

DESIGN D

(2nd MOTOR)

DESIGN E

(OTHER)

Version 1

Version 2

Stop Here

Version 1

Version 2

Version 3

Stop Here

Version 1

Stop Here

Version 1

Stop Here

Version 1

Version 2

Final Design

Design E (Version II) will be the final configuration choice for the report
TYPE : A

VERTICAL ACCELERATOR FOR STEPPED AUGER

VERSION : 1

TITLE OF DESIGN :

DESIGN COMPONENTS :

Rotating Cam Mechanism

Secondary Motor and Support Brackets

Rotating Cam and Gearing Struts

Clutch and Internal Spring

Engaging Device for Rotating Cam

lVIE22-IOD OF OPERATION: Following completion in drilling, clutch

disengauges to allow for control rod separation.

Rotating Cam is powered up by means of secondary

motor. Stepped Auger is able to vertically oscillate

downward due to the Cam and return upward due to

compression spring.

DESIGN ADVANTAGES :

- Cam design provides method of jerking

Auger particles in an abrupt fashion.

- Good Reliability

- Gear Simplicity

DESIGN DISADVANTAGES :

- External Motor and Power Up required

- Complexities involved in engaging motor and cam

structure to control rod

- Difficulties for internal spring mount

- Addition controls required for clutching rods and

engaging motor

- Continual Cam alignment with control rod

- Spacing availabilities with regard to rod changer

and storage racks

- Rotating Cam to control rod lip contact yields metal

to metal wear

COMMENTS : Due to complexities involving present

location of Rotating Cam and difficulties regarding

secondary motor and clutch controls, design is not

feasible as shown.

However, if a cam was designed powered and located

differently to control rod rotating cams may prove to

be effective.

FINAL RECOlVIMEaNDATIONS :

Discontinue all futher developments l_gmxiing existing design
w

Design A.1 - Rotating Cam Mechanism

Rotary Motor

/

Skitter

Clutch

Secondary

Separation Space _ _ Motor

Rotating Cam

Moon Surface I

Footplate

-. Drill

String / /

Stepped

Auger //

D .rill /

B,t //
VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : A VERSION : 2

TITLE OF DESIGN :

DESIGN COMPONENTS :

Retractable Cam Mechanism

Secondary Motor and Motor Supports

Rotating Cam and Control Rod Clamp

Linkage and Retractable Strut Supports

Possibility of third motor to engage and disengage

retractable cam mechanism ',

_OD OF OPERATION : Following completion in drilling, control string is

disengaged from motor and vertical accelerator

extends inward to "snap" onto control rod. Rotating

Cam is powered up by means of secondary motor.

Stepped Auger is able to vertically oscillate due to

inner configuration of cam track.

DESIGN ADVANTAGES : - Cam design provides method of jerking

Auger particles in an abrupt fashion.

- Good Reliability

- No crude springs or gear wearing

- No clearance difficulties or complex controlling

- Implements existing gripping present in footplate

- Maximum number of oscillations can be used as

needed.

DESIGN DISADVANTAGES : - External Motor(s) and Power Up required

- Possible friction occurring on inner cam tra'_l_

Potential difficulties could exist in alignment'

COMMENTS : Although complexities are involved with additional

motor(s) operation and controlling, cam design as

shown will have good reliability and is a very feasible

and realistic system.

FINAL RECOMMENDATIONS :

Develop existing design as a feasible vertical accelerator for the stepped auger

; _;Ji

I

,I

[!

, : :, '.:{

i

r

i :
- Design A.2. Auxilliary Motor Driven Cam

Rotary Motor Skitter

Cam

Moon Surface

J

%.

Solenoid

Motor Support

Arm

Footplate

Drill

String / /

Stepped /

Auger

Drill

Bit _
VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : B (Springs) VERSION : 1

TITLE OF DESIGN : Above Ground Compression Spring Mechanism

DESIGN COMPONENTS :

$

100 mm _ompression Spring

Spring I_busing

Control Rod Piston

M:ETHOD OF OPERATION: As Skidd i 10.Wei,s its.eft during drilling process,

control r_.dpl§_:0_lwil1!ower into spring housing

compress ihg _iidt_g spring under a reduced

rotationWith i_§i_bcttticontrolrod. After drillingis

completed, footplatewillgrip control rod and Skidder

will raiseitself.When instructed footplatewill

release grip of rod and spring will extend rod down

forcing particlesupward through the Auger.

DESIGN ADVANTAGES :

- Few components and design simplicity

- No secondary motor required

- Single compression spring no gearing required

- Minimal controlling required

- Implements existing gripping present in footplate

and forceable compression by means of Skidder

DESIGN DISADVANTAGES : -Minimal verticaloscillationseach time spring is

enabled.

- Abrupt oscillationswillminimize the number of

particlesjumping between steps.

- Low precision due to crudeness of spring design

- Spring housing and Piston interferewith rod

changer

- Turning Spring during compression creates

unnecessary spring stress

COMMENTS : Skidder and footplate operation benefits indicate

potential for e_sting design. Try to eliminate space

constraints imposed by rod changer and spring

rotation.

FINAL RECOMMENDATIONS :

Continue to develop or modify the existing design as needed

I
m

Design B.1 - Above Ground Compression Spring

Rotary Motor

Piston

Moon Surface
- VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : B (Springs) VERSION : 2

TITLE OF DESIGN: In-Ground Compression Spring Mechanism

DESIGN COMPONENTS : 80 mm Compression Spring

Spring Hot_Mhg

Control Rod _iiston

METHOD OF OPERATION: As Skidder lowers itself during drilling process,

control rod piston will lower into spring housing

compressing and turning spring under a reduced

rotation with respect to control rod. After drilling is

completed, foq_pl_,te will grip control rod and Skidder

will raise itself. When instructed footplate will

release grip of rod and spring will extend rod down

forcing particles upward through the Auger.

DESIGN ADVANTAGES : - Few components and design simplicity

- No secondary motor required

- Single compression spring no gearing required

- Minimal controllingrequired

-Implements existinggripping present in footplate

and forceablecompression by means of Skidder

- No interference with rod changer

-No additional stresscaused by twisting spring

DESIGN DISADVANTAGES : - Minimal vertical oscillations each time spring is

enabled.

- Abrupt oscillations will minimize the number of

particles jumping between steps.

- Low precision due to crudeness of spring design

- Underground spring forces control rods downward

thereby subjecting the drill bit to damage from lunar

surface

- Minimal oscillations due to footplate interference

- Footplate damage resulting from impacted piston

- Additional controls required for a piston clutch

COMMENTS : Use of inground application for compression spring

poses serious clearance problems. Existing

compression spring design is not feasibleeither

above or below footplate.Remaining alternatives

leave only extension spring designs.

FINAL RECOMMENDATIONS :

Discontinue allfux_er research regarding compression springs.

Analyze opportunities involving extension springs.
- Design B.2 - In-Ground Compression Spring

Rotary Motor

Piston

Moon Surface

Footplate /

Compres

Drill /

String / _

Stepped

Auger //

D.rill /

B_t //
VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : B (Springs) VERSION : 3

"w

TITLE OF DESIGN : Extension Spring Mechanism

DESIGN COMPONENTS : Extension Spring

Spring Housing

Threaded Shaft

METHOD OF OPERATION: After drilling is completed, Skitter raises up and

footplate grips control rod. Motor reverses direction

to unscrew rod assembly. Skitter raises again to

extend spring. Footplate releases and rod oscillates.

DESIGN ADVANTAGES : - Few components and design simplicity

- No secondary motor required

- Single compression spring no gearing required

- No addition stress caused by twisting spring

- No clearance difficuties

- Implements existing gripping present in footplate

and forceable compression by means of Skidder

- No interference with rod changer

- No additional stress caused by twisting spring

DESIGN DISADVANTAGES : . Minimal vertical oscillations each time spring is

enabled.

- Abrupt oscillations will minimize the number of

particles jumping between steps.

- Low precision due to crudeness of spring design

- Possible danager can exist in using extension

springs

- Reversability of drill shaft required

COMMENTS : Research possibilty of implementing reversable drill

motor. Continue further analysis of feasibility

regarding extension spring

FINAL RECOMMENDATIONS :

Continue to develop existing design as needed.
Design B.3 - In-Ground Compression Spring

W

Rotary Motor Skitter

Piston

Moon Surface - :/

Footplate /

Compression sp_

_'_/"--j /

J, Stepped

Auger //

__._21 Drill /

\ _0/-_ Bit /
VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : C (Gears) VERSION : 1

-w

TITLE OF DESIGN : Bevel Gear Mechanism

DESIGN COMPONENTS : 2 Bevel Gears

2 Retractable Arms for Device

Control Rod Pivit and Guide

Link conecting Gear to Control Rod

METHOD OF OPERATION: Following completion of drilling, Skitter will raise

itself to detach from control rods. Retractable arms

will engage beveled gear device onto primary motor.

Gears will be powered and will vertically oscillate

conecting link quickly shaking Auger particles.

DESIGN ADVANTAGES : - No secondary motor required

- Accuracy due to beveled gears

- No crudeness due to springs

- Good Reliability

- Maximum number of oscillations can be used as

needed.

DESIGN DISADVANTAGES : - A great number of controlling devices are required

- Simetrical Gear will cause difficultly in allowing

enough time to force particles up the Auger

- Retractable arms will interfere will rod changer

- Lunar dust will cause excessive gear wear

- Torque overload exists on pivot

- Difficulties involving control link connection to gear

COMMENTS : Retractable arm interference and a great degree of

controlling will be required. However, beveled gear

design does pose many benefits of simplicity

oscillatory motion.

FINAL RECOMMENDATIONS :

Modify the existing design as needed to achieve satisfactory gear device
Design C.1 - ]_evel Gear Mechanism

Skitter

Retractable arm

Rotary Motor

arin

Link

Pivot

/

/

Bevel Gear

Guide

Moon Surface

m m

Drill

String

Stepped

Auger

Drill

Bit

Footplate
VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : D VERSION : 1

TITLE OF DESIGN : Secondary Motor Configuration

DESIGN COMPONENTS : Secondary Motor and Support Brackets

Clutch and Detachable Housing

Link conecting Secondary Motor to Control Rods

METHOD OF OPERATION : Following completion of drilling, clutch disengauges

to allow for control rod separation. Secondary motor

and control link attach to lower control rod. Link

will vertically oscillate Auger shaking particles.

DESIGN ADVANTAGES : - Few components

- No strain on primary motor

- Secondary motor provides ability for accurate

oscillations

- No crudeness due to springs

- Good Reliability

- Maximum number of oscillations can be used as

needed.

DESIGN DISADVANTAGES : - Additional control devices and maintenence are

required for secondary motor

- Space limitations exist with external mounting of

secondary motor

- Larger operational costs due to additional motor

COMMENTS : Without further analysis, mounting locations and

control rod conections presently inhibit the

effectiveness of the secondary motor. Secondary

motor assembly dimensions and interconnection

with regard to the control rods would have to be

studied further in order to render a more accurate

decision.

FINAL RECOMMENDATIONS :

Continue to develop the existing design as needed
w

Design D.1 - Secondary Motor Configuration

Primary Motor Skitter

w

w

and _ X

Detachable _ = tx

Housing t._, ...... _.__

......................... ,__i Secondary

Vertical Travel Motor

Dist. Rotating Cam

Moon Surface I _ _-jA'k_-_

Footplate

-2

... Drill

Stnng / /

Stepped /

Auger Drill

Bit //
VERTICAL ACCELERATOR FOR STEPPED AUGER

TYPE : E VERSION : 1

OF DESIGN : Footplate Mountable Shaker Device

DESIGN COMPONENTS : Twelve Control Rod Cups per 2 meter rod

Clutch and Detachable Housing

Footplate mounted Roller Ball and Spring Device

METHOD OF OPERATION: Following completion of drilling, clutch disengauges

to allow for control rod separation. As Skitter raises,

control rods will be dr_wn upward through footplate.

The roller ball mounted to the footplate link will

catch each conrol rod cup as it travels upward

through the footplate. After a critical point is

achieved the roller ball will slip off each cup abruptly

jerking rods causing Auger to shake.

DESIGN ADVANTAGES :

- No rotary of motor driven parts

- No strain on primary motor

- No maintenence

- No controls needed for shaking

- Consistent adjustable travel distance for rods

DESIGN DISADVANTAGES :

- Vast complications develop when altering design of

control rods due to cups

- Oscillations are performed by Skitter lifting up,

only minimal shakes will for each drilling cycle

- Control cups will affect rod changer and storage

racks design

- Control cups will affect footpate design and pose

potential gipping difficulties for the group

- Multiple cups will add significant weight to control

rods and drill motor, especially at 50 meters

COMMENTS : Although simple in design, control rod cups and

footplate shaker pose significant difficulties due to

interaction with other groups and major design

changes on existing parts.

FINAL RECOMMENDATIONS :

Discontinue all further developments regarding existing design
- Design E.1. Footplate Mountable Shaker Device

7-

Primary Motor

Control

Rod

Detach

Point

Moon Surface

|

_ Con

U

:i: i

ii

u

q.

ii

nl

Z

_"_ R__oller BFa°_lplate /

Spring De

't:'n./

Stepped /

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B.10 MICRO-CORING DRILL

B.10.1 Drilling mechanism

Several drilling mechanisms were considered before a decision

was made to use the Diamond-Tipped coring bit. These alternative

modes are listed below:

• Ultrasonic Drill: when the ultrasonic vibrations are

applied to a solid at a certain point, plastic deformation

will occur, and if the cycles continue at a high

amplitude, the sample will break.

• Wedge Impact Sampling: this method uses a chisel like

instrument to break of a sample. This method is very

effective when ample space is available.

• Vibro Drilling: drilling by vibration involves the use of

pulsations from a high frequency vibration source and

the weight of the drilling apparatus to penetrate the

soil.

B.10.2 Bracing system

Two alternative methods of bracing were considered but

were deemed undesirable for this application. The first method

would have used two screws on opposite sides of the rod, driven

into the wall by a shaft powered by the main motor. The other

bracing mechanism we considered was a truncated cone that would

slide into the hole until its diameter was too large to slide any

further, thus bracing the rods firmly at the top.

B.10.3 Power source

An alternative source of power for the drill bit could be the

main motor from Skitter. The motor mould rotate a shaft, and a

series of gears would be used to convert this rotational motion into

rotational and translational motion for the drill bit.
A bevel gear attached to the drive shaft would mate with a

spline to transmit the necessary power. Two spur gears, located on

opposite sides of a separate motor, would rotate along two pairs of

racks to translate the drill bit into the lunar rock.

B.10.4 Sample protection

The sample could be protected by a device similar to the one

found in the final design. In this alternative case, an entire section of

the cylinder would move down, but this would leave no place to

mount the tracks for the translation.
APPENDIX C

PROGRESS REPORTS
WEEK OF APRIL 10, 1989

• The cutting teeth configuration for the main drill bit and a shoveltype mechanism to remove rock cuttings away from this bit were

discussed.

• Commercially pure grade-5 titanium was selected as the rod

material due to its good high temperature behavior and superior

strength to weight ratio.

• A set of ring sleeve bearings was adde_l to the rod deisgn to

improve buckling resistance.

• Reliability and performance constraint< for the lunar footplate

were determined.

• Nuclear fission and hydrazine were ruled out as potential power

sources for the lunar drilling apparatus. Employing a solar power

source with a rechargeable battery was determined to be the most

feasible method of providing power for this project.

• The materials selection and method of attachment for the drilling

platform mechanical interface were discussed.

• Research was conducted on grippers and robot arms feasible for

use in the rod changer.

• Martin Marrietta's Report on Lunar Apparatus and an Acker Drill

Company catalog were examined to find a potential method of

removing the macro-core produced by the main drilling process.

• Each member of the five man controls team took responsibility for

interfacing with two of the groups designing the other subsections of the lunar drilling apparatus.

C1
Continued

• The interaction of the vertical accelerator with the overall lunar

deep drilling apparatus was discussed.

• Using a diamond-tipped coring bit was determined to be the best

method of obtaining a pristine, micro-core sample. Ultrassonic

drilling, drilling by vibration, and wedge impact sampling were

alternatives that were eliminated.

C2
WEEK OF APRIL 17, 1989

It was determined that the main drilling operation needs to be

continuous for depth increments of only one-half inch. This will

make the removal of cuttings from the drilling process more

reasonable.

The rod interface and coupling collar were designed to isolate

torsional and and tensional loads in separate elements.

Rather than designing the footflate mechanism to handle an

uneven surface, it was decided have Skitter smooth out a surface

in the lunar regolith before the footplate is set down. A fishtail

bit or a scraping mechanism will be employed to carry out the

smoothing process.

A CSM motor was obtained from the V A hospital on April

20,1989. This will be used to power the drill bit for the main

coring operation, the macro-core removal process, and possibly

part of the micro-coring process.

Various alternatives for removal of the macro-core were

considered. Most of these would employ a "post hole digger" type

mechanism.

A list of potential vendors for electrical sensors, temperature

sensors, switches, actuators, velocity sensors, torque sensors, and

force sensors was compiled by the controls team.

Several potential methods for bracing the bottom rod for the

micro-coring process were analyzed. It was determined that pads

at the end of a spring loaded piston should be driven by an ovalshaped cam.

C3
Continued

• The mechanical and thermal properties of basalt were researched

by the micro-coring group to be used in a heat transfer analysis to

determine the maximum power input to that drill.

C_
WEEK OF APRIL 24, 1989

The design of the cutting teeth and the inner stepped auger and

core bit connection was begun this week. Contact was also made

with several dry core drilling companies for information.

A one-fourth scale model was built of the entire drilling operation.

Preliminary drawings of the rod and interface mechanism were

completed.

A chuck-type mechanism was proposed for the design of the lunar

excavation assistance foundation. An impeller type arrangement

is also under consideration currently.

Specifications for the CSM motor were obtained, and the sizes of

heat pipes and radiators were calculated.

Information on materials for the space environment was collected.

Research is also under way for information pertaining to the

method of attaching the platform to Skitter.

From force analysis, the exertable force by the walker is

estimated to be 450 lbs. in the upward direction, and 350 lbs

downward. This will limit the size of the core that walker

can pull up.

Currently, five preliminary designs are being considered for the

vertical accelerator.

• Three preliminary methods for bracin_ the micro-coring drill

operation were presented during ihi_ week.

C5
WEEK OF MAY I, 1989

• A drilling increment of 0.4" was decided for the main drill bit.

• Possible materials for the rods are currently magnesium and

titanium, or possibly some form of metal matrix composite. Final

selection will depend on temperature requirements and strength.

° The chuck-type mechanism proposed last week has been replaced

with a simpler, lighter design. The previous design was scrapped

due to anticipated problems in sealin_ from dirt and dust and the

complexity of the clamping device

• Batteries as a source of power for the drilling operation continue

to cause problems due to the temperature limitations. Fuel cells

are a possibility for an alternative power source.

° The interface shape was changed from triangular to hexagonal to

accommodate the large rotary motor, power cells, and rod holders.

The platform's depth was also increased.

• From the five preliminary designs for the vertical accelerator, the

two most feasible are cams and locking ring mechanisms.

Research is continuing of the propertie,_ of lunar rock.

• Issues in the control system of the lunar deep drilling operation

were discussed this week with Dr. Mike Kelly of the Georgia Tech

Research Institute.

• A mathematical model for the vertical accelerator is completed

and can be implemented to optimize the required amount of force.

C6
WEEK OF MAY 8. 1989

The final dimensions of the cutting holders on the drill bit were

completed. The core bit can drill to a depth of .43 in with out

stopping for rock chip removal. Looking at the possible use of

removable teeth. The implementation of heat pipes is being

investigated.

Force and deflection calculations for the clamping spring on the

footplate were done. The selection of the micromotor that will

provide a lifting motion for the spring. This motor will

incorporate a built-in, sealed gear reducer and it will be insulated.

The motor specifications were changed to a 5 hp CSM motor that

will operate at 100 rpm. A clutch was incorporated into the

design along with a solar radiator. The power supply will come

from batteries stored in the rods that will last for 10 hours. The

storage rack clips need to be modified for battery contacts.

The height of the mechanical interface reduced due to only 2" of

height needed for the motor.

Cams or a locking ring mechanism will be used in the macrocore

retrieval and storage process.

A mathematical model of the vertical accelerator for the stepped

auger is complete. This is capable of optimizing the force required

in determining the design parameters

The motor and cam system utilized in bracing the micro-coring

drill are complete. The current design is also completed, using

only one motor to drive the drill translation.

C7
WEEK OF MAY 15, 1989

Preliminary analysis of the heat developed at the cutting interface

was developed and this will allow the calculation of the

temperature at the teeth. The final problem was sectioned in

smaller subsections.

The rod interface was redesigned to accomodate the connection of

two stepped augers. The selection of possible materials was

narrowed to four.

• Calculations for the deflection of the foot plate were almost

finalized. Modifications to the flat spring are still being discussed.

A material was selected for the structural frame of the drilling

platform. The material selected was aluminum reinforced with

40% silicon carbide fibers.

The design of the rod changer and storage rack were almost

completed. A1 2219-T37 was used as the main material and MoS2

was selected for the lubrication.

Computer operating screens were developed for the control

console. Time calculations for the length of the entire drilling

process were conducted.

• The bearings for the vertical accelerator were sized and designed.

The micro coring drill will be powered by two seperate motors

that are independent of the main moto_ The drill will be

protected by a sliding door.

C8
WEEK OF MAY 22, 1989

The final week was almost entirely dedicated to the

preparation of the final report and the models for the formal

presentation of the project.

A new method of power supply was adopted. Due to the

excessive weight of a battery powered system, a hydrogenoxygen fuel cell system was adopted.

For the macro core retrieval a ne,_ system was designed. A

cutting device similar to a pipe cutter, powered by a spring,

will be used to cut the core. The device will no longer be

attached to the drill bit, but be contained in a separate

cylinder.

• The design of the rack clip was finalized and the force

analysis was completed on the robot :_rm.

C9
APPENDIX D

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