Aerial Lift Platform Design for Engine Testing Service

MECH 4860 Ȃ Engineering Design A01 Team #13 Ȃ The ALPS

Faculty Advisor Ȃ Kathryn Atamanchuk Submitted: December 2, 2013

Jessica Biggs ______

Genadi Borisov ______

Jared Halpin ______

Cameron Verwey ______

[Company Address] Clent Approval

i Genadi Borisov Team 13 ± The ALPS

Winnipeg, Manitoba Canada

December 2, 2013

Dr. Paul E. Labossiere Mechanical and Manufacturing Engineering Department Room E1-546, EITC Building University of Manitoba Winnipeg, Manitoba Canada R3T 5V6

Dear Dr. Labossiere,

Please find enclosed a copy of the final design report from team 13 ± The ALPS. The team consists of Jessica Biggs, Genadi Borisov, Jared Halpin, and Cameron Verwey. The project is a design of an aerial lift platform for engine service for the GLACIER facility in Thompson, Manitoba.

In this project we designed, modeled, and made engineering drawings of an aerial lift platform as SHUWKHFOLHQW¶VQHHGV7KLVILQDOUHSRUWGRFXPHQWVWKHFRQVLGHUDWLRQVWDNHQLQWR the design such as the stress analysis that was conducted, and part selection and sizing.

Our work in this project fulfills the objectives of this course. Our design meets the major requirements set by our client. We have managed to meet all of the deadlines for our deliverables. We got the opportunity to use our mechanical engineering knowledge to design a product for a real life problem. We developed our engineering design skills as well as our communication skills as a result of this course.

Feel free to contact me anytime for any questions you may have on this project. My email address is

Sincerely,

Genadi Borisov Team Manager Team 13 ± The ALPS

ii Contents

List of Figures ...... vi

List of Tables ...... viii

Executive Summary...... ix

1 Introduction ...... 1 1.1 Project Stakeholders ...... 1 1.2 Problem Statement ...... 2 1.3 Project Needs and Specifications ...... 3 1.4 Project Objectives ...... 8 1.5 Design Concepts...... 8

2 Design Methodology ...... 10 2.1 System of Units ...... 10 2.2 Material Selection ...... 10 2.3 Procurement of Materials ...... 12 2.4 Tolerances, Fits and Welds ...... 12 2.5 Stress Analysis ...... 12 2.6 Design Drawings...... 13

3 Details of the Design ...... 14 3.1 Platform ...... 15 3.1.1 Platform Frame ...... 16 3.1.1.1 Hollow Structural Sections...... 16 3.1.1.2 Angle Iron Perimeter...... 17 3.1.1.3 Stress Considerations ...... 18 3.1.2 Platform Surface ...... 19 3.1.2.1 Grating Geometry ...... 20 3.1.2.2 Stress Considerations ...... 20 3.1.2.3 Fabrication and Installation...... 21 3.1.3 Guardrails, Gates and Harness Anchors ...... 22 3.1.3.1 Drop Down Gates...... 23 3.1.3.2 Permanent Side Rails and Corner Posts ...... 27 3.1.3.3 Horizontal Lifelines ...... 29

iii 3.1.3.4 Stress Considerations ...... 31 3.1.4 Summary of Lifting Platform...... 33 3.2 Lifting Scissors ...... 34 3.2.1 Scissor Members ...... 36 3.2.2 Pivots and Slides ...... 39 3.2.3 Stress Considerations ...... 44 3.2.4 Hydraulics ...... 48 3.2.4.1 Hydraulic Pump and Electric Motor...... 49 3.2.5 Summary of Lifting Scissors ...... 50 3.3 Base Frame ...... 50 3.3.1 Structural Frame Components ...... 51 3.3.1.1 Central Frame ...... 52 3.3.1.2 Intermediate Access Platforms ...... 56 3.3.1.3 Hitch...... 57 3.3.2 Steering Components ...... 59 3.3.2.1 Wheel Brackets and Spindle Assembly ...... 59 3.3.2.2 Braking System ...... 65 3.3.3 Stress Considerations ...... 65 3.3.4 Base Frame Summary ...... 67 3.4 Tires and Wheels ...... 67 3.4.1 Design Considerations ...... 68 3.4.2 Operating Temperature ...... 68 3.4.3 Load Distribution ...... 68 3.4.4 Tire Wear and Maintenance ...... 70 3.4.5 Tire Types ...... 70 3.4.6 Pneumatic Tires ...... 70 3.4.7 Solid Tires ...... 71 3.4.8 Multipurpose Tires...... 71 3.4.9 Final Selection...... 71 3.4.9.1 Tires ...... 71 3.4.9.2 Wheels ...... 74 3.4.10 Cost Analysis ...... 75

4 Bill of Materials and Cost Analysis ...... 76

iv 5 Recommendations ...... 86 5.1 Adjustable Stairs...... 86 5.2 Top of Engine Access Ladder ...... 86 5.3 Electrical Connections ...... 86 5.4 Control System ...... 87 5.5 Ergonomic Mat ...... 87

6 Summary ...... 88

7 Works Cited ...... 92

Appendix A: Detailed ƌĂǁŝŶŐƐ͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͘͘͘ϵϰ

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Appendix C: Lifting Mechanism Data͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͙͘ϭϮϴ

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v List of Figures

Figure 1. Engine test stand...... 2 Figure 2. The Complete ALP Assembly...... 14 Figure 3. Complete lifting platform with railings and access gates...... 15 Figure 4. Render of assembled frame...... 16 Figure 5. Illustration of serrated bar grating...... 19 Figure 6. Recommended method of securing aluminum grates to steel box beams...... 22 Figure 7. Render of platform frame with complete guardrail system installed...... 23 Figure 8. Image of the sliding gate sleeve with associated installation hardware...... 26 Figure 9. Corner post supporting plates and caps...... 28 Figure 10. Miller Xenon permanent horizontal lifeline (used with permission)...... 30 Figure 11. Extended scissor pair...... 35 Figure 12. Completely collapsed scissor pair with inner cross-bracing visible...... 35 Figure 13. Holland Lift M-250DL27 at collapsed height...... 38 Figure 14. Render of scissor stack with stabilizing cross members (blue)...... 38 Figure 15. Scissor member with bushing bores and PTFE bushings installed...... 40 Figure 16. Close up view of installed bushing bore with support plates...... 41 Figure 17. Clevis connection with PTFE slide bushing...... 42 Figure 18. Sloped lower clevis mounts...... 43 Figure 19. Translating scissor ends with pins and clevis mounts installed...... 44 Figure 20. Basic scissor structure...... 45 Figure 21. Scissor lift coordinate system...... 46 Figure 22. Detailed view of hydraulic actuator...... 48 Figure 23. Eaton hydraulic power unit...... 49 Figure 24. Assembled base frame with wheels and hitch...... 51 Figure 25. Assembly of the aerial lift platform base...... 52 Figure 26. Sloped fixed clevis mounts and illustrative hydraulic mounts...... 54 Figure 27. Details of the wheel arm with the reinforcement (upper right)...... 55 Figure 28. Intermediate access platform...... 56 Figure 29. Pivoting hitch arm...... 58

vi Figure 30. Hitch used on engine transporter...... 58 Figure 31. Renders of the wheel brackets and axle retainer...... 60 Figure 32. Fully machined spindle...... 61 Figure 33. Machined spindle housing...... 63 Figure 34. Cover plate with bearing retainers installed...... 64 Figure 35. Assembled front caster wheel...... 65 Figure 36: Continental Tire Press on Band Dimension Guide (used with permission). .. 73

vii List of Tables

TABLE I HIERARCHY OF CUSTOMER NEEDS WITH PRIORITY RATINGS...... 5 TABLE II IDEAL AND MARGINAL VALUES OF TARGET SPECIFICATIONS ...... 7 TABLE III BASIC MATERIAL DATA ...... 11 TABLE IV SUMMARY OF SPINDLE BEARINGS ...... 62 TABLE V WEIGHT ESTIMATE OF ALP ...... 69 TABLE VI LOAD CAPACITY SUMMARY ...... 70 TABLE VII DIMENSIONS AND RATED LOADS OF 28X16X22 MH20 TIRES ...... 72 TABLE VIII COST ANALYSIS FOR TIRE AND WHEEL ASSEMBLY ...... 75 TABLE IX SUMMARY OF PLATFORM MATERIALS ...... 78 TABLE X SUMMARY OF SCISSOR ASSEMBLY MATERIALS ...... 80 TABLE XI SUMMARY OF BASE FRAME MATERIALS ...... 82 TABLE XII SUMMARY OF ACTUATION COMPONENTS...... 85

viii Executive Summary

The ALPS (Aerial Lift Platform Specialists) were tasked with designing an aerial lift platform for the GLACIER facility in Thompson, Manitoba. This project consisted of three major phases: project definition, concept design, and final design. The first major project phase required identifying the problem statement and project objectives, and outlining the target specifications. We identified the constraints and limitations associated with designing an aerial lift platform (ALP) for use at the GLACIER facility, and outlined a project schedule. The second project phase was concept focused, and involved a great deal of research. We investigated all aspects of the required design, and identified current market solutions for the specific areas of design associated with our concepts, such as platform design, lifting mechanisms, and steering geometries. Employing the use of concept screening and scoring matrices, our team filtered a preliminary list of 62 concepts, spanning the eight functional categories of design, down to 23 suitable designs. Moving into the final project phase, we began final concept design, focusing on each of the eight functional design categories individually, and combining these optimized groups into one coherent final design.

The GLACIER facility is in need of an ALP capable of meeting their size requirements. In order to meet facility requirements, a suitable lift must have a working surface area large enough to allow for access to all components of the engine while performing maintenance, troubleshooting, and sensor connection. The approximate size required ʹͲǯʹͲǯǤǡ require maintenance personnel to ascend and descend, or reposition the lift, in order to access engine components. These restrictions result in increased costs associated with engine testing, longer turn-time, and reduced efficiency.

The aerial lift platform we developed resembles that of a traditional scissor lift platform. Four wheels situated at the corners of the base platform consist of 28-16-22 ʹͲ ʹʹǳǡ wheels pivoting freely to give provision for steering of the platform. Wheels spindles

ix are mounted high to allow for larger diameter wheels, while the base frame is built low to the ground in order to minimize the retracted height of the platform. Four sets of scissors are mounted on the I-beam base, fixed in pairs on opposing ends of the platform. The scissors consist of ͳͶǳͳͲǳ rectangular beam sections, with three scissor tiers per set. Occupants are supported by the upper platform, which consists of ʹͲǯʹͲǯ-beam frame, with L-channel tracks welded to the top rim. The walking

3 platform is comprised of 19-Ͷʹǳ /16ǳǡ͸ǳ posts permanently fixed at each corner. The left and right sides of the platform have permanent hand rails, with horizontal life lines running the length of the platform at three separate heights on each side. The fore and aft ends of the platform each consist of five adjustable height railings, with the middle three functioning as platform access ladders when lowered fully.

In summary, we designed and aerial lift platform that meets all primary project objectives aside from that pertaining to project budget. Due to the extremely large number of parts associated with this design (>1000), it was not possible to obtain a cost estimate for the entire design, and as such our team is unable to conclude as to the final cost of the ALP. The final design easily integrates into the GLACIER facility, and pertains to the outlined physical constraints. There is room for future improvement with respect to additional components that will help the ALP integrate even further into the facility. Recommendations are given as to the development of various ALP accessories, including the ergonomic mat, engine riser, and electrical outlets.

x 1 Introduction

MDS AeroTest/EnviroTREC have requested that Team 13 design an aerial lift platform (ALP). The ALP will be used to support personnel and equipment required to perform facility related tasks. These tasks include access to the engine while on the test stand and maintenance work around the facility. An in-depth overview of the stakeholders and the project description and problem statement is given in the following sections.

1.1 Project Stakeholders Many organizations work together to ensure success of the Global Aerospace Center for Icing and Environmental Research (GLACIER) facility located in Thompson, MB. For this design project there is also support from the University of Manitoba.

MDS AeroTest operates the GLACIER facility. They will benefit from the design of the ALP by increasing profit. GLACIER is a joint venture between Pratt & Whitney Canada and Rolls-Royce Canada. Both parties of GLACIER will benefit from development of the ALP by being able to certify more engines in a shorter time period. The GLACIER facility performs icing certification testing through the use of a fixed-tunnel assembly, in which the engine is connected directly to the test tunnel. Thompson, MB is an ideal location for the test facility due to the cold temperatures experienced for much of the year.

The Canadian Environmental Test Research and Education Center (EnviroTREC) is a nonprofit organization that helps to aid in research and development, as well as provides human resources support at GLACIER. Additionally, the National Research Council (NRC) provides funding for the facility to promote research and development in the Canadian Aerospace community. As this design project was assigned through the University of Manitoba, academia are benefiting from the ALP design.

We were given the opportunity to work on this design project by being enrolled in MECH4860 Ȃ Engineering Design at the University of Manitoba. Dr. Paul Labossiere, the

1 course facilitator is responsible for the planning and organization of the course. Kathryn Atamanchuk, an Engineer in Residence at the University of Manitoba, is our project advisor. Both Paul and Kathryn provide guidance to the team throughout the design process.

1.2 Problem Statement The GLACIER facility requires an ALP to perform tasks on an engine when mounted in the test stand. These tasks include sensor connection, troubleshooting and maintenance on all parts of the engine.

In order to get a better understanding of where the ALP will be used, the gas turbine engine test stand is shown in Figure 1. For engine related tasks, the ALP will be located under the test stand, between the pillars.

Figure 1. Engine test stand.

While on a site visit to the GLACIER facility it was determined that the ALP will be centered between the two yellow lines visible on the pavement. The client expressed 2 interest in designing the lift to be versatile enough to perform facility related tasks, such as facility maintenance [1]. The design of the ALP is geared for use for under the engine test stand, as this will be the primary use and is the area with highest space constraint.

Upon discussion with Rob Howitt, Chief Facilities Engineer of MDS AeroTest, it was found that currently, commercially available scissor lifts are rented from United Rentals [1]. These scissor lifts do not provide an adequate work area to accommodate more than two workers with the required equipment for the aforementioned tasks. This limited space decreases productivity, as the lift must be raised and lowered many times in order to obtain the tools required for the tasks at hand. As a result, the time required to complete the certification tests on the engine increases, which in turn decreases the number of engines that can be certified per year, resulting in decreased profits for MDS AeroTest.

The ALP to be designed must safely support 4-6 workers, and up to 2000 lb of equipment. The lift platform must also maintain full functionality at temperatures as low as -40°C. The ALP should easily integrate with the current facility equipment and layout, and will be stored outdoors at the test site.

1.3 Project Needs and Specifications Prior to the concept development phase, a list of needs and target specifications was created. By consulting with the client and doing research into current technologies, the team was able to develop a realistic and thorough list of needs and specifications. This list helped to ensure that concept development was performed in an efficient and organized manner. The concepts that were developed were then modified and improved to create the final designs of each component.

The needs were organized into five categories, and subsequently given a priority rating on a scale from one to three. A rating of three signifies the highest priority. If the ALP design did not meet one of the needs corresponding to a rating of three, it would be

3 considered a failure. Any need that dealt with safety was given a rating of three. TABLE I displays an organized listing of the needs with their prioritized ratings.

TABLE I HIERARCHY OF CUSTOMER NEEDS WITH PRIORITY RATINGS

Physical Safety The platform safely supports all workers and equipment 3 The platform conforms to all relevant health and safety guidelines and codes 3 The platform is stable under rapidly shifting loads 3 The platform incorporates anchor points for fall protection harnesses 3 The platform remains stationary while set in desired position 3 The platform features a non-slip working surface 3 The platform features control mechanisms on movable surface and at ground level 3 The platform comes equipped with safety placards and operating procedures 3 Geometry and Motion The platform provides a wide range of lift height capability 3 The platform provides a generous working area 3 The platform is capable of smoothly lifting and lowering 2 The platform features variable lifting rates 1 The platform has a surface capable of stopping in close proximity to engine 2 Additional Functionality The platform is fully functional at very low temperatures 3 The platform provides workers with quick access to the ground from a fully lifted position 2 The platform is easily moved into working position 2 The platform provides electrical connectivity for tools 2 The platform is designed to be stored outdoors 3 The platform features a self-contained lifting mechanism 2 The platform has multiple points of entry 2 The platform enables access to the upper areas of test engines 1 Maintenance and Durability The platform functions as designed throughout a long lifespan 2 The platform is easily maintained 2 Subjective Needs The platform creates enthusiasm on the job site 1 The platform inspires a feeling of confidence and safety for its users 3

It is evident from TABLE I that the most important needs deal with the aspects of providing a safe working area for personnel over a wide range of heights. Also of importance, is that the ALP must be functional at very cold temperatures and stored outdoors.

Each need has an associated target specification and metric. These target specifications have ideal and marginal values associated with them. These values will be used to ǯǤ ǡ and ideal value for each can be found in TABLE II. Evaluations of each function have been carried out and are summarized throughout Section 3 - Details of the Design, of Ǥ ǯ needs were met.

TABLE II IDEAL AND MARGINAL VALUES OF TARGET SPECIFICATIONS

MARGINAL IDEAL METRIC UNITS VALUE VALUE Minimum and maximum lift heights ft 6 - 25 5 - 35 Time to begin lifting process s 20 10 Working fluid loss per use ft3 0.006 0 Integrity of mechanical systems visual safe as-new Usable surface area of platform ft2 400 400 Time to fully extend lift platform s < 120 < 60

Stress and strain analysis under load NVLİ ıy ıy Timed trials of workers exiting the lift s 30 20 Compare to safety standards pass/fail pass pass Number of workers necessary to begin the horizontal moving process # 4 3 Time necessary to begin the horizontal moving process s 60 30 Power available at outlet W 4400 8800 Durability predictions yr. 25 35 Maintenance downtime/usage ratio # 1/30 1/50 Observation of maintenance personnel performing necessary tasks subj. medium easy Range of angular motion ° 2.70 1.10 Reduction of protective surface finish in2 / yr. 1000 0 Anchor compatibility list yes yes Motion observation and measurement ft 0.5 0 Feedback from client and end users subj. good excellent Stand-alone lifting test pass/fail fail pass Coefficient of friction between boot sole and working surface ȝ > 0.4 > 0.6 Rate of precipitation permeability through working surface in/min 0.1 Entrance test from multiple sides pass/fail fail pass Reach distance necessary to reach upper-most engine location ft 2 1 Minimum and maximum safety railing heights relative to working surface in 20 - 42 0 - 42 Control test from ground and lifting platform pass/fail pass pass Inspect inclusion of relevant documentation pass/fail pass pass

1.4 Project Objectives In an effort to fully understand that design project, the team has split the project objectives into primary and secondary. The primary objectives relate directly to the design elements of the ALP. The primary objectives are as follows: x Design an ALP to safely support up to 6 workers and 2000 lb of equipment x ʹͲǯʹͲǯ x Ensure ALP is fully functional in the climate of Thompson, MB x Remain within budget of $350,000 x Allow for design to be built or assembled in Thompson, MB

ǯǤ will still function and be a success if the secondary objectives are not fully met. Lastly, the secondary objectives also deal with professional development of the stakeholders. The secondary objectives are as follows: x The platform provides easy access to the ground from a fully extended position x The ALP requires little to no maintenance over a 25 year lifespan x The platform has multiple access points x The lifting mechanism is variable speed x Develop a positive working relationship between MDS AeroTest and the University of Manitoba x Develop oral and written report skills

1.5 Design Concepts A functional decomposition was utilized in an effort to tackle the large scope of the project. The design was split into eight functional categories, as listed below:

1) Translational Motion Source 2) Steering and Maneuverability 3) Safety Harness Attachment 4) Platform Surface

5) Lifting Mechanism and Actuation 6) Access Openings 7) Lowering of Railings 8) Primary Power Source

Translational motion source refers to how the ALP will move, while steering and maneuverability refers to how the ALP will be positioned into place, such as self- propelled or towed. Safety harness attachment refers to how the current safety harnesses in use will integrate in to the ALP design. The platform surface refers to the material that makes up the supporting area. The lifting mechanism and actuation refers to how the ALP will move up and down. Access openings refer to one or multiple entry points to get onto the ALP. Lowering of railings refers to the method by which the railings will adjust in height to account for the engine size. Primary power source refers to what will drive the lifting mechanism.

Design concepts specific to each function were then developed both individually and as a team. A screening process was then developed to determine the most suitable design concepts. The concepts were evaluated on pre-determined evaluation criteria such as ease of use, safety, cost and ease of manufacturing. Each criterion was then assigned an importance rating relative to one another in preparation for the scoring procedure. Lastly, each design concept was scored and compared against one another. The top three to four design concepts from each function was then selected for a further evaluation in preparation for the final design selection.

2 Design Methodology

The design of a large scissor lift to meet the needs of MDS AeroTest and EnviroTREC involved many aspects, including material selection, procurement of available components, machining and joining methods and stress analysis. The above aspects, as well as unit convention, tolerances and design drawings are introduced in this section.

2.1 System of Units At the request of the client, imperial units are used throughout. Thousands of pounds per square inch (ksi) is used for all stress and strength calculations. Dimensions represented in inches are expressed in fractions, where applicable. Unless otherwise noted, the abbreviation lb refers to pound-force, not pound-mass.

2.2 Material Selection The vast majority of the lift platform is constructed of two materials: aluminum and structural steel. Structurally, aluminum was utilized where weight was a concern and steel was used for everything else. Of special importance was the use of low- temperature, high-strength steel for load bearing components, such as the scissor mechanism and pins. The guideline used for selecting appropriate material is CSA Standard B354.1-04, which governs the design of portable elevating work platforms [2]. Some of the major material selection points outlined in this document are listed below. x Load carrying members must not become brittle at low temperature operating conditions. A structural steel with a Charpy impact test value of 20 J or greater at a temperature of -18°C will meet this standard. x All ductile and load supporting materials must have a structural safety factor of no less than 2:1. Both minimum yield strength and column buckling resistance must be taken into account.

Aluminum was never used as a load bearing material. It was primarily utilized for the guardrail structure (in standard pipe form) and grating surface. Both of these applications were very weight sensitive. Aluminum 6061 T6 is used exclusively (however bar grating composition may vary across manufacturers). Regarding steels,

ASTM A572 steel is used in the less demanding situations due to its high availability, suitable Charpy numbers and reasonable strength. A572 steel is specified for the platform girders and base I-beams. For applications requiring greater load capacity, such as the scissor members and scissor pins, CSA G40.21 70T steel or equivalent is mandatory. 70T steel yields at 70 ksi, while A572 steel has a minimum yield of 50 ksi. It is found that in certain situations, simply adding more material to make up for lack of mechanical strength is not sufficient; stronger steel is needed. 70T steel is also weldable and recommended for use in cold weather. These characteristics make it ideal for the scissor arms and pins. Procurement of this material may be more difficult due to the required HSS shape instead of the more common I-beam and C-channel shapes.

For demanding situations requiring machining from a round bar, AISI-C1144 stress relieved steel is recommended. The yield strength of 100 ksi is exceptional, and a necessity for the caster wheel spindles. A summary of the major engineering materials is provided in TABLE III.

TABLE III BASIC MATERIAL DATA

Yield Charpy Density Modulus Material Required shape Availability strength V-notch (lb/in3) (ksi) (ksi) (J)

6061 T6 Aluminum Russel / Pipe, grating, L angle 41 N/A 0.098 10,100 [3] [4] Brunswick HSS, L angle, I-beam, ASTM A572 Steel [4] C-channel, pipe, Brunswick 50 28.70 0.284 29,000 [5] plate CSA G40.21 70T HSS, round bar Russel 70 N/A 0.284 29,000 Steel [6]

AISI-C1144 SR Steel Round bar Brunswick 100 N/A 0.284 29,000 [4]

2.3 Procurement of Materials A major goal of this project was to design a lift platform that could be manufactured without the use of a unique and expensive process. Because this is not a production- class machine, casting and similar processes that require large initial investments are inappropriate. Therefore, every component in this scissor lift is either pre-fabricated and purchased directly from a supplier (wheels, lifelines, hydraulics and motor) or made from readily available metal stock. Fabrication will involve cutting, drilling, welding, mechanical assembly and machining. Every attempt was made to reduce components requiring extensive machining; however in some cases this was unavoidable. Pre-fabricated components can be a challenge to source due to the scale of this lift. The result of this situation will often necessitate additional machining processes to customize the stock piece as necessary. The overall intent is that the vast majority of parts can be fabricated and installed locally in Thompson.

2.4 Tolerances, Fits and Welds There are some areas in this scissor lift that will require additional analysis by experienced professionals, including all bearings and scissor pins. This report provides the nominal dimensions for all such interfaces. Achieving the proper fit will typically entail machining a shaft or housing to a different tolerance; it is assured that no such alterations will affect the functionality of the design. All bearing and seal providers will be listed in the references, with catalog page numbers specified where applicable for ease of sourcing. Welding stress and distortion may be another area that will require additional investigation. It is assumed that all welding will be performed to CSA W59 standards.

2.5 Stress Analysis In order to ensure that the scissor lift is safe under all operating conditions, components are evaluated on a worst-case basis. By ensuring that a component remains safe under the most severe conditions, and then applying a reasonable safety factor (based on standards or engineering decisions), the design team is confident that all parts are sized and specified safely. Often, the worst-case scenario involves an 12 impossible situation, but one with theoretical implications which lend themselves to fairly simple stress analysis (for example, modeling a vertical railing as a cantilever beam, even though it derives some degree of support from other members). By always designing for the highest stress scenarios, even if they are not feasible, the design will be inherently overbuilt and safe. In order to significantly optimize the lifting platform, advanced methods such as finite element analysis would be required as component interactions become quite complex at this level. Because cost is not considered a driving constraint, but time is, the design team feels an overbuilt design with more material cost is acceptable for the client. Safety, of course, is the number one priority, not overall weight or lifting speed. Future extensions of this project could involve advanced analysis methods; however these are beyond the scope of the current design phase.

In the design details section, stress considerations are provided after the main design points pertaining to a specific mechanism have been discussed. Naturally, the material and geometry specifics were developed in conjunction with the stress analysis. In some areas, a great number of options were evaluated before the optimal design was discovered. Placing the stress analysis section after the component has been discussed is simply viewed as the preferred way to present our product, not as an indicator of the design process path.

2.6 Design Drawings A major deliverable featured in this report is detailed engineering drawings. Every single feature of this lifting platform can theoretically be manufactured using the generated drawings. Larger scale drawings are available upon request.

3 Details of the Design

This section covers the various design details of the ALP. Figure 2 shows the complete ALP assembly which consists of the platform, scissors, base frame, and the wheels. Section 3.1 covers the details of the platform assembly. Section 3.2 covers the details of the lifting scissor mechanism. Section 3.3 covers the details of the base frame including the secondary platform, hitch, and steering system. The height of the platform in its collapsed state as shown in Figure 2 iͻǯͷǳǤǡ platform located at the front and back, to assist workers with getting on and off the main platform. A towing hitch is located in the front of the ALP to accommodate being towed by the CAT 966H. Therefore it must be able to be steered by the CAT. The featured front caster wheels will allow a wide range of turning radii since they can rotate 360 degrees. Finally, section 3.4 covers the details of the tires and the wheels.

Figure 2. The Complete ALP Assembly.

3.1 Platform The platform of the aerial lift consists of the frame, surface, guardrails, harness anchors and access points. The frame is the central component of the platform as it provides the structural base for all the other aforementioned parts. Additionally, the frame must connect with the lifting scissors, and thus carries a great deal of stress. The surface sits on top of the frame and is the primary interaction point with the users of the lift. Guardrails are installed on all sides in accordance with safety standards and horizontal lifelines are integrated with the side railing. The front and back railings can be dropped down, which provides the necessary engine clearance as well as the primary access openings. Components that connect to the frame, but serve only as an interface for the scissors, will be separately discussed in Section 3.2. For the current discussion, the platform is broken down into sections for each of the three main components; the frame, surface and guardrail/harness system. Detailed engineering drawings of the platform can be found in Appendix A. The fully assembled platform is pictured in Figure 3.

Figure 3. Complete lifting platform with railings and access gates.

3.1.1 Platform Frame The platform frame is simple in design and purposely overbuilt. Constructed primarily of rectangular hollow structural sections (HSS), the frame is designed to withstand the bending induced by non-distributed loads and the torsion caused by loading the platform at the two corners lacking direct scissor support. For lateral surface support, ǯǤ steel beams, with no hardware necessary, and is pictured in Figure 4.

Figure 4. Render of assembled frame.

3.1.1.1 Hollow Structural Sections ͷ͹ʹͺǳ͸ǳ

1 /4ǳ -sectional profile. The perimeter sections are located in the same horizontal plane while the inner four cross spans Υǳ grating. In all cases, the beams are arranged such that there is increased depth in the vertical direction, leading to greater bending resistance. Box girders were chosen for this application due to their superior resistance to torsion when compared to the 16 varieties of structural I-beams. Although bending is considered to be the primary mode of loading, torsion in the frame as an entire unit could be experienced if the platform is loaded in the extremities not supported by scissors. The perimeter section can be cut and welded at 45° angles on the corners, thus leaving no open spaces for moisture to accumulate within the beams. The total outer dimension of the frame is 241-Φǳ 240-Φǳn corresponding to the side which is perpendicular to the cross spans.

ǡͶ͹ǳ ǤͶ͹-Φǳ separates the outermost cross span from the perimeter beam, with all measurements being made center to centerǤ Υǳ the grating. This unique feature is due to the grating being placed in an angle iron ǡ Υǳȋ ȌǤ be no open space for rain or snow to enter the cross spans, hence any gap created by raising the center spans relative to the perimeter must be fully filled during the joining process.

3.1.1.2 Angle Iron Perimeter The final aspect of the main platform frame is the angle iron perimeter. The unequal leg ͵ǳʹ-ΦǳΥǳǡ top of the perimeter in such a position that the outside of the short leg will be vertically flush with the outside of the box beam. The angle iron can be fillet welded on the inboard side, ensuring the weld does not exceed the thickness of the leg. Exceeding the thickness would cause problems with the seating of the grating. Because the angle iron does not experience significant stress aside from the transfer of compressive force directly to the boxed frame, a full seam weld is not required. It is recommended that the judgment of an experienced welder be utilized in determining the required weld pattern for securing the angle iron to the box beam. The material of the angle iron is not as critical as that of the box beams; however A572 steel is still recommended. As seen in Figure 4, the angle ir Ǥ ͸ǳ ͸ǳ͸ǳ Ǥ 17 welded directly to the frame, the corner posts will be discussed in Section 3.1.3 - Guardrails, Gates and Harness Anchors.

3.1.1.3 Stress Considerations The worst-case scenario for the platform is concentrated mid-span loading. While the grating will always be evenly distributed, the constituents of the rated load (personnel and equipment) are free to position themselves in any possible position. If we assume the worst possible scenario is the entire rated load (3500 lb) acting as a point mass along the center of a single beam and factor in the contribution from the grating, we may arrive at a maximum stress value for each beam in the frame. The maximum bending occurs when the scissors are fully contracted, such that they provide pinned support at the outer edges of the platform. This corresponds to an approximate span of ʹͶͲǳǤ ǡ ǡ basis for conservative stress analysis.

The chosen grating has a mass of approximately 4.5 lb/ft2. ʹͲǯʹͲǯ thus has a weight of 1800 lb. The grating weight will be distributed mainly on the cross spans, although a certain portion of the weight will certainly be supported by the outer perimeter. For the sake of a worst-case scenario analysis, it is assumed that the four cross beams each take a quarter of the grating weight, or 450 lb. Performing analysis under these circumstances, it is found that the maximum shear in the beam is 1975 lb and the maximum bending moment is 18,625 ft-lb. Correspondingly, the average shear stress (under maximum shear) is 1.18 ksi and the maximum bending stress is 14.3 ksi. Clearly, the shear stress is negligible when, coupled with the fact this small maximum shear will occur at the welds, indicates we need not be concerned about the welded joints (from a shear strength perspective) for the platform. The bending stress is much more significant. Given that the yield strength for A572 steel is 50 ksi, this represents a 3.5 factor of safety on the bending stress. We arrived at this value of stress considering an absolute worst-case scenario; practical loadings would therefore push our actual safety factor well over 3.5. This exceeds the requirement of a safety factor of 2 on all ductile load-bearing materials set forth in [2]. Furthermore, the contributions of the 18 channel sections that allow the scissors to actuate have not been considered; their inclusion will only serve to strengthen the entire frame. Appendix B provides the appropriate background information and calculations regarding the above values.

3.1.2 Platform Surface In order to provide excellent grip and maximum permeability of snow and rain, the selected surface is swaged aluminum grating with a serrated top. This type of grating is commonly available in long sections and is easily cut to size, ensuring custom fabrication can be done on site. Grating of this type often has an open area of up to 80% which provides almost no opportunity for snow accumulation. The frame design ensures that the unsupported span for any particular section remains minimal and thus Ǥ ʹͲǯgrating span is shown in Figure 5, with a close up view of the serrated bearing bars.

Figure 5. Illustration of serrated bar grating.

Steel grating was the initial choice, but it was found that choosing aluminum could save almost 3000 lb of weight. This became an absolute necessity when the analysis of the scissor forces showed that the platform weight was simply too great.

3.1.2.1 Grating Geometry Geometry in the horizontal plane of the platform is governed by the overall requirement for a total platform area of 400 ft2. The depth of the grating, however, is largely dependent on the unsupported spans and the tolerance for deflection under high loads. One of the primary reasons for including four cross members of rigid box ͵ǯ͸ǳǤ such that the bearing bars run perpendicular to the supporting cross members. Sectio ʹͶ͵͸ǳǡʹͶǳ size due to greater compatibility with our platform size. The length of the grating is ǡʹͲǯǤ tolerances are +/- Υǳ Ǣ further in the following section.

3 The required profile of the grating is 19-Ͷʹǳ /16ǳǤ

19 the bearing bar centers are located /16ǳ

3 ͶǳǤʹǳ /16ǳǤ size, and the particulars regarding the selection of this geometry are discussed in the following section. Aluminum retains an attractive appearance throughout its life cycle but could also be painted with the ǯ r scheme.

3.1.2.2 Stress Considerations Proper grating geometry was selected using manufacturer-provided load tables. With knowledge of our maximum unsupported span and a worst-case loading scenario, it is a simple matter to select an appropriate grating style. The load table for aluminum grating can be located in Appendix B.

As established by the fixed frame geometry, the maximum unsupported span falls under ͵ǯ͸ǳ Ǥ concentrated mid-span load per foot of grating width. It is conceivable that three ͵ǯ͸ǳͳǯǢ d load but rather a well distributed one. Regardless, if we use a purposely overstated 20 weight of three workers plus equipment (~1000 lb) as a concentrated load baseline and select the grating based upon this assumption, we may feel confident that the grating will easily handle any realistic load with a large margin of safety. This is the

3 3 methodology used to arrive at grating with a 19-4 1- /4ǳ /16ǳǡ support 1105 lb of concentrated mid-span load per unit width. This exceeds our target of 1000 lb. Furthermore, the maximum deflection under 1105 lb of mid-span loading is only a quarter of an inch.

Due to the serrated nature of the bearing bars, grating bar manufacturers advise to Υǳ of bearing bar depth. This

3 ʹǳͳ- /4ǳ acceptable in the above analysis.

3.1.2.3 Fabrication and Installation Additional considerations affecting the installation of the grating include dimensional tolerances, cutting procedures and fastening protocol. In the length-wise direction, there will be no tolerance build-up as each sheet comes pre-cut to the desired length of ʹͲǯ ΪȀ- ΥǳǤ yond the nominal ʹͲǯ ȋ Ȍ ΥǳʹͲǯǤ However, the potential for tolerance build up exists in the width direction, since 10 panels will be stacked side by side. To counter this possibility, it is recommended that one section of angle iron (one of the sections parallel to the bearing bars) is not installed until all the panels of grating are laid onto the platform. Then, once the tolerance build up can be physically observed, the final angle iron section may be installed in the appropriate position; snug to the grating. If the tolerance build up is such that the angle iron begins to hang off the outboard side of the platform, it is recommended to trim the final panel in order to keep the angle iron within the perimeter of the boxed frame. There is no possibility of the grating being too far ǡ ͸ǳs a contact point.

Cutting of the grating will be necessary even if the panels are dimensionally perfect in order to accommodate the side railings. It is recommended to lay up the panels prior to cutting and place the assembled railings in the appropriate position to visually assess Ǥ ͸ǳ ͸ǳ ȋͳǤͻͲǳȌǤ and aft sections of the platform will not need to be cut, as all railing and gate components are located outboard of the platform.

Finally, the grating must be secured to the intermediate beams. Welding is the best option for steel grating; however aluminum cannot be welded to the steel cross beams. It is therefore recommended to use saddle clips throughout. This type of fastener is illustrated in the following graphic.

Figure 6. Recommended method of securing aluminum grates to steel box beams [7].

3.1.3 Guardrails, Gates and Harness Anchors Due to the interdependency of design and function, guardrails, gates and harness anchors will be discussed together in this section. The fore and aft sides feature guardrails that drop down completely below the level of the platform surface to allow the lift to operate in close proximity to a mounted engine. It was identified early in the

22 design process that it would be advantageous to incorporate the access gates into these drop-down sections; the three center sections at both the front and back (identified by their two center railings and absent toe board) serve exactly this purpose. The side railings remain fixed at all times, which allows for the safe and convenient running of horizontal life lines. Figure 7 depicts the entire guardrail section in high visibility yellow paint.

Figure 7. Render of platform frame with complete guardrail system installed.

3.1.3.1 Drop Down Gates In order to allow the lift surface to reach mere inches below the bottom of the engine, the front and back railings need to be adjustable, in some form or another, from their current orientation. In earlier phases of the project, it was decided by the design team that these gates should drop down instead of swinging horizontally or being removed entirely and stored on or off the platform. The drop down design requires little compromise, and has the following advantages: x The guardrails are light and can be lowered by a single worker 23

x The guardrails function as access ladders in the lower position, with adjacent rails providing gripping surfaces to maintain three point contact x When dropped down, the railings will not impede any other function of the platform. Other designs could interfere with lifeline systems, create trip hazards, etc. x ͸ǳǡ shaping of the railing height around the contours of the engine x The drop down feature is controlled by simple mechanical devices with little chance of failure due to exposure to harsh elements

As evident in Figure 7, there are two distinct forms of front and rear guardrail sections. They do, however, share several common characteristics. Both are fabricated from 6061 T6 aluminum which is widely available in pipe form, has good weldability and exhibits some of the most impressive mechanical properties amongst the aluminum alloys with a yield strength of 40 ksi. The pipe sections for all vertical tubular members which slide through sleeves is NPS 1-½ schedule 80, which has an outer diameter of

1 ͳǤͻͲǳ ͲǤʹͲͲǳǤ ͳ- /2 schedule 40 profile, as they are not subjected to the same theoretical stress. Vertical portions of the ladder sections also use schedule 40. Weight is a major factor in this area of the platform design, with 6061 T6 having a linear weight of 0.935 lb/ft for schedule 40 and 1.250 lb/ft for schedule 80.

The method of attaching both is also identical. The vertical posts are allowed to slide up and down in aluminum sleeves which have a slightly larger inner diameter than the outer diameter of the posts themselves. This ensures there will be no binding under any temperature condition, but the clearance is kept tight enough that there the rail retains a sturdy feel. As both the railing pipe and sleeve are made of the same material, the difference in clearance between the two walls due to thermal expansion and contraction is minimal. NPS 2 schedule 40 is used for the sleeve, which has an inner ʹǤͲ͸͹ǳǤ ͲǤͳ͸͹ǳǡ ͲǤͲͺ͵ͷǳ 24 coaxial. To minimize the inward/outward rocking movement when the railings are at ǡͺǳter constrain the motion of the railing. In their unpainted condition, it is estimated that the very top of the railing will move

7 through a lateral distance of approximately /8ǳ being dead centered in the sleeve. While this movement is substantial, it is expected that the durable paint used to coat all members will remove a majority of the slack, and brief experimentation may be necessary to determine the optimal coat thickness to achieve negligible rocking motion before the lift is put into service. Another option is to move from a schedule 40 pipe to a schedule 80 for the sleeve, which increases the wall ͲǤͲ͵ͻǳǢ͹͹Ψ Ǥ If the tubing is to be left bare, this should be considered an option.

The sleeves are attached to the platform frame using angled aluminum as an interface. The aluminum sleeve is welded to one end, while the other is bolted to three tapped holes in the box beam using 5/8-ͳͳͳǳrade 5 bolts (along with appropriate washers). The aluminum angle section allows for a reasonable outboard offset of the gates, hence reducing the pinch hazard which exists between horizontal gate tubes and the platform edge as the gate is lowered. The aluminum legs match the length of the ͺǳ ͵ǳʹ-ΦǳΥǳǡ to the frame. The general concept can be seen in Figure 8.

Figure 8. Image of the sliding gate sleeve with associated installation hardware.

Also visible in Figure 8 are the holes that govern the various incremental heights the Ǥ͸ǳ Φǳ pin. The locking spring latch is located on the upper portion of the sleeve and when engaged, provides a solid mechanical connection between the sliding rails and the fixed sleeve. Each leg has an associated pin; therefore the correct procedure to drop the gate is to pull back and lock the first pin, securely grasp the gate to prevent falling, then pull back and lock the second pin with the free hand. Having opened and locked the pins, the gate may be lowered carefully. Applying light outward or inward pressure on the side of the gate prior to pulling back the last pin is recommended to jam the gate in its sleeves temporarily. Removable clevis pins below the sleeve prevent the railing section from being completely removed unintentionally but still allow the capability of detaching the railing from its sleeves if so desired. Pulling the gate up to the point of clevis pin contact will automatically line up the latch pin with the final hole for easy gate locking.

Access points do not require toe boards [2]. This is an advantage, because toe boards would require additional weight to be added to the center sections. The outer segments, while still capable of dropping down, are specified to include toe boards (which must be

ͶǳȌ the lift (they lack the ladder capability of the center sections). To compensate for the outboard offset of the sleeves, the toe boards are mounted to short horizontal tubes that move them very close to the frame with just enough clearance to slide down. Also outlined in [2] are the requirements for the ladder bar spacing. The distance between any two adjacent rungs must not be more ͳʹǳȋ Ȍ more than ͳͺǳǤ ǡ feet, an intermediate platform is integrated with the base in order to utilize the drop down gate. The drop down access gates weigh approximately 20 lb each while the weight of the outer drop down railings is approximately 29.5 lb; a value still within the realm of one-man raising and lowering. It is recommended that adhesive strips be installed on the rungs of the access gates to facilitate slip-free entry and exit.

3.1.3.2 Permanent Side Rails and Corner Posts The side rails are not required to drop down, which simplifies the design considerably.

1 NPS 1-Φ ͶͲǡ͸ǳ͸ǳ /2ǳ box beam is placed at all corners. Both the corner posts and the tubular members are made from A572 steel. Aluminum is not required here since the weight of the railings is no longer an issue. The large corner posts are necessary to provide suitable anchorage points for the horizontal lifelines, discussed in the following section. Each corner post is Ͷ͸ǳǡ sides are flush with the frame box beams. Because of the critical nature of the posts relating to the fall arrest systems, it is recommended to attach additional support plates on the outboard side of the frame to the two perpendicular sides. The plates,

3 ͳͲǳͶǳ /8ǳǡͷ͹ʹ fastened. Welding is recommended to avoid the introduction of stress concentrations in areas of high bending stress, should a fall ever occur. It is also recommended that the open tops of the corner posts be capped to prevent moisture accumulation. Thin sheet metal can be used here and bent into the required shape. The cap can then be mechanically fastened to the post; the particular method of fastening is not specified as

27 the cap bears no load. Self-tapping screws on all four sides are recommended. Both the support plates and capped posts are shown below in Figure 9.

Figure 9. Corner post supporting plates and caps.

Three intermediate vertical posts are welded directly to the angle iron and provide support for the horizontal guardrails, with equal horizontal distances between each one ȋͷ͹ǳȌǤ to be cut at these points. The horizontal guardrails may then be attached between the corner posts, using the intermediate vertical posts as support. All connections are welded, with the exact pattern at the junction of tubular members left as a decision for ǤͶʹǳ ǡ a fairly wide margin of error (+/- 3) [2]. The middle rail would then be placed at the halfway point between the upper rail and the grati Ǥ ǡͶǳ

1 high must be run from corner post to corner post, with a thickness of /8ǳǤ can also be viewed above in Figure 9. The gap between the bottom of the toe board and Υǳǡhowever butting it directly against the grating surface is acceptable. 28

3.1.3.3 Horizontal Lifelines Incorporated into the side railings are several horizontal lifelines. Using multiple lifelines instead of fixed anchor points has definite advantages, including increased mobility and the ability for workers to pass over each ǯǤ essentially limitless within the confines of the platform using this type of set-up. Furthermore, such systems are commercially available, easy to install, and guaranteed to meet important safety standards. One such system is the Miller Xenon permanent horizontal lifeline kit. It is recommended that this kit, or equivalent, is used for the following reasons: x Simple anchors can easily be installed into the corner posts by drilling holes and fastening a screw to the anchor bolt x Low profile shuttles will easily clear the vertical guardrail posts, which are in relatively close proximity to the lifelines x ͵Ͳǯ x This system incorporates a shock absorber, tension indicator, tension adjuster and fall indicator all in one unit

5 The system uses a /16ǲ rope and D-bolt anchorage connections, which can serve up to four workers (with an additional shock absorber). However, our design calls for Ǥ͵Ͳǯ ǡ size during installation. The Miller Xenon system is pictured in Figure 10.

Figure 10. Miller Xenon permanent horizontal lifeline (used with permission) [8].

The corner posts are designed to handle up to three lifelines per side, thus giving us a total of six lifelines (one for each worker). The D-bolts would not be mounted as shown in the above figure but rather rotated 90° such that the bolt is loaded in tension and not shear. The bolts would then run parallel to the horizontal guardrails (as well as the

5 lifeline itself). Each bolt is /8ǳͷǳǡ Ǥ͸ǳ͸ǳ ugh space to thread the nut onto the bolt where it comes through the wall. Naturally, the installation of the lifeline must precede the attachment of the corner post caps. Vertical spacing of the lifelines is entirely up to the users of the lift, as the posts have been selected on the basis of allowing for any loading condition. The only major requirement is that enough space is provided to allow the passing of one shuttle over another, however more space is recommended to reduce any interference between the harnesses attached to the shuttles. The lifeline should not exist in the same horizontal plane as the upper or middle guardrail for clearance related issues. Horizontally, the lifelines anchors should be bolted as close to the center of the corner post as possible ȋ͵ǳȌ Ǥ instructions can be found in [8].

3.1.3.4 Stress Considerations There are two important stress requirements which must be met by the guardrail and lifeline systems. Every rail must be capable of withstanding a concentrated load of 300 lb applied at any point, in any possible direction, without yielding [2]. Similarly, each anchorage point must withstand a static force of 3600 lb for each person attached to the lifeline without reaching ultimate strength [2]. For the horizontal guardrail system, it is clear that the most stressful load is one which is applied at the center of a rail, in any direction, thus inducing the maximum bending moment. For the lifeline anchor, the worst-case scenario is one which maximum bending and torsion are present. Bending is a result of the lifeline anchor being placed high up on the corner post, while torsion is induced due to the fact that the anchor is located at the perimeter of the corner post, thus introducing a moment arm about the center of the thin walled section.

For the side horizontal guardrails subjected to a mid-span load of 300 lb, the resulting bending stress would be 13.3 ksi, leading to a safety factor of 3.76 based on the yield strength of 50 ksi for A572. The situation becomes more complicated if we apply the force at the top of a vertical post, which would theoretically place the entire moment on the base of the tube, resembling a cantilever type load. Under this scenario, the bending stress is 38.6 ksi, which results in a less than desirable safety factor. However, considering the fact that each vertical rail is supported by both the top and mid railings, the above situation of pure bending about the base of the post is not possible since one cannot load a vertical post without transferring some of that load to the attached structure. We thus concluded that the side rails will remain safe under the guidelines of [2].

For the aluminum drop down gates, the safety factor is not as large due to the necessity of keeping the weight reasonable. The outside sections, which are longer than the middle gates, are very close (less than two inches) to the same length as each span of the side rail. Hence the mid-span bending stress is similar with the only differences being that we are now dealing with aluminum instead of steel and using a slightly

31 thicker wall for the vertical posts (schedule 80). With 6061 T6 yielding at 41 ksi, the safety factor remains over 3. The same logic applies to the ladder portion, however the span is even smaller and thus the bending stress lower.

Applying 300 lb directly to the top of a vertical tube represents the most dangerous situation: a cantilever beam. If we assume that the single vertical tube carries this force alone, the bending stress will be 30.6 ksi, representing a safety factor of 1.34. Because these access points are not technically load carrying members, they are not subjected to the same requirement of a safety factor of 2 [2]. The holes through which the spring latch locks represent stress concentrations, however they are located along the neutral axis (if the tube is bent in or out) and thus should not play a significant role. It is impossible to bend the gate side to the side without engaging more members of the structure. Finally, the likelihood of applying this force to a single post, given the close proximity of the other vertical posts, is small. For these reasons, it is not deemed necessary to increase the strength requirement of the front and back gates, as it already exceeds the standard for side rails and weight is a serious concern here [2].

The most critical component of the platform is the corner posts, which are subjected to demanding requirements. Because these posts serve as harness anchor points, they are required to handle 3600 lb for each person attached without reaching ultimate strength [2]. Previously, we had explained how the vertical position of the horizontal lifeline ǯǤ ǡ corner post will in large part be determined by how high each anchor point is, we must establish a worst case scenario such that any vertical mounting pattern of the anchors is acceptable. Recalling that three lifelines run between the corner posts, it is easy to see that if the post can be safely loaded by all three anchors located at the very top, any configuration is possible since only one anchor can actually be at the top. The remaining two anchors must be located further down the post; a less stressful situation for bending (the dominant stress consideration). An equivalent situation to our worst case scenario would see three workers attached to the single top lifeline. This is possible

32 with additional shock absorbers on the lifeline, which provides yet another way to customize the lift.

The D-bolt anchors are supplied by the company which produces the lifeline and are Ǥ ǡ Φǳ considered a likely failure mode with an appropriate sized backing washer, which is supplied with the lifeline kit. The bottom of the beam is welded to the frame and secured with two additional thick plates, hence the likely failure mode is the beam itself. The anchor attachment must be capable of handling 3600 lb of force per person in any direction [2]. Logically, the D-bolt anchor can only be realistically loaded in a finite number of ways, considering how a fall would occur. However, to comply with the safety standards, the most aggressive loading situation is considered: one that places the beam in bending and torsion. This would correspond to the D-bolt being loaded perpendicular to the lifeline and parallel to the platform. Bending occurs because the ȋͶʹǳȌ torsion is also a factor because the force is applied at a radius from the sectional centroid. Applying 10800 lb in this manner (equivalent to 3 x 3600 lb) produces a maximum normal stress due to bending of 26.9 ksi while the torsion results in 1.43 ksi. These forces compound at the extremity of the box beam, but given the small torsional stress value, the maximum stress will not deviate much from the bending value. Given that A572 has an ultimate tensile strength of 70 ksi, a safety factor exceeding 2 is achieved on the worst-case scenario. Thus the corner post has been validated as a suitable anchor point for three workers (one on each lifeline, or any equivalent ǯspecifications). Appendix B may be consulted for additional details on the stress analysis if desired.

3.1.4 Summary of Lifting Platform To summarize, the lifting platform consists of three main aspects; the platform frame, the surface and the guardrail/lifeline system. The frame is fabricated from HSS and angle iron, and is overbuilt to deal with unpredictable dynamic loads. The surface is serrated aluminum grating, which keeps the platform relatively light, provides excellent 33 grip and allows rain and snow to pass through with ease. The guardrails, which serve dual purposes of fall prevention and access openings, are manufactured from steel and aluminum and meet all relevant safety standards [2]. Side guardrails are fixed in place and attach to large corner posts. The corner posts provide suitable anchorage for three horizontal lifelines, and because the side railings do not need to be removed, the lifelines are permanent in nature. Because each worker is allocated his/her own lifeline, they can pass each other without having to unhook and re-attach. The front and back railings are made of aluminum and can be dropped down in increments of six inches to accommodate an engine close to the platform surface. The three center gates are smaller in nature and when dropped down serve as access ladders. A simple locking spring bolt provides the mechanism to move the gates up and down. The vast majority of the platform is welded together, with some mechanical joining of dissimilar metals required. The entire platform is maintenance free and will provide years of service for the client.

3.2 Lifting Scissors The lifting scissors were the most challenging part of the lift to design. Although they appear mechanically simple, the design process involved dozens of different approaches ranging from different actuator placement, alterations of scissor member length and considerations of different member cross sections. The large required rated ͵ͷͲͲʹͲǯʹͲǯ capacity of traditional scissor lifting mechanisms to the limit. The final product is two paired scissor stacks, with each pair supporting opposite corners of the platform. The scissors feature three levels, and each stack requires its own hydraulic ram. PTFE composite bushings at all pins ensure maintenance-free pivoting action, and PTFE sliding pads allow smooth linear motion of the unfixed members as the platform moves up and down. Minimum and maximum angles, with respect to the horizontal, are 9 and 45 degrees respectively. The ʹͻǯ-ʹǳ ǡǤ ͻǯ- ͷǳǤ Renders of the extended and collapsed scissor stack pair are provided in Figure 11 and Figure 12. Detailed drawings may be accessed in Appendix A. 34

Figure 11. Extended scissor pair.

Figure 12. Completely collapsed scissor pair with inner cross-bracing visible.

The following sub sections will delve into the main aspects of the scissor design. This includes scissor members, pivots and slides, and stress considerations. A summary will then be provided to tie the concepts together.

3.2.1 Scissor Members

5 ͳͶǳͳͲǳ /16ǳǤ These large and heavy beams are necessary to resist the incredible bending moments produced by the scissor mechanism. The box beams must be fabricated from steel with high yield such as CSA 70T, which yields at 70 ksi. A typical structural steel yielding at 50 ksi will not provide the mandatory safety factor of 2 on all ductile load supporting members [2]. Procurement of this material might prove difficult or costly, simply because 70T is commonly used for I-beams but not HSS.

ͳͷʹǳǡ ͶǳǤ ͳʹǯ-to-pivot length used in the sizing calculations. The advantage of using shorter members includes a lower collapsed height, more room on the platform to place hydraulic rams and less bending stress. The downside of short members is that less support is available in the fully extended mode (for a given number of levels). Short members require more angle (measured relative to the horizontal) to achieve the total required lift height and although stress in the members reduces substantially as angle increases, the supporting distance also Ǥͳʹǯ n the fully extended position. This might not be acceptable in a conventional scissor lift, as over half the platform would be overhanging the inboard pivot. However, our design spreads out the load by using four scissor stacks, with each pair supporting opposite corners of the platform. Hence there is no serious overhang issue, and the platform frame has been overbuilt to account for any torsional effects incurred by free hanging corners.

One option discussed was decreasing of scissor member cross-sectional size at higher levels of the lift. This strategy is plausible since the bottom level deals with the greatest 36 amount of force and the stress decreases substantially for every higher level. Although the weight savings are attractive, this optimization requires difficult folding geometries and was scrapped when analysis showed that using smaller upper levels did not significantly reduce the stress. Theoretically, a steel yielding at 50 ksi is acceptable beyond the bottom level of the scissors, which could be an ideal option if 70T is very costly in HSS form or has unrealistic lead times for fabrication of every single member.

Collapsed height became a very important concern during scissor design. Obviously, the lower to the ground the lift collapse, the easier it is to mount. The inherent problem is that of angle; it can be shown that collapsed angle has a dramatic effect on the force necessary to begin the lifting process. As an example, trying to lift a single scissor requires approximately 120,000 lb of force at seven degrees from the horizontal; this value skyrockets to 170,000 lb if the angle is decreased to five degrees. Not only do the stresses become unacceptable in the scissors, but the size and operating pressure of the hydraulics also increase beyond reason. Thus, if a premium is placed on the collapsed height, a significant price will be paid in terms of material and design complexity. The scissor lift may require four pairs of scissor stacks, with a single actuator operating each stack. To keep the situation reasonable, the design team feels that a higher collapsed height is acceptable, so long as adequate mounting provisions are provided. The scissors themselves are theoretically capable of collapsing to approximately 5.5 degrees, which is a function of member cross-section and pivot placement. However, they will need to be held at a higher angle in the lowest position for the reasons provided above. Access ergonomics will be discussed further in Section 3.3.

Research into large lifting platforms revealed that collapsed height will always be a sacrifice. An example of a lowered heavy-duty platform is provided below in Figure 13 to illustrate the scale. While this particular lift is capable of operating at much higher heights, it also carries far less payload. This makes it adequate for a rough visual comparison.

Figure 13. Holland Lift M-250DL27 at collapsed height [9].

ǡ ͶǳͶǳ

3 /16ǳ Ǥ ʹͶǳ at convenient intervals, four per member. The two pairs of scissors are not tied together as they do not move upward in the same plane. The material for these braces does not need to be 70T; more common steels such as A572 are recommended so long as welding compatibility exists. Cross member location is illustrated in Figure 14.

Figure 14. Render of scissor stack with stabilizing cross members (blue). 38

3.2.2 Pivots and Slides The pivots that connect scissor members are loaded in single shear and thus must be sized appropriately. The size determination simply involves looking at the joint reactions of the scissor structure, which will be discussed in the following section. A stronger steel, such as CSA 70T, is necessary. Center pins, in general, carry much more ͶǳǤ͵ǳǡ including the upper- and lower-most connections that attach to the platform and base frame clevis mounts, respectively. Along a similar line of reasoning, the upper pivots do not need to be as large as the smaller ones, but the weight savings associated with moving to progressively smaller pins is minimal when factoring in the overall weight of the scissor structure. To maintain symmetry and ease assembly, constant pin sizes are specified on all levels of the scissor mechanism.

Both the center and end pins at scissor-to-scissor joints are 23-ΦǳǤnsion provides 3-Φǳ ȋ ͳͲǳ wide). This is necessary for two reasons; space must be left between scissor members such that they do not rub against one another and enough free shaft on either side must exist to attach retaining rings. The retaining rings are not load bearing elements, and simply exist to ensure that a pin does not try and work itself out of the scissor bores. The standard recom͵ǳͶǳ (depending on which pivot we are talking about) should be followed. No stress concentrations will be created since the grooves will be machined on the free ends of the pin. Pins at scissor-to-clevis joints will be shorter in length (14-ΥǳȌ needs to span one box beam and two clevis plates.

Simply drilling holes in the box beams and inserting pins would not provide enough bearing surface to spread the immense load across the PTFE composite bushings. To compensate for the small surface area of the box beam walls, bushing bores made from mechanical tubing is first installed. This allows for full bushing contact on the pin throughout the box beam. Figure 15 below illustrates the concept.

Figure 15. Scissor member with bushing bores and PTFE bushings installed.

The bushing bore is easily visible at the open end of the scissor beam. End pins, which ͵ǳǡ͵-ΦǳͲǤͳͷ͸ǳǤ are welded to the beam, inside and out. The split PTFE bushing, available in 3-Χǳ ͲǤͲͻ͵͹ͷǳ ͵ǳǤ bushings equals the length of the bushing bore tube (11-ΥǳȌǡ eads out the load. Naturally, the center pin requires a larger bore tube, with an outer

5 3 diameter of 4- /8ǳͶ- /16ǳǤǡ

1 available size, but pipe with a 4- /8ǳ Ǥǡhe

1 middle bushing bore pipe needs to be bored out by /16ǳǤThe bushings themselves retain the same length and thickness. Clevis connections are mechanically and geometrically similar to scissor-to-scissor joints.

The pins should not experience any axial thrust due to the constraints imposed by the clevis pivots and the sliding guides. This means that the bushing bores, which may touch one another, are not likely to wear away and if reasonable tolerance is provided, they may not even touch in the first place. PTFE thrust washers, which would need to be

ǯ ǡ bores to ensure an almost frictionless lifting and lowering action. There are companies available to manufacture bushings to custom dimensions [10].

Looking closely at the open end of Figure 15, it is evident that an additional plate has been added. This is to prevent a shearing pull-out situation of the pin bore and is only

3 necessary at the ends of the beam. The plate is /8ǳ steel and must be securely welded in position. These support plates are clearly shown in Figure 16.

Figure 16. Close up view of installed bushing bore with support plates.

PTFE bushings were chosen for the pivots due to recommendations provided by an experienced power equipment mechanic [1]. Composite PTFE bushings also have excellent thermal stability and can be operated in temperatures as low as -200° C [11]. Under low velocities and high specific loads, the coefficient for PTFE can fall below 0.05; lower than common bronze bushings [11]. For these reasons, this type of bushing is also used as the sliding pad on the scissor arm ends that must move horizontally as the platform lifts and lowers. The clevis connections remain the same as the fixed pivots; however when sliding is required a PTFE pad is welded to the bottom of the clevis bracket. Conversely, fixed pivot brackets are simply welded to the appropriate frame or platform section. An example of a sliding clevis with a PTFE pad is provided in Figure 17. 41

Figure 17. Clevis connection with PTFE slide bushing.

Welding of the PTFE slide bushing is possible because they generally come with a steel backing plate. Although the scissor lift generates tremendous horizontal forces at the lower level, the pad itself need only overcome the force of friction. Because friction is product of normal force and coefficient of friction, the actual frictional force exerted on the pad surface is quite low. The maximum downward force on any particular scissor will rarely exceed 5000 lb. This, of course, is highly dependent on load distribution but even in the case of high load, the frictional force would only amount to approximately 250 lb.

The pads slide in C-channel shapes rigidly attached to the base and the platform. The C- channels provide lateral support and prevent the scissors from extending in a crooked fashion. Supporting the weak C-channels are full lengths of W8x15 I-beam fabricated from A572. These run all the way across the platform, as do the C-channels. For proper sliding action, thin hardened metal plates must be attached to the C-channels. This provides the proper interface for the PTFE slide bushings, and the plate need only be as long as the maximum stroke of the hydraulic ram. Hardened plates should be welded on the inside of the top and bottom C-channels.

Because the pins rotate in the bushing bores of the scissor members, no relative rotation between the pins and the clevis bores is desirable. Figure 19 shows retaining rings on the very ends of the clevis pivots, similar to the joints between scissors. Regardless of how the pins are secured, the pin should never rotate in the clevis bore because it would much rather rotate about the PTFE bushing. Instead of retaining rings, which simply ensure the pin does not work its way out, the pins could be welded to the clevis on the outside. Thrust washers may be needed here between the scissor and the clevis plate on both sides. The fixed clevis mounts at the bottom, including those that secure the hydraulics, are substantially larger than the sliding ones due to the tremendous forces developed in the lateral direction.

Figure 18. Sloped lower clevis mounts.

Maximizing the contact surfaces via a sloping backside (Figure 18) provides greater weld density, reducing the chance of failure. All lower clevis mounts are fabricated from

3 ͳǳǡ /8ǳǤThe actual size of the hydraulic mounts, both at the frame and on the scissor, can be specified via collaboration with the hydraulic cylinder manufacturer keeping in mind that this mount experiences the exact same force as the fixed ones.

Figure 19. Translating scissor ends with pins and clevis mounts installed.

3.2.3 Stress Considerations Analysis of the scissor members relied heavily on the methods developed in [12]. It is highly recommended that this document be consulted for any additional scissor analysis. Using equations of static equilibrium, it is possible to find the forces in all joints for every level of the scissor lift above or below the level which houses the actuator. Because our actuator placement is at the ground level acting horizontally on the pin, it can be shown that actuator force must equal the static equilibrium joint force if the scissor assembly was otherwise constrained by a fixed pivot. It is possible to solve for actuator force for any general placement, however the systems of equations become large and must be solved by computer. Several different actuator placements were tested using a spreadsheet which calculated the stress in the highest loaded members. The general conclusion was that while changing the placement of the actuator could dramatically reduce the actuator force requirement, the stress in the scissor members did not decrease enough to justify the more complex arrangement. The current placement is made possible by the large base frame, which allows the actuator to be mounted outside of the scissor stack. If mounted inside, the actuator would have to pull and not push. This is certainly possible with a double acting cylinder; however the area available to develop force is reduced due to the piston rod. The basic scissor structure for mathematical modeling can be seen in Figure 20.

Figure 20. Basic scissor structure [12].

The major difference between the basic scissor model in [12] and our lift is the fact that we have twice as many scissor stacks. This is easily accounted for by imagining that we have two scissor lifts acting on the same mass, and we may proceed with the same type of analysis.

The forces in the members increase at levels closer to the bottom. The lowest level has the greatest amount of stress and thus must be considered the critical component. One of the reasons for this is the fact that scissors are not massless and in fact weigh approximately twice as much as the loaded platform. Lower levels of scissors naturally have more mass on top of them, generating more stress. However, even if the scissors were massless, force in the members would still increase as we move down. Refer to Appendix C for the relevant equations.

One of the biggest factors in how much stress exists in the scissors is angle. Because the tangent of the angle (measured with respect to the horizontal) appears in the denominator of both the x and y force components at the joint, it is a mathematical fact that the force in the joints (and connected members) will approach infinity as the angle approaches zero. This places a very real limit on how low we can drop the scissor stack.

Although our scissors are capable of collapsing to approximately 5.5 degrees, analysis showed that we had to limit the minimum angle to 9 degrees. This results in a higher- than-desired platform height. Considering the alternatives, which included adding more scissor pairs and even a complete redesign of the lifting system, a design decision was made to accept the increased height. Provisions have been provided to make accessing the lift as simple as possible at this higher height, including the intermediate access platforms.

Under most circumstances, the lift can only be loaded in three ways. This includes a direct force down, a moment about the x axis and a moment about the z axis. To help visualize, Figure 21 provides the coordinate geometries.

Figure 21. Scissor lift coordinate system [12].

The load Hy is by far the most dominating factor. Moments Mx and Mz occur due to the load not being perfectly distributed in the center of the scissors. By accounting for these three phenomena, it is possible to accurately calculate the forces in all joints of the scissor lift under a worst case scenario. This scenario corresponds to the entire rated load being placed at a corner.

The weight of the scissors takes into account not only the member weight themselves, but all pins, bearing bores and cross members as well. Because these are evenly distributed throughout the scissor levels, it is possible to accurately account for all of these effects.

Using the equations in [12], a spreadsheet was constructed to calculate the load at the most demanding joints. The spreadsheet takes all the inputs discussed above and returns a single value; the safety factor in the highest stressed members. Based upon this value, we validated the scissor assembly presented earlier in this section. The maximum horizontal force at the lower pins was found to be 95,000 lb. This is the force the actuators are required to provide to begin moving the lift in a quasi-equilibrium fashion since they are aligned horizontally with the lower joints. Values of 120,000 lb were used in subsequent calculations to be conservative and account for the fact that the bearings are not truly frictionless. At the center pins of the lower members, the ͳͶͺǡͲͲͲǤ Ͷǳ ͵ǳdiameter. Vertical forces are extremely small in comparison, with the highest vertical joint force equaling 5,200 lb. This occurs at the outer pins; there is no vertical reaction at the centers.

The total combined stress produced by axial loading and bending is 30.65 ksi, while the yield strength of CSA 70T steel is 70 ksi. Applying a stress concentration factor of 1.10, as outlined in [2], results in a safety factor of 2.08 for the worst case scenario. This exceeds the necessary requirement. The dynamic load factor is assumed 1 due to the fact that the analysis already accounts for extreme loading situations.

The outer pins at the base are loaded in double shear due to the clevis mounts. The 120,000 lb horizontal force only produces 8.5 ksi of shear in the pin. Similarly, the center pin which is always loaded in single shear experiences 11.8 ksi. Both of these values are well below the shearing yield of 70T steel, providing a high safety factor.

3.2.4 Hydraulics In order to facilitate lifting of the upper platform, hydraulic actuators were required. These actuators apply a driving force to the scissor sections in order to raise the height of the platform. The final placement for ram location was determined to be along the bottom of the scissors, parallel to the horizontal. By pushing the bottom scissor members together, the entire scissor section actuates vertically, raising the platform. The actuators are mounted to the base frame via brackets similar to those used for the scissor mounts, as shown in Figure 22.

Figure 22. Detailed view of hydraulic actuator.

After speaking with Hypower Systems Inc, a local hydraulics supplier, it was determined that the optimal actuator type for our application would be a hydraulic cylinder that utilizes internal position feedback sensors. It is crucial to maintain the level orientation of the upper platform, meaning that each scissor section must move in unison. This can be achieved through position feedback actuators. They use an internal

48 sensing device to determine the extended length of the actuator, and adjust internal valves to achieve the desired position.

In order to calculate the required actuator size, hydraulic power calculations were performed, as seen in Appendix C. It was determined that the required ram force for each actuator could be up to 120,000 lbs. This corresponded to a hydraulic pressure of 2387 psi. In order to achieve the desired total platform extension height, an actuator stroke of 40 inches was required, which corresponds to a total hydraulic volume of 34.7 gallons.

3.2.4.1 Hydraulic Pump and Electric Motor In order to provide hydraulic pressure to the actuators, it was necessary to specify a ǡ ǲ ǳaerial lift platform.

Manufacturers, such as Eaton Hydraulics, commonly sell hydraulic power units as a pump-motor combo, eliminating the need to spec out a pump and motor separately, and subsequently design a coupling device to mate these two components. A typical power unit can be seen in Figure 23.

Figure 23. Eaton hydraulic power unit [13].

ǡ ǲ ǳ that produces a continuous 125 HP. These units are designed to be extremely compact due to their close-coupled motor pumps, and minimize noise to 10 dBA or less. The enclosed design is ideal for the harsh climate the ALP will be facing, as it eliminates the possibility for dirt and debris to enter the coupling unit, and contaminate the junction between motor and pump. Furthermore, the lack of dynamic seals eliminates the possibility of leaks in the system, reducing the maintenance required for this system.

Following the values obtained in the previous section, we determined that the hydraulic system would require 49 total HP. However, due to inefficiencies of 85% in the pump and actuators, a total power input of 57.6 HP would be required from the electric motor. This is ideal, as the Eaton power unit would be capable of providing a safety factor of over 2 for hydraulic demand, in the event that there were unforeseen restrictions in the sliding assembly.

3.2.5 Summary of Lifting Scissors In conclusion, the lifting scissors are hydraulically driven, and an electric motor is used to power the pump. The scissors themselves are large box beams with a high yield strength due to the extreme amounts of bending produced by collapsing the lift. PTFE bushings have been used at all pins to provide maintenance free pivoting action. This same material is used to facilitate the sliding action between the scissors and the platform/frame. Extensive research into modeling scissor lifts mathematically lead to the validation of the chosen design from all stress aspects. With an appropriate motor, the scissors will be capable of lifting from fully collapsed to fully extended mode in one minute.

3.3 Base Frame The base frame of the lift is responsible for supporting the scissors, allowing translational movement, and housing the hydraulic pump and electric motor. In this section, the various main components of the base frame will be discussed individually, but with an emphasis on how they integrate with one another to yield a safe and

50 reliable foundation for the entire lift. First, the structural components will be discussed, which consist mainly large steel hollow structural sections. Auxiliary structures, such as the intermediate access platforms and the hitching mechanism will also be included here. Next, the steering system will be explained. The majority of the frame components are welded together. An isometric view of the base frame is pictured below in Figure 24, and detailed drawings for all frame components can be found in Appendix A.

Figure 24. Assembled base frame with wheels and hitch.

3.3.1 Structural Frame Components The structural components of the base frame required a complete redesign after the tremendous lateral forces generated by the scissors were discovered. Originally, the base utilized I-beams to deal with the bending induced by the weight of the platform and scissors. Once it was found that this force component was much less than the horizontal force at the pins, the platform was specified to use large box beams as the main structural backbone. The scissor connections at the outer box beams, interfaced by clevis mounts, are the highest stressed areas of the base. The long arms that connect 51 to the wheels also generate significant bending stress. Side platforms are included to facilitate the mounting of the hydraulic pump, reservoir and electric motor; the side utilized for this purpose is at the discretion of the client. Running of hydraulic lines from either side is equally simple. The opposite side can then be used for storage of items such as wheel chocks, extension cords, and tools. The base is shown in Figure 25. ǡǮǯǡ Ǯ ǯ beams run up and down, based on the orientation of Figure 25.

Figure 25. Assembly of the aerial lift platform base.

3.3.1.1 Central Frame ͳͶǳͳͲǳ thicknesses. A572 steel is specified throughout. The front and rear main lateral beams

5 are /16ǳ ǡǤ of the main beams, which are 4-ΧǳǤ

3 /8ǳ Ǥ Four of these beams are used, with the outer two tying the wheel arms together. The 52 sliding scissor members do not operate on these beams; rather, they slide on I-beams placed lower in the frame. The I-beams and associated C-channel sliding guides which mount on top are not considered crucial to the structural integrity of the frame. One of the main design goals for this base was to keep the scissor connections as low as possible. This became especially important once it was realized that the scissors must not be allowed to collapse completely for stress reasons.

Overall, the frame is 2ͶͲǳʹʹ͸ǳǡ ǤʹͶͲǳ width is considered the maximum allowable, for reasons of easy maneuverability under the test stand and removal from the machine shop where it is being built. The length dimension is not as critical, and when the hitch and intermediate access platforms are ǡ ʹʹ͸ǳǤ

The major detail of the scissor-to-frame connection was discussed in Section 3.2. In regards to the base frame itself, all lower fixed clevis mounts must be welded directly to the main lateral box beam. The four large cross beams are aligned such that the neutral axis of the beam falls very close to the pivot location. This feature was incorporated to reduce bending in the frame. The two inner cross beams run directly under each pair of scissors, respectively. It is expected that these beams will provide the majority of structural support, and they have been designed in this capacity. Mounting details are shown in Figure 26.

Figure 26. Sloped fixed clevis mounts and illustrative hydraulic mounts.

Additional supports beyond the two large beams that run under the scissors are added to increase the rigidity of the frame. This includes the outer cross beams, which attach to the vertical wheel arm beams (left to left, right to right), as well as the center cross beam. Space is very limited between the scissor pairs, thus the main center cross beam

3 Ǥͺǳ͵ǳ /8ǳ are welded on top to line up with the pivots. Again, the theory here is to reduce bending by avoiding offset of the neutral axis. All cross beams which sit on top of the lateral main beams have reinforcing plates to provide more welding surface area (visible in Figure 26). Adding gusset plates, two for each connection, further increases the integrity of the beam joints.

The wheel arms feature the same level of extreme stress as the clevis mounts. The need ǲǳ Figure 26. Furthermore, the required caster ability of the front wheels means that the horizontal offset is significant. Coupled with the fact that each wheel supports a quarter ǯ ǡ Ǥ

5 ͳͶǳͳͲǳ /16ǳ well. Additional gusset plates help support the joint between horizontal and vertical 54 beams; three at each corner. The horizontal offset provides a generous three inches of clearance when the caster wheel is swung towards the frame, which ensures any foreign material caught in the tire does not hit the frame and impede the caster action.

Structurally, there is little difference between the front and rear wheel arms. The front arms must accept the caster assembly and therefore need large holes cut through them, ȋͻǳȌǤ ǡͳͲǳͳͲǳΥǳ HSS reinforcing collars are welded in the arm outside the perimeter of the spindle bore. These should be welded in place before the horizontal arm is joined to the vertical arm in order to access both sides. The reinforcement box beam is only loaded in compression and does not require an involved welding pattern to keep it in place. Detail of the wheel arm is provided in Figure 27.

Figure 27. Details of the wheel arm with the reinforcement (upper right).

Only the front wheel arms need to have a significantly cantilevered horizontal beam to allow for a full range of caster movement. Regardless, the rear wheels must feature the

55 same tread (distance from tire centers along a common axle line) as the front, and therefore the same arms are used such that the front and back tires line up perfectly. As pictured in Figure 27, the rear wheel brackets are simply welded to the arm.

3.3.1.2 Intermediate Access Platforms Front and rear intermediate platforms have been added due to the inability of the platform to collapse sufficiently close to the ground. It is not at all uncommon for large scale lifts to have a high collapsed height, but an acceptable method of mounting and dismounting must be provided. Unlike many other lifts, which incorporate long ladders, the proposed design includes an additional platform to stand when operating the drop down gates and provides a great location for stockpiling tools before loading them on the platform. The intermediate platform, equipped with ladders and guardrails, is shown in Figure 28.

Figure 28. Intermediate access platform.

The intermediate platform is fixed to the base and does not move with the platform. Thus there is no concern with engine clearance or any similar issues. The same assembly techniques are used here as for the main platform. The grating is the same

3 size (19-Ͷʹǳ /16ǳȌǤthree-foot wide section is Ǥͳʹǯǡ ͶͲ 56 surface area. All support girders are ͷ͹ʹ ǡ ͶǳͶǳ

3 /16ǳǤ ͳ-1/2 SCH 40 steel tube, and the ladder geometry conforms to the standards regarding rung spacing and width [2].

As a featured added late during the design process, stress analysis has not yet been performed. However, many comparisons can be made to the well-analyzed main platform. For example, the railings and surface geometries are almost identical. In fact, the surface has even less unsupported span in this application. Additional bracing has been added due to the current uncertainty, including the diagonal support beams from the outer edge joining back to the frame and the short tubes which attach the ladder horizontally to the intermediate platform frame. Both of these features can be seen in Figure 28. While it is likely that the platform can support as many workers and equipment that can fit on it, additional analysis will be necessary to confirm these preliminary estimations.

3.3.1.3 Hitch The hitch consists of two main sections; the fixed triangulated section and the pivoting beam with the hitch eye. The fixed center beam is built to the same specification as the main center beam, with added side struts to distribute the load during turning maneuvers. These features are easily visible in Figure 25Ǥǯǡǡ the actual connecting bar with the hitch eye. The pivot allows this part to move up and down to engage the pintle hitch. Like the intermediate access platform, the hitch is still in a preliminary design phase and the exact loads placed on the hitch are unknown. They are likely to be quite dynamic, and FEA is recommended. The pivoting segment of the hitch connection is pictured below in Figure 29.

Figure 29. Pivoting hitch arm.

The four holes at the front plate are for attaching the hitch eye, which is a commercially Ǥ ǯ be implemented here. Additionally, a trailer jack stand device should be included with this arm due to the extreme weight of over 400 lb. This would allow the arm to be smoothly lifted and lowered for pintle engagement.

Figure 30. Hitch used on engine transporter (used with permission).

It is envisioned that the end product would closely resemble the engine transporter, pictured in Figure 30 in terms of hitching capability.

3.3.2 Steering Components In order for the lifting platform to move into position a loader must tow it. To facilitate efficient navigation through turns, the lift has been designed with a set of caster wheels that automatically set themselves to the most stable angle. The only input from the loader is the connection at the hitch, and the platform will follow the desired path without any additional steering control. The wheels have been specified from an existing manufacturer and must be quite sizeable to safely handle the load of the entire machine. The caster system was designed by the team and requires significant machining to fabricate. Undertaking this design was deemed necessary after research into existing systems proved futile. Both the caster bracket and spindle are subjected to incredible amounts of bending and are thus built out of thick plates and high strength steel, respectively.

3.3.2.1 Wheel Brackets and Spindle Assembly Our discussion of the steering system begins with the wheel brackets. These are ͳǳd provide an interface to connect the wheel hub shaft to the frame. For the non-pivoting wheels, the bracket is simply welded to the frame in a fixed position. However, the wheels which are allowed to caster must have an offset assembly to ensure enough trail is provided, where trail is the distance between where the steering axis projection hits the ground and where the tire contacts the ground. This provides the mechanism for caster movement. Both the fixed and offset brackets, as well as the axle retainer, are shown in Figure 31. The wheel axle is preliminary at this point, and discussion with the rim manufacturer is recommended (Appendix E). If stress concentrations ǡʹǳǤ Like the spindles, the wheel axle will likely need to be made from machine steel with superior yield strength. The axle simply slides through the hub and sits in the semi- circle cut-out visible in Figure 31. The retainer bolts to the bottom of the bracket arms and pinches the axle in place.

Figure 31. Renders of the wheel brackets and axle retainer.

The additional length of the caster bracket allows for the spindle to connect behind the plane of the wheel hub. The spindle shaft, which will feature a plate welded to the bottom, attaches to the bracket using mechanical fasteners. The system is designed to allow a maximum of 9-Φǳǡ provided depending on the performance of the caster system.

In order to provide effective caster movement, the wheel must be allowed to rotate 360°. This way, the platform assembly can be driven in either direction while still allowing the self-aligning action required. The spindle, which provides this rotational freedom, rotates in its housing via a pair of bearings. The lower bearing is a tapered roller bearing, which takes the substantial axial thrust generated by the weight of the platform. The upper bearing is a radial ball bearing, which helps control the extreme bending moment found in the shaft. These bearings are spaced far apart to help reduce the radial reaction forces. Because neither of the bearings is sealed, the entire cavity between the spindle shaft and housing must be filled with the appropriate low- temperature grease. A grease fitting is installed in the housing to allow for occasional re-greasing. To seal the bearings from the elements, a lip seal is installed in the housing bore below the tapered bearing. Similarly, a bolted cap with gaskets protects the top of the housing. The cap also features a bearing retainer that helps locate the upper ball

1 bearing against the housing shoulder. Because the gaskets are thin ( /16ǳȌǡ stacked on top of one another to provide the correct amount of force on the bearing race while still maintaining a positive seal. The selected bearings have exceptional load

60 bearing characteristics and should not require replacement for the life of the aerial lift platform.

The spindle is machined from AISI-C1144 round steel bar with an initial outer diameter greater than 5-ΦǳǤngth of the spindle is 16-Υǳǡ get enough bearing spread. The lower portion, which accepts the caster bracket, is 5-Φǳ

3 in diameter and 1- /16ǳǤ ǯ inner cone. The shaft is then stepped down to 4-Φǳ remainder of the length, with a fillet radius smaller than that of the bearing race corner to allow for proper seating. The tapered bearing seats against this shoulder. The spindle is shown below in Figure 32.

Figure 32. Fully machined spindle.

The radial ball bearing will be installed near the opposite end of the shaft. The retaining ring groove is placed slightly above the bearing to ensure that the spindle does not fall out of the housing in the extremely unlikely event that the wheel loses contact with the ground. Retaining rings of this size have more than enough shear capability to handle the load induced by a free hanging wheel and bracket. Aside from the retaining ring, the shaft can float axially in the ball bearing while the lower tapered bearing provides shaft location. The upper bearing is located in the housing by a shoulder below and a bearing retainer above. All bearing interfaces should be specified with a clearance fit; interference should not be necessary to locate any bearings and the magnitude of the 61 reaction forces will prevent the shaft from spinning relative to the inner race. The important features of the bearings and related hardware are summarized below. For more complete data, refer to Appendix D.

TABLE IV SUMMARY OF SPINDLE BEARINGS [14]

Timken TS single row tapered Timken BIC single row roller bearing (64450 inner / radial ball bearing 64700 outer) (45BIC206)

Bore Diameter, d (in.) 4.5000 4.5000 Outer Diameter, D 7.0000 6.2500 (in.) Width, T (in.) 1.625 0.875 Dynamic Radial Load, 16000 12900 C90 (lb) Dynamic Axial Load, 14100 N/A Ca90 (lb) Weight (lb) 7.64 3.18

The key aspect to note from the above table is the load bearing capacity. It will be shown in the section regarding stress considerations that the bearings are more than capable of carrying the load. The dynamic load ratings given in the table are based on millions of revolutions; the bearings in the spindle may only complete a few revolutions per use. In general, static load ratings are much greater than dynamic ones (for example, the tapered roller bearing has a basic static radial load rating of 94,200 lb). Using the dynamic load ratings in sizing calculations ensures that the bearings are capable of handling loads much greater than those applied by our platform. In fact, such large bearings were only necessary due to the shaft size. Radial ball bearings are not always given an axial load rating, but it is of no consequence as the tapered bearing will be responsible for the entire thrust force generated by the spindle.

Gaskets, retaining rings, and seals are all off the shelf items. Only the gaskets will require an additional modification, which simply involves cutting bolt holes in them. All 62 parts have been sized according to the shaft and housing and should never require replacing.

In the unlikely event that the spindle needs to be removed for any reason, such as bearing replacement, the procedure is simple. After the wheel bracket and has been removed, the upper cap is then unbolted. Removing the retaining ring will allow the shaft to be removed from the bottom of the housing.

Like the spindle, the housing is also machined out of a solid round bar of steel such as A572 since the housing does not require the same high yield strength as the spindle itself. Typically, this type of part would be cast and then the critical areas, such as the bearing housings, would be machined to exacting tolerances. However, for the reasons stated earlier regarding the feasibility of casting one-off parts, it will be necessary to machine the housing. A cutaway view of the spindle housing is provided in Figure 33.

Figure 33. Machined spindle housing.

7 The total height of the spindle housing is 15- /8ǳͻǳǤ machining required is more of a brute force approach rather than intricate; once the main bore has been milled only the outer ends need to be machined to a wider diameter. There are no retaining ring grooves in the housing. In addition, the upper surface will need six holes drilled and tapped tͳǳ plate bolts. The cover plate, shown below, uses sections of standard pipe as a bearing retainer. The pipe is cut, welded to the plate, and serves the function of holding the upper bearing against the housing shoulder.

Figure 34. Cover plate with bearing retainers installed.

The complete spindle assembly is a very large, strong and crucial part of the aerial lift platform. It has been designed to provide years of trouble-free service. For assembly, a load-bearing ͳǳ housing, which then contacts the bottom of the box beam arm and transfers the load over a wider area. The entire assembly can now be welded to the box beam, and the spindle installation is complete. With the spindle assembly now in place, the front wheels can rotate a full 360°, allowing for effective maneuvering of the platform under the test stand. The entire wheel assembly is shown in Figure 35.

Figure 35. Assembled front caster wheel.

3.3.2.2 Braking System Because this lift is not self-propelled, it does not require power brakes [2]. It is our recommendation to use appropriately sized wheel chocks to prevent the lift from moving. The sheer weight of the platform assembly makes it unlike to move at all, but to keep it secure, wheel chocks are recommended. Because the lift does not move once in place, chocks are viewed as a cost effective and simple method of preventing unwanted motion. Wheel chocks can be placed on one of the side decks for convenience.

3.3.3 Stress Considerations The base frame features several areas of concern regarding stress. The main components analyzed include the spindle and bearings, the wheel arm, the main cross beams and the I-beams under the sliding rail. Detailed derivations can be found in Appendix C.

The spindle is subjected to a large bending moment of 142,500 in-lb due to the offset of the caster assembly, which is simply calculated from the expected maximum wheel load ȋͳͷǡͲͲͲȌ ȋͻǤͷǳȌǤ 65 is one quarter of the entire weight of the platform (disregarding any items below the spindle) with an added cushion to account for uneven distribution of load. Each bearing sees almost 11,000 lb of radial force, which is less than their dynamic load ratings provided by the manufacturer. The spindle shaft itself is more than capable of operating under this load with a bending stress value of 39.8 ksi. Since the shaft material is AISI- C1144, which yields at 100 ksi, a safety factor greater than 2 is realized. Compression forces at the lower section of the spindle are negligible at 0.631 ksi.

Calculation of the wheel arm stress involved the front wheel only, since the caster capability can put more stress on the arm. Since the rear arm is essentially the same as the front, a validation of the front under the worst case scenario also validates the rear. The maximum possible bending stress in the wheel arm, which occurs when the wheel is parallel with the arm pointing outward, is only 14.2 ksi. Maximum torsion, occurring when the wheel is perpendicular to the arm, is even less at 1.72 ksi. These stresses, despite acting together on the upper and lower surfaces, allow for a safety factor greater than 2 due to the yield of A572 at 50 ksi.

The main cross beams placed under each scissor stack are expected to handle most of the horizontal force generated by the clevis connections. In fact, a single beam could theoretically be loaded alone by a scissor stack and not fail. Both of the beams have 14.1 ksi of axial stress and 8.8 ksi of bending stress. Although these stresses compound at the top of the beam, the combined value of 22.9 ksi provides a safety factor of 2.18. Furthermore, there are three additional cross beams of similar size to help increase the rigidity and safety of the platform base.

Finally, the lower I-beams upon which the bushing pads slide are loaded only in the vertical direction. This force is much lower than the horizontal component and under the most extreme loading situations a bending stress of 24.8 ksi is generated at the mid- span. Once again, this represents a safety factor over 2, without considering the additional support provided by the C-channel on the upper flange.

Since every highly stressed member has a safety factor over 2, we conclude that the platform base is safe and conforms to all relevant regulations [2].

3.3.4 Base Frame Summary To summarize, the base frame can be broken down into two major sections; the structural frame members and the steering mechanism. Both areas are highly stressed but have been proven safe. The frame is manufactured primarily from large hollow structural sections welded together, which provides excellent mounting points for the scissor pins and hydraulic rams. The force generated is distributed across the platform by large cross beams, and the lift height is kept low by separating these higher cross beams from the sliding rails. By using different sections for each function, we were able to optimize the placement to best deal with stress at the expense of adding a relatively small amount of weight. Auxiliary platforms which use the center frame as a mounting point include the intermediate access platforms and the side decks used for installing the hydraulic pump and motor.

The front wheels can caster, which provides the steering mechanism needed to maneuver the lift into place. Once the platform is hitched to the loader, it will follow whatever path the loader takes. The system relies on bearings to provide free rotation of the spindle, and the wheels can rotate 360 degrees such that the aerial lift platform can be towed forward or backed up with equal ease. Once the platform is in place, wheel chocks may be installed to prevent it from moving unwantedly.

3.4 Tires and Wheels Traditional, industrial type tires were chosen to support the ALP in the concept design phase. The tires had to be carefully selected to ensure that the large weight could be safely supported. The design considerations, available tires and the final tire selection are outlined in the following sections.

3.4.1 Design Considerations There are a wide variety of tires available commercially. In order to narrow down the tires for selection it is important to fully understand all design criteria. Some of the factors to be considered for tire selection include the operating temperature, load distribution and tire wear and maintenance. Each of these design considerations will be discussed in more detail in the subsequent sections.

3.4.2 Operating Temperature The ALP will be used mainly in the winter months where temperatures dip as low as - 40°F. The ALP will be stored outdoors year round so the tires must also be able to withstand summer temperatures up to 75°F. It is well known that rubber becomes soft as it heats up and stiff as it is cooled.

The European Tyre and Rim Technical Organisation (E.T.R.T.O) works to align the national standards for tires, rims and valves. They also work to establish common engineering dimensions, load characteristics and operational guidelines [15]. In their 2013 report the E.T.R.T.O. recommends that tires be stored below 95°F and preferably below 77°F. The report goes on to saying that cold temperatures do not permanently deteriorate the tires, however, will cause the rubber to stiffen. Prior to operating for distances at cold temperatures it is recommended that the tires be warmed [15]. Glass transition temperature (Tg) is a property that describes when a material transitions into a glassy solid. Tires are commonly manufactured of a styrene-butadiene copolymer, which exhibits a Tg of -60°F [16]. The minimum expected operating temperature is -40°F and therefore it is unlikely that full glass transition will be experienced. The ͳͲͲͲǯ operations, therefore, the stiff rubber will not be a major area of concern.

3.4.3 Load Distribution The arrangement, number of tires and type of tires used is dependent of the total weight and load distribution of the ALP. An estimate of the ALP weight that the tires

68 will have to support is given in TABLE V. The addition of personnel and equipment are considered in TABLE VI.

TABLE V WEIGHT ESTIMATE OF ALP

ALP Component Weight Estimate [lb] Platform 15,000 Scissor Mechanism 10,000 Frame 10,000 Total 35,000 Total with Safety 70,000 Factor

A search was done to determine if tires are manufactured with a pre-determined safety factor for their rated load. This is important to know to ensure that the tires are not over-engineered. The search was unsuccessful in determining a definite built in safety factor, therefore the advertised rated load is taken to be the maximum load the tire can support. The team decided to incorporate a safety factor of 2. This safety factor will take into account uneven loading, extra weight from fasteners and welds and decreased strength due to tire wear.

The ALP is designed to carry load in the form of passengers and equipment. This additional load increases the overall weight that the tires have to support. For safety reasons, it is assumed that no people will be on board the ALP while it is being towed. However, the equipment may be placed on the platform at this time. The personnel are better utilized as spotters at this time due to space restrictions under the test stand. A summary of the load capacity under varying conditions is given in TABLE VI.

TABLE VI LOAD CAPACITY SUMMARY

Loading Condition Weight [lb] Weight/Wheel [lb] Stationary ȂMax 38,500 9,625 Moving/Stationary Ȃ Min 35,000 8,750 Moving Ȃ Max 37,000 9,250

The tires will be placed in each of the four corners on the frame. A single tire is preferred over a dualie arrangement. This helps to decrease the overall width of the tire, which allows for more room for the scissor mechanism to sit in the frame. Although, it is not exactly the case, a uniformly distributed load is assumed. Each tire will therefore have to support one quarter of the total weight.

3.4.4 Tire Wear and Maintenance The tires will wear from repeated use due to the friction between the rubber and the ground. The maximum load a tire can support decreases as the tire wears. The frequency and type of maintenance varies depending on the type of tire that is chosen. A tire with the low maintenance requirements is desired.

3.4.5 Tire Types There are many different types of commercially available industrial tires. The three most popular types are pneumatic tires, solid tires and multipurpose tires. The benefits of each of these types of tires are discussed in the next three sections.

3.4.6 Pneumatic Tires Pneumatic tires are air-filled. They can have a tube that becomes inflated with air, but more popular today, they can be tubeless. Pneumatic tires are available in a variety of sizes and rated loads. A key aspect of pneumatic tires is the ability to absorb road irregularities [16]. The tires will act as a spring and damper system to absorb impacts and vibrations caused by driving over rough surfaces.

3.4.7 Solid Tires Solid tires are best for tough applications on slow moving vehicles. Since there is no air contained in a solid tire, they are ideal for situations with a high risk of cut or puncture damage. Solid tires are highly stable, have a high load rating, puncture resistant and maintenance free [17]. They are commonly found on forklifts. Solid tires are available with or without a tread. Solid tires may also be purchased in the form of a press on band.

3.4.8 Multipurpose Tires Multipurpose tires are designed to be used both on-road and off-road. They often have a large tread that provides high traction, making them ideal for winter operations [17].

3.4.9 Final Selection The tires were sourced from Continental Tire, while the wheels were sourced from Precision Products. Both companies commonly work together to provide mating products. The tires and wheels that were chosen, along with their specifications are discussed in the following two sections.

3.4.9.1 Tires Solid tires were selected as the type of tire. This was due to their high load capacity and durability. MDS AeroTest specified that the ALP should last 25 years [18]. Although the test area will be free of foreign objects so as to prevent FOD, the risk of a tire puncture still exists. Solid tires cannot be punctured allowing them to be more durable than pneumatic and multipurpose tires. The temperature in Thompson, MB will vary from - 40°F in the winter months to 75°F in the summer months. This broad temperature range will cause the air pressure in pneumatic or multipurpose tires to change. The tire pressure would have to be constantly monitored and adjusted as required to ensure even tire wear. Solid tires will require far less maintenance and should have a longer lifespan.

Continental Tire of Continental Corporation was chosen as the tire supplier. Continental ranks in the top five automotive suppliers in the world. Continental can be 71 found in 46 countries throughout the world. The Canadian location is in Mississauga, ON [19]Ǥǯ ǤǤǤǤǤ ǡ Canadian sales contact, was contacted via telephone and e-mail for recommendations [20]. For the loading condition, operating environment and durability requirements, Tim suggested press on bands. Press-on bands refer to rubber seated on a steel base, or similar, that would then be pressed onto a wheel.

The selected tire is the 28x16x22 MH20. These tires are press on bands that have a steel base as opposed to steel wire reinforcement. Dimensions and loading capacities of the 28x16x22 MH20 can be found in TABLE VII with reference to Figure 36.

TABLE VII DIMENSIONS AND RATED LOADS OF 28X16X22 MH20 TIRES

Dimension (From Figure Measurement [in] 36) A 28 B 16 C 22 Speed [miles/hour] Rated Load [lb] Safety Factor Stationary 20,460 2.1 4 20,460 2.2 6 16,800 1.8 10 15,345 1.5

Figure 36: Continental Tire Press on Band Dimension Guide (used with permission) [17].

The safety factors in TABLE VII were calculated by dividing the rated load by the maximum weights for each speed as specified in TABLE VI.

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The rated load of the tire is a function of the speed of the vehicle. As the speed is increased, the rated load is decreased. For safety reasons the ALP will not travel at speeds greater than 6 mph when personnel are on the lift. From TABLE VII it can be seen that there is a safety factor of 1.8 or greater for all regular operations applications. The initial goal was to have a safety factor of 2, which is achieved when stationary and at low speeds. The ALP may experience speeds of 10 mph or greater when being towed behind the CAT 966H. These will be times of transit and the ALP will be at its minimum weight. The tires still experience a safety factor of 1.5 at this time as seen in TABLE VII.

All rated loads are based off of intermittent use. The tires may travel a maximum ͸ͷ͸Ͳǯ at one time. It is recommended that the tires sit for at least three hours between maximum drive cycles [21]. The ALP will primarily be maneuvering around the test site and sho͸ͷ͸Ͳǯ at one time.

A smooth tire versus a treaded tire was selected. Treads would provide a higher traction, which although would be more ideal, the team did not think was necessary for this application. The ALP will primarily be operating in the winter where there is a risk of snow and ice. These times would require high traction tires, however the test site is kept clear [18]. Lastly, the drive power of the ALP is achieved from the towing capabilities of the CAT 966. Since the ALP will be towed and not driven under its own power, the traction on the tires is not at as critical. The slightly decreased traction should allow for the wheels to more freely turn and follow the CAT 966 while being towed. The smooth tires should not experience too much sipping or skidding.

More detailed product specifications can be found in the datasheet for the 28x16x22 MH20 tires. This datasheet was made available by Tim Hubbert from Continental Tire and can be found in Appendix E.

3.4.9.2 Wheels Continental Tires does not sell the wheels that the tires would be pressed onto. Precision Products, located in Guelph, Ontario is the supplier they work with. Tim referred the team to David Hartig at Precision Products [20].

ͲǤ͹ͷǳͳǤͲͲǳǤ provided the team with two preliminary drawings that can be found in Appendix E. The first wheel is a heavy-duty wheel that supports high loads. It is one that Precision Products commonly manufactures. The second wheel was for a lighter-weight application. The second wheel had an ideal layout as it easily allowed for a shaft to be inserted through the center of the wheel, as the team had envisioned. After further consultation with David, he produced another drawing. This last design combined the high load capability of the heavy-duty wheel with the layout of the lighter-weight wheel [22].

Precision Products does not supply bearings, but David said that they could source one from a supplier. The Timken 22210 or Timken 22211 are recommended for this application. The customer will supply the shaft, as requested by Precision Products.

There are not many dimensions given on the drawing. This is to allow for further customization by the customer. Precision products will run FEA on the product prior to customer sale for $500. It is recommended that non-destructive testing be conducted on the wheels to ensure they are sufficient [22].

All of the drawings provided by Precision Products can be found in Appendix E.

3.4.10 Cost Analysis Shipping and freight costs are additional to the quotes as given in TABLE VIII. The detailed quotation for the tires can be found in Appendix E. No formal quotation was given for the wheels; however a quote was given via email.

TABLE VIII COST ANALYSIS FOR TIRE AND WHEEL ASSEMBLY

Product Price/unit Quantity Total Cost Supplier [CAD] [CAD] 28x16x22 MH20 $619.49 4 $2,477.96 Continental Tire Tire Custom Wheel $1832.00 4 $7,328.00 Precision Products Tire and Wheel Assembly $9,805.96

4 Bill of Materials and Cost Analysis

The following material tables summarize the majority of parts required to fabricate the scissor lift, with cost estimates provided when available. The only items not accounted for are the hitch and intermediate access platforms. Due to time constraints, they were not incorporated into the main summary. These components use similar shapes and sizes of materials compared to the remainder of the lift. Furthermore, specific information on the hydraulic and electrical drive system, such as manufacturer and model number, are omitted due to these items not being completely specified at the time of this report. General details, such as size, are provided where necessary.

Material tables are broken down by major component, and further subdivided into minor components. Major sections include platform, base, scissors, and actuation. Each table corresponds to a major component, while the subsequent minor sections can be found on the left hand side of the associated table. Parts are shown in the length required for assembly, unless otherwise noted. Certain components require significant machining operations to achieve the final shape; these items will be listed in stock form with a relevant explanatory superscript. The majority of the material is stock steel shapes which only require cutting.

ǤǮ ǯ ǤǮǯ rial is unique in its assembly and cannot be combined with other components. A total length or area for that component, dependent on how many exist, is also provided. This is simply the dimension of the ǡ Ǯǯ ǡmultiplied by the quantity. Bold values represent bulk steel and aluminum orders.

Steel and aluminum quotes are pending. A quote request has been sent out to relevant suppliers, and the information will be forwarded to the client as soon as it becomes available. The material list sent represents all stock steel and aluminum shapes 76 required to assemble the aerial lift platform, excluding the hitch assembly and the intermediate access platforms. Any cost cells which are empty are pending quotes.

TABLE IX SUMMARY OF PLATFORM MATERIALS

Component Material Supplier Quantity Weight (lb) Recommended Order Estimated Cost (CAD)

Outer beams - perpendicular to ͺǳ͸ǳΥǳ Ȃ A572 steel Brunswick Steel / 2 902 cross spans 241-1/2" length Russel Metals HSS Outer beams - parallel to cross ͺǳ͸ǳΥǳ Ȃ A572 steel Brunswick Steel / A572 Steel 2 900 spans 240-1/2" length Russel Metals 8" x 6" x 1/4" 1878" length ͺǳ͸ǳΥǳ Ȃ A572 steel Brunswick Steel / Cross span beams 4 1708 228-1/2" length Russel Metals FRAME 1 Perimeter for grating - ͵ǳʹ- /2ǳΥǳȂ A572 Brunswick Steel / Angle Iron 1 2 78 perpendicular to cross spans steel 229- /2" length Russel Metals A572 Steel

3" x 2-1/2" x 1/4" Perimeter for grating - ͵ǳʹ-1/2ǳΥǳȂ A572 Brunswick Steel / 2 78 916" length perpendicular to cross spans steel 228-1/2" length Russel Metals

19-Ͷʹǳ3/16ǳswaged aluminum serrated edge grating panel (bare Russel Metals / Grating1 10 1800 as shown, 400 sq. ft. finish) 24" width / 240" Grating Pacific

SURFACE length

6" x 6" x 1/2" HSS - A572 steel Brunswick Steel /

Corner posts 4 540 as shown, 184" length 46" length Russel Metals Brunswick Steel / Corner post support plates 10" x 4" x 3/8" plate - A572 steel 8 35 as shown, 320 sq. in. Russel Metals Brunswick Steel / Corner post caps 8" x 8" x 1/16" plate - A572 steel 4 5 as shown, 256 sq. in Russel Metals

NPS 1-1/2 SCH 40 - A572 steel Brunswick Steel / Side guardrails - top horizontal 2 104 229-1/2" length Russel Metals Standard Pipe AND HARNESS ANCHORS Side guardrails - middle NPS 1-1/2 SCH 40 - A572 steel Brunswick Steel / A572 Steel 8 104 horizontal 57-1/2" length Russel Metals NPS 1-1/2 SCH 40 1182" length NPS 1-1/2 SCH 40 - A572 steel Brunswick Steel / Side guardrails - vertical 6 60 43-3/4" length Russel Metals

4" x 1/8" plate - A572 steel Brunswick Steel / Side toe-boards 2 65 as shown, 459" length 229-1/2" length Russel Metals

GUARDRAILS, GATES 3" x 2-1/2" x 1/4" unequal leg - Brunswick Steel / Sleeve to frame connectors 20 21 as shown, 160" length 6061 T6 8" length Russel Metals

NPS 2 SCH 40 - 6061 T6 Brunswick Steel / Sleeves 20 17 as shown, 160" length 8" length Russel Metals Miller Fall Horizontal lifeline Miller Xenon HLL kit - 30' span 6 N/A individual order Protection Miller Fall Lifeline shuttles Miller Xenon shuttle 6 N/A individual order Protection 5/8-11 x 1" grade 8 bolt w/ grade 8 Sleeve to frame hardware McMaster-Carr 60 2 6 packs of 10 $52.80 plate washer

Front / back guardrails and NPS 1-1/2 SCH 40 - 6061 T6 Brunswick Steel / 1 93 as shown, 1184" length gates2 1184" length Russel Metals

Front / back guardrails and NPS 1-1/2 SCH 80 - 6061 T6 Brunswick Steel / 1 151 as shown, 1448" length gates2 1448" length Russel Metals

4" x 1/8" plate - 6061 T6 Brunswick Steel / Front / back toe boards 4 10 as shown, 204" length 51" length Russel Metals

Spring latches 3" body with locking steel bolt Princess Auto 20 35 Individual order $259.80

1/4"-20 x 5/8" pan head phillips Spring latch hardware McMaster-Carr 80 N/A packs of 100 machine screw w/ matching nut Brunswick Steel / Backing plates for spring latches 3" x 1-3/4" x 1/4" plate - 6061 T6 20 3 as shown, 105 sq. in. Russel Metals

5/8" diameter, 2-1/4" length w/ Clevis pins McMaster-Carr 20 N/A 10 packs of 2 $53.80 hairpin cotter

TOTALS 6711 $366.40

1 Grating is mesaured without considering the various cut-out pieces in standard lengths 2 Due to variety of shapes, this item is measured in total required length without sectional specification

TABLE X SUMMARY OF SCISSOR ASSEMBLY MATERIALS

Component Material Supplier Quantity Weight (lb) Recommended Order Estimated Cost (CAD)

14" x 10" x 0.313" HSS - CSA 70T steel Scissor arms Russel Metals 24 14853 as shown, 3648" length 152" length

4" x 4" x 3/16" HSS - A572 steel Russel Metals / Short cross member 20 378 HSS 24" length Brunswick Steel A572 Steel 4" x 4" x 3/16" 4" x 4" x 3/16" HSS - A572 steel Russel Metals / Long cross member 2 73 46-1/2" length Brunswick Steel 1046"

3-1/2" mechanical tubing, 0.156" wall - A572 Russel Metals / Outer bushing bores 48 251 as shown, 540" length steel 11-1/4" length Brunswick Steel SKF 3" I.D. / 3.1875" O.D. split PTFE composite Outer bushings SKF Bearings 144 137 individual order 3.75" length (PCZ 4860 B)

3" diameter solid round bar - CSA 70T steel Round bar Long outer pivot pins 1 Russel Metals 16 773 23- /2" length CSA 70T Steel 3" diameter solid round bar - CSA 70T steel 3" diameter Short outer pivot pins Russel Metals 16 469 604" length 14-1/4" length 3" shaft - standard black finish - steel McMaster-Carr (or Outer pivot retaining rings 64 N/A individual order $113.00 (97633A480) similar) SCISSOR STRUCTURE

4-5/8" mechanical tubing, 1/4" wall - A572 steel Russel Metals / Middle bushing bores 24 263 as shown, 270" length 11-1/4" length Brunswick Steel SKF 4" I.D. / 4.1875" O.D. split PTFE composite Middle bushings SKF Bearings 72 91 individual order 3.75" length (PCZ 6460 B) 4" diameter solid round bar - CSA 70T steel Middle pivot pins Russel Metals 12 1031 as shown, 282" length 23-1/2" length 4" shaft - standard black finish - steel McMaster-Carr (or Middle pivot retaining rings 24 N/A individual order $115.00 (97633A500) similar) Russel Metals / Bushing bore plates 4-1/2" x 4" x 3/8" plate - A572 steel 96 189 Brunswick Steel Plate A572 Steel 3/8" thick

Upper clevis mounts, vertical Russel Metals / 12-1/2" x 8" x 3/8" plate - A572 steel 16 174.6 plates1 Brunswick Steel 4256 sq. in. MOUNTING HARDWARE

Upper clevis mounts, horizontal Russel Metals / 14-1/2" x 8" x 3/8" plate - A572 steel 8 101 plates1 Brunswick Steel Lower fixed clevis mounts, vertical Russel Metals / 14" x 11-1/2" x 1" plate - A572 steel 4 187.5 plates1 Brunswick Steel Lower fixed clevis mounts, Russel Metals / 14" x 12-1/4" x 1" plate - A572 steel 4 199.5 Plate horizontal plates1 Brunswick Steel A572 Steel Lower translating clevis mounts, Russel Metals / 1" thick 12-1/2" x 8" x 1" plate - A572 steel 4 116 vertical plates Brunswick Steel 2194 sq. in. Lower translating clevis mounts, Russel Metals / 14-1/2" x 8" x 1" plate - A572 steel 4 135 horizontal plate Brunswick Steel

PTFE steel-backed slide bushing - 7-3/4" x 3- Bearing pad Fluorotec 8 N/A order as sheet, 233 sq. in 3/4" x 0.2165" (total thickness) C4 x 5.4 C-channel - A572 steel Russel Metals / Bearing pad mounts 8 29 as shown, 64" length 8" length Brunswick Steel C6 x 8.2 C-channel - A572 steel Russel Metals / Upper bearing pad guide rail 4 270 C-channel 99" length Brunswick Steel A572 Steel C6 x 8.2 C-channel - A572 steel Russel Metals / C6 x 8.2 Lower bearing pad guide 4 541 198" length Brunswick Steel 1188" length

6" x 2" x 3/16" HSS - A572 steel Russel Metals / Upper slide support 8 258 as shown, 328" length 41" length Brunswick Steel

4-1/2" x 1/16" plate - hardened steel Russel Metals / Plate Upper slide plate 4 20 60" length Brunswick Steel Hardened Steel 4-1/2" wide x 1/16" 4-1/2" x 1/16" plate - hardened steel Russel Metals / Lower bearing pad slide plate 4 18 thick 54" length Brunswick Steel 456" length

8" x 6" x 1/4" HSS - A572 steel Russel Metals / Fixed upper mount support 4 308 as shown, 165" length 41-1/4" length Brunswick Steel W8 x 15 I-beam - A572 steel Russel Metals / Lower slide support 4 990 as shown, 792" length 198" length Brunswick Steel

TOTALS 21856 $113.00

1 Given as square sections, cutting required

TABLE XI SUMMARY OF BASE FRAME MATERIALS

Combined Order for Component Material Supplier Quantity Weight (lb) Estimate Estimated Cost (CAD)

14" x 10" x 5/16" HSS - A572 steel Russel Metals / Front / rear main lateral beams 2 1353 166" length Brunswick Steel

14" x 10" x 5/16" HSS - A572 steel Russel Metals / Vertical wheel arm section - front 2 310 HSS 38" length Brunswick Steel A572 Steel

14" x 10" x 5/16" 14" x 10" x 5/16" HSS - A572 steel Russel Metals / Vertical wheel arm section - rear 2 283 34-3/4" length Brunswick Steel 666"

14" x 10" x 5/16" HSS - A572 steel Russel Metals / Horizontal wheel arm section 4 766 47" length Brunswick Steel

Outer cross beams / low center cross 14" x 10" x 3/8" HSS - A572 steel Russel Metals / 3 2876 HSS beam 198" length Brunswick Steel A572 Steel

14" x 10" x 3/8" 14" x 10" x 3/8" HSS - A572 steel Russel Metals / Thick inner cross beams 2 2188 226" length Brunswick Steel 1046" FRAME

10" x 10" x 1/4" HSS - A572 Russel Metals / Wheel arm reinforcements 4 102 as shown, 40" length 9-3/8" length Brunswick Steel BASE

8" x 3" x 3/8" HSS - A572 steel Russel Metals / Thin center cross beam 2 939 as shown, 452" length 226" length Brunswick Steel Russel Metals / Main beam-to-wheel arm gusset plates1 8" x 6" x 1/2" plate - A572 steel 6 42 as shown, 228 sq. in. Brunswick Steel Russel Metals / Cross beam gusset plates1 8" x 6" x 3/8" plate - A572 steel 4 21 Brunswick Steel Plate Russel Metals / A572 Steel Thick inner cross beam end plates 17-5/8" x 9-1/4" x 3/8" plate - A572 steel 4 71 Brunswick Steel 3/8" thick 525 sq. in. Russel Metals / Thin center cross beam end plates 19-5/8" x 2-1/4" x 3/8" plate - A572 steel 2 10 Brunswick Steel

3 2- /16" diameter shaft - A572 steel Russel Metals / Wheel axle2 4 94 as shown, 88" length 22" length Brunswick Steel 2" shaft - standard black finish - steel McMaster -Carr (or SYSTEM Axle retaining rings 8 N/A packs of 10 STEERING (97633A420) similar)

Front caster wheel mount, vertical Russel Metals / 23" x 16-1/2" x 1" plate - A572 2 194 plates3 Brunswick Steel Front caster wheel mount, horizontal Russel Metals / 21-1/2" x 18" x 1" plate - A572 2 225 plate Brunswick Steel Plate Russel Metals / A572 Steel Rear wheel mount, vertical plates3 20" x 16-1/2" x 1" plate - A572 2 192 Brunswick Steel 1" thick 2823 sq. in. Russel Metals / Rear wheel mount, horizontal plates 21-1/2" x 10" x 1" plate - A572 2 125 Brunswick Steel Russel Metals / Spindle to wheel mount plate 10" x 10" x 1" plate - A572 steel 2 58 Brunswick Steel 5/8-11 x 2.75" grade 8 bolt McMaster-Carr (or Thru-axle pinch bolts 8 N/A packs of 5 (91257A805) similar)

5-1/2" diameter shaft - AISI-C1144 steel Russel Metals / Spindle2 2 226 as shown, 33" length 16-5/16" length Brunswick Steel 3/4-10 x 3" grade 8 bolt w/ nut and two McMaster-Carr (or Spindle to wheel mount hardware 8 N/A individual order washers (92620A847) similar) 9" diameter round bar - A572 steel Russel Metals / Spindle housing2 2 592 as shown, 32" length 16" length Brunswick Steel

9" I.D. / 10-1/2" O.D. mechanical tube - Russel Metals / Spindle housing-to-arm support ring2 2 13 as shown, 2" length A572 steel - 1" length Brunswick Steel

Spindle seal Timken N 24600-3189 lip seal Timken Bearings 2 N/A individual order

Lower tapered roller bearing Timken TS 64450 inner / 64700 outer Timken Bearings 2 15 individual order

Low temperature synthetic grease McMaster-Carr (or Grease 1 N/A individual order (bulk) similar) Straight, 1/4"-28 taper thread, 35/64" McMaster-Carr (or Grease nipple 2 N/A individual order $6.10 height (1095K41) similar)

Upper ball bearing Timken 45BIC206 Timken Bearings 2 6 individual order

4-1/2" shaft - standard black finish - McMaster-Carr (or Retaining ring 2 N/A individual order steel (97633A519) similar) Russel Metals / Cover plate2 9" x 9" x 1/8" plate - A572 steel 2 6 as shown, 162 sq. in. Brunswick Steel

5-3/4" I.D. / 6" O.D. mechanical tube - Russel Metals / Bearing retainer2 1 7 as shown, 1" length A572 steel - 13/16" length Brunswick Steel McMaster-Carr (or Flanged cover bolts 3/8"-16 x 1" grade 5 zinc plated flanged 12 N/A pack of 25 $8.31 similar) McMaster-Carr (or Gaskets Buna-N Flange Gasket, 1/16" thick 6 N/A individual order $41.16 similar)

6" x 2" x 3/16" HSS - A572 steel Russel Metals / Storage rack - longitudinal beams 2 230 146" length Brunswick Steel HSS 6" x 2" x 3/16" HSS - A572 steel Russel Metals / A572 Steel Storage rack - middle beams 4 46" length Brunswick Steel 145 6" x 2" x 3/16" 668" length 6" x 2" x 3/16" HSS - A572 steel Russel Metals / Storage rack - outer beams 4 48" length Brunswick Steel 151

SIDE PLATFORMS Russel Metals / Storage rack - gusset plates 10" x 8" x 1/8" plate - A572 steel 2 6 as shown, 160 sq. in. Brunswick Steel

Precision Products Rims4 Per DWG X13-11-19 Rev 0 4 1872 Individual order $7,328.00 Ltd. 28 x 16 x 22 MH20 (Smooth) Continental Tire

WHEELS Tires 4 N/A Individual order $2,477.96 Article #1375005 Canada

TOTALS 11246 $9,861.53

1 Gusset plates are triangular, therefore a single rectangle represents two gusset plates 2 Nominal stock sizes shown, significant machining is required 3 Indicated area allows for both sides of a single caster housing, including hub retainer 4 Associated hardware such as bearings, retaining rings and seals integral to the rim are included

TABLE XII SUMMARY OF ACTUATION COMPONENTS1

Component Material Supplier Quantity Weight (lb) Recommended Order Estimated Cost (CAD)

Electric motor 575 V 3 phase AC / 65+ HP Eaton Hydraulics 1 Combined as unit Hydraulic pump 35 gpm @ 2500 psi Eaton Hydraulics 1

Hydraulic lines 3000 PSI capability 1

Fluid reservoir 40 gallon capacity 1

Hydraulic fluid Arctic 15 / 40 gallons 1

3000 PSI / 8 inch piston / 42 inch Hydraulic rams Eaton Hydraulics 4 stroke / double acting / clevis ends

ACTUATION AND RELATED

1 Preliminary information only TOTALS 0 $0.00

5 Recommendations

Some ǯ restrictions. Other requirements that require electrical engineering were not addressed because of the lack of electrical engineering students in our team. In this section, these requirements are mentioned and discussed. Recommendations for the integration of ǯs design are discussed as well.

5.1 Adjustable Stairs The client requested a design of a staircase that will move up and down with the platform. The staircase is to provide a quick and safe access to the platform and the ground, while the platform is in a raised position, at any height. This staircase must be located at the front of the platform (the side where the hitch connection is located). That is due to the test stand being in the way on the sides, and the tunnel being in the way at the back. One of the drop-down gates at the front could be lowered and access to the staircase would be then provided.

5.2 Top of Engine Access Ladder The client requested a design of a ladder that is to be flexible and be able to accommodate different engine sizes. The ladder is to provide an access from the platform to the top of a jet engine mounted on the test stand. The ladder could be mounted to the grating, pǯs frame.

5.3 Electrical Connections The client requested for the ALP to have 110V electrical plugs for connecting power tools while working at elevated heights. The plugs at the platform level would eliminate the need of long extension cords that must reach from the ground connection all the ǯ ǤAn electrical engineer or a certified electrician can easily add this feature. The plugs can be mounted anywhere on the platform where it is safe to do so. The extension cord required for connecting the platform plugs to the plugs located 86 at the ground level can run below the platform, down the scissor members, and through the base frame. The extension cord could be stored at the base frame level with easy access and retraction.

5.4 Control System The ALP must have a system to control the ascent and descent of the platform. The controls must be located at both the ground level and the platform level. The control system is to have variable speed control capability to lift and lower the platform at multiple rates. The control module at the platform level could be stored at any corner of the platform. At the ground level, the control module could be mounted somewhere on the base frame. An electrical engineer should design this feature.

5.5 Ergonomic Mat It is our recommendation that a semi-permanent ergonomic mat be installed on the upper platform. For this application, the mat should be installed on a roller assembly, similar to that of a common pool cover, with this assembly fixed to one side of the platform (sides containing the horizontal lifelines). This would allow the mat to be rolled out when needed, but retracted into a relatively compact form-factor when stored. The mat should be made from a material similar to that of the anti-fatigue mats commonly found at cashier stands in most department stores.

6 Summary

MDS AeroTest requires a large ALP to access the engine when mounted in the engine test stand at GLACIER. Scissor lifts are currently rented to achieve this task. These scissor lifts only provide enough surface area to accommodate one to two workers and a small amount of equipment. This results in the need for numerous trips up and down to obtain the required equipment thereby increasing the amount of time required to certify an engine. MDS AeroTest would like the ability to support four to six workers and 2000 lb of equipment on the ALP. As icing certification testing can only be performed in ideal conditions, the testing time is limited. Therefore, increasing the time required to certify an engine decreases the overall profit for MDS AeroTest.

Our team designed a large scissor lift, primarily made of 6061 T6 aluminum, A572 steel and 70T steel. These materials were chosen for their ability to retain high strength properties in low temperature environments and are available from Russell Metals or Brunswick Steel in Winnipeg, MB. Design features of the main components of the ALP will be discussed below.

The first major component of the ALP is the platform. Rectangular HSS and angle iron were combined to create the frame of the platform. A swaged aluminum gratiʹǳ depth was selected for the platform surface. Snow, ice and oil will not easily build on the surface due to the open nature of grating. This will in turn reduce the slipping hazard, which is of utmost importance at an elevated height.

When in ǡͻǯͷǳǡ ʹͻǯʹǳǤ incorporated to the fore and aft sides of the lift. This makes the large access height more ergonomic and feasible. Additionally, the guardrails on the fore and aft sides of the lift have the ability to be dropped completely below the surface of the platform. When in a dropped configuration these guardrails serve as a means of entering or exiting the platform. The adjustable height guardrails are not limited to a fully up or a 88 fully down position. A spring latch has been incorporated into the drop down mechanism to allow for a configuration over various heights. This allows for the railings to be shaped to the contour of the engine, which will prove to be beneficial when the ALP is being operated in close proximity to the engine.

The side guardrails remain fixed in position to allow a safe and convenient way of mounting horizontal lifelines. Three horizontal lifelines can be installed on each side, for a total of six or one for each worker. This allows for minimal entanglement between lifelines as workers move about the lift.

The next major component of the ALP is the lifting mechanism. A scissor mechanism was selected for its ability to become compact when collapsed, but still reaching large heights when extended. Traditional scissor mechanisms were not feasible due to the large weight of the ALP and in turn the large stressed that will be subject to the scissors. Two sets of paired scissor stacks extending toward opposite sides of the platform were selected. This allows for the scissors to be supporting opposite corners of the platform when extended, thereby reducing bending moments and increasing stability. Each ͹ͲͳͷʹǳǤ composite bushings are used at all pins to ensure maintenance-free pivoting action.

A hydraulic system driven by an electric motor (running off of grid power) was selected to power the lifting mechanism. A hydraulic pressure of 2387 psi per actuator is required to lift the load of the platform, scissor mechanism and load. Accounting for inefficiencies in the hydraulic pump and electric motor, a 64 HP electric motor is required.

Grid power was selected for its availability at the test site and reliability at low temperatures. Grid power of 575 V and 30 A is currently available at the GLACIER facility, however, the client is able to provide higher amperage. With higher amperage,

ʹͻǯʹǳ minute, which is desired by the client.

Due to time restrictions, a specific hydraulic pump and electric motor were not selected. However, companies such as Hypower Systems Inc, Eaton Hydraulics and Parker were found to have pumps and motors that would be acceptable for this application. An integrated motor pump unit, which combines the hydraulic pump and electric motor can also be purchased to alleviate the need to purchase and interface the two separately.

The last major component is the base frame. The base frame was designed to sit as low as possible in an effort to reduce the overall height of the ALP when collapsed. Varying wall thickness of Aͷ͹ʹǡͳͶǳͳͲǳ structural backbone of the base frame. The scissor mechanisms slide in a C-channel guide that sits upon I-beams, all of which are welded to the box beam frame.

For simplicity and in an effort to reduce cost and maintenance, the ALP will be towed behind the CAT 966H currently owned by MDS AeroTest. The hitch assembly of the ALP was modeled closely off the engine transporter in use at GLACIER. Due to time restrictions, the hitch assembly is in the preliminary design phase and FEA is recommended to ensure that applied loads are sufficiently supported.

Connected to the base frame is the steering mechanism. For simplicity and reliability at varying turning radii, a caster wheel assembly was selected for the front two wheels, while the rear two wheels do not have the ability to rotate. This set-up will easily allow the ALP to follow the turn of the CAT966H that is towing it. Solid rubber, smooth tread, ʹͺǳ Ǥ ensures durability, requires less maintenance and has less deflection than standard pneumatic tires. Precision Products created a customized wheel for our high load

90 situation. The wheel has a double center gusset and allows for shaft connection through the middle.

That concludes the main design features of the ALP. All primary design objectives of the client were met, however, due to time restrictions, not all secondary objectives could be designed and are recommended for further review as follows.

A means of transitioning from the platform (at elevated height) to the ground, without lowering the lift was desired. We envision an adjustable height staircase that would be connected to the hitch end of the ALP.

Additionally, a ladder used to access the top of an engine mounted in the test stand to accommodate sensor connections was desired. We believe this ladder could be mounted to the grating above one of the box beams that make up the platform frame.

Two desired design features relate to electrics. As the design team is comprised of mechanical engineering students, electrical design is outside of our expertise. It is recommended an electrical engineer be used to design electrical connections and control systems. The electrical connections are to be placed on top of the platform to allow for tools to be used at elevated heights without the requirement to run long extension cords. In the meantime, extension cords may be used. A control system is required to control the ascent and decent of the platform and must be located both at ground level and on top of the platform.

Lastly, the team recommends that a temporary ergonomic mat be installed on the platform to reduce worker fatigue and catch falling tools. The mat could be rolled up and placed on the side of the platform for storage.

7 Works Cited

[1] R. Howitt and D. Pereira, GLACIER Site Visit, Thompson, Manitoba, 2013.

[2] Canadian Standards Association, B354.1-04 Portable elevating work platforms, Canada, 2004. [3] F. P. Beer, E. R. Johnston, Jr., J. T. DeWolf and D. F. Mazurek, Mechanics of Materials, 5th ed., New York, NY: McGraw-Hill, 2009.

[4] Brunswick Steel, "Reference Catalogue". [5] ASM International, Metals Handbook Volume 1, 10th ed., 1990. [6] Russel Metals, "Stock List and Reference Book," Canada.

[7] Grating Pacific, "Bar Grating Catalogue," 2004. [Online]. Available: http://www.gratingpacific.com/metal_bar_gratings/metal_bar_gratings.html. [Accessed 30 October 2013].

[8] "Zenon Permament Cable Horizontal Lifeline System," Miller Fall Protection, [Online]. Available: https://www.millerfallprotection.com/fall-protection- products/horizontal-lifeline-systems/soll-xenon-permanent-cable-horizontal- lifeline-system. [Accessed 23 October 2013].

[9] Vertikal.net, "Bo-Rent takes big Holland Lift," Holland Lift, 16 September 2005. [Online]. Available: http://www.vertikal.net/en/news/story/1876/. [Accessed 27 November 2013].

[10] "Engineered Plastic Bearings," Bunting Bearings, [Online]. Available: http://catalog.buntingbearings.com/category/engineered-plastic- bearings?&plpver=10. [Accessed 27 November 2013].

[11] SKF Bushings, "Bushing Catalog," UP Print, Estonia, 2003.

[12] H. M. Spackman, "Mathematical Analysis of Scissor Lifts," U.S. Marine Corps, Quantico, VA, 1989.

[13] Eaton Hydraulics, "Integrated Motor Pumps," [Online]. Available: http://www.eaton.com/Eaton/ProductsServices/Hydraulics/index.htm?wtredire ct=www.eaton.com/hydraulics. [Accessed 29 November 2013].

[14] Timken Bearings, "Roller Bearing Catalog," [Online]. Available: http://www.timken.com/en-us/Pages/Home.aspx. [Accessed 10 November 2013].

[15] European Tyre and Rim Technical Organisation, 7 June 2013. [Online]. Available: http://www.etrto.org/page.asp?id=1594&langue=EN. [Accessed 9 November 2013].

[16] ǡ ʹͲͲ͸ǤȏȐǤǣ ǤǤȀ Ȁ ȀȀ ̴ -ͺͳͲ-ͷ͸ͳǤ. [Accessed 9 November 2013].

[17] Continental Tires, April 2008. [Online]. Available: http://www.continental- specialty- tires.com/www/download/industry_de_en/general/downloads_media/3_technic al_information/downloads/td_full_version_english_uv.pdf. [Accessed 7 November 2013]. [18] R. Howitt, Conference Call - Project Definition, 2013.

[19] Continental Corporation, 2013. [Online]. Available: http://www.continental- corporation.com/www/portal_com_en/. [Accessed 13 November 2013].

[20] T. Hubbert, Re: Continental Tire - Price Quote, Mississauga, Ontario, 2013. [21] Continental Tire, "Datasheet - Elastic Tires for Industrial Vehicles," Mississauga, 2011.

[22] D. Hartig, Re: Continental Tire - Price Quote, Guelph, Ontario, 2013.

Appendix A: Detailed Drawings

List of Drawings Drawing 01 ± ALP Assembly««««««««««««««««««««««..««« Drawing 02 ± Base Frame Assembly««««««««««««««««««««..«« Drawing 03 ± Scissor Assembly«««««««««««««««««««««««..« Drawing 04 ± Platform Assembly«««««««««««««««««««««««..« Drawing 05 ± Secondary Platform Assembly««««««««««««««««««..« Drawing 06 ± Caster Assembly«««««««««««««««««««««««« Drawing 07 ± Fixed Wheel Assembly«««««««««««««««««««««« Drawing 08 ± Hitch Assembly««««««««««««««««««««««««« Drawing 09 ± Spindle Assembly«««««««««««««««««««««««« Drawing 10 ± Scissor Beam Assembly««««««««««««««««««««« Drawing 11 ± Various Pin Assemblies««««««««««««««««««««« Drawing 12 ± Base Frame Weldment«««««««««««««««««««««« Drawing 13 ± Scissor Beam Weldment««««««««««««««««««««« Drawing 14 ± Platform Frame Weldment««««««««««««««««««««« Drawing 15 ± Secondary Platform Weldment«««««««««««««««««« Drawing 16 ± Caster Bracket Weldment««««««««««««««««««««« Drawing 17 ± Spindle Housing«««««««««««««««««««««««« Drawing 18 ± Spindle Weldment«««««««««««««««««««««««« Drawing 19 ± Drop Down Gate Weldments««««««««««««««««««« Drawing 20 ± Sleeves Weldments««««««««««««««««««««««« Drawing 21 ± Hitch Weldment«««««««««««««««««««««««« Drawing 22 ± Various Parts and Weldments««««««««««««««««««« Drawing 23 ± Various Shafts and Pins«««««««««««««««««««««« Drawing 24 ± Grating Cuts««««««««««««««««««««««««««18

100

101

102

103

104

105

106

107

108

109

110

111

112

113

114

115

116

117

118

Appendix B: Detailed Platform Analysis

Table of Contents

1 Platform Frame Stress Analysis ...... 120

2 Grating Validation ...... 122

3 Guardrail Stress Analysis ...... 122

4 Corner Post Anchorage Stress ...... 124

5 Works Cited ...... 127

List of Figures

Figure 1. Worst case loading scenario shear and moment diagrams...... 120

Figure 2. Aluminum grating load table from Grating Pacific catalogue...... 122

Figure 3. Bending of guardrails...... 123

Figure 4. Anchor loading on corner posts...... 125

119

1 Platform Frame Stress Analysis This section provides the stress analysis used to validate the beam selection for the main platform frame. A shear ± moment diagram is included to illustrate the loading stress induced by our worst case scenario. Formulas are then utilized to arrive at the actual stress numbers presented in the body of the report.

Figure 1. Worst case loading scenario shear and moment diagrams.

120

ݕܯ ȁߪȁ ൌ Equation 1.1 ܫ

)RURXU´[´[´ER[EHDPVHFWLRQ

ܫ ൌ ͸ʹǤ͸ͶͲ͸݅݊ସ

ݕ ൌ Ͷ݅݊

Based on our maximum moment illustrated in Figure 1:

݅݊ ሺ ሻ ሺ ሻ ൬ͳʹ ݂ݐ൰ Ͷ݅݊ ݏܾ݈ כ ͳͺǡ͸ʹͷ݂ݐ ȁߪȁ ൌ ൌ ͳͶʹ͹ʹ݌ݏ݅ ൌ ͳͶǤ͵݇ݏ݅ ͸ʹǤ͸ͶͲ͸݅݊ସ

Because the beam is symmetrical with respect to the neutral axis, this value represents the maximum tension and compression on the lower and upper-most faces of the beam, respectively.

ܸܳ ߬ ൌ Equation 1.2l ݐܫ ௔௩௚

The maximum average shear across the wall will be found along the neutral axis of the beam, since this is where Q will be highest and t is consistent throughout the section. I remains the same as above.

ܳ ൌ ͻǤ͵ʹͺ݅݊ଷ

ݐ ൌ ͲǤʹͷ݅݊

Based on our maximum shear found in Figure 1:

ሺͳͻ͹ͷ݈ܾݏሻሺͻǤ͵ʹͺ݅݊ଷሻ ߬ ൌ ൌ ͳͳ͹͸݌ݏ݅ ൌ ͳǤͳͺ݇ݏ݅ ௔௩௚ ሺ͸ʹǤ͸ͶͲ͸݅݊ସሻሺͲǤʹͷ݅݊ሻ

121

2 Grating Validation For ease of reference, the grating load table utilized during the selection of the platform surface is provided below [1]. Comparisons have shown that these values are fairly consistent across various manufacturers.

Figure 2. Aluminum grating load table from Grating Pacific catalogue [1].

3 Guardrail Stress Analysis Both the side rails and front/back rails must be capable of withstanding a 300 lb static force placed anywhere without yielding. Bending is the only concern here; shear and axial loading are miniscule for a 300 lb load. Either mid-span loading, where the beam is considered simply

122 supported on both ends, or cantilever loading of the vertical posts are possibilities. Equation 1.1 is used throughout. Figure 3 below illustrates the loading scenarios.

Figure 3. Bending of guardrails.

Mid-span bending stress of the side rail is computed as follows:

ݕ ሺͳͷͲ݈ܾሻሺʹͻ݅݊ሻሺͲǤͻͷ݅݊ሻܯ ȁߪȁ ൌ ൌ ൌ ͳ͵͵͵ͷ݌ݏ݅ ൌ ͳ͵Ǥ͵݇ݏ݅ ܫ ͲǤ͵Ͳͻͻ݅݊ସ

Cantilever loading results in:

123

ሺ͵ͲͲ݈ܾሻሺͶʹ݅݊ሻሺͲǤͻͷ݅݊ሻ ȁߪȁ ൌ ൌ ͵ͺ͸ʹͷ݌ݏ݅ ൌ ͵ͺǤ͸݇ݏ݅ ͲǤ͵Ͳͻͻ݅݊ସ

For the outer drop down gates which do not function as access openings, the horizontal span is essentially the same as the side rails. Since the horizontal posts still use schedule 40, the area moment of inertia remains the same and thus so does the stress. 6061 T6 yields at 41 ksi, so the safety factor is slightly lower yet still over three. Mid-span loading on the access gates will LQGXFHOHVVVWUHVVDVWKHPD[LPXPVSDQLVPXFKVKRUWHUDWDSSUR[LPDWHO\´All ladder rungs DUH´ZLGHZKLFKFDQKDQGOHthousands of pounds of force before yielding:

݈ܾ ͲǤͷܲ ሺͺ݅݊ሻሺͲǤͻͷ݅݊ሻ ߪ ൌ ͶͳͲͲͲ ൌ ௠௔௫ ௬ ݅݊ଶ ͲǤ͵Ͳͻͻ݅݊ସ

݈ܾ ସ ቀͶͳͲͲͲ ଶቁ ሺͲǤ͵Ͳͻͻ݅݊ ሻ ܲ ൌ ݅݊ ͵͵ͶͶ݈ܾ ௠௔௫ ሺͲǤͷሻሺͺ݅݊ሻሺͲǤͻͷ݅݊ሻ

Finally, we consider the cantilever loading of a vertical post. Both the outer drop down railings and the central access openings use schedule 80 as vertical posts, which increases the area moment of inertia to 0.3912 in4+HLJKWUHPDLQVDW´DQGWKHVWUHVVLVFDOFXODWHGDVIROORZV

ሺ͵ͲͲ݈ܾሻሺͶʹ݅݊ሻሺͲǤͻͷ݅݊ሻ ȁߪȁ ൌ ൌ ͵Ͳͷͻͺ݌ݏ݅ ൌ ͵ͲǤ͸݇ݏ݅ ͲǤ͵ͻͳʹ݅݊ସ

The drop down gates attach to the frame via large sleeves and long angle sections. There is two square inches of contact between the angle aluminum and sleeve for a solid connection. The ORFNLQJSLQLVò´LQGLDPHWHUDQGZRXOGRQO\H[SHULHQFH.53 ksi of shearing stress if a 300 lb load were applied directly down on a vertical tube.

4 Corner Post Anchorage Stress The corner post will be subjected to bending and torsion. There will also be horizontal shear, but as is the case for most beams, it is negligible and has a value of zero at the outer edge of the box beam, where bending and torsion stress is maximized. The worst case scenario implies that all three anchorage loads are applied at the upper most level of the corner post, in a direction illustrated below in Figure 4.

124

Figure 4. Anchor loading on corner posts.

Equation 1.1 provides the formula for bending stress in the corner post, while torsional shear stress in a thin walled closed shape is defined in Equation 4.1 below. T is the applied torque, t is the wall thickness and ऋ represents the internal area enclosed by the wall centerline.

ܶ ߬ ൌ Equation 4.1 ݐऋʹ

Solving for the bending stress first:

125

ݕ ሺͳͲͺͲͲ݈ܾሻሺͶʹ݅݊ሻሺ͵݅݊ሻܯ ȁߪȁ ൌ ൌ ൌ ʹ͸ͻͶ͹݌ݏ݅ ൌ ʹ͸Ǥͻ݇ݏ݅ ܫ ͷͲǤͷ݅݊ସ

Unsurprisingly, the bending stress is quite large. Now for the shear produced by torsion:

ܾ݈ כ Ͷ͵ʹͲͲ݅݊ ܶ ߬ ൌ ൌ ൌ ͳͶʹͺ݌ݏ݅ ൌ ͳǤͶ͵݇ݏ݅ ݐऋ ʹሺͲǤͷ݅݊ሻሺ͵ͲǤʹͷ݅݊ଶሻʹ

The insignificant value of shear indicates that we need not proceed with a combined stress analysis, especially considering that A572 features a tensile strength of 70 ksi. The safety factor will exceed 2 based on the ultimate strength for our worst case scenario.

126

5 Works Cited

[1] Grating Pacific, "Bar Grating Catalogue," 2004. [Online]. Available: http://www.gratingpacific.com/metal_bar_gratings/metal_bar_gratings.html. [Accessed 30 October 2013].

127

Appendix C: Lifting Mechanism Data

Table of Contents 1 Scissor Analysis ...... 129

2 Hydraulic Actuators ...... 134

3 Hydraulic Pump Information ...... 136

4 Hydraulic Fluid ...... 138

5 Works Cited ...... 140

List of Figures Figure 1. Force determination for a vertical load...... 129 Figure 2. Force determination for a moment about the z axis...... 129 Figure 3. Force determination for a moment about the x axis...... 130 Figure 4. Gear pump schematic...... 136 Figure 5: Piston pump schematic ...... 137 Figure 6: Schematic of vane pump ...... 138 Figure 7. Typical performance data for Arctic 15 hydraulic fluid...... 139

List of Tables TABLE I SPREADSHEET INPUT AND BASIC OUTPUT ...... 131 TABLE II OUTPUTS FROM APPLIED MOMENTS ...... 132 TABLE III TOTAL JOINT LOAD AT BOTTOM LEVEL ...... 133 TABLE IV OVERALL STRESS CALCULATION ...... 134

128

1 Scissor Analysis

The following section provides the formulas used to calculate joint and scissor stress, which were derived in [1]. An example of the spreadsheet utilized to streamline the process is also included.

Figure 1. Force determination for a vertical load [1].

Figure 2. Force determination for a moment about the z axis [1].

129

Figure 3. Force determination for a moment about the x axis [1].

In all the above figures, subscripts R and L indicate right and left sides. M indicates a middle position, while B and F represent back and front. In general, the values do not change much from left to right and back to front. Variables i and Ĭ designate the lift level (with 1 being the uppermost level) and the angle the scissor arms make to the horizontal, respectively. The following table is the spreadsheet input, with basic output assuming a center-center load.

130

TABLE I SPREADSHEET INPUT AND BASIC OUTPUT Number of scissor stacks 4 Level, i 3

Total load, Hy0 (lb) 11000 Effective load (lb) 5500

Level weight, by (lb) 6000 Effective level weight (lb) 3000 $QJOHĬ 9 Member length, d (ft) 12 Width (ft) 20 Moment of inertia (in4) 413 Centerline to perimeter, y (in) 7 14"x10"x0.313" Section area (in.2) 15 Linear weight of members, lb/ft 62.5

Center - Center Load

XBi 94706

XFi -94706

Yi 3625

XMi 148373

Y M i 0

Z i = Z M i 0

Additional outputs include the loads induced by adding moments, which result from offset loads. This output is shown in TABLE II.

131

TABLE II OUTPUTS FROM APPLIED MOMENTS

Induced moment, 54250 35000 Induced moment, Mx (lb-ft) Mz (lb-ft)

Offset to Front, Mz Offset to Side, Mx

YBi 1144 X i 0

YFi -1144 YRi 438

YMi 2289 YLi -438

X M i = X Bi = X Fi 0 X M i 0

Z i = Z M i 0 Y M i 0

X i = 0 0 Z M i = Zi 0

Adding all the contributions together will give us a final value for joint stress under the worst case scenario. This is illustrated in TABLE III.

132

TABLE III TOTAL JOINT LOAD AT BOTTOM LEVEL

Center - Center Load

XBi 94706

XFi -94706

Yi 3625

XMi 148373

Y M i 0

Z i = Z M i 0

Summary

XBRi 94706

XBLi 94706

XFRi -94706

XFLi -94706

YBRi 5207

YBLi 4332

YFRi 2918

YFLi 2043

XMRi 148373

XMLi 148373

YMRi 2289

YMLi 2289

Finally, the overall safety factor on the highest stressed areas is automatically computed. This simply entails resolving the joint reactions into tangential and perpendicular loads on the scissor member. With knowledge of angle, this is a simple computational task.

Not shown in this appendix are the various other spreadsheets used to evaluate alternate hydraulic actuator locations. These typically became quite complex with matrix operations

133 involved. The equations generated from different actuator placements generally cannot be solved by hand.

TABLE IV OVERALL STRESS CALCULATION Stress at back corner:

Axial force (lb): 94355 Axial stress (ksi): 6.29

Bending moment (lb-in): 1436977 Bending stress (ksi): 24.36

TOT 30. 65 S.F. 2.08

The scissor lift must also be designed to resist buckling at a safety factor of 2 to 1 [2]. Due to time constraints, this calculation is not shown. However, these calculations were performed and it was determined that there was no concern regarding buckling.

2 Hydraulic Actuators

The following section contains the calculations used to determine the appropriate hydraulic components for the ALP assembly.

The required piston pressure was determined from the load incurred in each hydraulic cylinder at the beginning of the lifting cycle, as this value is the maximum throughout the entire lift.

ܨ ͳʹͲǡͲͲͲ݈ܾ ܲ ൌ ൌ ൌ ʹ͵ͺ͹݌ݏ݅ ͳ ܣ௣௜௦௧௢௡ ሺ ሻଶ ቀͶቁ ߨ ͺ݅݊

With a required actuator stroke of 40 inches, individual cylinder volume was determined to be:

134

ଷ ͶͲ݅݊Ǥ ൌ ʹͲͳͳ݅݊ כ ௣௜௦௧௢௡ܣ ௖௬௟௜௡ௗ௘௥ ൌܸ

Total hydraulic volume was determined according to the following equation:

ͲǤͲͲͶ͵ʹͶ݈݃ܽǤ ൬ ൰ ൌ ͵ͶǤ͹ͺ݈݃ܽǤ כ ൌ ͺͲͶͶ݅݊ଷ ܸ כ ൌ Ͷ ܸ ௧௢௧௔௟ ௖௬௟௜௡ௗ௘௥ ͳ݅݊ଷ

Hydraulic horsepower was determined according to the following equation:

ሺܩܲܯሻሺܲܵܫሻ ܪܲ ൌ Equation 2.1 ௛௬ௗ௥௔௨௟௜௖ ͳ͹ͳͶ

To check the feasibility of lifting the platform in under 1 minute, 35 gallons per minute is used. This corresponds to the cylinders being completely full (totally extended) in one minute. This value of necessary horsepower is above what is actually required, since the demand on the hydraulic system decreases as the lifting angle increases. An electronic control system could modulate these parameters to provide a relatively constant lifting rate. The design and implementation of such as system is beyond the scope of this project.

݅ݏͶͲͲ݌ʹ כ ͷ݃݌݉͵ ܫܵܲ כ ܯܲܩ ܪܲ ൌ ൌ ൌ Ͷͻܪܲ ͳ͹ͳͶ ͳ͹ͳͶ

Due to inefficiencies in the hydraulic rams and pump, we must account for a typical 15% loss in hydraulic power [3]. Therefore, rated hydraulic horsepower is was determined to be:

Ͷͻܪܲ ܪܲ ൌ ൌ ͷ͹Ǥ͸ܪܲ ௥௔௧௘ௗ ͺͷΨ݂݂݁݅ܿ݅݁݊ܿݕ

Assuming the motor to be 90% efficient, we determine the required electric motor power to be:

ͷ͹Ǥ͸ܪܲ ܪܲ ൌ ൌ ͸Ͷܪܲ ௘Ǥ௠௢௧௢௥ ͻͲΨ݂݂݁݅ܿ݅݁݊ܿݕ

135

3 Hydraulic Pump Information

A hydraulic system was selected to drive the lifting mechanism. An introduction to hydraulic systems and the analysis used to size the hydraulic system is outlined below.

All hydraulic systems are based on pressure and flow requirements. The flow rate determines the speed that the hydraulic cylinder, in this case, actuator, extends. The pressure determines KRZPXFKIRUFHZLOOEHH[HUWHG3DVFDO¶V/DZJRYHUQVZK\K\GUDXOLFVDUHDYLDEOHRSWLRQ 3DVFDO¶VODZVWDWHVWKDWDSUHVVXUHDSSOLHGWRDFRQILQHd fluid will be transmitted adiabatically and with equal intensity throughout the fluid and the surfaces of the container [3].

Hydraulic pumps convert mechanical energy from a motor into fluid energy in terms of flow. Pressure is created when the flow produced by the pump meets resistance. Hydraulic pumps can be either uni-rotational or bi-rotational. Gear, vane and piston are the three typical types of hydraulic pumps [3].

Gear Pumps Relative to other types of pumps, gear pumps are inexpensive, are more tolerant to contamination and have few moving parts making them easy to service. Gear pumps operate by trapping fluid between the teeth of the gears and forcing the fluid out of the gear cavity to the outlet port [3]. A schematic of a gear pump is shown in Figure 4.

Figure 4. Gear pump schematic [3].

136

Gear pumps have a minimum along with a maximum operating speed. The maximum operating speed is based mainly on the maximum limit the rotating gears can fill the gear cavity without causing cavitation. Gear pumps have a minimum operating speed due to the inefficiencies that arise. Gear pumps are more efficient in the upper third of their operating speed. Below this, they are very inefficient and generate an immense amount of heat [3].

Piston Pumps Piston pumps can generally tolerate more pressure than gear pumps and are therefore ideal for high-pressure requirements. Piston pumps generally have a higher initial cost, are more prone to contamination and are more complex. Piston pumps have a cylinder block containing four, nine or eleven pistons. As these pistons move in and out they draw fluid from the inlet and force it through the outlet [3]. A schematic of a piston pump is shown in Figure 5.

Figure 5: Piston pump schematic [3]

7KHYROXPHRIWKHSLVWRQ¶VF\OLQGHUGHWHUPLQHVWKHGLVSODFHPHQWRIWKHSXPS3LVWRQSXPSVDUH available in both fixed and variable displacement. A fixed displacement pump has a fixed swash plate angle. In a variable displacement pump the swash plate angle is not fixed, but is controlled by a compensator. This angle is adjusted by a pressure signal from a directional valve [3].

137

Vane Pumps A vane pump operates similar to a gear pump. The different being there is only one set of vanes as opposed to two gears. The area on the inlet side of the vanes will increase, while the area on the outlet side will decrease. This change in area allows for the hydraulic fluid to be drawn in from the inlet and released at the outlet [3]. A schematic of a vane pump is given in Figure 6.

Figure 6: Schematic of vane pump [3]

4 Hydraulic Fluid

While on a site visit to GLACIER in Thompson the team was able to speak with Ben, a mechanic at United Rentals. Ben recommended Arctic 15 hydraulic fluid for operations in cold environments such as Thompson, MB. This fluid remains relatively thin at cold temperatures compared to other hydraulic fluids. The properties of Arctic 15 hydraulic fluid are given in Figure 7.

138

Figure 7. Typical performance data for Arctic 15 hydraulic fluid [4].

139

5 Works Cited

[1] H. M. Spackman, "Mathematical Analysis of Scissor Lifts," U.S. Marine Corps, Quantico, VA, 1989.

[2] Canadian Standards Association, B354.1-04 Portable elevating work platforms, Canada, 2004.

[3] Munchie Power Products, "Understanding Truck Mounted Hydraulic Systems," Munchie Power Products, 2006.

[4] Petro Canada, 2013. [Online]. Available: http://lubricants.petro- canada.ca/resource/download.aspx?type=TechData&iproduct=162&language=en. [Accessed 20 November 2013].

140

Appendix D: Base Frame and Steering Component Data

Table of Contents 1 Bearing and Seal Data ...... 142

2 Spindle and Bearing Stress ...... 144

3 Wheel Arm Stress ...... 148

4 Main Frame Stress Considerations ...... 150

5 Works Cited ...... 153

List of Figures Figure 1. Definition of geometries for single row tapered roller bearings...... 142 Figure 2. Tapered roller bearing catalog entry...... 142 Figure 3. Ball bearing geometries...... 143 Figure 4. Radial ball bearing catalog entry...... 143 Figure 5. Lip seal data...... 143 Figure 6. Model specific seal data...... 143 Figure 7. Spindle loading schematic...... 145 Figure 8. Forces acting on wheel arms...... 149 Figure 9. Frame beam loading...... 151

List of Tables TABLE I SUMMARY OF LIFT WEIGHT ...... 146

141

1 Bearing and Seal Data

All rolling element bearings and seals are Timken, although equivalent items from other manufacturers could be used. Specifics of each item are provided below.

Figure 1. Definition of geometries for single row tapered roller bearings [1].

Figure 2. Tapered roller bearing catalog entry [1].

142

Figure 3. Ball bearing geometries [1].

Figure 4. Radial ball bearing catalog entry [1].

Figure 5. Lip seal data [1].

Figure 6. Model specific seal data [1].

143

2 Spindle and Bearing Stress

Due to the weight of the aerial lift platform and the offset necessary to achieve caster, the spindles and support bearings experience high levels of stress. The forces involved work primarily to bend the spindle and load the bearings radially. The tapered roller bearing counters the entire vertical load, and both bearings assume equal but opposite loads in the radial direction.

Other components subject to these forces include the wheel axle shaft and the caster bracket itself. The wheel axle dimensions have been provided by the rim manufacturer and feature a high factor of safety. Furthermore, the caster assembly loads the shaft in double shear, reducing the shearing stress by half. Bending of the shaft between the caster bracket and bearing is a concern. It is recommended to consult with the rim manufacturer to find the best shaft design for the application. Rims will come loaded with bearings and seals of an appropriate safety factor.

Due to the difficulty of performing stress analysis on the caster bracket, this component has been SXUSRVHO\RYHUEXLOWXVLQJ´WKLFNVWHHO7RFOHDUWKHZKHHOWKHEUDFNHWUHTXLUHVDQRYHUDOOGHSWK of 15-ò´Zhich helps resist bending. Bearing stress between the axle and the bracket is minimal due to the available rectangular projected DUHDRI´SHUVLGHWRVXSSRUWWKHVKDIWIf we assume the allowable bearing stress is half of the yield for A572:

ͳ ቀ ቁ ܨ ʹͷǡͲͲͲ݌ݏ݅ ൌ ʹ ʹ݅݊ଶ ܨ ൌ ͳͲͲǡͲͲͲ݈ܾ

Thus, the theoretical axle load, F, is 100,000 lb; almost an order of magnitude greater than required. It is assumed that the wheel housing, as a welded unit, can handle the bending stress placed on it. FEA could be performed to verify this assumption.

144

A diagram of the spindle loading is located in Figure 7. Analysis is provided to verify that the spindle size and bearing loads are acceptable for the application.

Figure 7. Spindle loading schematic.

The force F in the above figure represents the load each wheel is required to handle. The platform is designed with a fairly neutral center of gravity, but there are some unequal distributions of weight resulting from scissor placement. The total weight of the unit, not

145 including rims and caster housings (items below the spindle) will be no more than 50,000 lb at its rated load. There is some uncertainty due to the fact that two structures have not yet been fully VSHFLILHGWKHLQWHUPHGLDWHDFFHVVSODWIRUPVDQGWKHKLWFKDVVHPEO\7KHUHPDLQGHURIWKHOLIW¶V weight can be accurately calculated from the material summary. This is summarized in TABLE I.

TABLE I SUMMARY OF LIFT WEIGHT Component Weight (lb)

Platform 6711 Scissors 21856 Base 8638 Rated load 3500 Actuation 3500 Intermediate platforms 2400 Hitch 2000

TOTAL 48275

Italicized weights are the estimated values and are purposely overstated where large uncertainty exists. Actuation includes the rams, hydraulic lines, reservoir, pump and electric motor. We take the maximum possible weight as 50,000 lb for the purpose of calculating spindle loads.

If the material was perfectly distributed, each wheel would be responsible for one quarter of the total load (12,500 lb). Because this is not entirely true, we will assume a worst case scenario of 15,000 lb at a spindle. This could occur for reasons beyond uneven loading, such as towing forces and other dynamic loads. Thus, F becomes equal to 15,000 lb. The moment created on the spindle assembly is:

ሺͳͷǡͲͲͲ݈ܾሻሺͻǤͷ݅݊ሻ ൌ ͳͶʹǡͷͲͲ݈ܾ Ȉ ݅݊

146

5HJDUGOHVVRIWKHFDVWHUURWDWLRQWKH´offset will always stay constant. The bearing spread is ´VSHFLILFDOO\PDGHODUJHWRKHOSUHGXFHUDGLDOIRUFHV Performing a moment summation about the lower tapered roller bearing:

ሺ ሻ ሺ ሻ ෍ ܯ஺ ൌ ͳͶʹǡͷͲͲ݈ܾ Ȉ ݅݊ െ ܨோೣ ͳ͵݅݊Ǥ ൌ Ͳ

ܨோೣ ൌ ͳͲͻ͸ʹ݈ܾ

Since the only forces in the x direction are the two radial reactions at the bearings, it is clear from a force summation that:

ܨ்ೣ ൌ ͳͲͻ͸ʹ݈ܾ

Positive values indicate that the forces are shown in the correct direction in Figure 7. It is clear that both bearings need the same load capacity in the radial direction; something greater than 10962 lb. Referencing this value with Figure 2 and Figure 4 proves that the bearings are acceptable in this application. To increase the safety factor of the upper bearing, which is near its limit, it is recommended that two of the same bearings be stacked on the spindle. The only required modification would be machining the housing shoulder lower; everything else remains the same.

The spindle shaft itself is fabricated from AISI C1144 stress relieved steel, which yields at 100 ksi. There is also a stress concentration at the shoulder; a stress concentration factor K of 2.5 will be applied to be conservative [2]. Bending is a maximum at the first lower tapered bearing and linearly decreases to zero at the roller bearing. Applying a safety factor of 2, we can find the required yield strength of the shaft and compare it to the 100 ksi actual value.

ݕܯܭ ͳ ߪ ൌ ൬ ൰ ߪ ൌ Equation 2.1 ௔௟௟ ʹ ௬ ܫ ͶǤͷ݅݊ ሺ ሻሺ ሻ ʹ ʹǤͷ ͳͶʹǡͷͲͲ݈ܾ Ȉ ݅݊ ቀ ʹ ቁ ߪ௬ ൌ ߨ ൌ ͹ͻ͸Ͷ͵݌ݏ݅ ൌ ͹ͻǤ͸݇ݏ݅ ቀ ቁ ሺͶǤͷ݅݊ሻସ ͸Ͷ

A material yielding at 79.6 ksi would theoretically provide a safety factor of 2 under this scenario. Since this value is less than 100 ksi, we can state that the spindle has a safety factor

147 greater than two on the bending produced by the caster offset. The only other force to consider is compression, which ends at the tapered bearing prior to the shaft sWHSSLQJGRZQWR´.

ͳͷǡͲͲͲ݈ܾ ߪ ൌ ൌ ͸͵ͳ݌ݏ݅ ൌ ͲǤ͸͵ͳ݇ݏ݅ ௖௢௠௣ ͲǤʹͷߨሺͷǤͷ݅݊Ǥ ሻଶ

This is obviously negligible. Similarly, shearing stresses in the spindle due to the radial bearing loads will also be small due to the large cross-sectional area of the spindle. Generally, shear will not compound bending extensively since the maximum values of each occur along planes where the other is zero.

3 Wheel Arm Stress

The wheel arm will be subjected to bending and torsion due to the caster offset. Maximum bending will occur when the wheel is parallel to the arm and has its contact patch outboard of the spindle. Here, the offset of the caster housing will add to the moment arm created by the wheel arm itself. Maximum torsion will occur when the wheel is perpendicular to the arm, in either direction. Here, the torsion is caused by the caster offset alone and thus is relatively small. Clearly, the situations for maximum bending stress and maximum torsion are mutually exclusive. It can be shown that the wheel arm section is strong enough to handle both maximum values at the same time, even though this cannot actually occur during operation. Both scenarios are shown schematically in Figure 8.

148

Figure 8. Forces acting on wheel arms.

The upper diagram illustrates the maximum bending moment arrangement, while the second drawing represents maximum torsion. We are not considering the gusset plates here, but they will undoubtedly add to the strength of the entire assembly. They will also reduce the angle of twist in the horizontal arm. The bottom figure gives the moments of inertia for the HSS beam.

149

The maximum moment is simply the wheel load, F, multiplied by combined the moment arm.

ܯ ൌ ሺͳͷǡͲͲͲ݈ܾሻሺ͵͸݅݊ ൅ ͻǤͷ݅݊ሻ ൌ ͸ͺʹǡͷͲͲ݈ܾ Ȉ ݅݊

The maximum bending stress can be found using the basic form of Equation 2.1:

ͳͲ݅݊ ݕ ሺ͸ͺʹǡͷͲͲ݈ܾ Ȉ ݅݊ሻ ቀ ቁܯ ߪ ൌ ฬ ฬ ൌ ʹ ൌ ͳͶʹͳͻ݌ݏ݅ ൌ ͳͶǤʹ݇ݏ݅ ܫ ʹͶͲ݅݊ସ

Torsion in the member can be found using the following:

ܶ ߬ ൌ Equation 3.1 ݐऋʹ

The torsion in the member is the force F multiplied by the caster offset:

ܶ ൌ ሺͳͷǡͲͲͲ݈ܾሻሺͻǤͷ݅݊ሻ ൌ ͳͶʹǡͷͲͲ݈ܾ Ȉ ݅݊

Taking this value into Equation 3.1, we find the following torsional shear:

ܶ ͳͶʹǡͷͲͲ݈ܾ Ȉ ݅݊ ߬ ൌ ൌ ൌ ͳ͹ͳͻ݌ݏ݅ ൌ ͳǤ͹ʹ݇ݏ݅ ͷ ݐऋ ሺ ଶሻʹ ʹ ቀͳ͸ ݅݊ቁ ͳ͵ʹǤ͸݅݊

Torsional shear is negligible. If it had been more significant, the principal stresses would have been found and the von Mises stress calculated evaluate the safety of the beam. When the shear is minimal, the von Mises stress remains close to the normal stress found above (14.2 ksi). Even a von Mises stress of 16 ksi represents a safety factor greater than 3.

4 Main Frame Stress Considerations

It is impossible to accurately calculate the stress distribution throughout the frame by hand. Some simplifying assumptions can lead to the absolute certainty that the frame is overbuilt. One example of this is the main cross beams which run under each scissor pair. Each of these beams

150 is placed such that the neutral axis offset, as compared to the pivots, is only 2-ò´%HFDXVHWKH horizontal force is so dominant, it is important to minimize the distance allowed for bending. Having a clevis mount on either side of the beam serves another purpose; cancel the bending moment in the horizontal plane about that center beam. Thus we are only worried about axial force, which is tensile, and the bending due to the slight neutral axis offset. The upper two diagrams in Figure 9 provide views of the horizontal loading on the clevis mounts, while the lower schematic illustrates the downward force on the I-beam supporting the slide channel.

Figure 9. Frame beam loading.

The horizontal load at the clevis pins in Figure 9 are above and beyond the likely maximum force, adding an increased safety factor on all calculations. Determination of the forces in scissor members and at the joints can be found in Appendix C.

151

Each main cross beam is exposed to 240,000 lb of axial tension force. Furthermore, the slight offset of 2-ò´SURGXFHVDPRPHQWRI600 in-lb acting to bend the beam in a concave down 3 2 IDVKLRQ7KH´[´[ /8´EHDPVKDYHDFURVVVHFWLonal area of 17 in and a moment of inertia of 480 in4 about the bending axis. The stresses are then resolved as follows:

ʹͶͲǡͲͲͲ݈ܾ ߪ ൌ ൌ ͳͶͳͳͺ݌ݏ݅ ൌ ͳͶǤͳ݇ݏ݅ ௔௫௜௔௟ ͳ͹݅݊ଶ

ͳͶ݅݊ ሺ͸ͲͲǡͲͲͲ݈ܾ Ȉ ݅݊ሻ ቀ ቁ ߪ ൌ ʹ ൌ ͺ͹ͷͲ݌ݏ݅ ൌ ͺǤͺ݇ݏ݅ ௕௘௡ௗ௜௡௚ ͶͺͲ݅݊ସ

These forces will compound on the top of the beam in tension, yielding a total value of 22.9 ksi. Since the yield strength of A572 is 50 ksi, this represents a safety factor of 2.18 which exceeds the requirement of 2 [3]. Buckling is not a concern as the beam is not loaded in compression.

As the scissors raise the platform, there will be a point where the movable member is at the mid- span of the supporting I-beam (bottom of Figure 9). The greatest possible vertical force is just below 6000 lb, which considers extreme load offset as well as all scissor member and platform component weights (Appendix C). A load of 6000 lb is used to be conservative. The moment generated about the mid-span is 297,000 in-lb, and the W8 x 15 I-beam features a moment of inertia of 48 in4. The resulting bending stress is calculated as follows:

ͺ݅݊ ሺʹͻ͹ǡͲͲͲ݈ܾ Ȉ ݅݊ሻ ቀ ቁ ߪ ൌ ʹ ൌ ʹͶǡ͹ͷͲ݌ݏ݅ ൌ ʹͶǤͺ݇ݏ݅ ௕௘௡ௗ௜௡௚ Ͷͺ݅݊ସ

Once again, this value is less than half of the yield of A572, meeting the requirement for allowable stress. It should also be noted that this value does not take into account the contribution from the C-channel slide, which will further strength the entire sliding support.

152

5 Works Cited

[1] Timken Bearings, "Roller Bearing Catalog," [Online]. Available: http://www.timken.com/en- us/Pages/Home.aspx. [Accessed 10 November 2013].

[2] R. L. Mott, Machine Elements in Mechanical Design, 4th ed., Upper Saddle River, NJ: Pearson Prentice Hall, 2004.

[3] Canadian Standards Association, B354.1-04 Portable elevating work platforms, Canada, 2004.

153

Appendix E: Detailed Tire and Wheel Data

1 Table of Contents

1 Tires ...... 155 2 Wheels ...... 159 2 Works Cited ...... 163

Table of Figures Figure 1. Technical datasheet for 28x16x22 MH20 smooth tread tires from Continental (used with permission)...... 156 Figure 2. Deflection and ground pressure values due to tire load (used with permission from Continental)...... 157 Figure 3. Detailed quotation from sales representative, Tim Hubbert, of Continental Tire Canada (used with permission)...... 158 Figure 4. Preliminary heavy duty wheel drawing, from Precision Products...... 160 Figure 5. Preliminary drawing of lighter duty wheel from Precision Products...... 161 Figure 6. Drawing of selected wheel from Precision Products (used with permission)...... 162

154

This appendix provides the detailed drawings and product specification sheets for the tires and wheels that were selected.

1 Tires

Continental Tire was chosen as the supplier for tires. Continental Tire is located in Mississauga, ON. A Canadian supplier was ideal in an effort to reduce shipping costs. Tim Hubbert, the Canadian sales representative from Continental was contacted for a quote.

Tire size 28x16x22 MH20 smooth tread, press on band tires were selected. These were selected for their high load capability, size and durability [1].

The technical data sheet outlining the load capabilities at various speeds can be found in Figure 1. The deflection and pressure imposed on the ground due to the tire load is given in a graphical format in Figure 2.

A detailed quotation on the tires is given in Figure 3. A quote of $619.49 per tire is given. This equates to a total of $2477.96 for all four tires. Taxes and freight costs would be additional to the quoted value [2].

155

Figure 1. Technical datasheet for 28x16x22 MH20 smooth tread tires from Continental (used with permission) [1].

156

Figure 2. Deflection and ground pressure values due to tire load (used with permission from Continental) [1].

157

Figure 3. Detailed quotation from sales representative, Tim Hubbert, of Continental Tire Canada (used with permission) [2].

158

2 Wheels

The initial hope was to purchase the tire and wheel as a set. However, Continental Tire does not sell the wheel for the selected tire. Tim Hubbert referred the team to David Hartig of Precision Products. They specialize in wheel manufacturing and have worked with Continental Tire many times.

After making David aware of the loading requirements and dimensions of the tire, he created two preliminary drawings which can be found in Figure 4 and Figure 5. Figure 4 displays a heavy duty wheel and is commonly produced by Precision Products. Figure 5 is a wheel of similar dimensions, however, it is intended for lighter weight applications, which is evident by the thinner steel wheelbase and centre support.

After discussions with David of Precision Products, he created a new drawing of a wheel shown in Figure 6. This wheel design incorporates the centre wheel attachment of the lighter weight wheel with the thicker wheelbase and supports of the heavy duty wheel. Not all of the firm dimensions have ben given in the drawing to accommodate customer customization. Precision Products will give the customer some options in terms of shaft size and bearing type. Lastly, Precision Products will run finite element analysis (FEA) on the product prior to the sale. They also recommend doing some non-destructive testing on the product [3].

No formal quotation was given as for the tires, however, through email correspondence, one was given. The price per wheel is $1832.00, putting the total for the four wheels at $7,328.00 [3].

159

Figure 4. Preliminary heavy duty wheel drawing, from Precision Products [3].

160

Figure 5. Preliminary drawing of lighter duty wheel from Precision Products [3].

161

Figure 6. Drawing of selected wheel from Precision Products (used with permission) [3].

162

2 Works Cited

[1] Continental. (2011, Novemebr) Datasheet - Elastic Tires for industrial Vehichles. Technical Data Sheet.

[2] Tim Hubbert, Tire Quotation, November 14, 2013, Tire quotation for tires from Continental Tires Canada. Tires were selected for specific application.. [3] David Hartig, Re: Continental Tire - Price Quote, November 18, 2013, E-mail communication.

163