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Indiana-University – Purdue University Fort Wayne

Department of Engineering

ME 487 - ME 488

Capstone Senior Design Project

Final Report

Project Title: Design of Heavy Duty Rail Test Machine

Members: Manuel Ramirez, Cody Huber,

Matt Shultz, Nick Moynahan Advisor: Zhuming Bi

Date: Tuesday, May 5, 2015

Sponsor:

Table of Contents

List of Figures and Tables 4 Acknowledgements 6 Abstract 7 1. Problem Statement 8

1.1 - Introduction 9 1.2 - Requirements & Specifications 10 1.3 - Design Parameters 11 1.4 - Design Variables 11 1.5 - Limitations & Constraints 12 1.6 - Additional Considerations 12

2. Conceptual Designs 14 2.1 - Design of Machine Frame 15 2.2 - Hydraulic System Design 17 2.3 - Energy Absorption Design 18

3. Conceptual Design Evaluation 21 3.1 - Concept Evaluation Overview 22 3.2 - Frame Concept Evaluation 23 3.3 - Hydraulic System Concept Evaluation 23 3.4 - Absorption Material Concept Evaluation 23 3.5 - Conceptual Design Evaluation 24

4. Detailed Design 25 4.1 –Detailed Design Overview 26 4.2 –Final Frame Design 26

4.2.1 - Front Faceplate 26 4.2.2 - Rear Faceplate and Inner Frame Supports 30

4.3 –Hydraulic System 33 4.4 – Energy Absorbing Material 34

4.4.1 - Calculations 39 5. Design Modifications 42

5.1 - Overview of Changes 43 5.2 - Frame Design Modifications 43 5.3 - Hydraulic Design Modifications 43 5.4 - Energy Absorbing Design Modifications 46

6. Design Testing & Results 48 6.1 - Testing Preparation: Hydraulic System Calibration 49 6.2 - Testing Procedure 49 6.3 - Frame Results 51 6.4 - Energy Absorption Results 52

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7. Cost Analysis 54 8. Conclusion 57

8.1 - Recommendations 58 8.2 - Project Summary 58

References 60 Appendix A: Tables & Figures 63 Appendix B: HPU Component List 69

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List of Figures and Tables Figure 1.1:Front view of the rail bending test machine done in SolidWorks Figure 1.2: Cross-section of rail specimen (actual size not shown here) Figure 1.3: 141 Weld slow bend test Figure 2.1: Minimum factor of safety of original frame design with pressure increase Figure 2.2: Hydraulic system schematic; highlighted pump cannot output 400 bar Figure 2.4: Example of corrugated plastic Figure 2.5: Example of aluminum honeycomb Figure 2.6: Example of Sorbothane sheet Figure 2.7: Example of a shock absorber Figure 4.1: Faceplate body diagram Figure 4.2: Bending moment and shear force diagrams Figure 4.3: Free body diagram input values and calculation results Figure 4.4: FEA analysis of frame with front faceplate thickness of 3 inches Figure 4.5: Rear view of FEA analysis of frame with front faceplate thickness of 3 inches Figure 4.6: Assembly of all added components that were added to the frame Figure 4.7: Rear view of FEA analysis of final frame assembly Figure 4.8: Shock response of Sorbothane compared to similar products Figure 4.9: Dimensions of Sorbothane sheet for bottom basin Figure 4.10: Dimensions of Sorbothane sheet for side slots Figure 5.1: Front plate modifications Figure 5.2: Top plate modifications shown in SolidWorks Figure 5.3: Minimum FOS of 2.05 after modifications Figure 5.4: Corrugated cardboard configurations Figure 5.5: Corrugated cardboard block dimensions Figure 5.6: Finished corrugated cardboard blocks Figure 5.6: Corrugated cardboard blocks initial testing Figure 6.1: HPU test setup Figure 6.2: a) Layer of wood acting as a base for the b) corrugated cardboard blocks. c) The rail still has enough clearance for testing after cardboard blocks are added. Figure 6.3: Machine loaded and ready to run tests (guard screen is up to show rail fitment) Figure 6.4: Machine after test was completed Figure 6.5: Minimum factor of safety at max force achieved during testing, 656,315 lbf Figure 6.6: Damage on the side corrugated cardboard blocks Figure 6.7: Wood layer is barely damaged thanks to cardboard layer Figure 6.8: Severely damaged corrugated cardboard block Figure 8.1: Channels suggested for housing the corrugated cardboard blocks

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Table 1.1: Rail hardness table for standard chemistry rail steel Table 1.2: Tensile properties of standard chemistry rail steel Table 3.1: Category weighted decision matrix showing the total score and rank of each category

for frame design concepts Table 3.2: Manuel’s grading matrix for frame conceptual designs Table 3.3: Decision matrix for the frame concepts Table 3.4: Decision matrix for the hydraulic system concepts Table 3.5: Decision matrix for the energy absorbing material conceptual designs Table 3.6: Conceptual design selection Table 4.1: Mechanical properties of A36 steel Table 4.2: Selected pump specifications highlighted. Table 4.3: Selected pump output flow specifications highlighted. Table 4.4:Weldable motor dampening bars schematic with table of dimensions. Table 7.1: Cost analysis of the frame. Table 7.2: Cost analysis of the hydraulic system. Table 7.3: Cost analysis of the energy absorbing material. Table 7.4: Total cost of project.

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Acknowledgements The design team would like to thank Steel Dynamics, Inc. (SDI) for sponsoring and making the rail bending test machine project possible. Special thanks goes to Andrew Weber of SDI for his support throughout the design process in providing necessary information and specifications. Thanks also to Dr. Zhuming Bi for his assistance and expertise throughout all phases of the design process. We would also like to thank Scott Williams (Flodraulic Group), Doug Cope (Kelly Box), and Power Components for their help and advice throughout the project.

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Abstract Steel Dynamics, one of the largest steel producers in the United States, needs a rail bending test machine that can operate at a higher pressure to test rails. Originally, the test machine was not able to deliver the pressure required to test new rails. The strengths of new rails are higher than those of previous rails, which require a high pressure to perform the tests. The objective of this project was to redesign the test machine to meet new operating requirements in sense that (1) higher strength rails can be tested; (2) better materials can be used to absorb energy. The original machine uses wood to absorb the energy released upon rail fracture, but the wood does not successfully protect the machine from damage. In this report, we provide the successful detailed design for the following tasks: (1) designing a hydraulic system that was capable of delivering the necessary pressure for rail tests, (2) strengthening the machine frame to withstand the increased loads, and (3) selecting cost-effective materials to absorb the energy released when the rail fractures.

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1. Problem Statement

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1.1 – Introduction

Steel Dynamics, Inc. has a test machine for testing the strengths of welds on rails. The machine structure is shown in Figure 1.1. Originally, the test machine ran at the maximum pressure of 350 bar, but the company was looking to increase the force delivered by their test press for welded joints. SDI contacted the hydraulic cylinder manufacturer and confirmed that the cylinder could run continuously at a higher pressure of 406 bar. However, increasing the hydraulic pressure required (1) upgrading the hydraulic system, excluding the cylinder, (2) reinforcing the machine frame for a higher load, and (3) using better material to absorb energy when test specimens experience fracture. SDI requested the redesign of the original test machine to meet these requirements.

Figure 1.1: Front view of the rail bending test machine

made in SolidWorks

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1.2 - Requirements & Specifications To clarify the objectives of the project, the requirements of the new design were quantified as follows.

● Maintain frame/support integrity after the required increase in hydraulic pressure to 400

bar. The frame originally supported a pressure of 350 bar with a factor of safety of 1.98. An increase in pressure was needed to fully deform a test specimen composed of stronger materials than those in standard rail. If an increase of pressure was applied on the original machine, the factor of safety of the machine would decrease to 1.62. SDI expected to better the factor of safety of 1.98 for the new test machine. The increase in stress may also cause undesired deformation in the frame and compromise its integrity.

● The mainframe of the machine could not be modified except for material additions. SDI

gave us the liberty to add material to the base frame as we saw fit, so long as the original frame remained intact. Modifications to the original frame would have introduced more uncertainties as well as undesired expense. Therefore, material had to be added to the frame in the form of structural reinforcements. In addition, the material added could not interfere with the operation of the machine.

● The test press was to be able to deflect the rail a minimum of 0.75 in. according to AREMA standards. The Figure 1.2 shows the cross-section of the rail, Table 1.1 describes the hardness requirements of the rail and Table 1.2 shows the minimum yield strength for different rail types.

● Any changes were to be easily integrated into the original machine. Table 1.1: Rail hardness table for standard chemistry rail steel

Type of Rail Minimum Surface Brinell Hardness, HB Standard Rail 310 High Strength Rail 370 Note 1: Hardness specified above shall be maintained in the head area only. Note 2: A fully pearlitie microstructure shall be maintained in the head. Note 3: If 410 HB is exceeded, the microstructure through the head shall be examined at 100X or higher for confirmation of a fully pearlitic microstructure in the head. Note 4: No untempered martensite shall be present within the rail. Table 1.2: Tensile properties of standard chemistry rail steel

Description Standard High-Strength Yield Strength, ksi, minimum 74.0 120.0 Tensile Strength, ksi, minimum 142.5 171.0 Elongation in 2 inches, percent, minimum 10 Note 1 10 Note 1

Note 1: Up to 5% of the order may be less than 10% elongation if purchaser’s authorized representative and supplier agree, but in no case may the elongation be less than 9%

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Figure 1.2: Cross-section of rail (actual size not shown here), by Harmer Steel Products Co.

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1.3- Design Parameters

Design parameters correspond to the system parameters that are not subjected to changes. As illustrated in Figure 1.3, the following are the design parameters of the test machine: ● Two point loading ram to apply the force was determined by test standards and

it could not be modified. ● The support locations with respect to the load could not be altered. ● The original hydraulic 13 inch cylinder had to be used. ● The basic frame structure had to remain unchanged. ● The rail welds had to be able to pass a minimum slow bending test as

prescribed by AREMA standards for head hardened rail: ○ AREMA Minimum Specification

■ Modulus of Rupture: 125 ksi ■ Deflection: 0.75 in.

Figure 1.3: 141 Weld slow bend test

1.4 - Design Variables

Design variables correspond to the system parameters that are subjected to changes. In this project, the followings were selected as design variables: ● Locations, geometries, and types of added materials

○ Any steel added could not not interfere with the testing operation ○ Additions could not require change in orientation of the machine and

should allow machine to remain in its original location

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● Energy absorption materials

○ Material had to be easily replaceable and cost-effective ● Hydraulic system changes had to allow for flow of 400 bar on a long-term, high

demand basis 1.5 - Limitations & Constraints Limitations and constraints are restrictions of the system. In this project, the followings limitations and constraints were identified:

● Original frame had to remain intact ○ Original overall machine dimensions - width 79.52 in (2020 mm) x depth 41.34

in (1050 mm) x height 91.50 in (2324 mm) ○ Structural integrity could not be compromised ○ Factor of safety was originally 1.98 ○ For high strength steel minimum deflection is 0.75 in

● Hydraulic cylinder of original design had to be used (max pressure of 406 bar). See Figure 2.2

● Additions to reinforce frame had to be made from ASTM A36 ● Had to make use of on-hand materials at SDI ● Had to retain a minimum factor of safety of 2 for the frame

○ Could be higher ● $15,000 budget for purchased components and steel produced by SDI

1.6 - Additional Considerations

The following were design considerations not mentioned above:

● Safety: The design could not add new safety hazards to the work

environment ○ OSHA 1910.212(a)(1) - One or more methods of machine

guarding shall be provided to protect the operator and other employees in the machine area from hazards such as those created by point of operation, ingoing nip points, rotating parts, flying chips and sparks. Examples of guarding methods are-barrier guards, two-hand tripping devices, electronic safety devices, etc.

■ Any changes to the frame should not decrease the effectiveness of the original guarding, which protects the operator and surrounding employees while a test is being

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conducted

● Maintenance: The design could not add new maintenance requirements ○ OSHA 1910.219(p)(1) - All power-transmission equipment shall

be inspected at intervals not exceeding 60 days and be kept in good working condition at all times

The entire hydraulic system should be inspected at the specified intervals in the standard above to check for leaks in fluid lines and hoses, secured hose connections and fittings, and overall safety

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2. Conceptual Designs

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The first step in generating conceptual designs was separating the machine system into three separate subsystems, i.e., the machine frame, the hydraulic system, and the energy absorbing material/design. Several brainstorming sessions were held to produce multiple design options that fulfill the specific design requirements and specifications for each system component.

2.1 - Design of Machine Frame

The original test press, manufactured by the Nencki Ltd. company, delivered a force of up to 2000 kN for testing, corresponding to the cylinder pressure of 305 bar. SDI later modified the press by increasing the hydraulic pressure to deliver a maximum force of 3000 kN. This was achieved by installing a larger 13 inch bore cylinder. The frame had to be modified by adding structural supports so that it was able to withstand the increased force. Changes made to the frame with the increased hydraulic pressure produced a minimum factor of safety of 1.98. SDI was seeking to increase the hydraulic pressure to deliver a maximum force of 3426 kN, which is the maximum force the newer cylinder can deliver. As shown in Figure 2.1 below, when the new force was applied the factor of safety decreased to 1.62.

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Figure 2.1: Minimum factor of safety of original frame design with pressure increase The increase in hydraulic pressure due to the introduction of stronger rail materials required the frame to be modified a second time to be able to withstand the increased forces that it will endure. The brainstorming sessions yielded four plausible design alterations for the frame that was analyzed more in depth.

● Add supports to frame - Find crucial stress points and add necessary supports to increase strength or more efficiently distribute load in those areas

○ Advantages: Low-cost, ensures structural rigidity ○ Disadvantage: Parts may need added in areas that would be difficult to weld

unless the front or back large plate was removed and welded back in place after the smaller parts are added.

● Expand frame - Make an exoskeleton around the frame to increase rigidity ○ Advantage: Only components on the front and back faces of the machine would

need to be removed and placed onto the new plates that sandwich the machine. ○ Disadvantage: Making plates this large could be very expensive.

● Relocate cylinder - Move the cylinder to side to reduce the bending loads ○ Advantage: Stresses at critical cross sections can be reduced without an increase

of cross section areas or adding extra components ○ Disadvantage: Unsymmetrical structure, relocate the test bed as well

● Mechanical change - Use additional mechanism to reduce force magnitude ○ Advantage:Reduce direct forces on cylinder, distribute forces over structure ○ Disadvantage: Would require more in depth redesign of structure and

components

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2.2 - Hydraulic System Design

The hydraulic system that feeds the 13-inch hydraulic cylinder could deliver a maximum pressure of 350 bar. Unfortunately, it was not possible to increase the pressure in the original system to supply the required pressure of 400 bar. Only a few options were available to meet the new requirements.

● Force existing system - Attempt to run the original high pressure pump at the

required pressure and replace necessary hoses and fittings to accommodate the increased pressure

○ Advantage: No cost to modify ○ Disadvantage: Pump could fail to deliver required pressure and be

damaged ● Replace existing system entirely - Purchase an entirely new system built

specifically for the application with the ability to accept later changes ○ Advantage: New system guaranteed to supply required pressure and

possible future upgrades ○ Disadvantage: Highest cost compared to other options

Figure 2.2: Highlighted pump and motor rated at 315 bar.

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● Replace high pressure pump - replace the original pump with one that was

guaranteed to deliver the required pressure and replace necessary hoses and fittings to accommodate the increased pressure

○ Advantage: Reduces the cost of replacing the entire system by only purchasing a pump that will solve the pressure issue of the original system

○ Disadvantage: Pump may not be integratable into the original system (a more powerful pump may require a larger power input to be able to run at desired pressure)

2.3 - Energy Absorption Design

An important part of the overall design was the ability to absorb energy when a rail weld fails. The energy needs to be safely dissipated within the confines of the test machine. Wood was being used to absorb the energy in the original design, but a greater energy absorbing option was desired to reduce vibrations and prevent the broken rail pieces from causing any damage to the machine components. The wood was placed into 4.25 inches slots on either side of the machine and also placed underneath the rail on steel support bars. Any new material/design needed to fit into the same spots that the wood was originally inserted. With unlimited options available, a large selection was made and reduced to the following list. The success of the new material was to be based on price, durability, and machine protection. Price would be compared to the cost of the original material. Durability was based on how well the new material could resist splintering or fracturing, compared to wood when a rail specimen fails. The new material had to protect the machine from damage of fractured rail pieces. ● Different type of wood

○ Advantage: Other woods are more suitable for absorbing energy ○ Disadvantage: Cleanup will still be a concern

● Plastic ○ Advantage: Can be cut to fit required space, recyclable ○ Disadvantage: May break upon impact

● Corrugated Cardboard ○ Advantages: Cheap, biodegradable, recyclable, strong static loading capabilities ○ Disadvantage: No impact testing data

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Figure 2.4: Example of corrugated cardboard ● Plascore (Aluminum Honeycomb)

○ Advantage: High-energy absorption ○ Disadvantages: Limited runs and possible high cost

Figure 2.5: Example of aluminum honeycomb ● Sorbothane (Visco-elastic polymeric solid)

○ Advantage: Can be customized to application ○ Disadvantages: May not hold up to rigors of use

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Figure 2.6: Example of sorbothane sheet ● High density foam

○ Advantages: Made to size and cheap ○ Disadvantage: Durability may be an issue

● Shock absorbers ○ Advantage: Could take rigorous use at high pressures ○ Disadvantages: Costly and would require maintenance

Figure 2.7: Example of a shock absorber

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3. Conceptual Design Evaluation

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3.1 - Concept Evaluation Overview

Multiple conceptual designs were developed for each of the three subsystems and then were evaluated to determine which conceptual design was the best solution for each subsystem. The first step in this evaluation was creating a weighted decision matrix of important categories that applied to that specific subsystem. The categories that were selected are cost, aesthetics, performance, safety, reliability, durability, implementation, ergonomics, and repairability. The categories were compared to each other on the basis of importance and then a weighting factor was calculated for each category. Table 3.1 shows the resulting category weighted decision matrix for the frame design. The tables for the other subsystems are located in the appendix. Table 3.1: Category weighted decision matrix showing the total score and rank of each category for frame design concepts Cost Aesthetics Performance Safety Reliability

Cost X 1 9 9 9 Aesthetics 9 X 9 9 9

Performance 1 1 X 9 4 Safety 1 1 1 X 4

Reliability 1 1 4 4 X Totals: 12 4 23 31 26

Weight: 0.13 0.04 0.24 0.32 0.27 Next, the conceptual designs were placed into a table and graded by how well they fit each subsystem category. Each team member did these tables individually and once each team member completed the grading of the conceptual designs for each subsystem, the scores were averaged as can be seen in the following tables. Table 3.2 below shows an example of the grading of the frame conceptual designs. Table 3.2: Manuel’s grading matrix for frame conceptual designs

Attribute weights 12.50% 4.17% 23.96% 32.29% 27.08%

Frame Cost Best=4

Aesthetics Best =4

Performance Best = 4

Safety Best = 4

Reliability Best=4 Score

Add Supports to Frame 4 3 4 4 4 3.96

Expand Frame 3 4 3 3 3 3.04 Relocate Cylinder 1 1 1 1 1 1.00

Mechanical Change 2 2 2 2 2 2.00

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3.2 - Frame Concept Evaluation

The decision to increase the structural integrity of the frame through added supports was a unanimous one. All of the group members agreed that increasing the number of supports boosts the safety of the machine while keeping costs low. Table 3.3: Decision matrix for the frame concepts

Conceptual Designs Manuel Score Nick Score Cody Score Matt Score Average

Score Add Supports to Frame 3.96 4.00 3.73 3.72 3.85

Expand Frame 3.04 1.69 3.23 3.28 2.81 Relocate Cylinder 1.00 2.11 1.96 1.96 1.76

Mechanical Change 2.00 2.20 1.08 1.04 1.58 3.3 - Hydraulic System Concept Evaluation

Choosing to replace the entire system would be the most comprehensive solution and give the ability to foresee and address future design changes. Based on price, however, replacing the high pressure pump and necessary fittings may also be a good solution. Based on the evaluation, forcing the existing system was not a viable option. Table 3.4: Decision matrix for the hydraulic system concepts


Score Force Original System 1.00 1.50 1.15 1.15 1.20

Replace Original System Entirely 3.00 2.85 2.93 2.85 2.91

Replace High Pressure Pump 2.00 1.65 1.93 2.00 1.89 3.4 - Absorption Material Concept Evaluation

Many different energy absorbing concepts were generated during the brainstorming process. This large selection allowed for vast differences in conceptual design scoring. The final average showed Sorbothane to be the best solution for the rail bending test machine.

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Table 3.5: Decision matrix for the energy absorbing material conceptual designs


Score Different Type of Wood 1.88 3.74 2.74 2.88 2.81

Plastic 1.42 2.48 2.25 3.83 2.49 Corrugated Cardboard 3.21 2.79 3.39 4.71 3.52 Plascore (Aluminum

Honeycomb) 5.31 3.65 4.05 3.61 4.15

Sorbothane (Visco-Elastic

Polymeric Solid) 6.82 5.70 6.00 5.75 6.07

High Density Foam 4.94 5.91 5.53 4.99 5.34 Shock Absorbers 4.42 3.74 4.05 2.24 3.61

3.5 - Conceptual Design Evaluation

Using the results from the above tables, we arrived at the optimum design concept for each of the subsystems. The table below shows the top conceptual design for each subsystem as well as the best alternative design. Table 3.6: Conceptual design selection

Subsystem Primary Conceptual Design Secondary Conceptual Design

Frame Add Supports to Frame Expand Frame Hydraulic System Replace Original System Entirely Replace High Pressure

Pump

Energy Absorbing Material Sorbothane (Visco-Elastic Polymeric Solid) High Density Foam

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4. Detailed Design

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4.1 - Detailed Design Overview Based on the results of the design evaluation, detailed designs were selected and analyzed. Support members were constructed to optimize frame strength and increase the factor of safety. The resistance of Sorbothane to tearing during repetitive test runs remains a concern, and will need further evaluation. The decision of whether to replace the entire hydraulic system or simply upgrade the high pressure will be based on the outcome of Flodraulics analysis. 4.2 - Final Frame Design

The project’s requirements and specifications dictate that the frame must be able to withstand the higher force caused by the increase in pressure in the hydraulic system. This goal was achieved by increasing the thickness of the front face plate as well as adding supports to the rear faceplate. The faceplates undergo higher stress levels than the rest of the frame components according to Figure 2.1 and were therefore, the main frame design focus. Various other supports were added to the frame to increase structural rigidity. Another requirement was to use A36 steel for the parts added to the frame. The material properties of A36 steel may be found in Table 4.1.

Table 4.1: Mechanical properties of A36 steel Mechanical Properties Metric Imperial Tensile Strength, Ultimate 400-550 MPa 58000-79800 psi Tensile Strength, Yield 250 MPa 36300 psi Elongation at Break (in 200 mm) 20.0% 20.0% Elongation at Break (in 50 mm) 23.0% 23.0% Modulus of Elasticity 200 GPa 29000 ksi Bulk Modulus (typical for steel) 140 GPa 20300 ksi Poissons Ratio 0.260 0.260 4.2.1 - Front Faceplate The first step in arriving at the final design for the front faceplate was to find the principal stresses acting on the pre-existing faceplate due to the increase in pressure delivered by the hydraulic cylinder. Since the faceplate was being subjected to different types of stresses, the maximum stress has to be determined to ensure that the design meets the design requirements. The principal stresses (σ1,2) were calculated using the following equation:

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± σ1,2 = 2σ +σx y √( )2

σ −σx y 2+ τxy

2 (4.1)

Since there were no forces acting along the horizontal direction, the stress in the x-direction (σx) is 0. The normal stress and bending stress both act along the vertical direction, thus the stress in the y-direction (σy) is the summation of these stresses. The shear stress acts at both of the locations of the concentrated loads (τxy). Taking this into account, the inputs for equation one reduce to

; σ − ) ; τ σx = 0 y = ( ZP a* + Z

M xy = 3 V*2 A* rect

(4.2)

The following body diagram was used to find the moment and shear force the original front and rear faceplates were subjected to.

Figure 4.1: Faceplate body diagram

The bending moment and shear force diagrams and calculation results shown in Figures 4.2 and 4.3 were made using an online calculator found on the SaECaNet website. By inputting the information found in Figure 4.3 into the online calculator, the shear force and bending moment were easily obtained. These results are for the original frame faceplates with a thickness of 2 inches.

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Figure 4.2: Bending moment and shear force diagrams

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Figure 4.3: Free body diagram input values and calculation results

Now that bending moment and shear force are known, we use those values in equation two and use the results in equation one, as shown below.

− ) − 6, 33 psi σy = ( 370.8 in3385,098 lbf 14.73 in* + 370.8 in3

−4,277,103 lbf in* = 2 8 (4.3)

, 60 psi τxy = 2 66.7 in2*3 385098 lbf* = 8 6 (4.4)

± σ1,2 = 20−26,833 √( )2

0+26,833 2+ 8, 606 2

Where

; − 9, 85 psiσ1 = 2 3 , 52 psiσ2 = 2 5

(4.5)

(4.6)

The original faceplate was made from ASTM 1020 steel with a yield strength of 42,748 psi. From eq. (4.6) it can be concluded that the maximum principal stress was and that the factor σ1 of safety under the increased pressure was therefore,

OS .44F = Max StressY ield Strength = 29,385

42,748 = 1 (4.7)

As can be seen from eq. (4.7), the factor of safety for the 2 inch faceplate is well below the required value of 2 upon increasing the system pressure. This result is 11.1% off of the value calculated using SolidWorks Simulation as shown in Figure 2.1. This discrepancy in our calculations may be attributed to the simplification of the faceplate body diagram. Following the same procedure, but using a 3 inch thick faceplate made from the required ASTM A36 steel which, according to MatWeb, has a yield strength of 36,300 psi, the factor of safety resulted in

OS .83F = Max StressY ield Strength = 19,590

36,000 = 1 (4.8)

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Figure 4.4: FEA analysis of frame with front faceplate thickness of 3 inches

From this analysis one can conclude that adding one inch of A36 steel to the front faceplate will satisfy the design requirements of the frame. The difference between our calculation and the FEA analysis was that of 3.1%. Figure 4.4 shows that the new factor of safety for the front plate was 2.088 at it’s weakest location. 4.2.2 - Rear Faceplate and Inner Frame Supports After adding an additional inch of thickness to the front faceplate, the location of the maximum stress shifted to the rear faceplate as it is shown in Figure 4.5 below.

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Figure 4.5: Rear view of FEA analysis of frame with front faceplate thickness of 3 inches

For the rear faceplate, we were faced with two options, we could either increase the thickness or add supports to distribute the load more evenly and alleviate stress at critical locations. Making another faceplate would prove expensive, thus the option for adding supports was chosen. The rear faceplate supports can be seen in Figure 4.7. In addition to the rear faceplate supports, other supports were added to the frame between the two faceplates as shown in Figure 4.6. These supports are objects four, three, and five. After adding all the supports to the frame, FEA analysis was performed to ensure that the final assembly met the required goals for the factor of safety. The result was a minimum factor of safety of 2.22, which fully satisfies the main requirement for the frame design. Based on our previous calculation deviations from the software predicted values, we can safely assume that the design will be able to safely operate with the intended increase in pressure.

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Figure 4.6: Assembly of all added components that were added to the frame

Figure 4.7: Rear view of FEA analysis of final frame assembly.

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4.3 - Hydraulic System

The requirement of the hydraulic system was to safely provide 400 bar of pressure to the hydraulic cylinder. This was an increase of 50 bar from the original operating pressure of 350 bar. After brainstorming possible solutions to obtain the increased pressure we decided, by design matrices, to replace the entire hydraulic system. SDI recommended that we contact their preferred hydraulic company, Flodraulic, and seek assistance in upgrading the hydraulic system. Flodraulic informed us that it would not be necessary to replace the entire system. Instead, we will be replacing the high pressure side of the system, which was our second conceptual design. The first step to upgrading the hydraulic system was to evaluate the motor. The reason for this was to ensure that the required power to drive the new high-pressure pump was met by the motor. The power required to drive a pump operating at 400 bar while maintaining the original motor’s flow rate of 1.45295 GPM was calculated in eq.4.9, below.

(4.9)P PM×P SI×0.0007 .45295×5, 02×0.0007 .91≈6 horsepower or 4.4 kWH = G = 1 8 = 5 The original motor in the system was rated at 3.6 kW (4.83 hp), which was not enough to drive a pump operating at 400 bar. The maximum operating pressure the original motor can operate is 4,750 psi (328 bar). Working together with SDI’s prefered hydraulic consultant, a motor rated at 7.5 hp was chosen, which is capable of driving a pump operating at pressures up to 7,374 psi (508 bar). The motor chosen was the WW7.5-12-254TC. The high-pressure pump was the next component to be selected. The system required a new pump that was able to operate at the desired pressure while maintaining the output flow rate of the original pump of 1.18877 GPM. The DynexRivett PF-1003-10 checkball piston pump, which can deliver up to 1.45 GPM at up to 6000 psi rated working pressure, was chosen for this operation. Tables 4.2 and 4.3 show the specifications for the selected pump.

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Table 4.2: Selected pump specifications highlighted.

Output Flow at

1500 rpm Output Flow at 1800 rpm

Rated Pressure

Maximum Intermittent Pressure

Pump Models

U.S gpm

L/min

U.S gpm L/min psi bar psi bar

Rated Speed rpm

Maximum Speed

rpm PF1002-10 1.2 4.7 1.5 5.7 6000 420 8000 560 1800 3600 PF1003-10 1.7 6.6 2.1 7.9 6000 420 8000 560 1800 3600 PF1004-10 2.1 8.2 2.6 9.8 6000 420 8000 560 1800 3600 PF1005-10 2.9 11.0 3.5 13.2 6000 420 8000 560 1800 2800 Table 4.3: Selected pump output flow specifications highlighted.

Output Flow (At 1800 rpm, at rated pressure) 02 1.5 gpm (5.7 L/min) 03 2.1 gpm (7.9 L/min) 04 2.6 gpm (9.8 L/min) 05 3.5 gpm (13.2 L/min)

Now that a pump and motor were selected, the next component to upgrade was the C-face adapter, which mounts the pump to the motor. The selection of the adapter was made easy due to the web application found on the BSF, INC. website, which builds an adaptor for you based on the pump and motor of your system. The application also allows you to choose the coupling halves for the motor and the pump. Figure 4.8 shows a schematic of the C- face adapter and coupling halves assembly generated by the web application.

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Figure 4.8: C-face adapter and coupling halves assembly.

The following components that were selected were motor dampening bars. These bars help reduce the vibrations and noise produced while the motor is operating. From the Vescor catalog, the correct bars were selected based on the size of the motor and the bonding method used to attach the dampening bars to the motor. We opted for the weldable option based on our motor 254TC motor size. Figure 4.9 shows the dampening bars that were selected along with dimensions on table 4.4. Table 4.4:Weldable motor dampening bars schematic with table of dimensions.

Motor Dampening Bars - Weldable Part No. Motor Size L W H B C T1 (NC)

VSM-56-W 56C 4.75 2 1.56 3 0.88 5/16-18 VSM-143-W 143TC 5.5 2 1.56 4 0.81 5/16-18 VSM-145-W 145TC 6.5 2 1.56 5 0.81 5/16-18 VSM-182-W 182TC 6 2 1.56 4.5 0.94 3/8-16 VSM-184-W 184TC 7 2 1.56 5.5 0.94 3/8-16 VSM-213-W 213TC 7.75 2 1.56 5.5 0.94 3/8-16 VSM-215-W 215TC 9 2 1.56 7 0.94 3/8-16 VSM-254-W 254TC 10.5 2 1.75 8.25 0.94 1/2-13 VSM-256-W 256TC 12.25 2 1.75 10 0.94 1/2-13 VSM-284-W 284TC/TSC 12 2.75 2.38 9.5 1.12 1/2-13

Figure 4.9: Weldable motor dampening bars schematic with table of dimensions.

Counterbalance valves, or overcenter valves, are throttle valves which are commonly used for a well controlled movement of loads connected to double acting hydraulic consumers (lifting and swivel cylinders). They prevent a load from dropping in case of hose or tube failure, and from

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drifting caused by directional control valve spool leakage. They also provide smooth, modulated motion when the load is in a lowering or run-away mode, or modulated motion when the directional control valve is suddenly closed. For best results, these valves are placed as close to the consumer as possible. The counterbalance valve in the original system was not rated for the 400 bar of pressure that upgraded system will operate. Figure 4.10 shows a schematic of the counterbalance valve that was chosen to replace the original valve. It is a Hawe LHT 50 G-11 counterbalance valve designed for direct pipe connection. This valve will be able to operate at the original flow rate, and it is rated for 450 bar.

Figure 4.10: Schematic of LHT 50 G-11 counterbalance valve by Hawe.

The next items that were selected are: the high pressure directional control valve and subplate, in-line check valves for low and high pressure pumps, high pressure proportional pressure relief valve, and the reverse free flow check valves. The original components were not rated for the 400 bar pressure that was intended for the new system. The components were chosen such that they were able to handle the increase in pressure at the original flow rate. A list of all the components suggested for replacement may be found in appendix B along with pricing, part numbers, and specs for all the components. 4.4 - Energy Absorbing Material

As previously mentioned, the extreme force exerted when the rail fractures can cause major damage to the frame. To reduce the amount of damage, an energy absorbing material is used in specific areas of the machine where the most damage is caused; directly underneath the rail and on either side of the rail. With the space limitation, a material was needed that is compact enough to fit in that space and also capable of absorbing extreme amounts of force. The original material

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used is wood. Wood does not absorb the energy effectively and also splinters on impact leaving a mess that is difficult to clean up. The solution to this problem was Sorbothane. Sorbothane is a proprietary, viscoelastic urethane polymer used as a shock absorber and vibration damper. Sorbothane will replace the original material used, wood, to absorb the energy released upon rail fracture and serve as a means of protection for the frame. Sorbothane will fulfill the design variable for this corresponding subsystem if the material remains intact for multiple cycles. The wood has three designated places: the basin below the rail and two side slots positioned at the ends of the rail. Sorbothane is the only material discussed here since it was the primary selection. While sorbothane is a very good energy absorber, it is also very expensive. If the rail fractures during a test it would easily damage the material rendering it unable to function properly for the next test. To prevent this damage, a sandwich-style assembly was created for the basin below the rail using two 0.375 inch steel plates on the top and bottom of the Sorbothane. The bottom plate serves as a resting platform and the top plate serves as a damage prevention material for the Sorbothane. Generally, continually needing to replace the top plate of steel would not be cost effective, but since SDI is a steel manufacturer the plates can be recycled in their own facility. The same style assembly was used for the sides of the machine.

Figure 4.11: Shock response of Sorbothane compared to similar products

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Figure 4.11 shows the time delay effect of impulse response of selected materials, including Sorbothane. As can be seen, Sorbothane does the best job of dissipating the impulse and reaches steady-state the quickest. Figure 4.12 and Figure 4.13 show the dimensional layouts of the two sizes of Sorbothane quoted, one size for the side slots and another size for the basin beneath the rail. A sheet thickness of 1 inch is shown because this was the highest thickness offered by Sorbothane without needing custom thickness sheets created, which would dramatically increase the price. Instead, smaller thickness sheets were used and placed on top of each other to achieve the sheet thickness desired.

Figure 4.12: Dimensions of Sorbothane sheet for bottom basin

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Figure 4.13: Dimensions of Sorbothane sheet for side slots

4.4.1 - Calculations The calculations for this section were performed to determine whether or not the energy absorbing material will be able to withstand the energy released when specimen failure occurs. Using the dimensions of the space where the material will be placed we can calculate the area (A0) and the perimeter area (Ap). A shape factor can then be calculated with those areas; eq. (4.10). This would correspond to the shape factor and percent of free height given by the manufacturer. A deflection can then be calculated with the percent of free height and the thickness of the material; eq. (4.11). Next we can calculate how much force the material can handle; eq. (4.12). The capable force is used to determine the material's effectiveness, as well as the elastic potential energy; eq. (4.13).

Shape Factor = ApAo (4.10)

eflection percent of free height)×(thickness)D = (

(4.11)

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apable Force C = Lo(young s Modulus)×Ao× L′

(4.12)

lastic Potential EnergyE = 2×Lo(Y oungs Modulus)×Ao× L2

(4.13)

After performing these calculations it determined that the material that will be placed on the sides will be able to withstand an impact of 45,000 lbs, and have an elastic potential energy of 6,700 lb*in. This will not be enough energy absorbing capacity, and will need to add additional material. The steel plates mentioned previously will provide the additional capacity needed.

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5. Design Modifications

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5.1 - Overview of Changes As with every design project, alterations to the original design were made to address concerns that were not evident during the development stages. The frame design remained intact apart from a minor modification done to one of the top supports, which was necessary for cable management, and holes cut on the front plate to help with the welding process. The hydraulic system also received minor changes as some of the parts that were previously thought to need replacement were kept upon further inspection, and others that didn’t appear to need replacement were replaced. The energy absorbing materials that were chosen for our final design turned out to be too expensive to be used in the long run and were replaced in favor of corrugated cardboard. 5.2 - Frame Design Modifications Modifications were made to the plates originally proposed, and all modifications were done to aid in their installation. The front faceplate underwent the most significant changes, as can be seen from Figure 5.1. The overall dimensions of the plate were decreased to allow for better bonding area when welding the new faceplate to the original frame. Another change to the front faceplate was the addition of oblong holes cut into the plate, this again was done to increase the bonding area between the original frame and the new material. Finally, the top plates, which are denoted by bubble 3 in Figure 4.6, were cut diagonally at a 45 degree angle to allow for better access to the hydraulic cylinders ports. This change may be seen in Figure 5.2. These changes were then subjected to F.E.A. analysis using Ansys and Solidworks with a maximum force of 770,034 lbf or 400 bar of pressure. As seen in Figure 5.3, the factor of safety decreased to 2.05, which still satisfies the minimum requirements set in the problem statement. 5.3 - Hydraulic Design Modifications In addition to the components originally ordered to upgrade the high pressure side of the system, a high pressure transducer and a new hydraulic tank were added to the needed components list. Upon further inspection of the high pressure proportional pressure relief valve originally in the system, it was concluded that a replacement would not be necessary, which saved SDI $4,981. An updated HPU components list may be found in Appendix B.

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Figure 5.1: Front plate modifications

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Figure 5.2: Top plate modifications shown in SolidWorks

Figure 5.3: Minimum FOS of 2.05 after modifications

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5.4 - Energy Absorbing Design Modifications During the implementation stage of the project, the energy absorbing materials chosen as our top conceptual designs, sorbothane and high-density foam, were substituted with a composite design consisting of wood and corrugated cardboard. While sorbothane and high-density foam performed very well, both of these materials were easily damaged when not shielded by another medium. In addition, the cost of fabricating the material to the sizes specified in Figures 4.12 & 4.13 was no longer viable for our sponsor due to the amount of money that was spent on the HPU. Corrugated cardboard provided a comparable energy absorption alternative, and showed some advantages over the previously chosen materials, such as being biodegradable, recyclable, and low cost.

Figure 5.4: Corrugated cardboard configurations

The strength of the corrugated is dependent on its flute configuration, as well as flute density. Figure 5.4 shows various types of cardboard configurations. Corrugated cardboard is stronger when loaded vertically along the flutes due to the contribution of the flute medium. The team worked with a local box manufacturer, Kelly Box, for the development of cardboard “blocks” such as those seen in Figure 5.5 and Figure 5.6.

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Figure 5.5: Corrugated cardboard block dimensions

Figure 5.6: Finished corrugated cardboard blocks

The strength of the blocks was tested on the rail press before any modifications were done to it. A layer of cardboard blocks was placed on top of the original energy absorbing material, wood, and a few tests were conducted. The results of the tests showed that the corrugated cardboard blocks help surprisingly well when subjected to the impact forces of the broken rail pieces. The corrugated cardboard deformed under the impact of the rail, but it was not destroyed and fared well, as shown in Figure 5.7. The maximum deflection of the blocks was found to be 1.33 inches.

Figure 5.6: Corrugated cardboard blocks initial testing

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6. Design Testing & Results

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6.1 - Testing Preparation: Hydraulic System Calibration Succeeding the hydraulic system overhaul, the hydraulic system arrived at Power Components Corporation for re-plumbing and calibration. Power Components connected the hydraulic unit to a hydraulic cylinder in order to test the units capabilities. Calibration of the Hydraulic Power Unit Pressure Upgrade was to dead head the cylinder and watch the pressure gauge until it reached 400 bar (5,800 psi) and then secure with a lock down nut on the adjustment knob on Dynex Rivett High Pressure Relief Valve to maintain the set point. Further adjustments were then done at SDI once the system was connected to the company’s own hydraulic cylinder. SDI monitored and confirmed the final forces being applied to test and adjusted per the feedback from the pressure transducer.

Figure 6.1: HPU test setup

6.2 - Testing Procedure

The first step of the test procedure was loading the energy absorbing material into the frame. As can be seen from Figure 6.2, the frame was first loaded with the usual layer of wood, followed by a layer of corrugated cardboard blocks. The combination of cardboard and wood made for a tight finish, although there was still enough clearance for the rail to be maneuvered into place. After the energy absorbing material was in place, the rail was loaded onto the machine and the guard screen lowered. Figure 6.3 shows the completed setup ready for testing.

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a) b) c)

Figure 6.2: a) Layer of wood acting as a base for the b) corrugated cardboard blocks. c) The rail still has enough clearance for testing after cardboard blocks are added.

Figure 6.3: Machine loaded and ready to run tests (guard screen is up to show rail fitment)

Two types of rail were used for these tests: standard rail and high speed rail. The high speed rail has a hardened head, which previously made the rail unable to be tested on the machine. The machine was originally not able to deflect the rail enough to meet industry standards nor was it able to break the rail for failure studies. Since the machine is also used to study how rails fail, the goal of the tests was to run the machine until the rail samples failed. Figure 6.4 shows the result of testing the high speed rail.

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Figure 6.4: Machine after test was completed

6.3 - Frame Results During the tests, the frame was closely monitored for irregularities in the welds that held the new supports onto the original frame. No issues occurred when the machine ran at full capacity, so it was concluded that the strengthening members added to the interior and exterior of the rail bending test machine fulfilled the desired result. Although the HPU is rated to deliver 770,034 lbf, only 656,315 lbf were required to cause the high speed rail to fail. SDI decided to calibrate the HPU to this max pressure for their tests. As a result, the final factor of safety of the frame is 2.38 and this is shown in Figure 6.5.

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Figure 6.5: Minimum factor of safety at max force achieved during testing, 656,315 lbf

6.4 - Energy Absorption Results

Initial testing of the corrugated cardboard occurred in the original 328 bar max pressure system. The testing resulted in a successful dampening effect toward the broken rails and machine as a whole. This gave the group confidence to use corrugated cardboard in the final design. The corrugated cardboard was further tested at the machine’s updated HPU operating pressure of 400 bar. Even under these conditions, the corrugated cardboard proved effective in absorbing the impact force of the rail, and it retained exceptional structural integrity. Another benefit to using corrugated cardboard can be seen in Figure 6.6 & 6.8. The deformation of the cardboard helps limit the rail movement after rupture. This decreases the resulting damage to other areas of the machine by absorbing the energy. It also aids in the inspection of the fracture face of the broken rail since it is hitting the soft cardboard instead of the hard wood or steel frame from ricocheting. After inspection of the corrugated cardboard’s ability to absorb the energy of the high speed rail, Figure 6.7 shows slight damage to the base wood layer. This result shows the cardboard’s exceptional absorbing ability as this was an extreme case in which SDI ran the test until rail fracture. Normally, SDI would have stopped testing prior to rail fracture.

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Figure 6.6: Damage on the side corrugated cardboard blocks

Figure 6.7: Wood layer is barely damaged thanks to cardboard layer

Figure 6.8: Severely damaged corrugated cardboard block

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7. Cost Analysis

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The initial budget for this project was set at $15,000 by Steel Dynamics. However, after speaking with a sales representative from Flodraulic Group, it was determined that replacing only the high pressure side of the HPU would put us well above our budget, as can be seen from Table 7.2. This high cost for a crucial part of the project required cutting costs in other areas. As the material additions to the frame were required for safety, that left the energy absorbing material as the only cost cutting option. This resulted in the selection of corrugated cardboard as a low cost alternative to sorbothane and high-density foam. Despite its short life expectancy of one run, it was found that when designed modularly, few pieces needed replaced per run. This combined with the purchase in bulk would decrease the cost per run below the $75 shown in Table 7.3. Discussions with SDI’s representative Andrew Weber concluded with the project budget being given a flexible $20,000. Table 7.1: Cost analysis of the frame.

Item Number Description Price

1 Steel $1,091.00

2 Installation $6,126.41

Total Cost $7,217.41 Table 7.2: Cost analysis of the hydraulic system.

Item Number Description Price

1 Hydraulic Components $7,616.00

2 Pressure Transducer $455.45

3 Hydraulic Tank and Assembly $11,156.48

Total Cost $19,227.93 Table 7.3: Cost analysis of the energy absorbing material.

Item Price

Corrugated Cardboard $75.00

Total Cost $75.00

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Table 7.4: Total cost of project. Subsystem Description Cost

Frame Includes manufacturing parts and installation

$7,217.41

Hydraulic System Includes high pressure component replacement

$19,227.93

Energy Absorbing Includes material purchase, modification, and installation

$75.00

Total $26,520.34

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8. Conclusion

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8.1 - Recommendations

Upon completion of the project design and evaluation, some recommendations were given to SDI to improve functionality and aesthetics of the machine. The first recommendation is to incorporate channels into the frame that help secure the corrugated cardboard in place during testing. Adding channels would retain the cardboard during rail fracture and increase its effective absorbing abilities. The second recommendation is to repaint the frame. Due to the addition of supports to the frame, there remains bare metal that is not aesthetically pleasing. Painting the frame will also help protect the metal from corrosion and help maintain structural integrity in the long run. The final recommendation for SDI is to replace the hydraulic cylinder. If SDI ever wishes to raise the testing pressure closer to the cylinder’s max rating, the efficiency of the machine will suffer and the life of the cylinder will drastically decrease. This, again, would prove costly in the long run. However, we found that replacing the hydraulic cylinder would not be cheap as one was quoted by Flodraulic Group at a price of $40,815 (higher rated cylinders are more expensive).

Figure 8.1: Channels suggested for housing the corrugated cardboard blocks

8.2 - Project Summary

SDI, a manufacturer of steel products was in need of modifications to their rail bending test machine. The machine setup was unable to properly test a new high strength rail that is being manufactured. The company had to send their test samples elsewhere to be tested and as a result they were losing money and time. The task entrusted to us by SDI was to make necessary modifications to the machine to ensure proper testing of the new rail. A problem statement was formulated, which outlined requirements & specifications, design parameters, design variables, and limitations & constraints. These were formulated to ensure the machine was properly analyzed and fully understood. The problem statement also served as the project guidelines during the entirety of it. The project was divided into three subsystems: Design of Machine Frame, Hydraulic System Design, and Energy Absorption Design. These subsystems were identified to have the most importance on functionality, performance, and safety. Several designs

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were considered and evaluated for each subsystem until a top design was chosen for each. The project continued with the implementation phase in which the team worked with several top tier companies to deliver SDI a finished and working product. Revisions were made to the original designs chosen due to the cost to performance ratio. After the implementation of the revised designs, testing was performed to ensure fulfillment of the requirements set forth in the problem statement. The testing phase determined that all the systems fulfilled the specified requirements set by SDI and the formulated problem statement: frame is able to support a pressure of 400 bar with a minimum factor of safety of 2, the energy absorbing material successfully absorbed the rail’s impact forces, and the HPU can safely deliver 400 bar.

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References

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"141-lb/yd AREMA Rail | Harmer Steel Products Company." 141-lb/yd AREMA Rail | Harmer Steel Products Company. 1 Jan. 2014. Web. 25 Sept. 2014. <http://harmersteel.com/catalog/tee-rails/141-lbyd-arema-rail/>. "Adaptor C-Face." BSF Inc.N.p., n.d. Web. "ASTM A36 Mild/Low Carbon Steel." ASTM A36 Mild/Low Carbon Steel. N.p., n.d. Web. Atkinson, Samuel, and Randy L. Bowman. The Millennium Rail 141 AB (n.d.): n. pag. Arema.org. Web. Budynas, Richard G., J K. Nisbett, and Joseph E. Shigley. Shigley's mechanical engineering design. New York, NY: McGraw-Hill Education, 2014. Print. "Calculation of Bending Moment,ShearForce,Amount of Deflection,Angle of Inclination Slope. Single Span Beam-Beams with Both End Clamped-Two Concentrated Load at Arbitrary Position." saecanet.com.N.p., n.d. Web. 04 Dec. 2014. ‘CVH’ Check Valves (n.d.): n. pag. Lifco Hydraulics. Web. "Dampening Bars." LDI Industries. LDI Industries, n.d. Web. 04 Dec. 2014. Dynex HP03 Pattern Directional Control Valves (n.d.): n. pag. Web. "Dynex Hydraulic Pressure Control Valves." Dynex Hydraulic Pressure Control Valves.N.p., n.d. Web. Dynexpespf1000specs.pdf (n.d.): n. pag. Web. Over-center Valves Type LHT (n.d.): n. pag. Web. FCHH Flow Controls (n.d.): n. pag. Web. "Fluid Power Formulas." Fluid Power Formulas.N.p., n.d. Web. Fox, Robert W., Philip J. Pritchard, and Alan T. McDonald. Fox and McDonald's introduction to fluid mechanics. Hoboken, NJ: John Wiley & Sons, Inc, 2011. Print. Moaveni, Saeed. Finite element analysis: theory and application with ANSYS. Upper Saddle River, N.J: Pearson Prentice Hall, 2008. Print. Riley, William F., Leroy D. Sturges, and Don H. Morris. Mechanics of materials. Hoboken, N.J: John Wiley, 2007. Print.

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"Sorbothane Technical Data." Sorbothane.com. N.p., n.d. Web. Thomson, William T., and Marie D. Dahleh. Theory of vibration with applications. Upper Saddle River, N.J: Prentice Hall, 1998. Print. Voland, Gerard. Engineering by design. Upper Saddle River, N.J: Pearson/Prentice Hall, 2004. Print. "WW7.5-12-254T." WO - DATA SHEETS.xls (n.d.): n. pag. Web.

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Appendix A: Tables & Figures

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Table A.1: Category weighted decision matrix showing the total score and rank of each category

for hydraulic system design concepts Cost Performance Safety Reliability Repairability

Cost X 9 9 9 4 Performance 1 X 4 4 1

Safety 1 4 X 1 1 Reliability 1 4 9 X 1

Repairability 4 9 9 9 X Totals: 7 26 31 23 7

Weight: 0.07 0.28 0.33 0.24 0.07 Table A.2: Category weighted decision matrix showing the total score and rank of each category

for the energy absorbing material design concepts Cost Implementation Function Ergonomics Safety Durability Availability

Cost X 9 9 9 9 9 9 Implementation 1 X 4 9 9 9 1 Performance 1 4 X 9 9 9 4 Ergonomics 1 1 1 X 9 9 9

Safety 1 1 1 1 X 1 1 Durability 1 1 1 1 9 X 4 Availability 1 9 4 1 9 4 X

Totals: 6 25 20 30 54 41 28 Weight: 0.03 0.12 0.10 0.15 0.26 0.20 0.14

Table A.3: Nick’s grading matrix for frame conceptual designs


Frame Cost Best = 4

Aesthetics Best = 4


Safety Best = 4

Reliability Best = 4 Score




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Table A.4: Cody’s grading matrix for frame conceptual designs


Frame Cost Best = 4

Aesthetics Best = 4


Safety Best = 4




Mechanical Change 1 3 1 1 1 1.08 Table A.5: Matt’s grading matrix for frame conceptual designs


Frame Cost Best = 4

Aesthetics Best=4

Performance Best=4

Safety Best=4





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Table A.6: Manny’s grading matrix for hydraulic system designs

Table A.7: Nick’s grading matrix for hydraulic system conceptual designs

Table A.8: Cody’s grading matrix for hydraulic system conceptual designs

Table A.9: Matt’s grading matrix for hydraulic system conceptual designs

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Table A.10: Manny’s grading matrix for energy absorbing material conceptual designs

Table A.11: Nick’s grading matrix for energy absorbing material conceptual designs

Table A.12: Cody’s grading matrix for energy absorbing material conceptual designs

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Table A.13: Matt’s grading matrix for energy absorbing material conceptual designs

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Appendix B: HPU Component List

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1. (Print Position #25) 1 pc. Worldwide Electric Motor pt.#HHI7.5-12-254TC, Hundai Heavy Industries Electric Motor, 7.5 HP, 1200 RPM, 230/460 Volt/3 Phase/60 Hertz, TEFC, 254TC Framed, C-faced and Foot Mounted, Delivery: 1-2 Weeks ARO Net Price Each: $1,304.00

2. (Print Position #13) 1 pc. DynexRivett pt.#PF1003-10, Checkball Piston Pump to

deliver up to 5.5 l/min., at up to 6000 PSI rated working pressure, Delivery: 12-14 Weeks ARO Net Price Each: $1,944.00

3. (Print Position #12) 1 pc. BSF, INC., pt.#PD1-B105-A2-R01, Pump/Motor C-face

Adapter, Delivery: 3-4 Weeks ARO Net Price Each: $464.00

4. (Print Position #11) 1 pc. Rotex, Inc., pt.#RT3801ST 1-5/8" x 3/8" Keyed, Electric

Motor Coupling Half, Delivery: 3-4 Weeks ARO Net Price Each: $150.00

5. (Print Position #11) 1 pc. Rotex, Inc., pt.#RT3801ST 7/8" x 1/4" Keyed, Pump Coupling

Half, Delivery: 3-4 Weeks ARO Net Price Each: $150.00

6. (Print Position #11) 1 pc. Rotex, Inc., pt.#RT38/98 'Purple' Rubber Spider Insert,

Delivery: 3-4 Weeks ARO Net Price Each: $24.00

7. (No Print Position) 2 pcs. Vescor pt.#VSM-254-W, Electric Motor Vibration

Dampening Bars, Delivery: 2-3 Weeks ARO Net Price Each: $150.00 x 2 pcs. = $300.00 For Set

8. (Print Position #16) 1 pc. DMIC, INC., pt.#CVH050750B, In-Line Check Valve, #12

(3/4") BSPP Female Ported, 10,000 PSI Rated Working Pressure, for Low Pressure Pump Kit, Delivery: 1-2 Weeks ARO

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Net Price Each: $70.00

9. (Print Position #17) 1 pc. DMIC, INC., pt.#CVH050500B, In-Line Check Valve, #8 (1/2") BSPP Female Ported, 10,000 PSI Rated Working Pressure, for High Pressure Pump Kit, Delivery: 1-2 Weeks ARO Net Price Each: $60.00

10. (Print Position #29) 1 pc. DynexRivett pt.#VR6110-31-00, Manually Adjustable Relief

Valve, Delivery: 10-12 Weeks ARO Net Price Each: $660.00

11. (Print Position #19). 1 pc. DynexRivett pt.#Z-5052, High Pressure Proportional Pressure

Relief Valve, Rated Flow: 15 GPM Working Pressure: 15,000 PSI, 24 VDC, Delivery: 12-14 Weeks ARO Net Price Each: $4,981.00

12. (Print Position #22) 1 pc. DynexRivett pt.#6553-HP03-24/HDC-20, High Pressure

Directional Control Valve, Solenoid Operated, 4-Way, 3-Position, Spring Centered, P Port Blocked and A and B Ports Connected to Tank Port in Center Position, 24 VDC Solenoid Voltage, Rated Flow: 19 LPM (5 GPM), Rated Pressure: 10,000 PSI, Delivery: 12-14 Weeks ARO Net Price Each: $960.00

13. (Print Position #20) 1 pc. DynexRivett pt.#PSO-29-BSP-6, HP03 Footprint, Subplate,

Side Ported, #6 (3/8") BSPP Ports, Delivery: 2-3 Weeks ARO Net Price Each: $204.00

14. ( No Print Position) 1 pc. DynexRivett pt.#P11-BK, Bolt Kit, (4) 1/4" - 20 UNC

Threaded x .75" Long, Delivery: 1-2 Weeks ARO Net Price For Kit: $30.00

15. (Print Position #21) 2 pcs. DMIC, INC., pt.#FCHH1000B1141, In-Line Style, Flow

Controls with Reverse Free Flow Check Valves with 5 PSI Cracking Pressure Springs, #16 (1") BSPP Ports, 10,000 PSI Operating Pressure, Delivery: 2-3 Weeks ARO Net Price Each: $353.00 x 2 pcs. = $706.00

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16. (Print Position #24) 1 pc. Hawe pt.#LHT50-G-11-7-E6-100, In-Line Mounting, Counterbalance Valve, Rated For: 450 BAR (6,525 PSI), G1" (1" BSPP) Ports, Flow Rating: 50 LPM (13.21 GPM), 1:6 Ratio For Load Release,

17. 1 pc. AST20HAC00500B5Y0H, AST20HAC00500B5Y0H000, 500 Bar, 0-10VDC

Output, 1/4" MBSPP, M12 Connector Net Price Each: $395.00

18. 1 pc. DR34015P0 Cable, 15 Ft. with 90 degree exit D34 connector to leads. Net Price Each: $49.60

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Figure B1: HPU Schematic 73| 73

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