MD2 JEDR Assignment PDF

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JEDR Assignment MD2...

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48650 Mechanical Design 2

Junior Engineer Design Report Critique and Redesign

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Table of Contents Introduction ...................................................................................................................................... 3 1. Critique of Junior Design Engineer Report ...................................................................................... 3 1.1 Specifications ........................................................................................................................... 3 1.2 Design Configuration ............................................................................................................... 4 1.3 Shaft Speed, Torque and Gear Ratios ....................................................................................... 5 1.4 Pulley Belt System.................................................................................................................... 5 1.5 Shaft Layout ............................................................................................................................. 6 1.6 Bearing Specification................................................................................................................ 7 1.7 Idler Pulley and Spring ............................................................................................................. 8 1.8 C-Channel Bar .......................................................................................................................... 9 1.9 Drum Brake............................................................................................................................ 10 1.10 Optimisation ........................................................................................................................ 10 2. Redesign...................................................................................................................................... 11 2.1 Gears and Hub Selection ........................................................................................................ 12 Shaft Speed ............................................................................................................................. 12 Gear Selection ......................................................................................................................... 12 Torque..................................................................................................................................... 14 Gear Force Analysis.................................................................................................................. 14 2.2 Pulley and Belt System ........................................................................................................... 15 2.3 Shaft A Design........................................................................................................................ 18 Estimate Weight of Pulley A ..................................................................................................... 18 Critical Bending Moment ......................................................................................................... 21 Minimum Shaft Diameter ........................................................................................................ 21 2.4 Bearings Selection.................................................................................................................. 21 2.5 Optimisation .......................................................................................................................... 22 Cylindrical bearings.................................................................................................................. 22 Ball bearings ............................................................................................................................ 23 Appendix ......................................................................................................................................... 23

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Introduction This report has been commissioned to analyse the existing Junior Engineer Design Report (JEDR) as initiated by a previous intern, who had left at the conclusion of their internship. To reiterate, the objective of this JEDR is to present the testing of the life expectancy of different compounds used for drum brake shoes and to compare the frictional component of different compounds of belt materials. Official testing procedures are scheduled to take place in six months’ time at time of writing. Three months have been allocated to the redesign process and three months for ordering/supplying parts and setup of the test rig. The first major deliverable of this report is a critique of the existing JEDR, containing points of criticism and suggested improvements for distinct design elements. The second major deliverable relate to the redesign of the test rig due to revised specifications. This report will also detail optimisation of relevant parameters in the previous design and revised design.

1. Critique of Junior Design Engineer Report 1.1 Specifications •

This statement assumes that the entire wheel, including the tyre, to be 406mm (16 inches) in diameter. For most modern-day vehicles, the rim size alone would be 16 to 17 inches in diameter to accommodate for improved safety technology such as larger brake calipers and rotors. As a result, a 482.6mm (19 inches) wheel, including tyre, would be a more reasonable assumption.

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In the above assumption, the intern has made a mathematical error. When taking the original 406mm diameter wheel into account, 60km/h does not equate to 835RPM as evident in the calculations as follows: 60km/h = 16.67m/s 406mm diameter = 0.406mm diameter = 0.203 radius

ω= RPM = ω ×

60 2𝜋

𝑉 → 𝑟

ω=

→ RPM = 82.12 ×

60 2𝜋

16.67 0.203

→ ω = 82.12 rad/s

→ 784 RPM (rounded to nearest whole number)

Additionally, it is recommended to test the system at a different range of speeds to better simulate real-world conditions such as city vs. highway driving or under emergency situations. If this isn’t possible, then testing at higher speeds is recommended to help determine the limits of the drum brake assembly.

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The intern has identified the testing period as shown in the above statement. The lack of specificity in this statement may pose significant design challenges as there is no mention of how many times the system experiences heavy shock due to operation or if the system is allowed to cool down in between each repetition. As such, it may be difficult to design for and select the correct parts and materials to ensure system reliability at a reasonable cost.

1.2 Design Configuration The intern has provided an initial design of the test rig as shown here:

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The placement of the motor, as illustrated to be situated between both mounts for shaft A, is inefficient and unnecessary. By shifting the motor to the side and just adjacent to shaft A, a shorter shaft can be selected thus reducing cost. Furthermore, a shorter shaft means the mounts and bearings can be moved closer together to make the system more compact. The proposed design doesn’t allow for easy access to the belt and pulleys, which can complicate relevant maintenance procedures. It is strongly suggested that the pulleys be moved to outside the shaft mounts and bearings, allowing users to easily replace the belt and pulleys.

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The figure doesn’t detail how the drum brake assembly is mounted to the system. During normal operation the drum is expected to rotate along with shaft B, whilst the brake shoes are secured to prevent rotation. As this is not reflected properly in the figure, it may create issues with setup of the test rig. Additionally, translation of the drum brake assembly means a shorter shaft B can be selected to reduce cost.

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The figure doesn’t provide a coordinate system, which can make it difficult to identify or locate various machine elements. The addition of an alphanumeric grid would be sufficient for most readers.

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1.3 Shaft Speed, Torque and Gear Ratios •

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The intern has stated that “The motor will spin at 1800rev/min”. However, as specified in the corresponding data sheet, the motor will only spin at 1784 RPM when operated at 4/4 (full) load. The motor is assumed to be operating at full load torque as given in chapter 2. Since the intern has used 1800 RPM for their calculations, recalculations should be carried out in order to obtain accurate values for shaft speed, torque, and gear ratios. The intern has calculated the final speed of shaft B as being 57.6km/h and has justified their selection of gears through the statement “This speed is still close enough to the original 60km/h so the gears are still acceptable”. However, this isn’t a viable practice. For testing purposes, the intern should have selected a gear that can handle 60km/h or greater than 60km/h to account for a factor of safety. Doing so will prevent premature failure due to gears not capable of handling the required system demands. The intern had concluded that TS590 from the Martin gear catalogue would be the most appropriate to use for this system, based on their gear teeth calculations. However, having assumed in chapter 2 that 75kW (100HP) is transmitted to shaft B, there are no gears in the entirety of the catalogue that can handle 100HP at the given 1800 RPM as shown in the horsepower rating charts. Therefore, it is advised that a new motor with a lower power output be selected as the current motor is unsuitable for this system.

1.4 Pulley Belt System

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In the above figure, the intern has specified the smaller pulley (Pulley A) as 340mm in diameter and the larger pulley (Pulley B) as 1020mm in diameter. However, on the following page the intern states pulley A as 110mm in diameter and pulley B as 330mm in diameter and uses the latter set of measurements in subsequent calculations.

This discrepancy may cause readers to question the consistency of the JEDR. Additionally, it cannot be determined which set of pulleys the intern intended to use and can result in selection of incompatible pulleys.

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When calculating torque for the belt, the intern states that the service factor (Ks) is equal to 1.4 in reference to table 17.15. However, because this table hasn’t been provided the validity of the supplied information and calculations should be questioned. When calculating the contact angles for the pulley belt the intern utilised trigonometry, as seen in the following figure, which isn’t the correct method for this application.

This assumes that the belt passes through the centre of the pulley which in this case isn’t true. The correct formulae are as follows:

1.5 Shaft Layout

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As mentioned previously, moving the motor so it’s situated just outside the test rig will allow for a shorter shaft A to be selected. As seen in the above figure, it is possible to shorten shaft A by approximately 1.53m, which will not only increase the bending stress the shaft can withstand but also cut costs on material.

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When the intern was calculating the minimum shaft diameters, they opted to use the DEGoodman criterion thus neglecting Australian standards.

Seeing as how we’re based in Australia, it is critical that the Australian standard for

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minimum shaft diameters is used to ensure consistency amongst reports and that readers can easily understand what is presented. The formulae that should’ve been used are as follows:

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The intern has not included any information regarding the tolerances for the shafts. At points along the shafts where the gears, bearings and pulleys are positioned, tolerances are critical to ensuring these elements fit properly to the shaft. When calculating the factor of safety for the shaft, the intern has made a mathematical error.

The first figure shown above displays the equation to calculating the factor of safety for shafts. As evident in the working out, the intern has instead computed the alternating stress and midrange stress as the denominator and the endurance strength (Se) and ultimate strength (Sut) as the numerator. The Se and Sut have already been computed to be 533 MPa and 848 MPa respectively in preceding pages.

1.6 Bearing Specification

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As shown in the figure above, the intern has estimated the bore and width sizes of the bearings. However, this isn’t necessary as these measurements are always available on the manufacturer’s website via the data sheets for each specific bearing.

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After having decided that ball bearings were not suitable, the intern opted to use cylindrical bearings and had determined that these bearings would experience a load of 129.4kN. The cylindrical bearing that the intern decided upon, based on bore and width sizes, is rated at 129kN. However, despite the intern’s justification for their decision, this isn’t an acceptable practice as the bearings would be subjected to excessive loads and can result in premature failure. In this situation, the intern should have selected a bearing with a greater load rating to ensure system reliability.

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In the above statement, the intern had decided that bearings with a design life of 5000 hours should be used, as based on the projected life span of the test rig. However, in chapter 2 the intern had stated that testing is proposed to occur for 6 months, 5 days a week, 6 hours a day. If we assume there are 4 weeks in a month, this equates to 720 hours for the specified testing period. The intern also states that the test rig will be in use for the next 3 years. If the testing period remains constant for each of the next three years, this equates to a total of 2160 hours which is significantly less than the aforementioned 5000 hours. If this difference is to account for a factor of safety, 5000 hours is a gross overestimation and will only contribute to greater material costs.

1.7 Idler Pulley and Spring

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As seen in the above figure, the chosen setup of the idler pulley will reduce the contact angle of the belt on the pulley. A reduced contact angle lowers efficiency as it decreases the amount of torque that is transmitted through the belt. It is recommended that the idler pulley is situated below the belt so it pushes the belt upwards, with a compression spring used instead of an extension spring to maintain tension.

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The intern has again used trigonometry to calculate the contact angle of the belts on the pulleys. This method is incorrect as it assumes that the belt will contact the centre of each pulley on both shafts A and B.

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The intern had decided in their JEDR to utilise 4 extension springs in parallel to maintain tension and to support forces exerted on the idler pulley. However, this design may be difficult and overly complicated to implement especially if the springs are solely supported by the cantilevered C-channel bar. Thus a redesign is recommended; either incorporating a compression spring instead or the open belt system is replaced with a cross belt system that eliminates the need for an idler pulley and tensioner spring.

1.8 C-Channel Bar •

The cantilever design of the C-channel bar doesn’t yield any significant benefits over a beam that is fixed and supported by both mountings for shaft A and shaft B. Moreover, a bar fixed at both ends will increase rigidity of the test rig and minimise risk of bending due to excessive deflection.

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The length of the C-channel bar is given as 1.5m. However, 1.5m is given as the centre distance between pulley A and pulley B, also the distance between shaft A and shaft B. If the intern had opted for a cantilever design, it wouldn’t be necessary for the beam to extend beyond the positioning of the tensioner spring. Additionally, at 1.5m long the beam could be fixed to both shaft mountings for improved rigidity.

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When calculating for the shear stress at point C, the intern has made a mathematical error. They had substituted in 40mm (0.04m) as opposed to the 4mm (0.004m) used in the previous two calculations. The beam thickness of the C-channel bar has previously been established as 4mm in the intern’s drawings.

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1.9 Drum Brake •

In both chapter 9 and the original design configuration, the intern has not specified how the drum brake assembly will be fitted to shaft B. As mentioned before, the drum should rotate with the shaft, whilst the brake shoes remain stationary. If the brake shoes rotated in conjunction with the drum, the braking function of the overall system would be negated.

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In the parameters for the drum brake assembly, c has been specified to be 165mm. However, c is not reflected anywhere in the corresponding diagram. This means that there is no clear indication as to what c represents. Also in the parameters, b has been given as 75mm into the page. However, b is not reflected anywhere in the corresponding diagram nor has a side profile of the drum brake assembly been provided. Whilst readers can assume b to be the thickness of some component, it’s not clear which component this measurement would pertain to.

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1.10 Optimisation •

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If the motor and drum brake assembly are repositioned, both shafts can be shortened which will contribute to material cost savings. Optimisation in this regard will require calculating for the viably shortest shaft without needing a substantial increase in minimum shaft diameter to cope with bending moments and torques. The idler pulley and tensioner spring setup can be optimised through a redesign of both components. Moving the idler pulley under the belt so it pushes up against the belt will increase the contact angle and improve torque transmission through the belt. Also, replacing the extension springs with a compression spring will simplify the design. Increasing the width of the belt can increase the amount of torque that is transmitted through the belt.

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Brake shoes attached to shaft mounting to prevent rotation Drum allowed to rotate freely in conjunction with shaft

2. Redesign

Parallel pulley system to reduce size of pulleys

Shaft B

Gears Motor moved outside of shaft A

Pulley and belt system moved to outside of shaft mountings to ease maintenance Shaft A

Open belt system replaced with cross belt system

The above figure reflects the proposed redesign of the intern’s original configuration. Primary key differences include: • •

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Motor moved to outside of shaft A. This allows for a shorter shaft to be selected. Original pulley and belt system moved to outside of shaft mountings to ease maintenance. The open belt system has been replaced with a cross belt system, negating the need for a tensioner. This setup has been supplemented with a parallel system to reduce the size of the pulleys and to distribute load more evenly. Brake shoes for the drum brake assembly are now attached to one of the shaft B mountings to prevent rotation, enabling the system...