drive pin analysis

7/24/2019 Drive Pin Analysis

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Drive Pin Analysis

Nicholas Taylor

Tuesday 12 th May 2015

MEC3098


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h l

Declaration

This Report is submitted as part of the requirements for the Degree of Mechanical

Engineering (MEng Honours) at the University of Newcastle upon Tyne and has not

been submitted for any other degree at this or any other University. It is solely the

work of Nicholas Taylor except where acknowledged in the text or the

Acknowledgements below. It describes work carried out at the University of Newcastle upon Tyne which is entirely recorded in a Project Logbook which has

been made available for examination. I am aware of the penalties for plagiarism,

fabrication and unacknowledged syndication and declare that this Report is free of

these.

Signed:

Date:

Acknowledgements


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Abstract

A current 160mm Test Rig is utilised to evaluate contact fatigue testing using power

recirculating gear boxes. Gears are mounted on tapered shafts using a high

interference fit to transfer up to 6000 Nm of torque around the system. This poses the

risk of adding to stress from gear loading resulting in reduced gear capacity.

The project is to evaluate and analyse the use of pins to transfer the torque through

the system.

The optimum number of pins and size were found based on mechanical properties of

the chosen material. Interference fit and differing hole edge profiles were analysed to

provide the best reduction in stresses.

A gear blank was manufactured where sets of holes were drilled by plunge slot

drilling and interpolating. These methods were measured and compared for accuracy.

From this pin contact was evaluated to better understand the performance based on

hole drilling methods.


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Table of Contents

1 Introduction ..................................................................................... 1

2 Material Specification ..................................................................... 2

3 Pin Design ....................................................................................... 4

4 Calculations ..................................................................................... 5

4.1 Initial pin calculations ............................................................... 6

4.2 Results forming initial pin choice ............................................. 8

4.3 Interference fit calculations .................................................... 10

4.4 FEA calculations validation .................................................... 11

4.4.1 Initial calculations correlation .............................................. 12

4.4.2 Interference fit calculations correlation ................................ 14

4.5 Pin separation due to bending stress ....................................... 16

4.6 Hub hole edge profile ............................................................. 17

5 D i f bl k 18


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List of Figures

Figure 1. Gear mounting on 160 Test Rig. .................................................................. 1

Figure 2. Diagram showing differing modes of bending for pin ................................ 4

Figure 3 . Pin design specification ............................................................................... 4

Figure 4. Stress Convention showing orbiting element. ............................................. 5

Figure 5. Chart showing calculated bending, shear and maximum equivalent von

Mises stresses for varying critical angle ................................................................... 7

Figure 6 . Socket joint under transverse load with linear distribution of pressure ..... 10

Figure 7 . Visualisation of initial setup used in ANSYS. ........................................... 11

Figure 8. Coordinate systems used for analysis ........................................................ 12

Figure 9 . Gap size for s6max interference fit with force applied on pin................... 16
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List of Tables

Table 1. Mechanical Properties for chosen materials. ................................................. 3

Table 2. Given values for empirical calculations. ....................................................... 5

Table 3. Worksheet showing optimum Pin specification. ........................................... 9

Table 4 . Final variable specification chosen ............................................................... 9

Table 5 . Values of the fundamental deviations for shafts k to zc.............................. 10

Table 6. Values for tolerances ................................................................................... 11

Table 7. Pressure for considered interference fit ranges ........................................... 11

Table 8. Mesh element sizing for correlation iterations ............................................ 13

Table 9. Correlation results from the seven iterations within ANSYS ..................... 14

Table 10 . Table of pressure values comparing ANSYS and calculated values for
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1 Introduction

Within the Design Unit, a 160mm Test Rig is utilised to evaluate contact fatigue

testing using power recirculating gear boxes. As part of the design, gears are

mounted on tapered shafts as shown in Figure 1 using a high interference fit to

transfer up to 6000 Nm of torque around the system. By using such a high

interference for pull -up of the gears onto the shafts , induces internal hoop stressinto each gear. This poses the risk of adding to stress from gear loading resulting in

reduced gear capacity.

This project sets out to evaluate and design

the use of pins to transfer the required torque

around the system. By using pins, this would

allow using less pull -up when mounting the

gears on the shaft, thus reducing the hoop

stress within the gears ensuring the results

from the fatigue tests are more accurate. The

i ld b i i l f h i d i i f ill

Figure 1. Gear mounting on 160 Test Rig.


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2 Material Specification

Ensuring the correct material is used in the manufacturing of the pins is critical to

how they will perform. Using a sensible and readily available material to match the

specification will allow optimum performance and longevity of the pins. Metals give

the required properties to work within the test rig set up.

They type of metal needs to have the ability to be surface hardened and provide ahigh hardness and yield strength. The pins actual profile will be key to optimising

performance and thus a metal with good machinability is also important.

By using CES EduPack (Granta, 2014) it allows to input parameters which narrows

down a database to show materials which match. By using the below limits within

Level 2:

Hardness: Minimum of 208 HV

Yield Strength: Minimum of 900 MPa

Youngs Modulus: Minimum of 200 GPa

M hi bili Mi i f 3 5 (1 P d d 5 E ll


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S156 steel is a common material used in gear manufacturing in the Automotive and

Aerospace industries and has capability of nitride hardening.

The chosen material for the hub was 15NiCr13 (Also known as En36A) the

mechanical properties were attained from Bhler Uddeholm (En36A Case hardening

steel datasheet, 2015).

The mechanical properties for the materials are detailed below in Table 1.

Table 1. Mechanical Properties for chosen materials.

Mechanical properties34CrNiMo6 15NiMoCr13

S156 CAPi(En24) (En36)

R e min - Yield Strength (MPa) 1000 785 1030

R m - Tensile Strength (MPa) 1400 1280 1520

A - Elongation at fracture (%) 9 8 11

Z - Reduction in cross section at

f (%)40 35 40


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3 Pin Design

The design of the pin to be used is critical to

successful operation of the Test Rig. The pin

has to have a sufficient stiffness and strength

to withstand the stresses involved, but also

the ductility to yield elastically to ensure allthe pins contact and share the load.

In a real world application, it would be extremely difficult to ensure all the variability

was controlled to an extent to ensure each pin shared the load equally. A good pin

design must incorporate sufficient flexibility to account for tolerances in

manufacturing of the respective parts.

Usually the simplest of designs is the best but not always and this case is one. Having

a simple cylindrical pin of the required length and diameter is a nice simple solution,

but actually adds complexity to the design. A cylindrical pin will have a double point

of bending as shown in Figure 2 and would result in much more complex contact

resulting in complicated calculations and analysis.

Figure 2. Diagram showing differing modes of bending for pin


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4 Calculations

The specification for the design concepts were derived from empirical calculation

largely based on formulae working towards calculating the von Mises mean stress for

the pins (American National Standards Institution/ American Gear Manufacturers

Association, 2008).

Initial assumptions made; there is no torque on the pin; there is no axial force andthere is no hoops stress. The pins will only undergo bending and shear forces during

normal operation. The shafts will not induce any alternating stress through

misalignment and vibration. Lastly the pins will be an idealised solid cylinder for

purposes of the calculations.

In Table 2 are the given values from the project brief used for the calculations.

Table 2. Given values for empirical calculations.

Property Value Unit

Torque acting on gears 6000 Nm


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4.1 Initial pin calculations

To find the bending stress:

=32 (1)

where F is the tangential force, l is the bending arm length, d she is the outer pin

diameter and N is the number of pins.

To find the shear stress:

=4 1+21+

(2)

where V is the transverse shear force and where is the Poissons ratio of

34CrNiMo6.

As this load case represents a simplified case then the von Mises stress has to be

l d f d ff l f f d h h l l


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In finding the critical angle, arbitrary values for, the number of pins, pin diameter

and the bending arm length were chosen to allow calculations to be made to find the

stresses. From these, the maximum equivalent von Mises stress can be found.

Arbitrary values used:

Number of pins: 4, Pin diameter: 15mm and Bending Arm length: 15mm

From Figure 5 it can be shown that the bending stress is maximum at = 0 and the

shear stress is maximum at when = . As the bending stress is by far the dominant

stress, then the maximum equivalent v on Mises stress occurs when = 0.


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This will give a deflection based on each pin sharing the load equally. To find the

deflection, the number of pins N was removed from equation (7).

Finally a safety factor was considered. As the pins are not a critical element within

the 160 Test Rig, a relatively low value of 1.5 is used.

To find the safety factor:

= (8)

where y is the allowable yield strength of 34CrNiMo6.

The standards details the calculation of the allowable yield strength as

=0.94500 12500 (9)


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Individual columns for each respective stress, pin deflection and safety factor were

created for each pin diameter value to easily evaluate differing variable

combinations.

Table 3. Worksheet showing optimum Pin specification.

Torque (Nm) 6000 Calculation Constant for Shear Stress 1.23

Tangential force (N) 112150 Young's Modulus (GPa) 209

No of Pins 6 In X Condition with H Bmin 341Allowable stress (MPa) 1019

Bending arm length (mm) 12

PinDiameter(mm)

BendingStress(MPa)

ShearStress(MPa)

Deflectionsharedover pinsequally(m)

Deflectionfor 1 pin(m)

MaxEquivalentStress(MPa)

SafetyFactor

8 4462 458 256 1537 4462 0.23

9 3134 362 160 960 3134 0.33

10 2285 293 105 630 2285 0.45

11 1717 242 72 430 1717 0.59

12 1322 203 51 304 1322 0.77

13 1040 173 37 220 1040 0.98


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4.3 Interference fit calculations

As the pins will be interference fit into the hub, calculations needed to be made to

evaluate which fit would needed due the transverse loading of the pins.

The pressure distribution as shown in Figure

6 where Dubbel (1994, p. F29) noted that for

pressure p b

This equated to a pressure of 1.35 GPa.

To calculate the interference fit needed to match the calculated pressure:

= 2( ) (12)

= +26 (11)

Figure 6 . Socket joint under transverse loadwith linear distribution of pressure


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As can be seen in the table, the largest interference fit is for a zc tolerance which

would give +150 m.

In choosing an interference fit, part of the consideration has to be, what fit can be

sensibly achieved. With the pin diameter

only being 15mm, using a very high fit

such as zc tolerance would need

heating/cooling of parts. This endangers

deforming the pin which would be critically

damaging to the project. With this in mind,

a choice of either a p6 or s6 interference fit

would be investigated with values as shown in Table 6.

By rearranging equation (12), it was found

that pressures for the tolerances are as

shown in Table 7.

FEA software would be used to model the

four interference fits to view how the

Table 6. Values for tolerances

Table 7. Pressure for considered interferencefit ranges


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A bonded contact was set between the hole in the bottom block and the pin. Whereas,

a frictional contact was used for the interaction of the pin and the hole in the top

block.

Fixed and Frictionless support constraints were used to limit the degrees of freedom

on the blocks. A remote displacement constraint was used on the bottom of the pin to

restrict rotation.

Mesh settings were input to concentrate refinement in the area of bending on the pin.

An overall size was added in the main mesh settings. Two Sphere of Influence were

centred on a created Coordinate Sphere of Influence system as shown in Figure 8.


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These values acted as the midpoint of the iterations. The element sizes were factored

by h=1.5 in both increasing and decreasing the mesh sizing. The actual mesh sizes

used are detailed in Table 8.

Correlation was deemed to have occurred when the maximum Equivalent Stress as

shown in ANSYS was within 1% of the calculated value.

Table 8. Mesh element sizing for correlation iterations

Iteration1 (mm)

Iteration2 (mm)

Iteration3 (mm)

Iteration4 (mm)

Iteration5 (mm)

Iteration6 (mm)

Final(mm)

Main Mesh 10.13 6.75 4.50 3.00 2.00 1.33 0.89

Sphere ofInfluence 1 10.13 6.75 4.50 3.00 2.00 1.33 0.89

Sphere ofInfluence 2 20.25 13.50 9.00 6.00 4.00 2.67 1.78Face Mesh 2.53 1.69 1.13 0.75 0.50 0.33 0.22

The calculated stress was 676.95 MPa for parameters of 6 pins sharing the load

equally with a pin size of 15mm.


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Table 9. Correlation results from the seven iterations within ANSYS

PIN Mesh Iteration Max EquivalentStress (MPa)% error with

previous mesh% error withCalculatedvalue

Iteration 1 629.57 -7.53

Iteration 2 649.78 +3.11 -4.18

Iteration 3 658.92 +1.39 -2.74

Iteration 4 665.35 +0.97 -1.74

Iteration 5 669.20 +0.58 -1.16

Iteration 6 688.79 +2.84 +1.72

Final 681.67 -1.04 +0.69

With an allowable stress calculated as 1019 MPa, the ANSYS result would equate to

a safety factor of 1.495 which is slightly below the 1.5 limit set previously.

4.4.2 Interference fit calculations correlation


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The interference fit was modelled by applying a Pinball Radius to the outer edge of

the pin circumference. The size of the radius was slightly greater than the tolerance

value for each pin size. The Contact Detection Method was set to Nodal -Projected

Normal from Contact. T his minimised contact pressure spikes at the nodes if a mesh

that is adequately discretised the mesh is used on either side of the contact.

An overall mesh size was used to ensure good discretisation across both components

using sizings of 3mm, 2mm, 1.33mm and 0.89mm again , using an element step

factor of h=1.5.

In formulating the results, both the two tolerance sizes looking at each of the 4 mesh

sizes detailed above. In viewing the pressures, a Contact Tool was added to the

results tab of ANSYS. The probe tool was utilised at 7 equidistant points around the

circumference of the pin from 0 to 180. The mean of the 7 values was compared

with the calculated values.

Correlation was set to when the error between ANSYS and the calculated value was

below 3% and the range of each of the 7 probed values was under 5 MPa.


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In conclusion, for both the pin size and interference calculations, correlation between

the ANSYS equivalent stress and the calculated values were both found using the

same overall mesh size of 0.89mm producing an acceptable error. The set up and

mesh sizing would be suitable to use in modelling and analysis of a final model

which would be derived from measurements of the manufactured gear blank.

4.5 Pin separation due to bending stress

As the force is applied to the pins, bending

occurs. The onset of this bending is below the

plane of the side wall of the hub. Due to this, a

gap opens between the edge of the hole in the

hub and the pin exterior wall as detailed in

Section 4.3.

An interference fit will reduce the gap distance.

Whilst evaluating the pressure correlation within

ANSYS, gap measurements were also taken for

each of the four fits modelled.

Figure 9 . Gap size for s6max interference fitwith force applied on pin


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4.6 Hub hole edge profile

As each pin undergoes bending, the subsequent force in the pin is reacted by the

respective side wall of the hole in the hub. This produces high localised stresses with

the hub.

In trying to manage this stress,

differing edge profiles were

analysed to reduce the stress

concentration factor at the opening

of each hole.

Each profile would be evaluated for

max equivalent stress as in Figure

10 and the number of elements

above the yield stress of 785MPa as

shown in Figure 11 which was

graded between 1 and 10 for performance.

Figure 10 . Stress distribution for Gear profile + chamfer

Figure 11. Elements above the yield stress of 785 MPa for gear


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5 Design of gear blank

The design of the gear blank had to closely match the specification of the actual

gears used in the 160 Test Rig. By using the specification in Figure 12 a design was

created for the blank represented in Figure 13.

Figure 12. Actual 160 Test rig gear specification. Figure 13. Gear blank design specification


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6 Manufacture of gear blank

The gear blank was initially turned

using a Colchester Mascot 1600 to form

the approximate shape of the gear blank

without the holes. It was then machined

finished using a 5-axis CNC machine

(Hurco VMX60SR) where the hole sets

were drilled. To allow machining of the

holes a mandrel was designed which

would hold the gear blank rigidly on the

CNC Machine bed. Figure 15 details the

setup of the Gear blank and mandrel.

The mandrel design allows repeatability of machining the holes with the two stated

methods. After Set 1 holes were created, the gear blank was removed from the

mandrel and then placed back on with a 30 offset. Sets 3 and 4 holes were produced

Figure 15. Gear Blank Manufacturing assembly crosssection


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7 Measurements

The gear blank hole sets were measured using

both a Klingelnberg P65 Gear Measurement

Machine, as shown in Figure 17 and a Zeiss

WMM850 Coordinate Measurement Machine,

as shown in Figure 18.

Four measurement datasets were produced. The

first three gave data on:

PCD of each hole set

True diameter of all twenty four holes Centre coordinates for each of the holes

These were found using the P65 machine. The

fourth and final measurement found was the form of two chosen holes which was

measured using the WMM850 machine.

Figure 17. Klingelnberg P65 GearMeasurement Machine


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7.1 Measurement Results

A full set of the first 3 data set measurements can be found in the appendix, whilst a

condensed version is represented in Table 13.

Set 1

This set was completed by plunging an 15mm Slot Drill.

The PCD of the holes was 160.8 m under specification, however the centre point in

terms of X & Y coordinates was very accurate with deviations of 3.6 m and 2.3 m

respectively.

All six holes diameters in this set were largely out of tolerance. The mean deviationover the specified H7 tolerance of the holes was 385.58 m with a range of 112.1 m

between the holes. The largest diameter measured was 15.4812mm.

The coordinate centres of the holes showed excessive deviation in the X axis while

comparatively small deviation in the Y axis. Holes one and four which had the


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Again the coordinate centres of the holes showed excessive deviation in the X axis

while comparatively small deviation in the Y axis. As with set 1, holes one and four,

which had the largest nominal X value, showed the largest deviation in X axis. The

mean deviation in the X direction was 116.28 m, whereas, in the Y direction the

mean deviation was only 10.63 m.

Set 3

This set was completed by interpolating an 10mm Slot Drill to a diameter of

15mm. Set 3 showed significantly greater accuracy than the first two sets, however,

tolerance was still not met.

The PCD of the holes was only 0.4 m over specification, however the centre point in

terms of X & Y coordinates was less accurate than sets one and two, with deviations

of -31.8 m and -52.8 m respectively.

However, the accuracy of the six holes diameters were greatly improved. The mean

deviation over the specified H7 tolerance of the holes was 14.2 m with a range of


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However, the accuracy of the six holes diameters were greatly improved. The mean

deviation over the specified H7 tolerance of the holes was 11.4 m with a range of

3.2 m between the holes.

Again, the coordinate centres of the holes showed much greater accuracy and showed

no mean difference in the X axis compare to the Y axis. The mean deviation in the X

direction was 4.2 m over tolerance, whereas, in the Y direction the mean deviation

was also 4.2 m.

Form

Two form measurements were also taken as can be seen in Table 12 below. The

measurements were taken with 16 points of contact around the circumference of each

hole. The mean deviation of those points are shown.

As can be seen, the form of the Set 1 hole shows much greater deviation than that of

the hole from Set 3. This was expected due to the inaccuracy of the drilling method

used in Set 1.


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24

Measurement Ideal Measurement Ideal Measurement Ideal Measurement Ideal

mm mm mm m mm mm mm m mm mm mm m mm mm mm m

P CD 1 06 .8 39 2 1 07 0.1608 160.8 P CD 1 06 .8 27 5 1 07 0.1725 172.5 P CD 1 06 .9 99 6 1 07 0.0004 0.4 PCD 107.0085 107 -0.0085 -8.5

x 0.0036 0 0.0036 3.6 x -0.0019 0 -0.0019 -1.9 x -0.0318 0 -0.0318 -31.8 x -0.0287 0 -0.0287 -28.7

y 0.0023 0 0.0023 2.3 y -0.0236 0 -0.0236 -23.6 y -0.0528 0 -0.0528 -52.8 y -0.0461 0 -0.0461 -46.1

H ol e 1 1 5.38 40 1 5.01 8 0.3660 366.0 H ol e 1 1 5.53 97 1 5.01 8 0.5217 521.7 H ol e 1 1 5.03 77 1 5.01 8 0.0197 19.7 H ole 1 1 5. 02 99 1 5. 01 8 0.0119 11.9

x 53.3407 53.5 -0.1593 -159.3 x 53.3607 53.5 -0.1393 -139.3 x 53.4967 53.5 -0.0033 -3.3 x 53.5084 5 3.5 0.0084 8.4

y -0.0073 0 -0.0073 -7.3 y 0.0201 0 0.0201 20.1 y -0.0046 0 -0.0046 -4.6 y 0.0056 0 0.0056 5.6


x 26.6810 26.75 -0.0690 -69.0 x 26.6631 26.75 -0.0869 -86.9 x 26.7471 26.75 -0.0029 -2.9 x 26.7559 26.75 0.0059 5.9

y 46.3244 46.3324 -0.0080 -8.0 y 46.3142 46.3324 -0.0182 -18.2 y 46.3382 46.3324 0.0058 5.8 y 46.3379 46.3324 0.0055 5.5

H ol e 3 1 5.48 12 1 5.01 8 0.4632 463.2

H ol e 3 1 5.45 31 1 5.01 8 0.4351 435.1

H ol e 3 1 5.02 85 1 5.01 8 0.0105 10.5

H ole 3 1 5. 02 75 1 5. 01 8 0.0095 9.5

x -26.6588 -26.75 0.0912 91.2 x -26.6584 -26.75 0.0916 91.6 x -26.7494 -26.75 0.0006 0.6 x -26.7453 -26.75 0.0047 4.7

y 46.3281 46.3324 -0.0043 -4.3 y 46.3318 46.3324 -0.0006 -0.6 y 46.3428 46.3324 0.0104 10.4 y 46.3393 46.3324 0.0069 6.9


x -53.3687 -53.5 0.1313 131.3 x -53.3337 -53.5 0.1663 166.3 x -53.4856 -53.5 0.0144 14.4 x -53.4983 -53.5 0.0017 1.7

y 0.0019 0 0.0019 1.9 y 0.0094 0 0.0094 9.4 y 0.0086 0 0.0086 8.6 y 0.0029 0 0.0029 2.9


x -26.6578 -26.75 0.0922 92.2 x -26.6464 -26.75 0.1036 103.6 x -26.7376 -26.75 0.0124 12.4 x -26.7461 -26.75 0.0039 3.9

y - 46. 3145 - 46.3324 0.0179 17.9 y - 46. 3329 - 46.3324 -0.0005 -0.5 y - 46. 3293 - 46.3324 0.0031 3.1 y -46.3335 -46.3324 -0.0011 -1.1


x 26.6887 26.75 -0.0613 -61.3 x 26.6400 26.75 -0.1100 -110.0 x 26.7539 26.75 0.0039 3.9 x 26.7506 26.75 0.0006 0.6

y - 46. 3074 - 46.3324 0.0250 25.0 y - 46. 3174 - 46.3324 0.0150 15.0 y - 46. 3304 - 46.3324 0.0020 2.0 y -46.3270 -46.3324 0.0054 5.4

Deviation Deviation Deviation

Set 1 Set 2 Set 3 Set 4

Deviation

Table 13. Measurements for the 4 sets including the comparator offset


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7.2 Measurement observations

The most important thing to note is that all twenty four holes were outside tolerance.

That being said, there was significant difference in deviations between the first two

hole sets and the third and fourth. Clearly, using the method of plunging a slot drill

of the same diameter is largely inaccurate compared to the interpolating method.

The head of the 5-Axis machine is direct drive which may help explain the larger

deviation in the X direction for Sets 1 & 2. Having the direct drive has less rigidity

than a conventional belt driven 3-Axis machine. Also one of the degrees of freedom

in the 5 Axis machine is rotational along the X axis which again, partly explains the

larger deviations.

The measurements significantly show how drilling methods and skilled technicians

play a vital role in 21 st century manufacture of highly accurate components. To gain

tolerance of holes, it is standard practice to interpolate holes to a known offset under

the required diameter. Then the hole is measured and the accuracy of the offset is

accounted for in the final stage of producing a hole to tolerance. While this was not


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8 Pin Contact Analysis

With the true measurements known for each set of holes, accurate 3D models were

built to analyse how each pin would contact relative to each other within each

respective set.

As the gears within the test rig can be set to run in both directions, pin contact was

analysed for both a clockwise and counter-clockwise rotation. The distance each pin

had to bend to fully contact and in which order they would contact was evaluated.

By calculating the maximum deflection using equation (7) for a range of torques

from 6000Nm to 1000Nm, a full picture can be built up on how each hole set would

perform. The distance to contact vs deflection was compared using Table 14.

Table 14 . Deflection values needed for pin contact

Number of pins

Deflectionat 6000

Nm (m)

Deflectionat 5000

Nm (m)

Deflectionat 4000

Nm (m)

Deflectionat 3000

Nm (m)

Deflectionat 2000

Nm (m)

Deflectionat 1000

Nm (m)


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For Set 2, in the counter-clockwise rotation, the contact distances ranged 235.9 m,

with the largest being 94.0 m for the third pin to contact. In the clockwise rotation,

the contact distances ranged 126.7 m, with the largest being 76.3 m for the third pin

to contact.

Again at the highest torque, in the both the counter-clockwise and clockwise rotation,

only the first tow pins will ever contact.

For Set 3, in the counter-clockwise rotation, the contact distances ranged 17.8 m,

with the largest being 5.8 m for the third pin to contact. In the clockwise rotation,

the contact distances ranged 18.9 m, with the largest being 9.9 m for the third pin to

contact.

In the counter-clockwise rotation, all pins will contact above a torque of 2000Nm,

where the sixth pin starts to contact at a torque of approximately 1500Nm. In the

clockwise rotation, all the pins will contact from a torque above 1000Nm.

Set 4 had the best performance. In the counter-clockwise rotation, the contact

distances showed the smallest range of any set at 5.7 m, with the largest being


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9 Discussion

The ultimate aim of the project was to fully analyse a pin drive design using FEA

once measurements were taken after a Nitriding heat treatment process.

Unfortunately due to time constraints, the heat treatment and full FEA of a final drive

setup could not be carried out. That being said, the results obtained in the project

have successfully shown optimum requirements for a pin drive to work successfully.

Knowing the gear sizing and torque values, it was determined using 34CrNiMo6

Steel (En24) that 6 pins of 15mm with a bending arm length of 12mm would create

the right balance of sharing the load and giving a necessary flexibility for bending to

allow all the pins to contact as long as components are manufactured to correct

tolerance.

Using an interference fit with a profiled pin design allowed many complexities of

calculations and analysis to be removed, whilst providing functionality. Whilst a high

enough interference fit would be difficult to achieve, using a sensible choice of s 6

minimised gap separation down to 17 m of the pin contact the hub. A fillet of 1mm


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10 Conclusion

With sensible material choice for the components, it was shown, that by calculating

the stresses and evaluating the von Mises criterion, that using six pins with a design

of 15mm and a bending arm length of 12mm would provide a relevant safety factor

of 1.5 in transmitting 6000Nm of torque through the gears in the test rig.

A sensible interference fit of s 6 was shown to provide a balanced pressure to reduce

the pins separation to 17 m, whilst negating the need for costly process to use a high

interference fit.

Using a fillet of 1mm on the hole edges was shown to be the most effective profiling

to reduce stress concentration where pins are bending against the hub.

Manufacturing and creating the holes accurately proved to be the biggest factor in

achieving a workable design. Plunging a slot drill proved highly inaccurate with a

maximum deviation of 519.2 m for the diameter. In comparison, interpolating the

holes proved much greater accuracy with the largest deviation of 19.7 m for the


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11 Further Work

The next part of the process would be for the gear blank to undergo a Nitriding heat

treatment cycle as is normal with the gears used in the test rig.

Another round of measurements would be needed to evaluate any deviation in the

holes during the heat treatment.

At this stage, a full model of the pins, hub and gear blank could be accurately

modelled and analysed using FEA software. This would help understand how the

pins would share the load and see how the stress was distributed across the 6 pins.

Once that knowledge was acquired, a prototype could be used on an existing rig to

prove performance. Strain gauges could be placed on the hub hole entrance to

measure if plastic deformation was occurring.


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Bibliography

Granta (2015) CES EduPack 2014 [Computer program]. Available at:

http://www.grantadesign.com/education/edupack/ (Accessed: 18 March 2015)

Macreadys Standard Stock Range of Quality Steels and Specifications (1995)

London: Royal Print Limited. Seventh Edition

British Standards Institution (1983) BS970: Part 1:1983: Wrought steels for

mechanical and allied engineering purposes . Available at:http://www.bsigroup.co.uk/ (Accessed: 13 February 2015).

British Standards Institution (2006) BS EN 10083-3:2006: Steels for quenching

and tempering. Available at: http://www.bsigroup.co.uk/ (Accessed: 16 February

2015).

British Standards Institution (1976) BS S 156:1976 : 4% nickel-chromium-molybdenum case-hardening steel (vacuum arc remelted) billets, bars, forgings and

parts. Available at: http://www.bsigroup.co.uk/ (Accessed: 13 March 2015).

Bhler Uddeholm (2015) En36A Case hardening steel Datasheet Available at:

http://www.buau.com.au/media/EN36A.pdf (Accessed: 05 May 2015).


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British Standards Institution (2010) BS EN ISO 286-1:2010 Geometrical product

specifications (GPS) ISO code system for tolerances on linear sizes. Available at:

http://www.bsigroup.co.uk/ (Accessed: 13 March 2015).

ANSYS (2015) ANSYS Workbench 15.0 [Computer program]. Available at:

http://www.ansys.com/Products/Workflow+Technology/ANSYS+Workbench+Platfo

rm (Accessed: 24 March 2015)


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12 Appendix

Full Calculations for final design

Initial Calculations

Calculating Force:

=

=6000 0.0535 = 112150

Calculating Bending Stress:

=32

=321121500.0120.015


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= 1.3458 Interference needed:

= 2( )

=20.0150.031.345810

20910 0.050.015

= 212 Pressure for s 6max interference fit:

= ( )

2

=209100.009257.527.525 = 114.12


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Measurements


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Form measurements


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Manufacture

Drawings


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drive pin analysis

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