EXAMENSARBETE INOM MASKINTEKNIK, Innovation och Design, högskoleingenjör 15 hp SÖDERTÄLJE, SVERIGE 2017

Design of a tilting test rig for automotive parts

Johan Jönsson

SKOLAN FÖR INDUSTRIELL TEKNIK OCH MANAGEMENT INSTITUTIONEN FÖR TILLÄMPAD MASKINTEKNIK

Design of a tilting test rig for automotive parts

av

Johan Jönsson

Examensarbete TMT 2017:4 KTH Industriell teknik och management

Tillämpad maskinteknik Mariekällgatan 3, 151 81 Södertälje

Examensarbete TMT 2017:4

Konceptstudie lutningsrigg

Johan Jönsson

Godkänt

2017-03-13

Examinator KTH

Mark W. Lange

Handledare KTH

Nils Gunnar Ohlsson Uppdragsgivare

BorgWarner Sverige Företagskontakt/handledare

Johan Jönsson

Sammanfattning BorgWarners olika system för fyrhjulsdrift och deras förmåga att överföra vridmoment testas på olika sätt för att man ska kunna förutsäga deras prestanda. Då tester utförda på väg kan ge osäkra resultat på grund av skiftande körförhållanden och olika förare, körs ofta testerna i rigg. BorgWarner står i begrepp konstruera en ny tiltrigg. En tiltrigg används för att luta testobjekten och därigenom simulera förändringen av oljenivån som uppstår då en bil accelererar eller lutas. Projektet innebär att upprätta en kravspecifikation för denna rigg och att därefter konstruera den. Sedan några olika idéer testades, baserades det slutgiltiga konceptet på att rotationsaxlar och masscentrum för riggens roterande del sammanfaller. Rotation av riggen kräver därigenom minimalt vridmoment vilket ger mindre påfrestningar. Riggens bord monteras inuti ett svängkranslager, vilket ger rotation kring ena axeln som sammanfaller med masscentrum. Svängkranslagret roteras sedan kring en annan axel, så att rörelse kring två axlar blir möjlig. Konstruktionen kräver endast hälften så mycket vridmoment som en av tidigare utförande. Riggbordet är likväl lättåtkomligt för montering av testobjekt. Nyckelord provutrustning, maskin konstruktion

Bachelor of Science Thesis TMT 2017:4

Design of a tilting test rig for automotive parts

Johan Jönsson

Approved

2017-03-13 Examiner KTH

Mark W. Lange Supervisor KTH

Nils Gunnar Ohlsson Commissioner

BorgWarner Sverige Contact person at company

Johan Jönsson

Abstract In order to predict the performance of their all-wheel drive systems and the torque transferring capabilities of these, BorgWarner carry out different tests. Road testing can depend on change of conditions and different drivers, so rig testing is often preferred when comparable results are required. BorgWarner wants to design new rig for tilt tests. A tilting rig simulates accelerations and inclinations of the test object, with possible effect on oil level. This project concerns the definition of requirements for the rig. Finally, a design is proposed. After a couple of design iterations, the final concept was based on the idea that the rotational axes should coincide with the center of gravity of the moving rig and its components in order to minimize the torque needed for rotation. This was accomplished by mounting the rig table inside a slew drive providing rotation around on axis. This slew drive, in turn, can rotate around another axis, thus providing dual axes rotation. This design only requires half the torque compared to the current solution. Still, easy access to the rig table is ensured. Key-words test equipment, machine design

Preface As part of the Degree Programme in Mechanical Engineering, Innovation and Design at the Royal Institute of Technology (KTH) in Stockholm, this project was made in collaboration with BorgWarner PowerDrive Systems AB in Landskrona.

It has been a true pleasure to work alongside the talented employees at BorgWarner: their insight and experience have been invaluable to the result of this project. I would in particular like to thank my supervisors Johan Karlsson, Adam Eliasson and Jonas Jönsson at BorgWarner for their guidance and support.

Table of contents Introduction ......................................................................................................................................................... 1

Background ...................................................................................................................................................... 1 Function of the AWD .................................................................................................................................. 1 Testing.......................................................................................................................................................... 1 The current rig ............................................................................................................................................. 3

Problem definition .......................................................................................................................................... 4 Method ............................................................................................................................................................. 4

Data Acquisition .......................................................................................................................................... 4 Concept generation ..................................................................................................................................... 4 Execution ..................................................................................................................................................... 4

Requirement specification .............................................................................................................................. 4 Limitations ....................................................................................................................................................... 5

Analysis ................................................................................................................................................................ 7 Market analysis ................................................................................................................................................ 7 Questionnaire .................................................................................................................................................. 7 Simulation of acceleration in a test environment .......................................................................................... 7

Determining how tilting angles correlate to acceleration ......................................................................... 8 Accelerations achieved when driving ......................................................................................................... 8 Considering what rotational acceleration is viable ................................................................................... 9

Torque demand ............................................................................................................................................. 10 Two concepts ................................................................................................................................................. 19 Accomplishing the rotations ........................................................................................................................ 20

Pneumatic ................................................................................................................................................. 20 Hydraulic .................................................................................................................................................. 20 Electric ....................................................................................................................................................... 21

Table height ................................................................................................................................................... 21 Finding a way to temper the test objects ...................................................................................................... 23

Results ................................................................................................................................................................ 24 Frame ............................................................................................................................................................. 25

Bending stress and deflection ................................................................................................................... 25 Steel plate for mounting the frame to the slew drive .................................................................................. 27 Slew drive ....................................................................................................................................................... 27 Servomotors ................................................................................................................................................... 27

Rotation around the y axis ........................................................................................................................28 Rotation around the x axis ........................................................................................................................28

Connection of slew drive to the stand .......................................................................................................... 29 Stand .............................................................................................................................................................. 29

Bearing ....................................................................................................................................................... 29

Test table ........................................................................................................................................................ 31 Centering fixtures .......................................................................................................................................... 32 Motor running the test objects ..................................................................................................................... 32 Torque sensor ................................................................................................................................................ 33 Drive shaft ...................................................................................................................................................... 33 Routing the cables ......................................................................................................................................... 34

Conclusion and recommendations ................................................................................................................... 35 Simulating in real time .................................................................................................................................. 35 Motor and gearbox selection ........................................................................................................................ 35 Climate chamber ........................................................................................................................................... 35

References .......................................................................................................................................................... 37 Appendices ......................................................................................................................................................... 39

Appendix A: Market analysis ........................................................................................................................ 39 Appendix B: Questionnaire ........................................................................................................................... 41

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Introduction This report will study the design of a tilting test rig for automotive parts. The following chapter will provide an introduction for the chapters to follow.

Background BorgWarner is a world-renowned company active in powertrain solutions. It consists of two divisions: engine and drive train. The BorgWarner plant in Landskrona belongs to the second, providing BorgWarner with state of the art all-wheel drive systems (AWD-systems). Formerly known as Haldex Traction, BorgWarner purchased the division from Haldex AB back in 2011. Their main product that has been continuously developed since the first product launch in 1996 is the on demand all-wheel drive system, a system that provides torque to a secondary wheel pair when needed.

Function of the AWD The technology used in the different configurations of AWD-systems BorgWarner supplies are all very

similar: As the car’s different sensory systems detect the need of more torque, an oil pump is actuated, pressurizing a clutch pack which in turn connects the secondary (front or rear) axle with the primary axle.

Testing The development of these advanced systems requires extensive testing. Tests are either done on the road or in a test rig. There are pros and cons of both ways of testing, but both are invaluable for the development. Testing on the road will provide test conditions which are not achievable in a test rig. The elements of nature are hard to reproduce in a test rig. The problem with road testing is that a driver of a car would have a hard time reproducing exactly the same procedure time and again. There is also an issue in the fact that several testers perform the same tests, but often with different results. For repeatability, rig testing offers outstanding performance.

Figure 1. Schematic layout of the BorqWarner AWD

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When driving a car, inclination of the oil in the test objects happens during cornering and/or acceleration/deceleration and when the car itself is driving on a slope. The rig tests what consequences these different driving conditions have on the test object. In order to simulate the inclination of the oil level in the test object, the rig is tilted to different angles which correlates to the accelerations achieved during driving. A motor is used to simulate the rotation normally provided by the prop shaft.

The types of test performed in the rig are typically the following, all run at different speeds, angles and temperatures:

Drag torque test To measure the internal friction, and thereby torque loss, through a test unit a drag torque test is performed. A torque meter is used to measure the torque. Preferably the torque lost is zero, but due to internal friction of preloaded bearings, seals and the friction between the lamellas this is never achievable. Typically, a couple of Nm is recorded, but during some tests involving cold conditions combined with tilt, more than 20 Nm of drag torque has been recorded.

Lubrication test To make sure that the lubrication of the test object is sufficient it is often run with a fully or partially clear housing, enabling the tester to evaluate the lubrication in real time. If lubrication is lacking, the bearings and seal are at risk of running dry. Proper cooling of the clutch pack is also dependent on proper lubrication. If the pump pressurizing the lamellas end up without oil, it will start pumping air. As air is compressible it is not capable of pressurizing the clutch pack, rendering a complete loss of torque transmission.

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The current rig The image below identifies the different components of the current rig and the coordinate system defining relevant degrees of freedom.

Figure 2. The current tilt rig

1. Motor 2. Torque sensor 3. Test table 4. Rig table 5. Actuator

The rig tilts around its x and y axis. The origin of the coordinate system is located at the rigs center of rotation. This coordinate system will be used throughout the report to reference the different motions.

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Problem definition BorgWarner needs to push new technology to the market in an effective way. This means making sure that development and testing of concepts is as streamlined as possible. The tilting tests of today are not automated and there is only one rig available for the tests. This rig is also used for other tests that do not involve tilting, further prolonging the que for other tilting tests.

Some tests require tempering of the test object, which is done by placing the test object in a separate climate chamber. Since the test object cannot be continuously tempered throughout the tests, an uncertainty regarding what temperature the tests are actually run at occurs, and a lot of time is spent carrying out the tempering.

The time of the test engineers is spent unwisely. At times a test is run in just a couple of minutes, but in between time-consuming adjustments has to be made to the rig due to the geometric differences of the test objects. For longer test the test engineer has to spend his time at the rig, as the tests are not automated.

The current rig’s characteristics will provide the basis for the problem definition. Areas in need of improvement were identified as well as what key features should be implemented in the new one.

These are some of the areas in need of improvements:

1. The tilting is limited to ±30 degrees of combined tilt, whereas test specifications require ±51 degrees of combined tilt.

2. The table can be rocked back and forth using your bare hands, impairing the certainty that tests are performed at correct angles and in a repeatable manner. The rig is prone to tipping over with heavy loads combined with steep tilting.

3. The adjustment of the tilt is manual, angles are measured by hand and adjustment is not possible during the tests.

Despite these shortcomings, many features of the current rig are satisfactory: the rig table is accessible from all directions and it has a good working height. These features should be passed on to the new rig as well.

Method Throughout the project, different phases have required different methods. These methods are described in the following segments.

Data Acquisition Initially time was spent understanding the purpose of the testing through studying of the test engineers using the current rig and through reading test reports. A market analysis was performed, studying current solutions and similar products. A questionnaire regarding sought for functions by the testers was used. The experiences from different courses at KTH was used to calculate necessary tilt angles and needed torques and accelerations for rotating the table. MATLAB was used to effectively use these calculations for varying geometries.

Concept generation Brainstorming was used to come up with suitable concepts. The foundation for creating new concepts were the current solutions found during the market analysis. As brainstorming progressed, new iterations of current products were generated and some innovative concepts were drafted.

Execution The concepts selected were modeled using Creo Parametric 2.0. Animations were created to facilitate an easier interaction with suppliers and to visualize the concepts for a presentation at BorgWarner.

Requirement specification As work progressed, the requirement specification became clearer. It developed throughout the project as the demands kept changing, but the following reflects its final iteration:

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Inclination The rig should be able to tilt the rig table ±55 degrees around both the x and y axis simultaneously. Any combination within this range should be attainable with an accuracy of ±0,5 degree.

Acceleration The time the table needs to adjust from -55 to +55 degrees on both axes should be no more than 10 seconds.

Weight capacity Maximum allowed weight of test object with fixture: 150 kg.

Size The size of the testing cell limits the size of the rig: 3.5 by 3 meters. The rig should be sized so that working around it is practical. The working height should be ergonomic and therefore not exceed 950 mm.

Installation The rig is to be permanently mounted to the floor in the test cell.

Motor The motor should be able to continuously output 35 Nm of torque and be able to reach 10,000 rpm.

Test table The test table of the rig should, to the greatest extent, be kept free from obstacles blocking easy access to it. The table should have a guide that centers the fixtures on the table.

Adaption of temperature chamber The rig is to be fitted with a temperature chamber, capable of cooling and heating the test object’s oil to between -30° to +140°.

Standardized interface for the axles All axels used should be interchangeable between different test objects. Adapters should be made for each test subject to fit onto said axles.

Adapting the distance between the motor axle and the ingoing axle The distance in the z direction between the outgoing axle of the motor and the ingoing axle of the test object needs to be either easily adjustable or have a fixed distance.

Automation The rig’s motions should be automated and controlled from a computer outside the test cell.

Synchronized recording The rig and its computer should be able to automatically record test scenarios.

Limitations With the limited time available, some areas of the design and construction must be left for later consideration. These areas are the following:

• The development of the automatic control for the rig.

• The calibration of the sensors.

• The final selection of components. Suggestions will be made though.

• The budget for the construction.

• If no suitable way of constructing the climate chamber is found, it is to be omitted.

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7

Analysis This chapter will focus on the acquisition of data and the processing of it to conclude the demands of the rig and the design of the first concepts.

Market analysis The possible sources of inspiration were broad, ranging from advanced six degrees of freedom Stewart platforms to tilting rigs capable of tilting a full car. There were no products found that would fit for the application out of the box. The alternatives from CFM Schiller (CFM Schiller GmbH, n.d.) are all too big for the intended test space, whereas other alternatives from Bosch (Rexroth Bosch Group, n.d.) and Intertek (Intertek, 2016) based on the Stewart platform are limited in the achievable tilt angles. An inspiration was found in gimbals for cameras, as they provide the same rotational capabilities as the tilt rig from CFM Schiller, but with the potential of having most of the rig table free from obstruction by the frame. The full market analysis can be found in Appendix A.

Questionnaire The desires of the test engineer were collected by having them answer a questionnaire. The most sought for attribute was to have the test table adjustable from the rig computer outside of the test cell. The second most requested features were to automate the test flow and to have a rigid and dependable testing rig. The questionnaire can be seen in Appendix B.

Simulation of acceleration in a test environment

The motions a test object and its oil are subjected to when driven on a road are very complex to simulate. In relation to Figure 3 the motions are rotation around and acceleration along the x, y and z axes. This corresponds to six degrees of freedom.

There are no available solutions today to fully simulate all of these motions properly. There are systems that can do it momentarily, like the Stewart platform, but not for a sustained time.

What BorgWarner is interested in is simulating four out of these six degrees of freedom: rotation around and acceleration along the x and y axes. Rotational motions are simulated by rotation around the corresponding axis. To properly simulate the accelerations however, movement along the axes are necessary. In a confined area, continuously moving along an axis to simulate a constant acceleration is not possible. But since the only thing of interest is the inclination of the fluid within a test object created by the acceleration, this can be simulated with a rotation around the perpendicular axis in the horizontal

Figure 3. The six degrees of freedom of a car

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plane. Simulating acceleration along the x axis is done by rotating around the y axis and vice versa. This is a simplification, as during real time driving the oil would slosh around in the housing. The simplification is agreed upon by both BorgWarner and its customers and therefore, BorgWarner only needs a tilt rig that can rotate around the x and y axes.

Determining how tilting angles correlate to acceleration The accelerations affecting the inclination of the oil’s surface is the acceleration of gravitation, the centripetal acceleration during cornering and the longitudinal acceleration during acceleration/braking of the car. The oil used in the test objects is a low viscosity oil (Statoil , 2007), and it is therefore shifting quickly from side to side with changed acceleration. The normal to the surface of the oil coincide with the resulting acceleration vector according to Figure 4.

With an acceleration, a and the acceleration of gravity, g, the necessary tilt angle 𝜑𝜑 of the rig is calculated using equation 1.

𝑡𝑡𝑡𝑡𝑡𝑡(𝜑𝜑) = 𝑎𝑎𝑔𝑔⇒ 𝜑𝜑 = 𝑡𝑡𝑡𝑡𝑡𝑡−1 �𝑎𝑎

𝑔𝑔� (1)

Accelerations achieved when driving To simulate the oil inclination in the test units, an understanding of what accelerations a car could achieve was needed. A suitable reference was found in the Bugatti Veyron, as it is the most capable car the company has developed an AWD-system for. The logs obtained were from a run around a dry track.

Figure 4. Schematic correlation between the centripetal acceleration and gravity

a

g

φ

9

Figure 5. Bugatti Veyron accelerations

It was concluded that the maximum accelerations achieved were approximately 12 m/s2 along the x axis and 11 m/s2 along the y axis. The time between maximum negative and positive acceleration was 0,5 seconds. This is an indication of how fast the driver of the Bugatti was able to change direction from fully developed acceleration in one direction until it was achieved in the other direction. Using the largest acceleration and equation 1, the inclination angle was calculated:

𝜑𝜑 = 𝑡𝑡𝑡𝑡𝑡𝑡−1 � 129.82

� ≈ 51 𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑 (2)

Some margin was desired. It was therefore decided that the table should be able to tilt ±55 degrees on both the x and y axis.

𝜑𝜑𝑥𝑥 = 𝜑𝜑𝑦𝑦 = ±55 𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑𝑑 (3)

Considering what rotational acceleration is viable As mentioned before, the time needed for a car to go from fully developed negative lateral acceleration to fully developed positive lateral acceleration was 0,5 seconds. The initial goal was to be able to simulate this by tilting between the corresponding extreme angles in the same amount of time. Questions began arising as to what BorgWarner would actually accomplish by adhering to these acceleration demands. What would be simulated by rotating the rig between its extreme angles in the same amount of time the driver could change acceleration from side to side? Would it actually simulate what was happening with the oil in the housing? Since tilting the rig to simulate acceleration is a simplification in its own, the team at BorgWarner decided that no emphasis should be put into making the rig simulate the changes of acceleration. It would complicate the construction of the rig and, based on some initial quotes, make it at least twice as expensive. The decision made was therefore to allow the rig to alter its tilt angles in a reasonable time, preferably in less than ten seconds between its extreme angles.

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Torque demand The rig is going to rotate around two axes. The most demanding rotation is the rotation from one extreme

angle to the other, corresponding to the change of acceleration from one side to the other in a car.

It is assumed that for this rotation, the table is to be accelerated with a constant acceleration during the first half of the rotation, and then decelerated the other half of the rotation. In reality the movement would be achieved by ramping the acceleration to control the jerk, but as this ramp would be quite rapid, and to simplify the calculations, constant acceleration is assumed. The angular movement is dependent on the acceleration and the time according to equation 4 for rotational movement:

𝜑𝜑 = 𝜔𝜔𝑜𝑜𝑡𝑡 + �̈�𝜑𝑡𝑡2

2 (4)

With the starting angular velocity of the rig being zero and the degrees converted to radians, the equation can be solved for angular acceleration:

�̈�𝜑 = 2𝜑𝜑𝑡𝑡2

= 2∗55𝑡𝑡2

∗ 𝜋𝜋180

(5)

An objects resistance to changing its angular velocity is called the mass moment of inertia (MMOI). To change an objects angular momentum, a Torque is needed. This can be calculated by using the rotational version of Newton’s second law:

𝜏𝜏 = 𝐼𝐼�̈�𝜑 (6)

Where τ is the torque, I is the MMOI and �̈�𝜑 is the angular acceleration.

Figure 6. Simplified model of rig table used for torque calculations

φ

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To calculate the torque with the acceleration known, the MMOI of the rig needs to be calculated. To simplify calculations, the motor, test object and rig table were considered homogeneous cuboids with a CG in their geometrical center according to Figure 7.

The MMOI can therefore be calculated for each component using the MMOI equations for homogenous cuboids:

𝐼𝐼𝑥𝑥 = 𝑚𝑚12

(𝑏𝑏2 + 𝑐𝑐2), 𝐼𝐼𝑦𝑦 = 𝑚𝑚12

(𝑡𝑡2 + 𝑏𝑏2) (7)

Where m is the mass of the object and a, b and c is the length of each side according to Figure 7.

Equation 7 applies when the object is rotated around an axis going through its center of gravity (CG). According to Figure 7, if an object is rotated around a parallel axis y0, the MMOI is calculated using the parallel axis theorem:

𝐼𝐼𝑦𝑦0 = 𝐼𝐼𝑦𝑦 + 𝑚𝑚𝑑𝑑2 (8)

Where d is the perpendicular distance between the rotation axis y0 and center of gravity axis y and m is the mass of the component.

Figure 7. The cuboid used for the mass moment of inertia calculations

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Figure 8. Schematic drawing of rig

Seen above is a schematic drawing of the moving parts of the tilt rig and their corresponding CG:s and their distances to the rotational axes (RA). These dimensions are used with a combination of equation 7 and 8 to find the MMOI around the RA of the components. For example, the MMOI for the motor around the rotational axis going through y (RAy) is calculated:

𝐼𝐼𝑀𝑀𝑜𝑜𝑡𝑡𝑜𝑜𝑜𝑜𝑅𝑅𝑅𝑅𝑦𝑦 = 𝑚𝑚𝑀𝑀𝑀𝑀𝑀𝑀𝑀𝑀𝑀𝑀12

(𝐿𝐿𝐿𝐿2 + 𝐻𝐻𝐿𝐿2) + 𝑚𝑚𝑀𝑀𝑜𝑜𝑡𝑡𝑜𝑜𝑜𝑜 ∗ 𝐶𝐶𝐶𝐶𝐿𝐿_𝑅𝑅𝑅𝑅2 (9)

Presented hereafter are the calculations made for the rotation around the y axis RAy, but the same theory applies to rotation around the x axis.

After calculating the MMOI for all components around RAy they are then summed up:

𝐼𝐼𝑅𝑅𝑅𝑅𝑦𝑦𝑡𝑡𝑜𝑜𝑡𝑡 = 𝐼𝐼𝑀𝑀𝑜𝑜𝑡𝑡𝑜𝑜𝑜𝑜𝑅𝑅𝑅𝑅𝑦𝑦 + 𝐼𝐼𝑅𝑅𝑅𝑅𝑔𝑔 𝑡𝑡𝑎𝑎𝑡𝑡𝑡𝑡𝑡𝑡𝑅𝑅𝑅𝑅𝑦𝑦+ 𝐼𝐼𝑇𝑇𝑡𝑡𝑇𝑇𝑡𝑡 𝑜𝑜𝑡𝑡𝑜𝑜𝑡𝑡𝑜𝑜𝑡𝑡𝑅𝑅𝑅𝑅𝑦𝑦

(10)

With both angular acceleration and the MMOI of the rig known, the torque needed due to acceleration demands can be calculated using equation 6.

𝜏𝜏𝑅𝑅𝑅𝑅𝑦𝑦 = 𝐼𝐼𝑅𝑅𝑅𝑅𝑦𝑦𝑡𝑡𝑜𝑜𝑡𝑡�̈�𝜑 (11)

The center of gravity (CG) for the entire rig is offset from the RA both in the x and z direction, CGX_RA and CGZ_RA. The position of CG causes an additional torque depending on the tilt angle 𝜑𝜑 of the rig.

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Figure 9. Offset between CG and RA

The distance CG_RA and the angle α does not change when the angle 𝜑𝜑 changes. The added torque is a function of the vertical force due to gravity, mg, with an origin in CG, and the horizontal distance between RA and CG:

𝜏𝜏𝐶𝐶𝐺𝐺𝑅𝑅𝐴𝐴𝑦𝑦 = 𝑚𝑚𝐶𝐶𝐺𝐺 ∗ 𝑑𝑑 ∗ 𝐶𝐶𝐶𝐶_𝑅𝑅𝑅𝑅 ∗ 𝑐𝑐𝑐𝑐𝑑𝑑(𝜑𝜑 + 𝛼𝛼) (12)

To minimize this torque, the length of CG_RA should be minimized. The distance CGX_RA (see Figure 8) is determined by the size and weight of the test object, and is therefore hard to change. The distance CGZ_RA (see Figure 8) can be altered in the design of the rig by moving the rotational axes above the rig table, and is minimized by finding the point where CGZ_RA=0 when performing the calculations using the dimensions of the heaviest test object.

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Figure 10. The two different rotational axes

Using the current rig as a reference for axis placement, axis number one represents the current setup with a rotational axis below the table and axis number two the rotation above the rig table (see Figure 10).

Plotted below are the results for the added torque depending on the axis placement. For every 𝜑𝜑 between -55 degrees to 55 degrees the added torque was calculated using equation 12.

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With the rotation axis in position one (Figure 10) the torque 𝜏𝜏𝐶𝐶𝐺𝐺𝑅𝑅𝐴𝐴1 changes according to Figure 11 through its tilt around the y axis.

Figure 11. Added torque when rotating around RA1

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Figure 12. Added torque when rotating around RA2

With the CG_RA minimized by placing the rotational axes in position two (Figure 10) the torque 𝜏𝜏𝐶𝐶𝐺𝐺𝑅𝑅𝐴𝐴2 changes according to Figure 12 through its tilt around the y axis. The maximum torque needed to hold the rig is more than halved.

The total torque needed to tilt the rig is found by adding the torque needed due to acceleration demands (equation 11) and the maximum torque from the offset of CG (equation 12):

𝜏𝜏𝑡𝑡𝑜𝑜𝑡𝑡𝑅𝑅𝐴𝐴𝑦𝑦 = 𝜏𝜏𝑅𝑅𝑅𝑅𝑦𝑦 + 𝜏𝜏𝐶𝐶𝐺𝐺𝑅𝑅𝐴𝐴𝑦𝑦 (13)

Plotted below are the results for the calculations of necessary torque around the y axis depending on in what time the rotation is to be performed. The time represents the completion of half the rotation, from its extreme angle 55 degrees of tilt to zero degrees of tilt. The two graphs represent the different positions one and two of the RA (Figure 10).

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Figure 13. Total torque demand for rotation around RA1

Figure 14. Total torque demand for rotation around RA2

As can be seen by comparing the two graphs, the torque needed to rotate the rig is roughly halved by minimizing the distance between the CG and the RA. Notice that after about 1.5 seconds, the total torque

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needed stabilizes. Thus, the results are based on the torque needed to perform half the rotation in 1.5 seconds, and the rotational acceleration is determined using equation 5.

�̈�𝜑 = 2𝜑𝜑𝑡𝑡2

= 2∗55𝑡𝑡2

∗ 𝜋𝜋180

= 2∗551.52

∗ 𝜋𝜋180

= 0.853𝑑𝑑𝑡𝑡𝑑𝑑/𝑑𝑑2 (14)

The calculated maximum torque needed to rotate the rig around the y axis is 900 Nm.

Since minimizing CGZ_RA is based on approximations, it was decided to perform the calculations using CGZ_RA±100 mm. This variation added roughly 100 Nm to the result, bringing the torque demand 1,000 Nm.

𝜏𝜏𝑡𝑡𝑜𝑜𝑡𝑡𝑅𝑅𝐴𝐴𝑦𝑦 = 1,000 𝑁𝑁𝑚𝑚 (15)

The corresponding calculations for the rotation around the x axis are based on the same principles and will therefore not be described in detail. The difference is that the offset CGX_RA does not affect the torque needed. Doing the calculations with CGZ_RA±100 mm, the resulting torque needed for the rotation around the x axis was 400 Nm.

𝜏𝜏𝑡𝑡𝑜𝑜𝑡𝑡𝑅𝑅𝐴𝐴𝑥𝑥 = 400 𝑁𝑁𝑚𝑚 (16)

With the rotational axes located at CGZ_RA=0, the distance RA_CGR is 330 mm.

19

Two concepts The first concept was designed taking cues from the current rig, eliminating the drawbacks of it while retaining the pros.

Figure 15. The first concept

The rotational axes were separated so that the motion of one actuator would not affect the other. The tilt around the x axis would cause a force pulling/pushing the base of the actuator in the y direction. To counter this deflection that would cause the actuator to hit the stand of the rig, a frame is made that runs along the side of the stand and is connected to the base of the actuator. The friction caused by the rubbing of the frame could either be solved with a coating of a low friction material like PTFE or by fitting a rolling bearing to the stand.

The drawback of this concept is that the rotational axes are located below the rig table, which, as seen in the torque demand calculations, is not ideal.

The second concept revolved around designing a rig that would allow for the rotational axes to be placed above the table without restricting the access to the rig table like the alternatives seen in the market analysis.

A solution was found in mounting the frame through a slew drive. A slew drive is a big bearing capable of simultaneously handing axial, radial and bending loads. As it is made for having a rotary actuator driving one of the two bearing raceways, it also handles the rotation of the frame along the x axis.

20

Figure 16. The second concept

The axis of rotation coincides with the mass center along the x axis and the y axis coincides with the center of mass in the z direction.

A presentation of the two concepts was made in front of the team at BorgWarner. This team choose the second concept for further development.

Accomplishing the rotations Movement is accomplished by using an actuator of some kind. The two main designs are rotary and linear actuators. The slew drive is to be controlled by a rotary actuator. Choosing a linear actuator for the rotation around the y axis would therefore lead to an unnecessarily complicated system. As a result, a rotary actuator was selected for the y axis rotation as well. To decide whether the rotation was going to be done by means of a pneumatic, hydraulic or electric system, the pros and cons were compared.

Pneumatic Pros

• Simple system. • Good for extreme temperatures. • Cheap.

Cons • Need to constantly pump air pressure to keep its position due to leakage and air being a

compressible gas. • Not suitable for closed loop applications, meaning that usually no information about what

happens during the motion is provided and it ends at a pre-determined mechanical stop.

Hydraulic Pros

• Great for high torque demands. • Can hold position without the pump adding more fluid. • The pump and motor can be located elsewhere. • Possible to setup as a closed loop system with full control over acceleration, velocity and position

throughout the movement regardless of the applied load.

21

Cons • The risk of a fluid leak. • Need a lot of extra parts like pump, valves, heat exchanger and tank. • The other rigs do not use hydraulics, leading to a more complexity and a need to acquire

additional knowhow.

Electric Pros

• Great precision with a servo electric motor. • Quieter than hydraulic and pneumatic motors. • Available in house knowledge and experience of working with and controlling electric motors. • No leakage. • Closed loop systems available. • Minimal maintenance required.

Cons • More expensive than the other solutions. • The size of an engine with its gearbox might be bigger than its pneumatic or hydraulic

counterpart. The drive controlling the motor can be placed remotely though.

With the need to accurately move to specified but variable positions, a closed loop system that continuously communicates the position of the actuator was needed. Having a closed looped hydraulic system involves a more complex solution with expensive servo valves or with electric servo motors controlling the pump for the hydraulic motor. As no hydraulic systems are used today at BorgWarner for controlling the rigs, there were no obvious advantages in a hydraulic system.

Therefore, an electric solution was selected and the different electric motors were evaluated. After discussions with several suppliers, it was decided that for the load and speed conditions at hand, a servomotor was preferable.

Table height To be able to tilt the table to the required angles, the height of the rig needs to be adequate. Since the table height is specified to be below 950 mmm, this limits the length and width of the table. As the table tilts on both the x and y axis, they will both affect the necessary table height. To calculate the minimal length of the table, the longest test object was used as reference, seen in the Figure 17.

Figure 17. The longest test object

The objects are, from left to right: the Porsche torque tube, the driveshaft, the torque meter and the motor. This adds up to the theoretical total length of all units on the rig table:

𝐿𝐿 = 1,130 + 150 + 50 + 480 = 1,810 𝑚𝑚𝑚𝑚 (17)

For good measure, a calculation of the necessary table height with a table length of 1,900 mm and a width of 800 mm was performed, using the ideal placement of the rotational axis RA-CGR=330 mm found in the “Torque demand” chapter and both 𝜑𝜑𝑥𝑥 and 𝜑𝜑𝑦𝑦 set to 55 degrees.

1,130 150 50 480 [mm]

22

Figure 18. Tilt angles defined

To calculate the necessary height of the table to accommodate the tilts of both axes, the following equation was derived through triogonometry.

𝑁𝑁𝑑𝑑𝑐𝑐𝑑𝑑𝑑𝑑𝑑𝑑𝑡𝑡𝑑𝑑𝑁𝑁 ℎ𝑑𝑑𝑒𝑒𝑑𝑑ℎ𝑡𝑡 = 8002∗ 𝑑𝑑𝑒𝑒𝑡𝑡(55) ∗ 𝑐𝑐𝑐𝑐𝑑𝑑(55) + 1,900

2∗ 𝑑𝑑𝑒𝑒𝑡𝑡(55) − 330 + 330 ∗ 𝑐𝑐𝑐𝑐𝑑𝑑(55) = 825 𝑚𝑚𝑚𝑚 (18)

φx

φy

23

The first term is the table's width addition to the height when tilted around the x and y axis. The second term is the calculation of the table’s length addition to the height. The third term calculates the lift due to the rotation occuring above the rig table. Below is the result of the calculation potted with 𝜑𝜑𝑦𝑦 on the x axis and the necessary table height on the y axis.

Figure 19. Necessary table height

Finding a way to temper the test objects Other rigs at BorgWarner are fitted with an external temperature chamber that is connected to a climate test cabinet. The connection is made with thick insulated tubes that are heavy and inflexible. These tubes were deemed too inflexible to use for the tilting rig, as it was to move around a lot. An alternative would be mounting a climate test cabinet onto the rig or having the entire rig inside a temperature chamber. The first option was ruled out as no such systems were found that would fit the rig. The second was ruled out due to other objects of the rig not being able to withstand the extreme temperatures, especially the torque meter. Different distributors and manufacturers were contacted to find a solution to the inflexible tubes, but none seemed to have an idea on how to continuously temper a test object on the rig. Because of this, the decision was made that no more time should be invested in designing the external temperature chamber, but that the design of the rig should allow for one to be mounted later.

24

Results This chapter will present the final design and, where applicable, the suggestion of a component. Custom parts are described with material selection and manufacturing method.

Figure 20. The final design

25

Frame

Figure 21. Comparative picture of frame design

The first concept had a rather intricate frame that was bent in both the x, y and z direction. This was done in order to fit the frame through a narrow slew drive while still keeping the rotation axis coincide with the center of mass and keeping cost to a minimum by not using an unnecessarily big slew drive. The implications of this decision were a more complex and weak frame and a design that limited the length of the test table. The latter is rather important, as this in turn limits the size of the future temperature chamber. It was therefore decided to use a larger size slew drive and a frame that passes straight through it. The parts of the frame are welded together and are then mounted to the slew drive using a bolted joint.

The following section will describe the calculations made on the strength of the frame that the motor and test table attaches to (see Figure 22).

Bending stress and deflection The rig frame is designed using strength calculations. The desired width of the frame was 160 mm or more, to allow for rigid mounting of the engine. The maximum deflection at the end of the table was not to exceed 1 mm, as any deflection would affect the alignment of the connection between the test object and the motor. The main part of the frame is constructed out of a rectangular hollow structural section (HSS) profile made from S355J2S steel, a readily available and reliable construction steel. With no need for custom manufacturing, money and time is spared.

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Figure 22. Deflection constraints

The simplifications made are that the analysis of the bending of the frame is done on the side that is subjected to the greatest weight. The connection of the frame to the slew drive is considered fixed and the load is simplified to a point load, see Figure 23. The approximate weight of the heaviest test object plus the temperature chamber is 300 kg, adding a force 𝐹𝐹𝑚𝑚𝑔𝑔 at the CG of the test object. The acceleration �̈�𝜑 of the rig contributes to the total force, as it corresponds to a linear acceleration adding a force 𝐹𝐹𝑎𝑎𝑜𝑜𝑜𝑜𝑡𝑡𝑡𝑡𝑡𝑡𝑜𝑜𝑎𝑎𝑡𝑡𝑅𝑅𝑜𝑜𝑎𝑎:

Figure 23. Schematic picture of table movement

𝐹𝐹𝑎𝑎𝑜𝑜𝑜𝑜𝑡𝑡𝑡𝑡𝑡𝑡𝑜𝑜𝑎𝑎𝑡𝑡𝑅𝑅𝑜𝑜𝑎𝑎 = 𝑚𝑚 ∗ 𝐶𝐶𝐶𝐶𝐶𝐶𝑥𝑥 ∗ �̈�𝜑 (19)

Giving the total force:

𝐹𝐹𝑡𝑡𝑜𝑜𝑡𝑡 = 𝐹𝐹𝑎𝑎𝑜𝑜𝑜𝑜𝑡𝑡𝑡𝑡𝑡𝑡𝑜𝑜𝑎𝑎𝑡𝑡𝑅𝑅𝑜𝑜𝑎𝑎 + 𝐹𝐹𝑚𝑚𝑔𝑔 = 𝑚𝑚 ∗ 𝐶𝐶𝐶𝐶𝐶𝐶𝑥𝑥 ∗ �̈�𝜑 + 𝑚𝑚𝑑𝑑 = 300 ∗ 0.475 ∗ 0.853 + 300 ∗ 9.82 = 3,068 𝑁𝑁 (20)

The bending moment is:

𝐿𝐿 = 𝐹𝐹𝑡𝑡𝑜𝑜𝑡𝑡 ∗ 𝐶𝐶𝐶𝐶𝐶𝐶𝑥𝑥 = 3,068 ∗ 475 = 1,457,087 𝑁𝑁𝑚𝑚𝑚𝑚 (21)

The bending resistance Wb defines how much bending moment M a beam can withstand before reaching its yield bending stress σs. The greatest bending stress of the chosen material is 355 MPa, leading to the required bending resistance being:

𝜎𝜎𝑇𝑇 = 𝑀𝑀𝑊𝑊𝑏𝑏

⇒𝑊𝑊𝑡𝑡 = 𝑀𝑀𝜎𝜎𝑠𝑠

= 1,457,087355

= 4,104 ≈ 4,100 𝑚𝑚𝑚𝑚3 (22)

The weakest rectangular profile with the desired width 160 mm is a profile measuring 160x80x5 mm which has a bending resistance of 62,300 mm3, giving us a safety factor ns:

𝑡𝑡𝑇𝑇 = 62,3004,104

= 15.18 ≈ 15 (23)

Ftot

27

Figure 24. Schematic picture of total deflection

To calculate the deflection of the tube, the beam deflection equation for a cantilever beam was used:

𝛿𝛿𝑚𝑚𝑎𝑎𝑥𝑥 = 𝐹𝐹𝑀𝑀𝑀𝑀𝑀𝑀∗𝐶𝐶𝐺𝐺𝑇𝑇𝑥𝑥2

6𝐸𝐸𝐸𝐸�3𝐿𝐿𝑅𝑅

2− 𝐶𝐶𝐶𝐶𝐶𝐶𝑥𝑥� (24)

Where E is the Young's modulus for steel and I is the second moment of area found in the data table for the beam.

𝛿𝛿𝑚𝑚𝑎𝑎𝑥𝑥 = 3,068∗507.52

6∗2.1∗105∗2.07∗106�3∗1,900

2− 507.5� = 0.7195 ≈ 0.7 𝑚𝑚𝑚𝑚 (25)

A beam deflection below 1 mm for the given load was accepted by BorgWarner, hence validating the selected frame dimensions.

Steel plate for mounting the frame to the slew drive The interface on the slew drive uses 36 M12 holes 20 mm deep equally spaced around a 785 mm diameter circle. Eleven of these holes are used to attach the mount to the slew drive. The mount itself is made of a 15 mm steel plate that is plasma cut to the right dimensions. Plasma cutting provides decent accuracy and is a low-cost alternative to water jet cutting (ESAB, 2016). To brace the frame, triangular ribs are welded between the frame and the plate.

Slew drive The size of the slew drive was determined by the desired distance of 330 mm between the test table and the rotational axis in the z direction. The selected slew drive is a IMO SP-I0741. Mechanically the slew drive is very over dimensioned. It can withstand a dynamic radial load of 238 kN, a dynamic axial load of 278 kN, a tilting moment of 120 kNm and transfer a torque of 7,800 Nm. Connecting a motor to the slew drive is easy, the interface follows industrial standard, the spur gear acts as a hub with a parallel key connection that fits a 25 mm shaft.

Servomotors After a discussion with colleagues at BorgWarner, the preferred drive for the servo motors was one that was compatible with the software LabVIEW developed by National Instruments (NI). Representative at NI suggested Kollmorgen servo motors.

Finding suitable motors for the rotation requires knowing the rotational speed of the rig. To find this, the longest duration the rig would accelerate with a more or less constant acceleration was considered: from -55 degrees to 0 degrees in 1.5 seconds. With the angles and time known, equation 26 was used to find the acceleration:

�̈�𝜑 = 2𝜑𝜑𝑡𝑡2

=2∗55∗ 𝜋𝜋

1801.52

= 0.853 ≈ 0.9 𝑑𝑑𝑡𝑡𝑑𝑑/𝑑𝑑2 (26)

The speed was found using the equation for rotational speed during constant acceleration:

�̇�𝜑 = �̈�𝜑𝑡𝑡 = 0.853 ∗ 1.5 = 1.28 ≈ 1.3 𝑑𝑑𝑡𝑡𝑑𝑑/𝑑𝑑 (27)

In rpm:

28

�̇�𝜑 = 1.28 𝑜𝑜𝑎𝑎𝑟𝑟∗𝑜𝑜𝑜𝑜𝑡𝑡𝑎𝑎𝑡𝑡i𝑜𝑜𝑎𝑎∗60𝑇𝑇𝑇𝑇∗2𝜋𝜋∗𝑚𝑚𝑅𝑅𝑎𝑎

= 1.28 ∗ 602𝜋𝜋𝑑𝑑𝑟𝑟𝑚𝑚 = 12.2 ≈ 12 𝑑𝑑𝑟𝑟𝑚𝑚 (28)

The servo motors from Kollmorgen typically have a rated maximum speed of between 1200 and 8000 rpm (Kollmorgren, 2014, p. 16). Considering the difference in desired speed and motor speed, and that no motors are available with a torque output matching the torque demand, the servo motors need to be coupled to a gearbox. The next sections will cover the selections of these combinations for both rotational axes.

Rotation around the y axis The y axis rotation is to be provided by a servo motor and a gearbox. The Kollmorgen representative in Helsingborg, SDT, was contacted and provided with the demands of the rig. They suggested the following combination: a Neugart PLN190, i=40:1, planetary gearbox. It handles 1,800 Nm nominal output torque and has an efficiency of 95%. It is matched to a Kollmorgen AKM73M AN C2R-00 servo motor, with nominal speed of 1,500 rpm and nominal torque output of 42 Nm. As insufficient torque stalls the rig, a safety factor Ω of 1.5 was agreed upon. To make sure that the suggested combination is suitable, the specifications of the components were used to calculate the torque and speed demand using the equation for gear ratio:

𝑒𝑒 = �̇�𝜑𝑖𝑖𝑖𝑖�̇�𝜑𝑀𝑀𝑜𝑜𝑀𝑀

= 𝜏𝜏𝑀𝑀𝑜𝑜𝑀𝑀𝜏𝜏𝑖𝑖𝑖𝑖

(29)

With basis in equation 29, the necessary motor speed and torque is calculated:

�̇�𝜑𝑅𝑅𝑎𝑎PLN190 = �̇�𝜑𝑒𝑒PLN190 = 12.2 ∗ 40 = 487.2 ≈ 490 𝑑𝑑𝑟𝑟𝑚𝑚 (30)

𝜏𝜏𝑅𝑅𝑎𝑎PLN190 =𝜏𝜏𝑀𝑀𝑀𝑀𝑀𝑀𝑅𝑅𝐴𝐴𝑦𝑦𝛺𝛺

𝑅𝑅PLN190𝜂𝜂PLN190= 1,000∗1.5

40∗0.95= 39.47 ≈ 40 𝑁𝑁𝑚𝑚 (31)

If a gearbox with a higher gear ratio could be found, it would be possible to use a lower torque motor and still comfortably be within the motors speed range. Unfortunately, when going up to the next higher ratio of 64:1, the nominal output torque of the gearbox is limited to 1,000 Nm (NEUGART, n.d., p. 83).

The motor is equipped with a mechanical brake that holds the table during stationary tests and in the event of a power outage.

Rotation around the x axis The output speed for the rotation around the x axis is the same, but as the slew drive has the gear ratio i =13.73:1, the input speed into the slew drive is:

�̇�𝜑𝑅𝑅𝑎𝑎𝑇𝑇𝑡𝑡𝑡𝑡𝑠𝑠𝑟𝑟𝑜𝑜𝑅𝑅𝑠𝑠𝑡𝑡 = �̇�𝜑𝑒𝑒𝑇𝑇𝑡𝑡𝑡𝑡𝑠𝑠𝑟𝑟𝑜𝑜𝑅𝑅𝑠𝑠𝑡𝑡 = 12.18 ∗ 13.73 = 167.2 ≈ 170 𝑑𝑑𝑟𝑟𝑚𝑚 (32)

The efficiency of the slew drive is 90% according to a representative at Bengtssons maskin AB. The necessary torque input into the slew drive is:

𝜏𝜏𝑅𝑅𝑎𝑎𝑇𝑇𝑡𝑡𝑡𝑡𝑠𝑠𝑟𝑟𝑜𝑜𝑅𝑅𝑠𝑠𝑡𝑡 =𝜏𝜏𝑀𝑀𝑀𝑀𝑀𝑀𝑅𝑅𝐴𝐴𝑥𝑥

𝑅𝑅𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑀𝑀𝑖𝑖𝑠𝑠𝑠𝑠𝜂𝜂𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑀𝑀𝑖𝑖𝑠𝑠𝑠𝑠= 400

13.73∗0.9= 32.37 ≈ 32 𝑁𝑁𝑚𝑚 (33)

The suggested combination for the x axis rotation was this: a Neugart PLQE80 planetary gearbox, i=20:1. It handles 120 Nm nominal output torque and has an efficiency of 95%. It is matched to a Kollmorgen AKM44G AN C2R-00 servo motor, with nominal speed of 4000 rpm and nominal torque output of 5.48 Nm. As with the rotation around the y axis, the necessary motor speed and torque is calculated using a safety factor Ω=1.5:

�̇�𝜑𝑅𝑅𝑎𝑎PLQE80 = �̇�𝜑𝑅𝑅𝑎𝑎𝑇𝑇𝑡𝑡𝑡𝑡𝑠𝑠𝑟𝑟𝑜𝑜𝑅𝑅𝑠𝑠𝑡𝑡𝑒𝑒PLQE80 = 167.2 ∗ 20 = 3,344 ≈ 3,300 𝑑𝑑𝑟𝑟𝑚𝑚 (34)

𝜏𝜏𝑅𝑅𝑎𝑎PLQE80 = 𝜏𝜏𝑖𝑖𝑖𝑖𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑠𝑀𝑀𝑖𝑖𝑠𝑠𝑠𝑠𝛺𝛺

𝑅𝑅𝑃𝑃𝑃𝑃𝑃𝑃𝑃𝑃80𝜂𝜂𝑃𝑃𝑃𝑃𝑃𝑃𝑃𝑃80= 32.35∗1.5

20∗0.95= 2.556 ≈ 2.6 𝑁𝑁𝑚𝑚 (35)

As can be seen, both the gearbox and the motor is over dimensioned for this rotation. A better selection would be the PLQE60 handling a nominal torque output of 44 Nm. With the same gear ratio, an AKM33E

29

would be sufficient, having a nominal speed of 4500 rpm and nominal torque output of 2.8 Nm. In any case, the motor is equipped with a brake for the same reasons as mentioned above.

Connection of slew drive to the stand The slew drive has got two interfaces, one on each side along the x axis, for attaching it to its components. The stationary interface is used for mounting the slew drive to the stand. It consists of 40 holes with a diameter of 14 mm equally spaced along a diameter of 698 mm.

The mounts that attach to these holes have got five holes each, and will be fitted with 5 M12 Allen screws with a washer and a locknut on the back with additional Loctite to stop it from coming loose during the dynamic loads that are occurring. The material chosen for the construction is a S355J2 steel, because of its combination of its decent yield strength and great weldability and price (Tibnor, 2012, p. 11). Due to the complex geometry of these mounts, they are milled and lathed from a solid piece of steel.

Stand Given the fact that the two stands are not part of the rotation and that mass therefore was of no importance, the stand was made sturdy. It is made of two square tubes, with a cross section measuring 120*120*5 mm. It is welded to a base plate measuring 250*250*7 that has got four holes drilled in it for mounting the base plate to the floor. On the top of the stand are square mounts that slots into the tube and are fixed by four M12 screws. These are milled out of steel and differ, as one is used to mount the servo motor and its gearbox tilting the table around the y axis and the other is supporting the bearing, acting as a bearing housing. The side with the servo motor does not need a bearing as the motor itself is capable of handling the loads.

Bearing A spherical roller bearing is selected, to be able to withstand misalignment due to deflection, improper mounting of the mount to the slew drive or poor machining of the components. It is also capable of handling the axial forces that occurs during the rotation around the x axis (SKF, 2013, pp. 48-52).

As this side of the slew drive does not need to handle the torque from the motor rotating the y axis, a thinner shaft can be used to connect the slew drive to the bearing. This is great as a smaller housing can be designed, but also because the weight of the rig might not be enough to exceed the requisite minimum load of a larger bearing (SKF, 2013, p. 86). In a situation where the minimum load is not met, sliding instead of rolling of the rollers in the bearing could occur, drastically increasing friction and reducing its lifetime.

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Figure 25. Bearing housing

The SKF bearing catalogue was used as a reference whilst doing the calculations of the bearing.

Figure 26. Schematic representation of the two stands and the loads on the bearing

To find the maximum equivalent dynamic bearing load, the axial and radial force needs to be calculated.

↑: 2𝑅𝑅𝑚𝑚𝑔𝑔 − 𝑚𝑚𝑑𝑑 = 0 ⇒ 𝑅𝑅𝑚𝑚𝑔𝑔 = 𝑚𝑚𝑔𝑔2

= 400∗9.822

= 1,964 𝑁𝑁 (36)

The only axial force is the force created by the linear acceleration:

→:𝑅𝑅𝑦𝑦 − 𝐹𝐹𝑦𝑦 = 0, 𝐹𝐹𝑦𝑦 = 𝑚𝑚 ∗ 𝑅𝑅𝑅𝑅-𝐶𝐶𝐶𝐶𝐶𝐶 ∗ �̈�𝜑 ⇒ 𝑅𝑅𝑌𝑌 = 𝑚𝑚 ∗ 𝑅𝑅𝑅𝑅-𝐶𝐶𝐶𝐶Z ∗ �̈�𝜑 = 400 ∗ 0.1 ∗ 0.853 = 34 𝑁𝑁 (37)

To find out what equation to use, the relationship between the radial and axial force needs to be calculated.

𝐹𝐹𝑎𝑎𝐹𝐹𝑀𝑀

= 342,066

= 0.017 (38)

The calculation factor e is found in the product tables and has the value 0.22. As Fa/Fr≤e SKF uses this equation to find the maximum equivalent dynamic bearing load:

31

𝑃𝑃 = 𝐹𝐹𝑜𝑜 + 𝑌𝑌1𝐹𝐹𝑎𝑎 = 1,964 + 3 ∗ 34 = 2,066 ≈ 2.1 𝑘𝑘𝑁𝑁 (39)

Where Y1 is the highest calculation factor found for the bearings.

With a dynamic basic load rating of 48 kN for the smallest spherical roller bearing, any of the bearings can be used. But as mentioned above, the minimum equivalent load Pm needs to be considered using the equation from the catalogue:

𝑃𝑃𝑚𝑚 = 0.01𝐶𝐶0 ⇒ 𝐶𝐶0 = 𝑃𝑃𝑚𝑚0.01

= 20660.01

= 206,600 ≈ 207 𝑘𝑘𝑁𝑁 (40)

This means that the selected bearing needs to have a static basic load rating C0 below 207 kN. As the catalogue is referenced, it is found that as long as a bearing with a bore diameter 45 mm or smaller is selected, all criteria will be met.

As the force Rmg is the same force as the force acting to shear the axle off, the minimal diameter of the axle can be calculated, using the safety factor Ω and S355J2 as the material:

𝜏𝜏𝑎𝑎𝑡𝑡𝑡𝑡𝑜𝑜𝑠𝑠𝑎𝑎𝑡𝑡𝑡𝑡𝑡𝑡 = 𝜎𝜎𝑎𝑎𝑡𝑡𝑡𝑡𝑜𝑜𝑠𝑠𝑎𝑎𝑡𝑡𝑡𝑡𝑡𝑡 ∗ 0.6 = RmgΩ𝑅𝑅

⇒ 𝑅𝑅 = 𝑅𝑅𝑚𝑚𝑚𝑚Ω0.6∗𝜎𝜎𝑎𝑎𝑠𝑠𝑠𝑠𝑀𝑀𝑠𝑠𝑎𝑎𝑏𝑏𝑠𝑠𝑠𝑠

= 1,964∗20.6∗355

= 18.44 𝑚𝑚𝑚𝑚 (41)

𝑅𝑅 = 𝜋𝜋𝑟𝑟2

4⇒ 𝑑𝑑 = �4𝑅𝑅

𝜋𝜋= �4∗18.44

𝜋𝜋= 4.294 ≈ 4.3 𝑚𝑚𝑚𝑚 (42)

Given these results, the bearing selected is the 22205/20 E with a bore diameter of 20 mm.

Test table

Figure 27. The test table

The test table will be made of an aluminum plate with milled T-slots along the x axis (see Figure 27). In these slots T-slot nuts will be fitted. The new fixtures will be made with milled grooves in the baseplate going in the perpendicular y direction. This allows for full x and y axis flexibility while mounting the fixture. The aluminum is attached to a 30 mm thick plastic plate made out of polyoxymethylene (POM). This is used for the future climate chamber to be mounted on, providing insulation from the frame. POM has a good combination of material characteristics and price. It can withstand the temperatures achieved during tests (-30° to +140°) while having a low thermal conductivity of 0.31 W/mK.

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Centering fixtures Since the driveshaft rotating the test objects rotates with speeds of up to 10,000 RPM, the shaft alignment is crucial to limit vibrations. The alignment along the x axis in the z direction is done by making sure the fixtures are manufactured to the correct height. The alignment in the y direction is done through the use of a milled keyway in the x direction in both the table and the fixtures.

Figure 28. The keyways

Motor running the test objects The requirements of the motor turning the test objects are these: the maximum speed of the motor needs to be 10,000 rpm, it needs to be able to withstand the dynamic forces induced by the motion of the rig, have a great low speed accuracy and be of a suitable size to the rig. Vascat has an appropriate motor called the MAC HS4 100 S. It meets all the requirements with some modifications recommended by Vascat, namely a reinforced fan structure and additional axial preload of a bearing. It delivers a nominal torque of 38 Nm, while still being compact to not impair the overall length of the rig.

Figure 29. The motor mount

The motor is mounted on a frame made from a plasma cut steel plate that is then bent to its final shape. It has got lathed plastic wheels made of PTFE, due to its low friction, mounted on a steel axle that is press fitted into four holes in the side of the frame. This allows for the position of the motor to be changed along the x axis, allowing for a shorter drive shaft for test objects that do not use the full length of the rig. The motor is fixed by two bolts attaching to two nuts on the opposite side of the frame, providing a clamping force that through friction holds the frame in place during the tests.

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Torque sensor As the motor could be fitted with a torque sensor, Vascat was asked to include a T12 by HBM. They have a version rated at up to 100 Nm, which is more than the peak torque output of the motor. This is a sensor previously used by BorgWarner with good results.

Drive shaft The driveshaft connecting the motor to the test object is designed to not reach its critical speed while still being able to transfer the required torque without shearing. The minimal wall thickness 𝑡𝑡 of a 30 mm diameter tube is calculated based on the peak torque of the motor, 80 Nm, and using a steel tube with a yield strength of 250 MPa:

𝜏𝜏𝑡𝑡𝑜𝑜𝑜𝑜𝑇𝑇𝑅𝑅𝑜𝑜𝑎𝑎 = 0.6 ∗ 𝜎𝜎𝑎𝑎𝑡𝑡𝑡𝑡𝑜𝑜𝑠𝑠𝑎𝑎𝑡𝑡𝑡𝑡𝑡𝑡 , 𝜏𝜏𝑡𝑡𝑜𝑜𝑜𝑜𝑇𝑇𝑅𝑅𝑜𝑜𝑎𝑎 = 𝑇𝑇𝑜𝑜r𝑞𝑞𝑞𝑞𝑡𝑡𝑀𝑀𝑀𝑀𝑀𝑀𝑠𝑠𝑖𝑖𝑀𝑀𝑖𝑖𝐽𝐽𝑀𝑀𝑀𝑀𝑀𝑀𝑠𝑠𝑖𝑖𝑀𝑀𝑖𝑖

, 𝐽𝐽𝑡𝑡𝑜𝑜𝑜𝑜𝑇𝑇𝑅𝑅𝑜𝑜𝑎𝑎 = 𝜋𝜋𝑟𝑟2𝑡𝑡2

(43)

𝑡𝑡 = 2∗𝑇𝑇𝑜𝑜𝑜𝑜𝑞𝑞𝑞𝑞𝑡𝑡𝑀𝑀𝑀𝑀𝑀𝑀𝑠𝑠𝑖𝑖𝑀𝑀𝑖𝑖0.6∗𝜎𝜎𝑎𝑎𝑠𝑠𝑠𝑠𝑀𝑀𝑠𝑠𝑎𝑎𝑏𝑏𝑠𝑠𝑠𝑠∗𝜋𝜋𝑟𝑟2𝑡𝑡

= 2∗80,0000.6∗250∗𝜋𝜋∗302

= 0.377 ≈ 0.4 𝑚𝑚𝑚𝑚 (44)

Where J is the torsion constant. Given the resulting wall thickness is very thin and below any found during research, a thicker wall was desired. Not a lot of sources were found, but a thin wall thickness found was 1.65 mm (Team Tube, n.d.). This was considered a reasonable thickness, and was used for the critical speed calculations.

Figure 30. The drive shaft

With the different geometries of test objects, two drive shafts are needed. One designed to fit the Porsche torque tube in Figure 17 and the other one longer. With the motor placed according to Figure 30, the axle would roughly reach the middle of the table with a length of 45 cm. Since the critical speed decreases with the length of the shaft, the critical speed calculations are made on this longer axle:

𝑡𝑡𝑜𝑜𝑜𝑜𝑅𝑅𝑡𝑡 = 1.22 ∗ 107 ∗�𝐷𝐷2+𝑟𝑟2

𝐿𝐿2= 1.22 ∗ 107 ∗

�32+2.672

45= 24,196 ≈ 24,000 𝑑𝑑𝑟𝑟𝑚𝑚 (45)

Where D is the outer and d the inner diameter in cm and L is the length of the shaft in cm. With a safety factor above two, the critical speed is not an issue.

As there is a risk of not mounting the joints of the axle parallel to each other, it was decided that constant velocity (CV) joints should be used. The alternative would have been cardan joints, but as these produce vibrations and pulsating output speed if they are not parallel, they were deemed a poor choice. The CV-joint also allows some play along the x axis, simplifying the mounting of the axle to the test object. As no CV axels with speed capabilities above 8000 rpm were found, GKN was contacted (GKN, 2014, p. 7). They replied that the critical speed of the axle is not the only important aspect of axle selection, but also the

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lubrication of it at high speeds. The grease is affected by the centripetal forces and is drawn away from the center of the joint. A solution was found in using axles developed for racing, as these are designed for the same speeds as the rig.

The drive shaft is connected at both ends to an adapter that is will be used in most future rigs, making drive shafts interchangeable between different rigs.

Routing the cables As the table rotates around two axes, making sure the cables to the rig run smoothly and without fatiguing over time is therefore challenging. The cables are suppling the three motors with power and control and as well as the rig computer with information from the different sensors. The current solution has the cables just hanging off the edge of the table, but having them rubbing against the rig for an extended time would shorten their life. A cable carrier was therefore deemed the best solution. Available are both single and multi-axis cable carriers. The Swedish distributor of cable carries from igus was asked for a possible solution, and a triflex R energy chain was their recommendation (igus, n.d.). It is a multi-axis chain capable of dealing with the rotations of the rig using only one carrier.

Figure 31. Igus triflex R

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Conclusion and recommendations The concept fulfils the specification requirements developed during the project. With basis in current test orders, the rig will be able to perform all tests. It manages to retain the core features of the current rig, namely the comfortable working height and easy access to the test table, while enabling greater tilting angles and operators to control the rig tilt from the rig computer. The rig table is constructed to have a temperature chamber fitted to it, which will enable tests with different testing temperatures to be run without interruptions.

Simulating in real time One of the biggest drawbacks with the current concept is that real time simulation will not be possible. Dimensioned as today, it simply would not cope with the accelerations involved in real time simulations. But, beef it all up and it wouldn’t be impossible at all. The real question is, and the reason that we settled for a slower acceleration, what do you actually simulate by moving fast? The conclusion is that a rotation of the test subject is not the way to do it. If you want to simulate the cars movement, you cannot substitute real latera or longitudinal acceleration with a simple rotation. The model works great for simulating constant acceleration, but the shift between negative and positive acceleration is a different matter. For this you might want to build a completely different tilt rig, one that can simulate all six degrees of freedom. The most used way of accomplishing this is a Stewart platform. This platform would allow for a more realistic flow of the oil, as it can move along the x, y and z axis creating the actual accelerations experienced in a car. It can only do this with short pulses, as the duration of the acceleration is limited by the length of the rod in the linear actuators controlling the platform. But with a combination of both moving along the axis and rotating around them, it would be possible to initiate the simulation of a change in acceleration by creating movement along the corresponding axis, and then follow up with a slower rotational transition to simulate the potential constant acceleration. There are some obvious disadvantages with this design though, the biggest being that no current platform can achieve more than 41 degrees of inclination. The platforms themselves are also quite bulky, making working around them a hazard and an inconvenience.

With the uncertainties of how to simulate the flow of oil in real time, my recommendation would be investing more time in the theory behind the tilt rig. The test orders of today are rather focused on simulating two specific driving conditions, by holding the rig stationary at different tilt angles:

• Accelerating at a constant acceleration for an extended time. • Driving on a steady slope with a smooth surface.

These are the two scenarios simulated today. In real life conditions, neither would typically be a calm ride. The car would vibrate, shake and move around, thus making the tests performed oversimplifications of reality. In a test environment you of course need to compromise, but it should be of utmost importance that the test preformed have a basis in reality. I would therefore recommend redefining the test orders to allow for a rig with different properties to be developed, for instance a rig with six degrees of freedom to fully be able to simulate all movements of a car. Finding out how the oil actually behaves in the housing during driving should also be assessed.

Motor and gearbox selection The servo motor and gearbox for the rotation around the x axis was over dimensioned. The solution would be excessively expensive and the recommendation would be to use the later suggestion found in the results. However, the CAD-model and quote was not updated with this update due to time restraints.

Climate chamber The rig table is constructed to have a temperature chamber fitted to it, which will enable tests with different testing temperatures to be run without interruptions. This climate chamber was an initial requirement, but no solution was found for supplying the chamber with tempered air from the climate test cabinet. A last-minute meeting with representatives from LaboTest AB, a supplier of climate test cabinets and accessories, turned things around. They had been promised by Vötsch, a large manufacturer of climate test cabinets, that the motion of the tilt rig would not harm the tubes supplying the tempered

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air and that they had used them in similar designs. Because of the late nature of this revelation, it was not incorporated in the design.

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References Anton Paar, n.d. Defining Viscosity. [Online] Available at: http://www.viscopedia.com/basics/defining-viscosity/ [Accessed 3 November 2016].

Bosch Rexroth Group, n.d. 6Dof Motion Platform Technology. [Online].

CFM Schiller GmbH, n.d. Engine Tilting Test Rig. [Online] Available at: http://www.cfm-schiller.de/index.php?zeige_rubrik=30&dbase=produktdetails [Accessed 3 November 2016].

Csere, C., 2011. Tilting Engine Dynamometers Explained. [Online] Available at: http://www.caranddriver.com/features/tilting-dynamometers-explained-tech-department

ESAB, 2016. Vilket är det bästa sättet att skära stålplåt?. [Online] Available at: http://www.esab.se/se/se/education/blog/what-is-the-best-way-to-cut-steel-plate.cfm [Accessed 8 November 2016].

GKN, 2014. Catalogue of CV-shafts. [Online] Available at: http://www.gknservice.com/fileadmin/user_upload/Brochures/English/CV_Katalog_10_2014_web.pdf [Accessed 8 November 2016].

igus, n.d. triflex® R 3D e-chain® specially made for robotic applications. [Online] Available at: http://www.igus.eu/wpck/1783/overview_TriflexR [Accessed 8 November 2016].

Intertek, 2016. Engine Tilt Rig. [Online] Available at: http://www.intertek.com/WorkArea/DownloadAsset.aspx?id=34359738567 [Accessed 3 November 2016].

Kollmorgren, 2014. Kollmorgen AKMTM Servomotor. [Online] Available at: http://www.kollmorgen.com/en-us/products/motors/servo/akm-series/akm-series-ac-synchronous-motors/_literature/akm_selection_guide_en-us_revb.pdf/ [Accessed 8 November 2016].

NEUGART, n.d. Precision gearbox catalog. [Online] Available at: https://www.neugart.com/fileadmin/user_upload/Downloads/Product_Catalogs/Neugart-Product-Catalog-EN.pdf [Accessed 8 November 2016].

Rexroth Bosch Group, n.d. eMotion-1500. [Online] Available at: https://www.boschrexroth.com/en/xc/industries/machinery-applications-and-engineering/motion-simulation-technology/products-and-solutions/6dof-motion-platform/emotion-1500/index [Accessed 3 November 2016].

SKF, 2013. Rolling bearings. [Online] Available at: http://www.skf.com/binary/77-121486/SKF-rolling-bearings-catalogue.pdf [Accessed 8 November 2016].

Statoil , 2007. LSC TRANSMISSION FLUID 301. [Online] Available at: http://www.mekster.se/fileuploader/download/download/?d=0&file=custom%2Fupload%2FFile-1391256172.pdf [Accessed 3 November 2016].

Team Tube, n.d. Driveline Tubing. [Online] Available at: http://www.teamtubellc.com/en/products/round-tubing-products/mechanical-

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tubing/driveline-tubing.aspx [Accessed 8 November 2106].

Tibnor, 2012. Stålvalsguiden. [Online] Available at: http://www.tibnor.no/media/c25cfaba-14d1-4c08-b692-7952108c060c/3YypnQ/pdf/No/St%C3%A5lvalsguiden%20SE.pdf [Accessed 08 November 2016].

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Appendices Appendix A: Market analysis To get an understanding of what systems were currently in use for similar applications a market study was conducted. The possible sources of possible inspiration were broad, ranging from stabilization gear for cameras to tilting machines capable of tilting a full car. The concepts were evaluated according to their potential of fulfilling all the requirements.

Industrial robot An industrial robot is capable of providing the motions needed and more. It would be a versatile solution, possibly capable of performing tests not performed today. The working height of the test table would be very flexible. After discussing the matter with a representative at FANUC, finding a robot that would be capable of performing the motions while still being of a size that would fit in the test cell was deemed impossible.

Two-axis test rig A solution commonly used is based around the concept of having a test table rotating within a rotating frame. The table rotates around the x axis and the frame around the y axis. This concept would allow for the rotational axes to coincide with the mass center, but the frames make the rig bulky and the access to the test table is restricted. Porsche uses this kind of rig for their engine simulations.

(Csere, 2011)

Gimbal for cameras Similar to the two-axis test rig, but by using a L-shaped outer frame three quarters of the outer frame is removed. Although only used on small gimbals used to stabilize cameras, the concept could be scaled to suit the requirements of a tilt rig. With the frame minimized, the rig table would be lot more accessible. The challenge would be finding a viable solution for supporting the loads at the connections.

Stewart platform The Stewart platform is a widely used design for reproducing actual conditions, and would therefore be suitable for simulating different driving scenarios. Using six linear actuators connected between a base a top plate, the actuators combined linear movement provides six degrees of freedom. No products found on the market was capable of producing the necessary tilt, and was deemed unnecessarily complicated for our use. (Bosch Rexroth Group)

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Goniometer The goniometer is worm gear driven unit capable of tilting ±45 on one axis. By placing two units on top of each other in two different sizes, a two axes tilt device with coinciding rotation axes is created. As a concept it could be worth looking into, but the found devices do not allow the weight, torque or inclination BorgWarner needs.

Racing simulator Most racing simulators look compelling as concepts, but none of the simulators today allow for any more than ±41 degrees of tilt simultaneously on both the x and y axis. The concepts are interesting in that the simulations they perform are very similar to what BorgWarner intends to do.

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Appendix B: Questionnaire

Frågeformulär Tiltrigg Namn:________________________________________________

Varför behövs en ny testrigg?

Vilka egenskaper fungerar bra i dagens tiltrigg?

Vilka egenskaper behöver förbättras?

Finns det test som idag inte kan genomföras på grund av bristande funktionalitet i de befintliga riggarna? Om ja: vilka test och vad behövs för att kunna genomföra dem?

Ranka och namnge de fem viktigaste funktionerna/egenskaperna den nya tiltriggen borde ha:

1:

2:

3:

4:

5:

Finns det några nyckelvärden som du anser att riggen måste uppfylla? Exempel: motorn ska klara minst 10000 rpm, riggen måste kunna vinklas 50 grader i alla ledder, motorn måste leverera minst 50 Nm, riggen ska tåla en vikt på minst 100 kg.

Ska riggen kunna överföra moment eller endast mäta ”drag torque” och kontrollera smörjning som idag?

Övriga önskemål?