Design 2012 Cip Frerichs DH

8/13/2019 Design 2012 Cip Frerichs DH

1/116

The University of the Witwatersrand

School of Mechanical, Aeronautical and Industrial Engineering

Design of a High Performance Diesel Engine

i

DESIGN PROJECTFinal year project report

Project title: Design of a High Performance Diesel Engine- Cylinder head

Project supervisor: Prof. D. Cipolat

Date: 22 August 2012

Student: Darryn Frerichs

Student number: 0600945H

In conjunction with: Kurt Crossman (0700001J)

Noordeen Sing (0601701G)

Fuaad Abdool (304124)


2/116




ii

University of the Witwatersrand, Johannesburg

School of Mechanical, Industrial and Aeronautical Engineering

Declaration

Name: Darryn Frerichs Student no: 0600945H

Course no: MECN4005 Course Name: Design Project

Submission Date: 22 August 2012 Project Title: Design of a High Performance Diesel Engine

I hereby declare the following:

I am aware that plagiarism (the use of someone elses work without their permission and/or without

acknowledging the original source) is wrong;

I confirm that the work submitted herewith for assessment in the above course is my own unaided work exceptwhere the I have explicitly indicated otherwise;

This task has not been submitted before, either individually or jointly, for any course requirement, examination

or degree at this or any other tertiary education institution;

I have followed the required conventions in referencing the thoughts and ideas of others;

I understand that the University of the Witwatersrand may take disciplinary action against me if it can be shown

that this task is not my own unaided work or that I have failed to acknowledge the sources of the ideas or words

in my writing in this task.

Signature: ___________________________ Date: _________________


3/116




iii

Executive Summary

It was required that a high performance compression ignition racing engine be designed which

complies to the Federation Internationale de lAutomobile (FIA) World Touring Car

Championship rules and regulations. A production BMW N47 engine was chosen to be

modified in accordance with the regulations with a design goal of achieving a maximum engine

power of 180kW and 400Nm of torque. The design was distributed amongst 3 other individuals

and components designed and modified to achieve the design requirement. The shape, length

and size of the inlet and exhaust ports in the cylinder head were optimized in order to improve

the volumetric efficiency of the engine. Furthermore, the valve train and its components wereanalysed and redesigned in accordance with the regulations, including both dimension and

material specifications. The designed components were modelled and analysed using ANSYS to

evaluate the thermal and stress forces occurring in the cylinder head components during the four

stroke power cycle of the engine. Additionally, analytical analysis was performed on designed

components to verify the ANSYS solutions and to gain insight into the fatigue life of the

designed component. The designed engine theoretically fell marginally short of the design

requirement and theoretically would produce 175kW of power and a maximum of 396Nm of

torque.


4/116




iv

Table of Contents

Declaration __________________________________________________________________ ii

Executive Summary ___________________________________________________________ iii

Table of Contents _____________________________________________________________ iv

List of Figures ________________________________________________________________ ix

List of Tables _________________________________________________________________ xii

Nomenclature _______________________________________________________________ xiii

1 Statement of Task ________________________________________________________ 1

1.1 Task as Given _______________________________________________________________ 1

1.2 Interpretation of task statement _______________________________________________ 1

2 Literature Survey _________________________________________________________ 1

2.1 FIA World Touring Car Championship Rules ______________________________________ 1

2.2 Four-Stroke Cycle Turbocharged Compression Ignition Engines ______________________ 2

2.3 Current motor specifications __________________________________________________ 3

2.4 Inlet and Exhaust Processes ___________________________________________________ 5

2.5 Combustion Chamber ________________________________________________________ 6

2.6 Valve gear _________________________________________________________________ 7


5/116




v

2.7 Valves ____________________________________________________________________ 8

2.7.1 Material __________________________________________________________________________ 8

2.7.2 Valve Guides ______________________________________________________________________ 9

2.7.3 Valve Seats _______________________________________________________________________ 9

2.7.4 Valve face angle ___________________________________________________________________ 9

2.8 Camshafts ________________________________________________________________ 10

2.8.1 Lever tappets ____________________________________________________________________ 12

3 Product Requirement Specification __________________________________________ 13

3.1 General Requirements ______________________________________________________ 13

3.2 General Constraints ________________________________________________________ 13

3.3 General Criteria ____________________________________________________________ 13

3.4 Specific Requirements ______________________________________________________ 13

3.5 Specific Constraints _________________________________________________________ 14

4 Design Development _____________________________________________________ 15

4.1 Concept Development ______________________________________________________ 15

4.1.1 Cylinder head and cover ____________________________________________________________ 15

4.1.2 Combustion chamber ______________________________________________________________ 16

4.1.3 Valve train layout _________________________________________________________________ 18


6/116




vi

4.1.4 Valves __________________________________________________________________________ 19

4.1.5 Camshaft ________________________________________________________________________ 24

4.1.6 Lubrication system ________________________________________________________________ 25

5 Design Development _____________________________________________________ 26

5.1 Valve Size ________________________________________________________________ 26

5.2 Port and Manifold size ______________________________________________________ 28

5.3 Camshaft _________________________________________________________________ 29

5.3.1 Valve lift ________________________________________________________________________ 29

5.3.2 Lever tappets ____________________________________________________________________ 30

5.3.3 Camshaft profile __________________________________________________________________ 31

5.4 Valve springs ______________________________________________________________ 32

6 Performance Prediction ___________________________________________________ 35

6.1 Estimated analytical performance calculations ___________________________________ 35

6.2 Performance curves ________________________________________________________ 36

6.3 Estimated cylinder pressures and temperatures __________________________________ 37

7 Detailed Design Analysis __________________________________________________ 39

7.1 Analytical stress analysis ____________________________________________________ 39

7.1.1 Valve Spring _____________________________________________________________________ 39


7/116




vii

7.1.2 Valves __________________________________________________________________________ 41

7.1.3 Camshaft ________________________________________________________________________ 43

7.2 Finite element heat transfer analysis ___________________________________________ 43

7.2.1 Thermal analysis with combustion modelled as constant heat added ________________________ 47

7.2.2 Heat transfer analysis with variable temperatures applied ________________________________ 53

7.2.3 Stress analysis with temperature _____________________________________________________ 56

8 Assembly Drawings ______________________________________________________ 58

9 Detailed Component Specifications __________________________________________ 63

9.1 Bill of components _________________________________________________________ 63

9.2 Cylinder head _____________________________________________________________ 64

9.3 Camshafts ________________________________________________________________ 65

9.3.1 Lever tappet _____________________________________________________________________ 66

9.4 Valves ___________________________________________________________________ 67

9.4.1 Intake valves _____________________________________________________________________ 67

9.4.2 Exhaust valves ____________________________________________________________________ 67

9.4.3 Valve clearance adjusters ___________________________________________________________ 68

9.4.4 Valve guides _____________________________________________________________________ 68

9.4.5 Valve seats ______________________________________________________________________ 69


8/116




viii

9.4.6 Valve springs _____________________________________________________________________ 69

9.4.7 Valve spring retainers ______________________________________________________________ 70

9.4.8 Valve collets _____________________________________________________________________ 71

10 Recommendations for Future Work _________________________________________ 73

11 References _____________________________________________________________ 75

Appendix A _________________________________________________________________ 78

Appendix B _________________________________________________________________ 79

Appendix C _________________________________________________________________ 87


9/116




ix

List of Figures

Figure 1: Otto cycle plotted on P-V and T-S curves [12] .......................................................... 2

Figure 2: N47 Engine camshaft timing diagram [7] .................................................................. 5

Figure 3: Cylinder Head assembly [3] ....................................................................................... 7

Figure 4: DOHC valve gear layout [13] .................................................................................... 7

Figure 5: Cam cross-section [7] ............................................................................................... 11

Figure 6: Combustion chamber shapes: a) wedge b) hemispherical c) Pent d) flat ................. 17

Figure 7: BMW N47 Valve train layout [7]............................................................................. 18

Figure 8: Plot showing the variation of Z with inlet flow area ................................................ 27

Figure 9: Modified camshaft timing chart ............................................................................... 32

Figure 10: Final predicted engine torque curve [20] ............................................................... 36

Figure 11: Predicted engine power curve [20] ......................................................................... 37

Figure 12: Variation of pressure per degree crankshaft revolution ......................................... 38

Figure 13: Estimated temperature per degree crankshaft revolution ....................................... 38

Figure 14: Exploded view of heat transfer analysis components ............................................ 45

Figure 15: Meshed cylinder head assembly ............................................................................. 46

Figure 16: Temperature fields on cylinder head assembly: view 1 ......................................... 48




10/116




x

Figure 19: Temperature variation on valves and seats ............................................................. 50

Figure 20: Temperature variation of cylinder head with valve components removed ............ 51

Figure 21: Inlet port temperature variation .............................................................................. 51

Figure 22: Exhaust port temperature variation ........................................................................ 51

Figure 23: Heat transfer with increased heat added ................................................................. 52

Figure 24: Heat transfer with increased water jacket heat transfer coefficient ........................ 53

Figure 25: Temperature variation after 0.002963s .................................................................. 54

Figure 26: Temperature variation after 0.00333s .................................................................... 54






Figure 32: Temperature variation after 0.00963s .................................................................... 56

Figure 33: Stress concentration predicted in valves ................................................................ 57

Figure 34: Stress concentrations predicted in valve faces ....................................................... 57

Figure 35: Plan view of cylinder head assembly ..................................................................... 58

Figure 36: Front view of cylinder head assembly .................................................................... 58

Figure 37: Rear view of cylinder head assembly ..................................................................... 59


11/116




xi

Figure 38: Left side view of cylinder head assembly .............................................................. 59

Figure 39: Right side view of cylinder head ............................................................................ 60

Figure 40: Cylinder head assembly.......................................................................................... 60

Figure 41: Exploded view of cylinder head assembly ............................................................. 61

Figure 42: Fully assembled engine .......................................................................................... 62

Figure 43: Variation of cylinder volume with crankshaft angle .............................................. 85

Figure 44: Instantaneous piston speed / mean piston speed variation with crank angle.......... 86


12/116




xii

List of Tables

Table 1: General engine restriction rules ................................................................................... 2

Table 2: Engine variant per model [7] ....................................................................................... 3

Table 3: Technical data of engine [7] ........................................................................................ 4

Table 4: Engine timing data [7] ................................................................................................. 5

Table 5: Weighted matrix selection table of cylinder head cover material ............................. 16

Table 6: Weighted matrix of Combustion chamber shape criteria .......................................... 18

Table 7: Weighted selection matrix table of valve design ....................................................... 20

Table 8: Weighted matrix selection table of valve guide material .......................................... 21

Table 9: Weighted matrix table of valve spring selection ....................................................... 23

Table 10: Calculated cycle state data ....................................................................................... 35

Table 11: Estimated effective pressures, power and efficiencies ............................................ 35

Table 12: Spring data for stress analysis .................................................................................. 39

Table 13: ANSYS mesh setting ............................................................................................... 45

Table 14: Assigned boundary conditions ................................................................................. 47

Table 15: Calculation of valve spring configuration ............................................................... 78

Table 16: Tabulated values of variables used in state calculations ......................................... 79

Table 17: Pressures and temperatures per degree crankshaft revolution ................................. 83


13/116




xiii

Nomenclature

Symbol Description UnitsAp Piston area m

2

Ai Inlet area m2

Mean piston speed m/s

Ci Flow coefficient past poppet valve -

a Speed of sound in medium m/s

L Stroke of engine m

N Speed of engine m

Z Valve index number -

d Diameter mA Area m2

Minimum frequency Hz

k Spring stiffness constant N/m

g Acceleration due to gravity m/s2

W Weight N

Number of turns in spring -

Spring deflection m

Alternating spring force average NMean spring force N

C Spring index -

D Mean coil diameter of spring m

Bergstrasser factor -

Alternating shear stress MPa

Midrange shear stress MPa

Ultimate tensile strength MPa

Ultimate shear strength MPa

Factor of safety -Heat flux between points 1 and2 W/m

2

k Conduction heat transfer coefficient W/m2K

Convection heat transfer coefficient W/m2K

T Temperature K

QHV Heating value of fuel kJ/kmol

R Ideal gas constant kJ/kmolK

A/F Air to fuel ratio -

Compression ratio -

CP Coefficient of pressure kJ/kgK


14/116


15/116




1

1 Statement of Task1.1 Task as Given

Design a high performance compression ignition (CI) engine as a group.

1.2 Interpretation of task statementIn a group of four students, individually design components of a high performance BMW 2l

turbocharged diesel racing engine for use in the Federation Internationale de lAutomobile

(FIA) World Touring Car Championship. Perform an accurate heat transfer analysis on the

components designed and provide a final assembly of all components designed by the group

in an overall design report.

2 Literature Survey2.1 FIA World Touring Car Championship Rules

The Federation Internationale de lAutomobile (FIA)sets out a set of competition rules with

regards to the engines used in the World Touring Car Championship. The engines used mustbe modifications of the standard production engine provided by the manufacturer within the

modification limits set out by the FIA. All modifications performed on the engine must be

homologated by the FIA and also produced in kit variant form for sale to the public. Article

263D from the FIA with regards to the specific regulations for modified diesel engines on

circuits outlines the limitations of the allowed modifications to the engines and a summary of

important regulations is outlined below in Table 1. For a more in depth analysis of the

regulations please consult references [1] and [2], or consult Section 3 for constraints specific

to individual components.


16/116




2

Table 1: General engine restriction rules

Engine restriction 4 Cylinder turbocharged diesel

Capacity 2000cm

Pulleys Free

Valves 4 maximum per cylinder of steel material

Variable valve timing Not allowed

Bore Circular

Compression ratio at least 16:1

Materials

Titanium, ceramics, magnesium, composites or reinforced fibreglass is

forbidden unless corresponding exactly to original parts

Fuel Diesel limited to 50ppm sulphur

2.2 Four-Stroke Cycle Turbocharged Compression Ignition EnginesTurbocharged diesel engine output is usually constrained by the maximum stress levels that

can be withstood by critical engine components as well as the thermal loadings of these

components, which limits the cylinder pressure which can be tolerated under continuous

operation, which is magnified by continuous racing operation. Boost pressure from the

turbocharger increases the pressure difference in the ports of the cylinder which

proportionally increases the stresses and thermal loadings in engine components. Toovercome the increase in stress problem, the compression ratio of the engine may be reduced

as well as the maximum fuel/air equivalence ratio. [3] Figure 1 below shows the otto-cycle

describing the diesel cycle plotted on P-V curve and T-S curves.

Figure 1: Otto cycle plotted on P-V and T-S curves [12]


17/116




3

2.3 Current motor specificationsThe motor being modified is a BMW 2l diesel engine from the modern 120d model. The

engine model is known as the BMW N47 engine from a BMW E87 sedan. Table 2 below

shows an overview of the engine variants of models with the N47 engine.

Table 2: Engine variant per model [7]

M

odell

Modelseries

Cylind

ercapacity

(cm)

Bore/stroke(mm

PowerinkW/bhp

atrpm

Torqu

einNmat

rpm

118d E81 1995 90/84 105/143 4000 300 at 1750

120d E81 1995 90/84 130/177 4000 350 at 1750

118d E87 1995 90/84 105/143 4000 300 at 1750

120d E87 1995 90/84 130/177 4000 350 at 1750

320d E92 1995 90/84 130/177 4000 350 at 1750

Table 3 overleaf shows the technical data of the M47 engine to be modified to meet the

specific rules outlined in Section 2.1 previously.


18/116




4

Table 3: Technical data of engine [7]

Engine type 4 inline

Cylinder capacity (cm) 1995

Stroke/bore (mm) 90/84

Output at engine speed

120/163

4000

Torque (1st gear at engine speed) 280 /2000

Torque remaining at engine speed 320 /2000

Cut-off speed (rpm) 4600

Power output per litre (kW/l) 60.15

Compression ratio 17

Cylinder gap (mm) 91

Valves/cylinder 4

Inlet valve diameter (mm) 27.2

Exhaust valve diameter (mm) 24.6

Main bearing journal diameter of

crankshaft (mm) 60

Big end journal diameter of

crankshaft (mm) 45

Engine management DDE604Emissions standard EURO4

The firing interval of an engine is the angle of crankshaft rotation between two successive

ignitions. The engine operates using a four-stroke cycle and therefore the crankshaft turns

720 to complete one cycle. Having the same firing interval between all ignition points

ensures that the engine runs evenly at all speeds where the firing interval is determined by

dividing the angle required to complete one cycle (720) by the number of cylinders,

producing a firing interval of 180 in the four-cylinder engine. The firing order is the orderin which the cylinders are ignited and is directly responsible for how smooth the engine runs.

The firing order used in the N47 engine is 1-3-4-2. [7]

The camshaft timing is imperative when modifying an engine and all modifications to other

engine systems are complemented by adequate camshaft or valve timing. Figure 2 w shows

the standard valve timing of the N47 engine while Table 5 beneath it shows the timing data

for the engine.


19/116




5

Figure 2: N47 Engine camshaft timing diagram [7]

Table 4: Engine timing data [7]

inlet exhaustValve diameter (mm) 27.2 24.6

Max. valve lift (mm) 7.5 8

Lobe separation (crankshaft) 100 108

Valve opens (crankshaft) 352 140.7

Valve closes (crankshaft) 568 362.5

Valve open

duration (crankshaft) 216 221.8

2.4 Inlet and Exhaust ProcessesThe engine must be constantly supplied with air during its four-stroke cycle, while the

exhaust gases must be expelled. The intake of the fresh air and the expulsion of exhaust

gases are known as the gas exchange cycle and the inlet and exhaust ports are periodically

opened and closed by the inlet and exhaust valves. The timing and sequence of the valve

movements are determined by the camshaft. The entire mechanism for transferring cam lift

to the valve is known as the valve gear, comprised of the camshafts, transmitting elements,

valve assembly, and valve clearance adjustment.


20/116




6

2.5 Combustion ChamberThe combustion chamber is the region located in the cylinder head where the combustion of

the fuel initially takes place when the piston is at top dead centre. The region is a small

volume and the shape can directly affect the quality of the combustion process and therefore

the performance output of the engine and the volumetric efficiency. The chamber also

contains the valves and the fuel injector and affects the compression ratio of the engine

because of its volume.

The combustion chamber shape can help improve the swirl of the fuel/air mixture which will

promote the mixing of the gasses for improved combustion efficiency. The shape of the

chamber will affect the flame propagation as the diesel fuel combusts spontaneously under

compression. A round combustion chamber may be the most advantageous due to the

concentration it provides of the forces generated by the combustion of the fuel onto the top of

the piston, however, this design presents other complications such as the location of the

valves and how a radial valve design would be required to achieve this design which imposes

a complicated camshaft design. The most common design is a triangular shaped design

where the intake valves are located at an angle to the exhaust valves with the injector situated

between the valves. The best shape is one with the shallowest angle between the valves

which allows a high compression ratio to be achieved without introducing high domed

pistons which would affect the flame path. The edges of the combustion chamber should be

flat and shaped and sized similar to the piston edges to produce a region known as squish,

which forces the fuel towards the centre of the combustion chamber with added turbulence

upon compression. The clearance between the wall of the cylinder head and the top of the

piston in a squish area is usually about 0.75mm. The top of the combustion chamber may be

ceramic coated to promote heat transfer should it be required. Figure 3 overleaf shows a

cross section of a cylinder head with a 27 valve angle and also shows the squish areas. [3]


21/116




7

Figure 3: Cylinder Head assembly [3]

2.6 Valve gearThere is a number of valve gear types used in engine design. Valve gear types are usually

distinguished according to the number and position of the valves and camshafts, the method

of actuation of the valves and the method of valve clearance adjustment. The BMW N47engine uses a double overhead camshaft (DOHC) valve gear layout which means that the

engine has overhead valves with two camshafts located above the cylinders where one

camshaft is used for the intake valves, and the other for the exhaust valves. [7] Cam

movement is transferred to the valve by roller cam followers in the DOHC configuration.

Figure 4 below is a schematic of a DOHC valve gear layout.

Figure 4: DOHC valve gear layout [13]


22/116




8

2.7 ValvesThe valves directly affect the volumetric efficiency of the engine as the air must navigate past

the valves both when entering the combustion chamber, and when exiting through the exhaust

ports. Valves which allow an increase in air flow will produce a higher volumetric efficiency

in the engine and an engine that produces more horsepower. Valves with thinner valve stems

provide less restriction to air flowing past it and are lighter allowing a higher revving engine

with less stress on valve components. [5]

2.7.1 MaterialThe valves allow the combustion gases to enter the combustion chamber during the inlet

stroke of the crankshaft and allow the exhaust gases to exit the chamber after the combustion

process has occurred, while providing an air-tight seal in the chamber during the combustion

process to maximise the power obtained during each cycle. The valves are opened and closed

by camshafts at the appropriate time which is discussed below in Section 2.6. The valves can

open and close up to fifty times per second in a diesel engine operation at 6000rpm, and are

therefore a highly stressed item in the engine that incur large fluctuation forces and need to be

designed to achieve a very high fatigue life.

Most performance engines use a good quality grade of stainless steel valves such as 21-4N

stainless steel, however, titanium may also be used when costs are a lesser a factor to be

considered and mass is important. The fatigue life of valves depends largely on their

material, as well as their application; a valve being used in a low rpm situation will most

certainly last longer than a valve used in a formula 1 engine. [4]

Sodium filled valves are hollow valves that are filled with sodium in their stems which

promotes heat transfer and is useful for use in exhaust valves. The sodium in the stem melts

as the valve heats up and moves around, promoting heat transfer from the valve head area, up

the stem, which in turn removes heat from the valve seat area, producing a better seating,

cooler operating valve which lasts longer. [4]


23/116




9

The stems of the valve can be hollowed out to reduce the weight of a steal valve in order to

achieve a similar mass to a titanium valve. Only the upper part of the valve stem may be

removed, where the transverse stresses are minimal. [4]

A cheaper alternative to making a new valve out of a different material is to coat the valve

with a nitride coating or a derivative thereof, using metal treatment, to create a hard coating

which is resistive to wear and also provides lubricating properties to improve the operation of

the valve. [4]

2.7.2 Valve GuidesValve guides in production diesel engines are often made of cast iron and coated with a

phosphate coating whereas bronze guides are the guides of choice in racing applications

because of their high wear, lubrication, and heat transfer properties. Valve guides are

important in the engine as they provide protection to the valve and the cylinder head while

the valve moves up and down, they provide a seal from the top of the head to the ports and

account for approximately 40% of the heat transfer from the valve. The clearance between

the valve and the valve stem is important as it affects the sealing of the valve, and theconduction of heat from the valve to the cylinder head. [4]

2.7.3 Valve SeatsThe valve seats provide a contact area between the valve and the cylinder head. The seats

seal off the combustion chamber from the ports when the valves are closed and transfer heat

away from the valve and into the cylinder head. The valve seats are usually pressed into the

cylinder head and are made from different materials with different thermal conductivity and

hardness properties. Race engines are often built with beryllium-copper alloy or copper-

nickel alloy seats to provide a faster heat transfer medium from the titanium valves. [4]

2.7.4 Valve face angleMulti angle valve jobs improve the airflow through the port and into the combustion chamber

by reducing the separation of the air as it passes over small changes in angle (almost curved)

on the valve face. Multi angle valve jobs are more expensive than single angle and take


24/116




10

longer to produce and while they may not last as long as single angle valves, they provide a

better seal against the valve seat, improved heat transfer to the valve seat from the valve, and

better airflow characteristics. [4]

2.8 CamshaftsCamshafts control the opening and closing of the intake valves with reference to the position

of the crankshaft. The shape of the lobes on the crankshaft determine the rate and time at

which a valve is opened or closed, known as valve duration, as well as the length by which it

is opened, known as the valve lift. Camshafts are specifically designed to achieve maximum

power within a certain range of engine speed and directly affect the fatigue life of valve train

components. The camshaft is the part which connects all the other modified parts of the

engine such as the turbocharger, piston and porting modifications.

Diesel engines require high compression ratios, resulting in the cam duration tending to be

short, with minimal overlap, to prevent the loss of compression because of an open valve as

the compression stroke begins. Valve lift is often limited in diesel engines because of the

tight clearances between the piston and valves as a result of the required high compression

ratio. The effectiveness of the turbocharger is directly related to the characteristics of the

exhaust camshaft which is required to keep the turbo spinning at peak efficiency in the

engines power band. The cylinder pressure is important when developing large, usable

power in an engine and in order to achieve this power the overlap of the valves should be

limited. A camshaft with asymmetrical profiles for the intake and exhaust valves will deliver

a broader and fatter power band. Camshafts with 105 to 110 degree centrelines seem to

produce a broader increase in power. [6]

The diesel engine operates between 2500 and 4500 rpm, and therefore does not reach the high

rpm values of a spark ignition petrol engine. The modifications to the valve train components

do not substantially increase the operating speed of the engine and therefore considerations of

secondary effects such as valve float (where the valve does not follow the profile of the

camshaft directly) are a lot smaller in diesel engines and valve spring stiffness often need not

be changed. [6]


25/116




11

The camshaft profile is described by the cam lift, cam tip, cam shoulder, and the base circle.

Figure 5 below shows a cross-section of a cam and identifies its various elements. The cam

follower follows the action of the cam and transmits it to the valve. The transfer of cam

movement to the valve by the roller cam follower depends on the ratio of the lever lengths.

The cam shoulder may be concave when used in conjunction with roller cam followers. The

N476 engine uses roller cam followers and the cam lift of the exhaust is greater than that of

the inlet, producing greater valve lift [7].

Figure 5: Cam cross-section [7]

The camshaft duration is the number of degrees of rotation on the crankshaft that the valves

are open or off their seats. Phasing is the relationship of the duration of the inlet and exhaust

cycles to each other. Duration and phasing are tied together when discussing camshafts.

Valve lift is the term applied to the maximum distance which the valve travels off its seat.

Camshaft lobe lift is different to valve lift as at varies with the ratio via the rocker. The lift

on the camshaft may be different to that of the valve but the correct amount and rate achieved

because of the rocker ratio. Overlap is the number of degrees of duration which the inlet and

exhaust valves are open simultaneously. The lift rate relates the speed with which the valve

is opened per degree of crankshaft rotation. [11]


26/116


27/116




13

3 Product Requirement Specification3.1 General Requirements

Assembled engine must produce at least 180 kW power

Assembled engine must produce at least 400Nm torque

3.2 General Constraints A minimum compression ratio of 16:1 must be used

Engine must contain four cylinders

Engine volumetric capacity must be no more than 2000 cm3

A maximum of four valves per cylinder can be used

A maximum boost pressure of 2.5 bar must be used for the intake manifold of the

engine

3.3 General Criteria Maximum power should be achieved at highest rpm

Maximum torque should be achieved at lowest rpm

Maximum volumetric efficiency should be achieved

Components should be designed for maximum durability and reliability

The mass of the engine should be minimised as much as possible

Engine should be as compact as possible

Cost restrictions may be neglected

3.4 Specific Requirements The original cylinder head must form the basis of the modified head.

Intake and exhaust ports and combustion chamber must be analysed and redesigned

with modifications by machining of standard cylinder head.

The minimum height of the cylinder must not be reduced by more than 2mm.

The number of valves per combustion chamber may not change.

The engine must not exceed four cylinders with an overall capacity of 2 litres.


28/116




14

The combination of the assembly must produce the maximum possible power from

the engine throughout all RPM, limited to 4500rpm.

The camshaft must be redesigned to attain the desired maximum power by changing

timing and lobe shape but not changing bearings on the shaft and number of shafts.

Valve stroke must not exceed 10mm

Valve seats and valve guides must be redesigned including materials.

Valve shape, size and length must be redesigned.

Valve springs must be redesigned but must be steel.

Upper section of the valve collar must have a stop collar with a diameter greater than

0.5mm.

Overall compression ratio must be greater than 16:1.

A thorough heat analysis must be done on all components and the assembly.

3.5 Specific Constraints All components must be designed to be made available to the public in kit form and

must be homologated by the FIA.

The cylinder head may have two threaded connections of maximum size M-14

machined in to it to improve cooling.

The head gasket is free but may not exceed 5mm thickness.

Titanium, ceramics, magnesium, and composite materials may not be used unless they

are the original material of the component but all chemical and heat treatment is

permitted.

All standard components may be ground, filed, or re-machined to achieve desired

modification.


29/116




15

4 Design Development4.1 Concept Development

4.1.1 Cylinder head and coverThe design of the cylinder head influences the output efficiency, fuel consumption, torque,

exhaust emissions and noise characteristics of the engine. The shape of the cylinder head

depends on the components accommodated by it. Components accommodated by the

cylinder head are; the valves, camshafts, glow plugs, injectors, and the inlet and exhaust

ducts. Furthermore, the cylinder head should be as compact as possible and depends on the

size, number and shape of each component it accommodates, the method of injection, and the

method of cooling to be used in the engine. The cylinder head of the standard N47 engine

comprises of two large cast parts with the camshafts integrated inside their own camshaft

carrier. The standard cylinder head configuration will be used as per the requirements set out

in Section 3.4 previously, where only modifications to the standard head may be performed

by grinding or welding.

The cylinder head cover covers the valves and camshafts and the top of the cylinder head in

the DOHC configuration. The cover seals the top of the cylinder head from the outside and

provides sound insulation between the noisy camshaft mechanism and the outside of the

engine. The N47 product information manual (reference 7) says that the cylinder head cover

of the 47 engine also performs the following tasks (verbatim):

Retention of the blow-by gas exhaust line from the crankcase, of the oil separation

system and of the pressure regulating valve of the crankcase breather

Retention of the fuel system rails

Retention of the camshaft sensor

Retention of the oil filler neck

Retention of line feed-throughs.

The cylinder head cover does not directly influence the performance output of the engine,

however, the cover does add mass to the engine and must provide adequate sound insulation

as well as seal off the top of the engine from the outside environment to keep the oil inside of


30/116




16

the engine and unwanted dust and water outside the engine. Elastomer seals can be used to

seal the join between the cylinder head and the cover, while a weighted matrix is used to

determine an adequate material to be used to construct the cover. Aluminium and plastic

materials are viable options for the construction of the cover and the criteria used to

determine the most suitable material were; sound insulation, weight, heat resistance, and

manufacturability. The design is to be used in high performance application where high

temperatures will be reached, and weight is a key issue with regards to performance,

therefore, these two criteria are assigned a larger weighting in the selection process than the

sound insulation and manufacturability. Table 5 below shows the weighted matrix table for

the cylinder head cover material selection.

Table 5: Weighted matrix selection table of cylinder head cover material

Criteria Weighting Material

Aluminium Plastic

Weight 40% 3 4

Heat resistance 30% 3 2

Manufacturability 15% 0.5 1.5

Sound insulation 15% 1 1.5

TOTAL 10 7.5 9

It was decided that plastic would be used to make the cylinder head cover due to its score of 9

out of ten in Table 5 above.

4.1.2 Combustion chamberThe combustion chamber is the space bounded by the cylinder head, the piston, and the sides

of the cylinder. The cylinder head forms the ceiling of the combustion chamber and its shapeaffects the combustion efficiency and the characteristics of the engine by affecting the

mixture preparation prior to combustion. The combination of the cylinder head and the

geometry of the piston determine the overall shape of the combustion chamber.

Shape of combustion chamber

The roof of the combustion chamber can take the form of many different shapes and sizes.

As per the requirements set out in Section 3.4 previously, the standard flat cylinder head can


31/116




17

be modified by grinding or welding. Four different combustion chamber shapes are

considered for the modified cylinder head, namely; hemispherical, pent, flat, and wedge. The

different shapes are shown below in Figure 6.

Figure 6: Combustion chamber shapes: a) wedge b) hemispherical c) Pent d) flat

The designs are evaluated against three criteria with different weightings to select the mostsuitable combustion chamber shape for the modified high performance cylinder head. The

size of the valve used directly affects the performance of the engine, discussed in detail later,

where in general, the larger the valve, the better the performance. Therefore the valve size

criteria are given the largest weighting. The standard production head is a cross-flow head,

where the inlet and exhaust valves are located on opposite sides to one another which

improves cooling and efficiency of the engine in comparison to counter-flow heads. The

straightness of the ports directly affects the power of the engine as the air entering or exiting

the combustion chamber has is obstructed less by a strait port where it need not negotiate a

turn twist and incur losses due to the changing of direction of the air. A formula 1 car, for

example, has an almost strait port as the size of the cylinder is not as an important

consideration as the power developed by the engine. The final criteria considered is the

manufacturability of the cylinder head and the size occupied by its construction. The final

criterion was allocated the smallest weighting as the design is for a high performance

specification. Table 6 overleaf shows the weighted matrix table used to select the shape of

the combustion chamber.


32/116


33/116


34/116


35/116


36/116


37/116




23

Table 9: Weighted matrix table of valve spring selection

Criteria Weighting Material

Standard single Dual Varying diameter

Weight 30% 3 1.5 2

Accuracy 35% 3 2 1.5

Stiffness 15% 1 1.5 1

Cost 10% 1 0 0.5

Simplicity 10% 1 0.5 0

TOTAL 10 9 5.5 5

Standard, single valve springs of suitable stiffness and material were chosen to be used in the

design as opposed to duel of varying diameter valve spring configurations.

Valve lash or clearance adjuster

The valves must be able to close properly under all engine operating conditions to prevent

loss of power due to a loss of compression or combustion pressure and to dissipate heat

generated through the cylinder head and coolant. Heat transfer is essential in compression

ignition engines because if an exhaust valve does not seal, hot combustion gases may flow at

high velocities through the narrow gap resulting in extremely high localised temperatures on

the exhaust valves which may cause pre-ignition of the diesel fuel. The two most popular

ways of controlling the clearance between the valves and the rocker arms to control the

perfect sealing of the valves at all times are by a hydraulic valve clearance adjuster systems,

or by a fixed, mechanical clearance adjuster system. The lever tappet rests on the valve

clearance adjuster which controls the clearance between the tappet and the valve. The

hydraulic clearance adjustment system maintains the valve clearance at zero under all

operating conditions and manual adjustments are easily and efficiently made, even after

substantial engine service time. The manual system involves manually inserting a shim

between the tappet and the lever to control the valve clearance. A manual system weighs less

than a hydraulic system and is more reliable, however, the hydraulic system can make fine

adjustments and all operating conditions and adjustments are efficiently performed. All

current BMW models use a hydraulic valve adjuster system and the system has proven to be


38/116




24

robust and reliable and therefore the same hydraulic adjuster system shall be used in the high

performance application [7].

4.1.5 CamshaftIn order to make the valve gear as rigid as possible the engine will make use of an overhead

camshaft layout to decrease the length of the linkage between the camshaft and the valves.

The engine will have separate camshafts for the exhaust and inlet valves respectively. The

main component of the camshaft is the cylindrical hollow shaft with the cams, or lobes,

arranged around it. The actuation forces are braced by camshaft bearings which are directly

mounted onto the camshaft tube in the high performance application. The surface of the

camshaft is to be ground at the mounting points of the bearings and an oil bore in the bearing

point will provide the necessary lubrication. In addition to the camshaft bearings on the

shaft, a thrust bearing will limit axial float of the camshaft. The gear of the inlet camshaft

should be easily adjustable in order to change the performance characteristics of the engine to

suite different racing circuits. A camshaft sensor will be installed in the gear of the inlet

camshaft as per the standard N47 engine specification [7].

Camshaft manufacture technique

Two variations of the camshaft manufacturing are considered in the design of the high

performance engine; forging, and composite constructions. Composite camshafts are made

by manufacturing the shaft tube, cams, and other functional elements such as drive gear

separately and joining them afterwards. A composite camshaft can weigh up to 40% less

than a forged or moulded shaft which in turn reduces engine fuel consumption, improves

vibrational and noise characteristics and allows the possibility of weight saving on other

system components such as the valve gear mechanism [7]. Furthermore, composite camshaft

manufacturing permits the possibility of combining different materials in the component and

a more economic production of the component. The advantages of a composite camshaft far

outweigh the disadvantages and a composite camshaft construction will therefore be used in

the high performance engine design.


39/116




25

There are many different methods by which composite camshafts can be made. Firstly, there

is the classic positively interlocking or frictional shaft and collar joint by which individual

components are attached to the shaft, secondly, components can be fixed by thermal shrink

fitting, or the shaft widened to obtain a frictional joint. Thirdly, the components may be fixed

to the shaft by welding or soldering. Finally, the tube can be widened by rolling, and given

pitchless thread at the relevant position intended for the seat of a cam or other components;

the appropriate part is then pressed on at the desired angle. The bore of the part being

pressed on has a longitudinal profile and this creates a positive and negative connection

between the shaft and the pressed part. The last method is known as the Presta method and is

a robust method that does not allow parts to come loose. The camshaft undergoes thousands

of stress and thermal cycles during operation and the most robust shaft must be used when

producing the high performance component, therefore, the Presta method shall be used to

produce the composite camshaft.

Lever tappets

Lever tappets are not bearing mounted on a shaft; they rest one end directly against the valve

clearance adjuster and the other in the valve stem, while the cam presses on the centre of the

lever tappet from above. The inertia and rigidity of the tappet depends heavily on its design,

where short levers make for low mass moments of inertia, and tappets made from sheet metal

rather than cast are much lighter. Roller cam followers reduce the internal friction of an

engine by following the action of the cam by a roller running on needle bearings and are a

must in a high performance application as a reduction of frictional losses will increase

performance, especially in the lower rev band.

4.1.6 Lubrication systemThe design and optimisation of the cylinder heads lubrication system was not conducted in

this document. It was assumed that the standard lubrication system in the N47 engine would

provide adequate lubrication and protection to the designed components. Any modelling of

the lubrication system was not performed in the design.


40/116




26

5 Design Development5.1 Valve Size

The volumetric efficiency of an engine is a quantifiable indicator of the performance of the

engine. An engine with a high volumetric efficiency is vital in a high performance engine.

The inlet and exhaust valve size affects the volumetric efficiency of an engine and the

optimal size needs to be calculated.

The inlet valve size affects the volumetric efficiency and power output of an engine more

than the exhaust valve as the exhaust valves are, in simple terms, used to expel gases from the

combustion chamber once they have been used to produce power. An inlet Mach number

index Z is formed from an average gas velocity through the inlet valve and is given by:

(1)

whereandare the piston and inlet areas respectively, is defined as the mean pistonspeed and

is the flow coefficient past the poppet valve while

is the speed of sound in the

medium in the inlet port [8]. Taylors correlations have been used to show that the

volumetric efficiency is drastically decreased beyond an index of 0.5 due to separation and

frictional effects passed the cylinder head components, furthermore, in high performance

racing applications, indices in the region of 0.4 are preferable. The calculation of the

discharge coefficient is a complicated process as it depends on the ratio of valve lift to

diameter as well as the seat angle and width [8]. John Heywood (reference 8) says that at full

open throttle when valve lift is at its maximum an approximation of =0.45 is sufficient. is calculated using the following formula:

(2)Where is the stroke of the engine in meters and is the operational speed of the engine inrev/second.

A plot was compiled of inlet Mach number indices for different flow inlet areas, while engine

speed is maintained at a maximum of 4500rpm and the piston area and flow coefficient are


41/116




27

maintained constant with reference to a bore of 81.5mm and a fully opened poppet valve.

The plot showing the variation of Z with flow area is shown below in Figure 8.

Figure 8: Plot showing the variation of Z with inlet flow area

From Figure 8 above, a Z index number of 0.4 requires an inlet flow area of approximately

0.0011mwhich corresponds to two inlet valves of diameter of approximately 27mm. From

Table 3 of Section 2.3 it is noted that the inlet valve diameter of a standard engine is 27.2mm

which corresponds to a Z index number of 0.38 which will be adequate in producing a high

volumetric efficiency engine.

The ratio of exhaust to inlet valve diameters lies between 0.85 and 0.92 for multivalve racing

engines [9]. The standard exhaust valve size is 24.6mm as shown in Table 3 of Section 2.3.

The use of this 24.6mm exhaust valve results in a ratio of 0.904 which is well within the

acceptable limits.

As discussed in Section 2.8.4, multi angle valve jobs would be beneficial in the performance

of a racing engine. A 5-angle valve job was selected for the high performance engine

0

0.1

0.2

0.3

0.4

0.5

0.6

0.0007 0.0009 0.0011 0.0013 0.0015 0.0017 0.0019 0.0021 0.0023

Z

index

Inlet flow area


42/116




28

application with the middle angle being the preferred 45. The angles chosen were 75, 60,

45, 30, 15 degrees respectively to promote a smooth transition. The 75, 60, and 45 degree

angles were cut into the valve seat while the 30 and 15 degree angles were cut into the

cylinder head for a perfect seal and to allow the use of the multi angle valve job.

5.2 Port and Manifold sizeThe dimensions of the intake and exhaust system play a significant role in the volumetric

efficiency of an engine. The high performance diesel engine is to be designed to operate over

a small rpm range between 2500 and 4500rpm, producing maximum power at maximum

speed where it would operate for the majority of a race.

When the intake and exhaust valves open and close pressure waves occur within the cylinder

ports. The pressure waves move through the ports and it is said that a port is tuned when

the moving waves can be used to force additional air into the combustion chamber prior to

the valves closing. The different piston speeds will cause different pressure wave phenomena

to occur and an engine can therefore only be tuned to produce the maximum power and

torque over a small rpm range. The standard cylinder head is to be modified and therefore it

was determined that the standard port lengths be used, with adjustments made to the lengths

by adjusting the lengths of the intake and exhaust manifolds, designed separately.

The throat diameters behind the valve in normal engines are usually in the range of 0.8-0.85

times the valve diameter [8]. In racing applications, experimental data has shown that throat

diameters of 0.86-0.89 times the valve diameter produce the largest power from the engine

[10]. The required throat diameters were calculated as follows:

The ports can be any shape; square or circular variation. The valves are round and therefore

the valve ports need to be round, because of this, the port shape was designed to be elliptical

to reduce flow disturbance regions and regions of high carbon build up in the exhaust ports in

small radius areas. The area of the port before it splits into the two ports respectively is


43/116




29

needed to be slightly larger than the sum of the areas of the individual ports. The areas of the

inlet and exhaust port are calculated below:

The dimensions of the ports were calculated using the equation for an oval where a is the

height and b is the width as shown below:

A venture shape was added to the inlet ports above the throat area to improve the distribution

and velocity of flow around the valve. The restriction incurred by the smaller bore is

overridden by the improvements gained by the flow improvements. The diameter of the

venture was calculated as 85% of the inlet valve diameter [12]:

Further optimization of port shape and size should be performed using a computational flow

dynamics analysis package (CFD).

5.3 Camshaft5.3.1 Valve lift

In most high performance applications the valve is required to be open for as long as possible

at its maximum lift to allow the maximum possible amount of air to pass through it in each

cycle. Production camshafts are designed to generate large amounts of power without adding

additional, unrequired stresses to the valve train components. Racing or high performance

camshafts do not take this into too much consideration as components such as valves and

valve springs can be replaced after each race. The performance camshaft was designed to lift

the valve as quickly as possible within the limitations of the valve gear design and close it as

quickly as possible, leaving it open for as long as possible. The lift rate was slowed down as


44/116


45/116




31

5.3.3 Camshaft profile

The diesel engine was designed to produce maximum power at only 4500rpm which isconsiderably less than that of a petrol engine. The camshaft modification possibilities and

inlet and exhaust processes were discussed in Sections 2.4 and 2.9. The standard engine

produces maximum power at about 3500rpm. The inlet valve opening angle (IVO) was

increased from 8 degrees before top dead centre (BTDC) to 12 degrees to allow more air to

be forced into the combustion chamber when the engine operates at a higher rpm, improving

volumetric efficiency, although performance at idle speed and initial acceleration will

decrease.

At high engine speeds, the air traveling into the combustion chamber, boosted by the

turbocharger, has a high moment of inertia, therefore, it is possible for air to continue

entering the chamber even after the piston has moved passed bottom dead centre (BDC) and

begins travelling upwards. The decrease in the compression ratio to 16:1 and the increase in

boost pressure allow the inlet valve closing angle to be increased from 28 degrees after

bottom dead centre (ABDC) to 31 degrees ABDC.

The changes to the inlet valve opening and closing angles change the inlet cam duration from

216 degrees to 219 degrees compared to the standard configuration.

The pumping work done by the engine during the exhaust stroke is decreased by decreasing

the pressures in the cylinder to as close to atmospheric as possible at BDC. The exhaust

valve opening angle was chosen to remain at 39.3 degrees before BDC as with the standard

camshaft configuration and an adjustment was made to the exhaust valve closing angle to

increase the duration of the cam and the valve overlap, improving the scavenging

characteristics of the engine at high speed operation. The exhaust valve closing angle was

increased marginally from 2.5 degrees after TDC to 3.5 degrees ATDC, as explained in

Section 2.4 previously.

The duration of the exhaust cam was increased to 222.8 degrees which is more than the inlet

cam as the exhaust gasses are under less pressure than the inlet gasses because of the

turbocharger.


46/116




32

The profile of the camshaft lobe was selected to produce maximum power by opening the

valve as quickly as possible and keeping it open as long as possible. The most rapid lift rate

occurs between 30 and 80% of the total valve lift. In the first 30% of a valves lift the valve

was accelerated from being stationary up to the highest rate and then slowed down in the last

20% of the maximum lift to maintain the control over the movement of the valve as the

camshaft reaches maximum lift. There is almost no lift produced 10 degrees of camshaft

movement on either side of the maximum lift position of the camshaft. The shape of the cam

was chosen to be symmetrical to achieve the same phenomena on closing the valve as

achieved when opening. [11]

The resulting camshaft timing is shown below in Figure 9. In comparison to Figure 2 of

Section 2.3, it can be noted that the modified design is more aggressive in terms of the speed

of the valve opening, as well as the valve overlap and durations of the valves.

Figure 9: Modified camshaft timing chart

5.4 Valve springsThe valve springs control the movement of the valves by maintaining a tensional force

ensuring that the movement of the valve follows the profile of the camshaft at all engine

speeds. They are subjected to high temperatures and cyclic loadings as the valve is opened

by the camshaft and closed by the tensional force of the spring. Fatigue life and spring

stiffness are therefore important properties to be considered when designing the valve spring.

The spring designed must operate within its elastic limits as over compression of the spring

will result in plastic deformation and unforeseen properties of the spring with undesired

operation. The frequency of the compression of the spring must not be close to the springs

0

2

4

6

8

10

-360 -260 -160 -60 40 140 240 340

V

a

l

v

e

l

i

f

t

Crankshaft Degrees

Int


47/116




33

natural frequency as the oscillations may become uncontrolled and undesirable, resulting in

undesired operation of the spring and premature failure.

The camshaft operates at half of the speed of the crankshaft and the valve spring is excited at

the same frequency of the camshaft. The camshaft therefore operates at a speed of 2250rpm

when the engine is at its maximum. The natural frequency of the spring was designed to be

at least 4 times the normal operating frequency of the spring:

The minimum frequency value of the spring was used to calculate the minimum stiffness to

weight ratio of the spring to determine the minimum spring stiffness required [14]:

The valve springs were required to be made of steel allow material as outline in Section 3.4.

Chrome Vanadium springs were chosen as they are used in applications where shock loads,

high stresses and elevated temperatures are predicted. The shear modulus and specific weight

of chrome vanadium springs are 77200MPa and 82g/cm3 respectively [15]. The spring

stiffness and weight ratios for different coil and wire diameter configurations were calculated

and an appropriate configuration was chosen, using the spring stiffness to weight ratio as a

reference. Table A-1 of Appendix A shows the comparison of different spring

configurations. A 3mm wire diameter and 16mm mean coil diameter configuration produced

a spring of 38.167N/mm stiffness with a stiffness to weight ratio of 261.1mm -1 which far

exceeds minimum stiffness to weight ratio determined and therefore was chosen as the spring

configuration for both inlet and exhaust valves. The length of the valve spring was calculated

taking into account the solid spring height for a valve spring [14] and the maximum

deflection that the spring incurs because of the maximum valve lift:


48/116




34

The above calculated spring length is an absolute minimum and does not account for

compressions incurred because of spring retainers or larger deflections incurred due to

unforeseen conditions. The spring length was therefore determined to be 40mm to ensure the

spring operates within its elastic deflection limits, taking into account all components in the

cylinder head.


49/116




35

6 Performance Prediction6.1 Estimated analytical performance calculations

In order to gain an estimate of temperatures and pressures after each process in the engine

cycle and to determine expected engine power output values the combustion process was

modelled as an air standard diesel cycle as explained in [8]. The cycle was greatly simplified

with idealizing assumptions made, such as the injection of fuel and combustion being a

constant pressure process, and therefore the calculations were used as indications of actual

values. The calculated conditions at the inlet, compression, injection and combustion, and

exhaust states are tabulated below in Table 10. For detailed calculations refer to Appendix B.

Table 10: Calculated cycle state data

State

Pressure

(kPa)

Temperature

(K)

1 350 374

2 16976.01 1133.75

3 16976.01 2453.63

4 1017.94 1098.08

The constant and calculated values in Tables 10 and 12 in Appendix B were used to

determine work, power, mean effective pressures and efficiencies using equations from [8].

Table 11 below shows the calculated values, detailed calculations are contained in Appendix

B.

Table 11: Estimated effective pressures, power and efficiencies

Gross work per cycle 152.55kW

Indicated mean effective pressure (IMEP) 2810.5kPa

Brake mean effective pressure (BMEP) 2529.36kPa

Maximum Output power 175.86kW

Specific fuel consumption 0.00256

Fuel conversion efficiency 83.15%

Volumetric efficiency 268.07%


50/116




36

6.2 Performance curvesThe final performance curves were calculated by Kurt Crossman [20] as they are heavily

dependent on the functionality of the turbocharger. The turbocharger was set by [20] to begin

boosting at 2000rpm and reach maximum boost at 2500rpm corresponding to the

turbocharger setting and properties.

The figures calculated in [20] are reproduced below in Figures 10 and 11. The engine

produced maximum power of 175kW at maximum engine speed with a torque of 396Nm at

2500rpm. The torque and power requirements outlined in Section 3.1 are higher than the

values attained with the modifications but were deemed acceptable as they were within 5% of

the requirements.

Figure 10: Final predicted engine torque curve [20]

0

50

100

150

200

250

300

350

400

450

500 1000 1500 2000 2500 3000 3500 4000 4500

Tourq

ue(N.m)

Engine Speed (rpm)


51/116




37

Figure 11: Predicted engine power curve [20]

6.3 Estimated cylinder pressures and temperaturesThe pressures and temperatures within the cylinder were approximated using the calculated

conditions at each state during the combustion process. The calculated values are tabulated

in Appendix B and sample calculations shown. It was determined that 180 crankshaft angle

would be bottom dead centre where inlet conditions are at state 1, 360 is top dead centre and

540 bottom dead centre. Temperatures and pressures were not calculated during the exhaust

and inlet strokes. The ignition delay time was calculated as 59.9765x10 -6(See Appendix B)

and the addition of heat from combustion was added after top dead centre, correlating to this

value. Figures 12 and 13 below show the estimated variation of pressure and temperature per

degree of crankshaft revolution respectively. The figures generated follow similar trends to

text book data in [8].

0

20

40

60

80

100

120

140160

180

200

500 1000 1500 2000 2500 3000 3500 4000 4500

Power(Kw)

Engine Speed (rpm)


52/116


53/116




39

7 Detailed Design Analysis7.1 Analytical stress analysis

7.1.1 Valve SpringThe valve spring undergoes thousands of cyclic loadings during a race weekend. At

maximum engine speed the valve spring is undergoing about 37.5 cycles per second and is

directly influencing the power produced by the engine by controlling the valve movement.

The valve spring was designed to be as lightweight as possible as explained in Section 5.4

and it was decided that the valve spring service life could be far less than infinite and will

need to be replaced after almost each race weekend. The factor of safety of the fatigue

loading of the valve spring was calculated using both the torsional Gerber and torsional

Goodman fatigue failure criterion with Zimmerli data. For a fatigue life that is not infinite,

the factor of safety was deemed to be satisfactory if greater than 1. Table 12 below shows the

data used for the calculation.

Table 12: Spring data for stress analysis

Operational Frequency 37.5Hz

Deflection 2-10.2mm

Uncompressed length 40mm

Wire diameter 3mm

Mean coil diameter 16mm

Minimum force applied 76.334N

Maximum force applied 389.3N

Spring constant 38.167N/mm

All equations used below are found in [14].

Spring index:

Bergstrasser factor:


54/116




40

Alternating shear stress component:

Midrange shear stress:

From Table 10-4 in [14]: m=0.263 and A=2065Mpa.mmmfor stainless steel springs.

Ultimate tensile strength:

Shearing ultimate strength: Load line slope:

For peened springs from [14], Ssu=398MPa and Ssm=534MPa

Gerber ordinate intercept for Zimmerli data:

Amplitude component of strength:

Gerber fatigue factor of safety:

Goodman order intercept:

Amplitude component of strength: Goodman fatigue factor of safety:

The two factors of safety of 1.324 and 1.25 indicate that the design of the spring is suitable

for the spring not to fail during an event if replaced regularly as the fatigue life was not

designed to be infinite.


55/116




41

7.1.2 Valves

As outlined in Section 3.2, the valves are restricted to being made of stainless steel. Thevalves were designed to be nitride coated to improve the hardness of the surface of the valve

and to reduce the friction between the valve and the valve guide. The groove area of the

valve where the collets support the valve was shot peened to improve the strength in this high

stress area.

The estimated expected maximum stress in the valve was determined by making assumptions

that the pressure force from combustion applies a uniform load to the valve and that the valve

rests equally in the valve seat with a contact length of 2mm throughout. The predicted stress

was calculated as:

for the inlet valve, and 69.57MPa for the exhaust valve

The stress concentration factor due to the collet grooves was calculated as follows [14]:

Taking into account fatigue loading [14]:

Endurance limit [14]:

The material selected was chosen by researching different performance valve types and

manufacturing processing. The chosen material has the following properties:


56/116




42

Type: Stainless steel (nitride hardened)

Tensile strength: 1030MPa

Yield strength: 760MPa

The factors were calculated using the material properties [14]:

The valve stems were predicted to operate in the region of 400C, especially in the exhaust,due to the 3000W of heat added in each combustion cycle, therefore the temperature factor

chosen was [14]:

A 95% reliability factor was selected as the components wold be regularly replaced in the

application: [14]

The endurance strength of the valve was then calculated as:

Fatigue factor of safety [14]:

The factor of safety calculated is above 1 and in line with that calculated for the valve springs

and therefore the material chosen and valve configuration is sufficient.


57/116




43

7.1.3 Camshaft

The dynamic loads experienced by the camshaft are complicated and fluctuating due to thevarying torsional and direct force loads applied by the rocker arms and the sprocket on the

camshaft. The calculation of the stresses in the camshaft below was highly estimated and

more than likely vastly exaggerated as forces were predicted using educated guesses. The

torsional force applied by the gear was assumed to be 50Nm. Moments induced by the

camshaft lobes were calculated using the rocker ratio to be 494N in the exhaust cam, which

experiences the larges forces, which were approximated to 500N. The diameter of the

camshaft was limited to 23.4mm in order to utilise the standard journal bearings and

lubrication system on the standard cylinder head. The camshaft diameter is 1.42 times larger

at the exhaust lobe than on its journal diameter which induces a stress concentration factor.

The notch sensitivity factor was estimated as follows for EN40B material [14]:

Therefore the factor of safety factor was calculated as follows [14]:

The safety factor calculated is fairly low but as mentioned above, forces in the camshaft were

approximated and exaggerated and a more in depth finite element analysis should be

performed on the camshaft to verify the forces estimated.

7.2 Finite element heat transfer analysisThe peak temperatures of combustion gasses inside a cylinder of a diesel engine are in the

region of 2500K, but the melting point of the aluminium composite cylinder head is far below

that. Therefore, the maximum temperatures of the metal surfaces enclosing the combustion

chamber must be limited to much cooler values by cooling the cylinder, cylinder head, and

piston to ensure trouble free operation of the engine. The rapid changes in temperature lead


58/116




44

to substantial thermal stresses in the cylinder head and its components which are increased

further by the application of the large pressure gradients induced by the compression and

combustion of the fuel mixture. The cylinder head is one of the most complicated parts of the

combustion engine as it houses valves, seats, intake and exhaust ports, water jackets, injectors

and glow plugs, as well as being directly exposed to the high pressures and temperatures of

combustion. When designing the cylinder head, the operation of all components must be

considered and often compromises are made with certain components to allow the desired

operation of another.

A finite element heat transfer analysis was conducted to provide information on the

temperature distribution in the overall assembly of the cylinder head. Problematic areas and

areas of interest where extreme thermal loadings are experienced were highlighted for a more

accurate examination. The valves are directly exposed to the peak temperature combustion

and the valve se

Design 2012 Cip Frerichs DH

Documents