6&o[070m > INSTITUTIONEN FOR VARME- OCH KRAFTTEKNIK FORBRiiNNINGSMOTORER LUNDS TEKNISKA HOGSKOLA I iliiism.1 ~
Design of a single cylinder optical access to the combustion engine Scania D12
Juergen Fuchs
ISRN/LUTMDN/TMVK – 5340 – SE
DMSION OF COMBUSTION ENGINES , 2000 DEPARTMENT OF HEAT AND POWER ENGINEERING LUND INSTITUTE OF TECHNOLOGY “J?iJlo!lOFmsIXICU!JENis UN!.;,” P.O. BOX 118, SE-221 00 LUND - FORQGNW3 FRWIB~ SWEDEN DISCLAIMER
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BengtJohansson Forfattare
JuergenFuchs JAN292001 Dokumenttitel och undertitel
Designof a singlecylinder optical access to the combustion engineScania D12
Referat (samrnandrag) In this work a maximumopticalaccessto a diesel engineis developed.The combustion-processin the engineshouldbe representativeto the one in a standardengine,so the geometryof the combustion chamberis modifiedas little as possible.A Scaniasingle cylinder,2-litreenginewas subjectedto modificationsallowingthe opticalaccess. Solutionsto theseproblems me obtainedby using the methodof Product-Development,mainlybased on the literatureby Prof. Dr.-Ing.Birkhoferat the TechnicalUniversityof Darmstadt,Germany. An opticalenginedesign of the Bowditchtype was the chosenmain workingprinciple.This engine containsan extendedcylinder,partlymade of glass, a glass piston-crownand a mirrorplacedinside the extendedpiston.The laser sheetis led into the combustionchamberthroughthe glasspart of the cylinder,then gets reflectedinsidethe combustionchamberandis led throughthe glasspiston crown and via the mirrorout of the engine. A redesignof the valve-train,using extendedpush-rods,is necessary.The demandto examinethe combustionat Top-Dead-Centre(TDC)and the necessityof supportingthe glass, give the reasonsto do workon the cylinderhead. This in returnbrings sealingproblems,whichhavebeen solved.Another problemthat occurswith that typeof engineis that is has to run without oil-lubrication.Piston rings madeof Rylonare used to solvethis problem.A specialfeatureof the enginethat has been constructed hereis that the inner surfaceof the glass maybe cleanedwithoutremovingthe cylinderhead.This is obtainedby a constructionwith a movablecylinder.In cleaning-statethe cylinderis drivenup and downtogetherwith the piston,whilethe headis supportedby an outer [email protected] engine,the cylinderis fixed to the structure. Furthermorethis report containsthe necessarycalculationsandintegity assessmentson the critical parts of the construction.All calculations,exceptthe deformationof the glass-platethat has been solvedwith the Finite ElementMethod,are done analytically.A maximumpressureof 200 bar and a wall temperatureof 300K are assumed.Such high loads requirea heavy construction,whichof course is wantedto be avoided here.So in order to handlethese loads,high qualitymaterialssuch as titanium for the piston-crownor high-qualitysteel have beenused.
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2 Abstract
In this work a maximum optical access to a diesel engine is developed. The combustion-process in the engine should be representative to the one in a standard engine, so the geometry of the combustion chamber is modified as little as possible. A Scania single cylinder, 2-litre engine was subjected to modifications allowing the optical access. Solutions to these problems are obtained by using the method of Product- Development, mainly based on the literature by Prof. Dr.-Ing. Birkhofer at the Technical University of Darmstadt, Germany. An optical engine design of the Bowditch type was the chosen main working principle. This engine contains an extended cylinder, partly made of glass, a glass piston-crown and a minor placed inside the extended piston. The laser sheet is led into the combustion chamber through the glass part of the cylinder, then gets reflected inside the combustion chamber and is led through the glass piston crown and via the mirror out of the engine. A redesign of the valve-train, using extended push-rods, is necessary. The demand to examine the combustion at Top-Dead-Centre (TDC) and the necessity of supporting the glass, give the reasons to do work on the cylinder head. This in return brings sealing problems, which have been solved. Another problem that occurs with that type of engine is that is has to run without oil-lubrication. Piston rings made of Rylon are used to solve this problem. A special feature of the engine that has been constructed here is that the inner surface of the glass maybe cleaned without removing the cylinder head. This is obtained by a construction with a movable cylinder. In cleaning-state the cylinder is driven up and down together with the piston, while the head is supported by an outer structure. When running the engine, the cylinder is fixed to the structure. Furthermore this report contains the necessary calculations and integrity assessments on the critical parts of the construction. All calculations, except the deformation of the glass-plate that has been solved with the Finite Element Method, are done analytically. A maximum pressure of 200 bar and a wall temperature of 300K are assumed. Such high loads require a heavy construction, which of course is wanted to be avoided here. So in order to handle these loads, high quality materials such as titanium for the piston-crown or high-quality steel have been used.
3 4 Acknowledgements
This Master-Thesis was written at the Division of Combustion Engines at the Department of Heat and Power Engineering at Lund Institute of Technology, Lund’s University. There are a number of people I would like to thank for help during this thesis. First of all I would like to thank my supervisor Bengt Johansson, Ph.D – Asc. Prof. and Anders Hultqvist for all their help and for organizing the project for me. I also want to thank Prof. Dr.-Ing. Birkhofer and Dr. Kloberdanz at the Technical University of Darmstadt for their good co-operation, taking care of the project and acknowledging it. You made it possible for me to get a very interesting and attractive project and to use my time as an exchange student meaningfidly! I want to thank my parents who always, during my studies and especially during my time as exchange student, supported me with all their help. Thanks to all the people at LTH who helped me, especially to Goran Haraldsson (I think he was the one who had to stand most of my questions), Olof Erlandsson for continuous help with Matlab programming and Bertil Andersson for a lot of tips and experience! Thanks to all the other people at the division that I have harassed with persistent questioning. You all saved mea lot of time and headache. A very big “Thank You” to all people I met for always helping me with the Swedish language and to Ralf Rank for the English-corrections of this Thesis. Last but not least I have to thank all the people of the Erasmus/Sokrates programme who offered the great possibility of an exchange-year in Sweden.
Lund, November 2000
Jiirgen Fuchs
5 >.—
6
— Table of Contents 1. k~oduction ...... 9 1.1. BACKGROW ...... 9 1.2. PURPOSE...... 9 1.3. h&THOD...... 10 2. Product.Development.Stiate~...... ll 2.1. W SOLUTIONCONC= ...... ll 2.1.1. Starting with the cylinder head ...... 13 2.1.2. Optical access through the piston ...... l5 2.1.3. Optical access through the cylinder wall...... l7 2.1.4. Combination of the different solutions ...... 2O 2.2. EXAMININGTHEDETAILPROBLEMS...... 22 2.2.1. Steering of the valves ...... 23 2.2.1.1. Examining theoriginaldesign...... 23 2.2.1.2. Discussionof thefoundsolutions...... 33 2.2.2. Fixing of the glass-windows, sealing, cooling, oil-supply, design as module; access through the qlinder.wall ...... 36 2.2.2.1. Designof theglass-holding-modules...... 36 2.2.2.2. The Glms.Wg.Modtie ...... 36 Functionsof thefig.modde ...... 37 2.2.2.3. Changingof themodtdes@ water-sealing-problem...... 47 2.2.2.4. ‘l’he Glms.Pti.Modde ...... 49 2.2.2.5. Designof theglass-piston-crown...... 5l 2.2.2.6. Functionalanalysisof thecylinderextension...... 60 2.2.2.7. Re-designof solution3)...... 69 2.2.3. Remaining Problems: ...... 74 2.2.3.1. Extendedpistonconnection ...... 74 2,2.3.2. Problem Oil maybe pressedup fromthecrankcaseandsoil themirroK...... 74 2.3. ~A-MtiYsB ...... 75 2.3.1. Association ofjimctions to the elements of the device ...... 75 2.3.2. Risk-Analysis ...... 76 2.3.3. Cause.Analysis ...... 77 2.3.4. Consequence.Analysis ...... 8O 2.3.5. Risk Rating and Summary ...... 83 3. Forces,LoadsAndWays ...... 86 3.1. DIMENSIONSANDwIG~s ...... 86 3.2. FORCESONTHEVALVE-TRAINANDTHEPUSH-RODS:...... 88 3.2.1. Loads on the valves: ...... 88 3.2.2. Acceleration-force of the valve-rockers ...... 92 3.2.3. Loads on thepush-rods: ...... 94 3.2.4. Deformation of thepush.rods: ...... 96 3.2.5. Buckling of the push rods: ...... 100 3.3. COARSECALCULATIONSONTHEGLASS~G: ...... 101 3.4. CALCULATIONSONTHEMIRROR...... 103 3.5. CALCULATIONSONTHESEALING...... 104 3.5.1. Sealing of the piston...... 104 3.5.2. Sealing of the glass.ting: ...... ll4 3.5.3. Sealing of the upper-plate: ...... ll5 3.6. LOADSON= G~SS.PETON.CRO~ ...... 116 3.6.1. Acceleration of thepiston: ...... ll6 3.6.2. Combustion-loads on the piston and the glass-plate inside the piston ...... 118 3.6.3. Combustion.courses: ...... ll9 3.6.4. Stresses on the glass-plate inside the piston: ...... 127 3.7. STRUCTURALINTEGRITYASSESSMENTFORTHETITANIUMPISTON-CROWN:...... 129 3.7.1. Structural integrity assessment for dynamic.loads: ...... 133 No specialtreatmentof thesurfaceis donehereso:...... 134 3.7.2. Structural integrity assessment for stationary-loads: ...... 136 3.8. LOADSONTHEPISTON-EXTENSION...... 137 3.9. STRUCTURALINTEGRITYASSESSMENTSONSCREWS:...... 140
7 ——
3.9.1. Screws@ing thepiston.crown: ...... 140 397 N...... 145 3.9.2. Screws&ing the glass.module: ...... l5O 3.9.3. Plate, taking the counter-force of the ring: ...... l57 4. References ...... l63 AppendixA MountingMmud ...... 164 AppendixB: Listof Demands...... 170 AppendixC MatlabPrograms...... 174 AppendixD: ~atigs ...... 187
8 1.Introduction
1.1. Background
The fuel efficiency of the diesel combustion engine makes it a frequent power source. Heavy-duty applications especially like heavy trucks are in use of diesel engines. Unfortunately the combustion process of the diesel engine creates several unwanted and harmful emissions. In the pas~ the emission levels have been strongly restricted and the future will require even lower emission levels. In order to fhlfiill these levels, the manufacturers of diesel engines have to concentrate on improvement and research of today’s engines. An alternative approach for a combustion engine with very low emissions and a working coefficient like the diesel engine is to run engines in HCCI, abbreviation for Homogeneous-Charge-Compression-Ignition. This new, just developing combustion process combines the advantage of a spark ignited (S1, Otto) engine that has very low emissions, with the very good fuel efficiency of a compression ignited (CI, Diesel) engine. In an HCCI engine, the fuel and air are premixed in the intake system just like in an ordinary S1 engine. The compression-ratio of the engine is raised to a higher level so that self-ignition is reached. The premixed aidfuel mixture provides homogeneous combustion gas inside the cylinder. That means the aidfuel-ratio A is distributed equally inside the cylinder. Combustion under such circumstances occurs with very low soot emissions. Soot is formed when the aidfuel ratio is too low at some points inside the combustion chamber and the combustion is not complete. This is the big problem of every diesel engine today! In order to avoid high NOX-ernissions, which occur at a high combustion temperature, the aidfuel ratio is even raised to about 3. The surplus air cools the combustion below the critical temperature where NO. is formed. Research into the combustion process requires that different properties are measured from the inside of the combustion chamber. One successful way to accomplish this is the use of laser-based technology. This technology however is in need of an optical access to the combustion chamber. To construct such an access is the main-purpose of this work.
1.2. Purpose
The purpose is to construct an optical access to a one-cylinder diesel combustion engine. The optical access should supply as large field of view as possible and allow laser-based measuring methods of the combustion process inside the engine. It is a major aim to examine the combustion under “normal” conditions, so the shape of the combustion chamber and of intake- and exhaust-system should be changed as little as possible. Examination of the operating conditions when running experiments, a construction allowing very easy and fast cleaning of the internal glass surface is required. All demands and requirements me shown in the List of Demands in the appendix. 1.3. Method
There are many options in constructing an optical access to a combustion engine. The biggest field of view, of course, is achieved by building the entire engine of glass. Since this is because of mechanical problems not possible, a compromise between a mechanically solvable design and the optical demands has to be found. In order to find the best solution for the task, the construction process has been done methodically, using Product Development Strategy. The strategy for the Product Development is based mainly on the literature of Prof. Dr.-Ing. Birkhofer from the Technical University of Darmstadt, Germany. It is described in detail in the section Product-Development-Strategy.
10 2. Product-Development-Strategy
This Thesis is about the construction of an optical access to a single cylinder diesel engine. Such engines have been built previously, using different methods of optical access. With the help of Product Development different possibilities of optical access have been developed and rated. The methodical way is divided into two parts. The first part was to find a main solution-concept of how to obtain optical access to the cylinder. The second part was to find solutions for the “detail-problems” that occurred for the chosen main solution-concept. The strategy of the Product Development mainly is based on the study literature of Prof. Dr.-Ing. Birkhofer [7] at the Technical University of Darmstadt, Germany. In order to get abetter overview over the whole report, the List of Demands, which of course was one of the first things during the solution-process is added to the appendix.
2.1. Main Solution concept
The first step completed was to decide a main solution-concept for the device. Here I want to decide how the principle working-sticture of the whole device should look like. Based on the instructor’s past experience, a design with an extended piston is favoured. Since there are in principle, as we know from the literature-search, not that many possibilities given the main functions @ solutions are developed in a fast way (Develop the fimctional-analysis only as far as necessary without senseless systematic). Possible solutions-principles are shown and compared to each other in this part. The next figure gives an overview over the main function of the device and the input and output quantities. _——-.
Main functions
Laser-lightthat Laserin, fi-om containsinformation light source F Lead laser-light into the combustion- aboutflow field and Movement combustioninside F chamber, so that as much as possible from piston thecylinder from the combustion process is viewable Heatfrom combustio~ Loadsfrom combusho~
00
Divides up into partial-functions:
The main taskof thedevice is to Ieadthelaser sheet,so thismainfunctionis examinedmore clearly
Leadthelight Lightgetsreflectedinside Laserin- intothe combustion- the chamberwith an angle light source chamber side chamber of 90 degrees H --H1’ 1 , 1 I I
L Leadthereflected light out :ide of the combustion-chamber ~ber - I 1 1
Figure 1: The main function of the opticaI device The movement of the piston and mechanical loads, as shown in Figure 1, also influence the device, but they differ with different building-principles. They are taken as criteria in choosing a suitable working principle from the found solutions and of course later to define dimensions of elements. So they are not divided up more here.
Taking a look at the structure for leading the laser sheet, three partial fimctions have to be fulfilled. It is obvious that the reflection of the light inside the chamber cannot be influenced. So principle solutions for leading the light into and out of the chamber are required. The problem can now be formulated as leading light into a cylinder. Examining the surfaces of the cylinder it can be said that in a very simplified view this can only happen through 3 areas. These are via: . The top of the cylinder (adjacent to the cylinder head in a combustion engine), ● The cylinder-walls (which are formed by the engine-block), . The bottom of the cylinder (that is the piston).
12 Drawing 1: Possible ways of how to view into the cylinder
2.1.1. Starting with the cylinder head Set out in the List of Demands is the use of a standard cylinder head in order to have a similar flow-field as in the original engine. This determines the building space for an optical access to a minimum.
. The use of lenses through the cylinder head requires replacing at least one valve in order to make space. This changes the flow field and is therefore out of the question. . Another way of viewing the flow field is to remove the entire cylinder head and to make a construction where the valves for intake and exhaust are standing or lying at the sides. This solution will of course give the biggest field of view of the combustion inside the engine. Unfortunately this also changes the flow field, so powerfidly that there will bean entirely different one. Therefore the use of a standard cylinder head is required in the List Of Demands. So this solution isn’t suitable either. Constructions with standing or lying valves at the sides have been made earlier in this century. They have been used until the fifties. Taking a look at them it’s obvious, that
13 —— .
the field of flow is not comparable with one where the valves are located over the piston inside the cylinder head. Examples are shown in Drawing 2. . The last possibility is to use a small endoscope that doesn’t need much building space. Using this, the flow-field isn’t disturbed and it’s possible to use a standard cylinder head. This is the only way of having optical access through the cylinder head while keeping the flow field more or less identical. Unfortunately the field of view is only very small, using this device. The possibility of using an endoscope is given in the List Of Demands as a wish, but it is not included in the main task!
m
optical W2&ss by Replace the vahx replacing cum Or and build an& more valveg and insert nmv cylinder head lenses into thent. with Strmiing or laying valves at tie Problem: tides. Then the flow field changes bm can be viewed ‘= not mitak~ whle througha moremx only gkss pllltq at the limited field d’ top, view PrcMems: this srkm v~ good the demandfm a big field oiTT&v.but changm the entire flow field. = not suitable
Figure 2: Solutions for leading the laser sheet through the cylinder head
Summarizing the solutions for optical access through the cylinder head it can be said that the only suitable solution is to use an endoscope. Since the field of view with this device is very limited, it can only be an added feature to the whole construction.
14 Zylinderdeckelmitstehenden Ventilen
ZylinderdeckelmitIiegendenVentilen Drawing 2: Different types of valve arrangement that allow optical access through the cylinder head. (Source: [9])
2.1.2. Optical access through the piston The piston forms the opposite site of the cylinder head. There are no moving parts like the valves at the surface of the piston, so the whole area can be used for optical access. Inside the piston the connecting rod is joined to lead the force from the combustion to the crankshaft. Making the piston crown of glass the light can be led out of the combustion chamber and into the piston. It is now possible to extend the distance between the piston crown and the bearing for the connecting rod so that a mirror can be added in between them. The light is then reflected to the side inside the piston. So far there is still the problem with au oscillating mirror inside the engine block. Moreover the engine block cannot be changed. A solution for this is, to also extend the cylinder. It can be made longer than the engine block allowing access from the side by leaving a hole under the combustion chamber, but still over the original engine block. Another advantage is that it’s then easily possible to fix the mirror from the outer side, so that it is not oscillating with the piston. This is a very costly construction with extensive changes within the engine. But it gives the ability to view the combustion process without disturbing it.
15 —— ——.
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Drawing 3: Working principle of an optical engine used at Dainder-Chrysler, (Source: [11])
16 Picture 1 shows the extended piston (K), the mirror (C) and the glass piston-crown (J) of an optical engine with extended piston. The mirror is fixed outside the cylinder, so that it is not oscillating. The working principle is obvious. Drawing 3 shows the schematic principle at an engine that it is used by the Dairnler- Chrysler Company. In this drawing light source and some measuring devices are also included. A principle difference from the engine studied here is that the Dairnler-Chrysler engine uses two camshafts lying over the cylinder head to steer the valves directly. The engine to be built uses one camshaft inside the engine block. Here push rods are used as contact between the valves and the camshaft.
2.1.3. Optical access through the cylinder wall Examining the combustion through the piston, the whole horizontal plane is viewable. The vertical examination is only possible through the cylinder wall. Moreover this is the biggest area where the combustion can be examined. There are no moving parts inside the wall that you have to take care of. So far it is very suitable for an optical access. The cylinder wall is part of the whole engine block. This has to withstand the highest mechanical loads and, as said while examinin g the possibility to use the piston as optical access, can not be changed. Extending the cylinder and the piston also can solve this problem. Here a construction with windows can be added between the cylinder head and the engine block, forming an extended cylinder. This construction is not as costly and extensive as access through the piston.
Taking a look at the shape of the windows there are two principle designs possible:
. Option 1: Straight formed windows. Taking them there are again 2 possibilities 1) build an entire square formed cylinder of glass @ square piston This gives very good optical access to the combustion chamber. The light can pass straight through the glass and is not distorted. The big disadvantage is that a sufficient sealing between piston and cylinder in the comers is very difficult to handle. Another disadvantage is that the flow field obtained differs from that in a normally shaped combustion chamber. A standard cylinder head cannot be used. @ this solution is not suitable 2) Small glass parts are inserted in a cylinder that has the same shape as the original. This has also the advantage that the light can pass through the glass without disturbance. The disadvantage is that only a very small field of view is possible. Another problem is, that there is always a small gap between piston and cylinder where the windows are (piston is round and windows are straight). This influences the flow field and the compression ratio. @ suitible solution, but with restrictions
. The other way is to build a window that has the shape of the combustion chamber. This involves using a glass-ring as a cylinder. The big advantages here are that the flow field is the same as in the original engine and that a large field of view is possible. Problems are the loads on the ring and the lubrication between piston and cylinder (to have abetter view of the combustion, no oil maybe used). Another disadvantage is that special lenses have to be used to compensate for the distortion of the light passing through the round glass.
17 —.. . .
9 this solution gives a big field of view together with the same flow field inside the cylinder @ most suitable
lead the sheet through the cylinder-walls
Solutions by examining the shape of the access
Window has the same form as cylinder => round
~ + A Use the combustion cham- Use a glass ring as cylinder: Build the entire combustion ber thatis in existence and chamber of class: (square- add windows atthe siales: Advantwes: 2?Z4! - the shape of the combus- Advantages: tion chamber is not Advantages: - only small glass parts are changed. - very big field of view used, so the loads of the => exact same flow field - no lenses for compensa- combustion are easier to as in the original tion are necessq handle engine! -” cheap”, because of the - very big field of view Problems: small glass parts into the engine - the glass windows have to - no 1ens systems for standthe high loads of the compensation are Problems: combustion necessary - the glass ring has to => pressure, temp. standthe high loads of - another shape of the com- Problems: the combustion process bustion chamber m cans - only avery small field of => pressure,temp. another flow field view is accessible through - no oil-lubrication be- - no suitable sealing these windows tween piston and between chamber andpis- - since the windows are glass ring is possible ton in the corners of the straightthereis always a => special sealing chamber possible small gap between windov rings required - standardcylinder head car and piston. This influence: - the glass ring acts as a not be used the flow field and the com lens, so the image is pression ratio. The last onf distorted. can be adjustedwith an => o~er lenses me extended piston. needed to compensate this effect
Figure 3: Solutions for leading the laser sheet through the cylinder walls Summarizing the solutions for optical access through the cylinder wall one can say that the two options “small straight glass parts” and “glass ring” are suitable. According to the main task of getting an optical access that is as big as possible it is the decision to use a glass ring for viewing the combustion from the side. Another argument is that a
18 construction with straight glass parts has already been built here at the department, a progression is desired.
A schematical drawing for access with a glass ring is given in Drawing 4. Here a Spark Ignition Engine is used!
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Drawing 4: Optical access through the piston, using a glass ring. (source: [4]) 2.1.4. Combination of the different solutions
Since there are only a few solutions remaining, not many combinations are possible. The endoscope-possibility is formulated as a wish from the instructor to be added later. It’s not included in the task. So there are only the solutions for access through the cylinder-wall and the piston remaining. Both solutions are in need of an extended piston; so a combination of them is possible (the glass piston nevertheless needs an inlet window, so that the laser sheet can be reflected with an angle of 90 degrees and caught again * List of demands). It’s the decision to use both possibilities, both the access through the piston and through the cylinder wall. For the access through the cylinder wall the solutions with the glass ring and small glass parts are possible. The glass ring gives the best field of view, so the decision is that this design should be built. On the other hand an optical engine with small glass parts as (the only) optical access, is in existence here at the institution and these parts can be used again. Because of the fact that the manufacturing costs for the steel-housing for the glass parts are relative low compared to the costs of the glass, this design should also be built. Doing so, the resources that are in existence are used and the expensive glass parts can be used for both engines. The task is now to build two optical “modules” for access through the cylinder wall, which can be easily changed. That means that the same connections to the extended cylinder have to be used. For the design with the glass parts, the fixing to the cylinder head should be done by using the same screws, which are normally used to fix the cylinder head to the engine. Another advantage is that in the case of window failure the measurements at the engine can be continued with the other modules.
Drawing 5 gives a basic overview of the chosen main working principle with the glass ring:
20 Head
N
I // / .,, ., ... . d
Bearing for
Drawing 5: chosen main working principle
21 .- —. .—— —-——.
2.2. Examining the detail problems
The main working structure is given now in Drawing 5. The next step is to examine detail problems and to find solutions for them. Figure 4 gives an overview over the detail- problems that are going to be examined. It shows that the solutions of some problems may influence several parts. Therefore the problems are ordered in groups and for every group solutions have to be found. i 1 I Detail Problems I
Detail Problems are: - sealing of the combustion chamber - steering of the valves - fixing of the glass windows=> easy to remove - build access through cylinder wall as modules - fixing of the whole device=> fast to build up . cooling of the cylinder - connection of the extended p iston to the connecting rod - oil-supplyh-eturn for the cylinder head - oil-return, for the case, that oil is pressed up from the engine-block beside the extended piston
These Problems can be structured into parts that influence each other:
- fining of the glass . oil-return for oil that windows can be pressed up - sealing of the from the ena@ne - steering of the valv combustion chamb ti - fixing of the - oil return for - cooling for the whoh whole device the cylinder head device - cooling of the whole . access through cylin- device der wall as module - oil supply for the cy- linder head connection of the extended piston to the bearing of the connecting rod ~
Figure 4: Examining detail-problems
22
— — 2.2.1. Steering of the valves
In the original engine the steering of the valves is done by push rods. They lead the force for opening the valves from the camshaft inside the engine to the valve-rockers on the cylinder head. In the basic working principle piston and cylinder are extended so that a new design has to be done. The function of steering the valves can be expressed by creating a linear force to open and to close the valves. This force has to be timed correctly to the combustion-process.
Energy open and close the valves at the right time according to the combustion process. valve- => createa lkear force on the valve-shaft in two movement -’l directions Signal from => steering signal has to steer the force crank-shaft--’i r Figure 5 function: steering the valves
2.2.1.1. Examining the original design Partial functions are found by examinin g the original design where push rods are used:
V* rocker
cam foHo Jer
Drawing 6: Design of the original engine using push rods
23 .——. —...
Element Function sortof fi.mction 1.) Original cam-shaft inside the ● create linear force Main- emzine ● create steerhw simal for the valves +funct. 2.) Spring ● create force to press the valve up again Main- (reset-force) funct. ● store reset-energy 3.) Push-rods . Lead the force/movement to press the Main- valves down funct. ● lead signal from cam-shaft
4.) Cam follower ● lead force/movement and signal to Sec. push rod Func. @ of no interest here 7 5.) Valve rocker . Lead the force/movement Sec. ● Change direction of force/movement Funct. 6.) Bearing of valve rocker ● Join rocker 7Sec. fimct. Table 1: Functions of steering the valves Table 1 shows the functions of steering the valves. They are divided into main- and secondary-functions. Main and secondary here is related to the function “valve steering”.
New solutions can be found by systematic variation of the physical effect. The results are shown in solution trees. In order not to get to many unnecessary solutions only the main- functions are examined.
Create force directly
rot.: - pneutnl. -pneu. -from -fi’amcam - hydr. -hydr. -dectr. -linear engine cylinder crank-shaft cylinder engine ewe ~~e -from cam- -electi. -pneu. shaft w ““ :$ ma~et cyiinder cylinder double double
Figure 6: Solution tree: create force directly
In Figure 6 the solutions for creating the force are shown. Only the suitable ones are getting inserted into the morphological box. Pneumatic devices cannot be used. The force on the valves from the combustion is too high and no compressible medium is suitable.
24
.- Create force for valve return from energy storage
-rotational -linearspring -pessure tenk -pressuretank -capacitor spring -massforee => turkine => cylinder -coil - massforce withrope added to wheel - force from momentof inertia Figure 7: Solution tree: create force from energy storage
Create steering signal
mech. ha electr. K<’ ‘~
/\ /\ /\ & transl.: rot.: transl - rot.: transl.: -eccentricmods -llFErle --campresses -zt@--ng- ~electr. diskon on cam-shatl createsflow m cylinder motarcon- engine crenk/cem relative => liquid flOW trolledby (ma~etdtnd.) d-aftwitha to fpm outof cylinder computer- controlledby connecting equalscam- signals computer- rod form sQnals -cam and camfollower m cm shaft
Figure 8: Solution tree: create steering signal Lead fm-ce/ movement for valve steerinjg
electr.
rot.: transl.: comb.: T —— _ -Gearbox -Pushrods -ccmnectingrod -oil Wlspm’t -currentflow -seesaw(VAppe) -chain t.hrou*pipes ibroughwires -belt -light-bow (absur~
Figure 9: Solution tree: lead force/movement
Solutions for the main functions are found in Figure 6, Figure 7, Figure 8 and Figure 9. In order to get a better overview which combinations are possible they are combined in a morphological box.
As shown in Figure 4 the design also influences the oil return. Of course the oil return should be designed as simple as possible. But since this is no major task and it is not examined in detail, only a short look at it is taken. In the original design the housing of the push rods is used for oil-return. This has the advantage, that no extra tubes are needed. Consequently the installation of the optical device goes faster and it can be expected to be arranged more clearly. There are now two ways of designing the oil return: ● using a housing for mechanical components, as in the original engine + this is possible, if mechmicd valve steering with a required housing is built + Advantage that there are no lubrication-problems with the valve-steering- system. ● using tubes for oil return to the oil sump Q extra co~ections to the oil-rem from the cylinder head have to be designed These two possibilities are added to the morphological-box as oil return “with housing” and “with tube”.
26
.—— >
27 ///
pa.xaalsQA~A “J@
(yIoptqos Luoy-qos The different solutions are now ordered in the morphological box (Table 2). The next step is to select unsuitable solutions and to eliminate them. The results are directly added into Table 2. To decide which mechanical demands are requested to the valve steering the loads and movements are examined. They are shown in the Part Forces, Loads and Ways. The course of the valves is given in part 3.2, Forces on the valve-train and the push-rods:. The time for one valve to open and close is about 200 degrees crank-shaft-angle. At a speed of 2000 rpm this equals 1/60s = 0.0167 sec = 16.7msec. It is of course demanded that the valve course is the same as in the original engine. So the steering device has to be able to create this course more or less exactly in the demanded time. This is one criterion to eliminate unsuitable solutions. Another criterion is the mechanical loads, especially the max force on the valves, also given in the Part Forces, Loads and Ways.
The next possibility to decrease the number of solutions is to combine different classes of solutions. These classes are also marked inside Table 2.
The number of classes remaining is: I Function I Number of classes 1A 14 I IB I c 16 ID 15 I
Since there are a lot of solution-combinations that are senseless, a combination-matrix is created.
IAIIIXIXIXIXI
I I 1 1 I ! 1 r ! ! 1 , 1 ‘3 1 x x’n’ I c c i 3 2 x
— [ I 1 1 I I 1 ! i , , , , I -4 -. -. -. C14 I I Ixlxlx N P P P P CI’5 1X1X-— -— P P P P P n n n n E ii x P r r r r D 1 x x x x x N D 2 x x x x N D 3 x x x N D 4 x x P D 5 x P E 1 x x E 2 x Table 3: Combination-matrix for valve-steering N: not necessary; P: not possible; S: no statement possible Solutionl: Al, (III,) C2, Dl, El Q/./j /-hQ-@ :—-—— —— 1’ Ci&iJ M
+7/ -P—. --
Solutiofi: Al, (El,) Cl, D2, El
SohXion3: Al, Bl, C3, D2, El
Solution4 : Al, Bl, C3, D3, El
30 Sohxtioti: Al, B2, C3,J#E2
Sohtionv: A2, Bl, C4, D4, E2 r
Solution8: A3, Bl, C5, D4, E2
.- Solut.iong: A3, Bl, C6, D5, E2
Solutionl(k A4, B2, C5, D4, E2
~“. WA. - f
SohXionl 1: A4, (B,) C6, D5, E2
Figure 10: Principle solutions for valve steering 2.2.1.2. Discussion of the found solutions 11 different solutions for the valve steering are given in Figure 10. It’s now the task to choose the suitable solution that fulfills the demands best. . The oil-return is not the most important function, so it is not weighted in the selection (the simple oil-return through the housing is just a suggestion). . Important is, that the mechanical loads can be handled (opening and mass forces of the valves) ● Another criterion that can be used to cross out unsuitable solutions from the morphological box is that the valve course has to be retained. . The solution should also not be too costly.
While examining the solutions it should not be forgotten that the optical access is the main task and not an ideal valve-steering, where a lot of nice features are given!
The solutions can be divided up into two parts: 1) Individual steering of the valves is possible. These are the solutions where the valves are steered electrically or hydraulically (solution8, 9,10,11). 2) Fixed valve-course that is given through a (partially @ solution 7) mechanical steering (solutionl, 2,3,4,5, 6).
Part one has the big advantage, that combustion with different valve-timings can be studied. To do so normally a new camshaft has to be produced, which is a very costly process. But since variable valve-timing is not the main-task it is not seen that important. More important are issues are the costs of the solution and that the mechanical demands are satisfied.
The hydraulic solutions (8, 10) are in need of valves that can produce a cylinder- movement representing the original valve-course (compare valve-movement in part Forces, Loads and Ways). Since these are very high demands special valves have to be used. Muug valves can handle these demands, but they are very expensive and moreover no experience with them is in existence here at the institution. For this solution additional hydraulically aggregates like tanks and pumps are needed. For information about problems and costs to be expected Thorbjom Brihmstrom at Rexroth Mecman in Malmo, a company that produces hydraulic components, has been contacted. The biggest problems are cooling (the oil gets very hot, because of the fast movement), leakage between hydraulic-cylinder and piston, and sealing. Since only 2 or 3 are needed and they are not built in a series the expected price is very high and may blow up the budget. An example for a hydraulic valve-steering done by Lotus-engineering can be found in [1]. The expenditure for such a system is enormous. Altogether the costs and the expenditure for this solutions are seen, since valve-steering is not the main task, as too high. Referring to Bengt Johansson, Ph.D. -Asc.Prof, and Lotus- Engineering the price is about 1Y2Mio. SEK. The decision is not to use hydraulically valve-steering. ——
m
Drawing 7: Hydr. valve-system, used by Lotus-Engineering (source: [1])
Remaining from pzmtone are the electrical solutions: Like said in the part Forces, Loads and Ways a maximum load of about 5.3KN has to be generated very fast. An electric valve-steering is no completely new idea. This has been done before by GM, published in [12]. Unfortunately big problems occurred. One problem was that it was very difficult to create the same course as the crankshaft. An example for a reachable course is given in Figure 11. Referring to this paper the expenditure is expected to be much higher than using mechanical valve-steering. The ordinary mechanical way is even preferred here. The decision is not to use electrical steering.
Exhatmt Intake
m w no m SW w 484 w Ooo Crank Angle (degrees)
Figure 11: Valve-course electrically valves (source: [12])
34 Part two arethe mechanically steered solutions. This is done either directly or by using a hydraulic cylinder. Using the hydraulic solution (7) the same problems with the cylinders as in (8) and (10) are expected. Here not that many additional aggregates and valves are needed. So this solution is not so extensive as the other hydraulic possibilities shown in part one. Nevertheless the problems with the cylinders remain and since we have a closed system here, the leakage-problem is of very high importance. If oil gets lost air can come into the system and so the valves cannot open properly. This is a major disadvantage. Even here the expected problems and costs are too high. This design will not be built.
A filly mechanical steering can be done by using an overhead camshaft or by using push- rods/connecting rod. The Problem with the overhead camshaft (3, 4, 5) is, that a new shaft has to be built and bearings for it have to be added to the cylinder head. This is a very extensive and costly solution and has no advantage compared to push-rods/connecting rod. This design is not favored either. The remaining solutions (1, 2, 6) can be distinguished between a steering using a return spring with an original valveklve-rocker design and using a connecting rod. Coming so far it is obvious that the design using push-rods and original valve-rockers is the easiest one to be build. Steering the push-rods from an external eccentric axis (1) has no advantage compared to steering them from the original camshaft (6).
The chosen solution is now to use extended push-rods and drive them from the original cam axis that is in existence (solution 6).
Advantages: . easy to build Q No eccentric axis, no extra-shafts... ● cheap @ a lot of original parts can be used, only extended push-rods are needed . oil-return through push-rod-housing is possible ● steering is as precise as the original steering @ real conditions . no external aggregates needed ● no maintenance-intervals . fast and easy to install
Expected Problems / Disadvantwes: ● no variable valve-timing possible ● Higher mass forces on the original cam because of the heavier push-rods that have to be accelerated. ● push-rods cover a part of the optical access (they stand before the window) ● Problems are expected, joining the extended push-rods. It could happen that they buckle under the expected loads. An additional bearing or other design of the push- rods could be necessary.
The chosen solution is, as mentioned earlier, not the ideal one for valve-steering, but the most effective one. It fulfills the demands, is easy to handle and the costs are low. The combination of these factors gives the best possible solution for the main task! A future possibility is the very nice feature of variable valve-steering with hydr. cylinders. This can be part of another construction, for example in another master-thesis.
35 ——. .
2.2.2. Fixing of the glass-windows, sealing, cooling, oil-supply, design as module; access through the cylinder-wall
This section contains the glass-modules, the extended-cylinder and the extended piston. The question is now, which part is done fnst. Design of the piston does not influence the other parts strongly, so this can be completed last. The two modules are chosen to be designed first. So they determine the connections for the cylinder-extension!
2.2.2.1. Design of the glass-holding-modules
Statements: The oil supply to the cylinder head normally is done through a small hole in the bottom of the head. The upper ring is expected to be so big that this hole will be covered by it. Taking a look at the drawings of the head, a very easy access from the side is possible. So the oil supply does not have to be examined longer in this section.
Because of the same reasons the water supply to the cylinder head for the glass-ring solution also has to be done from the sides. For the solution with the glass-parts a supply through the glass-socket is thinkable. This would mean two different supplies for the cylinder head resulting in high expenditure for a part that is not the main-task and will not be used so ofien. Moreover building space for leading the water around is needed. This would extend the whole construction, which isn’t favored. Another point is that the piston is the most heat-critical part. If it is not possible to cool down the piston enough, a water- cooling inside the glass-part-module makes no sense. Because of these reasons, water- cooling is given up in a first decision.
The glass ring is expected to be the critical part regarding to the loads, in any case more critically than the design with the glass parts. Because of that also a bigger building space can be expected (at least in diameter), and so this part determines the dimensions. The original connections of the screws to the cylinder head cannot be used here. These are the reasons to start designing the glass ring f~st. During the whole work it has to be reminded that both solutions have to fit to the same connections. It always has to be taken care of the other part! A change of them as modules has to be possible!
2.2.2.2. The Glass-Ring-Module
Basic dimensions of the ring are calculated in the part Forces, Loads and Ways. The height of the ring is expected to be about 30mm and the wall-thickness is expected to lie between 10rnm at lOObar and 29mm at 200 bar. First the “biggest case” with 29rnm wall- thickness is chosen for coarse dimensioning.
36
—.-— Functions of the ring-module
The main functions that have to be fulfilled are now examined. They can be formulated to:
~ Element Function to be fulfilled for this element
1.) Glass ring ● Sealing . Fixing the ring @ Leading the force from the combustion @ possibility to clean the ring easily 2) Steel-housing . Take up the combustion-forces ● Connect to the cylinder head/extended cylinder Q Build as module 9 Easy cleaning of the ring
● Sealing Table 4: Main-functions for the module ring
Solution-trees for rin~-moduIe-functions:
Like done previously in the part 2.2.1 Steering of the valves, solution-trees for the functions are created. The solutions are combined in the morphological box to different designs. These designs are discussed and the best one is chosen.
If one branch has solutions that are not possible this one is not drawn further. This has occurred for gluing of the ring (Figure 12) that of course makes a change impossible. Furthermore it is not possible to use screws for fixing glass. The remainin g solutions are shown in the morphological box.
37 — -.—————
Fixing of the ring
r&icJ elastic
/\ ~ .&!Ei w ~ \e — - no dismoun- - not possible ting possible if el astic -no dismoun- ting possible
screws/“’”into glass: clam pg lass: screws/“”into glass: Clamp Elass: mech. properties of glass not suitable mech. propetiies of glass not suitable
from the si~s from top andbottom M
Figure 12 solution-tree: fiing of the ring
Cxmnection to the extended cylinder/cylinder-head
screws horizontal screws vertical => clamp
use of oritial holes in cylinder head taking up counter-force beside the head
A A module remains when no connection between the module remains when no connection between the taking of the head partswhen takinp of head taking of the head partswhen taking of head
A A A A & E ~ ~
~ewsfor connection % threadbetween module to extended cylinder and extended cylinder
Figure 13 solution-tree: connection to extended cylinder/cylinder head
38 Connecting the modules to the extended cylinder, one solution was to use a big thread between the module and the extended cylinder. It’s shown in Figure 13 as possibility B. The problem using this connection is that the fixation has to be very precise. The compression-ratio has to be exactly the same as in the original engine. Another problem is the high pressure inside the engine (up to 200 bar) and the sealing problem. Because of this, the solution B is not suitable and is not examined further.
The solution-tree for sealing the ring is given in Figure 14. The possibility of cleaning and exchanging parts must be given, so all scalings have to be removable. The branch to the left has not to be continued anymore. If an elastic fixing of the ring is wished later, only elastic scalings, both at the bottom and top and at the sides are needed. This is examined later by combining all solutions after the morphological box.
Sealing
not or only limited removable removable sealing - gluing - welding - press fits - roll fits ~ all not suitable !
flat sealin~ ring sealing
hard el asticisoft metaS rubber
Figure 14 solution-tree: sealing of the ring
The solutions for the main-functions of the glass-ring-module are given now and they are combined in the morphological-box. The procedure is the same as done in the part 2.2.1.1 before. Here the number of solutions is fewer than earlier. It is not necessary to create a combination matrix to eliminate impossible combinations of solutions. It is obvious that for example a fixing from top and bottom (A2) cannot be combined with a sealing only from the sides (C3)! The question is rather to find a solution that gives optimal field of view and withstands the appearing loads from the combustion. Problems with the different solutions are examined and discussed. Not many possibilities remained and it was not necessary to create solution trees to keep everything clearly arranged. ———-. —— .—-
40
.- —.
——. —.
Discussion of the found solutions:
First the solutions for fiig areexamined: . Al Using a rigid fixation of the glass from the sides (Al) you have to take care of the expansion of glass and steel. The thermal expansion coefficient of glass is smaller than that of steel, so no problems should occur due to stresses on the glass, resulting from the expansion. But since the outer steel-parts are not as close to the combustion- source, they are not warmed up as much as the glass ring. This means the expansion cannot be precisely determined with a suitable calculation-expenditure. Since the glass ring is very expensive this solution is not favored. Another problem is that if the steel parts really expand more, the ring is in no fixed position any longer. But if the ring moves just a little bit off center, the piston may hit it and it will burst. Another disadvantage is that the field of view is decreased by the fixation. This solution is not suitable. . A3 The same reasons as for Al also count for A3. ● A2 Using rigid fixation only for top and bottom (A2) shear stresses are encountered by the glass ring. @ Obviously an elastic fixation gives few loads on the ring and is preferred here!
● A4, A6 Fixing the ring from the sides (A4, A6) the problem occurs that the field of view at the upper part of the ring is decreased. It is the aim in this construction to get a view that is as big as possible.
. A5 The above-discussed facts determine solution 5 to the most suitable one. * olution A5 is chosen most suitable.
. One problem that occurs when using this possibility is that the ring has to be fixed very precisely. If this is not done, the upcoming piston meets the ring and it will burst. Mounting the ring a plastic piston or tube is inserted into the cylinder so that the ring has its ideal position. After the ring is fixed, the tube can be removed.
42 Now that the fixation of the ring is determined, a suitable sealing for this fixation has to be found. The different possibilities and problems are discussed below.
● C3,C4,C7 Solutions C3, C4 and C7 cannot be combined with the chosen fixing of the ring A6. ● C6 Solution C5 is not suitable either. Work has to be done on the glass ring and a notch must be worked into the glass. A notch in a glass part is considered a stress-raiser and is always the starting point for cracks, making this sealing is absolutely undesirable. Moreover the demands on the sea.lings are very high. A high pressure under high temperature has to be sealed. Problems can occur using ordinary o-ring scalings. These demands can only be handled with special materials.
● cl, C2 C5 These scalings are the only possible for fining the glass. It has to be figured out which one is the most suitable one. C2 and C5 have in common that the glass is only joined on a small ring, but not over the whole possible surface. A simpler alternative is the use of flat scalings made of graphite. The glass is joined on a bigger area what decreases the stresses onto the glass. Experience with these scalings is in existence here at the institution and no problems occurred. They transmit nearly no shear stresses (@loads on the ring!) and they are easy to handle. Moreover the costs for them are acceptable.
@ The advantages for using flat graphite scalings are obvious. They will be used for sealing the ring. ~Solution Clis chosenj ..-—— ..—. - —
From function A and C the best suitable solution has been figured out now. The last remaining function is B connection to the cylinder head/extended cylinder.
Here one problem occurs that has not been discussed before. Solutions B4, B5 and B6 have in common that the glass ring is pressed directly against the original cylinder head. Doing so the combustion can be seen over the whole Top-Dead-Center (TDC), as required. Taking a look at the head, it can be seen that there is not much space to support the glass ring. The smallest distance between bore and outer-surface of the cylinder head is not much more than 10mm. It is not possible to join the ring only on such a small area. Here an additional plate like shown in solutions Bl, B2 and B3 is needed. In the List of Demands it is stated that an original head should be used. Also the combustion at TDC has to be observable. It’s not possible to fulfill these two demands at the same time! One of the demands has to be given up. Discussing the possibilities, problems and advantages with the instructor the decision was to modify the cylinder head. It was more important to have a good field of view, than to use an unmodified cylinder head. Another advantage is that using the module with the small glass parts, the upper zone where the injection nozzle is situated also can be viewed. Since the glass parts are in need of a steel-fixation always some space is lost above and below them. Now there is space for the fixation inside the cylinder head. Q Tlis means that the head has to be turned down, so that a steel-plate can be added there.
. The question is now whether to fix the whole device, using the original holes for fixation or to make a fmation beside the head. Using original holes is shown in solution B1 and solution B2, using a fixation beside the head in solution B3 (Here the plate also can be fixed using the original holes, but not for taking all the combustion- loads, only that it’s not falling down when the head is taken away). This depends on the maximum thickness that is usable for the plate. In the drawings for the cylinder head, the oil and water-channels can be seen. To get a better view, of what is possible one cylinder head was turned down and then sawed into parts. It is possible to turn down the head 10 mm without touching the water-shield or oil- drillings inside the head. If they are touched, they have to be sealed against 200 bar and high temperature in a very small area! This should be avoided. Picture 2and Picture 3 show the cylinder head that has been turned down 12mm. The critical places where the water shield is touched are marked. It can clearly be seen that it is not possible to turn down the head more due to water-leakage and stability of the head. Also the valve-seats and the intake and exhaust system can be seen clearly here.
44 Picture 2 Touching the water shield of the cylinder head
,.
Picture 3: Touching the water shield of the cylinder head A suitable form of the plate was examined, where the water shield is only touched at areas where no pressure form the combustion appear and only the water-shield has to be sealed against loosing water. These areas are easy to seal. The cylinder head is turned down 12mm maximum. In the part Forces Ways and Loads it has been calculated that a thread length of 10mm is enough for fixing the screws. The material of the plate has to withstand a Rm-value of at least 1080 N/mm*. Solutions B1 and B2 remain! . The next decision is whether to use a plate under the glass ring (Bl) or to support it directly from the extended cylinder (B2). This possibility can be estimated very easily. The advantage of a construction without a plate is that the extended piston is shorter. This results in a lower weight and the mass difference to the other original pistons is smaller. Another fact is, that the push rods do not have to be extended so much and the danger of buckling is smaller. The big disadvantage is that when the modules are changed, because of the length- difference new push rods have to be added. So two different pairs of push rods have to be manufactured. Another point is that when taking off the head the ring has no fixation. It may stick at the head what can result in problems. Taking a plate, the ring can be set into the module very precisely on the workbench. Also the head can be transported together with the module easily and may be handled on the workbench.
+ The solution with plate
A coarse drawing how the construction may look is given in Drawing 8. Here a circular plate for fixation is drawn. The final design will probably not have the same shape, but the working-principle can be seen here. The upper screw shows the height of the cylinder head.
46 Drawing 8: Working-principle for the glass ring module
2.2.2.3. Changin g of the modules @ water-sealing-problem:
One problem which touches the principle construction of the modules is, that turning down the cyIinder head 12mm the water-shield is touched. The problem is not to seal against the water, but the change of the modules in an easy and fast way without draining of the water. Four ways are possible here:
47 — — .
Module-change
/ \
RemovinEtheplate andsettin~ Leaviu theplateandfixing Two modules on in new modules with.dassrmrts ,glass-wttmodelunderthePlate: two heads:
-upperpartof combustion - high expenditure underinjection-nozzlecannot be viewed
Water-shied-sealinq Holes in water- removesAwith plate shielddurablesealed -cooling waterhasto be drainedoff, -sealinghasto be done on thehead! before themodules canbe changed Figure 15: Problem - Changing the modules
The solutions at both sides in Figure 15 have the big advantage that the combustion- process also with the glass-parts-module can be viewed directly under the cylinder head. The solution in the middle does not allow this possibility. This solution is not suitable. Sealing done inside the head by setting plugs into the holes of the water-shield, so that the plate can be removed without water leakage, isn’t possible either; the wall-thickness is not sui%cient. So this solution doesn’t have to be examined any longer. Two solutions remain: 1) water-shield-sealing removes with plate and 2) Two modules on two heads. The fwst solution means, that the cooling-water has to be drained off, something that should be avoided. With this solution much work is required to change the modules. This is therefore not a good solution. No 2) means, that the module has to be changed together with the whole cylinder head. This isn’t the perfect solution either, but the only possible one here. It is possible to connect the water-supply to the head with standard valve-connections, meaning that the water-supply can be removed very fast and easy without loosing water. The expenditure is higher, but it has a clean and easy way of changing everything. The water-shield will be sealed using sealing-paste. o 2) is the chosen principle.
48
— 2.2.2.4. The Glass-Part-Module:
The main-dimensions for this module are determined by the size of the windows. They have to be adjusted so that combustion at TDC can be viewed. Examining the old construction for another engine, at least 8 mm have to be turned down of the cylinder head. Doing so, unfortunately the water-shield is met and even no sealing of it inside the head is possible. The wall-thickness is too small to set in plugs. The same problem as with the glass-ring-module occurs here. Since 8mm seem to be very small and no further problems occur when turning down more, the building space is extended to 12mm; the same size as for the glass-ring module. The disadvantage is that the module can only be changed together with the head. This makes it heavier and not so easy to handle. But the field of view has to be retained and draining off the cooling-water for a module-change is not favored either.
The next step is to decide which parts of the old construction maybe used again: Here are of course the glass windows and even the fixations of them that maybe re- usable. They are only small parts, but they withstood the loads before and there is no need for redesign and new manufacturing.
Then the fixations to the cylinder head and possible scalings are examined. A design similar to the glass-ring-module is done. After that, the connection to the extended cylinder has to be examined. Here it is important that the gap between the cylinder and the module lies as much upwards as possible, so that the sealing-rings of the piston can be mounted as high as possible. The sealing-rings may not slide over a gap! Also the gap has to be as small as possible so that the compression-ratio is not affected. It is important, that the heat can be transported away. The above mentioned facts have been taken into consideration in the following Drawing 9, where the working-principle is shown. Preliminary dimensions are also given. — _.-— — _
Drawing 9 glass-module; side view Q
Drawing 10 glass-module; top-view In these drawings the water-cooling inside the module is not considered. The most critical part for cooling is the new piston. If there is no sufficient cooling possible, also no cooling for the glass-part-module is necessary. Moreover the water-cooling was only formulated as a wish in the List-Of-Demands. Another fact is that the glass-ring-module is accessible from the sides. Air-cooling is here also possible. This is much easier, the dimensions are smaller and no problems with water, while removing parts occur. Discussing these facts with the instructor, the f~st design will be without water-cooling for the glass-parts-module.
50 2.2.2.5. Design of the glass-piston-crown
The piston crown has to be made of glass. The possibility should be given to change between different glass-parts later. So the glass-piston-crown is designed as one part that is fixed to the piston extension. Using another glass shape the glass-crown can just be changed and the piston-extension can be reused. The engine also maybe driven with a standard piston with a standard combustion-bowl. The glass piston determines some dimensions for the cylinder-extension. So it has to be designed before the extension.
Requirements to satisfy are: . Using a glass-part with the biggest diameter possible . Fixing the glass-part so that it’s not clamped from the sides . Leading the heat away as well as possible . Making a change/cleaning of the glass part as easy as possible . Light design of the piston, so that the weight difference to the 5 other original pistons doesn’t get too big.
Cooling Principle: The piston is the part that has to withstand the highest loads from temperature. In a normal combustion-engine, spraying oil against the bottom cools down the piston. This can of course not be done here. Like also done in some special valves, liquid can be used inside the piston to lead away the heat. This has the disadvantage of high costs, reduces the glass-diameter and may increase the weight of the piston. The suitable solution here is to use air cooling of the piston. A tube can be inserted into the cylinder from the sides, for example where the mirror is located. Cool air can be pumped through this tube to cool down the bottom of the piston. The glass fixation has to be done in such a way that the heat from the combustion is led well to the bottom.
Sealing-rings of the ~iston: There are also special demands on the sealing rings of the piston. The piston has to be run without oil-lubrication, so normal sealing-rings cannot be used. Experience with sealing- rings made of Rylon is in existence here at the institution. An expert with these scalings is Prof. Gumar Lundholm. They worked well in the past and so there are no reasons to change them. The design of the Rylon-ring-sealing is given in the next drawing. The o- rings act as springs, which press the Rylon against the cylinder wall. They should be pressed together O.lmm after mounting. Important is that the surface of the cylinder-wall isn’t too smooth. It has to have a roughness-number of R= = 0.4 – 0.8 ~. While running the engine small Rylon-particles loosen themselves from the ring and fix on the rough cylinder-wall. The Rylon-ring then glides on the Rylon-particles on the Cylinder-wall. The sealing surface is formed only by Rylon. Particles that loosen from the ring and don’t fix at the wall fill the gap above the rings. The lower ring is needed to lead the piston inside the cylinder. There are no special demands for that ring. The dimensions are experience-values from Prof. Lundholm.
51 — .—. —
2.3 } P
L ‘i +y,+-<:. 4J- I—— ---- ‘“/ 0
\ /“ T ,-’ / “/,——’——/“’” I 0’
Drawing 11: Piston sealing-rings of Rylon
Fixation of the glass-pa~ The fixation of the glass-part is the most important task at this point. The requirements listed before have to be fuli311ed. In principle, 3 ways of fixing the glass part are thinkable: . Clamping the glass from the sides @ Not possible because of mechanical demands on the glass. o fixation from the upper side . fixation from the lower side The last two possibilities are shown in Drawing 12.
52
-.———— ..—— .—. -1
d
I
1 LJ./-’ c)i
Drawing 12: Different possibilities of glass-fmation in the piston
A possibility that is not shown in Drawing 12 is to use a big thread inside the extended cylinder to fasten the glass. This is also in the category fixation from the lower side. This possibility is not suitable because the space for using a tool for opening and closing the thread is too small and so it is not drawn.
53 — ...—-- — -. . . ..——.—
Discussion of the different solutions:
It is now time to rate the different solutions and figure the best one out. Advantages and disadvantages of the different solutions are discussed, suitable and unsuitable solutions are determined. Starting with solution I that describes the building principle of a smaller-car-size engine, the occurring problems are marked with 1) to 4).
. Drawing 12-1 shows the fixation from below. The inner-diameter is wished to be as big as possible to get a suitable field of view. In contrast with this wish are the mechanical demands on the wall-thickness. The numbers 1 to 4 mark the places, where the wall thickness is critical:
1) determines the supporting area of the ring. It cannot be too small, because of occurring stresses in the ring. 2) is the wall-thickness, where sealing or leading rings are added inside the piston wall. The remaining wall thickness has to stand the mechanical loads from the combustion and lead the force to the connecting rod. 3) is the wall-thickness, where the screws for mounting the cylinder-extension are set in. It also has to lead the force from the combustion to the connecting rod and it has to fix the screws. 4) is the wall-thickness of the piston-extension. Like 2 and 3 it has to lead the combustion-force. Since the piston-extension is expected to be manufactured of aluminum, this dimension may also become critical.
Different possibilities of avoiding these problems are shown in Drawing 12-II and –III. . In Drawing 12-IIa the glass-plate is fwed from the upper side with a ring. Problems that occur here are that the heat isn’t led away well from the ring. The ring may bum. When this happens, the glass-part can loosen and the financial damage is high because of the high costs of the glass. Another disadvantage are the many additional gaps right on top of the piston. (Inside the screws, between screws and metal, between ring and piston...) These gaps have a poor influence on the exhaust analysis. Fuel/air- mixture can come into them, but will not burn (the gaps are too small for a flame to reach). The hydrocarbon (HC)-emissions will rise and are not comparable with a standard engine!
. Drawing 12-IIb shows a connection of the upper-ring, using a big thread between ring and piston. Here still problems with gaps as in solution IIa occur. The heat- transfer may be better, but mounting the ring is a problem: A graphite-sealing has to be used between ring and glass-part (The glass may never touch metal!). Turning the ring for mounting it, the graphite sealing also turns and may not lie properly on the glass ! Q Solutions II have advantages with the wall-thickness at place 2. The piston sealing-rings may even lie under the glass-part. The wall-thickness in this place then is increased some millimeters. On the other hand there are, like shown earlier, big problems with these solutions. They are rated as not suitable.
. Drawing 12-III handles with the problems 1),2) and 4). Here the suggestion is to reinforce the wall-thickness of the extension and the ring inside the piston only at some places. Fewer screws are used for fixing the piston to the extension. The supporting area of the ring is bigger, even if the ring gets thinner. A thinner ring in
54 return can result in a higher wall-thickness of the piston, affecting problem 2). Also the problem with the fixation inside the extension [place 4)] gets smaller. The disadvantages are that the field of view is decreased in some places. Since the biggest field of view is wished this solution is not suitable either!
. Drawing 12-IV deals with the problems at places 3 and 4. Here a big thread is used to join the glass part inside the piston. To avoid the problem that the sealing is pushed away when turning the thread like in solution IIb, a screw is added to fix the distance- ring to the outer-piston. Also a groove on the ring maybe possible for this task, as shown beside. This solution has the advantage that not so many screws are needed and mounting and de-mounting goes faster than with solution I. Disadvantages are that additional gaps on the surface are necessmy which influence the combustion badly and a special tool for opening and closing the thread is needed. If the tool isn’t adjusted right and glides out of the run the glass surface maybe damaged! This can be avoided by placing the holes for the opening tool at the sides. Another danger may be that the thread may stick because of heat-stresses that may occur from the combustion; the thread may deform and it maybe impossible to open the thread again. These problems occurred before here at the institution while using changeable piston crowns made of aluminum. An advantage in this construction is that two different materials are used. The top window-holder will be manufactured of titanium and the piston extension of aluminum. So the risk of sticking is not so big. Also air-cooling will be used onto the under-side of the piston what decreases the risk of heat-deformations. The pistons where this problem occurred before had the thread much higher and nearer to the combustion than in this construction, so the risk here is not so high. But still it is not possible to predict surely if this problem may occur or not. The connecting rod is expected to withstand the torque for opening and fastening the thread (it has to take the counter-force while opening and closing the thread).
Rating of solutions I and IV:
In the end only solution I and IV remain as possible. The other possibilities have, because of fix demands or technical problems, been chosen as not suitable. The decision, which one of the two remaining principal solutions is to choose, cannot be failed just by discussing advantages and disadvantages, like done before. Solution IV has big advantages in changing the piston crown fast and easy, but also the risk that the thread sticks can not be neglected. This gives obviously the reasons to use a rating-method with the strategy of Product-Development. As a simplification the rating does not distinguish between technically and economically values. This is possible because only one unit is produced and the costs for manufacturing the piston are relatively low compared to the total costs. This part of the product does not have to be optimized to costs. (They are of course important and also a criterion.)
The chosen criteria for rating are: ● High security, the threat must not open. . Fast to mount and de-mount the piston-crown ● Easy and cheap to manufacture
55 The different criteria are different important. First, using pair-comparison, an importance- factor is created. To get more precise factors the rater-numbers are made more precisely than in the [7], as shown in Table 5.
0 1
Much less Less important Equal I More important Much more I important important Table 5: Rate-numbers
Pair-comptison gives the rate-factors shown in Table 6:
I Securitv Mounting Manufacturing ISecurity o 0 Mounting 4 Manufacturing 4
Z (absolute) 10 5 3 z (%) %?curity= 1 %founting= 0.5 gManuf@ring = 0.3 Table 6: Pair-comparison Here the security-factor doesn’t seem high enough, although the rate numbers have been made more precisely. New factors, which seem more reasonable, are created: Smrily = 1 I %o.nting= 0.2 I %fnnufacttig= 0.1
Table 7: Rate-factors
A rating scale now is created (VDI 2225):
Criterion 4 Security Very high, M no risk acceptable I Mounting Very fast Fast and w and easy easy time for mounting needed Manufactu- Very cheap Cheap and Suitable Higher Costly and ring and easy easy expenditure expenditure high ex enditure Table 8: Rating-factors d
The different solutions get values from the rating-scale-Table 8.
I Solution I (PI) I Solution IV (PIV)
Security \4 ,13 Mounting /1 13, I Manufacturing 11 14 1
56
— Rate-factors Verbal Solution I Solution IV description Normalized values Normalized values
not rated Rated not rated Rated gi PisO1.l [email protected] PisO1.w gioPisO1.~ 1 Security 4 4 3 3
0.2 Mounting 1 0.2 3 0.6
0.1 Manufacturing 1 0.1 4 0.4
---- .-#-- .A, .- .
WI gw~ Ww gww Solution-worth 0.3 ~ 0.75 ~ Ranking 2 1 1 2
The solution-worth gives now the percentage-value of the ideal solution that is reached. Once without the rate-factors (Wi), once considering them (gWi).
Decision: Solution I now is ranked the better one. Since the difference between the ranking isn’t so big the advantages and disadvantages, the driving situations and need of changing the piston crown often, are discussed with the instructor again and taking into consideration the experience when driving the smaller car-size-engine, Solution I was the choice. The main argument was the security that this system works well. It has been proved in the past and the risk of failure is wanted to be avoided.
Coarse dimensioning of the piston-crown:
The piston crown is designed in that way that it gives the biggest possible field of view. It was a major aim to design the inner diameter of the glass-holder as big as possible. A lot of dimensions, as for the sealing, are given before in Drawing 11. The recess inside the piston wall for the sealing package does not exceed 4mm. 0.5 mm is the gap between piston and cylinder wall. The wall thickness at place 2) in Drawing 12 is determined to 4 mm and the wall thickness for the inner ring (number 1) in Drawing 12) to 5 mm. This determines the glass diameter and the field of view to 100rnm (the outer piston diameter is 127mm). Examining another smaller car size engine at the institution with 100 bar maximurn- pressure and taking experience-values in account the glass thickuess is expected to lie between 50 and 65 mm. For coarse dimensioning it is set to 55 mm. Since the acceleration-forces work against the pressure-forces and so decrease the total-force the calculations are “on the sure side” with a lighter glass part (a heavier glass would result in higher acceleration-forces and lower total-forces !). Another important dimension is the distance of the piston-sealing rings from the top of the distance in Drawing 13 named 11.It is chosen so that the rings do not glide over the gap of the cylinder-extension. For the calculation of this value the distance between piston at TDC and clearance has to be known.
57 . . —.. .——4
The length between the piston at TDC and the clearance is calculated out of the compression-ratio. With a compression-ratio of &=Vt/VG18 the total volume is: Vt v, =154”~”y2 mm3 + Vc = 1.9508 1=1.9508. 10-3m3 + — Rc
e w. 1–+ I= 1.9508. 10-3m3
@ Vt =2.0656. 10-3m3 and the compression volume: Vc = 1.1475 “104ZZ?3 When all other gaps are neglected this results in a compression length of 9.5088mm =9.5mm. This dimension is also shown in Drawing 13. Since the gaps have an influence on the compression ratio and considering that different piston shapes with a combustion bowl eventually want to be driven, a constant c1 = 7mm is added later. Moreover a security distance constant C2= 3 mm also is added. e C.lomm
11now can be calculated to 11=38+6+ c – 9.5 = 44.5mm. — The worst case is defined in the List-Of-Demands that the distance between piston and top-clearance is 1.5 mm. This would result in a distance between piston-rings and gap of at least 1.5mm. This is always sufficient. L, the height for sealing and leading the piston is, according to Drawing 11, set to 19mm. The last dimension that has to be determined is 1s,the height of the ring, supporting the glass-window. The ring is needed because different glass-shapes want to be used later. This also means different thickness of the glass-window. The worst case is a window with a big hollow in the middle. This would obviously increase the window-thickness. The len h 13is set to 25mm. The total length of the titanium piston-crown is now determined tob. A drawing with coarse dimensions is given here:
58 !*
Drawing 13: Coarse dimensioning of the piston-crown with the glass-part module .—.
2.2.2.6. Functional analysis of the cylinder extension The extended cylinder is one of the major parts of the device. It has to fulfill a lot of requirements. Even if it “only” is a cylindrical tube a more precise examination of its functionality is required. The incoming and outgoing quantities of the extension are shown in Figure 16.
Cooling water +:’ixingoftiem”d”’esseahng of the combustion-chamber — Heat from - cooling of the device com ustion - cool-water-supply for the cylinder-head Counter- force horn pressure ilom (eventually, only for one module) ➤ com ustlon Combustion loads - fixing to the engine-block
+
Figure 16: Function analysis of the extended cylinder, using the incoming and outgoing quantities.
Other functions and important demands are given in the list of demands. The functions together are: . fixing of the modules ● sealing of the combustion-chamber ● cooling of the device . cool-water-supply for the cylinder head . fixing to the engine-block . leaving a side access for the mirror . leaving space for air cooling of the piston inside ● possibility of easy window cleaning has to be given . taking the loads of the combustion
Dimensions of the extended cvlinder: Before designing the extension the limiting dimensions have to be known. They are mostly determined by the piston-extension. Its dimensions can be seen in Drawing 14. Terms: 1.= extension’s length 1~= length of hole at the side 1,= mirror length 1,, = 100 mm; mirror dimension vertical= mirror dimension horizontal= expected maximum diameter of glass in piston rC= l%stroke = 77 mm lt = total length (inclusive piston) 1P= length glass-piston-crown = 80mrn
60 piston-extension
piston
Drawing 14: Extended piston in TDC and BTC
Coarse dimensions are determined with some simple calculations. Choose: ~ =2” rC+ lW + 1lmm = stroke + 1, + 1lmrn+154mm + 100mm+ 1lmm = 265mm (11 mm are added because of uncertainties.)
1== lh + 60mm = 325mm (60 mm are added for the fixation to the original and the glass-piston)
1, = 1,,+ 1P= 405mm (total length with titanium-piston)
Cooling of the device: Most combustion-engines today are water or air-cooled. Normal truck-engines are water- cooled. A simulation of the engine most resembling reality is preferred, so these 2 are the 2 possibilities. Defining the demand “cooling” more closely it can be formulated as “holding temperature on a constant level”! A short explanation follows here: Heat losses from the combustion-process depend primarily on the wall-temperature. So it should equal the one in a normal combustion engine to get suitable measurement results. Since the engine cannot be run very long because of heat-problems at piston and glass-ring, it
61 ————
probably has to be warmed up externally to get the right wall-temperature. Unnecessary wear is avoided if you can start your measurement quickly after starting the engine. Holding the temperature constant, respectively warming up the engine, with and airflow now is absolutely unsuitable. A design with cooling ribs is not possible and external water-cooling has to be done.
Spiral-formed tubes or direct water-shield: Possible designs are here to use a whole water shield around the cylinder or to use spiral- formed tubes, which wind around the cylinder. Disadvantages using the tubes are that the fixation to the cylinder isn’t so easy to create. Also the heat-transfer between cylinder and water inside the tube is of course not so good as with a direct water shield. Another point is that in the upper corner close to the glass- ring always space remains (lying a tube in a comer) and so at this critical place the heat transfer is very bad! The advantage is that the temperature gradient changes mostly from the top to the bottom. Using a water shield it differs more from one side to another (input and outlet of water). The water shield gives a better cooling at the critical places and even no problems in construction and manufacturing are expected, so this is the suitable solution.
Different uossibfities for changing or cleaning the glass-parts: One of the most important requirements to be thought of when designing the extended cylinder and the connections to the glass-modules and engine-body is that an easy cleaning of the windows and fast mounting and de-mounting has to be possible. This demand determines the principle shape of the construction. Solutions are found by examinin g the surfaces for a possible access. Three surfaces for cleaning access are possible (compare part 2.1 Main Solution concept, Drawing 1). They are cleaning from the top, the bottom or from the sides of the cylinder. Cleaning from the sides can be done by moving the cylinder up or down. They are shown in the following diagram and drawings:
Cleaning of the glass-parts
/ \
1) removing of cylinder-head 2) built piston that can be removed 3) moving the cylinder wall and cleaning on the work-bench downwards at the sides: down and getting access from the sides:
Figure 17: Possibilities for cleaning the glass-windows
1) The first one is cleaning from the upper side. That means that the cylinder head is removed together with the module and cleaning can be done on the working bench. This is the first and obvious solution. The smaller car-size engine is built that way. Advantages and disadvantages of this solution are:
62 A very simple construction for the extended cylindeq cheap and easy to build. The manufacturing expenditure is low! Much work for cleaning the windows, time losses while running experiments. The cylinder head has to be lifted up. This means that connections like for water supply, oil-supply etc. have to be removed. Oil may run out of the hole for the push rods when lifting the head.
Drawing 15: Working-principle of solution); ordinary design ——- .. .- .—
2) The second possibility is to clean from the under-side. The piston has to be removed. This can be done by building the extended piston from several parts, which can be de- mounted. These uarts then have to be removed through the hole for the mirror. @ One &sa&mtage is that an additional conne&on inside the force-way is created. But the biggest problem that occurs here is that the sealing-rings do not remain inside the cylinder during cleaning. It is very difficult to set them in again. This solution is not suitable!
+ glass window added here
.
/ ,original piston O
Drawing 16: Extended cylinders that can be demounted inside the engine
3) The third way of access is through the cylinder-wall. This is possible by moving the cylinder-wall down and leaving the head mounted. Doing so, the cleaning is done directly at the engine inside the lab and not on the working bench.
A first concept maybe seen in Drawing 17. Drawing 18 gives a preliminary 3D- design of the outer-tube that forms the structure for holding the construction. The holes for cleaning, the mirror and also space for the push rods is drawn. The movable cylinder is added into the outer tube and can be moved inside it.
64 n
Drawing 17: Movable cylinder for glass-cleaning
65 ————_
...... -.- .-.. -.,”” ,., .+..- -- ‘--, ---.4 ..._ _-+ ..-!+ .- .,, . --- .’, ... ~.. ~, . . ,.”.,,, ,,,. ”+- --. , . . ,$- -,.,- .__..- —---- .-
,. =- “Q.- ... ,.-, ..:-. ,. ..-’ ,4 .. “..f, . ,=Y .%.A --J’ ‘,.. , ‘. ,. ,+“.. ,“ .-. !.” >-- 4, .’ ,,
.,. “2 ~.~. “),,”
.“.- ‘----- .1”-__.--,—-.. .,./ ,--’ -“,...-+--’- ,,. ~ ,— ,-’- .7 ., -.. — . ..— ,- r“ -
.
,
,-,.. .—------k._ -.,. ,~-., / >,*,,, ,.,‘ _...-” -- -..------”—-.——., ,... :. ...-.’” .,- .... ““’”’ -., . ,. “...,,. --,.,.. ’--- --.’ J.- ..” ----
Drawing 18: Preliminary design outer-tube
66
— .—. — Occurrirw vroblems of solution): a) One problem are the extended push-rods. Their way lies inside the extension- dimensions and may not be changed. Work has to be done on one side of the extension like shown above, which results in a higher expenditure. b) Obviously a problem occurs opening the screws for lowering the cylinder (Drawing 17). There is not much space from above. Different solutions are examined: ● open the screws from the under side: 0 This has the disadvantage that additional holes inside the outer-tube are needed to open the screws, which decrease the stiffness of the construction. Another possibility is to open them with a long tool through the holes of the mirror. To examine the possibility of opening the screws from the under side, a paper model of the design has been built.
● use wedges and screws from the side to create sealing-force: @ Possibilities are shown in Drawing 19 and Drawing 20.
OR
I ,&T
~._ ,_ /“/’, ~’ I-..ny%xwxw G@-4Je e,,, Drawing 19 use of wedges to create sealing force Using the rectangular wedge like shown above there is the risk that the wedge sticks and the cylinder cannot be moved down. To avoid this, lubricating material has to be added to the surface of the wedge and perhaps a spring return has to be used as shown in Drawing 20.
,“
,,1 ,,, ,./ ,,
Drawing 20 spring-return for wedge
It should be added here, that these screws do not have to take the pressure-force from the combustion, but only the sealing force to seal the cylinder against the combustion-gases. In this design the outer structure takes the pressure force! c) Another disadvantage is the weight of the movable cylinder. It has to stand the combustion-loads and so a stiff design is needed here. A coarse calculation of the weight gives:
v metal=;. [(157.52 –127.52 )+ (177.52 –167.52)]. 160nzm3 = 1.508.000mnz3
V water = ; o[(d;~te,,OU,– d;a,e,,j~ )]. hW,e,
V water =:. (167.52 –157.52).130nmz3 = 330.000nznz3 = 0.331 and with p,t~ = 7.8 t/m3 and pW,l.,= 1 t/m3. m~t~~= 11.76kg mW~t~~= 0.33 kg mtOti = 12.09 kg
68 In this weight the flanges are neglected. A paper-model has been built to examine the accessibility of the model to move the part by hand. The result was that it would be, because of the high weight, nearly impossible to do so and a mechanical steering would be necessary.
Decision which design is to be built:
Possible solutions are 1) and 3). It is required to decide now which one to build. Solution 1) has low manufacturing costs, but more time for cleaning the glass is needed. Solution2) gives the great possibility of a fast glass-cleaning, but a higher expenditure for manufacturing has to be done and problems like mentioned above are expected.
Since it isn’t so easy to rank the different solutions and it isn’t obvious how easy the last design is to handle, a paper-model is created to get a better imagination of how to clean the glass parts. It can be proved to get with the hands into the combustion chamber or the expenditure for a special tool for cleaning maybe seen. This is a big help while getting the decision whether to use this design or not !
After building the paper-model it can be seen that dimensions of the whole construction got much bigger than without the outer-tube! The design was discussed with the instructor Bengt Johansson, the responsible person from SCANIA Anders Hultqvist, and a mechanic. The advantages and disadvantages, like named before, have been discussed and in the end the decision was to build solution 1). The strongest reasons against solution 3) and for solutions 1) have been:
. problems cleaning the glass properly ● not enough space to move the cylinder by hand. A mechanical motion for the cylinder is necessary and increases the manufacturing-expenditure! (Thinkable solutions are hydraulic steering or a gear-steering.) . a guidance of the cylinder in the outer-tube would also be necessary @ manufacturing-costs . in solution 1) the heat of the combustion can be led away better (air cooling) . removing the cylinder head isn’t much more work to do. Only a few screws have to be opened and it can be transported via a beam on the ceiling ● solution 3 is anew solution and it is not proofed that it works well
2.2.2.7. Re-desi gn of solution 3) Since the advantages of the new solution are not used, the decision to build the ordinary design 1) didn’t seem to be satisQing. The critical points were the outer tube and moving the cylinder easily. So in order to find a better solution the functions of these functions/elements are examined more clearly.
Moving the cylinder easily
Like before, a solution tree for the problem of easily moving the cylinder is shown below. The different solutions are shown and discussed. Moving the cylinder easily
/ \ 1)hydraulically: 2) mechanically: 3) pneumatically: Use of hydraulic cylinders Gear steering Use of Pneumaticcylinders Use Distonto move the cylinder Pump airfrom below into the cylinder to lift it up
Figure 18: Solution tree; Moving the cylinder easily
A very important point in choosing the suitable solution was the financial expenditure. Because of that a hydraulic/pneumatic solution and a gear-steering were out of the discussion very fast. One idea was to pump air from below into the cylinder and to lift it that way. This seems possible, but sealing has to be done.
The best possibility that can be seen here is to use the movement of the piston to move also the cylinder up and down. The cylinder maybe fixed to the piston in cleaning state so that each, piston and cylinder, can be moved together. Another possibility instead of fixing the cylinder to the piston is that the cylinder just may lie on screws during the cleaning. This is done, by setting 2 screws (called M4x15 stopscrew, part 30 in the drawings) horizontally into the cylinder extension. Opening the screws of the movable cylinder the stopscrews support the cylinder and it can be moved by turning the crank-shaft of the engine. This is a very cheap, easy and effective method to move the cylinder. It has no longer to be lifted by hand and the weight problem is solved easily with just 2 screws ! This is the chosen solution for this function!
Functions and demands on the outer-tube. design of solution):
I . withstanding the forces from the combustion I . holding and supporting the whole device ● guiding the extended cylinder . leaving space for moving the cylinder up and down
Examining these fimctions, a tube is not absolutely necessary. A more open design is more suitable here. Some solutions, which are not shown here, have been developed. The best one was to use 2 u-formed irons to replace the outer-tube. This provides good accessibility to the cylinder from the sides and also much space for air-cooling of the piston. A preliminary design is given in Drawing 21.
70 Drawing 21: U-Iron-Solution
Here the cylinder with the piston extension nearly in BDC is shown. A movement of the cylinder (drawn dimensions with water-shield) is possible. The u-irons can be fixed to the glass-module on the upper side and to the bottom-plate on the under side. The water supply to the water shield, the mirror and the push rods are not shown here.
71 Decision between U-Iron-Solution and Solutionl:
The new U-h-on-Solution is discussed again with the responsible persons. As a means of eliminating the decision, disadvantages and advantages of the both solutions are compared as shown below. The disadvantages of the new solution are that a higher expenditure is necessary for manufacturing the parts. It is a complete new solution, which always contains risks. The risk, which is seen here, is that the u-irons construction is not as stiff as the ordinary construction. The diameter and consequently the building space increases. On the other hand the new solution does not contain the criticized problems of the old solution 3) with the outer tube. Moreover it contains the great possibility of cleaning the glass in only some minutes. Just a few small screws have to be opened to lower the cylinder. Moving the cylinder is done by moving the piston up and down. Only a simple tool on the crank-shaft of the engine is necessary. All the connections that have to be removed, using solution 1) can remain when using the U-Iron-Solution. The cylinder head has not to be lifted and so also no problem with the oil-return occurs.
Comparing the advantages with the disadvantages, the solution is failed to use the U- Iron-Solution as structure for the new engine!
72 Old Solution: Movable cylinder:
Advantages
. certain that it works well no problems with oil- return when removing head @ clean and dry . easy to manufacture fast to mount and de-mount cylinder less time-losses while running experiments no connections have to be removed cylinder moved mechanically no longer weight problem
Disadvantages
. much work to change the head manufacturing is more costly than before . oil comes out of the hole for the push-rods when removing the head ~ heavy head ~ crane needed
. stiff construction risk that construction is not stiff enough . many connections have to be bigger in diameter removed when removing the head
air-intake air-exhaust fuel-supply water-supply water-return oil-supply
73 --——
2.2.3. Remaining Problems:
The functions which have been determined before (Figure 4), nearly all have been solved. The only remaining problems are:
1) Connection of the extended piston to the bearing of the connecting rod 2) Oil-return for oil that can be pressed up from the engine and then soils the mirror
2.2.3.1. Extended piston connection:
The problem is to connect the piston-extension to the bearing of the connecting rod. It has to be possible to open the connection again and the construction should be quite easy. The easiest way of doing so is to leave the original piston inside the engine and to mount the piston extension onto it. This is a very cheap, easy and fast way for the comection. The disadvantage is that the connection is not weight optimized. But since the original piston is made of aluminum and has been light constructed for its task, this solution is suitable and the chosen one. The disadvantage is compensated far by the advantages. One advantage that has not been mentioned before is that the engine also should be run without the extension and with an original shaped piston. In this case, only the extension has to be removed and the piston inside the engine-block has not to be changed!
2.2.3.2. Problem: Oil may be pressed up from the crankcase and soil the mirror:
Running the small car-size-engine, it occurred that oil has been pressed up from the crankcase beside the lower piston rings. The piston rings are obviously not designed for sealing against oil from the under side but only against combustion-gases fi-om above. It never has been much oil, but small particles inside the air. These particles then fasten rapidly on the mirror surface so that the taken pictures are disturbed. In order to get good pictures, the mirror has to be very clean and only very few oil causes a lot of cleaning work. However this problem is expected while running this engine. Different possibilities have been examined before to seal effective against the oil. Most of them failed and it became a real problem. The best and even cheap solution was to vacuum the air around the piston, and so take away the oil. Since good experience is in existence and the same device as for the small engine can be used here, the same solution for that problem is chosen. Most thinkable other solutions have to be tested if they really work well. This gives the reasons to choose the “conventional” way.
74
-— _— 2.3. FMEA-analysis
FMEA-analysis is abbreviation for “Fehler-Moglichkeits- und Einflussanalyse”, an analytically method from [7] to find mistakes and possible risks in the construction. The different parts and functions of the construction are examined, possible failure or damage is assumed. The causes for the damage are determined and the arising risks are estimated and rated.
2.3.1. Association of functions to the elements of the device
The function-stmcture of the device has been examined before in detail, so this has not to be done here again. To get abetter overview, the fimctions of the different elements are shown in the table below. From these functions/elements the critical ones are chosen.
Element Function Stop Disc Stopping the cylinder that it doesn’t touch the mirror when cleaning the glass Water shield Holding the engin e at a constant temperature Under Plate Holding the extension/ connecting the extension to the engine block U-Iron Taking the combustion loads / holding the whole construction Sleeve Protecting the glass ring from excessive loads from the screws Sealing Ring Sealing of the glass ring Rylon Piston-Ring Sealing of the piston/cylinder Rylon Piston Guidance-Ring Leading the piston inside the cylinder Ring2B Taking the loads from the combustion Supporting the glass-ring Connecting the glass-ring module to the extended cylinder RinglB Taking the loads from the combustion Supporting the glass-ring Comecting the glass-ring module to the cylinder head Push Rods Tube Covering the push rods Oil-return Push Rods Extension Lead steering-signal and –force to the valve-rockers inside the cylinder head (Cylinder) Head (Original Part, modified) Leading of the combustion gases Leading of the valves Leading of the cooling-water Plate Mirror Holding the mirror Piston Leading the combustion-force to the connecdng rod Taking the glass plate Mirror Leading the laser-sheet Graphite sealing Sealing and supporting the glass-plate inside the piston Glass Ring Leading the laser-sheet
75
.— ..----- —. .—
Taking the combustion-loads Glass Plate Leading the laser-sheet Taking the combustion-loads Extension-Plate Giving the possibility to change to compression-ratio [email protected] the engine to the right compression-ratio (manufacturing-tolerances) Extension Leading the force from the combustion to the connecting rod Creating space for placing the mirror in the cylinder-axis Cylinder Leading the piston Taking the combustion-loads Asbestos Sealing Supporting the glass-plate inside the engine Aluminum-Ring Leading the force from the combustion Adjusting the gap between titanium-piston and piston extension easily Giving the possibility to use different glass-plate with the same titanium-piston V14X15stopscrew Holding the cylinder when cleaning the glass k15x14 Insex Taking the predeforming force for the screws in the piston Table 9 Elements and Functions
2.3.2. Risk-Analysis
The risk-analysis determines the critical elements of the device which are then chosen for the complete ‘~A-Analysis. The probability of a damage W and the amount of the damage for adjacent elements and systems S is estimated. Elements that are obviously not critical are not taken into consideration.
Element Probability of damage W Amount of damage S stop disc SD 0.05 0.4 Water Shield Ws 0.05 0.3 U-Iron UI 0.25 1 Sleeve SL 0.5 0.8 Ring2B R2 0.2 1 RinglB R1 0.55 1 Push Rods Extension PR 0.55 0.5 Head HE 0.5 0.7 Piston PI 0.1 1 Mirror MI 0.1 0.6 Glass Ring GR 0.5 0.6 Glass Plate GP 0.7 0.6 Extension EX 0.1 0.8 Cylinder CY 0.05 0.8 Aluminum Ring AR 0.2 0.5 M4x15 stopscrew M4 0.6 0.8 M5x14 Insex M5 0.6 1 Table 10 Risk Analysis
76 Risk-Analysis
1 . . ..
t 0.9}.:. ””””;””””””’””}”””””””’”;”””””””””:””””” ““’ ...... I .~4 ......
......
...... :
......
...... ,. ...,, I 0.3 “WAS”:””’””””’”:””’””’”””:””” ““””’:””’”””””’:””””’””’ ...... :..,...... t
o.2F. """;"""''`";""""'"""":"`"""""`"";"""""""""":""""""""" ...... ,...... I 0.1 ...... :....,...... : t
“o 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 Probabilityof Damage W
Figure 19: Risk Analysis
The results from Table 10 are shown in Figure 19. Now the critical elements are easy to see. For a more precise FMEA-Analysis chosen are: ● RI ● M5 . M4
2.3.3. Cause-Analysis
The Cause Analysis helps to find the causes for the failure of an element or a function. The Cause Analysis is done for the previously chosen elements RinglB and M5x14 Insex. For the stopscrew M4x15, this is not necessary. 4 m
N .o. Cause-AnalysisRinglB
I I
Combustion-pressuretoohigh: Torque of attractiontoo high: Material fault Thread tears out
pressure sensor defect Wrong assembly Cavity/ Shrinkage Thread not manufactured engine steeringdefect Material chosentooweak properly Torque of attraction + efiernd reasons —
—
— —
Figure 21: Cause-Analysis M5x14 Insex — . ...—. —
Reasons and consequences for failure of M4x15 Stomcrew
LAe mentioned above, the M4x15 stopscrew is treated specially; no Cause Analysis is done here. The risk for failure or damage is during experiments with the engine. When cleaning the glass and moving down the cylinder, the screws have to be added to support the cylinder. It is now the risk that the screws are forgotten to remove when the cleaning is finished and the cylinder is fixed to Ring2B again. When the screws are not removed and the engine is started, they meet the under plate and burst. Metal-splinter may fall into the gap between piston and cylinder and destroy the surfaces or cause even more damage.
2.3.4. Consequence-Analysis In the Consequence-Analysis the consequences and effects of failure for adjacent systems are analyzed and examined. This is done for the above-determined critical elements. The consequences for failure of the M4x 15 stopscrews are mentioned above.
80
—— — ● U’&F Ring deforms because of high combustion- loads so that high forces act on the glass-ring w Slow deformation of threacl Thread tears out suddenly => glass damage at hiph 1oad Sealing of cylinder head against combustion-gases Sealing of cylinder head fails and cooling water does Worst case: not work longer other threads also get damaged => the failure gets detected and cylinder-head looses by maintaining-personal => glass-damage
03
. Consequence-Analysis M5x14 Insex
I
I I1 .I Screw-failure when mounting Screw-failure when (staticmaq ftilure) running the engine (dynmnic failure)
I I Thread tears out: Screw bursts: New thread has to be worked New screw needed Failure Rets detected Failure does not get detected into titanium piston if possil le Old screws has to be removed out (l?iston sealing failure Sealing losses If not possible: of titanium piston if possible => gas losses => wrong measuring data New titanium piston lfnot possible: pressure 1osses) Other screws get too high loads New titanium piston ! => more screws burst glass-plate damage piston damage cylinder damage glass-ring damage I I eventually vatve damage Thread tears out: Screw bursts:
New thread has to be worked New screw needed into titanium piston if possible, Old screws has to be removed out If not possible: of titanium piston if possible New titanium piston ! Knot possible: New tit anium piston! dEA Product: Bowditch Extension Scania D12 Responsible J.F. Date Sheet 1 15.07.2000
>ment Failure Consequences Reason Previous steps Risk analysis Suggested steps Carried out New risk analysis steps AEBRPZ ‘ AEBRPZ
glB Threadtears Slowdefer.: Extern: out Sealing failrne Combustion 5 3 8 120 Implement a waning sensor in => Failure gets pressure too software to give warning signal detected by high when pressure gets too high maintaining personal Torque of Calculation on the Extra Annotation in Mounting => new Ring has to attraction too thread length 5 8 5 200 Manual to use lubricating oil be produced high and to be very careful when Chose of high quality mounting the screws x 3 6 5 90 Internrd: steel Sudden terrr OUL Thread not Sealing fails mamrfactu- 4 7 5 140 red properly Cylinder head looses => glass-damage!
Ring deforms Glass- damage Extern: Chose of high quality toomuch Combustion steel 3 3 9 81 Implement a waning sensor in pressure too software to give warning signal high Screws distributed as whenpressuregetstoohigh suitableaspossible
Probability of failure E: Probability of discovery B: Importance for the customer Priority: 03 Improbable =1 high 1 consequences insignificant = 1 high = 1000 w Very low = 2-3 mean :2-5 little disturbance of customer= 2-3 mesm = 125 Low = 4-6 low = 6-8 mean/ hard = 4-6 no =1 Mean = 7-8 Very low = 9 hard /customer gets angry = 7-8 High = 9-1o Improbable = 10 very serious annoyance = 9-1o
. ~MEA Product: Bowditch Extension Scania D 12 Responsible J.F. Date Sheet 2 15.07.2000
~lement Failure Consequences Reason Previous steps Risk analysis Suggested steps Carried out New risk analysis steps AEBRPZ AEBRPZ
Internal: Material fault Chose of high quality 2 8 8 128 Use of strain gauges to steel measure not allowed deformations
X-my Ringl B to find material faults
15X14 Screws burst Matiormrv failure: External: Isex New screws needed Gap piston I Calculating the right 6 9 9 486 Usehydr. press for first x 3 2 9 54 If removal of rest extension too gap size assembly not possible big => new piston Torque of 5 4 9 180 Precise instructions in the x 3 3 9 81 dvnaruic failure: attraction too mounting-manual detected: see above High Internal: not detected: Screws too Screw Calculation 5 8 9 360 Use new screws after every x 3 8 9 216 total engine damage weak assembly possible
i: Probability of failure E: Probability of discovery B: Importance for the customer Priority: Improbable =1 high 1 consequences insignificant = 1 high = 1000 Very low = 2-3 mean 12-5 little disturbance of customer= 2-3 mean = 125 Low = 4-6 low = 6-8 mean/ hard = 4-6 no =1 Mean = 7-8 Very low = 9 hard /custonler gets angry = 7-8 High = 9-1o Improbable = 10 very serious annoyance = 9-1o .)
MEA ProducL Bowditch Extension Scania D 12+ Responsible J.F. Date Sheet 3 15.07.2000
lement Failure Consequences Reason Previous steps Risk analysis Suggested steps &irried out New risk analysis steps AEBRPZ AEBRPZ
Thread tears Statiormrv failure: External: out new piston Gap piston I Calculating the right 6 9 9 486 Use hydr. press for first x 3 2 9 54 necessary extension too gap size assembly Wlg dvtmmic failure: Torque of detected: see above attraction too 5 4 9 180 Precise instructions in the x 3 3 9 81 High mounting-manual not detected Internal: total engine damage Thread not Annotation in possible ! manufactu- Mounting-Manual 3 5 9 135 red properly how to manufacture the thread!
[4X15 Forget to Screws meet the Small screws Color the screws red topscrew remove the under plate and are easy to 4 5 7 140 screwsafterburst forget Manufacture gap into under- cleaning the => metal splinter plate so that the SCRWSdo not x o 0 9 0 engine may fall into the gap have to be removed between piston and (less work cleaning the cylinder and engine!) damage the surface
i: Probability of failure E: Probability of discovery B: Importance for the customer Priority: 00 Improbable =1 high 1 consequences insignificant = 1 high = 1000 ul Very low = 2-3 mean 12-5 little disturbance of customer= 2-3 mean = 125 Low = 4-6 low = 6-8 mean.1hard = 4-6 no =1 Mean = 7-8 Very low = 9 hard /customer gets rurgry = 7-8 High = 9-1o Improbable = 10 very serious annoyance = 9-1o
. ,. ..- ————-.-., ;.* . __u- —..-. “. “.:. ~ . .-. —.. /. : , .–..–__ . ... .
3. Forces, Loads And Ways
Calculations on the stresses inside important parts of the device area main-task. The values that are calculated depend much on the input quantities, that means the occurring loads of the combustion. In the List Of Demands a maximum pressure of 200 bar is wished, 150 bar are suitable and 100 bar just tolerable. Calculating on the different parts after some steps it was obvious that it would be difilcult to reach 200 bar. The decision has to be failed if a higher expenditure for a construction standing 200 bar is reasonable or if a construction that “only” stands 150 bar is sufficient. For that reason the calculations are often done both with 200 bar and 150 bar. The revolutions are also important. Higher revolutions result in higher mass-forces on movable-pats. The max-revolutions are given in the List Of Demands with 2000 rpm. They are taken for calculations of the valve-steering system and the piston-movement. Calculating on the piston acceleration, the acceleration-force acts against the pressure- force. So the resulting force gets smaller. Here the resulting-force, which is substantial for the calculations, decreases with higher revolutions. Both cases, with maximum- revolutions 2000 rpm and with minimum-revolutions of 700 rpm are calculated. Results are shown in 3.6.1 Acceleration of the piston:.
3.1. Dimensions and weights
Some basic dimensions and important weights of the engine parts are given in the table below.
Description: Dimensions [mm] Weight [g] Other Connecting rod 255 Stroke 154 Cylinder-diameter 127 Inlet-valve ● Seat-diameter 39.7 ● Weight 156 ● Max. way 14.82mm . Max. velocity 2.7mls 810rn/s2 ● Max. acceleration Spring for inlet- valve ● Weight 103
● Stmhw constant 64300 N/m Valve rocker inlet
● Weight 300 ● Moment of inertia Retainer and clamp 27 inlet-valve
86
—— —— Exhaust-valve
● Seat-diameter 38 . Weight 180 15.07mm ● Max. way . Max. velocity 2.73rnls . Max. acceleration 9oom/s2 , 2springs for exhaust- valve 196 (both together) . Weight . Spring constant 120700N/m Valve rocker exhaust . Weight 374 . Moment of inertia Retainer and clamp 34 exhaust-valve Ok (same for inlet and 140 exhaust) Push rods 200.10, round 130 S,,&I= 7800kg/m3 Table 11: Dimensions and weights ———. -_
3.2. Forces on the valve-train and the push-rods:
In the first part 3.2.1 preliminary calculations on the forces that appear on the valves are done. The results are used in the part Product-Development-Strategy for choosing a suitable design. They cannot be used for precise calculation like for example on the push rods. A more precise calculation, using a Matlab-program, follows later in parts 3.2.2,3.2.3 and 3.2.4.
3.2.1. Loads on the valves:
The loads on the valves are divided into loads from the combustion and loads from the acceleration. The total force is calculated by superposing them. The most critical case regarding inlet or outlet-valve and acceleration has to be determined.
● max valve-velocity: vi~l~,.m = 2.70 IdS vefiau~t-m = 2.73 m/S
● max valve-acceleration: ai~.,-~~ = 810 IIdS2 - %Xba”st-nlax = 900 ~s2
● max valve-opening: xid.t-- = 0.01482 m X~fi.”$~-m = 0.01507 m
● static force on the valves from the combustion: (the worst case is reached when the exhaust valve opens) Q suppose: p = 15 bar= 1.5Mpa; when exhaust valve opens @ dvdv. s 40mrn
Q &+Lcomb = PA = p“md2/4 ~ ~~
● static force on the valves from the spring return 1.) inlefi + k= 64300 N/m (spring-constant) * s = 0.015m (maximum valve opening) b FpE,id~t= 500N * 20N (reloading force from the spring on the valve) @ Roriruz.inlet-= Fpre,inlet + k-s G SZON + 9GSNs 1485N
1.) exhaust: b k = 120700 N/m (spring-constant) b s = 0.015m (maximum valve opening) Q FPm,efiau~t= 665N * 25N + 280 * 20N (reloading force from the spring on the valve) + F__pnng,e*au,L=Fp~,eti.u,t + k-s = 990N + 1810= 2800N
88 . dynamic force on the valve from the acceleration For calculation of the dynamic mass of the springs, 1/3 of the spring mass has to be chosen. A calculation for this is given in [2, B38]. An approach is done by using the kinetic energy. The displacement is given as a function of the spring- coordinate u(x) = (S/1)x (s: max. elongation; 1: length of spring). Then it’s obvious that dnz= (nz$Ptin~/l)&. The kinetic energy then is formulated by:
Here k = 1/3 which means that 1/3 of the spring mass has to be added to the oscillating mass. In these calculations here 50% percent of the spring mass is chosen. This maybe too much, but it’s taken because of possible inaccuracy measuring the weights and so the calculations are on the “sure-side”. Nevertheless the valve springs play a subordinate role to the whole force on the valves.
1.) inlefi 4 rn,d,ts 232g (mass of lvalve with Y2 spring-mass and retainer) @ %d.,.M = 810 m/s2 (maximum acceleration of valve) Q &C,hv.l,. = 2 mid~t”%d~t-m = 376N (2 inlet-valves)
2.) exhausti @ Qfi..,ts 312g (mass of lvalve with Y2 spring-mass and retainer)
* aexhaust.nlax = 900 In/s’ (maximum acceleration of valve) Q &cc.ExVaIve = 2 mefiau5~”ae*~”q-- = 561 N (2exhaust valves)
. total force on the valve: 1.) Fsm~CO~b= 1900N
2.) Fspring,etiaust = 2800N (highest force for exhaust valve) 3.) highest acceleration-force: FaCC,fivdVe= 561 N
@ Fvalve-max= Fsta~comb+ Fspri.g+FaccExVaive = W
The dynamic force occurring from the acceleration doesn’t play the most important role so it is only examined for 2000 rpm, not for 700rpm. ,. , /,,. . .-. —. .—’ ‘.+ . ...__{_.+ ~__ —. --,. - -.
0.01!
0.0 ......
F ~ 3
0.00! :. .,,, ,.,
( ) 300 350 400 450 500 550 600 650 angle
exhaust valve — way 0.016
. . ... 0.014
0.012 ......
0.01 ..
F ~ 0.008 .:, . 3
0.006 .:..
0.004 ......
0.002 ....
0 100 150 200 250 300 350 400 450 angle
90 exhaust valve — velocity
-“50 100 150 200 250 300 350 400 450 angle
exhaust valve — acceleration
800 -
600 -
400 -
*OO ...... j\ ...... t II1 “1
I
-600 ‘ I I I I I 1 ! I 50 100 150 200 250 300 350 400 450 angle —. —. .—. —
3.2.2. Acceleration-force of the valve-rockers
The valves rockers are different for exhaust and inlet-valve. Since the forces on the exhaust valves are the higher ones and even the valve-rockers are bigger for them, what results in a higher moment of inertia, this one is to be calculated. The motion is rotational and the load is determined with the moment of inertia. To do so, a mechanical model of the valve-rocker is shown in Drawing 22.
Drawing 22 Model for the exhaust valve-rocker The dimensions are: q = 17rnm r,= 28rnm h~n~= 16mrn (height in 3.dimension) h~id~= 12mm
Moment of Inertia:
For calculation, the valve rocker is divided into 3 parts: two rectangular parts on the outer sides and the ring in the middle. The short rectangle in Drawing 22 is located on the side of the push rods (Drawing 22 left) and the longer rectangle (right side) on the side of the valves. With a density for steel of p = 7.8 kg/dm3 the masses of the parts can be calculated. The formula for the moment of inertia is n’z(l-az+ ;2 ) . for a tube: Jting = 2 nz(az +bz) 0 for a rectangle: Jr,ctigi, = (with a and b dimensions 12 in the plane rectangular to the rotating axis)
According to the law of Steiner it is true for parallel axes two-dimensional, that Jx=Ji+d~”m Here d, equals the distance from the centre of gravity of the part to the rotational axis.
92
-. ———— The moment of inertia can be calculated now for every part and then be added, using Steiner’s law.
Part Rectangle left Ring Rectangle right Mass @cg] 0.079 0.194 0.108 Moment of inertia 34.5 105 47 Kg. mm’] Distance to centre 43 0 51 of gravity [mm] Moment of inertia 180.5 105 328 to the rotational axis pcg.mrn2] Total moment of 613.5 ~g”mm2] = 6.135”104 ~g. m2] inertia Kg. mm2]
Force out of the moment of inertia:
The dynamic force from the moment of inertia can now be calculated with
M= J” a = J” ~. The following drawing gives the geometrical relation between the way of the push rods x and the angle of the valve-rockers (p.
Drawing 23 Geometrical relation between push rod and valve-rocker The push rod moves the distance x and the so the valve-rocker rotates the angle q. For .. the calculation of the force out of the moment of inertia the angle-acceleration q has to be known. The angle Q can be written as: ,, .—, -— ——....’ ,.. f” ---- . ,’ ‘. -. J _-_. -.u...u. .,,
%) ~ sinq=— p=arsin ~ ll+r ll+r
The first derivative of (pis written as:
Now the second derivative ~ can be calculated.
.;= m For the calculation of @ the values for a,v and x of the push rods have to be taken.
The distance, velocity and acceleration of the valves are given. This gives the ll+r _— ‘push– lZ i- r ‘“”1”’ ll+r distance,velocity and acceleration of the push-rods to: Vpu’h= — lZ + r ‘“”1”’ ll+r a ‘“h = ~av”’ve
(lZis the corresponding length to 11in Drawing 23.)
Without proof it can be stated that: M = F. 1 and so the resulting force on the push-
rods from the mass-inertia is: Ftit,ti~ m
3.2.3. Loads on the push-rods:
The loads on the push rods are not as easy to determine as the analytical forces before. They are time-dependent. The acceleration-forces then can decrease or increase the total force on the push rods. In principle these forces can be divided into the force
94 resulting from the combustion-pressure on the valves and the acceleration-forces on the valves, the valve-rocker and the push rod. Since the maximum forces do not appear at the same time all together, the forces are added in a Matlab-prograrn and the resulting force-pattern is determined. In order to get the right sign the positive x-directions are given in the drawing below.
)%c
‘x
Drawing 24 valve-steering geometry
Using the equilibrium-condition at place L
M inertia + ‘valve d2-Fpdl=0
Minema i- FVa[ve“11 I+ FP= 12
TO this force the dynamic force of the push-rod acceleration FP,ti,tia has to be added to get the force on the bottom of the push-rod.
M inertia + ‘valve “ ‘1 + F Fp.max = p,inerria 1?
While doing the programming it has to be taken care that the pressure forces on the valves are only considered while the valve is opened. Otherwise, the valve-seat takes the forces ! ,’ —— ..— —...... , ---
The force that has not been determined now is Fp,i~~ti~.It is an acceleration-force and is proportional to the mass. Using the chosen coordinate system, its maximum is at the .. bottom of the push rod. It is determined to Fp,i~~tia= nz,O,~,. ~. a(f).
For the calculations against buckling the maximum dynamic force of the push-rods is assumed. The force is calculated at x = 1to FP.,i~emi~- = m,O,~,. a(,)
3.2.4. Deformation of the push-rods:
The push rods deform axially because of the loads of the combustion and the acceleration of the valve-mechanism. The deformation can be calculated, using the formula Al= f%= ] ~N~~l + a,AT dx. X=l) The push rods are not expected to warm up, so the second term can be neglected. Also the area of the push rods doesn’t change over x so that A(X)= A = const. The normal force N divides up in one part that is created inside the push rods because of their acceleration N1 = FP,i~~m”~.This part changes over x. The other (higher) part Nz that results from the combustion-pressure and the dynamic force of the valves and valve- rockers is constant over x.
‘ N,(r,,) N2(,)“ 1 * Az(r)= ~=o~&+f E.A
Equation 1 deformation of the push rods
The cinematic relation between valves and push rods is shown in Drawing 24.
xl –=’”123Hlz+r l,+r
All the input quantities for the calculation have been determined now. The deformation of the push rods is calculated in a Matlab-program, solving the equation above and determining the max-value over the combustion-process.
11= 30mm 11= 46mm r= 28mm 10,g= 230mm AOngiml= 7.854 .10-5 mz
96
— Results:
The following figures show the above-determined forces and the deforrnation, varying with the crti angle.
Acceleration and spring force on the exhaust valves 2000rpm, 200 bar 5000 1 1 1 I I I 1
4500 -
4000 -
3500 - ~ g g = 3000 - 0 al p o L 2500 -
2000 -
1500 -
t t 1 ! ! I t 1000 50 100 150 200 250 300 350 400 4 o CAD after TDC
Pressure-force on the exhaust valves (Fpvalve): 2000rpm, 200 bar
‘ooo~ 6000
5000 -
~ a) 4000 - ~ m > E
83000 - 2
2000 -
1000 -
I o I 1 I t I 0 50 100 150 200 250 300 350 400 450 CAD after TDC ———
Force on the exhaust valves (Fvalve,preas+acc): 2000rpm, 200 bar 60001 I 1 1 I , 0 ,
5500 -
5000 -
4500 -
~ 4000 - $ ?j = 3500 - 0 ~ LF3000 -
2500 -
2000 -
1500 -
1 1 1 ! I I 10001 ! 50 100 150 200 250 300 350 400 450 CAD after TDC
Force on the push rods (N2): 2000rpm, 200 bar
‘ooo~
3500 -
3000 -
~
$2500 - .$ 3 n ~ a)2000 - P f
1500 -
1000 -
I ! 1 I I I I 500 I 50 100 150 200 250 300 350 400 450 CAD after TDC
98
— Deformation of the original push-rock 2000rpm, 200 bar
‘“””~ 0.05 -
0.045 -
0.04 -
‘F ~ 0.035 - E = g 003 -
E
0.025 -
0.02 -
0.015 -
/ 001~ I t I I ,h 50 100 150 200 250 300 350 400 450 CAD afterTDC
For these results it was assumed that the total force from the combustion is conveyed to the valve-seat and not further into the push rods as long as the valves are closed. When the valves open the combustion force is added. Because of this, the force “jumps” at 60 CAD, when the valves open. This is of course not 100% correct, because also the valves deform elastically (what has been neglected here), what would result in a slower rise of the force. Nevertheless a more precisely examination of all forces and deformation results in a very high work-expenditure. The results are seen as precisely enough.
. The highest deformation occurs at 232 CAD to AIOn~i~.,,m= 0.05nvn. . The highest force on the push-rods occurs at 232 CAD to 3630 N
The first attempt to dimension the extended push rods is to demand that the deformation of the extended push rods equals the deformation of the original ones. With this assumption, the area is calculated to:
N* “len A&– 6“ 1: “apuh ‘lOriginaI,max “ E – 2
1.X, = 405.5mm -!-lo,g =635.5 mm
Since the area now is a function of the acceleration and of the force N2, the max-value also is calculated inside the Matlab-program to: ———— . , ______. .
Ae~ =20.7 .10-5 m2 at 232 CAD after TDC
The weight of push rods with this area is calculated to ca. 1 kg each. Since this is too much for the camshafi the requirement of having the same deformation has to be changed. The maximal deformation is determined to 12/100 mm. This gives a cross- area of A .,, =8.95 -10-5mz.
3.2.5. Buckling of the push rods:
The necessary cross-area of the push rods has been determined above. It is now to calculate the push rods against buckling. For that task the maximum force on the push rods is calculated to iV1+ Nz = 3550 N. The higher mass force of the push rods reduces the total force. Examining the acceleration-course shown before proves that. The acceleration force inside the push rods is determined as before in Equation 1 to 1 be: F.CC&= –m& “ap . This equals 50% of the acceleration force and so the force in 2 the middle of the push rods, where the critical place is. This assumption is suitable because higher acceleration forces decrease the buckling-risk. So the calculations are “on the sure side”.
The push rods are designed as a tube with an outer diameter of 16 mm and an inner diameter of 12 mm. The cross-area equals the one calculated above. The area moment of inertia is I=x. R~”t=2155mm4 with R~ = 7.0 mm t =2.Omm
4 different cases of loads are shown below. The push rods are joined like case 2. This determines the critical force to:
The maximum load has been calculated above to 3.55 KN. This corresponds to a factor of safety of 3.11.
The weight of the extended push rods is 440g each!
100
— F 1?. T IL If
3.3. Coarse calculations on the glass ring:
● !+@@ a view of 50° crank-shaft-angle required in the List Of Demands. stroke = 154mm @ 180° fl @50/180*154 mm = 42.8mm
An additional height of 9mm is added for the compression-volume (that would be the worst case without any combustion-bowl inside the piston), what dimensions the height of the glass ring to: hGl~, =51.8mm
The height is discussed with the instructor Bengt Johansson and glass-specialists from Heraeus, Hanau [3]. Since the glass is very expensive and the mechanical properties are worse the bigger the part gets, the height is limited to 30 mm. This equals, when also considering the compression-volume, 24.5° CAD. In the case that the piston has a combustion-bowl in the middle and goes up until 1.5 mm under the ceiling (List Of Demands), 33° CAD are viewable.
● A precise structural integrity assessment for the glass ring is due to shear stresses inside the glass not so easy to do. The appearing loads can be used for a Finite- Element-Calculation of the ring. It is not done in this work. For a coarse dimensioning the ring is calculated as a tube under inner pressure. Simptiled only radial-stresses occur. This can be seen in Drawing 25.
The force from the combustion is F = p.2. r. h (his the height of the window). From the equilibrium-equation follows: F F 2“r. h r Crp =-= .— A 2(R–r)lz =p.2(R-r)h=p R–r. —.——. _.__— —— — —.——
Pcmtibu5tiorl ‘Q u 111
Drawing 25 Stresses on the ring
The maximum allowed tension-stresses for glass are 50 N/mmz ([3]). Now the Radius R and the ring-thickness [t~dl = R-r] can be calculated to:
R=r o(1+ % with Oq= Ctin = 50 N/mm* Op Q Rtilw~ar =77 mm tw~l= 13mm 6 ‘Min,lsf)bo,=83 mm t.dl = 2omm * Rti2mbar =90 mm twdl= 25.5mrn.
As a comparison the stresses of the glass ring that is used for a smaller car-size-engine here at the institution are calculated. The stresses for our ring should not be higher than for this one. Since no problems occurred with the small ring it can be expected that the new one also sustains the loads.
The car-size ring is made for a pressure up to 100 bar. This gives a radial stress of r 90 0 .— = 100bar .— = 45 N/mmz , p.small-ring = p R–r 110–90 which lies little lower than the assumed one before.
Using this value the wall thickness are: Q twdl,lCObar= 15mm @ tw~~,~~ob~= 22mrn + wall,200bar= 29mm.
The height of the ring is now dimensioned to 30 mm. This higher value takes care of the bigger size (decreases the strength of the material) and the neglected appearing shear stresses.
102 3.4. Calculations on the mirror:
The mirror is expected to give the biggest possible field of view through the piston. It is mounted in an angle of 45 degrees, so the shape of it is an ellipse. The size and the equation of the ellipse are calculated here:
z A
lm
x ➤ Y
Figure 25: Mirror-geometry
The required geometry for calculations on the mirror is given in Figure 25. The cylinder-axis corresponds to the z-axis in the drawing. The X/y-plane is horizontal; rectangular to z. The mirror is mounted under an angle of 45°, it’s axis corresponds to the m-axis inside Figure 25.
The circle-equation in the X/y-plane is X2+ y2 = rz (l). A coordinate-transformation to the X/m-plane is done by projecting the y-axis:
y=m. cos45” =;. fi. m(2) inserting (2) in (1) gives:
x2+ ~.W.rnl’=r2
1 x2+ —.nz2=r2 (3) 2 Equation (3) gives the geometry for the mirror.
Its normalized form is: m
This is a parabola with the axis r and &. (r is the radius of the mirror)
103 ..—.—. .——. .— _____
3.5. Calculations on the sealing:
3.5.1. Sealing of the piston:
The chosen gasket for piston sealing is Statotherm@Item-Nr. 9591-1, graphite-sealing with stainless steel insert from Burgmann. The chosen thickness is 2 mm. The sealing is done at the upper clamp of the glass. The appearing forces are shown in the mechanical model in Drawing 26. In this part also the necessary screw-force and the necessary size of the gap lgap between Titanium-Piston and piston-extension for prestressing the gasket is calculated. An important demand is that this gap is always closed, when the engine is running. Otherwise it could happen that the loads on the glass are not distributed equally. Therefore the tolerances inside the piston-crown have to be determined precisely. The maximum-tolerance between the parts is set to 5/100 mm. This is reached by manufacturing the titanium part and the glass and measuring them precisely. The aluminum part then has to be manufactured according to the other dimensions.
For calculation, the under “gasket” or better-called “glass-bearing”, made of asbestos, is assumed as not deformable in the calculations.
Fp,piston 4
.—. 1 19P l-l F Seal,.h+Fp,gj- X F+ 7 screw
Drawing 26: Forces in the piston crown
104 The integrity assessment for the sealing and the prestressing-forces are calculated after [2]. The different steps are shown below:
Input muarneters:
102 mm Mean gasket diameter in mm
5mm Gasket with in mm
20 N/ mm2 Minimum contact pressure in Nhn.rn2
120 N/ mm2 Maximum contact pressure in installed state in N/ mmz
M 1.3 Loading coefficient
26 N/ mmz I Minimum contact pressure in operating state in N/mmz
> 6B0 100 N mm2 Maximum contact pressure in operating state in N/ mmz at 300”C P 200 bar Max. pressure to seal
sD 1.3 Safety factor
Fp,piston 88.4 ISN Max. force on the titanium-part of the piston, (pressure and acceleration, 200 bar and 700 rpm) Fp,~l&S 161.6 KN Max. force on the glass-part in the piston, (pressure and acceleration, 200 bar and 700 rpm) FSd,tin 54.2 KN Min. sealing-force required (FDBU)
Titanium Glass Aluminum Length (Drawing 1A= 24mm lT = 61mm lG =35mm 26) E-module EA = 72.2 KNhd ET= 105.2 KN/mm2 E~ = 70 KN/mm2
~T A = 24.510-6 K-l ~T,T = 8.510-6 K-l ~T.~ = 5.910-7 K-l Area (for AA= 2221 mm’ AT,l = 3308111111: AG,hdp =8221 mm’ deformation- AT,2= 1774 mm’ I calculation) (piston rings at A~.z !) Temp.-difference AT. = 100[K] 3K 4K ATT=— — .X2 ATG=— — . X2 490 [1mm2 405 [1mmz + 100[K1 I +11O[K1
. FDVO:accepted sealing force in installed state iv F~vo=dD. m. bD. avO=102mm. z.5mm.120— mmz F~vo = 1921CN ———— . . .. -. ——-_
. FDBu:seting force necessq inoperation (=Fs~.~in Drawing 26) N ‘Bu ‘m” P= 26_ mmz N F~Bu =dD”z”bD-~Bu “SD =102mm”x”5mm”26—. 1.3 mmz F~Bu = 54.2KN
● FDBO:accepted sealing force at operating temperature N F*BO=dD”r”bD”crBO = 102mm .Z. 5mm. 100— mm2 F~BO= 160KN
Equilibrium-conditions:
The equilibrium-conditions me according to Drawing 26 formulated to:
and
Influence of the Temperature: Since the temperature influences the strain (the glass must never loosen) and therefore also the forces in the piston, it is taken into consideration here. The temperature difference for the aluminum-ring is determined to be lOOK. For the titanium part a quadratic temperature-difference-distribution with 130K at the top and 100 K and the bottom is determined. Also for the glass-part a quadratic temperature- difference-distribution with max-value 130 K and rein-value 110 K is used. This gives a temperature-distribution to:
Titanium: ATT=a”xz+b ; ()< X<70 (x is the coordinate from the top to the bottom) ATT(O)=100K=b ATT(75)=130= a”70z+b 3K @ a=— 490 ~1mm- 3K ~ ATT=— ~l. x2+100K 0< XS70 490 mm- -Glass” ATG=c. xz+d ; OS XS45 (x is the coordinate from the top to the bottom)
106
— ‘TG(0) =lloK=d ATG(a~)=130= a”452+d 4K @ a=—— 405 mm2 I 4K @ ATG=— — 1.x2+110K 05X545 405 mm2
Strain of the different txu-ts:
With the normal-force N, the strain can generally be formulated as: du N —= —+cx~ “AT dx E*A or:
I N Hdu=’ —a!x+cx~” AT”& 0 o E*A J
The integral consists of a first part, which describes the deformation because of outer- forces and a second part, which describes the deformation because of temperature- differences. It is solved for Titanium to:
I 161 NT “ 1~ lK Al~imiu~= .— .13 +1 OO[K].1 -% “4 + a’” 490 mmz J , and for the aluminum to:
N~”l~ +a Al ‘,~ “100[K]” 1A I ‘1”-”” = EA “AA
To get more precise results, the grooves for the piston rings inside the titanium-part 1’ are also taken into consideration in this calculation. The factor in the formula ET “AT above is determined with the following drawing. ...— —. . .
Figure 26: Mechanical model to determine the stress-deformation of the titanium-part
F~l _ F~ 13.5 47.5 Aln,m,$,,,,, = — – — = FT “1.6398 ”10-4M KN ‘T* ‘T ~(118’ -108’)+ ;(126’ -1082)
This formula gives the deformation of the titanium part due only to the outer load FT!
The deformation of the glass because of outer forces cannot be calculated analytically. The problem has been solved by using the Finite-Element-Method. The deformation of the glass because of the temperature-influences can be calculated precisely enough. They are just to be added.
Al~,,,,~,,,, = 0.015mm
35 4K AlG1w~41G,~W,~~~$i- CZT,G— — .13 +llOIK]d 3”405 mmz J , “
Now that the deformation is known a help-area that takes all the stresses inside the glass is calculated. It is needed later for calculation of the necessary predeformation- force.
NG “ lG,~ Al=,,m,,,,, = — EG -AC .O,n ‘‘ - ,.-..l- “~ ZG,.is the mean length of the clamped glass and is determined to 35-1-45 1 = 40mm. G.m = 2
108 That this area is suitable for calculation can be seen in the following figure that shows the appearing stresses inside the glass. The dark part takes most of the stresses and can be used as a heIp-area, where I.inem-elasticity is assumed later. The critical place at the top left corner can also be seen well. Here a radius of lmm helps to reduce the stresses.
‘-+-+++-++-+-%”--5.: i I i f i ‘-i f 1 t i
1
Figure 27: Finite-Element calculation on the glass-plate
The sealing-deformation is strongly not linear (Figure 28). After a higher preloading force a linear deformation is assumed. For the interesting part at maximum-load E is set to 1.2 KN/mm2. A temperature-deformation is neglected. Then AI can be calculated to:
(use l~P= 0.94mm, predefomned! !) .——..._ —-_
50
40
30
V&fQrmurlg . 20
?0
Ruckfederung o 2 5 10 2(3 50 l(x) 200 Fl&henprewng N/mm2
Figure 28: Compressibility of the graphite sealing
Calculation-stratemc Since, because of the non-linearity of the graphite-sealing analytical calculation is not possible some iteration had to be done. Here only the results are shown. The E-module for the sealing is calculated at O= 34 N/mm* (minimum sealing-force) to EM = 560 N/mm*. At cT=60 N/mm* (expected maximum value) to E60 = 1200 N/mm*. This value is much lower than the E-modules of aluminum or titanium. For calculations an E-module of 1.2 KN/mm2 in the interesting field while running the engine is assumed. In this state the sealing still has a spring-return of 5%, which equals 0.1 mm. This is enough to withstand all the heat and load deformations while running the engine. A a-value of about 60 N/mm* equals a load of 96 KN onto the sealing. The calculations are split into four parts. In part 1 the engine is warm and running. Part 2 is a cold engine when mounting it. In part 3 a warm engine is examined during the intake and exhaust stroke. A cold engine under full load is calculated in part 4.
Part 1), warm engine. runnk under combustion When the engine is warmed up, the loads shown in Drawing 26 appear. The load on the sealing has to be at least the required sealing-force for every state! Examining the influence of the manufacturing tolerances, they increase the predeforming-force and the force on the sealing. Since the minimum predeforming- force is to be calculated here, they are not added to this calculation. This gives the force X to O. The tolerances are added to the length of the gap later and then the force X can be calculated. The extension of the parts can be calculated now.
110
.—. Nmdumwm = -Fp,pi~tO.+ F~~,ti. = -34.2 KN
NG1m,,wm = NAU,W- = Nsa.g,wm = Fp,glass + Fseal,min = -215.8 KN
Nsting,warm G -Fseal,min e -54.2 KN
This gives the deformation from the appearing combustion forces and the temperature-difference to:
Al~imiu~ = 0.04624nvn
and
Al~\~, = -0.01265nvn ,
Al Alu minurn = 0.0265nvn .
Al,,~\in~= -0.02707nvn . (< 0.1 mm sealing spring-return) d
The inner part is now compressed &lass+ Ah.+ ~sealing] = 0.01322 mm. The Outer titanium part instead extends Al~tiUrn = 0.04624 mm. The size of the gap between the titanium and the piston-extension, which has to be closed when the engine is cold, is chosen equal to the sum of these values. The material has to be predefonned at least this much!
*1 gap,load = 0.04624mm + 0.01322nvn = 0.05946nvn —
Part 2) cold engine, standstill Calculating the necessary force for the above-determined necessary predeformation lg,p,t.ti, a cold engine at standstill is assumed. It has to be taken into consideration that the deformation takes place in all parts; glass, aluminum, sealing and titanium. Because no outer forces are loaded to the device in this state the only force in the different parts is the predeformation-force Fsa.w. Taking the manufacturing tolerances into consideration, the gap length has to be increased with 0.05 mm. On the other hand may the sealing still be predeformed because of the manufacturing-tolerances with 3/100 mm (1.5 % of the sealing thickness). Now the determinable gap length and the resulting predeforming force can be calculated.
3 lgap= lgap,,oad+ lgap,,o,erm,e,–—mm = 0.05946mm + 0.05mm – 0.03mm = 0.07946mm 100 F “1~ F=r.,v,~.al“1A F=,~v,~..l“lG = F lT+lA+lG+l~ lgap= screwseal + -1- — — screw,seal ETA= EAA~ EGAG,~~~ E~A~ ‘Tub E~ “AA EG . AG,~.h ) . .. —— -. .—:..-
1 * F~c,Ov~~Q1= gap —=91m1= 1A ● lG 1~ - — — +— ET~ + E~A~ EGAG,~eh ‘SAS
With the above-calculated necessmy gap-size l~@oadand predeforming force the sealing is suitable to handle all the temperature and load-deformations while running the engine. Since this is the biggest gap-size of part 1) to 4), it determines the minimum required force.
The manufacturing tolerances increase the gap by 2/100 mm. This results in a higher predeforming force of X = 23 KN in this state. It equals a stress increase on the upper sealing of 14 N/mm*.
A linear material-law for the sealing was assumed. This assumption is only acceptable after a pre-compression. The pre-compression is taken from the diagram (Figure 28) at a force of F~CEW,~A=91 KN what equals a o-value of 57 N/mm*, to 53% of the sealing-thickness respectively 1.06 mm. This value has to be added to the gap-size. After the first mounting of the screws it does not appear any more. The total gap is now determined to &a..~o,~= l~a~,load+ 1.06mm = 1.120 mm
For manufacturing the parts the gap-size l~a~,lo~dhas to be used to determine the precise length of the parts and manufacture the aluminum ring. (l~.P,tOti appears only before the first pre-compression of the sealing!)
Since the predeforming-force does not exceed the assumed force on page 110, it was correct to use a working-point of o ~ 60 N/mmz.
Part3) Warm entine running during intake and exhaust-stroke:
The force that is needed here is the maximum force on the upper-sealing. It occurs during intake and exhaust-stroke of the engine. The deformations because of temperature-influences remain, but the pressure loads are seen as zero here. All the deformations, calculated only with temperature-influence and minimum sealing-force are:
<aniumJemp = 0.0607 mm
M4huninum,ternp = 0.0507 mm
~glass,temp =-0.00095 mm
@yaphitJemp =-O.0271mm (< 0.1 mm sealing spring-return) #
Now the inner part extends with 0.02265 mm. The gap only because of the temperature and minimum sealing-force would then be:
+ Igap,temp = 0.0607 mm – 0.02265mm = 0.03805mm
This gap is smaller than the one in partl). So the gap in partl) determines the preloading force.
112 Since this gap is less than in partl) an additional force X appears. This force is important while calculating on the glass-plate. It increases the appearing load. It can be calculated, using the difference between the gaps. The manufacturing tolerances do not have to be taken into consideration here because they appear in both gaps. Subtracting both gaps form each other also the tolerances are subtracted! Al~aP= 0.05946rnm -0.03805 mm = 0.02141 mm.
Algap x=— = 24.5KN lTIA+lG+l~ — — E~~ + EAA~ EGAG,hlp E#~
This equals a stress increase on the glass plate of 15 N/mm*. The force on the glass- plate in this case is then F,~,tirr + X = 54.2KN + 24.5KN = 78.7 KN.
Part4) Cold engine running with full load worst case:
In this state the sealing-return and the remaining sealing force are calculated. The sealing return must not decrease the maximum-value and the sealing force has to be bigger than the minimum sealing-force. The way of calculation is the same as before. The deformations due to the pressure and minimum remaining sealing-force are:
AITitaniunpressure = -0.0056mm
M41.minumpressure = -0.0322mm Alglass,pressure = -o.o131mm &l-aphiLpressure = -0.0271 rnm(< 0.1 mm sealing spring-return) #
This is also a control that the sealing has enough elastic-return! The gap only because of the pressure and minimum sealing-force would then be: * lg,p,pm,w = o.0668rnm
Since the predeforming-gap equals Igap,load = 0.05946 mm the gap-difference Algapgets Algap= 0.06681mn -0.05946 mm = 0.0073 mm. The sealing force in this case then is decreased with:
Algap AF cold,seal =— = —8.4KN M lT lG s —+ ~+1 +1 E~A~ EAA~ EGAG,hh EsAs
This means that instead of the required sealing-force of 54.2 KN only about 46 KN remain. Fpr=SuE,S~in this state is smaller than the minimum-serding-force. This means that the sealing may not seal properly!
The only possibility to secure sealing also in this case is to raise the predeforming- force. This in return results in higher loads on the glass-part which absolutely have to be avoided. Since the engine is expected not to be driven cold and with full-load and
113 .—— . —. .-. . __
moreover only skilled workers are going to use it, the predeforming force will not be increased! Another argument for not increasing the predeforming-force is the chosen security for the sealing that is 1.3. Neglecting this security, the sealing withstands the loads even in this case. Also bad manufacturing tolerances increase the gap and so the predeforming force. These calculations have been done for the worst case of sealing. That means with the minimal gap and the best manufacturing at nominal dimension.
Nevertheless this drive-case should be avoided!
Check for the gasket loadability in the installed and operating states:
The highest loads on the sealing are at engine standstill because of the preloading force F,CEW.Since the engine still may be warm at standstill, the parameters at operating temperature have to be taken.
F Screw,seal =91 KN 3.5.2. Sealing of the glass-ring: The sealing of the glass-ring is the same type as the sealing for the piston. Here also graphite sealing is used. The sealing also acts as a bearing for the glass-ring! The calculations follow in principal the calculations above. The influence of the temperature is not calculated here. This is done when calculating on the screws for clamping the glass-ring. Like mentioned above the sealing supports the glass-ring, which demands a large area for sealing. That in return results in a high minimum sealing force. Since the area is so large and the engine is not often run with 200 bar pressure the safety factor SD is chosen to SD = 1.15. 0 FDVORaccepted sealing force in installed state (d: - d:) =Z1662 –126Z .120 N FDVO =x 4 ‘v” 4 mmz F~vo = 1.lGN o FDBU:sealing force necessary in operation N ~Bu=m-p=26_- mmz =Z(d: -d:) .S _Z166’–126z N DBU “crBu D– .26—. 1.15 4 4 mmz FDBU= 276KN k 114 –—— . FDBO:accepted sealing force at operating temperature N o ~~ = loo— mmz i 1662 –126Z N =Z(H2) “ FDBO ‘o~o=n loo— 4 4 mm z F~~o = 917KN 3.5.3. Sealing of the upper-plate: For sealing of the upper-plate and the cylinder head against the combustion gases, also graphite sealing is chosen to be most suitable. Good experience with o-ring sealing made of Viton 75 is in existence here at the institution, but the graphite sealing is smaller and so more suitable for this task. The sealing that is used here is Statotherm@ Item-Nr. 9591, 0.5mrn from Burgrnan. Since the sealing is wanted to be formed here by hand, a sealing without metal-sheet is chosen. The safety factor SD is chosen like above to SD= 1.15. . FDVO:accepted sealing force in installed state N CTvo= loon; mm =Z(d:–d?) 138.52 –131.52 .1OO N F ‘ Ovo=n DVO 4 4 mmz F~vo = 148iKN ● FDBU:sealing force necessary in operation (m= 1.3; Burgmann data-sheet) crB~ =m. P=26~ mm (d: -d: ) 138.72 –131.72 .26 N ~ 15 F =Z .OBU. SD=X —. . DB1.J 4 4 mm z . FDBO:accepted sealing force at operating temperature N aBo = loon (~z~ d: ) . ~Bo =Z138.52–131.52 N F =7T . loo— DBO 4 4 mmz FDBO= 148KN This sealing force FDBUalways has to remain onto the sealing. It is used to determine the screw-force in part 3.9.3 and equals the force Fmti. 115 .—...... —— Commission of the sealing: It is now to determine the compression of the sealing that occurs when fastening the screws. With the determined load, a compression of 46% occurs (Figure 28), which equals 0.23mm. That means for the sealing, a space of l$,~lin~= 0.27nvn has to be retained. When fastening the screws further more, the force is led via the plate directly into the head and no bigger force occurs onto the sealing. This construction somehow equals the one inside the piston. 3.6. Loads on the glass-piston-crown: Decisive quantities for the mechanical stresses that occur inside the piston are the acceleration of the piston, the combustion-temperature and the combustion-pressure. The acceleration and the pressure course are examined individually fust and then the worst case, superimposing both values, is used for dimensioning. 3.6.1. Acceleration of the piston: The piston acceleration can be easily calculated from the piston path. Drawing 27 shows the necessary dimensions. Drawing 27: IFiston-movemet 116 — The piston-way can be written as: s=r.cos~+~m ,([5, page5-101) With ~ = fo(,), the first derivative gives the piston velocity to: ds = ds dp 1 r2”2”p”sinpcosq ——=–~rsinp–– ‘% dpdt 2~ To get the acceleration the second derivative is done. The velocity is divided up into two parts: Doing the second derivative one more simplification can be done. Since q changes .. linear with the time t, ~ = ~ is a constant and @ is O. So all the terms containing p can be neglected. .L 13 al =–rp cosp U’v —v’u A fraction y ~ is deviated with the formula y’= v V2 1“ u = ——z r2 psin(2~) u’= -~. 2. r2 fj2 cos(2p) = –r* fj2 cos(2p) _r2 ~2 c0s(29)” (12– r2 sin2 fp)+~rz sinz(2~) a2 = (1’ - r2 sin’ ~~” . .— ___ .. ——-— - .2 (12 - r’ sin’ p)cos(2q)+~r’ sin’(2p) * a==l +az=–r~ cosp+r (12 – r’ sin’ fo~” For every angle the acceleration can now be calculated, with ~ = 2nN = const. (N: rev. per second). This is done with a Matlab-program. The courses are plotted on the following pages. Here it easily can be seen that the acceleration-force acts against the pressure-force. Therefore the case with 700rpm and 200 bar is the one with the highest loads. It determines the stresses inside the parts! 3.6.2. Combustion-loads on the piston and the glass-plate inside the piston The acceleration forces can be determined now. For the total loads also the forces from the combustion-pressure have to be added. Representative pressure-courses for the engine are in existence here at the institution and so the pressure-force can easily be calculated. The pressure course also is shown in the following diagrams. Here the loads out of a maximum pressure of 150 bar and of 200 bar are calculated. The critical-case is determined as discussed in 3.6.1. The important values for the calculations are: Max. piston-velocity 700 rpm 5.89 111/S at –74.4”CAD max acceleration 700 rpm 539 m/s2 at -212.6”CAD max piston-velocity 2000 rpm 16.9 111/S at -74.4”CAD max acceleration 2000 rpm 4398 In/S2 at -212.6”CAD As control of the results, a numerical calculation for the values at 2000 rpm has been done in Matlab. The results are: max piston-velocity (numerically) 2000 rpm I 16.9 m/S at -74.4”CAD I I mix acceleration (numerically) 2000 mm I 4398111/S2 at -212.6”CAD Forces on the glass: max. abs. acceleration-force 700 rpm I (-)457N at –212.6”CAD 118 I max. abs. acceleration-force 2000 mm I (-)3.73KN I at -212.6”CAD I max. pressure-force 150 bar I 125.OKN at 8.2°CAD I 1 max. pressure-force 200 bar I 162.lKN at 3°CAD max. total-force 150bar, 700rpm 124.5KN at 8.2°CAD max. total-force 200bar. 700rpm 161.6KN at 3“CAD Forces on the piston: With PTitaoiu~s 4.5 kg/dm3, the piston mass is calculated to be: rn.isto. = rnTita. + mgla,, 1.02 kg -i- 0.848 kg = 1.87~ This results in acceleration and pressure forces (The area of the total piston is bigger than only of the glass) ofi I rn&. abs. Acceleration-force 700 mm I (-)1 .065KN I at -212.6”CAD I I max. abs. Acceleration-force 2000 mm I (-)8.22KN I at -212.6”CAD I max. pressure-force 150 bar 192.2KN at 8.2°CAD max. pressure-force 200 bar 249.31KN at 3°CAD max. total-force 150bar. 700rpm 191.2KN at 8.2°CAD max. total-force 200bar. 700rpm 248.3KN at 3“CAD 3.6.3. Combustion-courses: piston-way versus CAD 0.34 0.32 0.3 0.28 0.22 0.2 0.18 0.16 -4 I -300 -200 -100 0 100 200 300 400 CAD piston-velocity versus CAD; 700 mm 1 I [ I -300 -200 –100 o 100 200 300 400 CAD piston-acceleration versus CAD 700 rpm 300 1 200 100 -6001 ! 1 ! I , I 1 -400 -300 -200 -1oo 0 100 200 300 D CAD 120 piston-velocity veraus CAU 2000 rpm 20~ 15- 10- ~ 5 - .2 0 0 o - z~ 0c ~ n -5 - , -lo - -15 - t 1 1 I I t -20 I -400 -300 -200 -100 0 100 200 300 400 CAD piston-acceleration versus CAD; 2000 rpm 3000 1000 -3000 -500C 1 1 I 1 # I t 4 ) -300 -200 -100 0 100 200 300 CAD ..._ —.—.._._ numerically results piston-velocity veraus CAD 2000 rpm 20 I # o 1 , I i 15 10 -10 -15 _2rJ I , 1 I ! , 1 I -400 -300 -200 -100 100 200 300 400 C:D numem”cally results piston-acceleration versus CAD: 2000 mm 3000 1000 -3000 –5000 1 I 1 , 1 I I -4 -300 –200 –loo o 100 200 300 400 CAD 122 .—— .— Force on piston out of the acceleration; 700 rpm 600 I 1 1 I 1 n I I I 400 200 0 ~ ~ -200 ~ n 0c g -400 z IL -600 -800 -1000 I I I 1 t -1200 I I I I -400 -300 -200 -1oo 0 100 200 300 400 CAD Force on piston out of the acceleratio~ 2000 rpm 50 1 I I 1 1 I 1 ~ ~ .Q a c o ~ 5 L -50[ -1001 1 I I I I ! 1 I -300 -200 -100 0 100 200 300 400 CAD 105 Force on piston out of the combustion-pressure; 150 bar 1 I 1 # o 1 I , 200 -300 -200 -100 0 100 200 300 ~ CAD 105 Force on piston out of the combustion-preasurq 200 bar \ , 1 -!00 -300 -200 -100 0 100 200 300 400 CAD 124 Both forces on piston; 200 bar, 2000rpm x 105 2.5 [ a .5~ 400 -300 -200 -100 0 100 200 300 400 CAD 105 Both forces on pistow 200 bar, 700rpm 2.5 2 1.5 ~ 0c ~ a 1 0c ~a Ls 0.5 0 -0.5 I I 1 I 1 , 1 t -4 -300 -200 -100 0 100 200 300 400 CAD —.— . .— 105 Total Force on piston; 200 bar, 2000rpm L -/ -300 -200 -100 0 100 200 300 400 CAD 105 Total Force on piston; 200 bar, 700rpm 2. o 1 , -300 -200 –100 o 100 200 300 400 CAD 126 .;— e.———- 3.6.4. Stresses on the glass-plate inside the piston: The glass plate inside the piston is one of the most critical parts of the construction. Therefore the stresses on the plate are determined for different cases. These stresses are used for a numerical calculation with the Finite-Element-Method. The appearing stresses are shown in Drawing 28. 0 Ill below Drawing 28 stresses on the glass-plate 1) first case: standstill, mounting of the engine @stationm load The stresses are determined by the necessary preloading-force for the sealing and the heat-extension. In 3.5Calculations on the sealing:, the preloading force is determined to FSCRw,,~= 91KN. This gives with the bearing area a stationary load of ~abo,~=57 N/mrn2 and ~b~l~~=41 N/mm2 (the bearing area at the under-side is bigger!). This load always appears in standstill! . ~b~l~~= 41 N/mm2 ● ~ab~ve = 57 N/rnm2 ● ~glass = ON/mrn2 2) running the engine at max load, warmed up: The stresses that occur here are the ones for the max working-state. The load on the glass results in agl~s = 19.75 N/mrn2. Running the engine at full load only the minimum sealing-force on the upper sealing and the force X that results of the manufacturing-tolerances remains. This equals a 127 rnin stress Of ~abov~~n=48 N/mm*. The stresses on the lower bearing are then ~b~l~.,max= 107 N/mm2 (combustion pressure!). The stresses from the combustion- pressure are C@=,= 19.75 N/mm*. The lowest combustion loads occur during the intake or exhaust-stroke, when the pressure loads are set to O. The loads on the sealing have been calculated to FefiaU~t= F,~,ti. + X = 78.7 KN. This give the max-value for the upper and the minimum value for the lower to ~&lOW,fin– 35.4 N/rnm2 and ~~bov.,m = 49 N/rnm2. The stresses in this case are ● ~abovqm = 49 N/mrn2 (intake and exhaust) o ~belo~,~n= 35.4 N/mm* (intake and exhaust) ● ~&lOW,_ = 107 N/mm* (combustion) ● Og]ass= 19.75 N/mm* (combustion) ● ~abovqfin= 47 N/mm* (combustion) State ~above.~” ~abov~.m ~belo~,fi~ ~belo~.m %ss 1) standstill 57 N/mm* 57 N/mm* 41 N/mm2 41 N/mm* o 2) running at max 47 N/mm2 49 N/mmz 35.4 107 N/mm2 19.75 load, warm N/mm* N/mmz These values now can be used for a Finite-Element-Calculations of the glass-plate. This is not done in this Master Thesis, but from a PHD-Student at the Institution. 128 — 3.7. Structural integrity assessment for the titanium piston-crown: The critical part at the piston crown is the place where the wall-thickness is a minimum. This is where the Rylon sealing rings are added as shown in place 2) in Drawing 12 in the part Product-Development-Strategy. The load has been examined in the part 3.5 Calculations on the sealing:. It consists of a stationary load that equals the preloading-force on the gasket and a dynamic force during operation. Both cases are calculated here. Way of calculations: The structural integrity assessment is done with the nominal stress approach both for not stationary loads and for stationary loads ([2] C55). It is fulfilled when ~notin~ S cf~~ni~,~owed. For Titanium ~noniti,wow~tension = RPO.2 = 820 N/mm* ([2] El 13) is chosen (TiA16V4 F 89). Different factors for stress-concentration, material and geometrical influences are used. Geometrv: During the construction process the dimensions had to be changed several times. Unfortunately, because of other critical parts, the wall thickness had to be increased at the very end of the construction-process again to 9mm. This equals an inner diameter of 108mm instead of 111 mm like used in the calculations below. Since this also increases the maximum material-value, the calculations are not done again. The structural integrity assessment is done for a weaker construction so it withstands the loads also in a stiffer construction! The piston is simplified as a cylindrical tube with the inner-diameter 11 lrnm and the outer diameter 126mm. So the wall thickness is 7.5mm. The tube is thought as being a cut along the axis so that you get a flat iron as shown in Drawing 29. Drawing 29: Model of the stresses inside the piston Nominal stresses on the piston: The nominal stresses are determined with the geometry in the groove-bottom. The appearing loads from the combustion on the piston are (3.6.2 Combustion-loads on the piston and the glass-plate inside the piston): Max. abs. acceleration-force 2000 rpm F.CC= (-)1 O.68KN Max. total-force 150bar, 700rpm FtOti,150= 192.5KN Max. total-force 200bar. 700mm F,.*.1x-m= 250.OKN At the place, where the structural integrity assessment is done, the prestressing forces for the sealing appear. Most of the combustion-forces are led over the glass into piston-extension. The forces have been calculated in part 3.5. F Scnw.seal =91m Station ary load Ffim.ti. = -22.4 KN Min. value of force, during combustion F.fi.u~t = FTim,- =78 KN Max. value of force, during exhaust and inlet (temp.-extension and preloading) This gives with 0 = $ and Anoti ~[= . d_n “4=m.114.4mmz = 1432mm2the nominal stresses to: Intluence because of grooves, stress concentration-factor q: The groove of the sealing-rings increases the stresses inside the groove-bottom. o Therefore cz~= real (stationary) has to be added to the calculations and so OEI o nominal can be calculated. The stress concentration factor ~k is calculated with the formula 1 ak =1+ ([2] E97) I I . . a a 130 --——— The values for flat-iron ([2] E97), and the variables are: A 0.1 6 Groove-radius lmm B 0.7 t Groove-depth 4mm c 0.13 a Reminding wall-thickness 3mm k 1.00 1 2.00 H+m 1.25 Table 12: Values for q-calculation 0.5 k, t 0.8 1.6 @ “ Drawing 30: Detail, O-rings used as spring-return for the sealing rings The groove-radius can be determined with Drawing 30, examining the o-rings inside the groove. They have a diameter of 1.6 mm so on every side a radius of 0.8 mm can be worked into the material. An additional gap of 0.5 mm increases the radius to a=lmm. T@s gives an ~k-value of @ = 2.5. Since the geometry above doesn’t equal the geometry in this case precisely, an ~k– value from diagrams also is chosen for comparison. .—.. — _.- iaJr@%+---- f--. -.+-. -i.-.... .{. .-..{ 1*I II ,’ I . . . * { ....”.. ..* ..-,--:. ,., . +-&-T..-f .. . ~....:..-j -GH,.:.:,’.,,:D,.~’~‘“ ,’ ; ,+.~ ...... “ ...... I .. . . . <....1 ‘ , . 1 { --i--: --. ”-i.;- i ..[ ; . ~ --- *”--- - ~: : :- . ..-. . . . . 7-3 Figure 29: cx~-values(source: [13]) From this diagram the ~k-vdue is determined to ~k = 1.85. This diagram has a close resemblance to the geometry, but to be on the sure side the higher value of ~k = 2.5 is chosen! w-value I 2.5 132 3.7.1. Structural integri ty assessment for dynamic-loads: Required securitw Required security I VD = 2.5 (1.5 SVD<3) Geometrical size-influence-factor: [ Cdg.. 11 (tension and pressure) Geometrical shape-influence-factor: Assuming linear-elastic material, it is expected that the stresses in the bottom of the gap also rise with the ~k-vdue. Numerous tests have shown that depending on the radius of the gap, much higher strength can be reached than assumed. This is taken into consideration, using the geometry-influence-factors. The factor ~~t (dynamic support-factor) can be determined after [2, E18] and [6, Appendix Festigkeitsberechnung] using X~aP= 2 (stress-lowering-factor) to n.,mr=l+lo~O’33*).Z=Lo5 The geometrical shape-influence-factor Gm then is calculated to: G. = % = 2.38 ni,uf,Z“C~g,Z Ig -4 7.5 cd, = 1-0.25 = 1.05 log(20) Alternate-StrenWh of Titanium: The strength of Titanium for an alternating load is for a probe with a diameter of 7.5mIn Oti= 580 N/mm2. (TV/f) I 580 N/mmZ I This gives the max. alternate-strength to: 133 ——— .— O,. = cd, “C,VO= 609 N / mm2 Surface-influence-factor: The surfaceroughness of the piston is determined by the way of manufacture. The piston will be turned, a roughness easy to manufacture is Ra = 160 pm. co =1 – o.22(lgRa) Ig ~–l 1=0.776 Ra 160 ~m co 0.776 Influence of the surface-treatment: No special treatment of the surface is done here so: ICT.7 11 (only for tension and pressure) I Factor for non-isotropic material: Determine after [6]: I CA I 0.86 (600 s RpO.2 s 900 Total influence-factor: 1 1 G= = G.+—– 111—“— = 3.10 co Cv CA Shape-alternate-strength: K,, =~=177N z / mmz Medium-stress-factor: 134 -- Kw Vz = =0.18 2 “~Po.z – Kw Vz I 0.14 Shape-deviate-strength: The medium-stress value is determined hereto iv 554– 15.6— Gnoti ~,,= N + Cnomin al,min mm- mm2 0= = =19.77, z,m 2 2 mm and the shape deviate-strength to N Kti = KW –~z”om=193—. mm2 (YZJn 19.7 N/mmz KZA 193/mm2 Reached securitv of the part: The reached security is calculated to ‘D=EaGT=$=2=g (values for bending and torsion) an= ~nOfi~d,H -G-= 35.3 N/DIU12 Q@ 35.3 N/mm2 jD 5.47 (reached security) Utilization: 2.5 #’D- —=0.26<1 j“ 9.8 I AD I 26% > — .—. . 3.7.2. Structural integrity assessment for stationary -loads: Reauired securitv (factor of safety): Required security I VF= 1.8 (1.2 SVDS2) Maximal stress in the piston: Ia-r I 160 N/mmz I Stationary support-factor n~t: The stationary support-factor n,w, is calculated with q== 0.01 & = Rp~z = 820 N/mrn2 E = 105.2 KN/rnm2 nsraf‘fi=/~=1.132 G n~wt I 1.132 Size-influence-factors: Only the technological size-influence-factor has to be determined.The geometrical size-influence-factor is only needed when torsion or bending occurs. lg Al c~,,,,a,= 1– 0.05 7“5 = 1.01 lg(20) 136 — — _.. _. Strength of the piston: With ~= 1.2 (for 1.51 <%< 2.00) N K Fzd =C~,,,,~,“yF “RpO,z = 994— mm 2 I KFd I 994 Nknm2 Reached securitv of the part: The reached security is calculated to ‘F’EZR=*=%=%= CY~~=’Ckt=K~=KFt=() (values for bending and torsion) jF I 6.21 (reached security) Utilization: 1.8 A; = —–—=0.29<1‘F _ j~ 6.21 I AF 129% 3.8. Loads on the piston-extension: The piston-extension is a complex part and can not be calculated analytically. Only the weakest part of it is calculated. This is where space is left for the mirror. The calculations are done against buckling and inadmissible deformations. The length of the interesting part is assumed to 1= 270mm. The geometry is shown in section below. For the calculations the outer radius is chosen to 126mm and the distance between the two parts to 103mm. Later in the construction-process the inner distance decreases to 100 mm. Since this increases the maximal values of the extension, the calculations are not done again! . x Calculation of the deformation: It is now required to calculate the area of the material. It is a part of a circle and is 4 times as big as shown in Drawing 31. With the origin in the middle of the circle, the circle-equation can be written to: xz+y2=rz=63z (outer diameter: 126mm = 2.63mrn) or As stated earlier, the total area is 4 times as big as the sketched part and can be calculated to: 63 d= 63 63 A=4” ~ Jl”dx”dy =4” f~~dy=4.$ y.~~+63zarcsin&) y=s I.5 .t=O y=51.5 51.5 Drawing 31: Area of piston extension at weakest cut 138 With the maximum force, calculated before in part 3.6.2 to FM = 250 KN, and a module of elasticity for aluminum of EA =72.2 ISN/mm2, the deformation now can easily be calculated to The appearing stresses in this case are: F= 250KN = 220 N crm=— A = 1135mm2 mm’ s ‘p””’= 400N’mm2 The Rpo.2-value of the used aluminum lies at 400 N/mm2 and so the extension stands the loads easily! Calculation against buckling: The same forces also count for buckling. Here the areal moment of inertia EI has to be known. The calculations can be done here with the same values as in the calculations of the area before. Here again % of the geometry is calculated and then it is multiplied with 4. IX= ~y2dA I, = jX2dA IV = x.y. dA J Ix= ~y2dA = ~y2X dy =4 ‘/ y2~~ldy = 515 I, = 3.587 “10bmm4 1, = 3.588” 10emm4 > Since the geometry has two symmetrical axes, 1. and IYalso give the main-area moments of inertia. For checking the integral IXYis done and it is to equal to O! 36.29 .36.29 IV = jx Y M= jX y2ak=2 ~x. (632 -x’)dx=2 ~(63’x-x3)&= –36.29 –36.29 36.29 =2 63Z 1 —X2 –—X3 =—: (-36.293 - (-36.29)3)= O 2 3 I -36.29 The critical force for buckling can now be calculated after Figure 24. The piston never Z2EI sits stiff inside the cylinder and so case II with F~tif= ~ is assumed. This gives L F~tit= 35it4iV . ~ The extension never buckles! 3.9. Structural integrity assessments on screws: The structural integrity assessments on the screws have been done, using [14]. Some simplifications are taken from [6]. 3.9.1. Screws fiiing the piston-crown: For calculations on these screws, two different states have to be examined. The maximal forces on the screws have been determined in the part 3.5 to F~CmW,~~= 91 KN for the first state while mounting them. The integrity assessment for stationary loads has to be done here. The other state is when running the engine. The maximal force appears with the worst manufacturing tolerances. It is a swelling force with the max-value of 54.2KN (minimum sealing force) + 24.5 KN (X; force because of bad manufacturing tolerances)s 79 KN during intake and exhaust-stroke and the tin-value of O KN during combustion cycle. The calculations are now split in a stationary and a dynamic part * F,CRW,,4 =91KN (stationary) @ O< Fdw~C 579~ (dynamic load) First choose of screws: A first coarse dimensioning is done with [14, Table 7] The chosen screws are: 15 * M5*14 (inside hexagonal head screw) D~ = 5.5 mm (bore diameter) 140 d2 = 4.48 mm , do =4.134rnm lk =8mm dK = 8,5 mm P = 0.8 (gear-ratio of screws) ~min = 0.1 (steel thread on titanium) & =12.9 mm2 As =14.2 mm2 Choose clamp-length: I lk !8mm Choose friction-coefficient: steel on aluminum ~SA I 0.15 (0.08 Choose method of attraction: The chosen method of attraction is to use a torque wrench. This determines % to: q = 1.7 WI 11.7 (torque wrench) Strength-class of the screws: The chosen strength-class of the screws is, according to the high loads, chosen to 12.9 I Strength -class I 12.9 —- Yield-strength of the screws : With the chosen strength-class of 12.9 the yield-strength is determined to RpO.z=1080 N/mm2; Rrnti = 1200 N/mm*; Rmm = 1400 N/mm2; I Rpo.2 I 1080 N/mm2 Determine the minimum necessary clamwforce FK,eti,the axial force on the screw and the dvnamic force: Minimum necessary clamp-force: a) Fixing the part against a transverse force Appearing transverse forces are taken from the aluminum-ring that centers the piston-crown. * no force here b) Force needed against gaping of the parts: A force against gaping is not taken into consideration in these calculations. Two reasons account for this: 1) The parts that are fixed are assumed to be relatively stiff. 2) The loads on the piston when running the engine only decrease the predeforming force to a theoretical minimum of ON. So no gaping can occur here. Q no force here c) Sealing against a medium. The sealing force F,CEW,,~is determined before. It is created in part from the predeforming force, determined in 3.5 Calculations on the sealing:, and in part from the combustion-pressure. When running a combustion-cycle, then the combustion-pressure creates the sealing force and the force on the screws reduces until O at full load. * That determines that even in this state no minimum screw-force is needed. Axial screw-force when mounting: The above mentioned force Fscrew,s~ is the axial force on the Screws. lt is a predeforming force and appears only in a cold engine at standstill. It is a stationary force and so is used for the stationary calculations! This determines the axial force on the screw to: FA,~Ount= FsCRw,,@/15=91 KN/15 = 6.07 KN (for every screw) F&~Ou~t I 6.07 KN (stationary) I Axial screw-force when running the engine: 142 The maximum force when running the engine warm, occurs during intake and exhaust stroke, when the combustion-pressure does not decrease the predeforming force. It is only decreased by the thermal deformation of the piston. It also has been calculated in part 3.5 to F,~,ti, + X =79 KN. The minimum force occurring at high pressure is O. The screw force in this case is dyntic. It is O~ F&dwtiC ~ 79/15 = 5.27 KN. Elasticity of the clamped parts& Elasticity of the screw ~~; for every partof the screw (head threadand shaft) ~s ‘ E& 1 1SK + 11 1’ lM 8s= ls’ +Ej +E lGM + + Es”A~ ~“ ~ ~~-AG~ = Es “AN ‘ Es” Al ‘ Es” AG ‘ EP”A~ IK r 1 FiWre 30: Screw-geometry With the dimensions after Figure 30 1,~= 0.5” d = 2.5mm 11 = 1~= 8mm lG = 0.5” d = 2.5mm 1~ = 0.33” d = 1.65mm Al = AG = 12.69mm2 z“d: AN = — = 19.635mm2 4 8s = 4.95 “10-6 y Elasticity of the plate& For the elasticity of the plate a compensation-area S~ [6] is calculated to: —— with: DA= 10 mm DB = 5.5 ~ lk This gives tip to 6P = = 3.259“10-67 EP “Se,= Force-ratio cD: The force-ratio @is determined to: @ – ‘sA FA The force-introduction factor is relatively hard to determine. For this geometry, it can only be done with the Finite-Element-Method or by measuring forces and deformations of the part. To avoid this expenditure the factor can be precisely enough estimated with [14, 5.2.2.3]. Since the titanium part is relatively stiff, the ideal assumption of a centric force being introduced determines n to be n = 0.5. n I 0.5 This gives @nto: Change in the predeforming force because of setting and temperature influences: Force resulting from setting: “=(3,+fz($,) The factor f= is determined after table 5.4/1 [14] with a surface roughness of 10pm S Rz < 40~ to f== 3pm (thread) + 2.5pm (screw-head)+ 1.5pm (gap) t&3?21 a F== 397N 144 Force resulting from the temperature-influence: E E The fractions ~ and ~ describe the change in the E-modules of the screw and ES Em plate. This expansionary change is neglected here, because the screws are not expected to get warmer than 100”C. The temperature coefficients for steel and aluminum are: ~T,~IJ = @ = 24.5 “lO-GK-l ~T,~&l = as = 12010-6 K-l AT= 90K @ AF;,,, = -1100N Fz 397 N L@,, -11OON Minimum mounting force: The minimum mounting-force is determined to F~,ti = F~,~ti +(1 – @)F~ + F= + ~jf~ . Since ~~,~ decreases the minimum mounting-force and the screws are safety-parts, AF~,l,is set to ~~,~ = O. This secures a sufficient screw-force even when the engine runs cold. Moreover a possible wrong temperature assumption doesn’t influence the minimum force negatively! This gives F~,ti to F~,ti = 4780N FM,ti. I 4780 N I Maximal mounting force: The maximal mounting force now can be determined with%= 1.7 to F~,= = cx~“F~,ti = 1.7. 5850N = 9945Z? FM,- 8120 N 145 ——.— ..— — —.. Determine maximal allowed reduced stress ~~~d~~ and maximal allowed screw- force: The maximal allowed stresses and forces may not exceed 90% of the yield-stress of the material (v=O.9) iv N ~red,Mzu/ = V - RP0.2= 0.9 “1080 —=972— mmz mmz v “RPO.2 F = Mad ‘4 , 9 11280N FWU1 11280N ~red,Mml 972N/rnm2 Maximal load after mounting (&@: FSm =F~- +@”F~- –AFv,~ The thermal forces here are determined after [14, 5.4/10]. 8.+ 6, Lk-(as “ATS– Ctp”ATP) A&h = Fw 1– %rr + 6P E Pm + 6. — c$s~+6p& En Em En Em Making the same assumption as before that the E-modules do not change relevantly over the small temperature distribution of 100”C, these fractions can be set to 1. With this assumption the first part equals zero and AFW~= AF~,fi= –1 lOON. Q Fs- = 8120 N + 0.278 -6070N –(-11OON) =1 O.1KN Fs,~W 10.1KN N a P:,- = Fs,_ 1A = 780— mmz with 146 114G= F~,= ~ —+1.155p~tip = 3380Nmm 2 z“dz Wp =~d~ = 18.496mm3 k, = 0.5 N @ Zm =182.7— mm2 + 3(k, “r- )2 = 796— = 1080N/mm2 # mZ2 < ‘p0”2- 796 N/rmn2 ~red,B Calculation of the alternating load on the screws using [141: When calculating the alternating load on the screws, the extra screw-forces are the important ones. They determine: F~AO– F~All P.= 2A . s The forces F~~Oand F~~Uhave been determined in part 3.5 with F& = n. @. F~. They appear in a warmed up engine. The forces at maximum load and maximum manufacturing tolerances are taken for the calculations. 0< F~d@C <5.27 KN. N * pa=@n 5~~ = 51.2— (15 screws!) “s mm’ The materials maximal value is: PAS” = 0.85 y+45 I = 64-’$ # Pa 51.2 N/mm* ( Calculation of the alternating load on the screws using [61: Since the screw-loads do not equal precisely the ones in [14], alternatively the calculation of alternating load after [6] is done. FaO -Fa. ~=56.8~ = ~ F~,,ti -t @~ “ ‘n’ 4 2 mm’ Now KK can be chosen to 68 N/mm*. This value is a bit higher than before, but still lies about in the same range. .—— _ . . . A last view on the screws is taken by using the whole alternating force of the screws to determine the load and leaving away the factor @..This gives an alternating stress 0 s pa,q ~ 370 —.~[? The mean value 0. is 0. = 0,/2= 185N/mm2. The fatigue alternating stress for the screws is determined in the smith-diagram to ca. 740 N/mm*. This diagram unfortunately does not take care of gaps, which a thread is. So as the fatigue stress decreases, respectively the load has to be increased. Nevertheless this gives a coarse view, as to how the screws are loaded. Calculation of the surface-mwssure: Since the highest loads appear when mounting the engine respectively at standstill, this case has to be calculated here. d,,,. = d~ = 8.5mm No vahes for pGare available for the chosen ah.uninum. Since the pG–values are often near the yield-stress, Rpo.2is chosen. N pG = RPO.Z= 400— mmz F N N P- = A“’- = z50 ~s pG =400— p,mm mm- mm~ pmax I 250 N/mm2 Minimum length of the screws to set in the thread: N The shear-strength for the thread of the titanium is r~ = 0.6. Rnz = 534— mm2 N respectivelyr~ =2. HB = 600— . The lower value of 534 is chosen. After [14, mmz picture 5.5/3] this equals a length of m~ti = 1.17. d = 5.85mm for screws with strength class 12.9. This is less than the chosen length of 6mm. Determine the necessarv torque for mounting the screws: The necessary torque for mounting the screws is determinedto: D A4A =FMW 0.16. P+0.58. dz .fl~ti + ~#Kmin I = 5120Nmm = 5J2Nm with 148 —— ~ (4+%)=7mm Km= 2 lfJKmin= 0.07 (steeI on aluminum) Discussion on the strength of the screws: These calculations have been done using the VD12230 norm [14]. This is the standard for calculations of the screws. Yet this special case cannot be found in there, and some assumptions had to be done. They are mentioned while doing the calculations above. Since the forces of the screws nearly reach the maximum allowed values and neither more nor bigger screws can be used, a means of decreasing these forces has to be thought of! One way is to remove the forces inside the thread, which are created because of friction between screw and plate during attraction. This is done with a device that clamps the piston-crown and the piston extension together and thereby closing the gap that is needed for predeformation. The screws can then be set in easily by hand and no torque-wrench for the many screws is needed. This device also has the advantage that the gap is closed straight and the glass is loaded uniformly. Since the screws are loaded nearly to the maximum such a device is strongly recommended ! Another problem is that the remaining wall-thickness of the titanium is reduced when cutting in the thread. The wall may deform as shown in Drawing 32. This has to be taken into consideration when manufacturing this part. At first the thread has to be cut and then the piston is turned out of a titanium tube! Drawing 32 Cutting the thread may deform the piston-crown ——.— .—— — 3.9.2. Screws fining the glass-module: ! 1 4’ / ,,” /“ / ,/’ . ‘112. ‘ ,/’ ‘ v’ ,/’ (\ / x q ~ X-4 %4\A / / -30 / > / /’ , % , /“ ‘-. “\ ‘\ “\ ‘. ‘\ s’\“ Ic\ ~?:... <“; u r Drawing 33 screws fiing the glass-ring Chosen screws: 4* M22*90 (hexagonal head screw) D~ =24mm (bore diameter) dz = 20.376 mm do = 19.023 mm lk =62mm dK =36mrn P = 2.5 (gear-ratio of screws) P&in = 0.1 (steel thread on titanium) & =281.5mm2 As = 303 mrn2 da =34mrn (outer-diameter of the sleeve) di =24rnm (inner-diameter of the sleeve) lk,p =18rnm lsleeve = 32.28 mm lGl~~ =30rnrn l%ding = 2.28 mm Es~&l = 210000 N/mm2 EG1=, = 70000 N/mmz ESealing = 500 N/mm* 150 .— — The ring around the screws is needed to avoid deformations of the plate, which could destroy the glass. Moreover this ring takes the predeforming force of the screws, so that the load on the glass-ring is decreased. The minimum sealing force has been determined before to FK,ti = 267KN. This force always has to remain. With a minimum sealing force of 267 KN the sealing deforms 43Y0.The module of Elasticity on this state is then about E,~ = 500Nhmn2. FK~fi= FSC~~W,S~/4= 276 KN/4 = 69 KN (for every screw) 69 KN F Kef > Choose clamp-length: Ilk 162rnrn I Choose friction-coefficienti steel on steel: l-k I 0.1 (0.08 < v S 0.20) Choose method of attraction: The chosen method of attractionis to use an ordinarytorque wrench. This determines cx.to: ~= 1.6 1% I 1.6 (torque wrench) Strength-class of the screws: The chosen strength-classof the screws is, according to the high loads, chosen to 12.9 I Strength-class I 12.9 I Yield-strength of the screws : With the chosen strength-class of 12.9 the yield-strength is determined to Rpo.2=1080 N/mm2; Rmfi. = 1200 N/mm2; Rm- = 1400 N/mm2; I Rpo.2 I 1080 N/mm2 I Axial force on the screws when running the engine: 151 ——- ——. _ This force has been determined in part 3.6.2. It equals the maximum force on the piston. FA = FM,~OJ4= 250 KN/4 = 62.5 KN (for every screw) I FA I 62.5 KN (stationary) I Elasticity of the screw 6s; 1~~=0.5” d=llmm 11 =1~ = 61mm 1~ =0.5. d=llmm 1~ = 0.33. d = 7.26mm Al = A~ = 281.5mmz n“d: AN = ~ = 326.1mm2 4 1 1 lG 6, = lSK+ 1’ GM +’ +1 Es “AN Es “Al + EG~ . A~~ = Es TAN E~fAl+ES.AG E,”AN Elasticity of the plate ~~ li The elasticity of the plate also is calculated with the formula dP = — z E./li “ For the elasticity of the plate a compensation-area S.His calculated to: . h,p s ers.p =: d~+— – D; = 670mmz (DA 23 d~) 10 with l~,P= 18mm (plate-thickness) For the sleeve this area is determined to: A sleeve = f(d~ – d: )= 455mmz 152 — The area of the glass and the sealing is: A,,~= = A$,~,i~,= ;(df~ – d;)=: (1872 –127z)=14.8.103mm2 1 1 1 —= + 1 8P l,,p sleeve glass h%? +1 + Es “S,,~,P Es “ASleO~ E~l~~“A~l~~ E~e.li.~“A~..Ii.~ Force-ratio 0: The force-ratio @is determined to: F~A Q=— FA The height for the upper plate is set to infinity and case SV1 is chosen. This determines n to n = 0.7 n I 0.7 This gives O. to: ($, @,, = n = 0.054 c$p+ 6, I m, I 0.058 Change in the predefonning force because of setting and temperature influences: Force resulting of setting: fz ‘==(8, +3,) The factor fz is determined after [14, table 5.4/1] with a snrface roughness of 10pm < Rz < 40~m to fz = 3 ~m (thread) + 2.5 pm (screw-head) + 2”1.5 pm (gap2) @ Fz = 5.07 KN ——.—— -. i Sealing and glass-ring are not taken into consideration here. Force resulting from the temperature-influence: Since the glass has a temperature coefficient that is very low and screws nearly do not warm up, the temperature influence can be neglected here. IF= I 5.07 KN I Minimum mounting force: The minimum mounting-force is determined to F~,ti = F~,,ti +(1 – @)FA + Fz. This gives F~,& to F~,ti = 132.9KN FM,ti I 132.9 KN Maximal mounting force: The maximal mounting force now can be determined with Q = 1.6 to FM ,ma.x =aa “FM,ti = 1.6 “132.9N = 212.6KN I 212.6 KN Determine maximal allowed reduced stress ~r~d~ti and maximal allowed screw- force: The maximal allowed stressesand forces may not exceed 90% of the yield-stress of the material(v=O.90) iv ~r.d,Mzul= v “RpOz = 0.90 “1080— = 9724 mmz mm- ‘MW’=’F=Z===25 Fwul 251.4KN ~r,d.Mad 972 N/mmz 154 Maximal load after mounting &dJ?: FSnux =F~-+@”F~u Fs,m 216.2 KN N @ pZ,u = Fs,u f& = 768z Z’m =ik?Glwp with P MG = F~,u ~ —+1.155pGti = 340.4KNmm 2 z“dz WP = ~d~ = 1802mm3 k, = 0.5 ~r.d,B 784 N/mmz Calculation on the alternating load of the screws: F – F’ti P. = ‘A;A “ s The forces FSAOand Fs~uare calculated with F&O= n” @”F~O. The materials maximal value is: PAS” = 0.85y+45 1=44~ mm- 5.98 N/mm2 P. (s pMv =44N/mmz) I Calculation on the surface-pressure: The highest loads appear when runningthe engine at full load with 200 bar pressure. d,,,. = d. = 34mm AP,ti = ~(d~~ – D; ) = 455mmz pG = looo~ (42CrMo4) mnt2 F N Iv S=455— < pG = looo— / ‘- = APti mmz mm2 pmax 455 N/mm2 Determine the necessary toraue for mounting the screws: The necessary torque for mounting the screws is determined to: D MA =F~- 0.16” P+0.58. dz .~~fi + :#Kmin I= 733mmm = 733Nm with ~ (d, +DB) = 3omm Km = 2 pKti = 0.12 (steelon steel) —— —.. 3.9.3. Plate, taking the counter-force of the ME: It should be examined whether 10mm of thread length inside the plate is enough to withstand the counter-force from the combustion and the force for sealing. The maximum possible length is 10mm. If M12-screws can withstand the load, a thread- length of 10mm would also be sufficient. Later the original screws will be used. That means that the area of the thread increases and so the minimum length even decreases. The calculations on the screws are only done roughly. The same screws as in the original engine will be used. Since this engine will never exceed the loads of the original engine, it can surely be expected that the screws withstand the loads. Necess w sealing force: The necessary sealing force for the graphite sealing has been determined in part 3.5.3 44.4KN before to F~,ti = 44.4KN for all screws together, or to F~,ti = 6 = 7.4KN for each screw (6 screws). 7.4 KN FKerf Chosen screws: 6X M16X135 (hexagonal head bolt) d2 = 14.701 mm do = 13.835 mm P =2 (gear-ratio of screws) lk = 125 mm dK =24rnm ktnin = 0.1 (steel on steel) All = 144.1 mm2 As = 157 mm2 dB = 17.5 mm Choose friction-coefficienti steel on steel: l-k I 0.1 (0.08 S p S 0.20) Choose method of fiiation: The chosen method of fixation is to use an ordinary torque wrench. This determines Q to: ~= 1.6 1% I 1.6 (torque wrench) 157 —— Strength-class of the screws: The chosen strenati-class of the screws is, according to the high loads, chosen to 12.9 IStren gth-class I 12.9 Yield-strength of the screws : With the chosen strength-class of 12.9 the yield-strength is determined to RPO.2=1080 N/mm*; Rmti = 1200 N/mm*; Rmm = 1400 N/mm*; I RPO.Z I 1080 N/rnm2 Axial force on the screws when running the engine: This force has been determinedin part3.6.2. It equals the maximum force on the piston. FA = FH,,0J6= 250 KN/4 = 42 KN (for every screw) I FA 142KN (dynamic) Elasticity of the clanmed parts &: Elasticity of the screw 6s; 1GM 1+1, lG lM 6, = lSK+ “ + Es “AN Es “Al + E~M . AGM = Es TAN Es” Al+ Es. A~ E,-AN 1,~= 0.5. d = 8mm 11 = 125mm 1~ = 10mm lM = 0.33. d = 5.28mm Al = AG = 144.1mmz n.d~ A~=- = 196.74mmz A 158 — Elasticity of the Plate&; For the elasticity of the plate a compensation-area Sen [6] is calculated to: 2 =— s ers,p : d~+~ –D: = 806mm2 @*> 3 d~) Force-ratio @ The force-ratio @is determined to: F 0== a FA The introduction of a centric force is assumed which determines n to be n = 0.5; n I 0.5 This gives @nto: t$p Q,, = n = 0.07 c$p+ 6s I an I 0.07 Change in the predeforming force because of setting and temperature influences: Force resulting of setthux f. ‘z ‘(8, +6,) The factor fz is determined after [14, table 5.4/1] with a surface roughness of 10p.m S Rz s 4oprn to fz = 3 pm (thread) + 2.5 pm (screw-head) + 1.5pm (gap) m @ Fz = 1270N Force resulting from the temperature-influence: The screws are not getting very warm so the temperature influence is neglected here! IFz I 1.27KN 159 Minimum mounting force: The minimum mounting-force is determinedto FM~ = FK,& +(1 – @)F~ + F=. This gives F~,ti to FM,ti = 47.7KN I FM,ti I 47.7 KN 1 Maximal mounting force: The maximal mounting force now can be determined with ~=l.6to F~,m = (xO- FM,ti = 1.6. 47.7KN = 76.4KN Determine maximal allowed reduced stress ~,.d~ti and maximal allowed screw- force: The maximal allowed stresses and forces may not exceed 90% of the yield-stress of the material (v=O.90) N ~r.d,Mzd= v . RP0.2= 0.90.1080 ~=972~ ‘“”’=’&=128”3” FWU1 128.3 KN ~r.d,Mml 972 N/rnmz Maximal load after mounting (hed,B: Fs,- 79.3 KN with P MG = F~,= ~ —+1.155pGti = 89.lKNmm 2 z“dz wP = ~d~ = 693mm3 kr = 0.5 N * Zm =127.8— mm2 =1080N/mm2 W ~r,d,ll= zm +Z(kr “Tm~ =561~ ~ RPo.2ti mm2 pr.d,B 561 N/rn.rn2 Same loads on screws M12: It has now been proven that the original screws withstand the loads as expected. Now it is required is proof that also smaller screws can handle these loads. The maximum loads are determined before so only the maximum acceptable Ioads of the screws have to be determined. l?M,m I 76.4 KN Fs,- 79.3 KN Determine maximal allowed reduced stress ~~~de and maximal allowed screw- force for M12-screws: The maximal allowed stressesand forces may not exceed 95% of the yield-stress of the material(v=O.95) N Pred,Mud = V” RP02= o.95“1080—= 9724 mm2 mm- ‘“”’=4*=82”9” 6*M1’2*135 (inside hexagonal head bold, fine thread) d2 =11.188mm do = 10.106mm P = 1.25 (gear-ratio of screws) &l = 88.1 mm2 FWU1 I 82.9 KN I I 972 N/rnmz ~red.Mzul I So also M12 screws can handle the load. In [14, picture 5.5/3] the minimum length for screws is given as a function of the shear-strength of the material. For a length of 10mm, the shear strength has to be at least 700 N/mmz. N The shear strength is determined to ~~ = 0.65. Rm = 1080—. This determines the mm z chosen steel: 142CrMo4 @ Rm = 1100N/mm~ Since M16-screws will be used later, the area of the thread increases so that the calculations are on the sure side. I Indeed Irecomend to fasten the screws as precise as possible, using the torque I wrench ! Lubricating oil has to be used! Fine thread screws. I Determine the necessary torque for mounting M16 screws: The necessary torque for mounting the screws is determined to: D A4A=F~H 0.16” P+0.58”dz”#~ti+— ; #KminI with D _ (dk+‘B )=20.75mm Km — 2 ~Kmin= 0-12 (steel on steel) P~=2 P~=l.5 dz(fi.e~d) = 15.026 mm * ~A(fi) = 185Nml (standard thread) 162 ——. _.— —-— 4.References [1] Automotive-Engineering 12/93 [2] Burgmann: Sealing Datasheet [2] W. Beitz and K.-H. Grote, Dubbel Taschenbuch fuer den Maschinenbau, Springer-Verlag Berlin Heidelberg New York; jSBN 3- 540-62467-8 [3] Heraeus: Glass Datasheet [4] Institute for Thermodynamics L’IT, RWTH-Aachen, homepage [5] B. Johansson, Study literature LTH Forbr&mingsmotorer AK 2000 [6] Prof Dr.-Ing Birkhofer and Prof. Dr.-Ing. Dr.-Ing.E.h. F.G. Kolhnann, Study literature Technical University of Darmstadt, Maschinenelemente [7] Prof Dr.-Ing Birkhofer, Study literature Technical University of Darmstadt, Product Development [8] Motor Technische Zeitung (MTZ) [9] Shell Lexikon, Folge 52, Editor M’EZ [10] Schnell Gross Hauger, Technical University of Darmstad6 Technische Mechanik 2-Elastostatik, ISBN 3-540-58696-2 [11] SAE-paper 1999-01-3646 [12] SAE paper 940816 [13] R.E. Petersson, Stress Concentration Factors, ISBN 0-471 -68329-9 John Wiley and Sons [14] VDI Richtlinie 2230, systematic calculation of high duty bolted joints; Verein Deutscher Ingenieure, VDI-Gesellschaft Entwicklung Konstruktion Vertrieb, Postfach 101139,40002 Diisseldorf Appendix A: Mounting Manual Mounting Manual: This Manual concisely describes the important facts when mounting the engine. It gives an overview of the necessary torque for the screws and shows the critical parts of the engine. The critical Darts of the engine are: ● The glass ring . The glass plate inside the piston ● screws between the titanium piston crown and the piston extension . thread between the titanium piston crown and the piston extension . thread for the screws of the cylinder head inside the upper ring that supports the glass ring . thread of the screws that hold the glass module together Al :The glass piston: The glass piston is one of the most critical parts of the engine. Very high loads from the combustion in form of pressure and heat appear. Of course the glass plate inside is wished to have a diameter as big as possible. This results on the other hand in a small bearing area of the glass plate. The piston has been designed to get the biggest field of view through the glass and reaching nearly critical values for the stresses onto the glass at the bearing area. Since the biggest loads on the glass bearing appear when mounting the engine it is absolutely important not to exceed the given tolerances and torque-values. Remaining gap between titanium-piston and piston extension: When mounting the glass piston onto the piston extension, a gap between these two parts remains. This gap is necessary for predeformation of the sealing and to compensate the heat and pressure deformations inside the piston. 164 . . ———. lFP,Pistcm * 17771 F+ screw Drawing 34 Forces onto the glass piston Drawing 34 shows the forces on the different parts of the glass piston. When running the engine at high combustion-pressure, the inner part, which is made of the glass plate, the graphite sealing and the aluminum ring, is pressed together. At the same time the outer titanium parts extends because of a high temperature and different temperature-coefficients between the different parts ( FP,fi~im c FP.~la$). In the worst case the glass plate may loosen during the combustion cycle. During intake and exhaust it is pressed against the bearing again. This may cause glass-damage and absolutely has to be avoided! Because of thatall the partshave to be predeformed. The gap-size determines that predeformation and so is very important for a safe run of the engine! Unfortunatelypredeformation rises the load on the glass bearing very much. As a compromise between suitablepredefonnation, sure sealing and reliable operation of the engine had to be found. The gap size has been determined so thatsealing of the piston in every meaningful drive-situationis secured. Since even small predeformation causes high forces, the manufacturingtolerances play an important role here. As mentioned above, the manufacturing tolerances regarding the gap-size are very 1 important. — rnrn of manufacturing tolerances can increase the loads on the glass 100 plate with 10KN! Since not all parts maybe manufactured that precisely another solution has to be found. Most important is the relative-distance between extension and piston. So all the piston parts without the aluminum ring can be manufactured .——. with standard tolerances. The aluminum ring then has to be adjusted to get the right gap-size! The tolerances for the gap have been set to l~~P=1. 10~ nvn. It is desirable to manufacture the parts even more precisely, if possible. Manufacturing thread into the titanium ~iston crown: A problem is that the remaining wall-thickness of the titanium is reduced when cutting in the thread. The wall may deform as shown in Drawing 35. This has to be taken into consideration when manufacturing this part. At first the thread has to be cut and then the piston is turned out of a titanium tube! Drawing 35 cutting the thread may deform the piston-crown First mounting of the glass piston and extension: Calculating on the gap-size requires some assumptions to be made. Because of that uncertainties remain. The necessary force for closing the gap should be91 KN. The critical load for the glass plate is 96 KN. Since the possibility exists that some assumptions have not been correc~ a higher force than calculated maybe necessary. To avoid glass-damage the force has to be measured when closing the gap the first time. The force for closing may not be higher than 96 KN! This can be done by clamping the parts using a hydraulic press. Some other reasons, regarding the screws for fixing the gap also count for that method. If the gap still isn’t closed at 96KN, the 1 aluminum ring has to be turned down some — 100 ‘“ Screws fixing the glass piston to the piston extension: Chosen screws: 15x M5; [email protected] 166 Because of constructional reasons only small screws can be used here. The chosen screws are nearly loaded to the maximum. (Compare Calculation on the screws inside the main-report!) To decrease the apparent loads on the screws, an alternative to the ordinary mounting is shown here. + Ordinarv way of mountinz ● In order to clamp the glass plate straight the screws have to be closed over cross several times. ● Necessary torque on the screws ~~ Before closing the screws secure the gap size is correct! (Compare First Mounting of glass piston and piston extension!) ● The screws have to be fixed with Loctite! Be carefid with Loctite so that the screws are able to open again! ● Mount the screws smoothly and as precisely as possible ● Every time mounting the parts, new screws should be used! (Life length) @ Use of an external device to create the predefonnimz force One trick to decrease the loads on the screws is to decrease the friction inside the thread. Since this fiction plays an important role this increases the acceptable force of the screws considerably! This may be done by closing the gap not with these screws but with an external device such as a hydraulic press or some smaller device that maybe used at the mounted engine. The screws then only have to be set in with a very small torque and only negligible friction inside the thread appears. When removing the device and thereby the outer force, the clamped parts then extend again and the screws are fastened. Another advantage is that the screws do not have to be closed over cross several times like when mounting ordinarily. The outer device closes the gap straight so that the glass plate is loaded uniformly Here of course also Loctite is needed. Since the screws are loaded nearly to the critical values I highly recommend this way of mounting! The same facts as, mentioned before also count when remounting the screws again, where even an extra-force for opening Loctite is needed! .——. —– When using an outer device for mounting, which uses screws itself like a wrench, and you are uncertain about the necessary torque for this device you can contact meat my home address. A2: Upper ring supporting the glass ring: The cylinder head may not be turned down more that 12mm. This determines the thickness of the upper plate to 12mm. The plate is fastened to the cylinder head, using the original screw-size 6 x M16 through the original holes of the screws. Because of the thin plate the thread may not be longer than 10 mm. Calculations have been done, that this length is sufficient when using high quality steel. Nevertheless this is pretty short for an M16 screw. Some instructions have to be followed when mounting the screws so as not to destroy the thread inside the upper ring! Of course the same reasons with the friction inside the thread as mentioned before also count here. ● Use graphite or oil-lubrication inside the thread when mounting the screws! ● Necessary torque on the screws: 16.standard = 185 N standard thread 16.fine = 180N4 fine thread (recommended) . Fastening the screws the torque wrench should be used as precisely as possible and the screws should be mounted smoothly and not with too much force! Creating the hole for these screws it has to be taken care that it is not drilled through the plate. The glass ring has to be supported at every place. To decrease the depth of the hole, it has to be drilled first with an ordinary drill and then the bore-cone has to be removed with a slot-drill. This allows a hole-depth of 11 mm and later a thread- length of 10mm! A3: Screws fixing the glass-module: Chosen screws: 4xM22X64 These screws have to withstand the loads of the combustion pressure and secure necessary sealing force! Since the thread-length is limited to 20mm, lubricating oil, when mounting them has to be used! . Necessary torque on the screws: ~1 . Use lubricating oil! . Mount smoothly not with too much force! 168 -—— –,-— A4: Length of the extended Push-Rods: Manufacturing the extended push rods, “head” and “foot” of the original push rods are used. In the drawing, only the “head” is shown. The drawing for the head only has been done schematically. When manufacturing the extended push rods they should be manufactured 405.5 mm k than the origin~ on~” ~s has to be men into consideration when soldering the “foot” to the extension! A5: Screws between piston extension and original piston: When making the drawing for the piston extension and determining the distances for the screws, which fix the extension to the original piston, the precise dimensions of the original piston unfortunately have not been known. The dimensions of the screws are only guessed! Control the dimensions of the original piston, before drilling the holes for those screws into the piston extension and the original piston! The bearing inside the piston may not be damaged. A6: M5x14 Stopscrews: The M5x14 Stopscrews (part 30) have to be fixed using Loctite, so that they cannot loosen because of vibrations of the engine. They remain while running the engine, so less work has to be done while cleaning the engine. A7: Work on the original piston: Work has to be done on the original piston. Since there are no drawings added, it is shortly described here. In order to get the right compression-ratio, the piston has to be turned down 71nrn from the top. When turningit down it has to be taken care thatthe same shape as in the piston extension is worked into the piston, so thatthe two parts laterfit together. For comparison, you can also examine the extension-plate. —.- -——___ Appendix B: List of Demands List of Demands Date: 30.march 2000 struc- Demands ture Optical access to Scania Diesel engine D12, Department of Combustion Engines, Lunds Tekniska Hogskola, Sweden Term: Value, data, comment, explanation ● De-mounting ● The piston crown should be able to change 1.) easily. Different combustion-processes pro- want to be simulated. duc- ● Cylinder-head ● A standard cylinder-head has to be used tion, and easy change to another one should in principle be possible. ● Geometry ● The general shape has to be retained. That deve- means the intake system (valves) and the lop- cylindrical form of the chamber. ment, ● Original parts: ● There may no work been done at the original engine-block that is here at the con- department struc- ● Minimum distance between ● >1.5 mm tion, valves and piston at TDC (Top Dead Centre) e Geometrical demands because of ● light has to pass the combustion-chamber in the laser-light a rectangular way (it is reflected in the chamber) @ either horizontal or vertical ● numbers of cylinders of the ● 6 cylinders-engine (One to Use for optical engine access) ● length connecting rod: ● 255mrn ● stroke: ● 154rnm ● cylinder-diameter ● 127mrn ● inlet-valve-seat diamete~ ● 39.7mrn ● output-valve-seat-diameter: ● 38rnm o relation between piston area and ● ca. 1.47 (may be needed for heat transfer cylinder area: calculation) ● other important dimensions of ● drawings are at the institution the engine like 0 cylinder head @ distance to next piston + determine building space for optical access @ holes for screws of cylinder- head 170 -- —.— —.. B place of disposal B No borders for dimensions, but it should be easy to handle on the workbench Fix B Connection dimensions ~ Other devices are not going to be connected directly to the object @ lasers are adjusted later @ no demands Fix E water connections to the device ● standard connections for water cooling of the cylinder-head @ one water in one water out (wish) Fix m oil connections to the device ● standard connection for oil transport to the cylinder-head @ use the possibility to take the cam follower as a oil-return (wish) Wis s further optical connections ● possibility to add endoscope later h Fix ● pressure-measuring connection ● pressure-measuring-device has to be added inside the cylinder-head later @ no task here since use of standard cvlinder-head Aim ● Costs of the project Not more than 500.000 SEK Fix ● Deadlines ● Detailed time-table has been made Fix ● Patents s Patent office has been checked while doing the literature research with few success; . Engine will not be sold on the market @ no patents to take care of Wis . special wishes for the . Optical access from the sides has to be h construction because of a better fixed to the cylinder-head handling (installation) @ It should be possible to install this whole building group (Cylinder-head with side-windows) as one part on the working bench before mounting 2.) Fix ● special parts and resources to use . Quarz glass window work . Engine-Block and Cylinder-Head are here prepa- at the department ration, manu- Wis . Are there parts to be bought from . Quarz-Glass at “Kemicentrum LTH” factu- h somewhere else? Q is made after the drawings are ready, not ring my task of parts 171 _. —-— Fix ● special materials . Quarz-Glass: light wave-length from UV to Near IR 240-1000nm 3.) Fix . Installation of the device . Fast and easy to install and to remove the instal- whole device lation Fix . De-mounting ● Need of possibility to uninstall everything again Fix @ No welding Wis ● Cleaning of the device ● As easy cleaning of the windows as h possible 4.) Fix ● Max. weight of device . no special limitations of the whole-weight trans- Aim . weight of moving parts . mass of the moving parts (adjusted to the port, piston) has to be as low as possible @ the weight of the whole moving piston stora- with added parts has to be about the ge same as the other piston in order to avoid oscillations 5.) < ● prestige wishes, design . use non-corrosive-metals at the outer parts distri- -. % in order to avoid rust bu- ~ a little high-tech-look tion, sale 6.) Fix ● main-function of the device get an optical access to the combustion chamber of the engine that is as big as tion possible. Hereby the combustion process should be and disturbed as little as possible. This means that the shape of the combustion chamber stand- and the valves can not be changed! still Fiel e field of view: View of the entire bore is demanded, at d around TDC (Top Dead Centre) :-50 to +50 CAD (Crank Angle Degrees) Fix . measuring method a laser-based measuring-method will be used. The optical demands such as lenses or signal processing are not included in the task Fix ● demands on the way of the laser the laser light has to be led inside the combustion chamber and be caught again in an angle of 90 degrees 172 —. . . Fix ● cooling of piston B piston has to be air-cooled from the inner- side Fix . cooling of cylinder B cylinder: water-cooled Fix . lubrication E glass-piston has to go without oil- lubrication in order to have abetter view @ use special sealing rings made of Rylon Fiel . kinematics, revolutions ~ Min. speed: 1500rpm max speed d 2000rpm; Fix . piston mean-velocity D Sp = 2LN = lomls (L= stroke, N = speed [rev/s]) Fix ● stroke ● same as original engine Aim ● max temperature o has to stand a temperature up till 300C Fiel . min. demand on device according @ Has to stand at least lOObar! d to pressure in the combustion + Wish to drive with 150 bar chambec @ Doesn’t have to exceed 200 bar 100bar Wis . Rust-protection ● outer surface should have a nice view h Fix ● parts that have contact with cooling water of high-quality-steal ● Use of the device . No problem, experts are available @ Which knowledge is expected of the personal? Fix . wishes about managing the . optical control is demanded, so that You device can see, if mixture is burning or not @ how over-viewable ● security . A protective glass-shield so that in case of @ Protection of personal? window-failure no body gets hurt (normally engine runs in a separate room behind a glass window wish) 7.) . Maintenance ● long maintenance intervals are of course main- wished, but are not the main task. No tenan- optimisation. Wish that the windows are ce easy to change in order to clean them and @ 3.) installation repair Fix ● accessibility . possibility to change the piston crown @ 1.) prod., dev., contr. 8.) . just one unit is produced recyc- @ single piece ling @ no fimther demands on recycling 173 —- .—— _ .— Appendix C: Matlab Programs Cl: Acceleration % the program can be used for the calculation on the forces on the glass and on the piston. 3 values have to % be changed!! ! ! %1.) mglass = . . . . . [program head] ; line 24/25 %2.) F15CIp = . . . . . [calculation on the force out of the pressure, 150 bar] line 166/167 %3.) F200p = . . . . . [calculation on the force out of the pressure, 200 bar line 194/195 clear all; close all; dglass = 102e-3; %gasket diameter dpiston = 126.5E-3; hglass = 45E-3; Aglass . pi*(dglass)A2/4; Apiston = 0.25*pi*dpistonA2; Vglass =3.856e-4; %[mA31 glass-volume (5 mm added because of gasket diameter before rhoglass = 22oo; %[kg/mA3] density glass %mglass = Vglass*rhoglass; %values for glass mglass = Vglass*rhoglass + l.o2; %values for piston (density titianium =1.4 kg/m r = (154/2)*lE-3; % radius of crank-shaft is 1/2 stroke 1 =255E-3; % lenght of connecting rod nl = 700/60; % [rps] n2 = 2000/60; % [rpsl [phi] = linspace (-2*pi,2*pi,3601) ‘; dphil = 2*pi*nl; dphi2 = 2*pi*n2; s(i) = r*cos(phi(i)) + (1*2 - rA2*(sin (phi(i) ))A2) ‘0.5; %%%%%%%%%%%%%%%%%velocity %V= ds/dt; 174 vi(i) = -r*dphil* (sin(phi (i)) + 0.5* (r*sin(2*phi (i)) )/(lA2- rA2*(sin(phi (i)) )A2)A0 .5) ; v2 (i) = -r*dphi2* (sin(phi (i)) + 0.5* (r*si.n (2*phi(i)))/(l A2- rA2*(sin(phi (i.) ) )A2)A0 .5) ; %%%%+000000000000>$~~>~$~>> >acceleration %a = dv/dt al(i) = -r*(dphil) A2*(cos (phi(i)) + r*((l A2 - rA2* (sin (phi (i)))A2)*cos(2*phi (i))+ 0.25* rA2*(sin(2*phi (i)) )A2)/((lA2-rA2* (sin(phi (i)) )*2)*1.5) ) ; a2 (i) = -r*(dphi2) A2*(cos (phi(i)) + r*((lA2 - rA2*(sin (phi (i)) )A2)*cos (2*phi (i))+ 0.25* rA2*(sin(2*phi (i)))A2)/((lA2-rA2* (sin(phi (i)) )A2)Al.5)) ; end; [mist, highvll =max (vi) ; highvl = phi(highvl) *180/pi; [mist, highall =max (al) ; highal = phi(highal) *180/pi; [mist, highv2] =max (v2) ; highv2 = phi(highv2)*180/pi; [mist,higha21=max (a2); higha2 = phi(higha2)*180/pi; 9>999020s.000060aoOOOO0099>%%99>>.o0060a%~>o numerical test of results % at 2000 rpm dt = 1/60000 s %=> dt = 1/60000; vnum2000 = gradient(s)/dt; anum2000 = gradient(vnum2000)/dt; .0000000000...... 0000000000...... ,.,.-<,..0000000000 ...-..~~~~0000000000 ~~~~~~~.0000.0 o 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 00 e o 0 00 0 e o 0 0 0 0 0 0 0 disp(’max. velocity 700 rpm: ‘) max (v1 ) disp(’at CAD:’) highvl disp(’max. absolute acceleration 700 rpm: ‘) max(abs(al) ) disp(’at CAD:’) highal disp(’max. velocity 2000 rpm:’) max (v2 ) disp(’at CAD:’) highv2 disp(’max. absolute acceleration 2000 rpm: ‘) max(abs(a2) ) disp(’at CAD:’) higha2 disp(’numerically max. velocity 2000 rpm:’) max(vnum2000 ) disp(’numerically max. acceleration 2000 rpm:’) max(abs(anum2000)) figure(l); plot(phi*180/pi,s); title(’piston-way versus CAD’); ylabel (’piston-way [m] ‘); xlabel(’CADl); print -deps2 ‘d:\juergen\pictures\plots\piston_way.eps’ -tiff figure(2); plot(phi*180/pi,vl); title (’piston-velocity versus CAD; 700 rpm’) ; ylabel (’piston-velocity [m/s] ‘); xlabel(’CAD’); print -deps2 ‘d:\juergen\pictures\plots\ve1700rpm.eps ‘ -tiff figure(3); plot(phi*180/pi,al); title (’piston-acceleration versus CAD; 700 z-pm’); ylabel (’piston-acceleration [m/sA2] ‘); xlabel(’~’); print –deps2 ‘d:\juergen\pictures\plots\acc700rpm.eps I -tiff figure(4); plot(phi*180/pi,v2); title (’piston-velocity versus CAD; 2000 rpm’); ylabel (’piston-velocity [m/s] l); xlabel(’CADl); print -deps2 ‘d:\juergen\pictures\plots\ve12000rpm.eps ‘ -tiff figure(5); plot(phi*180/pi,a2); title (’piston-acceleration versus CAD; 2000 rpm’) ; ylabel (’piston-acceleration [m/sA2] I); xlabel(’~’); print -deps2 ‘d:\juergen\pictures\plots\acc2000rpm.eps ‘ -tiff figure(6); plot(phi*180/pi,vnum2000) ; title (’numerically results: piston-velocity versus CAD; 2000 rpm’); ylabel (’piston-velocity numerically [m/s] ‘) ; xlabel(’CAD’); print -deps2 ‘d:\juergen\pictures\plots\ve12000rpm_num.eps I -tiff figure(7); plot(phi*180/pi,anum2000) ; title (’numerrically results: piston-acceleration versus CAD; 2000 rpm’) ; ylabel (’piston-acceleration [m/sA2] ‘); xlabel(’~’); print -deps2 ‘d:\juergen\pictures\plots\acc2000rpm — num.eps! -tiff %%%%%%%%%%%%%%%%%% acceleration-force on the glass-part: F = m*a F700a = mglass*al; F2000a = mglass*a2; [mist,highF700al =max(F700a); highF700a = phi(highF700a)*180/pi; [mist,highF2000al =max(F2000a); highF2000a = phi(highF2000a)*180/pi; disp(’max. absolute acceleration force; rpm 700:’) max(abs(F700a) ) disp(’at CAD:’) highF700a disp(’max. absolute acceleration force; rpm 2000:’) max(abs(F2000a)) di.sp(’at CAD:’) highF2000a figure(8); plot(phi*180/pi,F700a); title(’Force on piston out of the acceleration; 700 rpm’) ; ylabel(’Force on piston [N] ‘); xlabel(’~’); print -deps2 ‘d:\juergen\pictures\plots\force_acc700rpm.eps I -tiff figure(9); plot(phi*180/pi,F2000a) ; title(’Force on piston out of the acceleration; 2000 rpm’) ; ylabel(’Force on piston [N] ‘) ; xlabel(’CAD’); print -deps2 ‘d:\juergen\pictures\plots\force_acc2000rpm. eps’ -tiff 176 e o . c! .3 0 0 0 0 0 0 0 0 0 0 >>>>% >%>9>B% %9>//,,..o 00 0 0 0 0/////0 e o 0 /H///0 0 0 0 0 e/<~< 0000.JOOOOOOOOO o.>>>>q>q>. sg>>>%gg>>>>>~>~>>calculation On the force <<66<~<6<~6<~<6< 00000000600000 0000000000000 out of the pressuress<~sss 00 0 a 0 0 0 0 0 0 a 0 0 ~>~>s%>>> 45<6666<666550 0 0 0 0 0 0 0 0 150 bar load Pre153bar.hrc; p150 = Pre153bar( :,3) ’*lE5; %F150P = p150*Aglass; %values for glass F150p = p150*Apiston; %values for piston [mist,highF150pl =max(Pre153bar (:,3)); highF150p = Pre153bar (highF150p, 1) ; disp(’max. pressure force; 150 bar:’) max(F150p) disp(’at CAD:’) highF150p figure; plot (Pre153bar (:,l),F150p) ; title(’Force on piston out of the combustion-pressure; 150 bar’) ; ylabel(’Force on piston [N] ‘); xlabel(’CAD’); print -deps2 'd:\juergen\pictures\plots\forceflressurel5Obar.eps’ - tiff ..0 00 0 0 0 0 0 0 0 0 0 0 0 0 >>>>> 6%<556656s55565550 0 0 0 0 200 bar load Pre198barshort .hrc; % datafile exists of only 721 instead of 3601 values so it has to be adjusted here! -.:=0; allways 5 values in after each other are taken the same values as the one before , for i = 1:720 % steps are so created, but the information is still precise enough! for counter = 1:5 j = j+l; Pre198bar(j, :) = Pre198barshort (i,:); end; end: Pre198bar(3601, :) = Pre198barshort (721,:); % finish of adjusting p200 = Pre198bar (:,3) ’*lE5; %F200p = p200*Aglass; % values for glass F200p = p200*Apiston; % values for piston [mist,highF200pl =max(Pre198bar (:,3)); highF200p = Pre198bar (highF200p,l) ; disp(’max. pressure force; 200 bar:’) max(F200p) disp(’at CAD:’) highF200p figure(n); plot(phi*180/pi,F200p); title(’Force on piston out of the combustion-pressure; 200 bar’) ; ylabel(’Force on piston [N] ‘); xlabel(’C!AD’); print -deps2 ld:\juergen\pictures\plots\forceJressure2OObar.eps' - tiff 177 ---- .,.,,,,,..-.,... >,..,,.. ..mq...-.-..<.;/.,, ,., [mist,highFpiston150] =max (Fpiston150); highFpiston150 = Pre153bar (highFpiston150 ,1) ; disp(’max. total-force; 150 bar, 2000rpm:’) max(Fpiston150) disp(’at CAD:’) highFpiston150 figure; hold on; plot(phi*180/pi,F150p) ; plot(phi*180/pi,F2000a) ; title(’Both forces on piston; 150 bar, 2000rpm’) ; ylabel(’Forces on piston [N] ‘); xlabel(’CAD’); hold off; print -deps2 'd:\juergen\pictures\plots\force_bothl5Obar2OOO~m .eps’ -tiff figure(13) ; plot(phi*180/pi,Fpiston150) ; title(’Total Force on piston; 150 bar, 2000rpm’) ; ylabel(’Force on piston [N] I); xlabel(’CAD’) ; print -deps2 'd:\juergen\pictures\plots\force_totall5Obar2OOOrpm. eps’ -tiff %%%%%%%%%%%%%%% 200 bar, 2000 rpm Fpiston200 = F200p + F2000a; [mist,highFpiston2001 =max(Fpiston200) ; highFpiston200 = Pre198bar(highFpiston200 ,1) ; disp(’max. total-force; 200bar, 2000rpm :’) max(Fpistzon200) disp(’at CAD:’) highFpiston200 figure; hold on; plot(phi*180/pi,F200p) ; plot(phi*180/pi,F2000a) ; title(’Both forces on piston; 200 bar, 2000rpm’) ; ylabel(’Forces on piston [N] ‘); xlabel(’CAD’); hold off; print -deps2 ‘d:\juergen\pictures\plots\force_both2 00bar2000rpm. eps’ -tiff figure; plot(phi*180/pi, Fpiston200); title(’Total Force on piston; 200 bar, 2000rpm’) ; ylabel(’Force on piston [N] ‘); xlabel(’CAD’); 178 print -deps2 !d:\juergen\pictures\plots\force_total2OObar2OOOqm.eps’ -tiff %%%%%%%%%%%%%%%150 bar, 700 rpm Fpiston150_700 = F150P + F700a; [mist,highFpiston150_700]=max(FPiston150_700) ; highFpiston150_700 = Pre153bar (highFpiston150_700 ,1) ; disp(’max. total force; 150 bar, 700rpm:’) max(Fpiston150_700 ) disp(’at W:’) highFpiston150_700 figure; hold on; plot(phi*180/pi,F150p) ; plot(phi*180/pi,F700a) ; title(’Both forces on piston; 150 bar, 700rpm’) ; ylabel(’Forces on piston [N] ‘); xlabel(’CAD’); hold off; print -deps2 !d:\juergen\pictures\plots\force — both150bar700rpm.eps ’ - tiff figure; plot(phi*180/pi,Fpiston150_700) ; title(’Total Force on piston; 150 bar, 700rpm’); ylabel(’Force on piston [N] ‘); xlabel(’CAD’) ; print -deps2 'd:\juergen\pictures\plots\force_totall5Obar7OOqm. eps’ -tiff %%%%%%%%%.%%%%%%200 bar, 700 rpm Fpiston200_700 = F200p + F700a; [mist,highFpiston200_700]=max(Fpiston200_700) ; highFpiston200_700 = Pre198bar (highFpiston200_700 ,1) ; disp(’max. total force; 200 bar, 700rpm:’) max(Fpiston200_700 ) disp(’at CAD:’) highFpiston200_700 figure; hold on; plot(phi*180/pi,F200p) ; plot(phi*180/pi,F700a) ; title(’Both forces on piston; 200 bar, 700rpm’); ylabel(’Forces on piston [N] ‘); xlabel (’CAD’); hold off; print -deps2 'd:\juergen\pictures\plots\force_both2OObar7OO~m.eps’ - tiff figure; plot(phi*180/pi, Fpiston200_700) ; title(’Total Force on piston; 200 bar, 700rpm’) ; ylabel(’Force on piston [N] ‘); xlabel(’CAD’); print -deps2 !d:\juergen\pictures\plots\force — tota1200bar700rpm.eps ’ -tiff . ..— .—— —— close all; clear; pack; load SCANIAinletv.dat; load Scaniaexhaust.dat ; Xin = SCANIAinletv(:,2)/1000; Xout = Scaniaexhaust(:,2)/1000; XinMax = max(Xin) XoutMax = max(Xout) Vinkelin = SCANIAinletv(:,l) ; Vinkelout = Scaniaexhaust (:,l); % calculation on dt % revolutions = 2000rpm = 100/3 rps % dt =’2 grad % 1 revolution =’ 3/100 [s1 > => 2 grad = dt = 3/100 [s1 *1/18CI = 1/6000 & = 1/6000; Vin = gradient(Xin)/dt; Vout = gradient(Xout)/dt; VinMax = max(Vin) %[m/sl VoutMax = max(Vout) %[I_fI/S] %%%%%~%~~%%%%%%acce~eration Ain = gradient(Vin)/dt; %[m/s’2] Aout = gradient(Vout)/dt; %[m/sA2] AinMax = max(Ain) AoutMax = max(Aout) hold on; figure(1) ; plot(Vinkelin, Xin) ; grid; title(’inlet valve ------way’) ; ylabel(’way(m) ‘); xlabel(’angle’) ; hold off; print -deps2 ‘d:\juergen\pictures\plots\inlet_valve — way.eps’ -tiff hold on; figure(2) ; plot(Vinkelout ,Xout) ; grid; title(’exhaust valve ------way’) ; ylabel(’way(m) ‘); 180 ..— xlabel (’angle’) ; hold off; print -deps2 Id:\juergen\pictures\plots\exhaust_valve_way.eps’ -tiff hold on; figure(3) ; plot(Vinkelout ,Vout) ; grid; title(’exhaust valve ------velocity’); ylabel(’velocity(m/s) ‘); xlabel(’angle’) ; hold off; print -deps2 'd:\juergen\pictures\plots\efiaust_valve_velocity.eps’ - tiff hold on; figure(4) ; plot(Vinkelout ,Aout) ; grid; title(’exhaust valve ------acceleration’) ; ylabel(’acceleration(m/sA2) ‘); xlabel(’angle’) ; hold off; print -deps2 'd:\juergen\pictures\plots\efiaust_valve_acceleration.eps’ -tiff hold on; figure(5) ; plot (Vinkelin,Xin) ; grid; title(’inlet valve ------way’) ; ylabel(’way(m) ‘); xlabel(’angle’) ; hold off; print -deps2 td:\juergen\pictures\plots\inlet_valve_way.eps’ -tiff % program push_rods Juergen Fuchs Lth, Lund 21.06.2000 Hauptstr.56 D-56829 Pommern Germany tel. : 0049/(0)2672/1771 e-mail: fuchs75@hotmail .com (currently) %This program calculates the deformation of the push-rods in the internal combusiton engine D12 from SCANIA %The calculations are done both with the original push-rods and with the extended one for the optical engine. %The diameter of the extended push rods then can be determined to have the same deformation. 181 .——— — %The program takes into consideration % the loads from the combustion-pressure ~o the lodas from the acceleration of valves >0 the loads from the acceleration of the valve-rockers % the dynamic force of the push-rods because of the acceleration ~o force because of the spring return of the valve-springs close all; clear; pack; %ingoing parameters: lorg = 0.230; %[m]lenght of original push-rods Aorg = 7.854e-5; %[mA2] area of original push-rods lext = 0.6355; %lenght of extended push-rods rho = 7800; %[kg/mA3] density steel morg = Aorg*lorg*rho; valvediameter = 38e-3; %[ml diameter exhaust-valve-seat Avalve . pi/4*(valvediameter) ‘2; %[mA2] area exhaust valve E = 210000e6; %[N/mmA2] E-module for steel Ppre_exhaust = 665 %[N] predeforming force of the springs , exhaust valve fexhaust = 120700 %[N/m] spring constant, exhaust-valve mvalve = 0.312 %[kg] mass of exhaust valve, retainer and dynamic mass of spring %dimensions valve- rockers 11 = 30e-3; % [m] 12 = 46e-3; % [m] r = 28e-3; % [m] J= 6.135e-4; %[kg*mA2] moment of interia of the valve-rocker o e o 0 0 0 0 0 0 0 e . 0 0 00 0 0 0 0 0 0 00 0 0 0 0 0 0 0 0000000000 0000000000 00000.0000 0000 ...... ,-.. %%%%,//////////.///////././././///.<5< 0000000000 0.00000000 0000000000 0 0 00 0 000 0 0 00 00 0 0000 0 0 0 0 0 0 00 000 0 0 0000000000 0000000000 0000000000 0 ...... 0000000000 0.00000000 0000000000 0000000%%%~~%%>%%%s>%%s0000000 00000-0000 0000000000 0000000000 0 ...... 0000000000 0000000000 0000000000 0 forces on the valves because of combustion-pressure and acceleration 0.00000000 0000000000 0000000000 0 ~%%~eo. oo...oOOOoOo . . ..o 00..00.0.0.000 ...... ,,,.,, ...... /...-,-...... 0000000000 0000000000 0000000000 0 0000000000 0000000000 0000000000 00000 Ooooeeoooo 0000000000 0000000000 0000000000 0000000 ...... /... .,....~~~~ ~~~~~~~ 0000000000 0000000000 0000000000 000000 182 Xin = SCANIAinletv(:,2)/1000; Xout = Scaniaexhaust(:,2)/1000; XinMax . max(Xin) ; XoutMax = max(Xout) ; Vinkelin = SCANIAinletv(:, l); Vinkelout = Scaniaexhaust(:, l) ; % calculation on dt % revolutions = 2000rpm = 100/3 rps % dt .“2 grad % 1 revolution =’ 3/100 [s1 > => 2 grad = dt = 3/100 [s1 *1/18CI = l/61300 & = 1/6000; %%%%%%%%%%%%%%%%%velocity Vi.n = gradi,ent(Xi.n)/dt; Vout = gradient(Xout)/dt; VinMax = max(Vin); %[m/s] VoutMax . max(Vout) ; %[m/sl Ain = gradient(Vin)/dt; %[m/sA21 Aout = gradient(Vout)/dt; %[m/sA21 AinMax = max(Ain); AoutMax . max(Aout) ; %%%%%%%%%%%%%%%%%%%%%acceleration.force on the exhaust valves: Fvalve —accl = mvalve*Aout; %%%%%%%%%%%%%%%%%%spring force on the exhaus-valves: Fspring = Fpre_exhaust+fexhaust*Xout; %%%%%%%%%%%%%%%%%total acceleration-force on the exhaust valves Fvalve_acc = Fvalve_accl + Fspring; 6565s666<66<5<56565s6565%<66...... 0. oeeoo...e. 00000..0 determine dynamic forces on the valves because of the combustion-pressure %%%%%%%% Fp = p*Avalve p200 = Pre198barshort (:,3) ’*lE5; Fp200 = p200*Avalve; ~64G<66<56<6<<5< % One more point to take into consideration here is that the combustion-pressure is led % into the valve-seat, when the valve is not opened! So the pressure may only be taken into % consideration in that part of the combustion, This gives the reasons to erase the first % part of the pressure-course and start at 360 CAD. The if loop-secures that the force on > 0 valves are only calculated when the valve has really opened and is not let into the valve-seat! for i=410:607 Fpvalve(i-409) = Fp200(i); end; Fpvalve = Fpvalve’ ; %%%%%%%%%%% take into consideration that the engine has 2 exhaust valves: Fpvalve =2*Fpvalve; Fvalve_acc=2 *Fvalve_acc; for i = 1:198 if Xout(i) > 0 Fvalve(i)= (Fvalve_acc (i)+Fpvalve(i.) ); else; Fvalve(i) =Fvalve_acc(i) ; end; end; Fvalve =Fvalve’ ; figure(l); plot (Vinkelout, Fvalve_acc) ; title (’Acceleration and spring force on the exhaust valves: 2000rpm, 200 bar’); ylabel(’Force on valve [N] ‘); xlabel (’CAD after T.DC’); print -deps2 ‘d:\juergen\pictures\plots\acc_force_valve’ -tiff figure(2); plot(Vinkelout,Fpvalve) ; title(’ Pressure-force on the exhaust valves (Fpvalve) : 2000rpm, 200 bar’); ylabel(’Force on valve [N] ‘) ; xlabel (’CAD after TDC’) ; print -deps2 ‘d:\juergen\pictures\plots\pressure_force_valve’ -tiff figure(3); plot (Vinkelout, Fvalve); title(’ Force on the exhaust valves (Fvalve,press+acc) : 2000rpm, 200 bar’); ylabel(’Force on valve [N] ‘); xlabel(’C!AD after TDC’) ; print -deps2 ‘d:\juergen\pictures\plots\force_valve’ -tiff 184 — %% % M=J*d2phi/dtA2 =J*phi2; %%%%% way, velocity and acceleration of the push-rods: Xp = (ll+r)/(12+r) *Xout; vp = (ll+r)/(12+r)*Vout; ap . (ll+r)/(12+r) *Aout; %%%%%%%%%%%%%%%%%% determine d2phi/dtA2 =phi2 for i.=1:198 phi2(i) =(ap(i)*((ll+r) ‘2-(xp(i))A2)+ ((vij(i.))A2)*ti(i))/( ((ll+r)A2--(xp i)A2))A 3/2) : end; phi2=phi2’; 9%9>>9>9> 000000000 Minertia = J*phi2; % N2 is smaller than the forces on the valves, because the of the geometrical relations of the valve-rockers % the forces decreases, but the way x2 increases! deforOrg = (morg*lorg*ap)/(2*E*Aorg) + (N2*lorg)/(E*Aorg); figure(5); plot(Vinkelout,deforOrg*lOOO) ; title (’Deformation of the original push-rods: 2000rpm, 200 bar’) ; ylabel (’Deformation [mm] ‘); xlabel(!CAD after TDC’) ; print -deps2 ‘d:\juergen\pictures\plots\deformation_org’ -tiff <<556<5<5656<66<555s556 [ma.xlleforg,vinkelmax] =max(deforOrg) ; disp(’max.deformation original push-rods at 200 bar, 2000rpm [mm] :’) max(deforOrg) *1OOO disp(’at CAD after TDC:’) 2*vinkelmax+50 %(every second CAD one measured point ! ) 99>9>9s>>...... ss...s009>90>>>>>99>99...... deformation of the extended push- rodsss~s<<< for i = 1:198 Aext(i) = (N2(i)*lext)/(defext*E - 0.5*rho*lextA2*ap (i)); end; Aext = Aext’ ; [maxAext,vinkelmaxA] =max Aext) ; disp(’Area necessary for the extended push-rods at max load [mA21 :’) max(Aext) disp(’at CAD after TDC:’) 2.vinkelmaxA+5 O figure(6); plot (Vinkelout,Aext); title(’Necessary area of the extended push rods’) ; ylabel (’Area [mA2] ‘); xlabel (’CAD after TDC’) ; ~6<<< for i = 1:198 Fext(i) = mext*ap(i)/2 + N2(i); end; figure(7); plot(Vinkelout ,Fext) ; title(’Force inside the extended push-rods’) ; ylabel (’Force [N] ‘); xlabel (’CAD after TDC’) ; print -deps2 ‘d:\juergen\pictures\plots\force_ext ‘ -tiff .-0000000.0...... 000000.000...... -0000000-...... 0000000000...... 0000000000...... 0000. . . o 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 . 0 0 0 0 0 0 0 0 0 0 0 e o 0 0 0 0 0 0 0 0 0 00 0 e 00 0 0 0 0 0 0 0 0 [maxFext,vinkelmaxF] =max(Fext); disp(’max force inside the extended push-rods [N] :’) max(Fext) disp(’at CAD after TDC:’) 2.vinkelmaxF+5 O %(every second CAD one measured point ! 186 Appendix D: Drawings