Automation of VW transmission 02k-DNZ
Contents
1. ER Tijd s Schakelactie 3 gt 4 abe Bak Rem Koppeling EE inept mas a erder Koppeling 0 02 0 4 06 Schakelactie 5 gt 6 Bak Rem Koppeling 44 Terugschakelen Schakelactie 2 1 Bake Rem Koppeling D 0 42 04 Schakelactie 3 gt 2 0 4 Schakelactie 3 8 Rem F Koppeling Sch kel ctie 4 gt 3 Bak Rem Koppeling B k Rem Koppeling H 02 0 4 06 6 4 12 dd 16 LE 5 Schakelactie 4 gt 1 Bak Rem Koppeling 45 Schakelactie 5 gt 4 armen pn vennen nnen 1 4 E gt EI Koppeling Schakelactie 5 2 4 222222 an 15 Schakelactie 6 gt ES i i 1 o DE D o So Loi eo 5 E T t L M Pi a e per t e HN 3 o 1o 5 5 KA a 9 EI Kc 8 Ser 3 o s os a 6 5 3 8 N N 5 ei d o We o ro Koppeling 0 1 46 Appendix Parameter Geschatte Waarde Gemeten Waarde Symbool Hoe te verkrijgen 10 mm s Meten ie pp p2. M i actuation t tmp 0 t omega omega tmp k 17 k tmp 0 while k tmp phi if omega tmp lt omega max 0 de de de de rad hoekverdraaing motor Nm rad s 2 s N omega tmp max accel t tmp k tmp omega tmp t tmp else k tmp omega max t tmp end omega omega omega tmp k k k tmp t tmp t tmp 001 t t t tmp end subplot 211 plot t k ylabel Angle rad subplot 212 plot t omega ylabel Angular velocity rad s xlabel Time s axis tight 40 Appendix A Rev_change_sti m Opschakel acties 1 gt 2 2 gt 3 3 gt 4 4 gt 5 5 gt 6 sequentieel o 0Opschakelen for k 1 length ratios 1 omega up k 1 omega k ratios k 1 ratios k omega k end omega up 0 omega up toevoegen van een nul om dat 0 gt 1 met de koppeling gebeurt ae Te bepalen terugschakel acties 2 gt 1 3 gt 1 4 gt 1 3 gt 2 4 gt 2 5 gt 2 6 gt 2 4 gt 3 5 gt 3 6 gt 3 de ge ae 5 5 gt 4 6 gt 4 8 6 gt 5 v car 30 25 40 50 50 60 60 80 80 60 100 80 140 3 6 g 2 3 4 3 4 5 6 4 5 6 5 6 6 r tire 316 Terugschakelen omegal for k 1 3 omega engine v car k r tire ratios g k i diff omegal tmp omega engine ratios 1 ratios g k omega engine omegal omegal omegal tmp end tl 0 73 0 92 1 62 dwdtl omegal tl omega2 for k 4 7 omega engine v car k r tire
3. kene 11 kijken Synchromesh afstand die schakelmof aflegt van neutral tot synchro ring aantal wrijvingsvlakken in synchromesh naar 1 of 2 e o I rd o I xd TI L re LN kj mtn ok I eme bekend mm 1 naar 4 naar 5 naar 6 Stijfheid veren op de schakelas Overbrengingsverhoudingen automotive notatie Hartafstand prise as en pignont as 0 0058 0 004 j PP 2 Kiken 1 Kiken 115 graden tea mee 5 70E 02 0054 m d meten 6 905 02 om m dc Meten 6 005 02 0065 m dN Meten PP aschatten 7800 EEE EE el er ee Be KE uw 22 Alle oa o PP 2 mee 0 0170 0 0085 17 3E 3 23865 m a TI 00141 0008 142 m ge 0 0141 1562 m d Dm 0 0141 1462 m Io Il 0 0141 1462 m P PI 0 0141 155E2 m ae 0 0141 1482 m a 0 0141 fm 8 PJ 0 0141 542 ee 0 0141 1452 m eco h h 5 Lis meten Los metn 0 0017 915 afschattenenuitrekenen red afschattenenuirekenen 0 0743 RE 11 T md a 0 0103 ri Paar 4 Seats Stifheid veren op de schakelas I Overbrengingsverhoudingen automotive notatie 3 45 1 94 1 29 0 97 0 8 650 rads omega TI 650 rads omega2 rads omega EE I 550 mds omg 555 rads ome Ee ee NG 332 rads omg T
4. 360 F1 Ferrari F1 Enzo Bugatti Veyron proposed Ferrari F1 575M 220 ms BMW SMG M3 E36 me oo 220 Ms Aston Martin Vanquish 2506 BMW SSG 3 series 250ms 150ms for 1st to 2nd Alfa Selespeed 156 Selespeed old 700 ms Table 1 Shift times specified by manufacturer In this table only hydraulically actuated systems are shown In the following figure a comparison concerning an unknown hydraulically operated gearbox and the electromechanically operated gearbox Easytronic from Opel hydraulisch maximale Kraft elektromotorisch maximale Kraft komfortabel oe aS cS Momenten Gang Synchro Gang Momenten abbau raus nisieren ein aufbau Figure 1 Shift time comparison hydraulic versus electromechanical According to Figure 1 the electromechanical actuated systems are slower than the hydraulically actuated ones The source of this picture doesn t mentions which hydraulic system is used in this comparison It can be concluded that for performance a hydraulic system is preferred and looking at Table 1 the fastest system is the SMG implemented in the BMW or the one the Ferrari F1 implemented in the Maserati 4200 GT These systems are obviously shift the quickest 2 2 Comfort This is a very subjective parameter but nevertheless it cannot be ignored Since have not been able to test myself have based my conclusions on the available
5. Conclusions An electromechanical actuation system has some obvious advantages compared to a hydraulic actuation system especially in the field of cost and packaging Its lower power density is its main disadvantage but we have seen from the calculations that it should perform quite well More over it is possible to increase performance which is desired at up shifts The increase in performance can be obtained by fitting a more powerful electromotor However thermal restrictions have to be taken into account since the performance of the synchromeshes will be met 9 2 Recommendations Some important parameters used in the calculations and which are stated in the specification sheet in Appendix H might differ from reality and these parameters should be checked Also the parameters of the specific work power and contact pressure are not from this transmission so they could be wrong 29 10 Bibliography 10 11 12 13 14 15 Antrieb und Getriebe Aral De complexe aandrijflijn Kluwer technische boeken Automatische Fahrzeuggetriebe Springer Verlag Automotive Transmissions Fundamentals selection design and application G Lechner H Naunheimer Springer De elektronische versnellingsbak Zelfstudieprogramma 221 Constructie en werking Technology survey on smartness added to automotive manual transmissions J D W de Cock Das tribologische Verhalten von Syncrhonisierungen unter Ber cksichtigung B
6. In this case this will be the outgoing axle J p D AE Equation 2 32 Inner diameter m Outer diameter m Thickness m Density kg m R200 Jim 7t J D Diu EL Using 32 Equation 2 and the data in Appendix H results in the following table 1 957E 5 1 551E 3 Agel TAMES 1 3416 4 6 gear OS 4 878E 4 A gear IS D 4 241E 4 4 gear OS Js 1 613E 4 5 gear IS Je 3 666E 4 5 gear OS Jio 1 927E 4 Clutch Jc 4 086E 3 Ingoing axle Jis 1 729E 3 Table 5 Inertia s OS means outgoing axle and IS ingoing axle In the table above the calculated inertia s are shown The dimensions of the gearwheels and axles and therefore the inertia s are based on the information of a picture from the service manual This picture is shown in Figure 11 The dimensions of the gearwheels were measured in this drawing and the information is represented in Appendix Because the inertia s were calculated and their dimensions were extracted from drawings the obtained values contain errors and should be considered estimates 1 gear IS means the following the gearwheel of 1 gear fixed to the ingoing axle The inertia of the clutch is quite a rough estimate because the drawing does not represent it very clear 6 2 Determining the relative angular velocities In the following table a survey of the angular velocities of the inertia s in each gear is presented This is necessary to lump the inertia s with respect to one angular velo
7. are integrated in the actuation modules which makes it a very compact and simple add on system 3 2 Implemented modifications Enttali Kupplungsweg Sensor Entrall Handschaltdom 66 65681 inkt Schaltbet tigung SAC Entfall Getriebeeingangs Entfall Kupplungspedal rehzahl Sensor Kupplungsaktor Figure 3 System components in the figure above the modifications made to the gearbox when compared to the standard manual gearbox are graphically shown as well as the sensors that should be added when a hydraulic system was chosen Opel removed the clutch pedal and there is no need of a sensor measuring the number of revolutions of the ingoing axle of the gearbox or an extra clutch position sensor The latter sensor is integrated into the clutch module However a different clutch is mounted a so called Self Adjusting Clutch from now on referred to as SAC The advantage of such a clutch is that it compensated for wear This results in a constant force during its life cycle and makes it easier to control Another advantage of the SAC is it requires less force to open This can be seen in Figure 5 conventional clutch SAC coil spring sensor DS cover stop Figure 4 Self adjusting clutch working principle Feensor Fieaf Feushion Sensor DS actuation travel Figure 5 Clutch actuation force System design of an electromechanical Automated Manual Trans
8. as well However the 23 specifications of the selection motor are unknown The reduction from rotation of the motor to displacement of the selection rod is also unknown and the only information about the required time can be derived from Figure 10 According to this figure it takes about 100 ms to complete th s action From this figure it can also be concluded that before initiating the selection procedure the shift lever does not have to be in neutral So it does not take an extra 100 ms when also a selection action is required For the length of the shift time no concrete information can be derived from the figure because the circumstances are unknown 7 4 Synchroniser performance limits In the preceding paragraphs the minimal shift time is calculated However it is not sure the synchromeshes can dissipate the heat that is generated at these shift speeds If the synchronizer has to process too much power the tapers will become too hot so the performance is determined by the thermal stress This causes the material properties to change which results in a lower friction coefficient and thus a damaged synchronizer which will not function properly any longer By calculating the specific frictional work which is defined as the absolute work divided by the various gross friction surface areas the maximum permissible work can be defined W La Equation 14 Ar With the gross friction surface defined as d d r ms 9 9 10 d Equatio
9. however the shift and selection actuator are not applied in this model LuK developed these modules and Bosch supplied the electro motors The motors are based on motors used to actuate door windows and are developed further to meet the requirements for gearbox operation The actuator consists of an electromotor worm wheel a gear wheel with cam and a piston plunger The rotation of the electromotor is translated in a translation of the plunger generating an oil flow and ultimately in the disengagement of the clutch The used worm is self locking so no power is needed to maintain a certain position According to publications from LuK the clutch actuator uses less then 10 W Figure 9 Clutch actuator The most important components in Figure 9 will be discussed here Component 101 represents the electro motor connected to the worm with number 112 by means of a shaft with number 102 This worm rotates gearwheel 113 the bearing of this gearwheel is indicated by number 114 A cam is mounted on gearwheel 113 and this cam translates the rotation of gearwheel 113 in a translation of the plunger 116 The oil moved by this plunger is situated in chamber 121 which is connected by tube 122 to a piston 123 that operates the clutch 3 2 4 Easytronic control signal 11 From patent GB2313886 an example for a possible control signal is obtained This control signal is shown in Figure 10 More information concerning the control of the Easytronic actuation s
10. information which was sometimes provided by the manufacturer The BMW SMG system which had the best performance as we have seen is also very uncomfortable Due to the fast gear changes the torque is not decreased gradually but it is suddenly interrupted This can be recognized by an excessive nodding movement of the heads of the people in the car and it is not comfortable Looking at Figure 1 again we can see that especially the decrease of the transmitted torque is essential in the determination if the gearshift is experienced as comfortable However if a gearshift takes too long the driver experiences it as irresponsive and annoying All of the automated manual gearboxes mentioned earlier have different characteristics varying from a sportive program to a comfortable or normal shift program From Figure 5 we can see that an electromechanical actuated system shifts fast enough to comply with a comfortable gear change 2 3 Cost The used components in the automated manual gearboxes especially determine the price of such a system The disadvantage of a hydraulic system is that it has quite a lot of components compared to an electromechanical one 2 4 Packaging Figure 2 Left shift and selector actuator Right clutch actuator There are only two modules that have to be added to the gearbox and they are shown in Figure 2 These modules contain all the needed sensors actuators and even the control unit which is integrated with the c
11. ratios g k i diff omega2 tmp omega engine ratios 2 ratios g k omega engine omega2 omega2 omega2 tmp end t2 0 46 0 72 0 98 1 dwdt2 omega2 t2 omega3 for k 8 10 omega engine v car k r tire ratios g k i diff omega3 tmp omega engine ratios 3 ratios g k omega engine omega3 omega3 omega3 tmp end t3 0 35 0 557 0 65 dwdt3 omega3 t3 omega4 for k 11 12 omega engine v car k r tire ratios g k i diff omega4 tmp omega engine ratios 4 ratios g k omega engine omega4 omega4 omega4 tmp end t4 0 25 0 47 dwdt4 omega4 t4 omega5 for k 13 41 end omega engine v car k r tire ratios g k i diff omega5 tmp omega engine ratios 5 ratios g k omega engine omega5 omega5 omega5 tmp t5 0 45 dwdt5 omega5 t5 pees Determining the most critical shift actions max dwdtl max dwdt2 max dwdt3 max dwdt4 7 D max dwdt5 omega down max omegal max omega2 max omega3 max omega4 max omega5 0 omega max omega_up omega_down adapting for synchromesh_sti m 42 Appendix F Check_ist m Checking the results of synchromesh_sti n 5 Parameters molybdenum WA 0 53 J mm 2 Specific frictional work WA peak 1 5 5 J mm 2 Specific frictional work at peak load PA 0 84 W mm 2 Specific frictional power pR 6 N mm 2 Conta
12. 22 plot W down t r xlabel Friction Work W J title Downshifts axis 5000 0 0 1 legend 4 gt 1 6 gt 2 6 gt 3 6 gt 4 6 gt 5 2 grid on subplot 121 plot W up t r xlabel Frictional Work W J ylabel Permissable slip time t r s title Upshifts axis 1500 1 legend 1 gt 2 2 gt 3 3 gt 4 4 gt 5 5 gt 6 1 grid on figure subplot 121 plot abs P up 1000 t r xlabel Power kW ylabel Permissable slip time t r s title Mean friction power at upshift axis 0 10 0 0 7 legend 1 gt 2 2 gt 3 3 gt 4 4 gt 5 5 gt 6 1 grid on subplot 122 plot abs P down 1000 t r xlabel Power kW title Mean friction power at downshift axis 0 10 0 1 legend 4 gt 1 6 gt 2 6 gt 3 6 gt 4 6 gt 5 2 grid on 35 figure plot F down t r title Downshift xlabel Gearshift effort at sleeve F N ylabel Permissable slip time t r s axis 3500 0 0 t eind legend 4 gt 1 6 gt 2 6 gt 3 6 gt 4 6 gt 5 2 hold on grid on figure plot T r down t r xlabel Friction Torque T r Nm axis tight hold on plot W_down t_r xlabel Frictional Work W J axis tight 1 Das tribologische Verhalten von synchronisierungen unter Berucksichtigung des Beanspruchungskollektivs KSN 97 LOE 2 blz 246 Lechner amp Naunheimer 36 Appendix D Sti m Parameters voor versnellingsbak Volkswagen 02K DNZ
13. 5 In Figure 17 and Figure 18 the values in the table are very obvious as well as the impact on the permissible slipping time At up shifts the only problem is the shift action from 1 to 2 gear limiting the permissible slipping time to a minimum of0 4 seconds When shifting down the thermal performance of the synchronizer is a bigger limitation especially when shifting from 4 to 1 gear 26 8 Interpretation of the results First we will compare the required shift times with the calculated shift times of the actuation system We can combine Table 3 with Table 7 and Table 4 with Table 8 however this is not a fair comparison because Table 7 and Table 8 represent only the synchronization time The time the actuator needs to move from the engaged to the chosen idler gear must be added Gearshift Minimum achievable slip time s Desired slip time s Table 11 Comparison achievable and desired downshift times When changing to a higher gear there are not a lot of problems only changing from first to second gear takes too long When shifting to a lower gear however a problem occurs when shifting from fourth to first gear This takes very long and this is due to the high inertia of the gearbox which is a result of the first gear ratio Now that the minima of the shift times in the worst case scenario are calculated and the minimum shift times looking at the thermal stress e
14. Automation of VW transmission 02k DNZ Internal traineeship DCT report 2004 21 Date 05 04 2004 Author A J Baeten Supervisor dr ir R M van Druten Index le ernennen EM 2 dt Itten emer ree bane e ER EE d 3 2 Choosing an actuation system eee eem nennen 4 2 4 Performance nno eater eere e e nennen e dla 4 22 Comfort EE 5 PAMELA 5 2 4 GE Un BEE 5 2 5 Goncl si n uu users AE S 6 3 Opel Corsa EasytrOn ies onu etia nnde HE nn 7 3 1 Reason for automation eiae eee nennen Rb Pn nenne 7 3 2 Implemented modifications 7 3 2 1 Shift and selector actuator deci can c I eU HERD der a SEU 9 3 2 2 Actuation lay QUE icis ctt does 10 2 re IE ni been 11 3 24 Easytronic control signal ienei eem nnne eene 11 A Transmission 02k DNZ overview cesses enn n hn en nens enn nene nennen 13 44 Transmission lay OUt 2 ires eret ener nhe e iR Hera EE XXR RR d uwen Lee eni 13 4 2 Properties of the synchromeshes E 15 5 Req ired specifications sassi naiean Naa ANEA aa Ee ERROR eve E e storys 16 5 1 Time available irr iiie eere ege dee tbe a PR p add eves an 16 5 2 Finding the most critical shifts ENEE 16 6 Reductio
15. I Nm TV ____geschatiechner _ N Fes geschatlechner KEN a TRUE maximale omwentelingssnelheid schakelactuator 5000 omwmin n LuKkapi3 pdf 0 813 2 505 05 132 5 LW P LuK kap 13 paf 0 2531 Me 1 1 1 berekend 2500 radim actuatie totaal patent schakelelasticiteitpdf Table 14 Dimension and parameter survey 47 Appendix 1 neuer Motor EC 267 Wikg heutiger Schaltmotor DC 165 Wikg heutiger Kupplungs motor DC Leistungsdichte 101 Wikg 268 Massentr gheit 25 0 10 kgm 6 5 109 kgm 400 mechanische 27 75 ms Zeitkonstante 400 Gewicht 6939 Volumen Bild 16 Kennwerte neue Elektromotoren 48
16. ble 3 and Table 4 we can calculate the corresponding angular velocity ofthe engine This is done using the following equation 30 3 6 i Lag gear Equation 1 motor T wheel The value of the wheel radius Leet as well as the differential ratio ig and the gear ratio Leer can be found in Appendix In the next section it will become clear why exactly these shifts will be discussed 5 2 Finding the most critical shifts First it will be shown that the gearshifts in Table 3 and Table 4 are the most critical ones Using O motor m 30 3 6 TT ug dar wheel Equation 1 and the data from Appendix the following two figures are made The first figure represenis the loss in angular velocity at up shifts and the following figure the increase in rotational speed when shifting down The shift actions at which the largest change in angular velocity occurs are the most critical ones In these situations a lot of power has to be dissipated However the change in rotational velocity alone is not the only parameter to determine which shift actions are the most critical ones The available time for the gearshift is important as well 16 6 Reduction of moments of inertia 6 1 Calculation of the inertia s Because of the construction of a constant mesh gearbox the gearwheels are subject to different angular accelerations In order to be able to use only one angular velocity for all the masses involved they will be lumped to one axis
17. city 17 Selected gear Ja Jis Mery Za 23 2 Zg 21 Table 6 z are the number of teeth on a gearwheel indicated with its inertia The clutch ingoing axle and inertia s J1 and J3 are all fixed to each other so they always have the same angular velocities The rotational speeds stated in Table 6 are reduced with respect to the outgoing axle This means that if multiplied with the angular velocity of the outgoing axle the actual angular velocity is returned 6 3 Lumping the inertia s When the transmission is shifted as in Figure 12 to first gear with inertia s reduced with respect to the outgoing axle of the transmission the corresponding lumped inertia is 2 2 Jaar Jr HI Jot JU tat E 5 Equation 3 24 re 2 On the assumption that the output shaft OS and the components connected to it are not subject to any change of angular velocity during synchronization their moments of inertia will be ignored This is a valid approximation when the road has no gradient 18 7 Calculating if the solution will work 7 1 Simplifications and assumptions The following simplifications are made to use this calculation method Oil temperature of 80 C The gearshift effort F is constant The friction coefficient u is constant Torque losses Ty are constant Friction torque Tg is constant Change in angular velocity is constant aon The errors resulting from the simplifications ar
18. ct pressure v perm 7 m s Permissible friction speed Calculations AR le6 pi d N 2 sqrt 9 9e 3 2 d N 2 2 W perm W A A R W perm peak W A peak A R P perm P A A R F perm p R A R Q Comparison results if W perm gt max max abs W_up disp Upshift frictional work is within limits else disp Upshift frictional work is too large end if W perm gt max max abs W_down disp Downshift frictional work is within limits else disp Downshift frictional work is too large end if P perm gt max max abs P up Momentarily transmitted power disp Upshift specific friciton power is ok else disp Upshift specific friction power is too large end if P perm max max abs P down disp Downshift specific friciton power is ok else disp Downshift specific friction power is too large end if perm gt max max abs F up disp Upshift contact pressure is ok else disp Upshift contact pressure is too large end if F perm gt max max abs F down disp Downshift contact pressure is ok else disp Downshift contact pressure is too large end 43 Appendix G Opschakelen Schakelactie 19 2 Bak Rem Koppeling 15 2 25 Schakelactie 1 3 Bak Rem Koppeling 05 1 15 ES i Schakelactie 2 3 25 3 55 4 Bak sat Rem Koppeling
19. e largely offset in the calculation by the acceptable stress values in the synchromeshes The acceptable stress values are obtained from Appendix H and are derived from experience 7 2 Required synchronization force In this section calculations will be made to estimate if the actuation system can comply with the required specifications The required specifications are a worst case scenario which means that shifts will be made when the engine is at maximum velocities so the axes in the transmission will have their maximum inertia Decelerate Accelerate Figure 13 Synchronisation of two equivalent rotating masses To estimate the shift times the torque equilibrium for a synchronizer as in Figure 13 must be solved d T T 7 0 Equation 4 When the master clutch is fully opened the load moment T is zero throughout the synchronizing process The torque losses Ty are the result of bearing losses oil churning losses oil drag losses and oil compression losses These losses are specific in each individual transmission and should be measured When shifting up the gearwheel to be shifted is decelerated with the rotating masses reduced to its axis 4 Friction torque and torque losses act in the same direction When shifting down the 19 gearwheel to be shifted is accelerated with the rotating mass reduced to its axis Friction torque do T Jna 1 TT 0 and torque losses act in opposite directions W
20. eanspruchungskollektivs Dipl Ing Tobias L sche 1997 Getriebe in Fahrzeugen 2001 VDI Berichte 1610 Patent number GB2316723 Patent number GB2313886 Patent number WO03087632 Patent number DE19725816 Patent number W003087628 Patent number WO03081091 30 d Appendix A Synchromesh dimensions Figure 19 Schematic representation of a synchromesh The symbols used in Figure 19 are d do dc F FN S a Effective diameter Nominal diameter Clutch diameter Gear shift effort Normal force shift movement at the gearshift sleeve Taper angle 31 The steps in the synchronization process Figure 20 Syncrhonisation process 32 Appendix 5 ue 2 m T N oo T s Figure 22 Position 2 33 Appendix C Synchromesh sti m Calculation of the shift force clear all close all clc format long Gearbox specific parameters sti rev change sti General and estimated parameters TV 2 Nm Torque losses F Hperm 100 N Permitted hand shift force t Rperm 0 25 s Permitted slip time F kies 60 N selection force 30 60 Slip time t begin 02 t stap 01 t eind 1 oe Start time time step size End oe oo are the tapers self locking if tan alpha gt mu disp No self locking occurs else disp Self locking of the tapers end Calculating opening torque Z F a d C 2 cos beta 2 mu D s
21. in beta 2 sin beta 2 mu D cos beta 2 T Z 1 2 F a d acos beta 2 vereenvoudiging negeren van de frictie coefficient mu D for u 1 2 u 1 upshift u 2 downshift maximum difference in sliding velocity v max omega max u d N 2 also possible with omega Tr 17 W I P li 5 for t_r t_begin t_stap t_eind Required friction torque T r tmp J red omega max u t r T V Tr T r T r tmp Friction work per shift action W tmp 1 2 omega max u t r T V 1 2 J red omega max u 2 W W tmp Mean friction power tmp W tmp t r P P tmp Power transmitted momentarily to the synchromesh Ps tmp T tmp omega max u de de HG tg ce 3i 34 Ps Ps Ps_tmpl end oo Results in the required axial force belonging to a certain slip time F T r 2 sin alpha d mu Correct for the number of friction tapers for 1 length ratios 1 F 1 F 1 3 1 end LE uv T_r_up T_r W_up W P up P Ps up Ps F up Fi end if u T r down T r W down W P down P Ps down Ps F down end clear F P T r Ps end Check ist Plotten the results t t begin t stap t eind Time axis for plotting figure plot F up t r title Upshift xlabel Gearshift effort at sleeve F N ylabel Permissable slip time t r s axis 0 1500 0 0 7 legend 1 gt 2 2 gt 3 3 gt 4 4 gt 5 5 gt 6 1 grid on figure subplot 1
22. in combinatie met STI scale 65 23 1e 3 Schaal van de tekening de do rho 7800 gebruikte tandwielen kg m 3 dichtheid materiaal Versnellingsbak specifieke parameters mu 1 gemiddelde wrijvingswaarde in het konusvlak 1 alpha 11 42 pi 180 rad halve kegelhoek van de synchronisers beta 115 pi 180 rad kegel hoek opening angle van de tanden op de synchronisers d 54e 3 m effectieve diameter halverwege wrijvingsopp synchro d C 70e 3 m clutch diameter v d synchromesh waar de dogs zitten d N 55e 3 m nominal diameter voor definities van bovenstaande drie zie 1 blz 234 i actuation 2500 rad m Total gear ratio of the actuation part 65e 3 m hartafstand tussen prise en pignont as omega 650 650 650 650 565 332 rad s toerental waarbij geschakeld wordt worst case maxvermogen 6200 rpm maxkoppel 3200 rpm s 10e 3 m Weg die de gearshift sleeve aflegt 10 13 mm f l l l TT Allemaal enkele wrijvingsvlakken volgens Christian z 11 38 18 35 28 36 32 31 41 33 55 31 5 3e versnelling wordt 6e tandwielen omdraaien nog niet gedaan ratios z 2 z 1 z 4 z 3 z 6 z 5 z 8 z 7 z 10 z 9 z 12 z 11 Wordt gevraagd door synchromesh m i diff 68 18 overbrenging differentieel oo Gemeten waarden uit tekening ae Traagheid ingaande as L as 29 4 8 1 11 10 9 scale van rech
23. ith all this in mind dt Equation 4 reduces to do Te Sad Equation 5 R d wv For up shifts the term de becomes negative and now it works in opposite direction of Ty With the friction torque the friction work can be calculated using 2 W ot Ty 4J pd Orax Equation 6 W tot T 5J red Equation 6 Dividing the friction work by the permissible slipping time results in the average friction power as in W r Equation 7 W P Equation 7 t The permissible slipping time is defined as the time it takes to engage an idler gear With the friction torque and the dimensions of the synchromesh the force practiced on the synchromesh sleeve can be obtained using the following equation T 2sin a F Equation 8 du do T T Using dt Equation 5 to T 2sin F Ke du Equation 8 several plots can be made with varying permissible slipping time When the permissible slipping time is known the corresponding values for T W P and F can be found The maximum permissible slipping time available for each gear change is stated in Table 3 and Table 4 20 Permissable slip timet 5 Permissahle slip time 1 s 1 3500 3000 Figure 15 Critical down shifts EDU 1000 Gearshift effort at sleeve FIN 2500 Downshift 2000 1500 1000 Gearshift effort at sleeve F N 600 T rie rn Tain E eg S zi T
24. le and then make a stop after second gear is engaged The specifications of the electromotor are given in Appendix they determine the performance One parameter is not presented yet and that is the maximum angular velocity being 5000 rev min which equals 524 rad s The length of the shift stroke is 10 mm from neutral position to when the gear is engaged so when another gear is engaged the stroke must be covered twice being 20 mm With the given 22 specifications of the electromotor the resulting acceleration is plotted in Figure 16 and is calculated with Equation 11 40 T T 30 20 10 1 0 BEN en L L 0 0 01 0 02 0 03 0 04 0 05 0 06 0 07 600 T T T T T 500 e 490 J 300 4 200 4 100 J 0 1 0 0 01 0 02 0 03 0 04 0 05 0 06 0 07 Tijd s Figure 16 Actuation speed The displacement is calculated with the following function do M zs Equation 11 dt M rad ll 10080 Equation 12 ot J Ss O emations Equation 13 After t 0 052 s the maximum velocity is reached M 0 252Nm n zm Equation 9 states it takes 25 radials to move from the engaged position to neutral From Figure 16 can be derived it would take 0 051 s to bridge this distance The time as stated above is the time required for a shift action as in a shift from 1 to 2 gear When shifting from 2 to 3 gear a selection movement must be made
25. lutch actuator as shown right in the figure above All sensors which are not integrated in these modules but are required for the controls are already present in a manual gearbox The control unit needs some of the signals from these standard sensors so the control unit will have to be able to receive data from them 2 5 Conclusion An electromechanically actuated system is preferred mainly because its low cost and its compactness It is also simple because it is a completely dry system it cannot leak and eliminates difficult sealing problems Another reason for choosing the Easytronic system is that it is well documented compared to the Sensodrive solution of Citroen Getrag which is a comparable solution A possible explanation why information concerning the Sensodrive system is rather scarce is that it is introduced recently whereas the Easytronic system was introduced in 2000 3 Opel Corsa Easytronic 3 1 Reason for automation Opel automated a manual transmission of the Corsa because they wanted to offer their costumers an extra bit of comfort Two electro motors power the actuators This is done because a system driven by two electro motors is a low cost solution compared with an hydraulic actuated system The following parts are superfluous when an electromechanical system is compared with an electro hydraulic actuation system pump accumulator and solenoids this makes it less expensive The extra sensors required for an automated system
26. mission system ignition switch Figure 6 Block scheme of the easytronic system In the block scheme in Figure 6 a schematic representation of the connections between the different components of the Easytronic system is shown 3 2 1 Shift and selector actuator The actuators of the system are working electromechanically as mentioned before A disadvantage of such a system is its lower force density compared to a hydraulic system By means of some innovative features this system can shift almost as fast One of these features is a so called build in shift elasticity The advantages according to LuK are Minimized free flight phases Nostop of electro motor while synchronizing Constant shifting comfort Protection of transmission and actuators The free flight phase is a part of the shift action where power is only needed to move the shifting rods No forces other then friction and inertia have to be overcome The electro motor does not have to stop and wait for synchronization because of the shift elasticity however the motor runs at reduced speed until the gears are synchronized An example of how the control signal could look like can be seen in Figure 10 The shift rod cannot move further until the rotating masses rotate at the same speed however a force exerted on the shift rod is need to push the gear against each other The increa
27. n 15 With the data given in Appendix H this results in an Ag of 2 5E 3 m The synchronizer ring is made of molybdenum This material can deal with 0 53 J mm resulting in a maximum permissible work of 1325 W Transient peak loads significantly higher than those given may be tolerated The peak value for specific frictional work Wa in the synchronizer ring friction linings for molybdenum is 1 5 J mm allowing a work of 3750 W Additional parameters are the permissible friction speed power and contact pressure All these values are calculated in a Matlab script in Appendix C Running this file leads to the following conclusion the synchromeshes are too small to shift this fast In the worst case scenario the maximum transmitted power for an up shift is 2 03E5 where the maximum power allowed is only 2 12 10 as in Table 9 In this situation the synchromeshes will be damaged Unity Permitted value WI 1 3453 PI 2 12E3 1 5254 Table 9 Values of check_ist m 24 Upshifts s r Permissable slip time t Et 1500 1000 500 Frictional Work W J Figure 17 Left up shift work Right downshift work Mean friction power at upshift Permissakle slip timet ts Power kVV Figure 18 Left up shift power Right downshift power gi i 1 4 5000 4000 2000 2000 1000 Downshifts Friction Work W J Mean friction power at downshift 5 Power kW 10 2
28. n of moments of inertia EEN 17 6 1 Calculation of the inertia s uk enne nnne nennen nnns 17 6 2 Determining the relative angular velocities AEN 17 6 3 Lumping the iriertia s cerei abe eri E rne seh E Dess de Eee 18 7 Calculating if the solution will work eee nee em nennen 19 7 1 Simplifications and assumptlons 7 19 7 2 Required synchronization force 19 7 3 Properties of the actuating system En 22 7 4 Synchroniser performance limits ENEE 24 8 Interpretation of the results eese nnne 5559396945 5599444 27 9 Conclusions and recommendations EEN 29 E ee Te DEI AEROSOL OD DIDIT IRI 29 9 2 Recommendatiors nece te t essen Pe d ces 29 10 Bibliography eer Age Dee disc 30 Appendix sne EE 31 Appendix Bi P Easo 33 Pioneer 34 APPSNdX Dr ns T 37 Appendix Er P ERES 40 Append E EET 43 Appendix Gude karene Auen 44 Appendix o ii anstehen Saeed ue 47 Appehdibx ASR EE sd 48 1 Introduction This report regards my internal traineeship the goal of the report is to look at the possibilities for automation of a manual transmission from Volkswagen indicated by VWO2k DNZ In this report will be studied if existing systems for automating a manual transmission are applicable to the Volkswagen gearbox These systems are implemented in for example the Alfa Romeo 147 and BMW M3 but also in ra
29. se in load practiced by the electro motors as they continue to rotate is represented in Figure 7 Because the electromotor did not have to stop it reaches its top speed again in the final phase of the gearshift so the gears will be locked faster SEN ZZ EIC N Nom E VE EES 0 ff d 210 Figure 7 Left effect shift elasticity Right implementation of shift elasticity 3 2 2 Actuation layout The layout of the actuation part is shown in the next figure 1100 Figure 8 Lay out actuation part In this figure the layout is orderly represented and the components are easy to recognize The components indicated with the numbers 1104 to 1107 are the implementation of the shift elasticity Part 1104 is connected to shaft 1140 by means of the springs required for the shift 10 elasticity a detailed representation is given in Figure 7 so is not connected rigidiy to this shift Worm wheels are fitted at the shafts 1103 and 1120 of the electro motors Shaft 1111 performs the selection of the desired gear At the top of this shaft a groove can be found the shift finger fits into this groove and converts a rotation into a translation of shaft 1111 Parts 1112a and 1112b are the shift fingers where 1112a operates the shift rods 1130 and part 1112b may operate for example a reverse gear 3 2 3 Clutch actuator The clutch actuator is also operated by an electromotor This actuator is also used in the Mercedes A class
30. th Symbol indicating number of teeth nel Table 2 Gear teeth Zs and Ze are not linked to a gearwheel in Figure 12 This is due to the already mentioned modifications made to the transmission with respect to the third gear This gear will not be realized by gearwheels but in combination with a planetary gear set However the same gear ratio as the original third gear is used only in a different way It takes to far to explain this in detail 4 2 Properties of the synchromeshes Very important parts of the transmission are the synchromeshes Their main objective is to match the speeds of two rotating masses to each other This is done by means of friction so we can think of synchromeshes as small clutches The dimensions of the synchromesh determine the performance of the gearbox as well as the shift comfort Especially the friction area is crucial for the performance the bigger the friction area the lower the shift force The dimensions of the synchromeshes of this transmission are measured and its specific values can be found in Appendix H In Appendix A the different stages of the synchronization process are explained as well as the symbols used in Appendix H 15 5 Required specifications 5 1 Time available The required specifications of the actuation system are given in Appendix G here a complete overview of the shift actions will be given 3 Table 4 Critical downshift times at corresponding vehicle speed From Ta
31. ther small cars like the Opel Corsa and the Citroen C3 These existing systems will be examined and then one system will be chosen based on the following criteria performance cost and complexity Because of the length of the report only the specific details of the chosen system will be included Then the specifications the gearbox has to comply with are presented and in the following chapters the shift time of the gearbox will be estimated In the following chapters the performance of the gearbox and actuating systems will be determined by solving the torque equilibrium on the gearbox masses Finally the performance of the gearbox and the actuation system will be compared with the desired specifications and some conclusions and recommendations will be made 2 Choosing an actuation system Now one of the before mentioned systems will be chosen based on the following criteria Performance Comfort Cost Packaging 2 1 Performance The performance of a system is represented by its shift times The faster a system can shift the better its performance However manufacturers are always vague under which conditions the given shift times are valid This makes it difficult to predict how they will perform in a different gearbox and different conditions which can be less favorable The following table with shift times is therefore no more than a rough indication Gearbox car l BMW SMG M3 E46 Ferrari F1 Maserati 4200GT Ferrari F1
32. ts naar links op de tekening plus synchromeshes D as 7 9 9 14 28 9 28 scale idem ook inclusief de synchromeshes D as i 4 scale inwendige diameter uitsparing voor pen van de koppelingsbediening J IS sum D as 4 L as D as i 4 sum L as pi 32 rho Traagheid Koppeling D clutch 0 03 0 14 210e 3 gedeeltelijk uit easydata D clutch i 0 02 0 03 0 14 uit easydata d clutch 0 04 0 011 0 0007 ook uit easydata 37 J C sum D clutch 4 d clutch sum D clutch i 4 d clutch pi 32 rho Traagheid J1 dl 6 3 scale D1 12 9 scale J 1 sum D1 4 d1 Traagheid J2 d2 5 1 scale D2 38 30 scale J 2 sum D2 4 d2 Traagheid J3 d3 5 scale D3 12 scale J 3 sum D3 4 d3 Traagheid J4 d4 5 scale D4 34 scale J 4 sum D4 4 d4 Traagheid J5 d5 5 scale D5 21 scale J 5 sum D5 4 d5 Traagheid J6 d6 5 scale D6 29 scale J 6 sum D6 4 d6 Traagheid J7 d7 5 scale D7 28 scale J 7 sum D7 4 d7 Traagheid J8 d8 5 scale D8 22 scale J 8 sum D8 4 d8 Traagheid 29 d9 5 scale D9 27 scale J 9 sum D9 4 d9 Traagheid J10 d10 5 scale D10 23 scale J 10 sum D10 4 d10 Traagheid van Jl t m J6 reduceren omdat ze hol zijn d i sum d1 sum d2 sum d3 sum d4 sum d5 sum d6 s
33. um d7 sum d8 sum d9 sum d10 Di 4 5 4 5 4 5 4 5 4 5 scale binnendiameters J1 J2 J10 D i 4 d i J J i pi 32 rho corrigeren van traagheden en opslaan in een rij Hr 96 C Berekende waarden van de diameters en hierbij behorende traagheden ter controle for k 1 2 10 Db k 2 z k z k 1 h zZ k Zz k 1 1 05 3 2 h z k z k 1 41 38 end Jb Db 4 5 pi 32 rho Gereduceerde traagheden bepalen zodat er met 1 hoeksnelheid gerekend kan worden J red 1 1 J 2 t J 8 3 1 2 3 2 4 z 3 z 4 2 z 2 z 1 2 J red 2 1 J 4 J 8 21 1 2 3 J 2 z 1 z 2 2 z 4 z 3 2 J red 3 1 J IStJ CeJ 1 J3 3 J 4 z 3 z 4 24J2 2 z 1 z 2 2 z 6 z 5 2 J red 4 1 J_IS J_C J 1 J 3 J 4 z 3 z 4 243 2 z 1 z 2 2 z 8 z 7 2 J red 5 1 J_IS J_C J 1 9 3 J 4 z 3 z 4 242 2 z 1 z 2 2 z 10 z 9 2 J red 6 1 I 8 1 J 3 c 2 4 z 3 z 4 242 2 z 1 z 2 2 z 12 z 1 1 2 39 Appendix E Actuatie m Actuatie berekeningen Gegevens van de actuator oo oo n 5000 m 813 V 162 omega max n 2 pi 60 J schalt 25e 6 P 163 m tau 7 46e 3 phi s i actuatie totaal M P omega max max accel M J schalt t v max omega max max accel F shift
34. x S z Z a 1500 21 7 3 Properties of the actuating system Now we can calculate the maximum force the actuators can practice on the gear sleeves The maximum torque can be determined by M 20 0 252Nm n 27 Equation 9 M Z o9 0 252 Nm Equation 9 n 27 With the maximum torque and the transmission ratio of the actuation system the maximum force can be determined The transmission ratio is 2500 rad m which is the total ratio including everything from rotation of the electro motor till translation of the gear sleeve The transmission ratio is found in Appendix I F M n actuation 630N Equation 10 Now the maximum force the actuator can practice is known we can look at Figure 14 again and make the following table 324 foo 0 045 Table 7 Upshift slip times And for downshifts Table 8 Slip time for downshifts The required synchronization force is one component that determines the shift time The other component is the speed at which the actuation system can move from point A to point B also referred to as the free flight phase When for example first gear is engaged and the system must go to second gear the gear sleeve must be moved twice the shift stroke When a gear is engaged the shift rod and gear sleeve stand completely still So the electromotor should accelerate then slow down wait for synchronization no full stop because of the shift elasticity move further as fast as possib
35. xpressed by permissible work and power these results should be compared The minimum slip time for the calcufated gear changes were given in Table 7 and Table 8 The minimum shift times when looking at the thermal stresses are shown in Figure 18 Gearshift 1 22 Slip time s 0 03 Less then 0 03 Less then 0 03 8 gt 6 Table 12 Minimum thermal slip times for up shifts Gearshift Slip time Is Table 13 Minimum thermal slip times at downshifts When looking at these tables we can conclude that the thermal restrictions are not exceeded yet When we take a look at the minimum thermal slip times for gear changes and compare them with the required specifications the following remarks can be made 27 stronger electromotor can be used to increase performance since the synchronizers do not carry their maximum load yet However shifting from fourth to first gear cannot be much faster and a more powerful electromotor would cause damage to the friction material of the first gear The remaining gear changes meet their desired specifications These calculations are done in case of a worst case scenario so when driving normal and shifting with a more likely engine speed like for example at maximum torque of the engine the Easytronic system would be applicable with an exception for the fourth to first gear change at maximum engine speed 28 9 Conclusions and recommendations 9 1
36. ystem can be found in this patent Figure 10 Easytronic control signal 12 4 Transmission 02k DNZ overview 4 1 Transmission lay out In this chapter the transmission 02k DNZ represented in Figure 11 will be introduced Figure 11 Cross section view of gearbox 02k DNZ The gearbox shown above is used in the Volkswagen Golf and Bora as well as in the audi A3 Originally it was designed as a four speed transmission but a fifth gear was added afterwards This is obvious when we take a look of the cross section view above The gearwheels of the fifth gear are added onto the original casing and an extra lid was added to cover the extra gearwheels and synchromesh In the following figure a schematic drawing of the gearbox is presented Here it is more obvious which gearwheel is belongs to which gear However a change is made compared to Figure 11 A sixth gear designed as an overdrive replaces the third gear So the schematic figure is not the same as the cross section view in Figure 11 13 58 48 6th 2nd Ist Figure 12 02k DNZ schematic In the figure above the symbols used to indicate the inertia s are given The indices will be used from now on to refer to the specified gearwheel So gearwheel i is indicated by index i and the inertia belonging to the gearwheel will be represented as Ji 14 The number of teeth on each gearwheel is given in the service manual of Volkswagen and are summarized in Table 2 Number oftee
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