LEE- Otomotiv-Yüksek Lisans

Bu koleksiyon için kalıcı URI

Gözat

Son Başvurular

Şimdi gösteriliyor 1 - 3 / 3
  • Öge
    Optimization design parameter of dual mass flywheel coupled with a non-linear elastic path
    (Graduate School, 2023) Karakuş, Gökay ; Şen, Osman Taha ; 807252 ; Automotive Programme
    Today, expectations from automotive industry are high in the fields of performance, comfort, economy and environmental protection. That is why, the automotive industry must have a knowledge, experience, development processes, and strong research. Hence, the design of the components that transmit power from the internal combustion engine to the wheels via axes in a vehicle are crucial research topics. The flywheel, which is the primary element in power transmission system, works in engine in order to reduce the speed fluctuations on the crankshaft. Along with the studies in the field of flywheel development, dual-mass flywheels have been started to be used as an alternative to the single-mass flywheel. Briefly, the dual mass flywheel can be defined as the combination of two single mass flywheels. The spring damper system is combined with these two single mass flywheels. The dual-mass flywheel, which is used in diesel engines, is thought to have favorable effects on the dynamics of the power transmission system compared to the single-mass flywheel. In 1985, the first dual mass flywheel (DMF) was manufactured in automotive sector. In begin, dampers in flywheel were not lubricated and the springs are stand away from outside and created some wear problems. In 1987, DMF is lubricated with grease oil for the first time. Service life problem was no longer an issue thanks to grease oil application. Around 1989, arc spring damper was innovation for the DMF, and it had pretty much solved all resonance problems. Alson manufacturing costs were continually reduced. The primary mass of flywheel was made by casting or forged steel at first production batches. Over time, the primary mass was formed from sheet metal parts by metal-forming specialists. In 1995, folded masses were developed from sheet metal in order to increase inertia moment of primary mass of flywheel. This development led to the widespread use of the DMF. However, there is some disadvantages such as cost, the achievable improvements are seen clearly, therefore, DMF are used widespread in vehicles. There are some advantages of dual mass flywheel such as isolation from torsional vibration, relief of transmission and crankshaft. Within the scope of this study, firstly, give information about flywheel, parts of dual mass flywheel and its advantages. Then, while engine is running, the working principles of dual mass flywheel is mentioned. Also, the single mass flywheel and dual mass flywheel are compared and it has been observed that the resonance regions are reduced below the engine idle speed with the use of dual-mass flywheel. After all crucial information was explained, mathematical model of DMF is built. The model was solved in Matlab for all optimization. Sensitivity analysis was done in order to determine the parameter effects on system. The design parameters were determined as stiffness ratio, damping ratio and inertia ratio. Also, the system was running on non-linear cubic path, that is why, non-linear stiffness ratio is also important parameter to examine system correctly. The primary body represent not only primary flywheel itself but also engine side dynamic properties.
  • Öge
    Investigation of regenerative braking efficiency in different drive cycles
    (Graduate School, 2023-06-23) Barın, Berkay ; Şen, Osman Taha ; 503201705 ; Automotive
    As being an important fact, climate change threatens the world population and its impact has become crucial recently. Thus, the world is being prepared for a very fresh era: green energy and electrification era. The European Union proposes the EU7 emission regulation, which has lower limits compared to prior regulations; and net zero emission policies are already adopted by several governments throughout the world. These developments increase the importance of electric vehicles (EVs), which have already start to replace the conventional vehicles due to their zero tailpipe emissions. Furthermore, the high efficiency, quiet operation and improved performance characteristics of EVs make them more preferable, and the recent advancements in battery technology bring this preference to the fore. Though, EVs still have significant disadvantages such as limited driving range and long charging time. The concept of regenerative braking becomes crucial in increasing the battery state of charge, which improves the driving range and reduces the required charging duration. Thus, the optimization of the regenerative braking system operation becomes critical. The chief objective of this study is to investigate the utilization rate of regenerative braking in different drive cycles, which depends on braking rate and road conditions. Consequently, seven well-known different drive cycles are selected, which vary with distance, urban / highway scenario, traction per kilometres, etc. An E class sport utility battery electric vehicle is selected for modelling purposes and it is subjected to all drive cycles. The driving resistances are calculated and the instantaneous electric motor torque demand is obtained for all drive cycles. Furthermore, the total braking and traction force ratios are evaluated for all drive cycles. It is observed that the comparison of these drive cycles based on total braking and traction force values does not provide reasonable conclusions due to the variation of drive cycle range. Thus, results are compared by normalizing the traction and braking forces by the range of each drive cycle. Finally, a braking to traction ratio is determined for all drive cycles, and the tendency of regenerative braking utilization rate change is investigated. Based on the results, it is observed that the braking to traction ratio of the vehicle significantly reduces when the vehicle moves from urban to highway drive cycle. Furthermore, the utilization rate of the regenerative braking also drops down to %13 from %53, when the drive cycle transition from urban to highway occur. The low braking to traction ratio shows that the recovered energy begins to be insufficient in highway driving conditions, and enough energy cannot be provided for the charging of the battery pack. In addition, the test vehicle's weight has increased to its gross weight. Subsequently, results were reanalyzed in the view of weight change. Finally, it is observed that weight change has not a significant effect on utilization rate in city conditions, whereas it increases the efficiency around %6 in highway conditions.
  • Öge
    Reduction of engine torsional vibrations via hydrodynamic dampers
    (Graduate School, 2022-02-10) Aslan, Yavuz ; Akalın, Özgen ; 503191710 ; Automotive
    Within the scope of this master's thesis, the vibration damper used to dampen the torsional vibrations of internal combustion engines has been examined. In this manner, the hydrodynamic damper, which is a special type of vibration damper, is investigated. The calculation methods of the stiffness coefficient and damping coefficient, which are the two main characteristics of the hydrodynamic damper, have been studied. CATIA and Hypermesh software programs are used in order to perform the needed 3D model creation and analysis. Values obtained from the calculation by analytical and numerical methods are compared with test results performed on measurement systems. After the determination of these characteristic parameters, a crank train model is built over the AVL EXCITE program to examine the effect of the hydrodynamic damper in an 8-cylinder diesel engine. This established model is considered as two different sub-models, with and without vibration damper, and the differences between both cases are determined. Angular displacement and stress values in different sections of the crankshaft are used as the output of the analysis. Thus, the effect of the hydrodynamic damper used on the crank train and the engine has been clearly demonstrated. Currently, efficiency of the engine is increased by increasing power density and reducing engine volumes in internal combustion engines. On the other hand, this trend increases the loads on the components and affect the strength limits. In an 8-cylinder diesel engine with a high power density, the vibrations observed with increasing cylinder pressures and loads also increase. These vibrations seen in the crank mechanism directly affect the operation of the engine and the lifetime of the parts. For this reason, vibration dampers are used for the damping of these vibrations and for the safe operation of the engine by reducing their amplitudes. In this thesis, a hydrodynamic vibration damper coupled to the free end of the crankshaft is used. The hydrodynamic damper is a vibration damper that consists of leaf springs arranged in an inner part and that creates damping by providing oil flow between these spring packages. Firstly, the stiffness coefficient of this damper is determined. For this purpose, finite element analysis is applied to the damper, which is 3D modeled, through the Hypermesh program. As a result of the analysis, the angular displacement values against the acting torque are obtained and the related stiffness coefficient is determined. Also, a test system is designed to measure the stiffness coefficient. The stiffness coefficient value is obtained by the measurement made on this test system. The stiffness coefficient calculated by the numerical method is compared with the stiffness coefficient measured from the test system. As a result of the comparison, it is seen that both values coincided with each other within a certain margin of error. Analytical calculations have also been made to determine the damping coefficient, which is another specific feature of the hydrodynamic damper. The passage of the oil in the damper between the chambers next to each other creates a damping effect. In this context, the damping coefficient is tried to be calculated with two different methods. The first of these methods is the control oriented transient method, which is used in a similar study before. With this method, the damping coefficient is calculated over different coefficients based on the oil flow. However, since the geometry is very small and there are many parameters affecting the flow, the damping coefficient calculated with this method is found to be quite different from the damper coefficient in the catalog information from the manufacturer. The second method is to reduce the damper to an equivalent dashpot system. With this reduction, the damping coefficient is calculated from the dashpot. With this method, the damping coefficient differs from the value in the catalog, depending on the assumptions and reductions used. In addition to the calculations, a test system is designed for a static damping coefficient and it is measured on this system. The damping coefficient is obtained from the measurement result using the logarithmic decrement method. Since the engine operating conditions cannot be reflected in the test system and the damper is tested statically, not dynamically, the value obtained from the measurement does not represent the dynamic damping coefficient. For this reason, the values obtained from the calculations and the values in the catalog have not been compared. Finally, after determining the characteristics, a crank train model is created using the AVL EXCITE program to examine the effect of the hydrodynamic damper on the engine and to reveal the damping effect of the hidrodynamic damper. In this model, the crankshaft, connecting rod, piston and flywheel are modeled under the forces after combustion pressure and inertia forces. In order to see the effect of the damper on the system, two models, with and without the damper, are created. From these two models, the angular displacement and the resulting stress values in different parts of the crankshaft are investigated. The results from the model with the damper obviously stated that the hydrodynamic damper effectively reduces the torsional vibrations in the engine compared to the model without the damper. It can be said that the lifetime of the components and therefore the engine is extended since the stress values on the parts are reduced at the same time. It is claimed that adding torsional vibration damper into the crank train has no negative effect on the engine in terms of torsional vibrations. According to the results of the torsional vibration analysis, almost every part of the crank train elements has minimized amplitude level of torsional vibration and angular displacements with hydrodynamic damper. Additionally, 7 different damper case studies are revealed in order to determine the possible optimized stiffness and damping coefficients of the current hydrodynamic damper. In order to compare the different cases, the existing stiffness and damping coefficients of the current damper are changed by 25%. Updated stiffness and damping coefficients of the damper are applied into the torsional vibration analysis and corresponding results are obtained, respectively. These results are compared with the results of the current hydrodynamic damper. In some cases, the angular displacements and the shear stresses are increased due to the change in the stiffness and damping coefficients. On the other hand, some cases have improved results due to the higher stiffness and damping coefficients compared with the current hydrodynamic damper. These improved cases are investigated in order to determine the possible optimized version of the existing damper. It is not hundred percent possible to claimed that the revised damper has better results in the all speed intervals, but it can be said that the updated dampers with high stiffness and damping coefficients have better torsional vibration results in the most of the engine operating speeds. Regarding this aspect, Case 5 and Case 7 have better results than the other cases and the current hydrodynamic damper. Moreover, Case 5 has less angular displacements and shear stresses in the most speed ranges than the Case 7. Although Case 7 has better results in some speeds than the Case 5, there are some speed intervals that the Case 7 has higher angular displacements and shear stresses than the current hydrodynamic damper. Also, the reduction of the results in Case 7 is less than the increase coming from the Case 7 for specific speeds. Thus, the hydrodynamic damper in Case 5 has better results than the Case 7. It can be concluded that the revised hydrodynamic damper in Case 5 can be the optimized version of the current hydrodynamic damper in terms of torsional vibrations. Lastly, the needed structural changes in order to obtaine the stiffness and damping coefficients of the hydrodynamic damper in Case 5 are determined by using the formulations derived in previous sections.