车身结构轻量化优化设计方法外文文献翻译、中英文翻译、外文翻译
汽车制造工艺的发展外文文献翻译、中英文翻译、外文翻译
( 4 )计算机模拟冲压成形及虚拟试模技术:(5)模具制造技术,模块式冲压的突出优点在于能把冲压加工系统的柔性与高效生产有机的结合在一起。
柔性的含义较广, 如冲压件的几何形状的多种要求, 只要通过自由编程就可获得, 体现了加工形状的柔性。
又如既适用大批量单品种冲压件的生产, 更对小批量多品种加工发挥也表性。
概括而言,模块式冲压的持点是:(1)在冲压成形过程中可快速更換组合模具以提高生产效率,(2)由于具有带材的供带和矫带装置, 可省却另设上料下料工序,(3 )实现了大工件的不停机加工;(4)既能独立又能成系列的控制组合冲模动作, 能连续进行冲压加工;( 5)冲模具有可编和的柔性特点。
一种模块式冲压加工系统由一台带有控制功能模块式冷冲压的压力机、卷材带材送进装轩、带材矫正机及可编程进给装置等构成。
这种冲压系统在运行时可进行冲模横向位移、带材进给定位、冲模重复运行及自动调整下工步的冲模调整等多项功能。
由于在冲压过程中进行可编程冲压, 使这种模块式冲压系统能柔性地适应生产需求,能在相同带材上进行曲不同工件及批次的混合生产,实现不停机的串接式加工,还同时在工件西面冲压加工,极大地提高了工作效率,有资料表明, 模块式冲压成形使加工费用能下降至40%-50%。
当前模块式冲压装置的集成度是很高的, 在宽度为300MM 尺寸范围内可安排达35个模具, 通过冲模上端的顶板可对冲模进行独立式系列控制, 即形成冲模的集成控制。
整个系统的编程可在windows用户界面和菜単下实现,编程涉及模具沿者横向定位納的何服驱动定位,带材的检验矫正及纵向进给定位, 冲模的质量跟踪检验, 冲模的调整及状况监控等多功能。
当冲模重新配置成更換时,这些变化则会被参数并被控制系统所贮存, 以务下次査询和调用。
冲模数据包括有冲头及其组合标记, 冲头组合在模具中的 X、Y坐标位置及模具轴编号等信息。
l毫米冲压是指汽车车身冲压件的精度控制在0-1. OMM 的范围内,与过去制造业通行的误差2MM 相比, 是个非常大的提高。
汽车结构的轻量化设计措施分析
AUTOMOBILE DESIGN | 汽车设计时代汽车 汽车结构的轻量化设计措施分析贾朝贝郑州科技学院 河南省郑州市 450000摘 要： 汽车工业要发展，在目前必须要满足环保要求，汽车轻量化设计可实现节能减排，但轻量化设计不是单纯减重，而是要保证安全性能的前提下去减重，因而如何进行轻量化设计值得探索，本文中重点对此进行了分析讨论，探析了目前市面上主流的轻量化设计方法措施，仅供参考。
关键词：汽车 轻量化设计 方法措施轻量化在当前汽车设计制造产业当中是一个比较主流的方向，与新能源车具有相当的地位，在传统发动机技术发展陷入瓶颈，新能源汽车受限于电池的情况下，轻量化成为了一种非常关键的解决手段，通过轻量化来实现节能减排。
但汽车轻量化，不是单纯减轻汽车的重量，而是在减轻重量的同时提升性能，因此分析讨论如何去进行轻量化设计，具有非常典型的价值意义。
1 轻量化设计概述1.1 轻量化产生背景轻量化设计是目前国内外汽车设计制造技术中的主要发展方向之一，与环保和安全具有同等地位，随着人们环保意识增强，汽车工业要发展，必须要走可持续发展道路，而可持续发展显然必须要实现节约资源、减少消耗，对于汽车工业而言，要达到相关要求，已经得到公认的路径包括提高发动机效率、新能源和轻量化。
汽车的节能环保通常情况下是降低油耗或提高燃油效率，降低或者清洁排放尾气。
在提高发动机效率方面，由于传统发动机不管是柴油机还是汽油机，实际上都已经达到了一个相当高的水准，现阶段主要是通过对发动机进行微量调整并利用汽车电子技术来提高发动机的效率，但效果并不是很理想，仅仅只能说达标。
而新能源汽车在环保上的效果最佳，但是问题在于由于电池的限制，新能源车的发展还需要走很长的一段路，而轻量化技术，在保证汽车安全性的基础上去降低汽车的自重来实现能耗的下降，它可以作为提高发动机能效，甚至是新能源车能效的一种基础技术手段，在当前发动机技术、新能源车技术尚未出现巨大突破之前，轻量化将是节能减排的主流技术手段。
汽车轻量化英语
汽车轻量化英语
汽车轻量化是指通过采用轻量化材料和设计优化等手段，减轻汽车总重量，提高燃油效率和节能减排。
以下是一些关于汽车轻量化的英语词汇和短语：
1. Lightweight materials: 轻量化材料
2. Carbon fiber: 碳纤维
3. Aluminum alloy: 铝合金
4. High-strength steel: 高强度钢
5. Design optimization: 设计优化
6. Weight reduction: 减重
7. Fuel efficiency: 燃油效率
8. Energy-saving and emission-reducing: 节能减排
9. Composite materials: 复合材料
10. Aerodynamics: 空气动力学
11. Intelligent manufacturing: 智能制造
12. Vehicle safety: 车辆安全
13. Cost-effective: 成本效益
如果您想更深入了解汽车轻量化相关的英语词汇和短语，可以通过阅读相关的英文资料和文献来进一步提高自己的英语水平。
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以轻量化为目标的汽车车身结构优化方法综述
2.2 尺寸优化 尺寸优化是在结构参数、材料分布确定的前提下，对各桁 架结构寻找梁最合适的横截面积、几何尺寸，使得车身质量最 小且满足刚度等要求的优化方法。相对来说，尺寸优化建立数 学 模型 较容 易，计 算简 单，在 实际 工 程 中可 以 较 快取 得 最优 解[9]。也可以说，尺寸优化是拓扑优化的进一步完善和发展。 2.2.1 尺寸优化的数学模型 尺寸优化以车身质量最小为目标，几何尺寸为设计变量， 刚度以及各变量尺寸限制作为约束条件。 2.2.2 尺寸优化的基本步骤和实例 利用有限元分析划分单元，再进行灵敏度分析，排除不参 与优化的单元。为了减少计算量，通常采用近似模型，然后对 近似模型进行求解。刘开勇[10]利用超拉丁实验设计方法，采集 车身的刚度和模态数据，在此基础上建立一阶响应面模型。潘 锋[11]通过建立组合近似模型，减少优化过程的计算量，提高效 率。 常用的近似模型有响应面模型、人工神经网络、径向基函 数模型、kriging 和支持向量回归模型等[10][12]。通过对一阶近似 模型进行分析，计算不同的权系数并进行加权叠加构成的组 合模型在满足模态和刚度要求的前提下，又兼顾了汽车碰撞 安全性、NVH 和疲劳等性能影响，且精度更高，因此组合近似 模型在多目标多学科优化方面更胜一筹。 张伟[13]等采用遗传算法，结合拓扑优化和车身尺寸优化，不 仅将质量降低 35%，而且使刚度提高了 80%以上。康元春等[14]采 用 DOE 及极差分析和方差分析，确定车身骨架梁截面最优尺 寸方案，使车身骨架质量减轻了 123.5kg。
汽车车身结构的轻量化设计
汽车车身结构的轻量化设计随着人们对环保意识的日益增强，汽车工业不得不针对汽车的能源效率、油耗和减排提出更多的要求。
然而，想要提高汽车的能效性，降低油耗和减少排放，就需要解决汽车车身结构过于笨重的问题。
为此，越来越多的车厂开始在汽车车身结构上采取轻量化的设计，以达到更高的油耗效益和碳排放的减少。
1.概述汽车的轻量化设计是指通过采用更轻的材料、更有效的结构和设计、更先进的制造技术等方式来减轻汽车的整车重量，提高汽车的耐用性和性价比。
一般来说，汽车的轻量化设计可以分为三个方面：车身材料的优化选择、车身结构设计的优化及加工工艺的优化。
目前，轻量化的设计已成为汽车工业的一个发展趋势，并在汽车性能、油耗以及碳排放等方面带来了巨大的改进。
2.轻量化材料的选择在轻量化设计中，材料的选择非常重要。
据调查，汽车的车身重量中有70%来自于铁和钢这两种材料，而这些材料也是目前唯一能够满足汽车的强度和安全性要求的材料。
因此，为了实现轻量化的设计，厂商可以从以下方面考虑：（1）铝合金：相比于铁和钢，铝合金密度更低，具有一定的强度和硬度，耐腐蚀性能良好，成本较高，但是可以提高车辆燃油效率及减少碳排放。
（2）碳纤维：碳纤维是当今车身轻量化的理想材料，密度仅仅只有铁和钢的1/5，而且具有很高的强度和拉伸强度。
但是碳纤维容易受潮湿及高温影响,而且成本非常高，所以在实际应用中用的较少。
（3）镁合金：镁合金是一种相对轻质的金属材料，密度比铝合金更轻，力学性能也很好，而且还具有良好的热传导和电导率。
不过，镁合金的腐蚀性也比较强，制造成本较高，所以仅在部分车型上应用。
3.车身结构设计的优化除了材料的选择，车身结构的优化设计也是轻量化设计的重要方面。
通常，车厂可以采取以下设计措施：（1）钢材件结构优化：对车身的各个零件加以精简或是部分区域的厚度减薄，将车身零件的功能和强度保持不变，同时将车身重量降低，否则加强，可以使用HSS及UHSS材料。
汽车车辆类悬架和框架设计外文文献翻译
附录AFormula SAE is a student competition sponsored by Society of Automotive Engineers (SAE), were students design, build, and compete with a small formula style race car. The basis of the competition is that a fictitious company has contracted a group of engineers to build a small formula car. Since the car is intended for the weekend autocross racer, the company has set a maximum price of $8,500. The race car is also limited to a single 610cc displacement engine with a single inlet restrictor. Other major rules require that the car must have a suspension system with a minimum wheel travel of 50mm and a wheelbase greater than 1524mm. The remainder of the rules define safety requirements such as side impact protection .The competition is separated into static and dynamic events. The static events include the cost nalysis, sales presentation, and engineering design. The dynamic portions of the competition are the 15.25 m diameter skid-pad, 91.44 m acceleration event, 0.8 km autocross, 44 km endurance race, and fuel economy.The FSAE competition has been established to provide an educational experience for college students that is analogous to the type of projects they will face in the work force. To participate in FSAE, student groups work with a project from the abstract design until it is completed. The aspects of engineering design, team work, project management, and finance have been incorporated into the basic rules of Formula SAE.This paper is intended to cover some of the basic concepts of suspension and frame design and also highlights the approach UM-Rolla used when designing their 1996 suspension and frame. The suspension section addresses the basic design parameters and presents specific examples. Next, the frame section discusses how to achieve a compromise with the FSAE design constraints. Finally, the design section gives a brief overview of the design methodology used by UM-Rolla for the 1996 race car.1.Suspension GeometryFSAE suspensions operate in a narrow realm of vehicle dynamics mainly due to the limited cornering speeds which are governed by the racetrack size. Therefore, FSAE suspension design should focus on the constraints of the competition. Forexample, vehicle track width and wheelbase are factors governing the success of the car's handling characteristics. These two dimensions not only influence weight transfer, but they also affect the turning radius.Not only do the kinematics have to be considered for FSAE suspension, but the components must also be reasonably priced for the cost analysis and marketable for the sales presentation. For example, inboard suspension could be a more marketable design, while outboard suspension might cost less and be easier to manufacture.The suspension geometry section concentrates on some of the basic areas of suspension design and highlights what the UM-Rolla design team selected for their 1996 race car suspension geometry. UM-Rolla chose to use a four wheel independent suspension system with push rod actuated inboard coil over shocks. This decision was mainly because of packaging constraints. Furthermore, the appearance of inboard suspension was considered important for both the design judging and the sales presentation because of its similarity to modern race cars.Also, this section of the paper was written with short-long arm suspension systems in mind. However, many of the concepts are valid for other suspension types.2.Track Width and WheelbaseThe definition of track width is the distance between the right and left wheel centerlines which is illustrated in Figure 1. This dimension is important for cornering since it resists theFigure 1. Track Widthoverturning moment due to the inertia force at the center of gravity (CG) and the lateral force at the tires . For the designer, track width is important since it is oneUpper Ball Joint Track Low Ball Joint Upper Control UpperControlcomponent that affects the amount of lateral weight transfer . Also, the designers must know the track width before kinematic analysis of the suspension geometry can begin.When selecting the track width, the front and rear track widths do not necessarily have to be the same. For example, track width is typically wider in the front for a rear wheel drive race car. This design concept is used to increase rear traction during corner exit by reducing the amount of body roll resisted by the rear tires relative to the front tires. Based on the corner speeds and horsepower to weight ratio of FSAE cars, this concept should be considered by the designer.The wheelbase also needs to be determined. Wheelbase is defined as the distance between the front and rear axle centerlines, and also influences weight transfer, but in the longitudinal direction. Except for anti-dive and anti-squat characteristics, the wheelbase relative to the CG location does not have a large effect on the kinematics of the suspension system. However, the wheelbase should be determined early in the design process since the wheelbase has a large influence on the packaging of components.For track width and wheelbase starting points, the designers should research the opposition's dimensions to serve as a baseline for their own calculations. FSAE car specifications for the competing teams, including track width and wheelbase, are available in the event program published by SAE.The 1996 design team selected a 1727 mm wheelbase, 1270 mm front track width, and a 1219 mm rear track width, which were based on previous UM-Rolla cars. Although this wheelbase was adequate for the FSAE competition size courses, the UM-Rolla design team has decided to increase the wheelbase for the next car to 1854.2 mm. This increase in wheelbase is an attempt to improve stability for high speed corner entry at the competition.3.Tire and WheelAfter track width and wheelbase considerations have been addressed, tire and wheel selection is the next step in the design process. Since the tire is important to the handling of the vehicle, the design team should thoroughly investigate the tire sizes and compounds available. The tire size is important at this stage of the design since the height of the tire must be known before the geometry can be determined. For example, the tire height for a given wheel diameter determines how close the lower ball joint can be to the ground if packaged inside the wheel.Tire Size - The designers should be aware that the number of tire sizes offeredfor a given wheel diameter is limited. Therefore, considering the importance of the tire to handling, the tire selection process should be a methodical process. Since the amount of tire on the ground has a large influence on grip, it is sometimes desirable to use wide tires for increased traction. However, it is important to remember that wide tires add rotating mass which must be accelerated by a restricted FSAE engine. This added mass might be more detrimental to the overall performance than the increase in traction from the wider tires. Not only does a wider tire add mass, but it also increases the amount of rubber that must be heated. Since racing tires are designed to operate most efficiently in a specific temperature range, this added material may prevent the tires from reaching the optimum temperature range . The UM-Rolla team used tires for the 1996 competition that were designed to work most efficiently at a minimum of 71°.During the selection process the designers must consider how the tires will influence the performance of the entire package. For example, the weather conditions for the FSAE dynamic events might determine which tire compound and tire size should be used for the competition. Another important consideration is the price of the tires since the cost can be a large portion of a team's budget.For the 1996 competition, UM-Rolla selected a 20 by 6-13 racing tire for both the front and rear of the car. Because of the low vehicle mass, a narrow tire was selected so tire temperatures would be greater than previous UM-Rolla designs. This tire selection increased the operating temperature from 48o to 60oC. For the competition, the weather was predicted to be cool, so the team brought a set of hard and soft compound tires. The team chose to use the harder compound since the weather for the endurance was predicted to be clear and warm.Wheel Selection - Once a decision has been made as to which tire sizes to use, the wheel selection should be next. Usually, the wheel dimensions are fixed and allow for little modification. Therefore, it is important to have some design goals in mind before investing in wheels. Generally, the upright, brake caliper, and rotor are placed inside the wheel which requires wheel offset for clearance. It is usually easier to design the suspension geometry if the wheel profile is known. For example, the ball joint location is limited to the area defined by the wheel profile. Some packaging constraints are shown in Figure 2.Other considerations for wheel selection include: cost, availability, bolt circle, and weight. For example, three-piece rims, althoughexpensive, have the distinct advantage of offering many offsets and profiles that can be changed during the design process .Figure 2. 1996 Front SuspensionUM-Rolla designed the 1996 suspension geometry around a wheel profile from a previous car and then acquired a set of three-piece rims to meet the design specifications. All four wheels selected for the 1996 competition were size 6 by 13. This wheel selection allowed for tire rotations, reduced cost, and a wide selection of tire sizes, compounds, and manufacturers.4.GeometryThe designer can now set some desired parameters for the suspension system. These usually include camber gain, roll center placement, and scrub radius. The choice of these parameters should be based on how the vehicle is expected to perform. By visualizing the attitude of the car in a corner, the suspension can be designed to keep as much tire on the ground as possible. For example, the body roll and suspension travel on the skid pad determines, to a certain extent, how much camber gain is required for optimum cornering. The amount of chassis roll can be determined from roll stiffness while the amount of suspension travel is a function of weight transfer and wheel rates.Once a decision has been made about these basic parameters, the suspension must be modeled to obtain the desired effects. Before the modeling can begin, the ball joint locations, inner control arm pivot points, and track width must be known.The easiest way to model the geometry is with a kinematics computer program since the point locations can be easily modified for immediate inspection of their influence on the geometry. Should a dedicated kinematics computer program not be available, then a CAD program can be used simply by redrawing the suspension as the points are moved.When designing the geometry, it is important to keep in mind that designing is an iterative process and that compromises will be inevitable. For instance, the desired scrub radius might not be possible because of packaging constraints. When modelingthe suspension, the designers should not aimlessly modify points without first thinking through the results. For example, the designer should visualize how the wheel will camber relative to the chassis when making the lower A-arm four times longer than the upper A-arm. One method that can be used to visualize the results is the instant center location for the wheel relative to the chassis. Another method is to use the arcs that the ball joints circumscribe relative to the chassis. For a complete explanation for determining suspension point locations from instant center locations refer to Milliken .Scrub Radius, KPI, and Caster - The scrub radius, or kingpin offset, is the distance between the centerline of the wheel and the intersection of the line defined by the ball joints, or the steering axis, with the ground plane which is illustrated in Figure 2. Scrub radius is considered positive when the steering axis intersects the ground to the inside of the wheel centerline. The amount of scrub radius should be kept small since it can cause excessive steering forces . However, some positive scrub radius is desirable since it will provide feedback through the steering wheel for the driver .Kingpin inclination (KPI) is viewed from the front of the vehicle and is the angle between the steering axis and the wheel centerline . To reduce scrub radius, KPI can be incorporated into the suspension design if packaging of the ball joints near the centerline of the wheel is not feasible. Scrub radius can be reduced with KPI by designing the steering axis so that it will intersect the ground plane closer to the wheel centerline. The drawback of excessive KPI, however, is that the outside wheel, when turned, cambers positively thereby pulling part of the tire off of the ground. However, static camber or positive caster can be used to counteract the positive camber gain associated with KPI.Caster is the angle of the steering axis when viewed from the side of the car and is considered positive when the steering axis is tilted towards the rear of the vehicle . With positive caster, the outside wheel in a corner will camber negatively thereby helping to offset the positive camber associated with KPI and body roll. Caster also causes the wheels to rise or fall as the wheel rotates about the steering axis which transfers weight diagonally across the chassis . Caster angle is also beneficial since it will provide feedback to the driver about cornering forces .UM-Rolla's suspension design team chose a scrub radius of 9.5 mm, zero degrees of KPI, and 4 degrees of caster. This design required the ball joints to be placed near the centerline of the wheel,which required numerous clearance checks in the solid modeling program.Roll Center - Once the basic parameters have been determined, the kinematics of the system can be resolved. Kinematic analysis includes instant center analysis for both sets of the wheels relative to the chassis and also for the chassis relative to the ground as shown in Figure 3. The points labeled IC are the instant centers for the wheels relative to the chassis. The other instant center in Figure 3, the roll center, is the point that the chassis pivots about relative to the ground . The front and rear roll centers define an axis that the chassis will pivot around during cornering. Since the CG is above the roll axis for most race cars, the inertia force associated with cornering creates a torque about the roll center. This torque causes the chassis to roll towards the outside of the corner. Ideally, the amount of chassis roll would be small so that the springs and anti-roll bars used could be a low rate for added tire compliance . However, for a small overturning moment, the CG must be close to the roll axis. This would indicate that the roll center would have to be relatively high to be near the CG. Unfortunately, if the roll center is anywhere above or below the ground plane, a "jacking" force will be applied to the chassis during cornering forces . For example, if the roll center is above ground, this "jacking" force causes the suspension to drop relative to the chassis. Suspension droop is usually undesirable since, depending on the suspension design, it can cause positive camber which can reduce the amount of tire on the ground. Conversely, if the roll center is below the ground plane, the suspension goes into bump, or raises relative to the chassis, when lateral forces are applied to the tires. Therefore, it is more desirable to have the roll center close to the ground plane to reduce the amount of chassis vertical movement due to lateral forces .Figure 3. Front Roll CenterSince the roll center is an instant center, it is important to remember that the roll center will move with suspension travel. Therefore, the design team must check the migration of the roll center to ensure that the "jacking" forces and overturning moments follow a relatively linear path for predictable handling . For example, if the roll center crosses the ground plane for any reason during cornering, then the wheels will raise or drop relative to the chassis which might cause inconstant handling.Theroll center is 35.6 mm below ground in the front and 35.6 mm above ground in the rear for UM-Rolla's 1996 car. Since none of the previous UM-Rolla cars had below ground roll centers, the selection of the 1996 points was basically a test to understand how the below ground roll center affected the handling. Because of the large roll moment, the team designed enough camber gain into the suspension to compensate for body roll associated with soft springs and no anti-roll bar. The team was very happy with the handling but decided, for the next car, to have both roll centers above ground for a direct comparison between both designs.Camber - Camber is the angle of the wheel plane from the vertical and is considered to be a negative angle when the top of the wheel is tilted towards the centerline of the vehicle. Camber is adjusted by tilting the steering axis from the vertical which is usually done by adjusting the ball joint locations. Because the amount of tire on the ground is affected by camber angle, camber should be easily adjustable so that the suspension can be tuned for maximum cornering. For example, the amount of camber needed for the small skid pad might not be the same for the sweeping corners in the endurance event.The maximum cornering force the tire can produce will occur at some negative camber angle . However, camber angle can change as the wheel moves through suspension travel and as the wheel turns about the steering axis. Because of this change, the suspension system must be designed to compensate or complement the camber angle change associated with chassis and wheel movements so that maximum cornering forces are produced.The amount of camber compensation or gain for vertical wheel movement is determined by the control arm configuration. Camber gain is usually obtained by having different length upper and lower control arms. By using different length control arms, the ball joints will move through different arcs relative to the chassis. The angle of the control arms relative to each other also influence the amount of camber gain. Because camber gain is a function of link geometry, the amount of gain does not have to be the same for both droop and bump. For example, the suspension design might require the wheels to camber one degree per 25mm of droop versus negative two degrees per 25mm of bump.Static camber can be added to compensate for body roll, however, the added camber might be detrimental to other aspects of handling. For example, too much static camber can reduce the amount of tire on the ground, thereby affecting straightline braking and accelerating. Similarly, too much camber gain during suspension travel can cause part of the tire to loose contact with the ground.Caster angle also adds to the overall camber gain when the wheels are turned. For positive caster, the outside wheel in a turn will camber negatively, while the inside wheel cambers positively. The amount of camber gain caused by caster is minimal if the wheels only turn a few degrees. However, FSAE cars can use caster angle to increase the camber gain for the tight corners at the FSAE competition.UM-Rolla designed for a relatively large amount of camber gain since anti roll bars were not used in the 1996 suspension design. The use of low wheel rates with a large roll moment required the suspension to compensate for the positive camber induced by chassis roll and suspension travel. The camber gain for UM-Rolla's 1996 car was from both the caster angle and the control arm configuration.5.Steering SystemThe steering geometry has a large influence on the handling characteristics of the vehicle. For example, if the system is not properly designed, then the wheels can unexpectedly toe in or out during suspension travel. This toe change is referred to as bump steer which is described in detail in both references . Bump steer is basically undesirable since the car changes direction when the driver does not expect the change .Ackermann steering must also be considered during the design process. Ackermann steering occurs when the outside wheel turns less than the inside wheel. This is possible since the amount of steering angle for each wheel is determined by the steering geometry. Reverse or anti-Ackermann occurs when the outside wheel turns more than the inside wheel during cornering .During a turn, the inside wheel travels around a smaller geometric radius than the outside wheel. Ackermann steering can be used so that the wheels travel about their corresponding radii, theoretically, eliminating tire scrub. However, designing for precise Ackermann steering might not provide the best handling since tire slip angles influence the actual turning radius . The designer must decide, based on the requirements, if the steering system design will include Ackermann geometry.UM-Rolla placed the rack and pinion in front of the axle centerline near the lower control arms because of packaging constraints. This placement required extra room in the frame design since the driver had to straddle the steering column. Afterbuilding a test car that was hard to steer because of a half a turn lock to lock system, the 1996 steering system was designed to be one turn lock to lock. This was accomplished by changing the rack and pinion ratio instead of increasing the steering arm length because of packaging constraints. The system specifications for the 1996 car are: 76mm steering arms, 250mm diameter steering wheel, and 51mm of rack travel per one pinion revolution. These specifications were retained for the next race car design because of the handling characteristics. The 1996 UM-Rolla design has a small amount of anti-Ackermann due to packaging.FSAE suspension not only has to be competitive on the racetrack, but the suspension must also perform well in the static events. For the dynamic events, the designers should concentrate on the geometry so that most of the tire will stay in contact with the ground for all normal driving situations: braking, accelerating, and cornering. The suspension system must also be designed so that it is easy to manufacture and is reasonably priced for the cost analysis. To reduce the cost and complexity of the 1996 race car, UM-Rolla designed the system so that the wheels, hubs, and bearings were the same for each corner of the car.Designing the suspension geometry is only a small part of building a vehicle. A well engineered suspension system does not automatically make a fast race car. Although this paper has concentrated on the design aspect, development is just as important to the success of the package. Because the design process must take place within a given time constraint, the first suspension design might not provide the best handling. It is not uncommon to make design changes after the car is completed. It is more important for FSAE teams to compromise the overall design so that the car can be completed and tested prior to competition.6.FrameThe purpose of the frame is to rigidly connect the front and rear suspension while providing attachment points for the different systems of the car . Relative motion between the front and rear suspension attachment points can cause inconsistent handling . The frame must also provide attachment points which are not going to yield within the car's performance envelope.There are many different styles of frames; space frame, monocoque, and ladder are examples of race car frames. The most popular style for FSAE is the tubular space frame. Space frames are a series of tubes which are joined together to form a structurethat connects all of the necessary components together. However, most of the concepts and theories can be applied to other chassis designs.Figure 4. UM-Rolla's 1996 Frame Design7.StiffnessThe suspension is designed with the goal of keeping all four tires flat on the ground throughout the performance range of the vehicle. Generally, suspension systems are designed under the assumption that the frame is a rigid body. For example, undesirable changes in camber and toe can occur if the frame lacks stiffness. Superimposed images of a frame subjected to a torsional load and an undeflected frame and can be seen in Figure 5.Figure 5. Chassis DeflectionUM-Rolla has found that in most cases, a stiff chassis will not have a problem with yielding. However, some care should be taken to ensure that the attachment points of the frame do not yield when subjected to design loads. For example, the engine mounts should be made stiff enough to reduce the possibility of failure.Torsional Stiffness - Torsional stiffness is the resistance of the frame to torsional loads . UM-Rolla used FEA to analyze the torsional stiffness of the 1996 chassis. The solution of the simple rod and beam element model for the frame was roughly 2200 foot pounds per degree of deflection. The 1996 frame weighed approximately 27kg, which UM-Rolla believes is heavier than needed for a two day racing series. However, some extra structure was added to the frame to increase its safety. Also, the drivetrainmounts were significantly strengthened so that the car would be able to serve as a driver training tool for several semesters.As the 1996 frame evolved, the stiffness to weight ratios of different designs were compared. A chassis can be made extremely stiff by adding significant amounts of material to the frame. However, this additional material might degrade the performance of the car because of the added mass.Obviously, torsional rigidity is not the only measurement for analyzing the stiffness of a chassis. Bending stiffness can also be used to analyze the efficiency of a frame design. However, bending stiffness is not as important as torsional stiffness because deflection due to bending will not affect wheel loads . Because the design time is severely limited in FSAE, UM-Rolla has found that a torsional analysis is adequate to determine the relative stiffness of different frame designs.Triangulation - Triangulation can be used to increase the torsional stiffness of a frame, since a triangle is the simplest form which is always a structure and not a mechanism. Obviously, a frame which is a structure will be torsionally stiffer than a mechanism . Therefore, an effort should be made to triangulate the chassis as much as possible.Visualizing the frame as a collection of rods which are connected by pin joints can help frame designers locate the mechanisms in a design . Designers can also evaluate their frame by checking to see if each pin jointed node contains at least three rods which complement the load path.UM-Rolla chose to use thin wall steel tubing for the 1996 frame design. This required significant triangulation of the frame, since thin wall tubing performs very well in tension and compression but poorly in bending. The components which produce significant amounts of force, for example the engine and suspension, were attached to the frame at a triangulated point.Figure 6. Frame Triangulation(Frame, Side View)Previous UM-Rolla frames have lacked adequate triangulation for highly loaded components. These components were attached to the frame with load bearing tabs which were welded at the midpoint of a single section of tubing. As expected, thistube bent like a simply supported beam and caused unwanted movement of the attached component. Although these designs worked for the duration of the competition, they invariably failed by fracturing the tube or breaking the tab. For the 1996 car, all of the highly loaded components were attached to triangulated points.Area Moment of Inertia - The area moment of inertia has a large influence on the stiffness of a structure. Therefore, the farther material is from the axis of twist the stiffer the frame will be in bending and torsion. This concept is implemented by adding structural side pods to the basic frame.Figure 7. Structural Sidepods(Frame Top View)Figure 7 shows the triangulated side pods which were used to increase the torsional rigidity of the 1996 frame. This material also increased the side impact protection. The sidepods add structure as far from the centerline of the chassis as possible which increases the area moment of inertia between the front and rear suspensions. Most of the successful FSAE cars have structural side pods for safety and increased torsional stiffness.In addition to using the sidepods to increase the stiffness of the chassis, UM-Rolla's 1996 entry used the roll hoop and down tubes to increase the rigidity of the frame. The 1997 FSAE rules state that the tubes from the top of the roll hoops to the base of the frame have to be 0.049" wall when fabricated from 4130 steel . Because these tubes are stiffer than 0.035" wall tubing, the frame stiffness can be substantially increased by properly placing the roll hoop tubes.8.Load PathDuring the design process, it is important to consider how loads are passed into the frame. A "Load Path" describes the path through which forces are dissipated into the frame. For example, Figure 8 shows how the vertical load generated by the weight on the wheel will travel through the upright, push rod, rocker, coil-over shock and into the structure of the frame. Of course, to properly investigate the forces involved, a freebody diagram for each component would have to be drawn. Nevertheless, this。
传统车身与现代车身设计方法外文文献翻译、中英文翻译、外文翻译
外文文献及译文文献题目：Traditional car body and modern car body design method 文献来源： Taylor&Francis文献发表日期： 2011年06月17日学生姓名：学号：系别：专业：指导教师：职称：2017年5月19日Traditional car body and modern car body design method1、Traditional car body design methodThe autocar car body shape is one of the three greatest fractions that are paralleled with bedrock and engine in the autocar structure, its development and capacity preparation cycle is the longest, and the preparation work of chart paper and craft is the biggest, and it is usually still need to remodel, unlike bedrock and engine attain which is easy to reach serialization and generalization.The characteristics of car body structure lies on the constitution of car body shape of each null parties be big for size more but shape complicated space curved face(namely so-called person type overlay piece), these space curved faces can not be used ordinary machinery drawing method will it the complete earth's surface appear, as a result have to create stereoscopic pattern as basis. For making these chart paper and change a type tangibly express shape and structure of car body, need to pass a complicated design process to complete.In the traditional car body design, the design information and the data of car body deliver mainly by two-dimensional chart paper and principal mode type, there are many design fraction of person's hand work. The defects that traditional design method can not overcome embody at as follows a few aspects:1）、The manpower resource consume greatly. The traditional car body designs a tapping requirement the art designing personnel, engineering the technical personnel and its operative make a concerted effort, such as the creation and conservancy, molding tool of the two-dimensional engineering chart (the plot of car body graph), soil pattern design, production, grind to go together with and adjust to try, capacity prepare medium process planning etc.s will expend a person quantitative manpower and resource.2）、Designing accuracy is low, the cycle is long. Accuracy low and main reason's lying in the information of each link that designs with born reserve to deliver is a kind of “move a shape", For example from lord draftion board manufacture principal mode Xing, is carried on by the principal mode type to process craft supplement, productioncraft pattern, from convex of the craft pattern turn over into concave craft pattern, again is depended by the craft pattern anti to process trimming die, after the convert of these links, raw data's the error of various factitiousness unavoidable, cause the molding tool accuracy of processing's canning not assure, only depend next move of the hand work grind to go together with to clinch.3）、Car body already set over speedy and the incapability carry on proceed together design. Traditional car body's designing is basically a kind of one-way anti reversible design status, product once already set, modification or remodel very difficult, the universal turns, the series turn a degree low.In addition, the car body of the modern autocar null parts because of the requirement of structure and beauty, the space curved face gradually complicates, traditional design method F very difficult satisfy design request.2、Modern car body design methodAs design and modification of autocar car body shape request gradually increasing higher requirements, the parameter designment match more and come closer to the conceptual design in modern CAD and proceed together design thought, and can develop the flexibility of design and analysis, therefore the parameter of the car body …craft shape design have already become a inevitable trend of vehicle production. The traditional hand work shape method has already been replaced by Computer aided modeling method which is base on various of advanced modeling soft wares, Parametric design technology can quickly absorb the foreign autocar car body designs of advanced technique, quickly respond to a market, upgrading the capability of independence design and improving autocar product competitive power in our country. Since 70's in last century, along with the increasing development of computer aided geometric design and Computer graphics, the car body designs process mid for decide or the whole distances lead into calculator supporting system (CAD/CAM software), the garbage se ts up car body 3D digitize pattern, with” hat you see is what you get of" alternate mode the perfect design scheme is the main features of modern car body design method. The basic process is following: After investigating the market, forming a car type of whole request; Manufacture hand-drawing effect chart, alsoavailable calculator assistance the software draw; The creation shrinks ratio and principal mode type; Set up car body figure pattern by the mode garbage of three coordinates scanning or laser scanning; Carry on a structure design, set up A class pattern; Born NC code produces a kind car. Draw an effect chart step, can hand work plot also the available calculator assistance software plot, concrete according to personal condition of designer and then settle, but use calculator plot can fasterly set up a car body figure pattern and make to modify an operate more convenient. But the use hand work plot rework make to shrink scale model with 1:a , again scan by three coordinateness or the laser scan of the mode garbage set up car body figure pattern, this is inner common use kind of method, because manufacture pattern in this way can even keep the style principle that the view earth's surface reaches to a designer. This is actually a process of contrary design. In a brief, the contrary engineering is to point the product pattern conversion that will have been already had to design in the process and apply very extensively in the car body for the process of the pattern of digitize engineering design in the calculator, equal hence 1 kind make copy and have to develop a cycle short, designing the accuracy being higher and easy to progress structure designing/ analyzing waiting merit.The famous soft wares, such as CATIA, IMAGEWARE and UG etc provided the strong reversal the design function. The car body figure pattern cans share after creating and is advantageous to each segment realize to proceed together an operate and thus and consumedly shorten to design a cycle, and used for various CAD/CAM systems of car body design all according to 3D solid pattern auto born two-dimensional car body chart paper and used for number to control the code of transform, at the same time when the 3D pattern takes place arbitrarily minuscule transform, the systems will of oneself modify with of related chart the paper and number control a code. Use a far-ranging calculator to lend support to software currently CATIA is the most famous, it is a CAD/CAM/CAE software that is reached Suo company and IBM by France to unite tapping, and the function is extremely strong. CATIA provides to have perfect CLASSA (A class curved face, this denomination also from reach Suo to propose first) to set up function. CLASS the Acurved face is in a brief a car body shape, instrument board, inside decoration piece etc. higher to curved face quality requirements type of curved face of parts, namely and only agreeable requests higher type of curved face, since wanting a surface is smooth and links to transfer smooth satisfy aerodynamics principle with the significant decrease air resistance, and then want to reserve the beauty of shape. Here point of A class curved face optics follow the sexual notion can in brief define from mathematics for: curve two ranks are several what continuous, have no redundancy to turn a dot, curvature change balance, the strain energy is smaller. Is a G2 at least sequential(curvature continuous) to curved face, also namely two curved faces follow public tie line to be placed in all directions to have common law curvature. While putting together connecting, car body shape different curved face slab's continuousness have smooth request to transfer, satisfy G2 at least, even G3 continuous. The CLASS A curved face describes with the B kind method, the nodal point vector adopts tired the long parameter applying string turn a law; In the U, curved face's V direction last time count at 3-7 times of, the maximum was 9 times no bigger than. In the actual design can from as follows a few aspects to evaluate A class curved face mass: the continuity, curved face whole cooperativeness, curved face of the inner mass, curved face of the design accuracy, curved face of curved face satisfy car body structure the design and car body make a request. Positive because garbage energy such convenience, efficiently sets up CLASS A curved face, it just lets car body solid the shape is accurate day by day and quick, design a personnel construct more Be approached to reality, curved face mass higher car body 3D pattern. Garbage not only can vs. car body external appearance and inside decoration create figure pattern, and can vs. engine, bedrock etc. other null parts create pattern, and directly carry on a limited kyat analysis, structure design/analysis, even conjecture assembly and virtual air opening test etc., making to design a personnel can set up virtual electron kind with the garbage car and carry on trial, can in advance discover the existent question in the design before physically producing, raised an efficiency, lowered cost. Usually each kit pattern sort is depositing in a large data bank, hereafter can go direct to match in the data bank while improving vs. the car type similar of nullparts, modify its related parameter then meet demand, have never needed to re- design, be advantageous to the series of car type to turn very much, produce company should to quick change of the capability of the market Be also strong to get poly.3、Domestic and international search present conditionCompare bigger autocar companies to widespread adopt the two-dimensional design of CAD system progress car body in the world, like the general arrangement(person machine engineering) of car body, inside the curved face structure of the outside overlay piece design, null parts structure and assembly design, molding tool design, and automatically born related design technique document. CAM/CAE is to apply to the CIMS technique also very universal. Each big autocar manufacturing companies all own their huge car body tapping of oneself troops, and don't hesitate to throw in the laboratory that the mint capital create a forerunner, develop or usher in the technical software progress car body design.Designing applied more earth of realm in the car body can the software mainly has the United States Auto studio of the ALIAS Company, CATIA of subsidiary Daussault of IBM Company, PRO/Engineer etc. of the UG of EDS Company, parameter Technique Company. These soft wares all belong to in general use machine design software and have stronger curved face design, parameter to turn designed capacity or mix to design function, and can favor the machinery manufacturing whole processes, then design a model →engineering analysis →the transform make, some even can better backing car body design of first step, namely conceptual design(Concept Design) step(is like Auto studio), and put together with model tie through a certain way it, design plane the convert as two-dimensional pattern.The local calculator is used for an autocar design to begin from 70's.The long spring vapor grinds agreeable procedure car body exclusive use function of the curved face optics developed board, created its car body graphics used in design library, developed the autocar car body surface shape and structure design process system, the system had to draw car body soil draftion board and car body part drawing and provide transform principal mode type NC data, manufacture the capability of soil pattern. The autocar motorcycle manufacturing company in Peking develops theBJA-BSM car body CAD system, successfully is applied to BJ124, the BJ125 waits the car type of various new car type operate. Shanghai G.M. and Shanghai together helped together Jie science and technology limited company etc. to design to develop flow process to carry on to a little bit thoroughly study vs. the car body and forming a more perfect car body tapping procedure. But still exist a lot of problems that need settle urgently in the aspects of clicking a data handling, curved face surface optics being agreeable, mathematical models converting process waiting. Is a lot of local win a small autocar business because of the restriction of capital, equipment and personnel, can not carry on the search and application of comparing the system to the car body design method, or only apply some one realm, applied get maximum of also just make vs. the product partial or little yellow of change.传统车身与现代车身设计方法1、传统车身设计方法汽车车身外形是汽车结构中与底盘和发动机并列的三大部分之一，其开发和生产准备周期最长，图纸及工艺准备的工作量最人，并且还经常要改型，不像底盘和发动机那样容易做到系列化、通用化。
优化结构设计外文文献翻译、中英文翻译、外文翻译
JOURNAL OF OPTIMIZATION THEORY AND APPLICATIONS: Vol. 6, No. I. 1970SURVEY PAPEROptimization of Structural Design I.~W. PRAGER 3Abstract. Typical problems of optimal structural design are discussed to indicate mathematical techniques used in this field. An introductory example(Section 2) concerns the design of a beam for prescribed maximal deflection and shows how suitable discretization may lead to a problem of nonlinear programming, in this case, convex programming. The problem of optimal layout of a truss (Section 3) is discussed at some length. A new method of establishing optimality criteria (Section 4) is illustrated by the optimal design of a statically indeterminate beam of segmentwise constant or continuously varying cross section for given deflection under a single concentrated load. Other applications of this method (Section 5) are briefly discussed, and a simple example of multipurpose design (Section 6) concludes the paper.1. IntroductionThe most general problem of structural optimization may be stated as follows: from all structural designs that satisfy certain constraints, select one of minimal cost. Note that this statement does not necessarily define a unique design; there may be several optimal designs of the same minimal cost.Typical design constraints that will be considered in the following specify upper bounds for deformations or stresses, or lower bounds for load-carrying capacity, buckling load, or fundamental natural frequency. Both singlepurpose and multipurpose structures will be considered, that is, structures that are respectively subject to a single design constraint or a multiplicity of constraints.The term cost in the statement of the design objective may refer to the manufacturing cost or to the total cost of manufacture and operation over the expected lifetime of the structure. In aerospace structures, the cost of the fuel needed to carry a greater weight frequently overshadows the cost of manufacture to such an extent that minimal weight becomes the sole design objective. This point of view will be adopted in the following.In the first part of this paper, typical problems of optimal design will be discussed to illustrate mathematical techniques that have been used in this field. The second part will be concerned with a promising technique of wide applicability that has been developed recently. Throughout the paper, it will be emphasized that the class of structures within which an optimum is sought must be carefully defined if meaningless solutions are to be avoided. The fact will also be stressed that certain intuitive optimality criteria of great appeal to engineers do not necessarily furnish true optima. For greater clarity in the presentation of design principles, the majority of examples will be concerned with single-prupose structures even though multipurpose structures are of far greater practical importance.2. DiscretizationTo explore the mathematical character of a problem of structural optimization, it is frequently useful to replace the continuous structure by a discrete analog. Consider,for instance, the simply-supported elastic beam in Fig. 1. The maximum deflection produced by the given load 6P is not to exceed a given value δ To discretize the problem, replace the beam by a sequence of rigid rods that are connected by elastic hinges. In Fig. 1, onlyFig. 1. Discrete analog of elastic beam.three hinges have been introduced; but, to furnish realistic results, the discretization would have to use a much greater number of hinges. The bending moment i M transmitted across the ith hinge is supposed to be related to the angle of flexure i θ byi M =i s i θ (1)where i s is the elastic stiffness of the hinge. Since the beam is statically determinate, the bending moments i M at the hinges are independent of the stiffnesses i s ; thus, 1M =5Ph=1s 1θ, 2M =3Ph=22s θ, 3M =Ph=33s θ. (2)In the following, the angles of flexure i θ, will be treated as small. In a designspace with the rectangular Cartesian coordinates i θ, i = 1, 2, 3, the nonnegativecharacter of the angles of flexure and the constraints on the deflections i u at the hinges define the convex feasible domain1θ，2θ，3θ≥0,51θ+32θ+3θ-6δ/h ≤0,31θ+92θ-33θ-6δ/h ≤0, （3）1θ+32θ+53θ-6δ/h ≤0，As will be shown in connection with a later example, the cost (in terms of weight) of providing a certain stiffness may be assumed to be proportional to this stiffness. The design objective thus is 1s +2s +3s =Min or, by (2),5/1θ+3/2θ+1/3θ=Min (4)Note that, for the convex program (3)-(4), a local optimum is necessarily a global optimum. This remark is important because a design that can only be stated to be lighter than all neighboring designs satisfying the constraints is of little practical interest. Note also that the optimum will not, in general, correspond to a point of design space that lies on an edge or coincides with a vertex of the feasible domain. This remark shows that the intuitively appealing concept of competing constraints is not necessarily valid. Suppose, for instance, that a design 1s ,2s ,3s has been found for which 3u <2u <1u =δ. If s ∆ denotes a sufficiently small change of stiffness, the design 1s +s ∆,2s -s ∆,3s , which has the same weight, might then be expected to have deflection 1u ，2u ，3u satisfying 3u <2u ,2u <1u <1u =δ, and all threestiffnesses could be decreased in proportion until the deflection at the first hinge has again the value δ. If this argument were correct, this process of reducing the structural weight could be repeated until the deflections at the hinges 1 and 2 had both the value &. In subsequent design changes, 1s and 2s would be increased by the same small amount while 3s would be decreased by twice this amount to keep the weightconstant. In this way, it might be argued that the optimal design must correspond to a point on an edge or at a vertex of the feasible domain, that is, that, for the optimal design, two or three of the constraining inequalities must be fulfilled as equations. This concept of competing constraints, to which appeal is frequently made in the engineering literature, is obviously not applicable to the problem on hand.Minimum-weight design of beams with inequality constraints on deflectionhas recently been discussed by Haug and Kirmser (Ref. 1). Earlier investigations (see, for instance, Refs. 2-4) involved inequality constraints on the deflection at a specific point, for instance, at the point of application of a concentrated load. In special cases, where the location of the point of maximum deflection is known a priori, for instance, from symmetry considerations, a constraint on the maximum deflection can beformulated in this way. As Barnett (Ref. 3) has pointed out, however, constraining a specific rather than the maximum deflection may lead to paradoxical results. For example, when some loads acting on a horizontal beam are directed downward while others are directed upward, it may be possible to find a design for which thedeflection at the specified point is zero. Since it will remain zero as all stiffnesses are decreased in proportion, the design constraint is compatible with designs of arbitrarily small weight.3. OptimalIn the preceding example, the type and layout of the structure (simply supported, straight beam) were given and only certain local parameters (stiffness values) were atthe choice of the designer. A much more challenging problem arises when type and/or layout must also be chosen optimally.Figure 2a shows the given points of application of loads P and Q that are to be transmitted to the indicated supports by a truss, that is, a structure consisting ofpin-connected bars, the layout of which is to be determined to minimize the structural weight. To simplify the analysis, Dorn, Gomory, and Greenberg (Ref. 5) discretized the problem by restricting the admissible locations of the joints of the truss to the points of a rectangular grid with horizontal spacing l and vertical spacing h (Fig. 2a). Optimization is then found to require the solution of a linear program. The optimal layout dependsFig. 2. Optimal layout of truss according to Dorn, Gomory, and Greenberg (Ref. 5). on the values of the ratios h/l and P/Q. Figures 2b through 2d show optimal layouts for h/l = 1 and P/Q = O, 0.5, and 2.0.For h/l = 1 and a given value of P/Q, the optimal layout is unique except for certain critical values of P/Q, at which the optimal layout changes, for instance, from the form in Fig. 2c to that in Fig. 2d. The next example, however, admits an infinity of optimal layouts that are all associated with the same structural weight.Three forces of the same intensity P, with concurrent lines of action that form angles of 120 ° with each other, have given points of application that form an equilateral triangle (Fig. 3@ A truss that connects these points is to be designed forσis prescribed for the magnitude of the minimal weight, when an upper boundaxial stress in any bar.Figures 3b and 3c show feasible layouts. After the forces in the bars of these statically determinate trusses have been found from equilibrium considerations, theσin cross-sectional areas are determined to furnish an axial stress of magnitudeeach bar.The following argument, which is due to Maxwell (Ref. 6, pp. 175-177), shows that the two designs have the same weight.Imagine that the planes of the trusses are subjected to the same virtual, uniform, planar dilatation that produces the constant unit extension e for all line elements. By the principle of virtual work, the virtual external work e W of the loads P on the virtual displacements of their points of applicationFig. 3. Alternative optimal designs. equals the virtual internal work i W =∑F λof the bar forces F on the virtualelongations ~ of the bars. If cross-sectional area and length of the typical bar are denoted by A and L, then F=0σ A and λ=εL. Thus,i W =0σ∑AL=0σεV （5）where V is the total volume of material used for the bars of the truss. Now, e W depends only on the loads and the virtual displacements of their points of application but is independent of the layout of the bars; therefore, it has the same value for both trusses. If follows from e W =i W and (5) that the two trusses use the same amount of material.If all cross-sectional areas of the two trusses are halved, each of the new trusses will be able to carry loads of the common intensity P/2 without violating the design constraint. Superposition of these trusses in the manner shown in Fig. 3d then results in an alternative truss for the full load intensity P that has the same weight as the trusses in Figs. 3b and 3c.Fig. 4. Alternative solution to problem in Fig. 3a.Figure 4 shows another solution to the problem. The center lines of the heavy edge members are circular arcs. The axial force in each of these members has constantσ. The other bars are magnitude corresponding to the tensile axial stressσand are prismatic, comparatively light. They are also under the tensile axial stressexcept for the bars AO, BO, and CO, which are tapered.The bars that are normal to the curved edge members must be densely packed. If only a finite number is used, as in Fig. 4, and the edge members are made polygonal rather than circular, a slightly higher weight results. This statement, however, ceases to be valid when the weight of the connections between bars (gusset plates and rivets or welds) is taken into account.The interior bars in Fig. 4 may also be replaced by a web of uniform thickness under balanced biaxiat tension. While fully competitive as to weight, this design has, however, been excluded by the unnecessarily narrow formulation of the problem, which called for the design of a truss. In this case, the excluded design does not happen to be lighter than the others. However, unless the class of structures within which an optimum is sought is defined with sufficient breadth, it may only furnish a sequence of designs of decreasing weight that converges toward an optimum that is not itself a member of the considered class.Figure 5 illustrates this remark. The discrete radial loads at the periphery are to be transmitted to the central ring by a structure of minimal weight.If the word structure in this statement were to be replaced by the expressionFig. 5. Optimal structure for transmitting peripheral loads to central ring is trussrather than diskdisk of continuously varying thickness, the optimal structure of Fig. 5 would be excluded. Note that Fig. 5 shows only the heavy members. Between these, there are densely packed light members along the logarithmic spirals that intersect the radii at 45o ±The problem indicated in Fig. 3a has an infinity of solutions, each of which contains only tension members. Figure 6 illustrates a problem that requires the use of compression as well as tension members and has a unique solution. The horizontal load P at the top of the figure is to be transmitted to the curved, rigid foundation at the bottom by a trusslike structure ofFig. 6. Unique optimal structure for transmission of load P to curved, rigid wall. minimal weight, the stresses in the bars of which are to be bounded by-0σ and 0σ.The optimal truss has heavy edge members; the space between themis filled with densely packed, light members, only a few of which are shownin Fig. 6. Note that the displacements of the densely packed joints of thestructure define adisplacement field that leaves the points of the foundation fixed. A displacement field satisfying this condition wilt be called kinematically admissible.There is a kinematically admissible displacement field that everywhere has theprincipal strains 1ε =0σ/ E and 2ε=-0σ/E , where E is Young's modulus. Indeed, if u and v are the (infinitesimal) displacement components with respect to rectangular axes x and y, the fact that the invariant 1ε+2ε vanishes furnishes the relation x u +y v =0， （6）where the subscripts x and y indicate differentiation with respect to the coordinates. Similarly, the fact that the maximum principal strain has the constant value e1 yields the relation4x u *y v -(x v +y u )( x v +y u )=-421ε （7）In view of (6), there existsa function (),x y ψ such thatu =y ψ,v =-x ψ （8）Substitution of (8) into (7) finally furnishes4 2xy ψ+()2xx yy ψ-ψ=421ε (9)Along the foundation are, u = v = O, which is equivalent toψ=0， n ∂ψ∂=0 （10）where n∂ψ∂ is the derivative of T along the normal to the foundation are.The partial differential equation (9) is hyperbolic, and its characteristics are the lines of principal strain. The Cauchy conditions (10) on the foundation arc uniquely determine the function ψ, and hence the displacements (8), in a neighborhood of this arc.These displacements will now be used as virtual displacements in the application of the principle of virtual work to an arbitrary trusslike structure that transmits the load P to the foundation are (Fig. 6) and in which each bar is under an axial stress of magnitude %. With the notations used above in the presentation of Maxwell's argmnent, e W =i W =∑F λ. Here, ｜F ｜=0σ Aand ｜λ｜()0/E L σ≤, because no line element experiences a unit extension or contraction of a magnitude in excess of 0σ/E . Accordingly,e W =∑F λ≤∑｜F ｜｜λ｜≤ (20σ/E)V , (11) where V is again the total volume of material used in the structure.Next, imagine a second trusslike structure whose members follow the lines of principal strain of the considered virtual displacement field and undergo the corresponding strains. Quantities referring to this structure will be marked by an asterisk. Applying the principle of virtual work as before, one has *e W =e W , but*F *=*0A σ±and *λ=()0/E L σ± with correspondence of signs. Accordingly,e W =**F λ∑=()2*0/E V σ (12)In view of *e W =e W , comparison of (11) and (12) reveals that the second structure cannot use more material than the first.The argument just presented is due to Michell (Ref. 7), who, however,considered purely static boundary conditions and, consequently, failed to arrive at a unique optimal structure. The importance of kinematic boundary conditions for the uniqueness of optimal design was pointed out by the present author (Ref. 8).Figure 7 illustrates an important geometric property of the orthogonal curves of principal strain in a field that has constant principal strains of equal magnitudes and opposite signs. Let ABC and DEF be two fixed curves of one family. The angle c~ formed by the tangents of these curves at their points of intersection with a curve of the other family does not depend on the choice of the latter curve. In the theory of plane plastic flow, orthogonal families ofFig. 7. Geometry of optimal layout.curves that have this geometric property indicate the directions of the maximum shearing stresses (slip lines). In this context, they are usually named after Hencky (Ref. 9) and Prandtl (Ref. 10); their properties have been studied extensively (see, for instance, Refs. 11-13).Figure 8 shows the optimal layout where the space available for the structure is bounded by the verticals through d and B. Because the foundation arc is a straight-line segment, there are no bars inside the triangle dBC. Here again, the edge members are heavy, and the other members, of which only a few are shown, are comparatively light. The layout of these bars strongly resembles the trajectoriat system of the human femur (see, for instance, ReL 14, p. 12, Fig. 6). For further examples of Michell structures, see Refs. 15-16.4. New Method of Establishing Optimality CriteriaThe beam in Fig. 9 is built in at A and simply supported by B and C.Its deflection at the point of application of the given load P is to have the givenvalue δ. The beam is to have sandwich section of constant core breadth B and constant core height H. The face sheets are to have the common breadth B, and their constant thicknesses 1T 《H and 2T 《H in the spans 1L and 2L are to be determined to minimize the structural weight of the beam. Since theFig. 8. Optimal layout when available space is bounded by verticals through A and B. dimensions of the core are prescribed, minimizing the weight of the beam meansminimizing the weight of the face sheets. Moreover, since the elastic bending stiffness s i of the cross section with face sheet thickness i T , i = 1, 2, is 2/2i i s EBH T =, where E is Young's modulus,1122W L s L s =+ (13) may be regarded as the quantity that is to be minimized.Fig. 9. Beam with spanwise constant cross section.Let i x be the distance of the typical cross section in the span i L from the Left end of this span, and denote curvature and bending moment at this cross section by i k and i i i M s k =. The prescribed quantity P δmay then be written asP δ=i i i i M k dx ∑⎰=2i i i is k dx ∑⎰(14)where the integration is extended over the span i LWithin the framework of the problem, a beam design is determined by the values of i s , i = t, 2. If s i and si are two designs satisfying the design constraint (given valueof P δ), and i k and i k are the curvatures that they assume under the given load, itfollows from (14) that2ii i i s k dx ∑⎰ =i i i is k dx ∑⎰ （15）Moreover, since the curvature i k is kinematically admissible (i.e., derived from a deflection satisfying the constraints at the support) for the design i s , it follows from the principle of minimum potential energy for the design i s that22i i i i s k dx P δ-≤∑⎰22i i i i s k dx P δ-∑⎰ (16) Suppressing the terms 2P δ in (16) and using (15), one obtains the inequality()0i i ii i s s L μ-≥∑ （17） where2(1/)i i i i L k dx μ=⎰ （18） is the mean-square curvature in thespan i L . If12μμ= (19) it follows from (17) and (13) that the design s~ that satisfies (19) in addition to the design constraint cannot be heavier than an arbitrary design i s that satisfies only the design constraint. The condition (19) thus is sufficient for optimality; that it is also necessary may be shown as follows. With the definition()i i i i s s L λ=- （20） the condition that the design s i should not be heavier than the design i s takes the form 0i i λ≥∑. (21)On the other hand, the inequality (17), which followed from the principle of minimum potential energy, becomes0ii i λμ≥∑. (22)The quantities 1λ, 2λ and 1μ, 2μ will be regarded as the components of vectors λ and μ with respect to the same rectangular axes. The inequality (21) states that the vector λ cannot point from the origin into the half-space below the bisectors of the second and fourth quadrants, and the inequality (22) demands that the scalar product of λ and μ be nonnegative.Now, the optimal design s i and its mean curvature i s are unknown but fixed. The design i s , on the other hand, is only subject to the design constraint, which prescribes the value of P δ and, hence, determines the magnitude of the vector )t when its direction has been chosen. Moreover, in the neighborhood of the optimal design i s , there are designs i s of structural weights that come arbitrarily close tothe minimum weight. The corresponding vectors λ are arbitrarily close to the boundary of the half-space defined by the inequality(21). If the scalar product of λ and μ is to be nonnegative for all feasible vectors λ, the vector μ must be directed along the interior normal of this half-space at the origin, that is, (19) is a necessary condition for optimality.This proof of necessity is due to Sheu and Prager (Ref. 17).5. Multipurpose DesignFigure 11 illustrates a problem of multipurpose design. Under differentconditions of loading, one and the same structural element is to serve as tie, beam, or column. In the first case, its elongation under the given longitudinal loadL is not to exceed the given value λ In the second case, its deflection under the given central transverse load T is not to exceed the given value δ; and, in the third case, itsbuckling load is to have at least the given value B. Note that the design constraints are stated in inequality- form, because the optimal design may be governed by only one or two of them. It will, however, be assumed in the following that all three constraints are relevant. Proceeding as in Section 4, one then obtains the inequalities2`()0s s u dx -≥⎰，2``2()0s s H v dx -≥⎰， 2``2()0s s H w dx -≥⎰, (23) where u(x) is the longitudinal displacement in the tie mode, and v(x) and w(x) are the deflections in the beam and column modes.Taken individually, the relations (23) would yield optimality conditions that could be written as2`21u α=，22``21H v β= ，22``21H w γ= （24）where ,,αβγ are constants. It is readily seen that these optimality conditions are not compatible. For the load L to produce the constant longitudinalFig. 11. Multipurpose design.strain u' required by the first optimality condition, the element would have to have constant cross section, but the curvature v" that this element assumes under thetransverse load T will not satisfy the second optimality condition.Since the inequalities (23) thus cannot be exploited individually, combine them with positive multipliers to obtain2`222``222``2(......)()0s s u H v H w dx αβγ++≥⎰（25） This inequality immediately shows that2`222``222``2(u H v H w αβγ++=Const （26） is a sufficient condition for optimality. It can be shown that this condition is also necessary. Note that it may be written in the alternative form222222L T B Const ασβσγσ++= （27） where L σ,T σand B σare face-sheet stresses at the typical cross section in the tie, beam, and (buckled) column modes.For other examples and general theory, see Refs. 32-33.6. Concluding RemarksIn conclusion, it should be stressed that the design constraints discussed inSection 4, while typical, are by no means the only ones to which this method ofestablishing optimality criteria can be applied. In fact, new applications are still being developed. For example, the criterion (31) for optimal design for given dynamicdeflection has for the first time been established in the present paper, and no examples have as yet been worked out. ~Optimal design for given stiffness in stationary creep is treated in Ref. 35.Similarly, the restriction to optimal design of beams has been introduced here to simplify the discussion, but is not essential.优化结构设计W. PRAGER 3摘要。
电动车的车身轻量化与结构优化分析
电动车的车身轻量化与结构优化分析随着世界各国对环保与可持续发展的日益重视以及汽车行业的快速发展，电动车作为一种环保、节能的交通工具，受到越来越多人的关注和选择。
在电动车的设计与制造过程中，车身轻量化与结构优化是一项至关重要的任务。
本文将对电动车的车身轻量化和结构优化进行分析和探讨。
一、车身轻量化的意义车身是电动车的重要组成部分，也是整车的基础结构。
在电动车的设计中，车身轻量化可以达到以下几个重要的目的。
首先，车身轻量化可以提高电动车的能源利用效率。
通过减轻车身的重量，可以减少电动车在行驶过程中的能量消耗，提高续航里程。
这对于电动车的普及和推广意义重大。
其次，车身轻量化可以提高电动车的操控性能和安全性能。
降低车身的质量可以使电动车在行驶过程中更加灵活，提高加速性能和车辆的操控性。
同时，车身轻量化还可以减少碰撞时对乘车人员的伤害程度，提高电动车的安全性能。
最后，车身轻量化可以降低电动车的制造成本。
较轻的车身可以减少材料的使用量，降低生产成本。
这对于电动车的大规模生产和市场价格的下调有着积极的促进作用。
二、车身轻量化的方法实现电动车的车身轻量化有多种方法和途径。
下面将介绍几种常见的方式。
首先，选用轻量化材料。
在车身的设计和制造中，选择轻量化材料是一种有效的方式。
目前，常用的轻量化材料包括铝合金、高强度钢材等。
这些材料具有较轻的质量和良好的强度，可在保证车身强度的同时降低车身的重量。
其次，优化车身结构。
通过对车身结构的优化设计，可以降低不必要的重量。
例如，在设计车身骨架时，可以采用更加合理的结构布局和连接方式，减少零部件的使用量，达到轻量化的效果。
另外，采用先进的制造工艺和技术。
随着科技的不断进步，新的制造工艺和技术为车身轻量化提供了更多的可能性。
例如，采用3D打印技术可以实现复杂结构的一体成型，减少连接件的使用，降低车身的重量。
三、车身结构优化的意义除了轻量化外，车身结构的优化也是电动车设计过程中的重要环节。
车身结构优化设计与轻量化技术研究
车身结构优化设计与轻量化技术研究车身结构优化设计是近年来汽车制造领域的重要研究方向之一。
目的是通过优化设计实现汽车轻量化，减少能源消耗及环境污染。
本文将从车身结构和轻量化技术两个方面来探讨相关内容，以期能够对读者有所启迪。
一、车身结构优化设计车身结构是汽车设计中非常重要的一环。
在车身设计中，科技人员注重了减轻车身重量和提高汽车强度的相互关系。
为实现这个目标，结构设计必备以下几个方面：1、形状设计汽车的设计应该注重其形状，力学分析及有限元计算方法。
及在这个过程中，对汽车的每个结构部分都要进行设计，通过最大限度地应用材料实现轻量化和节能。
设计对零件和所有连接部件的异形构造结构都应该进行分析，以确定最佳配置与连接点。
2、汽车整体性能分析在这个过程中，汽车工程师需要确定各个组件的尺寸和数量，如发动机、变速器、车轮等部分。
在确定尺寸及部件数量后，可以进行力学分析，检查汽车的整体性能并确定强度。
3、车身各部分连接点设计车身连接点是车身各个部件之间的关键部位。
设计应该力求减少连接点数量和闯关点。
在实现这个目标的同时，还需要确保所有连接点的强度和连接功能。
二、轻量化技术研究与应用1、材料轻量化目前，轻量化技术中应用最广泛的方法之一是材料轻量化。
这个方法就是将传统重质材料替换为轻质材料。
这些轻质材料包括钢铝、镁、碳纤维和高分子塑料等。
根据使用部位不同，部分材料有其专用应用，发展出特殊复合材料以满足各种需求。
2、结构轻量化结构轻量化是通过优化设计汽车的构造，减少不必要的重量而实现轻量化。
汽车工程师需要考虑各个车身构件之间的联系，以确保加强性能不牺牲轻量化的效果。
3、技术改进与可靠性提升在实际应用中，轻量化技术的研究也需要结合现有技术和装备进行改进。
例如，在电动汽车中大量使用轻量化材料，同时也需要优化其节能技术，以达到更好的性能和可靠性。
结语：车身结构优化设计与轻量化技术的研究一直是汽车制造业的核心领域之一。
在未来的发展中，将需要通过持续的创新来实现轻量化和环保。
轻量化的意义外文文献翻译、汽车车辆类外文翻译、中英文翻译
附录AThe significance of lightweightAutomotive lightweight main guiding principle: to ensure stability on basis of improved performance, energy-saving design of the assembly partscontinuous optimization model spectrum.The lightweight car, the car is to ensure strength and security of the premise, as far as possible to reduce the car's curb weight, thereby enhancing the car's power, reducing fuel consumption and reduce exhaust pollution. Experiments show that if 10% of automobile weight reduction, fuel efficiency can bincreased by 6% -8%; vehicle curb weight is reduced by 100 kg hundred kilometers can be reduced 0.3-0.6 liters; 1% of vehicle weight reduction, fuel consumption can be reduced 0.7 %. At present, due to the need for environmental protection and energy saving, lightweight car has become the world trend of automobile development.Lightweight car is the main way:①mainstream car models continue to optimize the size, specification size to retain the main parameters of the premise, to enhance the structural strength of the vehicle, reducing the amount of supplies;② use of lightweight materials. Such as aluminum, magnesium, ceramic, plastic, fiberglass or carbon fiber composite materials;③ The computer design. Such as finite element analysis, local reinforcement design;④ the use of load-bearing body, thinning the thickness of the body sheet metal.Among them, the current major automotive lightweight measures mainly the use of lightweight materials. for automotive lightweight design of the metal materialThe main vehicle through the car's lightweight materials to improve economy to contribute. Theoretical analysis and experimental results show thlightweight is to improve vehicle fuel economy and effective way. In order to equirements of lightweight vehicles, some new material emerged and expanded the scope of application.Nonferrous AlloysTo passenger cars, the 1973 vehicle used by the non-ferrous alloy of the total timber weight ratio of 5.0% in 1980 to 5.6%, while in 1997 it reachNon-ferrous alloys in automotive applications is the rapid growth trend of development of automotive materials. AluminumSteel aluminum density of about 1 / 3 is the most widely used lightmaterials. United States production of automotive products, for example, aluminum alloy 1976 per vehicle is only 39kg, 1982 reached 62kg, while in 1998 it reach100kg.(1) cast aluminum alloyMany elements can be used as cast aluminum alloy elements, only Si, Cu, Mg, Mn, Zn, Li in the mass production of great significance. Of course, widely used in the car is not thesimple binary alloy but also a variety of elements added to obtain a good overall performance.The automobile industry is the main market of aluminum castings, such as Japan, 76% of aluminum castings, aluminum die casting 77% of automotive castings. Mainly used in aluminum alloy engine block castings, cylinder heads, pistons, intake manifold, rocker arms, engine mount bracket, air compressor connecting rod, clutch housings, clutch housings, wheels, brake parts, handles and cover shell-type parts and so on.Aluminum casting inevitably flawed, casting can not be heat treated produce a result in the use of higher strength aluminum alloy castings is limited. On the casting for this production process has been improved, semi-solid casting forging forming method and method will be used to craft the future more.(2) deformation of aluminum alloyRefers to deformation of aluminum alloy aluminum alloy plate and extrusions and forging materials, mainly used in automotive body panels on the boengine radiators, air conditioning condenser, evaporator, wheels, trim and suspension parts.As the effect is obvious lightweight, aluminum alloy car body applications is expanding. September 1990 as the Japanese started selling cars with a Honda NSX Unibody aluminum, cold rolled steel plate than in the same light 200kg body attracted worldwide attention. NSX to the entire vehicle with aluminum 31.3%, as in the aluminum car body, using the 6000 series alloy outer plate, inner plate with 5052-0 alloy, most use the 5182-0 alloy frame; that the side door of the strength and stiffness requirements are high, the use of the 6N01 alloy-based, appropriate adjustments to the Mg and Si content of the alloy. Also useful in Europe and America for 2036 and 2008 alloy plates inside and outside the body. Aluminum radiator originated in Europe and then throughout the world. In Europe, after the 1980s, aluminum radiator has occupied 90% of the market. With? Car air conditioning, oil coolers, etc. The extensive use of aluminum heat exchangers market is rapidly expanding. From the material point of view, aluminum heat exchangers in a wide range of coating material applied to a large extent due to aluminum and aluminum cladding successful development.(3) aluminum matrix compositesLow density of aluminum matrix composites, strength and high modulus, good thermal fatigue properties, but in automotive applications by the price and production quality control and other constraints, have not formed a large scale. Currently, aluminum matrix composites in the connecting rod, piston, cylinder body holes, brake discs, brake calipers and other parts of the drive shaft tube test or showing superior performance, such as developed by Honda does not? Enhanced aluminum matrix composite wire rod steel rod down more than 30% weight on the 1.2L gasoline engine can increase fuel economy 5%; using chilled aluminum powder and SiC powder (weight percentage of 2%) and mixed extruded into rods, forged with the shape of the piston rod for high strength can be 20% lower weight, engine power has increased significantly; composite enhanced with aluminum piston head to cancel the first ring groove of austenitic cast iron inserts can drop weight 20%; aluminum matrix composite brake discs reduce weight than cast iron brake disc 50%.MagnesiumAluminum magnesium density of about 2 / 3, in the practical application of the metal is the lightest. Vibration absorption of magnesium alloy cutting ability is good, metal casting performance, it is suitable for manufacturing automotive parts.Most of the die casting magnesium alloy in the form of applications in the car, magnesium die casting aluminum production efficiency than the 30% to 50%. The newly developed non-porous casting method (Pore Free Diecast) can produce no holes and can be heat-treated magnesium die castings.Magnesium castings used in the car the earliest examples is the wheel rim. In the car on the trial or application instance there magnesium alloy clutch housing, clutch pedal, brake pedal mounting bracket, dashboard frame, seat, steering column parts, steering wheel round core, transmission housing, engine mounts, cylinder head and cylinder head cover and so on. With the traditional system of the steering column bracket compared to the zinc, magnesium and heavy parts down 65%; with traditional steel steering wheel core compared to 45 percent weight reduction of magnesium parts; compared with the aluminum cylinder head, magnesium parts 30% weight reduction; traditional stamping and welding structural steel brake pedal bracket compared to the overall weight of magnesium castings down 40%, while its rigidity is also improved.Magnesium matrix composites has also made progress, with SiC particles as reinforcement, mixing technology is the use of liquid magnesium-based composite materials with good performance and lower production costs. In the AZ91 alloy by adding 25% SiC particle reinforced composites than the matrix alloy with 23% higher tensile strength, yield strength increased 47%, elastic modulus increased 72%.TitaniumTitanium density 4.5g/cm3, with high strength, high temperature strength and corrosion resistance and so on. Because titanium is expensive, since I saw the car in the luxury car and a small amount of individual applications. However, the application of titanium in the test car has a lot of research work. For example, with α + β titanium alloy of the engine connecting rod system, the equivalent strength quenched and tempered steel 45 levels, and weight can be reduced 30%; β of titanium (Ti-13V-11Cr-3Al, etc.) and time by the strong cold treatment, intensity of up to 2000MPa, can be used to manufacture suspension springs, valve springs and valves, etc., and the tensile strength of 2100MPa of high strength steel than titanium springs can be 20% lower weight.The biggest obstacle to the application of titanium from its high price, Toyota Central Research Institute has developed a low-cost titanium matricomposites. The composite material to Ti-6Al-4V alloy matrix to the reinforcements powder metallurgy method of production, has been applied in the engine connecting rod.附录B轻量化的意义汽车轻量化主要指导思想：在确保稳定提升性能的基础上，节能化设计各总成零部件，持续优化车型谱。
车身结构优化设计与轻量化
车身结构优化设计与轻量化近年来，随着社会的发展和人们生活水平的提高，汽车已经成为人们生活中不可或缺的一部分。
然而，汽车的快速发展和普及也带来了一系列的环境和能源问题。
为了应对这些挑战，汽车制造商们开始关注车身结构的优化设计与轻量化。
本文将探讨这个领域的最新发展和技术。
首先，车身结构的优化设计是实现汽车轻量化的关键。
轻量化可以降低车辆的燃油消耗和排放，减少对环境的负面影响。
在设计过程中，利用计算机辅助工具，如有限元分析软件和优化算法，可以对车身结构进行模拟和优化。
通过对结构进行材料优化、强度分析和厚度分布优化，可以使车身更加轻巧但仍能满足安全性能的要求。
在材料的选择方面，新一代高强度钢和铝合金等轻质材料的应用成为车身轻量化的主要途径之一。
这些材料具有较高的强度和刚度，可以在减少重量的同时确保车辆的结构强度。
此外，也有一些新兴的材料，如碳纤维复合材料和镁合金，它们具有更高的比强度和比刚度，能够进一步减轻车身重量，但其成本相对较高，应用仍面临一些挑战。
此外，在车身结构的设计中，形状优化也是一项重要的技术。
通过改变车身的外形和曲面设计，可以减少空气阻力，提高车辆的行驶稳定性和燃油经济性。
借助计算机模拟和流体力学分析，可以优化车身形状，减少气动阻力，从而降低燃油消耗。
同时，车身结构的优化设计还要考虑乘员安全和碰撞保护的因素，确保车辆在发生事故时能够提供最大的保护。
除了车身结构的优化设计，轻量化技术也是实现汽车节能减排的重要手段之一。
减轻车身重量可以降低车辆的能量消耗和动力需求，进而减少汽车的排放。
除了上文提到的材料的应用，还有一些具体的轻量化技术可以采用。
例如，采用合理的螺栓连接和焊接技术，可以减少结构的重量和材料的使用；采用多材料混合设计，可以根据不同部位的要求选择不同材料，实现更好的轻量化效果；同时，还可以利用3D打印等新兴技术制造轻量化的零部件，实现个性化定制和优化设计。
然而，车身结构的优化设计与轻量化并非一帆风顺。
车身轻量化设计方法
10.16638/ki.1671-7988.2018.07.044车身轻量化设计方法郎勇（吉利汽车前瞻技术研究部，上海201501）摘要：轻量化技术已经成为汽车工业发展的重要研究课题之一，如何提高汽车轻量化水平已引起了汽车行业的广泛关注，文章主要从车身轻量化评价指标入手，根据车身结构特点，从车身设计角度，详细探讨了车身轻量化的实现方法。
关键词：汽车；轻量化；优化设计中图分类号：U462.1 文献标识码：A 文章编号：1671-7988(2018)07-142-03Design Method of Automobile LightweightLang Yong( Geely Automobile Prospective Technology Research Dept, Shanghai 201501 )Abstract:The lightweight technology has become one of the important research topics in the development of automobile industry. How to evaluate the automobile lightweight level has attracted the attention of the automobile industry. In this paper, the article mainly starts with the evaluation index for the lightweight of body, According to the characteristics of body structure, from the angle of body design, the realization method of body lightweight is discussed in detail.Keywords: Body; lightweight; Optimal designCLC NO.: U462.1 Document Code: A Article ID: 1671-7988(2018)07-142-03引言随着汽车设计技术水平的提高，轻质材料的推广，工艺制造水平的提高，通过合理的车身结构、新材料的应用、新工艺实施，以更低的重量，实现整车性能目标成为可能。
轻量化的意义外文文献翻译、汽车车辆类外文翻译、中英文翻译
轻量化的意义外文文献翻译、汽车车辆类外文翻译、中英文翻译by improving the design of assembly parts and continuously optimizing the model spectrum。
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基于CATIA知识工程优化的车身轻量化设计
基于CATIA知识工程优化的车身轻量化设计【摘要】本文提出一种基于CA TIA知识工程优化模块，对车身钣金件进行轻量化的设计方法，并以某车的顶棚拉手支架为例，来说明此设计方法的应用过程，采用此方法最终可以得到满足模型总质量最小，符合拉手支架刚度要求的零件，它们的材料利用率将到达最优，同时到达轻量化车身的目的。
此方法具有科学高效，易于掌握，利于经验积累等优点，并且可以广泛拓展到白车身总成的轻量化设计，底盘件的轻量化计算，以及发动机的优化计算等领域。
【关键词】知识工程优化，钣金件，轻量化，有限元A Lightweight Design Method of Auto Body Based on CATIA KEOAbstract: The present paper mainly introduces a lightweight design method of auto body based on CATIA knowledge engineering optimization, and the application example in bracket of grip of headlining is given. Thus you can gain the light and strong bracket and the efficiency of material of the bracket is high. This method is proved to be efficacious, convenient and easy to accumulate experience; it can be used on lightweight design of BIW assembly, chassis, engine and so on.Key words: KEO, sheet panel, lightweight, FEM在能源日趋紧张的今天，汽车节油新技术不断涌现，而采用车身轻量化技术来提高汽车燃油经济性的传统方法一直得到人们的推崇。
轻量化设计的汽车零部件用高强度钢来抗凹外文文献翻译、中英文翻译、外文翻译
Lightweight design of automobile component using high strength steel based on dent resistanceAbstractLightweight and crashworthiness are two important aspects of auto-b-ody d esign. In this paper, based on the shallow shell theory,the express-ion of dent resistance stiffness of double curvatured shallow shell is obt-ained under the co ncentrated load condition. The critical loads resulting i-n the local trivial dent i n the center of the shallow shell is regarded as the important index for the lig htweight of the automobile parts. This rule is applied to the lightweight design of bumper system by using high stre-ngth steel instead of mild steel. The cra shworthiness simulation of the li-ghtweight part proves the validity of the light weighting process.Keywords: High strength steel; Lightweight; Dent resistance1.IntroductionIn recent years, the retaining number of automobiles has been incre-asing steadily, which has impacted the society and human life greatly. S-uch situatio n leads to many severe problems such as fuel crisis, environ-ment pollution. T he international association of aluminum stated that petr-ol consumption can de crease by 8–10% with 10% reduction of car weig-ht [2]. Thus, automobile ligh tweight is a basic way to fuel saving.In order to reduce the automobile weight, there are two important metho ds [3]: One, automobile parts are redesigned to optimize the struc-ture. By usi ng thinning, hollowing, minitype, and compound parts, car we-ight can be redu ced. The other, more and more lightweight materials, s-uch as aluminum alloy, high strength steel, composite material, are wide-ly used as lightweight materi als to replace the traditional materials like mild steel [4]. These materials cou ld reduce the weight remarkably. Mat-erial replacement is generally more effect ive in automobile lightweighting than structure modification. With the introducti on of automobile safety leg-islation, crashworthiness and safety should be consi dered as preconditio-ns in lightweighting design of auto-body. High strength ste el is widely us-ed in automobile replacing the traditional material of mild steel.High strength steel sheet can be used in auto-body to improve com-ponent s_ impact energy absorption capacity and resistance to plastic def-ormation. Th e automobile weight can be reduced by use of high strength steel sheet of a t hinner thickness to replace the mild steel sheet of bo-dy parts [1,3]. Comparing with aluminum, magnesium,and composite mat-erials, high strength steel has better economy in that its raw material an-d fabrication cost are cheaper. Besid es, high strength steel can be direc-tlyused in product line including forming, wielding, assembling, and painting. The operating cost can be saved since there is no need adjusting the whole line Outside of automobile body, there are several sheet metal Panels, most of which are shallow panels. Dent resistance is the ability to retain- the shape a gainst sunken deflection and local dent under the external force. Dent resistan ce of automobile panels becomes an important issue and quality criterion. Ther efore, dent resistance stiffness of automobile p-anels should be tested and evalu ated in the process of panel design an-d manufacture.Some reported methods of testing are listed below [6–8]:1. Test the displacement of sunken deflection fp underfixed external forc-e2. Test the external force f to obtain fixed displacement of sunken defle-ction.3. Test the slope of force–displacement curve under external load.In this study, the second method will be used. The rest of this paper is organized as follows: In Section 2, the expression of dent resistance stiffness of double curvatured shallow shell is obtained under the concentrated load condition based on the shallow shell theory. The critical load resulting in the local trivial dent in the center of the shallow shell is regarded as the important evaluating index for the dent resistance of the automobile parts. This rule is applied in Section 2 to the lightweighting design of bumper system by using high strength steel instead of mild steelwith crashworthiness simulation.2.Dent resistance analysis of double curvatured shallow shell2.1.Dent resistance stiffness analysis of shallow shellShell with mid surface can be characterized into three features: thickness h, mid surface dimension L, curvature radius r, which satisfies h/r_ 1. When there exists h/L _ 1, the shell can be defined as thin shell. If L/ r _ 1 is added besides the above two conditions, the thin shell is regarded as shallow shell [10].As Fig. 1 shows, the plane x–y is the projection of the mid surface of shallow shell along the z-axis. Supposing M is an arbitrary point on mid surface, two planes QMN & PMN are made paralleling to coordinate plane OYZ and OXZ, respectively. The two edges PM and QM can be regarded approximately as vertical because of mid surface_s flatness. At the same time, the line MN is normal to mid surface. Thus, MN, QM, PM can constitute a perpendicular reference frame MPQN, whose difference from orthogonal coordinate system OXYZ can be ignored. And PM and QM are denoted by a and b, the curvilineal coordinate of MPQN.Assuming the Z coordinate of the point M is z, the analytical equation of mid surface is expressed as follows:(,)z F x y = (1) The following equations can be obtained because of the flatness of the shell:222()1,)1,)1z z z x y x∂∂∂≤≤≤∂∂∂ (2) The curvature and torsion of mid surface can be approximated to:22222,,x y xy z z z k k k x y x y∂∂∂=-=-=-∂∂∂∂ (3) The Lae coefficients of mid surface along a and b directions are deduced:121,1ds ds dx dy A B d dx d dyαβ====== (4) Applying concentrated force P along Z-axis and ignoring the influence of the transverse shear resultant forces, the balance differential equations of shallow shell are:120,0N N S S x y y x∂∂∂∂+=+=∂∂∂∂ 1212()(0,0)0,x y Q Q k N k N P x yδ∂∂-++++=∂∂ (5) 12112212,,M M M M Q Q y x x y∂∂∂∂=+=+∂∂∂∂ where d(0,0) is Dirac-d function.The compatibility equation of shallow shell is2212()0k N N Et w ∇+-∇= Where 222222222222,x y k k x y x y ∂∂∂∂∇=+∇=+∂∂∂∂ (6) Expressing the moment resultants M1, M2 and M12 by the function of transverse displacement w, the basic equations of shallow shell under concentrated transverse forces are:2212D ()(0,0),x y w k N k N P δ∇∇++=2212()0k N N Et w ∇+-∇= (7)221222N N x y ∂∂=∂∂ where N1 is the membrane stress resultant in X-direction; N2, the membrane resultantin Y-direction; D, the Fig. 1. Double curvature shallow shell. bending stiffness of shallow shell.It is very difficult to solve above equation. According to practical situation, sunken deflection will only concentrate on a small area around external force P, so infinite large shallow shell [5] is assumed in this study. Because w, N1, N2 are symmetric about x-, y-axis, all orders of derivatives of w, N1, N2 become to zero at infinity. The following equations can be achieved by Fourier transformation to Eq.(7):2212ˆˆˆ()()x y D w k N k N Pξη+++= 222212ˆˆˆ()()()0y xN N Et k k w ξηξη++-+= (8) 2212ˆN N ξη= Where(0,0)i x iny P e e dxdy P ξδ∞∞---∞-∞=⎰⎰00ˆ4cos()cos()w w x y dxdy ξη∞∞=⎰⎰100ˆ4cos()cos()N w x y dxdy ξη∞∞=⎰⎰ (9) 200ˆ4cos()cos()N w x y dxdy ξη∞∞=⎰⎰ From Eq.(8), ^w can be obtained. Reverse Fourier transformation to ^w and polar coordinates transformation to n, g, w under polar coordinate system can be gained/22042222cos(cos )cos(sin )12(cos sin )p x y P x y w d d D k k tπρρθρθρθπρθθ∞=++⎰⎰ (10) Put x = 0 and y = 0 in Eq.(8), the relationship between deflection fp and concentrated force P of rectangle shallow shell can be achieved as follows:224(1)3x yp Et k k P f μ=- (11) Finally, dent resistance stiffness of shallow shell K is obtained224(1)3x y p Et k k P K f μ==- (12) This equation explains synthetically the relationship between the dent resistance stiffness of double curvature shallow shell and all influencing factors including material properties, geometry parameters, which can be used to guide design, material select and manufacture.2.2. Analysis of critical load causing local trivial dentFor quantitative evaluation of critical load against local dent resistance of panels, several experience formulas have been brought forward by researchers. Based on large numbers of experiments, Dicellello [9] stated a formula that expresses minimum energy W causing visible trivial dent trace by thickness t, yield stress rs and basic dent resistance stiffness K24s t W C K σ=(13)where C is proportional constant. From Eqs. (12) and (13), the critical load Pcr resulting in the local trivial dent in the center of the shallow shell can be achieved, which is defined as the evaluating index2er s P C t σ=(14)From Eq. (14), there is a closely correlation between critical loads Pcr and thickness t, yield stress rs. The critical load can be a rule to carry out lightweight design of automobile parts by using high strength steel instead of mild steel.3. Example and crashworthiness analysis3.1. FE model of full car and its crash simulationA detailed finite element model has been established based on a passenger car refitted from a saloon car, which is showed in Fig. 2. To ensure the correctness and effectiveness of FE model, the following methods are adopted:1. Since the goal is to simulate the frontal impact of the car, the meshing of front car body is denser than that of the rear car body.2. Reduced integration method with hourglass control is taken for 4 noded shell element and 8 noded brick solid element to improve the efficiency of simulation.3. By using of the meshing and mass scaling technology, the characteristic length of the minimal element is ensured to improve the simulation efficiency.4. Materials constitutive with Cowper –Symonds strain rate item is used for steel parts.5. Automatic single surface contact algorithm is adopted in the simulation aiming at complexity of car impact simulation.6. Spot weld element with failure rule that considering the couple of normal force and shear force is used to simulate the spot weld connection between auto parts.Explicit dynamic FEM software LS-DYNA Version 950 is used to simulate the frontal impact of the car against a rigid wall at the speed of 50 km/s according to the National Crash Legislation CMVDR294. A real car crash experiment is done at Car Crash Lab settled in Tsing Hua University. By comparing the time history of acceleration of certain position on the A pillar within 0.1 s, the simulation gives areasonable fit to the experiment results, which guarantees the correctness of FE model and gives a nicer base for the next lightweighting optimized design.3.2. Lightweighting design and crashworthiness analysisThe use of high strength steel is one of the effective ways to reduce car weight. However, the performance (such as crashworthiness, stiffness, and dent resistance) of part made of new material should be assured. For example, the front parts of a car are major energy absorption parts in the process of car crash, so energy absorption performance without affecting the safety of passengers should be assured in the design of front parts of a car. In this research, the bumper of the passenger car is studied under different materials but remaining its dent resistance.The mechanical properties of mild steel and high strength steel are listed below (see Table 1).The evaluation index of dent resistance for bumper using mild steel is21111cr s P C t σ=When high strength steel is used to replace the mild steel remaining its primary shape and dent resistance performance, the new thickness t2 of high strength steel can be achieved112122s s C t t C σσ= From (16), the thickness of bumper that uses high strength steel is gained and updated in the full car FE model. The deformation history of bumper using new material is achieved after the car crash is re-simulated with updated part thickness (see Fig. 3).By simulation, the deformations of bumper made of two different kinds of material are similar in that plastic hinge and tensional plastic deformation appear in the middle part of bumper. And the energy absorption history is shown in the following for beam of the bumper. From Fig. 4 the difference of the energy absorption between twomaterials is small, about 4.1%for beamof the bumper,from which a conclusion can be drawn that it is feasible to reduce the thickness of the bumper panel based on the dent resistance evaluation index studied in this research.4. ConclusionDent resistance performance of small curvature shallow shell parts in automobile is studied in this paper, which enables the follows:1. Dent resistance stiffness under concentrated force is given for such parts.2. The critical load resulting in the local trivial dent in the center of the shallow shell has been deduced, which in turn becomes the index to evaluate the dentresistance of automobile parts.3. The validity of evaluating index is proven by applying the developed rule to the lightweight design of bumper system using high strength steel instead of mild steel through crashworthiness simulation.轻量化设计的汽车零部件用高强度钢来抗凹摘要：轻巧耐撞性是汽车车身设计的两个重要因素。
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到CAE方法它将节省大量的测试成本，及时指导车身的设计结构，提出优化方法; 有限元法已经成为在车身特性设计和分析过程中不可或缺的[3]。
在结构优化过程中，首先，敏感性分析应该是对的以设计变量的灵敏度来确定目标[4]然后根据灵敏度修改设计变量以获得最佳设计目标与约束条件。
设计变量的灵敏度与A轻量级优化方法相关其目标函数是由客观因素的变化来衡量的单位设计变量的变化，灵敏度分析是opti的基础可以根据敏感度进行结构分析，优化结构分析，这将节省大量的计算时间并改进优化因此，在工业中期望提供一种改进的轻型汽车车身结构，特别是车身框架，其保持所需的强度，刚度和稳定性特性，以满足乘客安全和车辆性能标准。
此外，身体结构应当可使用现有技术和材料制造，以实现减轻重量而不增加成本。
此外，主体结构应该能够减少总成分，以进一步降低成本和制造时间。
2灵敏度分析的基本理论
从结构分析可以分为动态分析和静态分析两个方面，结构敏感性分析也可以分为动力敏感性分析和静态敏感性分析。
动态灵敏度分析包括特征值灵敏度分析，传递函数灵敏度分析和动态响应敏感性分析。
静态敏感性分析可以是压力，位移等。
对于车辆，灵敏度分析是指车身刚度强度，自由模式和敏感性分析的应变能，部分结构参数包括材料厚度和横截面转动惯量[5 |。
有两种计算灵敏度的方法，推导方法和伴随结构的方法。
直接推导法，是亲由胡力构成R.Kapoor M.P，然后由许多人开发和推广人们在广泛的领域。
直接推导法具有明显的物理意义概念，简单的数学理论方便计算可扩展从一阶敏感度到高阶灵敏度，因此是广泛的在工业领域[6]。
3 BIW扭转刚度分析
汽车BIW的有限元模型由769 862组成元素包括三角形元素和四边形元素重量为371.2公斤。
为了比较车辆扭转刚度试验过程中的有限元分析模型需要使用直径为50 mm的梁单元进行模拟测试设备如图有限元边界条件分析模型如图1所示。
3，左右后方的震动塔都很硬点被限制在X，翻译DOF左右前方震动吸收塔受限于X方向DOF。
负载被施加到两个前冲击塔的中心在垂直方向这是平等的，但与相反的方向。
负载通过公式F = M / L获得，其中M是测试扭矩值，L是左右前冲击中心之间的距离塔。
刚度结果与施加的扭矩无关，但对于目的是与试验结果进行比较，施加的扭矩为4000 Nm，进行试验过程如图1所示。
4。
导出左右负载点的Z方向位移使用公式计算扭转刚度值，扭转刚度K如下： M施加扭矩; ts Z大号和AZ尺是负载点的Z方向位移;L是负载点之间的距离。
轻量级指标[7]，是评估的关键因素之一车身结构的性能，通过车身扭转刚度和质量计算该越少越好轻
体重指数与车身质量成正比，反之亦然与扭转刚度成比例因此减轻轻量级的指标最多有效的方法是增加扭转刚度，同时减少体重，车身轻量级指数计算公式如下。
在公式中：M是体重; K是扭转刚度; A是车身项目区（轴距x胎面）。
通过有限元软件的分析计算，刚度车身如表1所示，主体Z方向的轮廓显示如下图。
5，根据图车身的位移改变了总之，符合预期的负载情况。
4 BIW结构刚度灵敏度分析与优化
根据敏感性分析，各种零件的影响相关性能可以判断，灵敏度更大意味着部件厚度更多重要的是相应的属性，因此关键组件可以确定并优化。
部件的主要特点包括结构。
形状，厚度，材料性能，加工性等，用于线性分析如扭转刚度材料的非线性特性不会影响扭转刚度性能，部件的结构形状和尺寸是扭转刚度的主要因素。
车身扭转刚度主要取决于连接器部件属性和车身com几何特性部分厚度和部分厚度是确定几何的因素之一部件的特点，是影响扭转刚度的主要因素之一。
通过有限元分析，部件厚度对扭转的影响刚度可以方便地研究。
在本文中，50部分的车身（包括对称部分）设置灵敏度分析，在分析中只有一面的对称分量被选择用于分析最后39厚灵敏度分析中选择变量。
选择的组件显示在刚度灵敏度优化分析中，将最小质量设置为优化目标和左载荷点的Z位移为约束条件部件厚度相对于扭转刚度和整个车辆质量的敏感度。
较大的灵敏度对此有较大的影响目标性能和目标性能可以通过增强来有效增加厚度，同一部分可能对扭转刚度有很大的影响车辆质量，当增加厚度以提高扭转刚度也将导致车辆质量的提高这种类型的零件不适合通过增加厚度来优化扭转刚度。
不同部件对扭转刚度和质量的贡献是多样的如何选择合适部件用于厚度优化，以提高扭转刚度，而不会影响体重是指优化效率的问题，这可以通过扭转刚度灵敏度和质量的比较来表示灵敏度，一部分的比较灵敏度相对于其他部分较大的手段，改变其厚度可以提高扭转刚度，但较少增强车辆质量比其他人比较灵敏度优化效率。
为了提高扭转刚度和减少车辆质量，最终优化可以采取措施来增加部件厚度较大的部件比较灵敏度，减少组件的厚度比较灵敏度。
零件量规根据图1进行了优化。
9，扭转刚度为重量相对于原始模型，扭转刚度增加38 Nm /度整体重量减少3.1公斤轻量级指数下降0.061。
5车身结构优化
车身结构优化，优化钣金结构肋结构的分布和尺寸可用于优化零件强度和刚度，同时减轻重量，为片材提供优化方金属零件设计。
在设计领域，车身优化决定了最佳由元素节点扰动引起的位置和优化参数质量和体积对车身优化中的元素
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