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QD10t-31.5m箱形双梁桥式起重机起重小车设计【7张CAD图纸和说明书】

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目录

第1章 前言··········································································1

1.1 国内外起重机发展情况·······················································1

1.2 桥式起重机定义及特点·······················································4

1.3 实习地点及实习内容··························································4

第2章 总体设计···································································4

2.1 概述·············································································5

2.2 传动方案的确定·······························································6

2.3 基本参数······································································10

第3章 起升机构的设计计算···················································12

3.1 选择钢丝绳···································································12

3.2 滑轮和卷筒的计算···························································13

3.3 计算静功率···································································15

3.4 选择电动机···································································15

3.5 验算电动机的发热条件······················································15

3.6 减速机的初选································································16

3.7 校核减速机···································································16

3.8 制动器的选择································································17

3.9 联动器的选择································································17

3.10 验算起动时间·······························································18

3.11 浮动轴强度验算····························································19

第4章 运行机构的设计计算···················································21

4.1 确定机构传动方案···························································21

4.2 选择车轮与轨道并验算其强度··············································21

4.3 运行阻力计算································································23

4.4 选择电动机···································································24

4.5 验算电动机发热条件························································25

4.6 选择减速器···································································25

4.7 验算运行机构和实际所需功率··············································25

4.8 验算起动时间································································26

4.9 验算起动不打滑条件························································27

4.10 制动器的选择·······························································27

4.11 选择联轴器··································································28

4.12 验算低速浮动轴强度·······················································29

第5章 零部件的设计计算·····················································31

5.1 滑轮的尺寸计算与选择······················································31

5.2吊钩组的选择·································································32

5.3 车轮轴的设计计算···························································35第6章 零部件的设计计算·····················································38

6.1 梁Ⅰ···········································································38

6.2 梁Ⅱ···········································································40

6.3 梁Ⅲ···········································································42

6.4 梁Ⅵ···········································································44

6.5 梁Ⅴ···········································································48

第7章 毕业设计小节····························································53

参考文献············································································54

附:英文原文

英文译文

毕业实习报告

第1章 前言

1.1国内外起重机发展概况

起重运输机械行业在我国从上世纪五六十年代开始建立并逐步发展壮大,该行业已形成了各种门类的产品范围和庞大的企业群体,服务于国民经济各行业。随着我国经济的快速发展,起重运输机械制造业也取得了长足进步。2005年起重运输机械行业销售额达到1272亿元,“十五”期间平均每年增长超过30%,2006年依然保持着持续增长的态势,目前市场前景非常好。近年来,国家重点发展能源(其中煤炭工业迅猛发展,起重运输机械制造业将提供所需的竖井提升设备、斜井防爆下运带式输送机、防爆移置式带式输送机、装车机、露天矿连续开采输送设备、用于洗选设备的各种输送设备等)、电力(各种电站专用桥式/门式起重机、料场用物料搬运装卸设备、输煤给煤栈桥内物料输送设备、环保排灰输送设备、水电站用闸门启闭机械、升船机、核电站废料处理专用起重机等将有较大需求)、石化(起重运输机械制造业将提供所需的自动灌装和包装码垛设备、仓储专用设备、厂内和车间内物料搬运装卸设备等)、冶金(对各种冶金起重机、厂内和车间内物料搬运装卸设备、料场堆取料与混匀料设备等将有较大需求)、造船、交通等工业领域(需要大量的高效、节能、低污染、智能化、柔性化、成套化的物料搬运装卸设备)。


内容简介:
Vehicle System DynamicsVol. 44, No. 5, May 2006, 387406Control of a hydraulically actuated continuouslyvariable transmissionMICHIEL PESGENS*, BAS VROEMEN, BART STOUTEN, FRANS VELDPAUSand MAARTEN STEINBUCHDrivetrain Innovations b.v., Horsten 1, 5612 AX, The NetherlandsTechnische Universiteit Eindhoven, PO Box 513, 5600 MB Eindhoven, The NetherlandsVehicular drivelines with hierarchical powertrain control require good component controller tracking,enabling the main controller to reach the desired goals. This paper focuses on the development ofa transmission ratio controller for a hydraulically actuated metal push-belt continuously variabletransmission (CVT), using models for the mechanical and the hydraulic part of the CVT. The controllerconsists of an anti-windup PID feedback part with linearizing weighting and a setpoint feedforward.Physical constraints on the system, especially with respect to the hydraulic pressures, are accountedfor using a feedforward part to eliminate their undesired effects on the ratio. The total ratio controllerguarantees that one clamping pressure setpoint is minimal, avoiding belt slip, while the other israised above the minimum level to enable shifting. This approach has potential for improving theefficiency of the CVT, compared to non-model based ratio controllers. Vehicle experiments show thatadequate tracking is obtained together with good robustness against actuator saturation. The largestdeviations from the ratio setpoint are caused by actuator pressure saturation. It is further revealed thatall feedforward and compensator terms in the controller have a beneficial effect on minimizing thetracking error.Keywords: Continuously variable transmission; Feedforward compensation; Feedback linearization;Hydraulic actuators; Constraints1. IntroductionThe application of a continuously variable transmission (CVT) instead of a stepped transmis-sion is not new. Already in the 50s Van Doorne introduced a rubber V-belt CVT for vehiculardrivelines. Modern, electronically controlled CVTs make it possible for any vehicle speed tooperate the combustion engine in a wide range of operating points, for instance in the fueloptimal point. For this reason, CVTs get increasingly important in hybrid vehicles, see forexample 13. Accurate control of the CVT transmission ratio is essential to achieve theintended fuel economy and, moreover, ensure good driveability.The ratio setpoint is generated by the hierarchical (coordinated) controller of figure 1. Thiscontroller uses the accelerator pedal position as the input and generates setpoints for the localcontrollers of the throttle and of the CVT.*Corresponding author. Email: pesgensdtinnovations.nlMichiel Pesgens was previously affiliated with Technische Universiteit Eindhoven.Vehicle System DynamicsISSN 0042-3114 print/ISSN 1744-5159 online 2006 Taylor Fshift= Fp (rcvt,primes) Fs(10)An axial force difference Fshift, weighted by the thrust ratio results in a ratio change, and istherefore called the shift force. The occurrence of pin the model (10) is plausible because anincreasing shift force is needed for decreasing pulley speeds to obtain the same rate of ratiochange. The reason is that less V-shaped blocks enter the pulleys per second when the pulleyspeed decreases. As a result the radial belt travel per revolution of the pulleys must increaseand this requires a higher shift force. However, it is far from obvious that the rate of ratiochange is proportional to both the shift force and the primary pulley speed. kris a non-linearfunction of the ratio rcvtand has been obtained experimentally. Experimental data has beenused to obtain a piecewise linear fit, which are depicted in figure 6. The estimation of krhas392 M. Pesgens et al.Figure 5. Contour plot of (rcvt,primes).Figure 6. Fit of kr(rcvt); greyed-out dots correspond to data with reduced accuracy.Hydraulically actuated CVT 393Figure 7. Comparison of shifting speed, Ides model vs. measurement.been obtained using the inverse Ide model:kr(rcvt) =rcvt|p|Fshift(11)In the denominator Fshiftis present, the value of which can become (close to) zero. Obviously,the estimate is very sensitive for errors in Fshiftwhen its value is small. The dominant dis-turbances in Fshiftare caused by high-frequency pump generated pressure oscillations, whichdo not affect the ratio (due to the low-pass frequency behavior of unmodeled variator pulleyinertias). The standard deviation of the pressure oscillations and other high-frequency distur-bances has been determined applying a high-pass Butterworth filter to the data of Fshift.Toavoid high-frequency disturbances in Fshiftblurring the estimate of kr, estimates for values ofFshiftsmaller than at least three times the disturbances standard deviation have been ignored(these have been plotted as grey dots in figure 6), whereas the other points have been plottedas black dots. The white line is the resulting fit of this data. The few points with negative valuefor krhave been identified as local errors in the map of .To validate the quality of Ides model, the shifting speed rcvt, recorded during a road exper-iment, is compared with the same signal predicted using the model. Model inputs are thehydraulic pulley pressures (pp, ps) and pulley speeds (p, s) together with the estimatedprimary pulley torque (Tp). The result is depicted in figure 7. The model describes the shiftingspeed well, but for some upshifts it predicts too large values. This happens only for high CVTratios, i.e. rcvt 1.2, where the data of is unreliable due to extrapolation (see figure 5).3. The hydraulic systemThe hydraulic part of the CVT (see figure 3) consists of a roller vane pump directly connectedto the engine shaft, two solenoid valves and a pressure cylinder on each of the moving pulley394 M. Pesgens et al.sheaves. The volume between the pump and the two valves including the secondary pulleycylinder is referred to as the secondary circuit, the volume directly connected to and includingthe primary pulley cylinder is the primary circuit. Excessive flow in the secondary circuitbleeds off toward the accessories, whereas the primary circuit can blow off toward the drain.All pressures are gage pressures, defined relative to the atmospheric pressure. The drain is atatmospheric pressure.The clamping forces Fpand Fsare realized mainly by the hydraulic cylinders on the move-able sheaves and depend on the pressures ppand ps. As the cylinders are an integral part of thepulleys, they rotate with an often very high speed, so centrifugal effects have to be taken intoaccount and the pressure in the cylinders will not be homogeneous. Therefore, the clampingforces will also depend on the pulley speeds pand s. Furthermore, a preloaded linear elasticspring with stiffness kspris attached to the moveable secondary sheave. This spring has toguarantee a minimal clamping force when the hydraulic system fails. Together this results inthe following relations for the clamping forces:Fp= Ap pp+ cp 2p(12)Fs= As ps+ cs 2s kspr ss+ F0(13)where cpand csare constants, whereas F0is the spring force when the secondary moveablesheave is at position ss= 0. Furthermore, Apand Asare the pressurized piston surfaces. Inthe hydraulic system of figure 3, the primary pressure is smaller than the secondary pressure ifthere is an oil flow from the secondary to the primary circuit. Therefore, to guarantee that in anycase the primary clamping force can be up to twice as large as the secondary clamping force,the primary piston surface Apis approximately twice as large as the secondary surface As.It is assumed that the primary and the secondary circuit are always filled with oil of constanttemperature and a constant air fraction of 1%. The volume of circuit ( = p, s) is given by:V= V,min+ A s(14)V,minis the volume if s= 0 and Ais the pressurized piston surface.The law of mass conservation, applied to the primary circuit, combined with equation (14),results in:oil Vppp= Qsp Qpd Qp,leak Qp,V(15)Qspis the oil flow from the secondary to the primary circuit, Qpdis the oil flow from theprimary circuit to the drain, Qp,leakis the (relatively small) oil flow leaking through narrowgaps from the primary circuit and Qp,Vis the oil flow due to a change in the primary pulleycylinder volume. Furthermore, oilis the compressibility of oil. The oil flow Qspis given by:Qsp= cf Asp(xp) radicalBigg2|ps pp|sign(ps pp) (16)where cfis a constant flow coefficient and is the oil density. Asp, the equivalent valve openingarea for this flow path, depends on the primary valve stem position xp. Flow Qpdfollows from:Qpd= cf Apd(xp) radicalBigg2 pp(17)Here, Apdis the equivalent opening area of the primary valve for the flow from primary circuitto the drain. The construction of the valve implies that Asp(xp) Apd(xp) = 0 for all possible xp.Hydraulically actuated CVT 395Flow Qp,leakis assumed to be laminar with leak flow coefficient cpl, so:Qp,leak= cpl pp(18)The flow due to a change of the primary pulley cylinder volume is described by:Qp,V= Apsp(19)with spgiven by equation (4).Application of the law of mass conservation to the secondary circuit yieldsoil Vs ps= Qpump Qsp Qsa Qs,leak Qs,V(20)The flow Qpump, generated by the roller vane pump, depends on the angular speed eof theengine shaft, on the pump mode m (m = SS for single sided and m = DS for double sidedmode), and the pressure psat the pump outlet, so Qpump= Qpump(e,ps,m). Qsais the flowfrom the secondary circuit to the accessories and Qs,leakis the leakage from the secondarycircuit. Flow Qsais modeled as:Qsa= cf Asa(xs) radicalBigg2|ps pa|sign(ps pa) (21)where Asa, the equivalent valve opening of the secondary valve, depends on the valve stemposition xs. The laminar leakage flow Qs,leakis given by (with flow coefficient csl):Qs,leak= csl ps(22)The flow due to a change of the secondary pulley cylinder volume is:Qs,V= Asss(23)with ssaccording to equation (3).The accessory circuit contains several passive valves. In practice, the secondary pressurepswill always be larger than the accessory pressure pa, i.e. no backflow occurs. The relationbetween paand psis approximately linear, sopa= ca0+ ca1 ps(24)with constants ca0 0 and ca1 (0, 1).Now that a complete model of the pushbelt CVT and its hydraulics is available, the controllerand its operational constraints can be derived.4. The constraintsThe CVT ratio controller (in fact) controls the primary and secondary pressures. Severalpressure constraints have to be taken into account by this controller:1. the torque constraints p p,torqueto prevent slip on the pulleys;2. the lower pressure constraints p p,lowto keep both circuits filled with oil. Here, fairlyarbitrary, pp,low= 3 bar is chosen. To enable a sufficient oil flow Qsato the accessorycircuit, and for a proper operation of the passive valves in this circuit it is necessary that396 M. Pesgens et al.Qsais greater than a minimum flow Qsa,min. A minimum pressure ps,lowof 4 bar turnsout to be sufficient;3. the upper pressure constraints p p,max, to prevent damage to the hydraulic lines,cylinders and pistons. Hence, pp,max= 25 bar, ps,max= 50 bar;4. the hydraulic constraints p p,hydto guarantee that the primary circuit can bleed off fastenough toward the drain and that the secondary circuit can supply sufficient flow towardthe primary circuit.The pressures pp,torqueand ps,torquein constraint 1 depend on the critical clamping forceFcrit, equation (5). The estimated torqueTpis calculated using the stationary engine torquemap, torque converter characteristics and lock-up clutch mode, together with inertia effects ofthe engine, flywheel and primary gearbox shaft. A safety factor ks= 0.3 with respect to theestimated maximal primary torqueTp,maxhas been introduced to account for disturbances onthe estimated torqueTp, such as shock loads at the wheels. Then the pulley clamping force(equal for both pulleys, neglecting the variator efficiency) needed for torque transmissionbecomes:Ftorque=cos() (|Tp|+ksTp,max)2 Rp(25)Consequently, the resulting pressures can be easily derived using equations (12) and (13):pp,torque=1ApparenleftBigFtorque cp 2pparenrightBig(26)ps,torque=1AsparenleftbigFtorque cs 2s kspr ss F0parenrightbig(27)Exactly the same clamping strategy has been previously used by ref. 3 during test standefficiency measurements of this gearbox and test vehicle road tests. No slip has been reportedduring any of those experiments. As the main goal of this work is to an improved ratio trackingbehavior, the clamping strategy has remained unchanged.A further elaboration of constraints 4 is based on the law of mass conservation for theprimary circuit. First of all, it is noted that for this elaboration the leakage flow Qp,leakandthe compressibility term oil Vpppmay be neglected because they are small compared tothe other terms. Furthermore, it is mentioned again that the flows Qspand Qpdcan never beunequal to zero at the same time. Finally, it is chosen to replace the rate of ratio change rcvtby the desired rate of ratio shift rcvt,d, that is specified by the hierarchical driveline controller.If rcvt,d0 andQsp= 0. Constraint 4 with respect to the primary pulley circuit then results in the followingrelation for the pressure pp,hyd:pp,hyd=oil2parenleftbiggAp p max(0, rcvt,d)cf Apd,maxparenrightbigg2(28)where Apd,maxis the maximum opening of the primary valve in the flow path from the primarycylinder to the drain.In a similar way, a relation for the secondary pulley circuit pressure ps,hydin constraint 4can be derived. This constraint is especially relevant if rcvt 0, i.e. if the flow Qspfrom thesecondary to the primary circuit has to be positive and, as a consequence, Qpd= 0. This thenHydraulically actuated CVT 397results in:ps,hyd= pp,d+oil2parenleftbiggAp p max(0, rcvt,d)cf Asp,maxparenrightbigg2(29)where Asp,maxis the maximum opening of the primary valve in the flow path from the secondaryto the primary circuit.For the design of the CVT ratio controller it is advantageous to reformulate to constraintsin terms of clamping forces instead of pressures. Associating a clamping force F,with thepressure p,and using equations (12) and (13) this results in the requirement:F,min F F,max(30)with minimum pulley clamping forces:F,min= max(F,low,F,torque,F,hyd) (31)5. Control designIt is assumed in this section that at each point of time t, the primary speed p(t), the ratio rcvt(t),the primary pressure pp(t) and the secondary pressure ps(t) are known from measurements,filtering and/or reconstruction. Furthermore, it is assumed that the CVT is mounted in avehicular driveline and that the desired CVT ratio rcvt,d(t) and the desired rate of ratio changercvt,d(t) are specified by the overall hierarchical driveline controller. This implies, for instance,that at each point of time the constraint forces can be determined.The main goal of the local CVT controller is to achieve fast and accurate tracking of thedesired ratio trajectory. Furthermore, the controller should also be robust for disturbances. Animportant subgoal is to maximize the efficiency. It is quite plausible (and otherwise supportedby experiments, 3) that to realize this sub-goal the clamping forces Fpand Fshave to be assmall as possible, taking the requirements in equation (30) into account.The output of the ratio controller is subject to the constraints of equation (31). The constraintsF F,mincan effectively raise the clamping force setpoint of one pulley, resulting in anundesirable ratio change. This can be counteracted by raising the opposite pulleys clampingforce as well, using model-based compensator terms in the ratio controller. Using Ides model,i.e. using equation (10), expressions for the ratio change forces Fp,ratioand Fs,ratio(figure 8)can be easily derived:Fp,ratio= Fshift,d+ Fs,min(32)Fs,ratio=Fshift,d+ Fp,min(33)where Fshift,dis the desired shifting force, basically a weighted force difference betweenboth pulleys. As explained earlier, depends on primes, which in turn depends on Fs. This is animplicit relation (Fs,ratiodepends on Fs), which has been tackled by calculating from pressuremeasurements.It will now be shown that at each time, one of the two clamping forces is equal to F,min,whereas the other determines the ratio. Using equations (30), (32) and (33) the desired primary398 M. Pesgens et al.Figure 8. Ratio controller with constraints compensationand secondary clamping forces Fp,dand Fs,dare given by:Fp,d= Fp,ratioFs,d= Fs,minbracerightBiggif Fshift,d+ Fs,minFp,min(34)Fp,d= Fp,minFs,d= Fs,ratiobracerightBiggif Fshift,d+ Fs,minFp,minFp,min Fs,ratio= Fshift,dif Fshift,d+ Fs,minFp,max Fs,ratioFs,max(40)If either pressure saturates (pp= pp,maxor ps= ps,max), the shifting speed error inevitablybecomes large. The anti-windup algorithm ensures stability, but the tracking behavior willdeteriorate. This is a hardware limitation which can only be tackled by enhancing the variatorand hydraulics hardware. The advantage of a conditional anti-windup vs. a standard (linear)algorithm is that the linear approach requires tuning for good performance, whereas the con-ditional approach does not. Furthermore, the performance of the conditional algorithm closelyresembles that of a well-tuned linear mechanism.6. Experimental resultsAs the CVT is already implemented in a test vehicle, in-vehicle experiments on a rollerbench have been performed to tune and validate the new ratio controller. To prevent a non-synchronized operation of throttle and CVT ratio, the accelerator pedal signal (see figure 1) hasbeen used as the input for the validation experiments. The coordinated controller will track themaximum engine efficiency operating points. A semi kick-down action at a cruise-controlledspeed of 50 km/h followed by a pedal back out has been perfo
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本文标题:QD10t-31.5m箱形双梁桥式起重机起重小车设计【7张CAD图纸和说明书】
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