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QD10t-31.5m箱形双梁桥式起重机起重小车的设计【机械毕业设计含7张CAD图+说明书论文1.5万字56页】

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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


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 & Francishttp:/www.tandf.co.uk/journalsDOI: 10.1080/00423110500244088nts388 M. Pesgens et al.Figure 1. Hierarchical powertrain control.The CVT and its hydraulic actuation system are depicted in figures 2, 3. The hydraulicsystem not only has to guarantee good tracking behavior of the CVT but also has to realizeclamping forces that, on the one hand, are high enough to prevent belt slip but, on the otherhand, are as low as possible to maximize the transmission efficiency and to reduce wear. Inpractice, the clamping forces levels are kept at levels that avoid belt slip at all times, whilestill maintaining an acceptable degree of transmission efficiency.The main focus of this paper is on the ratio control of the CVT, using the hydraulic actuationsystem of figure 3. The presented control concept is based on the work of 3, 4. It enablestracking of the ratio setpoint, while guaranteeing at least one of the two pulley pressuresetpoints to be equal to its lower constraint. Even though the controller effectively changesfrom controlling one of the two pressures to the other, no actual switching between differentcontrollers takes place. Among the approaches seen in the literature, some incorporate aswitching algorithm 3, 5, whereas others control only one of the two (usually the primary)pressures 6, 7. Although the former approach cannot guarantee one of the two pressures tobe equal to its lower constraint, the latter cannot explicitly prevent the uncontrolled pressureto stay above its lower constraint.The rest of this paper is organized as follows. First, a mathematical model is derived forthe mechanical part of the CVT in section 2. Next, in section 3, the hydraulic part is modeled.The physical constraints, imposed by the hydraulic system, are discussed in section 4. Theseconstraints are taken into account by the CVT ratio controller, that is developed in section 5Figure 2. Variator.ntsHydraulically actuated CVT 389Figure 3. Variator with hydraulic system.and is based on the earlier derived models for the mechanical and the hydraulic CVT parts.The tracking performance of this controller is experimentally evaluated in section 6. Finally,section 7 gives some concluding remarks.2. The pushbelt CVTThe CVT (figure 2) considered here is equipped with a Van Doorne metal pushbelt. Thisbelt consists of a large number (around 350) of V-shaped steel block elements, held togetherby a number (between 9 and 12) of thin steel tension rings. The belt runs on two pulleys,namely the primary pulley at the engine side and the secondary pulley at the wheel side. Eachpulley consists of one axially fixed and one moveable sheave, operated by means of a hydrauliccylinder. The cylinders can be pressurized, generating axial forces (clamping forces or thrusts)on the belt, necessary for transmission of torque (without macro-slip of the belt) and for ratiochange. Here the distinction is made between micro-slip, needed for torque transfer betweenbelt and pulleys, and macro-slip, which should be avoided at all times for its negative effecton efficiency and especially the risk of severe belt and pulley wear 8.The bounded transmission ratio rcvtrcvt,LOW,rcvt,OD is defined here as the ratio ofsecondary pulley speed sover primary pulley speed p, so:rcvt=sp(1)In deriving the variator model, it has been assumed that the pulleys are rigid and perfectlyaligned, and that the V-shaped blocks are rigid and the steel rings are inextensible. The beltis assumed to run in perfect circles on the pulleys. Further, it has been assumed that theclamping forces are large enough to prevent macro belt slip. The effects of micro-slip arerelatively small with respect to the ratio change behavior of the CVT, and are, therefore,neglected in the model. The power transmission between the belt and the pulleys is modeledas Coulomb friction (which is assumed in the majority of CVT variator research 3).nts390 M. Pesgens et al.Using these assumptions, the running radii Rpand Rsof the belt on the primary andsecondary pulleys are functions of the ratio rcvtonly and are related by:Rp= rcvt Rs(2)The axial position s( = p for the primary pulley, = s for the secondary one) of themoveable pulley sheave of pulley is also completely determined by the ratio rcvt. Denotingthe taper angle of the conical sheaves by (see figure 4), it is easily seen that sis given by:s= 2 tan() (R R,min) (3)Subscript max (or min) implies the maximum (or minimum) value possible, e.g. Xmax=max(X), unless stated otherwise. Differentiation with respect to time yields the axial velocitysof the moveable sheave of pulley s= (rcvt) rcvt(4)where the function follows from the geometry of the variator.Assuming that the radial friction force component between the pulley and the belt is zero,the critical pulley clamping force (equal for both pulleys, neglecting the variators efficiency)is given by references 3, 5 (for pulley ):Fcrit=cos() |T|2 R(5)where Tis the net transmitted torque between belt and pulley and is the coulomb frictioncoefficient between pulleys and belt. The factor 2 appears, as there are two friction surfacesbetween pulley and belt.Radial forces between belt and pulleys can be mainly contributed to centrifugal forces andCoriolis forces. In the detailed thrust force model of ref. 9, it is reported that even if thesliding angle (and hence the friction force angle) between the belt path and the friction forcevector changes along the pulley circumference, its value converges rapidly towards values lessthan 10. As a result, the angle is assumed zero. The friction force angle would enter intoequation (5) as a multiplication factor cos(), which rapidly converges to 1 for small angles.For the choice of , a worst-case approach is applied. It is chosen as the maximum of thetraction curve (of which several have been presented in ref. 10), which is the point of transitionfrom micro-slip to macro-slip. The lowest value of all the maxima found in ref. 10, as wellas in ref. 8 (for both very similar variators) is 0.09, the value of that has been used here.Figure 4. Pulley sheave definitions.ntsHydraulically actuated CVT 391The torque ratio is the ratio of transmitted torque and maximally transmittable torquewithout belt slip for pulley :=TT,max=cos() T2 R F(6)As in a practical vehicle application a good estimate of the torques acting on the secondarypulley is not available, the following modified torque ratio is introduced:primes=cos() Tp2 Rp Fs(7)The estimated primary transmitted torqueTpcan be obtained from the dynamic drivelineequations together with engine and torque converter characteristics (also see section 4). Incase of a perfect torque estimation, i.e.Tp= Tp, it is easily seen that (using equation (6):primes=PpPs s(8)with transmitted power P= T . As it has been assumed that Pp= Ps, the modifiedtorque ratio becomes equal to the torque ratio for the secondary pulley.An important part of the model for the mechanical part of the CVT is the sub-model for therate of ratio change as a function of, for instance, the clamping forces. Sub-models of this typeare proposed, among others, by Guebeli et al. 11, Ide et al. 12, 13 and Shafai et al. 14.The blackbox model of Ide is preferred here, as it reasonably describes the results of a seriesof experiments with metal V-belt CVTs 3, 4.The steady state version of Ides model yields a relation for the primary clamping force Fpthat is required to maintain a given ratio rcvtwith a given secondary clamping force Fsand agiven primary torque Tp(through the modified torque ratio primes):Fp= (rcvt,primes) Fs(9)For obvious reasons, the quantity in equation (9) is called the thrust ratio. Some experimen-tally obtained results for this highly non-linear function of the CVT ratio rcvtand the torqueratio primesare given in figure 5.For instationary situations, Ides model states that the rate of ratio change rcvtis a functionof the ratio rcvt, primary pulley speed p, clamping forces Fpand Fsand torque ratio primes:rcvt= kr(rcvt) |p|Fshift; 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 krhasnts392 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.ntsHydraulically 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 pulleynts394 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.ntsHydraulically 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
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