資源目錄里展示的全都有，所見即所得。下載后全都有,請放心下載。原稿可自行編輯修改=【QQ：401339828 或11970985 有疑問可加】 附 錄A ABSTRACT High clamping force levels reduce the efficiency of the Continuously Variable Transmission (CVT). However, high clamping force levels are necessary to prevent slip between the belt and the pulleys. If a small amount of slip is allowed, the clamping force level can be reduced. To achieve this, slip in a CVT is investigated. From measurements on an experimental setup, Traction curve data and efficiency measurements are derived. A model describing slip in a CVT is verified using measurements with a belt with increased play. It is found that small amounts of slip can be controlled in a stable way on the setup. The traction curve was mostly dependent on the CVT ratio. Efficiency is found to be highest for 1 to 2% slip depending on the ratio. The model is in reasonable agreement with the measurements. 1. Introduction Applying a Continuously Variable Transmission (CVT) in an automotive driveline has several advantages. A CVT can operate at a wider range of transmission ratios, therefore the engine can be operated more efficiently than with a stepped transmission. Also, a CVT does not interrupt the torque transmission when shifting. This gives a more smooth ride than a stepped transmission does. A V-belt based Continuously Variable Transmission uses a belt or a chain to transmit torque from a driving side to a driven side by means of friction. The layout of the CVT and the V-belt are shown in figure 1. The variator consists of two pulleys which are wedge shaped. By changing the position of the pulleysheaves the ratio of the CVT can be adjusted. The V-belt consists of blocks which are held together by two rings that in turn exist of a set of bands. To achieve torque transmission sufficiently high clamping force levels are needed to prevent slip in the variator. Because the torque level is not exactly known at all times, since no torque sensor is used due to cost considerations, a safe clamping force level based on the maximum possible load is maintained at all times. This safety level is based upon assumed maximum shockload levels from the road, like bumps, and the engine torque. In order to maintain these safety levels higher clamping force levels are maintained then needed. Higher clamping force levels cause more losses in the CVT. These losses are caused by increases in power consumed by the hydraulic pump, by increases in the losses due to slip in the belt if a pushbelt is used, and by increases in deformation in the belt and in the pulleys. Furthermore wear is increased and fatigue life is reduced. In order to reduce these clamping force levels a method is needed to detect slip in the variator fast enough to prevent slip from reaching destructive levels. A method to detect and control slip is therefore needed. In this paper measurements are presented of the traction curve in a V-belt CVT. Figure 1. Layout of a CVT and a metal pushbelt 2. Traction curve The V-belt type CVT utilizes friction to transmit power from the primary pulley to the secondary pulley. The traction curve is the dimensionless relationship between transmitted torque and the slip. The maximum input torque that can be transmitted by the CVT is dependent on the applied clamping force. The traction coefficient is therefore chosen to be a dimensionless value. The traction coefficient μ is defined as: （1） In which represents the input torque, represents the secondary running radius of the belt on the pulley, represents the secondary clamping force and is the pulley wedge angle. Figure 2. CVT torque transmission scheme The second variable in the traction curve is the slip in the variator. Slip is defined as: （2） Where is the angular speed of the secondary axle, is the angular speed of the primary axle and is the geometrical ratio, which is defined by: （3） is the running radius on the primary pulley. 2.1 Tangential slip Slip is defined in equation 2. When the CVT transmits power a certain amount of slip can be measured almost linear with the applied torque. This is called the microslip regime of the CVT, because traction is still increasing in this regime with increasing slip. The microslip is caused by gaps between the blocks on the idle part of the driving pulley as shown in figure 3. On the driving pulley an idle arc exists where no slip occurs. Also an active arc exists (see figure 2), where slip occurs relative to the total play in the belt and the active arc length. However, when the maximum torque capacity of the CVT is reached slip will increase dramatically. This situation, macroslip, is not stable during normal operation of the CVT, because the traction coefficient decreases with increased slipspeed. It is assumed that the total gap dt is evenly distributed along the idle arc of the driving pulley. The traction Figure 3. Gaps in the belt curve (figure 5) shows that torque transmission increases almost linearly with an increase in slip, until a certain maximum torque is reached. dt can be estimated by adding an initial gap do to the increase in belt length due to the internal stresses in the bands and a decrease in length of the blocks due to the compressive forces. （4） To calculate the slip caused by these gaps we can use the following equations: （5） （6） In equation 5, a is the idle arc, d is the width of a belt element and dt is the total gap between the elements in the belt. To calculate the amount of slip the total gap dt has to be known. This effect has an influence on the traction coefficient in the macroslip regime. When macroslip occurs the traction will decrease with increasing slip. The Stribeck effect is modelled using equation 9. （7） （8） （9） Equation 7 gives a value for the friction caused by viscous friction component. Equation 8 gives a value for the coulomb friction component. a0,1, c0 and v1 are coefficients which can be chosen to match the measured values. With these equations we can derive slip and traction from measured data as shown in section 4. With Asayama [1995] we can obtain the tension and compression force distribution needed to calculate the lengthening of the belt. Also, we can calculate the idle arc from this model. From the idle arc, the length of the belt and the initial gap we can calculate an estimate for slip in the belt for a given load. 2.2 Radial slip Not only slip in tangential direction occurs, but also slip in radial direction. The first reason for radial slip is spiral running. When the belt runs along the arc of contact the radius at which it runs is not constant. This effect is caused by pulley deformation. One type of deformation is the bending of the axle between both pulley sheaves. The belt is not fully wrapped around the pulley, therefore the resulting normal force of the blocks on the pulley is not axial. This causes a bending moment in the axle. A second effect is the bending of the pulley itself. This effect is mostly dependent on the local normal force exerted on the pulley by the blocks. This effect is small when the belt is running on a small running radius, but on a large running radius this effect is significant. The second reason for slip in radial direction is due to shifting. When the CVT is shifted to a different transmission ratio, radial slip is forced. This is done by changing the clamping force ratio. The amount of radial slip that is forced depends on the shifting speed and the (primary) angular speed. 3. Experimental setup In the experiments the geometric cvt ratio is fixed and the clamping forces are constant, the traction coefficient then depends only on the slip in the system. The traction curve can be constructed from output torque and slip measurements. The test rig motors deliver a maximum torque of 298 Nm with a maximum speed of 525 rad/s. Both motors are equipped with a Heidenhain ERN1381 incremental rotary encoder with 2048 pulses/rev. The torque at both sides is measured using a HBM T20WN torque sensor. The maximum allowable torque is 200 Nm with speeds up to 1050 rad/s. A separate hydraulic unit is used to provide the required flow and pressure for the clamping forces. Figure 4 gives a schematic overview of the experimental setup. 4. Experimental results The geometric ratio of the CVT was fixed during the experiments using a so-called ratio ring and the limits of the primary pulley. This ratio ring limit the movement of the pulley. Primary and secondary pressure was held constant (clamping forces were held constant) during the experiments. Figure 4. Experimental setup 4.1 Traction coefficient The traction coefficient was measured at different ratios, at different primary speeds and at different pressures. In figure 6 and 7 can be seen that the traction coefficient depends little on primary speed or secondary clamping pressure, but mostly on the transmission ratio, as can be seen in figure 5. An increase in clamping force causes more slip (see figure 8). This is caused by an increase in tension in the bands and therefore in an increase in length of the belt. This causes the play to increase. Figure 5. Traction coefficient at 300rad/s, ratio low(0.4), Medium (1.1) and overdrive (2.26) 4.2 Efficiency The efficiency depends on pressure and on ratio. From figure 12 can be seen that an increase in pressure causes a decrease in efficiency. This effect is caused by the internal friction in the belt. Slip between the blocks and the bands also causes a strong dependency on ratio (see figure 9). Efficiency is clearly higher in medium than in overdrive or low. In medium no slip occurs between the blocks and the bands, but in overdrive or low the bands slip over the blocks. At high clamping levels this effect is greater, because the normal forces acting between the blocks and the bands increase linearly with an increase in clamping level. From figure 10 and 11 can be seen that input speed also has an influence on efficiency. Figure 6. Traction coefficient in overdrive, ws = 150,225,300 Figure 7. Traction coefficient in low, wp =150,225,300 Figure 8. Traction coefficient for resp. 5bar and 8bar secondary clamping pressure From figure 10 and 11 can be seen that input speed also has an influence on efficiency. 4.3 Play The microslip region is dependent on play in the belt. An experiment has been carried out with a belt with increased play. One block was taken out of the belt. The performance of the belt was measured with a total gap of 1.8mm. The cumulative gap in the belt was 0.3mm in the other experiments. A significant difference is measured in the LOW ratio of the CVT. In figure 4.3 the traction curve is shown for the low ratio of the CVT for the belt with increased play. Also the result of the numerical model is shown in figure 4.3. The results for overdrive show that in overdrive there is no significant change in the traction curve, see figure 4.3. However, the model is less consistent with the tractioncurve in overdrive than in low. Figure 9. Efficiency at 300rad/s, ratio low(0.4), Medium (1.1) and overdrive (2.26) Figure 10. Efficiency in overdrive, ws =150,225,300 Figure 11. Efficiency in low, = 150,225,300 Figure 12. Traction coefficient for resp. 5bar and 8bar secondary clamping pressure Figure 13. Effect of play in the belt, wp = 30rad/s, in low, with increased gap (1.8mm) Figure 14. Effect of play in the belt, wp = 30rad/s, in overdrive, with increased gap (1.8mm) 5. Conclusion The traction curve is mostly ratio dependent. This can be explained with the shown model as explained in section 4. Transmission efficiency is dependent on applied pressure, input speed and the CVT ratio. Gaps between the blocks of the belt cause at least part of the tangential slip of the belt. This was confirmed by the experiment with increased play in the belt. The consistency of the model is better in low than in overdrive. Future research will be directed at controlling slip in the CVT. This can enhance the efficiency of the CVT. 附 錄B 各級高夾緊力降低了無級變速器（CVT）的效率。然而，各級高夾緊力之間的必要措施，防止金屬帶和滑輪滑。如果滑少量是允許的，夾緊力水平可以降低。要做到這一點，在無級變速器滑動進行了研究。從上一個實驗裝置測量，牽引效率測量數據和曲線推導。一個模型描述無級變速器滑動驗證使用具有增加播放帶測量。研究發(fā)現，少量的滑可在道路上設置穩(wěn)定控制。牽引曲線主要是依賴于CVT的比率。效率是發(fā)現1％至2％的最高比例滑倒而定，該模型與測量合理的協(xié)議。 1. 簡介 應用在汽車傳動系統(tǒng)無級變速器（CVT）有幾個優(yōu)點。無級變速器可以工作在更廣泛的傳動比，因此該引擎可以使用，其傳輸效率比階梯。另外，CVT的不中斷換擋時的扭矩傳遞。這給出了一個比一個更平穩(wěn)的傳輸并加強。阿V帶無級變速器的使用金屬帶或鏈條傳送通過摩擦意味著從驅動側的扭矩到從動側。該無級變速器和V帶的布局見圖1。該變速器由兩個滑輪是楔形。通過改變位置的CVT的比例可以調整。 V型帶，其中包括分別由兩個環(huán)一起，在樂隊依次設置存在的塊。為了實現足夠高的扭矩傳遞夾緊力水平是需要防止變速器滑。由于轉矩是不完全知道在任何時候，因為沒有采用扭矩傳感器由于成本的考慮，一個安全級別夾緊力最大的可能是維持負載為基礎在任何時候。這是基于安全等級最高的假定像顛簸道路上，和發(fā)動機扭矩水平。為了保持這些安全級別較高的夾緊力水平維持不變，然后需要。夾緊力水平造成的CVT更多的損失。這些損失是由由液壓泵消耗功率提高造成的損失中帶滑，如果在一個帶使用的增加，并在帶變形和滑輪增加。此外磨損增大，疲勞壽命降低。為了減少這些夾緊力的方法檢測，需要足夠快的變速器，防止破壞性的水平失誤達到的水平。一個方法來檢測和控制，因此需要滑。本文介紹了測量中的V帶無級變速牽引曲線。 圖1 金屬帶式無級變速器布置 2. 牽引曲線 V型帶式無級變速器采用了摩擦，從主滑輪傳送到輔助電源滑輪。牽引曲線之間傳遞扭矩和滑量綱關系。最大輸入扭矩，可以通過發(fā)送的CVT的夾緊力的應用而定。牽引系數因此選擇是一個無量綱值。牽引系數μ的定義為： （1） 其中表示輸入扭矩，代表著對金屬帶輪二次運行半徑，代表了二次夾緊力，是滑輪楔角。 圖2 CVT的扭矩傳輸方案 牽引曲線中的第二個變量是在變速器滑移。 滑移的定義為： （2） 是次要軸角速度，是主軸角速度，是幾何比例，將其定義為： （3） 正在運行的主滑輪半徑。 2.1 切向滑移 滑移是指在公式2。當無級變速器傳遞動力滑移一定量的可測與施加的扭矩幾乎呈線性關系。這就是所謂的CVT的microslip政權，因為在此牽引仍隨滑移區(qū)增加。該microslip是由塊之間的間隙對傳動滾筒閑置部分，如圖3 所示。在主動輪弧存在其中一個空閑無滑移發(fā)生。也是一個積極的弧存在（參見圖2），其中發(fā)生相對滑移，在帶和總發(fā)揮積極的弧長。然而，當CVT的最大扭矩達到防滑能力將顯著增加。這種情況，macroslip，是不是在本無級變速器的正常運行穩(wěn)定，因為牽引系數降低與增加滑移率。據推測，總差距是均勻分布在主動輪的閑置弧線處。牽引曲線（圖5）顯示，傳遞扭矩增大幾乎呈線性增加，在滑動，直到達到一定的最大扭矩?？梢酝ㄟ^添加一個初始間隙做了帶長度的增加，估計由于內部應力的樂隊，并在適當的塊長度的壓縮力下降。 （4） 圖3 空白帶 計算的滑動造成這些差距的原因我們可以用以下的方程: （5） （6） 公式5是一個空閑的弧線，D是一個帶元素的寬度和是在金屬帶之間的元素的總差距。要計算總的差距滑的金額為已知。這種效應有一個關于在宏觀滑移政權牽引系數的影響。當宏觀滑移發(fā)生的牽引滑移的增加會降低。效果是模仿的摩擦模型使用公式9。 （7） （8） （9） 公式7給出了由粘性摩擦元件所造成的摩擦系數。公式8給出了庫倫摩擦組件的值。為1，和的是可以選擇匹配的測量值系數。有了這些方程，我們可以從測量數據下滑牽引，在圖4所示，我們可以得到的張力和壓縮力分布計算所需的金屬帶延長。另外，我們可以從這個模型計算出空閑的弧線。從閑置的弧線，金屬帶的長度和初始差距，我們可以計算出在金屬帶承保給定負載的估計。 2.2 徑向滑動 不僅在切線方向發(fā)生滑動，而且滑徑向方向。對于第一個原因是徑向滑動螺旋運行。當沿帶的接觸半徑，在它運行的弧線運行的不是恒定的。這種效應是由滑輪變形。一類是變形的兩滑輪軸彎曲。金屬帶是不完全的滑輪包左右，因此導致正常的區(qū)塊隊在滑輪是不是軸向，這會導致軸彎矩。 第二個效應是金屬帶輪本身彎曲。這種影響主要是對當地正常的由塊滑輪施加力而定。這種效果是小的時候帶運行在一個小半徑運行，但運行在一個大半徑這種影響是顯著。在徑向方向的滑動的第二個原因是由于轉移。當無級變速器被轉移到一個不同的傳動比，徑向滑移是被迫的。這是通過改變夾緊力比。對徑向滑動量是被迫依賴于移動速度和（主）角的速度。 3. 實驗裝置 在無級變速器的幾何比例是固定的實驗和夾緊力是恒定的，那么牽引系數只依賴于在系統(tǒng)中溜走。牽引曲線可以構造從輸出轉矩和轉差的測量。該試驗臺電機提供了最大的525拉德/速度為298牛米的最大扭矩秒這兩種發(fā)動機都配備了海德漢ERN1381增量與2048脈沖/轉的旋轉編碼器。在雙方的扭矩測量使用HBM的T20WN扭矩傳感器。允許的最大轉矩與速度高達200至1050弧度/ s的牛一個獨立的液壓裝置是用來提供所需流量和壓力的夾緊力。圖4給出了實驗裝置原理圖概述。 4. 實驗結果 幾何比例的CVT被固定在這個實驗中使用一個所謂的比率的極限環(huán)和主要滑輪。這一比率環(huán)限制的運動滑輪。初級和中級壓力保持不變(即夾緊力常數)舉行了在實驗。 液壓裝置 電機2 編碼器 扭矩傳感器 電機1 圖4 實驗裝置 圖5 牽引在300rad/ s時，比低（0.4）系數，中（1.1）和高速（2.26） 4.1 牽引系數 測量了牽引系數在不同比例、不同主要的速度和在不同的壓力。圖6、7,可以看出牽引系數取決于原發(fā)性或繼發(fā)性小速度夾緊壓力,但主要在傳動比,從中我們可以看到圖5。夾緊力的增加會引起更多的滑移(見圖7)。這是由于增加的緊張局勢,因此在樂隊的長度增加金屬帶。這使發(fā)揮增加。 4.2 效率 效率取決于壓力和比例。12從圖可以看出,增加減少壓力會使效率。這種效應是由于內部摩擦帶中?；瑝K和繩索之間也會有強烈的依賴比(見圖9)。顯然是更高效率中比在超載或低。在中等無滑塊之間發(fā)生的,但在超載樂隊,樂隊或低滑動的街區(qū)。這種效應在高夾緊程度更大,因為正常的有力作用之間呈線性增長,帶塊增加夾緊的水平。從圖10和11可以看出，輸入速度也對效率的影響。 圖6 牽引過載系數，是=150225300 圖7 牽引系數低,wp= 150225300 圖8 牽引系數8桿和5桿對照 從圖10和11可以看出，輸入速度也對效率有一定的影響。該微滑地區(qū)依帶上發(fā)揮而定。實驗已經進行了一個增加播放帶。一個塊被取出來的金屬帶。金屬帶的性能是衡量一個總落差為1.8mm。帶中的累積差距是在其他實驗0.3毫米。一個重要的區(qū)別是在測量CVT的低比率。在圖4.3的牽引曲線所示為用于增加播放帶無級變速器低的比例。也是數值模式結果顯示在圖4.3。對超載超速結果顯示，在沒有任何重大變化曲線的牽引，見圖4.3。然而，該模型比在低中超速不太一致。 圖9 效率在300rad/ s時，比低（0.4），中（1.1）和高速（2.26） 圖10 在高速的效率，是=15022.530萬 圖11 低工作效率，= 150225300 圖12 牽引系數8桿和5桿中級對比 圖13 帶的影響,工作wp=30rad / s低速時,增加了差距(1.8毫米) 圖14 帶的影響,工作wp=30rad / s超載時,增加了差距(1.8毫米) 5. 結論 牽引曲線主要是比依賴。這可以解釋模型解釋顯示第四節(jié)。傳動效率是依賴于應用壓力、輸入速度和無級變速比。塊體的間隙帶造成至少部分的切向滑移帶。實驗證實了這則以增加參加帶。該模型的一致性具有更好的低比高峰。未來的研究將針對控制滑在無級變速。這能提高CVT的效率。 16