帶式輸送機外文翻譯---輸送帶的二維動態(tài)特性

1、<p>　　附錄 英文文獻及翻譯</p><p>　　The Two-Dimensional Dynamic Behavior of Conveyor Belts</p><p>　　3.1.1 NON LINEAR TRUSS ELEMENT</p><p>　　If only the longitudinal deformation of the be

2、lt is of interest then a truss element can be used to model the elastic response of the belt. A truss element as shown in Figure 2 has two nodal points, p and q, and four displacement parameters which determine the compo

3、nent vector x:</p><p>　　xT = [up vp uq vq]             (1)</p><p>　　For the in-plane motion of the truss element there are three indep

4、endent rigid body motions therefore one deformation parameter remains which describes</p><p>　　Figure 2: Definition of the displacements of a truss element</p><p>　　the change of length of the a

5、xis of the truss element [7]:</p><p>　　where dso is the length of the undeformed element, ds the length of the deformed element and ξ a dimensionless length coordinate along the axis of the element.</p>

6、;<p>　　Figure 3: Static sag of a tensioned belt</p><p>　　Although bending, deformations are not included in the truss element, it is possible to take the static influence of small values of the belt s

7、ag into account. The static belt sag ratio is defined by (see Figure 3):</p><p>　　K1 = δ/1 = q1/8T           (3)</p><p>　　where q is the distrib

8、uted vertical load exerted on the belt by the weight of the belt and the bulk material, 1 the idler space and T the belt tension. The effect of the belt sag on the longitudinal deformation is determined by [7]:</p>

9、<p>　　εs = 8/3 K²s              (4)</p><p>　　which yields the total longitudinal deformation of the non linear truss element:</

10、p><p>　　3.1.2 BEAM ELEMENT</p><p>　　Figure 4: Definition of the nodal point displacements and rotations of a beam element.</p><p>　　If the transverse displacement of the belt is being

11、of interest then the belt can be modelled by a beam element. Also for the in-plane motion of a beam element, which has six displacement parameters, there are three independent rigid body motions. Therefore three deformat

12、ion parameters remain: the longitudinal deformation parameter, ε1, and two bending deformation parameters, ε2 and ε3.</p><p>　　Figure 5: The bending deformations of a beam element</p><p>　　The b

13、ending deformation parameters of the beam element can be defined with the component vector of the beam element (see Figure 4):</p><p>　　xT = [up vp µp uq vq µq]       

14、   (5)</p><p>　　and the deformed configuration as shown in Figure 5:</p><p>　　3.2 THE MOVEMENT OF THE BELT OVER IDLERS AND PULLEYS</p><p>　　The movement of a belt is const

15、rained when it moves over an idler or a pulley. In order to account for these constraints, constraint (boundary) conditions have to be added to the finite element description of the belt. This can be done by using multi-

16、body dynamics. The classic description of the dynamics of multi-body mechanisms is developed for rigid bodies or rigid links which are connected by several constraint conditions. In a finite element description of a (def

17、ormable) conveyor belt, where</p><p>　　Figure 6: Belt supported by an idler.</p><p>　　There are two independent rigid body motions for an in-plane supported beam element therefore five deformati

18、on parameters remain. Three of them, ε1, ε2 and ε3, determine the deformation of the belt and are already given in 3.1. The remaining two, ε4 and ε5, determine the interaction between the belt and the idler, see Figure 7

19、.</p><p>　　Figure 7: FEM beam element with two constraint conditions.</p><p>　　These deformation parameters can be imagined as springs of infinite stiffness. This implies that:</p><p&

20、gt;　　ε4 = D4(x) = (rξ + u ξ)e2 - rid.e2 = 0</p><p>　　ε5 = D5(x) = (r ξ + uξ)e1 - rid.e1 = 0                (7)</p><p

21、>　　If during simulation ε4 > 0 then the belt is lifted off the idler and the constraint conditions are removed from the finite element description of the belt.</p><p>　　3.3 THE ROLLING RESISTANCE</p

22、><p>　　In order to enable application of a model for the rolling resistance in the finite element model of the belt conveyor an approximate formulation for this resistance has been developed, [8]. Components of

23、 the total rolling resistance which is exerted on a belt during motion three parts that account for the major part of the dissipated energy, can be distinguished including: the indentation rolling resistance, the inertia

24、 of the idlers (acceleration rolling resistance) and the resistance of the be</p><p>　　ft = fi + fa + fb                  (8)&l

25、t;/p><p>　　where fi is the indentation rolling resistance factor, fa the acceleration resistance factor and fb the bearings resistance factor. These components are defined by:</p><p>　　where Fz is

26、distributed vertical belt and bulk material load, h the thickness of the belt cover, D the idler diameter, Vb the belt speed, KN the nominal percent belt load, T the ambient temperature, mred the reduced mass of an idler

27、, b the belt width, u the longitudinal displacement of the belt, Mf the total bearing resistance moment and ri the internal bearing radius. The dynamic and mechanic properties of the belt and belt cover material

28、play an important role in the calculation of the</p><p>　　3.4 THE BELT'S DRIVE SYSTEM</p><p>　　To enable the determination of the influence of the rotation of the components of the drive sys

29、tem of a belt conveyor, on the stability of motion of the belt, a model of the drive system is included in the total model of the belt conveyor. The transition elements of the drive system, as for example the reduction b

30、ox, are modelled with constraint conditions as described in section 3.2. A reduction box with reduction ratio i can be modelled by a reduction box element with two displacement paramete</p><p>　　εred = Dred(

31、x) = iµp + µq = 0               (10)</p><p>　　To determine the electrical torque of an induction machine, the so-called two axi

32、s representation of an electrical machine is adapted. The vector of phase voltages v can be obtained from: </p><p>　　v = Ri + ωsGi + L ?i/?t             

33、;       (11)</p><p>　　In eq. (11) i is the vector of phase currents, R the matrix of phase resistance's, C the matrix of inductive phase resistance's, L the matrix of ph

34、ase inductance's and ωs the electrical angular velocity of the rotor. The electromagnetic torque is equal to:</p><p>　　Tc = iTGi             &#

35、160; (12)</p><p>　　The connection of the motor model and the mechanical components of the drive system is given by the equations of motion of the drive system:</p><p>　　where T is the torqu

36、e vector, I the inertia matrix, C the damping matrix, K the stiffness matrix and ø the angle of rotation of the drive component axis's.</p><p>　　To simulate a controlled start or stop procedure a fe

37、edback routine can be added to the model of the belt's drive system in order to control the drive torque.</p><p>　　3.5 THE EQUATIONS OF MOTION</p><p>　　The equations of motion of the total b

38、elt conveyor model can be derived with the principle of virtual power which leads to [7]:</p><p>　　fk - Mkl ?²x1 / ?t² = σ1Dik            

39、0;   (14)</p><p>　　where f is the vector of resistance forces, M the mass matrix and σ the vector of multipliers of Lagrange which may be interpret as the vector of stresses dual to the vector

40、 of strains ε. To arrive at the solution for x from this set of equations, integration is necessary. However the results of the integration have to satisfy the constraint conditions. If the zero prescribed strain compone

41、nts of for example e.g. (8) have a residual value then the results of the integration have to be corrected,</p><p>　　3.6 EXAMPLE</p><p>　　Application of the FEM in the desian stage of long belt

42、conveyor systems enables its proper design. The selected belt strength, for example, can be minimised by minimising, the maximum belt tension using the simulation results of the model. As an example of the features of th

43、e finite element model, the transverse vibration of a span of a stationary moving belt between two idler stations will be considered. This should be determined in the design stage of the conveyor in order to ensure reson

44、ance</p><p>　　The effect of the interaction between idlers and a moving belt is important in belt-conveyor design. Geometric imperfections of idlers and pulleys cause the belt on top of these supports to be

45、displaced, yielding a transverse vibration of the belt between the supports. This imposes an alternating axial stress component in the belt. If this component is small compared to the prestress of the belt then the belt

46、will vibrate in it's natural frequency, otherwise the belt's vibration will follow the </p><p>　　The results of simulation with the finite element model can be used to determine the frequency of tran

47、sverse vibration of a stationary moving belt span. This frequency is obtained after transformation of the results of the transverse displacement of the belt span from the time domain to the frequency domain using the fas

48、t fourier technique. Besides using the finite element model also an analytical approach can be used.</p><p>　　The belt can be modelled as a prestressed beam. If the bending stiffness of the belt is neglected

49、, the transverse displacements are small compared to the idler space, Ks << 1, and the increase of the belt length due to the transverse displacement is negligible compared to its initial length, the transverse vib

50、ration of the belt can be approximated by the following linear differential equation, also see Figure 5:</p><p>　　where v is the transverse displacement of the belt and c2 the wave speed of the transverse wa

51、ves defined by, [1]:</p><p>　　c2 = √g1/8Ks                  (16)</p><p>　　The first natural transverse f

52、requency of the belt span of Figure 5 can be obtained from eq. (16) if it is assumed that v(O,t)=v(l,t)=0:</p><p>　　where ß is the dimensionless speed ratio defined by:</p><p>　　ß = Vb

53、 / c2                  (18)</p><p>　　The frequency fb is different for each individual belt span since the belt tensi

54、on varies over the length of the conveyor. The excitation frequency of an idler which has a single eccentricity is equal to:</p><p>　　fi = Vb / πD           

55、        (19)</p><p>　　where D is the diameter of the idler. In order to design a resonance free belt support the idler space is subjected to the following condition:&l

56、t;/p><p>　　The results obtained with the linear differential equation (16) however are valid only for low values of the ratio ß. For higher values of ß, as is the case for high-speed conveyors or low

57、belt tensions, the non-linear terms in the full form of e.g. (16) become significant. Therefore numerical simulations using, the FEM model have been made in order to determine the ratio between the linear and the non-lin

58、ear frequency of transverse vibration of a belt span. These relations have been determined </p><p>　　The results for the transverse displacements were transformed to a frequency spectrum using a fast-fourier

59、 technique. The frequencies obtained from these spectra were compared to the frequencies obtained from e.g. (18) which yielded the curves as shown in Figure 8. From this figure it follows that for ß smaller that 0.3

60、 the calculation errors are small. For higher values of ß the calculation error made by a linear approximation is more than 10 %. Application of a finite element model of the belt</p><p>　　For lower val

61、ues of ß the frequencies of transverse vibration can also be predicted accurate by e.g. (18). However it is not possible to analyse, for example, the interaction between the belt sag and the propagation of longitudi

62、nal waves or the lifting of the belt off the idlers as can be done with the finite element model.</p><p>　　The determined relation between the belt stress and the frequency of transverse vibrations can also

63、be used in belt tension monitoring systems.</p><p>　　Figure 8: Ratio between the linear and the non-linear frequency of transverse vibration of a belt span supported by two idlers.</p><p>　　4. E

64、XPERIMENTAL VERIFICATION</p><p>　　In order to be able to verificate the results of the simulations, experiments have been carried out with the dynamic test facility shown in Figure 9.</p><p>　　F

65、igure 9: Dynamic test facility</p><p>　　With this test facility the transverse vibration of an unloaded flat belt span between two idlers, as for example a return part, can be determined. An acoustic device

66、is used to measure the displacement of the belt. Besides that, also the tensioning force, belt speed, motor torque, idler rotations and idler space were known during the experiments.</p><p>　　5. EXAMPLE</

67、p><p>　　Since the most cost-effective operation conditions of belt conveyors occur in the range of belt widths 0.6 - 1.2 m [2], the belt's capacity can be varied by varying the belt speed. However before th

68、e belt speed is varied the interaction between the belt and the idler should be determined in order to ensure resonance free belt support. To illustrate this the transverse displacement of a stationary moving belt span b

69、etween two idlers have been measured. The total belt length L was 52.7 m, the idle</p><p>　　After transformation of this signal by a fast fourier technique the frequency spectrum of Figure 5 was obtained. In

70、 Figure 5 three frequencies appear. The first frequency is caused by the passage of the belt splice:</p><p>　　fs = Vb/L = 0.067 Hz</p><p>　　The second frequency, which appears at 1.94 Hz, is cau

71、sed by the transverse vibration of the belt.</p><p>　　Figure 10: Frequencies of transverse vibration of a stationary moving belt span supported by two idlers.</p><p>　　The third frequency which

72、appears at 10.5 Hz is caused by the rotation of the idlers. From the numerical simulations Figure 11 was obtained.</p><p>　　Figure 11: Calculated resonance zone's for different idler diameters D. Cross i

73、ndicates belt speed and idler space during experiment.</p><p>　　Figure 11 shows the zone's where resonance caused by the belt/idler interaction may be expected for three idler diameters. The idlers of th

74、e belt conveyor had a diameter of 0.108 m thus resonance phenomena may be expected nearby a belt speed of 0.64 m/s. To check this, the maximum transverse displacement of the belt span has been measured during a start-up

75、of the conveyor.</p><p>　　Figure 12: Measured ratio of the standard deviation of the amplitude of transverse vibration and the static belt sag.</p><p>　　As can be seen in Figure 12 the maximum a

76、mplitude of the transverse vibration occur at a belt speed of 0.64 m/s as was predicted by the results of simulation with the finite element model. Therefore the belt speed should not be chosen nearby 0.64 m/s. Although

77、a flat belt is used for the experiments and the theoretical verification, the applied techniques can also be used for troughed belts.</p><p>　　6. CONCLUSIONS</p><p>　　Application of beam element

78、s in finite element models of belt conveyors enable the simulation of the transverse displacement of the belt thus enabling the design of resonance free belt supports. The advantage of applying beam elements for small va

79、lues of ß instead of using a linear differential equation to predict resonance phenomena is that also the interaction between the longitudinal and transverse displacement of the belt and the lifting of the belt off

80、the idlers can be predicted from simul</p><p>　　輸送帶的二維動態(tài)特性</p><p>　　3.1.1非線性梁架（構(gòu)架）元</p><p>　　如果只有帶的縱向變形是主要素，那么梁架元就可用于模型的皮帶彈性反應(yīng)。梁架元組成部分有如圖2所示的兩個結(jié)點， P和Q ，四個位移參數(shù)確定部分載體X：</p><

81、;p>　　xT = [up vp uq vq]             (1)</p><p>　　對平面運動的梁架元有三個獨立的剛體運動，因此（這公式）仍然是描述一個變形的參數(shù)。</p><p>　　圖2 ：梁架元的精確位移</p><p>

82、;　　梁架元軸的長度變化， [ 7 ] ：</p><p>　　DSO是限元未變形的長度，DS是限元變形的長度，ξ是沿著有限元軸的無量綱長度。</p><p>　　圖3 ：張帶的靜態(tài)凹陷</p><p>　　雖然帶呈彎曲狀態(tài)，但梁架元并沒有變形，這可能考慮到帶小數(shù)值凹陷的靜態(tài)影響。靜態(tài)帶凹陷的比率是有定義的（見圖3 ） ：</p><p>　

83、　K1 = δ/1 = q1/8T           (3)</p><p>　　其中q是暴露在外面帶和散裝物料的重量在豎直方向上分布的荷載， 1是帶輪間距，而T是帶的張力。，帶凹陷的縱向變形影響取決于[ 7 ] ：</p><p>　　εs = 8/3 K²s  &#

84、160;           (4)</p><p>　　產(chǎn)生了非線性梁架元總的縱向變形。</p><p><b>　　3.1.2梁架元</b></p><p>　　圖4 ：節(jié)點的精確位移和旋轉(zhuǎn)的梁架元。</p><

85、p>　　如果帶的橫向位移是主要因素，那么梁架元就可以用來模擬皮帶。同樣對于擁有六個位移參數(shù)的梁架元的平面運動來說，相當于三個獨立的剛體運動。因此就剩下三個變形參數(shù)是：縱向變形參數(shù)ε1 ，兩個彎曲變形參數(shù)ε2和ε3 。</p><p>　　圖5 ：梁架元的彎曲變形的</p><p>　　梁架元彎曲變形的參數(shù)可以定義為梁架元的組成載體（見圖4 ） ：</p><p&g

86、t;　　xT = [up vp µp uq vq µq]          (5)</p><p><b>　　和如圖5的變形結(jié)構(gòu)</b></p><p>　　3.2繞過托輥及帶輪的帶運動</p><p>　　當繞過托輥或帶輪的時候，帶運動是受到

87、約束的。為了說明（弄清楚）這些制約因素，影響制約因素（邊界）的條件都必須添加到用來代模擬帶的有限元中來。這可以通過使用多體動力學進行描述。多體機置動力學的經(jīng)典描述，建立起由若干約束條件連接起來的剛體或剛性鏈接。在（變形）輸送帶的有限元描述里，帶被分離成多個有限元，有限元之間的聯(lián)系是可變形的。有限元是由節(jié)點連接的，因此分配了位移參數(shù)。要確定帶的運動，排除了剛體模型的變形模式。如果一個帶繞過托輥，，決定托輥上帶的位置（如見圖6）的帶長度為ξ

88、，被添加到組件矢量，如：式（6） ，因此產(chǎn)生了7個位移矢量參數(shù)。</p><p>　　圖6 ：由托輥支撐的帶</p><p>　　梁架元有兩個獨立的剛體運動，因此依然有五個變形參數(shù)存在。其中已經(jīng)在3.1中給出了ε1 , ε2和ε3 ，確定了帶的變形。剩下ε4和ε5 ，確定帶和托輥之間的相互作用，見圖7 。</p><p>　　圖7 ：兩個約束條件的梁架元有限元。&l

89、t;/p><p>　　這些變形參數(shù)可以假設(shè)成無限剛度的彈性。這意味著：</p><p>　　ε4 = D4(x) = (rξ + u ξ)e2 - rid.e2 = 0</p><p>　　ε5 = D5(x) = (r ξ + uξ)e1 - rid.e1 = 0         

90、0;      (7)</p><p>　　如果模擬的是ε4 > 0的時候，那么帶將脫離托輥，而描述帶的有限元上的約束條件也將去除。</p><p><b>　　3.3滾動阻力</b></p><p>　　為了使一種模型能應(yīng)用于帶式輸送機有限元模型的滾動阻力，已經(jīng)制定了

91、一種計算滾動阻力的近似公式， [ 8 ] 。帶運動中，暴露在帶外面的總滾動阻力的組成部分，這三部分是耗能的主要部分，可以區(qū)分為包括：壓痕滾動阻力，托輥的慣性（加速滾動阻力）和軸承滾動阻力（軸承阻力） 。確定滾動阻力因素的參數(shù)包括直徑和托輥的材料，以及各種帶參數(shù)，如速度，寬度，材料，緊張狀態(tài)，環(huán)境溫度，帶橫向負荷，托輥間距和槽角?？倽L動阻力的因素，可以表示成總滾動阻力和帶垂直負荷之間的比例，定義為：</p><p>

92、;　　ft = fi + fa + fb                  (8)</p><p>　　Fi是壓痕滾動阻力的系數(shù)，F(xiàn)A是加速阻力系數(shù)，而FB是軸承阻力系數(shù)。這些組成系數(shù)由下面的[9]確定：</p><p>　　FZ是

93、帶垂直方向上分布的負載和散裝物料的負載的總和， H是帶的覆蓋厚度，D是托輥的直徑，Vb是帶速，KN是帶負荷的名義百分之比，T是環(huán)境溫度，Mred是托輥的折算質(zhì)量，B是帶的寬度， U是帶的縱向位移，MF是總的軸承阻力矩和RI是軸承內(nèi)部半徑。在計算滾動阻力中，皮帶的動力性能及機械性能和皮帶上覆蓋的材料發(fā)揮著重要作用。這使得帶的選擇和帶上覆蓋材料，盡量減少由動力阻力引起的能源消耗。</p><p><b>　

94、　3.4帶驅(qū)動系統(tǒng)</b></p><p>　　在穩(wěn)定性的帶運動情況下，為了能夠測定帶式輸送機驅(qū)動系統(tǒng)的旋轉(zhuǎn)組件的影響，這個帶式輸送機的總模型必須是含有驅(qū)動系統(tǒng)模型。驅(qū)動系統(tǒng)的旋轉(zhuǎn)元件，就像一個減速箱，參照了3.2節(jié)中所述的約束條件。帶有減速比的減速箱，可以用帶兩個位移參數(shù)的減速元件來代替， μp和μq ，像一個剛體的（旋轉(zhuǎn)）運動，因此就剩下一個變形參數(shù)：</p><p>　　

95、εred = Dred(x) = iµp + µq = 0               (10)</p><p>　　要確定電式扭矩感應(yīng)式電機，是否適應(yīng)所謂的兩軸式電動機。該相電壓的矢量v可從（11）獲得：</p><p>　　v

96、= Ri + ωsGi + L ?i/?t                    (11)</p><p>　　在（11）式中I是相電流矢量，R是模型的相電阻， c是模型的相電感抗，L是模型的相感系數(shù)而ωs是電機轉(zhuǎn)子的角速度。電磁轉(zhuǎn)矩等于：<

97、;/p><p>　　Tc = iTGi               (12)</p><p>　　電機模型和驅(qū)動系統(tǒng)機械組件是由驅(qū)動系統(tǒng)的運動方程聯(lián)系著的：</p><p>　　其中T是扭矩矢量，I是模型的慣量，C是模型的阻尼，K是

98、矩陣剛度和ø是電機旋轉(zhuǎn)軸的角速度。 </p><p>　　模擬啟動或停止程序控制反饋的程序可以添加到帶式驅(qū)動系統(tǒng)模型中，用來控制驅(qū)動扭矩。</p><p><b>　　3.5運動方程</b></p><p>　　整個帶式輸送機模型的運動方程可以得出潛在功率的原則， [ 7 ] ：</p><p>　　fk -

99、Mkl ?²x1 / ?t² = σ1Dik                (14)</p><p>　　其中F是阻力矢量，M是模型的質(zhì)量而σ是拉格朗日乘數(shù)的矢量，可能解釋為雙重壓力矢量to張力矢量ε 。為了解決帶有X這一組方程，方程一體化是必要

100、的。但是一體化的結(jié)果，必須確保滿足約束條件。如果(8)式中應(yīng)變?yōu)榱?，那么必須糾正一體化結(jié)果，如見[ 7 ] ?？梢允褂媚Ｐ偷姆答佭x擇，例如限制提升物質(zhì)垂直方向上的運動。這種違逆動力學的問題可以用下面公式表示。鑒于帶模型及其驅(qū)動系統(tǒng)的提升運動眾所周知，根據(jù)系統(tǒng)自由度和它的比例（速度）可以確定其他元件的運動。它超出了本文所討論關(guān)于此項的所有細節(jié)范圍。</p><p><b>　　3.6實例</b>

101、;</p><p>　　為了在長距離帶式輸送機系統(tǒng)設(shè)計階段能夠正確設(shè)計，應(yīng)用了有限元法。例如帶強度的選擇，可以減少的盡量減少，使用模型模擬的結(jié)果確定傳送帶的最大張力。以有限元模型的功能作為例子，應(yīng)該考慮到在兩個托輥位置范圍之間穩(wěn)定移動帶的橫向振動。在運輸機的設(shè)計階段這必須被確定，才得以確保空帶的共振。 </p><p>　　對于皮帶輸送機的設(shè)計來說，托輥和移動帶間相互作用影響是很重要的。托

102、輥的及帶輪的幾何不完善性，導(dǎo)致帶脫離托輥和帶輪能支撐的位置，在帶和支撐帶輪之間產(chǎn)生一種橫向振動。這對帶施加了一部分的交互軸向應(yīng)力。如果這部分力是比皮帶的預(yù)應(yīng)力小，那么帶將在它的固有頻率中振動，否則帶將被迫振動。皮帶是會受迫振動的，例如受托輥的偏心率影響。在輸送帶返程中，這種振動特別值得注意。由于受迫振動的頻率取決于帶輪和托輥的角速度，因此對于帶的速度，確定在帶輪和托輥之間，帶在自然頻率狀況下，橫向振動中帶速影響，這個是很重要的。如果受迫

103、振動的頻率接近于皮帶橫向振動的固有頻率，將發(fā)生共振現(xiàn)象。 </p><p>　　有限元模型的模擬結(jié)果可用于確定穩(wěn)定移動的帶的橫向振動頻率范圍。該頻率是利用快速傅立葉技術(shù)從時域范圍到頻域范圍，帶橫向位移變換后得到的結(jié)果。除了使用有限元模型外也可以運用近似分析法。</p><p>　　皮帶可以模擬成一個預(yù)應(yīng)力梁。如果皮帶的彎曲硬度可以被忽略，橫向位移比托輥間距還小，Ks << 1

104、，并且?guī)г黾拥拈L度相對于橫向位移的原始長度來說是微不足道，帶的橫向振動可近似為下列線性微分方程，如見圖15 ：</p><p>　　其中V是皮帶的橫向位移和C2是橫向波的波速度，由（16）式定義：</p><p>　　c2 = √g1/8Ks            

105、60;     (16)</p><p>　　首先，圖5中帶的橫向固有頻率范圍可從公式（16）獲得，如果假定v(O,t)=v(l,t)=0：</p><p>　　ß是無量綱的速比，由（18）式確定：</p><p>　　ß = Vb / c2    

106、0;             (18)</p><p>　　FB是不同帶的各自獨立的頻率范圍，由于輸送帶長度方向上帶張力變化。托輥的受迫振動頻率，使托輥產(chǎn)生了一個偏心率等于：</p><p>　　fi = Vb / πD    &#

107、160;              (19)</p><p>　　其中D是托輥的直徑。為了設(shè)計一個在托輥間距中無支撐的共振，這受到以下條件限制：</p><p>　　由線性微分方程（16）所取得的成果不過是只適用于小數(shù)值的速比ß。對于大數(shù)值的

108、速比ß來說，如高速運輸機或低的帶張力，在（16）式中所有非線性條件就顯得重要的。因此，數(shù)值模擬的運用，有限元模型的開發(fā)，都是為了確定帶橫向振動線性和非線性頻率之間的比例范圍。這些關(guān)系已被確定適合不同的數(shù)值的ß，例如說一個功能凹陷的比率Ks。</p><p>　　使用快速傅里葉技術(shù)將橫向位移結(jié)果的轉(zhuǎn)化為頻譜。從這些頻譜中獲得的頻率與公式（18）獲得的頻率相比，其產(chǎn)生了圖8所顯示的曲線。從這一數(shù)字

109、可見，對小于0.3的ß來說，計算誤差很小。對于大數(shù)值的ß來說，運用線性近似值法產(chǎn)生的計算誤差達到10 ％以上。運用了皮帶采用非線性梁架元的有限元模型，因此可以準確地確定大數(shù)值ß的橫向振動。</p><p>　　對于小數(shù)值ß的橫向振動的頻率也可以用公式（18）準確地預(yù)測。然而，它不能分析，例如帶凹陷和縱向波的傳播之間的相互作用，或者同樣可以看成有限元模型的脫離托輥的皮帶。&l

110、t;/p><p>　　這決定帶應(yīng)力和橫向振動頻率之間的關(guān)系可以用于皮帶張力監(jiān)測系統(tǒng)。</p><p>　　圖8 ：由兩個托輥支撐的帶的橫向振動線性和非線性頻率之間的比例。</p><p><b>　　4 實驗驗證</b></p><p>　　為了使模擬的結(jié)果能夠得到驗證，實驗中使用了動態(tài)試驗設(shè)備，如圖9所示。</p&g

111、t;<p>　　圖9 ：動態(tài)試驗設(shè)施</p><p>　　使用這試驗設(shè)施能夠確定的兩個托輥的間距和卸荷扁帶的橫向振動，例如返程部分的。聲音裝置是用來測量皮帶的位移。此外，還有在試驗中為我們所知的張緊力，帶速，電機轉(zhuǎn)矩，托輥轉(zhuǎn)子與托輥的距離。</p><p><b>　　5 為例</b></p><p>　　由于最具有成本效益帶式輸