外文翻譯--動(dòng)力減振鏜桿結(jié)構(gòu)參數(shù)優(yōu)化

1、<p><b>　　附錄A</b></p><p>　　動(dòng)力減振鏜桿結(jié)構(gòu)參數(shù)優(yōu)化</p><p>　　摘要：深孔鏜削過(guò)程中，鏜桿不可避免產(chǎn)生振動(dòng)，影響孔的加工質(zhì)量，為了提高加工質(zhì)量，本文針對(duì)動(dòng)力減振鏜桿建立力學(xué)模型，通過(guò)對(duì)模型的研究得出減振器的最優(yōu)參數(shù)，應(yīng)用ADAMS動(dòng)力學(xué)仿真軟件和試驗(yàn)驗(yàn)證了理論優(yōu)化的正確性。通過(guò)和普通鏜桿對(duì)比分析，結(jié)果表明動(dòng)力減振鏜桿有效

2、地達(dá)到了減振效果。</p><p>　　關(guān)鍵詞：減振器結(jié)構(gòu)；動(dòng)態(tài)性能；參數(shù)優(yōu)化</p><p><b>　　1．引言</b></p><p>　　在深孔鏜削過(guò)程中，受到孔的尺寸限制，鏜桿長(zhǎng)徑比較大，剛度小，固有頻率低，在受到機(jī)床自身激勵(lì)和外部激勵(lì)時(shí)，很容易發(fā)生振動(dòng)，影響工件的加工精度和表面質(zhì)量。三菱公司通過(guò)減輕鏜桿頭部的的重量來(lái)提高鏜桿的剛度，

3、美國(guó)Kenametal公司生產(chǎn)的減振鏜桿 (最大長(zhǎng)徑比 L /D = 8 ) 主要采用特殊材料來(lái)提高鏜桿靜剛度,這些方法受到長(zhǎng)徑比的限制。</p><p>　　動(dòng)力減振鏜桿可以進(jìn)一步提高長(zhǎng)徑比，在深孔加工方面具有很大的優(yōu)勢(shì)。Warburton通過(guò)對(duì)附加在鏜桿上的減振器的參數(shù)進(jìn)行優(yōu)化來(lái)實(shí)現(xiàn)對(duì)主系統(tǒng)的減振，減振器包括彈簧，阻尼和減振塊。在載荷作用下， J iaJang W u研究了減振器螺旋彈簧的慣性效應(yīng)對(duì)鏜桿動(dòng)態(tài)特

4、性的影響。Felipe Antonio Chegury Viana等人基于蟻群算法設(shè)計(jì)出可調(diào)動(dòng)態(tài)減振器。這些方法所設(shè)計(jì)出的動(dòng)力減振鏜桿成本較高，結(jié)構(gòu)復(fù)雜，維護(hù)麻煩，當(dāng)前應(yīng)用不廣泛。</p><p>　　針對(duì)上述問(wèn)題，下面將采用虛擬樣機(jī)技術(shù)，在ADAMS環(huán)境下進(jìn)行減振器結(jié)構(gòu)優(yōu)化，最后進(jìn)行實(shí)驗(yàn)驗(yàn)證，通過(guò)對(duì)比分析，表明理論優(yōu)化的結(jié)果、仿真結(jié)果和實(shí)驗(yàn)結(jié)果基本一致，降低了設(shè)計(jì)成本。</p><p>

5、　　2．動(dòng)力減振鏜桿理論及建模</p><p>　　動(dòng)力減振是將主系統(tǒng)的能量轉(zhuǎn)移到減振器系統(tǒng)上，減小主系統(tǒng)的振動(dòng)。減振鏜桿結(jié)構(gòu)如圖1所示，建立的力學(xué)模型如圖 2所示。動(dòng)力學(xué)方程可表示為</p><p><b>　　主系統(tǒng)的振動(dòng)幅值為</b></p><p>　　對(duì)不同的ξ值所作出的主系統(tǒng)的幅頻響應(yīng)曲線如圖3所示，當(dāng)ξ =∞時(shí)，鏜桿和減振器之間沒(méi)

6、有相對(duì)運(yùn)動(dòng)，成為單自由度系統(tǒng)，時(shí)其幅頻曲線只有一個(gè)峰值，等效于普通鏜桿。當(dāng) ξ介于 0和∞之間時(shí)，系統(tǒng)為兩自由度，產(chǎn)生兩個(gè)共振點(diǎn)。阻尼的存在使主系統(tǒng)的共振幅值減少，但并不能完全消除主系統(tǒng)的振動(dòng)。圖 3中所有的曲線都相交于P、Q兩點(diǎn), 表明P、Q兩點(diǎn)的頻率和幅值與 ξ的變化無(wú)關(guān)，得出方程式為</p><p>　　求出 P、Q 兩點(diǎn)的頻率,帶入( 2 )式得到 P、Q兩點(diǎn)的幅值。從 ( 2 ) 、( 3 ) 式可以看

7、出，對(duì)確定的主系統(tǒng)而言，幅值和頻率取決于減振器的質(zhì)量和彈簧。減振器最理想的結(jié)構(gòu)參數(shù)應(yīng)該是在P、Q兩點(diǎn)達(dá)到峰值，并且數(shù)值相等。根據(jù)這種思路，可按下述步驟選擇減振器的最優(yōu)參數(shù)。</p><p>　　對(duì)于確定的主系統(tǒng)和選定的減振塊質(zhì)量，結(jié)構(gòu)最優(yōu)參數(shù)解為：</p><p>　　進(jìn)而確定減振器的剛度</p><p>　　在 P、Q兩點(diǎn)取駐點(diǎn)的條件下，求得減振器的阻尼率ξ<

8、;/p><p><b>　　3．動(dòng)力學(xué)仿真</b></p><p>　　為了驗(yàn)證所建模型的有效性，在ADAM S環(huán)境下進(jìn)行仿真。應(yīng)用ADAMS中有限元模塊將鏜桿桿體模型轉(zhuǎn)變成柔體，在刀頭端部創(chuàng)建輸入和輸出通道，然后進(jìn)行系統(tǒng)的振動(dòng)分析，通過(guò)仿真計(jì)算，在后處理模塊中得出系統(tǒng)的模態(tài)和頻響函數(shù)。</p><p>　　減振器初始參數(shù)，，。鏜桿桿體的結(jié)構(gòu)尺寸：

9、直徑D = 0. 016 m ,長(zhǎng)度L =0. 192 m ,長(zhǎng)徑比為12: 1；材料屬性：密度ρ= 7 801 kg/m，彈性模量E = 2. 07E + 011 N /m2，泊松比ν= 0. 29。根據(jù)結(jié)構(gòu)圖建立振動(dòng)模型。</p><p>　　減振塊質(zhì)量的變化對(duì)幅頻曲線的影響。當(dāng)m 2 = 0. 02 kg時(shí)，得到前兩階自然頻率為253 Hz和452 Hz,共振時(shí)的最大幅值為- 95. 16 dB 和- 10

10、3. 3 dB；當(dāng)m 2 = 0. 10 kg時(shí)，前兩階的自然頻率為128 Hz 和406 Hz，共振時(shí)的最大幅值為- 95. 2 dB - 95. 3 dB。對(duì)不同的質(zhì)量值繪制主系統(tǒng)的幅頻響應(yīng)曲線如圖4所示?？梢钥闯鲎匀活l率隨著減振塊質(zhì)量的增加而降低，當(dāng)外部激勵(lì)的頻率與主系統(tǒng)的自然頻率接近時(shí)，可以通過(guò)修改減振塊質(zhì)量的方法來(lái)避免發(fā)生共振，而減振塊質(zhì)量對(duì)幅值的影響不敏感。</p><p>　　圖 4 頻響函數(shù)隨質(zhì)量

11、變化曲線</p><p>　　阻尼的變化對(duì)幅頻特性曲線的影響。當(dāng)c2 = 10 N s/m時(shí)，前兩階自然頻率為 253 Hz和 452 Hz,共振時(shí)最大幅值為- 94. 75 dB 和- 103. 24 dB；c2 = 2 N s/m ,前兩階的自然頻率為 253 Hz 和 452 Hz, 共振時(shí)最大幅值為 - 90. 11 dB , 和 - 95. 49 dB。圖5為振動(dòng)分析后繪制的頻響曲線圖，表明阻尼的變化對(duì)

12、幅值的影響比較大，幅值隨阻尼的增大而減小，當(dāng)共振不可避免時(shí)，通過(guò)修改阻尼來(lái)減小振幅，而阻尼對(duì)自然頻率的影響不太明顯。</p><p>　　圖 5 頻響函數(shù)隨阻尼變化曲線</p><p>　　剛度的變化對(duì)幅頻特性的影響。當(dāng)剛度 k2 = 10 kN /m時(shí)，前兩階的自然頻率為 253 Hz和 452 Hz，共振時(shí)的最大幅值為 - 94. 71 dB 和 - 108. 20 dB; 當(dāng) k2

13、= 200 kN /m 時(shí)，前兩階的自然頻率為 284 Hz和 898 Hz, 共振時(shí)的最大幅值為 - 90. 27 dB和 - 110. 06 dB。圖6為繪制的頻響函數(shù)圖，表明自然頻率隨剛度的增加而增大，剛度的變化對(duì)幅值的影響比較大，通過(guò)修改剛度可避免共振和調(diào)整幅值。</p><p>　　圖 6 頻響函數(shù)隨剛度變化曲線</p><p><b>　　4．減振優(yōu)化</b&g

14、t;</p><p>　　根據(jù)動(dòng)力減振鏜桿振動(dòng)分析模型，以減振器的剛度和阻尼作為設(shè)計(jì)變量，使用ADAMS中View變量和振動(dòng)宏作為目標(biāo)函數(shù)，使目標(biāo)函數(shù)最小。約束條件為振動(dòng)幅值小于減振器和鏜桿內(nèi)腔之間的距離，優(yōu)化采用OPTDES-GRG廣義遞減梯度算法。參數(shù)優(yōu)化的目的就是在給定的鏜桿結(jié)構(gòu)和減振塊質(zhì)量一定的條件下，優(yōu)化出減振器的剛度和阻尼參數(shù)，當(dāng)采用最優(yōu)參數(shù)時(shí)主系統(tǒng)的振動(dòng)幅值最小。當(dāng)減振塊質(zhì)量 m 2 =0. 021

15、 44 kg，優(yōu)化后的曲線和普通鏜桿曲線如圖 7所示。</p><p>　　圖 7 普通鏜桿和優(yōu)化后減振鏜桿</p><p>　　優(yōu)化后減振器的參數(shù)是 k2 = 58 662 N /m，c2 = 22. 34N s/m，前三階的自然頻率為 228 Hz、309 Hz和 392 Hz,前兩階的自然頻率的比值 0. 7378，根據(jù)公式 ( 4)計(jì)算出前兩階自然頻率的比值為 0. 7376,相對(duì)

16、誤差為 0. 04%。仿真優(yōu)化的阻尼率為 0. 221,公式 ( 6)得出的阻尼率為 0. 216,相對(duì)誤差為 2. 2%。根據(jù)上述定量分析，得出仿真優(yōu)化和理論優(yōu)化結(jié)果基本一致，表明仿真優(yōu)化有效可行。</p><p>　　從圖 7中可以看出，在激勵(lì)條件不變的情況下，與普通鏜桿相比，減振鏜桿的振型得到明顯的改善，振型變得更加光滑，幅值也明顯減小。共振時(shí)最大幅值為 - 102. 33 dB，根據(jù)信號(hào)處理理論，實(shí)際幅值

17、和曲線幅值的對(duì)應(yīng)關(guān)系</p><p>　　M agnitude為仿真曲線幅值，根據(jù)上式得到實(shí)際振幅為0. 007 6 mm。普通鏜桿與優(yōu)化減振鏜桿對(duì)比見(jiàn)下表，表明在長(zhǎng)徑比較大的情況下，動(dòng)力減振鏜桿振動(dòng)幅值僅是普通鏜桿幅值的 23%，具有很好的減振效果。</p><p><b>　　5. 結(jié)論</b></p><p>　　在動(dòng)力學(xué)仿真技術(shù)的基礎(chǔ)上

18、，較為系統(tǒng)的探討了動(dòng)力減振鏜桿的動(dòng)態(tài)特性，以及減振器參數(shù)的變化對(duì)主系統(tǒng)的影響，并對(duì)參數(shù)進(jìn)行優(yōu)化，參數(shù)優(yōu)化結(jié)果和理論優(yōu)化結(jié)果吻合良好，最后通過(guò)和加工范圍。該方法對(duì)于進(jìn)一步提高深孔加工領(lǐng)域的水平和相關(guān)技術(shù)的研究具有十分重要的理論意義和實(shí)際應(yīng)用價(jià)值。</p><p><b>　　參考文獻(xiàn)</b></p><p>　　[1] D G Lee, H Y Hwang and J

19、 K Kim. Design and manufacture of acarbon fiber epoxy rotating boring bar [ J ]. Composite Structures,2003, 60 ( 1) : 115～124.</p><p>　　[2] SANJ I G TEWAN I, KE ITH E ROUCH and BRUCE L WALCOTT A study of cu

20、tting p rocess stability of a boring bar with ac2tive dynam ic absorber [ J ]. I Mach. Tools Manufact 1995, 35 ( 1) : 91～108.</p><p>　　[3] G B W arburton. Op tim um absorber parameters for m inim izing vibr

21、ation response[ J ]. Journal of Earthquake Engineering and Structural Dynam ics , 1981, 9: 251～262.</p><p>　　[4] J ia - Jang W u . Study on the inertia effect of helical sp ring of the absorber on suppressi

22、ng the dynam ic responses of a beam subjected to a moving load [ J ]. Journal of Sound and V ibration. 2006, 297 ( 3- 5) : 981～999.</p><p>　　[5] Felipe Antonio Chegury V iana, Giovanni Iam in Kotinda, Tunin

23、gdynam ic vibration absorbers by using ant colony op tim ization [ J ].Computers and Structures, 2008, 86 ( 13～14) : 1539～1549.</p><p>　　[6]　邵俊鵬 ,秦柏.基于ADAMS的動(dòng)力減振鏜桿仿真分析 [ J ].機(jī)械設(shè)計(jì)與研究 , 2008, 24 ( 1) : 84～88.&

24、lt;/p><p>　　[7]　師漢民. 機(jī)械振動(dòng)系統(tǒng) — 、分析 測(cè)試 建模 對(duì)策 [M ]. 武漢 :華中科技大學(xué)出版社 , 2004.</p><p><b>　　附錄B</b></p><p>　　A Study of Optimum Parameters of A Boring Bar with Passive Dynamic Absor

25、ber</p><p>　　Abstract: The vibration of the boring bar directly affects the processing quality in the deep hole machining In order to improve the processing quality, theoretical model of a boring bar with pa

26、ssive dynamic absorber has been developed and derived the optimum parameters of the absorber Both the dynamic simulation based on ADAM S and the experim ents were conducted to verify the theory Comparing w ith boring ba

27、r, numerical results reveal that boring bar with dynamic absorber has the effect of vi</p><p>　　Keywords: passive dynamic absorber structure; dynamic character; optimum parameter</p><p>　　Introd

28、uction</p><p>　　In the process of deep-hole boring, restricted by the size of holes, boring bar larger aspect ratio, stiffness of small, low natural frequencies. Inspired by the machine itself and external i

29、ncentives, it is prone to vibration, impact on the machining accuracy and workpiece surface quality. Mitsubishi boring bar by reducing the weight of the head of the boring bar to increase the stiffness, the United States

30、 produced Kenametal vibration boring bar (maximum aspect ratio L / D = 8) the main use of</p><p>　　Driving force for boring bar vibration can be further enhanced aspect ratio, and has great advantage in the

31、deep processing of. Through the pole attached to the parameters of the shock absorber,Warburton achieve the main system of the vibration, shock absorber, including springs, dampers and damping block. In the load, J iaJan

32、g W u studied coil spring shock absorber of the inertial effect on the dynamic properties of boring bar impact. Felipe Antonio Chegury Viana, who designed the Ant Colony Al</p><p>　　The following will be use

33、d virtual prototyping technology in response to these problems. In the ADAMS environment damper structural optimization, and finally to carry out experiments. By comparing the analysis results show that the theory of opt

34、imization, simulation results and experimental results are basically the same, lower design cost.</p><p>　　2. Driving force for boring bar vibration theory and modeling</p><p>　　Damping is the m

35、ain driving force for the energy transfer system to the shock absorber system to reduce the vibration of the main system. Boring bar vibration structure as shown in Figure 1, the establishment of the mechanical model sho

36、wn in Figure 2. Kinetic equation can be expressed as</p><p>　　1.the body of Boring Bar 2. rubber ring 3.gasket</p><p>　　4.damping block 5. damping 6.blocking 7.segment</p><p>　　F

37、ig.1 Boring bar vibration structure</p><p>　　Fig.2the establishment of the mechanical mode</p><p>　　The main system for the vibration amplitude</p><p>　　For different values of the

38、main system by the amplitude-frequency response curve as shown in Figure 3.</p><p>　　Fig.3 different damping ratio of vibration amplitude-frequency characteristic curve</p><p>　　When ξ = ∞, the

39、boring bar and there is no relative motion between the shock absorber, a single degree of freedom system, when amplitude-frequency curve is only one peak, equivalent to an ordinary boring bar. When the range of ξ between

40、 0 and ∞, the system of two degrees of freedom, resulting in the two resonance points. The existence of the damping of the resonance amplitude of the main system to reduce, but it does not completely eliminate the vibrat

41、ion of the main system. Figure 3 are all of </p><p>　　Calculated P, Q two points in the frequency Into (2) to be P, Q two points of the amplitude. From (2), (3) style can be seen that the main system for det

42、ermining, the amplitude and frequency depend on the quality shock absorber and spring. Structural parameters of the best shock absorber should be in the P, Q two points to reach the peak, and the same values. According t

43、o this line of thought, according to the following steps to select the optimal parameters of shock absorber.</p><p>　　For the determination of the main system and the selected block damping quality, the stru

44、cture of the optimal solution for the parameters:</p><p>　　To determine the stiffness of shock absorber</p><p>　　In P, Q two points from stagnation conditions, the shock absorber damping rate ob

45、tained ξ</p><p>　　3. Dynamics Simulation</p><p>　　In order to verify the validity of the model, ADAM S in the simulation environment. ADAMS application modules in the finite element model boring

46、 into flexible, in the head end of the creation of input and output channel, and then the vibration system analysis, through simulation, in the post-processing module to draw modal system and frequency response function.

47、</p><p>　　The initial parameters of shock absorber m2=0.02144，k2=10kN/m，c=10Ns/m。The size of boring structure: diameter D = 0. 016 m , length L =0. 192 m , aspect ratio of 12: 1. Material properties: density

48、 ρ = 7 801 kg / m, young's modulus E = 2. 07E + 011 N / m2, poisson's ratio ν = 0. 29.</p><p>　　Damping block changes in the quality of the effects of amplitude-frequency curves.When m 2 = 0. 02 kg,

49、the first two-order natural frequency of 253 Hz and 452 Hz, the maximum amplitude at resonance for the - 95. 16 dB and - 103. 3 dB;when m 2 = 0. 10 kg, order the first two natural frequency of 128 Hz and 406 Hz, the maxi

50、mum amplitude at resonance for the - 95. 2 dB - 95. 3 dB. The quality of the different values of the main system mapping amplitude-frequency response curve shown in Figure 4. As</p><p>　　Frequency/Hz</p&g

51、t;<p>　　Fig.4 With the quality of frequency response function curve</p><p>　　Changes in damping characteristics of the amplitude-frequency curves. When c2 = 10 N s / m, the first two-order natural fre

52、quency of 253 Hz and 452 Hz, maximum amplitude of the resonance for the - 94. 75 dB and - 103. 24 dB; c2 = 2 N s / m, the first two bands of 253 Hz natural frequency and 452 Hz, maximum amplitude of the resonance for the

53、 - 90. 11 dB, and - 95. 49 dB. Figure 5 after the draw for the vibration analysis of the frequency response curve, indicating that changes in damping the imp</p><p>　　Frequency/Hz</p><p>　　Fig.5

54、 Frequency response function with the damping curve</p><p>　　Changes in stiffness of the effects of amplitude-frequency characteristics. When the stiffness k2 = 10 kN / m, the first two natural frequency ban

55、d 253 Hz and 452 Hz, the maximum amplitude at resonance for the - 94. 71 dB and - 108. 20 dB; When k2 = 200 kN / m, the first two natural frequency band 284 Hz and 898 Hz, the maximum amplitude at resonance for the - 90.

56、 27 dB and - 110. 06 dB. Figure 6 The frequency response function for drawing maps showing the natural frequency with the increase of </p><p>　　Frequency/Hz</p><p>　　Fig.6 Frequency response fun

57、ction curve with the stiffness</p><p>　　4. Damping optimization</p><p>　　Damping according to driving force for boring bar vibration analysis model of shock absorber stiffness and damping as a d

58、esign variable, the use of ADAMS and vibration in the View macro variables as the objective function, so that the smallest objective function. Constraints for the amplitude of vibration damper and the boring bar is less

59、than the distance between the cavity and optimize the use of generalized OPTDES-GRG reduced gradient algorithm. The purpose of optimization is in a given struc</p><p>　　Frequency/Hz</p><p>　　Fig

60、.7 Ordinary boring bar and boring bar vibration optimized</p><p>　　Optimized the parameters of shock absorber is k2 = 58 662 N / m, c2 = 22. 34N s / m.Before the third-order natural frequency of 228 Hz, 309

61、 Hz and 392 Hz, the first two bands of the ratio of the natural frequency of 0.7378. According to the formula (4) to calculate the natural frequency of the first two bands for the ratio of 0.7376, the relative error is 0

62、.04%. Simulation and optimization of the damping rate of 0.221, the formula (6) derived from the damping rate of 0.216, the relative error is</p><p>　　From Figure 7 can be seen in the excitation conditions r

63、emain unchanged, compared with the ordinary boring bar, boring bar vibration of the vibration mode has been marked improvement in vibration mode becomes more smooth, the amplitude is also significantly reduced. Maximum a

64、mplitude for the resonance - 102. 33 dB, based on signal processing theory, the actual amplitude and the amplitude of the correlation curve</p><p>　　actual amplitude=10Magnitude/20 (7)</

65、p><p>　　M agnitude amplitude curve for the simulation, according to the real amplitude of type 0. 007 6 mm. General Boring Bar Boring Bar Vibration and Optimization of contrast in the table below that in the ca

66、se of larger aspect ratio, power vibration amplitude vibration boring bar is just an ordinary boring bar of 23% amplitude, with very good damping effect .</p><p>　　the results of comparative table ordinary b

67、oring bar and boring bar to optimize</p><p>　　5. Conclusion</p><p>　　In the dynamic simulation technology based on more dynamic system of boring bar vibration of the dynamic characteristics, as

68、well as the shock absorber parameters change on the impact of the main system, and optimization of parameters, parameter optimization to optimize the results and theoretical results good, and finally through the scope an

69、d process. The method for further improving the level of deep processing and related areas of research have important theoretical and practical application</p><p>　　References</p><p>　　[1] D G

70、Lee, H Y Hwang and J K Kim. Design and manufacture of acarbon fiber epoxy rotating boring bar [ J ]. Composite Structures,2003, 60 ( 1) : 115～124.</p><p>　　[2] SANJ I G TEWAN I, KE ITH E ROUCH and BRUCE L W

71、ALCOTT A study of cutting p rocess stability of a boring bar with ac2tive dynam ic absorber [ J ]. I Mach. Tools Manufact 1995, 35 ( 1) : 91～108.</p><p>　　[3] G B W arburton. Op tim um absorber parameters f

72、or m inim izing vibration response[ J ]. Journal of Earthquake Engineering and Structural Dynam ics , 1981, 9: 251～262.</p><p>　　[4] J ia - Jang W u . Study on the inertia effect of helical sp ring of the a

73、bsorber on suppressing the dynam ic responses of a beam subjected to a moving load [ J ]. Journal of Sound and V ibration. 2006, 297 ( 3- 5) : 981～999.</p><p>　　[5] Felipe Antonio Chegury V iana, Giovanni I