頻譜分析在轉(zhuǎn)子動平衡中的應(yīng)用#中英文翻譯#外文翻譯匹配

2009 屆本科畢業(yè)論文（設(shè)計） 相關(guān)中英文翻譯資料 資料 題目： 頻譜分析在轉(zhuǎn)子動平衡中的應(yīng)用 學(xué)生姓名：徐軍鋒 所在院系：機電學(xué)院 所學(xué)專業(yè)：機電技術(shù)教育 指導(dǎo)老師：付素芳 APPLICATION OF FREQUENCY SPECTRUM ANALYSIS IN THE ROTATOR MOVING EQUILIBRIUM ABSTRACT The experimental equipment is developed to simulate the rotator vibration. The running state of machine is simulated by using different running conditions. The vibration caused by non-equilibrium mass is analyzed and discussed for first order with focus load. The effective method is found out by using frequency spectrum analysis. INTRODUCTION In the conventional island of nuclear power plant, turbine generator set is a very important equipment in which the core thermal energy is transferred into electric energy. When the turbine generator set has run for long time, the original equilibrium of system would be upset because of the remnant deformed of the metal, abrasion or damaging of the components etc. As a result, the vibration will be increased. So it is necessary to adjust the equilibrium at spot. On the other hand, a large turbine generator set also needs the work for adjusting the equilibrium in the process of manufacture, debugging, installation and operation. The moving equilibrium technique at spot is an important means to eliminate the violent vibration of the turbine generator set. We could have a definitely view for the vibration type, vibration power source and vibration property by analyzing the vibration of the turbine generator set or doing some special experiment. When the vibration signal is obtained, the frequency spectrum could be used to analysis the vibration signal in order to diagnose quickly. Using frequency spectrum analysis, the electrical signal of vibration that is obtained by the vibration sensor and has a wide frequency range will be decomposed into several main frequency compositions. Different frequency compositions have different influence on the turbine generator set. The frequency spectrum analysis is a very useful method to study the vibration of turbine generator set. The vibration caused by rotator mass non-equilibrium with concentrated load is discussed and analyzed in this paper. And an effective method to prevent the vibration is presented by using frequency spectrum analysis. 1 EXPERIMENTAL EQUIPMENT The experimental system consists of the motor, shaft coupling and rotor etc. Its structure is very simple. The rotor is driven by the motor directly. Its rotating speed could be adjusted in a wide range. The system could be operated smoothly and reliably. The rated current of the motor is 2.5 A, the output power is 250 W. The field excitation of the motor is provided by the 220 V AC power source which is commutated by the speed regulator, the armature current of the motor is also provided by the same power source. It is adjusted by the compressor governor. Through adjusting the output voltage of compressor governor, the motor could be of step-less speed regulated at the range 0 10000 r/min, the rate of velocity increasing could be 800 r/min. The length of the experimental equipment is 1200 mm, the width is 108 mm, the mass is about 45 kg, the diameter of the shaft is 9.5 mm, the length of the shaft is 500 mm and the maximum deflection is less than 0.005 0.015 mm. Any position along the axial direction could be selected as experimental abutment point. The diameter and mass of the experimental rotating table is 7619 mm and 600 g, respectively. The arrangement of experimental equipment is shown in Fig.1. Fig. 1 The arrangement of experimental equipment The electrical vortex sensors are used to measure the relatively displacement or vibration for axis to bearing pedestal. They are installed in the x and y directions at the sensor support, respectively. They do not touch the shaft, and could be used to directly measure the vibration signal of the rotation shaft. The flashing phase-measurer is used to measure the rotator speed and the phase of the shaft. 2 THE FREQUENCY SPECTRUM ANALYSIS OF VIBRATION SIGNAL The real vibration of turbine generator set is the most of simple harmonic periodic motion. Its wave type is also made of many simple harmonic motion. In order to analysis the vibration, we should study the wave type, the frequency composition of the vibration, and the amplitudes. Frequency can be used as x-axis to describe the vibration in the frequency-domain. The method decomposing the vibration into its various frequency components in frequency components in the frequency domain is called frequency spectrum analysis. The purpose of frequency spectrum analysis is to decompose the signals into different compositions. So the vibration becomes the simple harmonic motion including different amplitudes, frequencies and phases. In the frequency spectrum of rotator vibration, the different frequency is often corresponding to the different reason. If we can find the frequency composition of the vibration signal, the reason of vibration will be discovered. There are about 80% accidents caused by rotator non-equilibrium in the vibration accidents of turbine generator set happened in the spot, and 90% accidents caused by rotator mass non-equilibrium. In the experiments described in this paper we study the vibration caused by rotator mass non-equilibrium by using the method of frequency spectrum analysis and the influence coefficient method of finding equilibrium to determine the position of rotator mass non-equilibrium. 3 EXPERIMENTAL RESULTS AND ANALYSIS The Bode diagram shown as Fig. 2 is about the horizontal vibration characteristics of rotator. It shows that the critical velocity of rotator is about 2605 r/min, the maximum amplitude of rotator corresponding to the bearing shell is 371 m, the amplitude of select frequency (AMP-1X) is 360 m, and the phase difference (PHA-1X) is -36. The mark position in Fig. 3 is the maximum value of vibration, and also is the position where the phase angle changes. When the rotator speed is smaller than 2152 r/min, the relative phase angle is -130. When the rotator speed increases to 2605 r/min, the amplitude will increase to the maximum value. That is to say, the vibration amplitude increases rapidly in the rotator speed range 2152 2605 r/min. Fig. 2 The Bode diagram of rotator horizontal vibration characteristic The phase change )94)130(36( is slightly greater than 90. According to the eccentric forced vibration theory for single free dimension, the case in the Fig.2 is caused by resonance. The maximum amplitude occurs at the position where the phase change is slightly greater than 90, because of the damping. In this case, the angle frequency of rotator speed equals to that of exciting force, that is, 1 . The angle frequency of rotator speed can be considered as first order critical rotator speed. The critical rotator speed is related to material of rotator, geometrical shape, size, structure, and supporting conditions etc. It is the inherent characteristics of the rotator system, and it is not related to the external conditions. The influence factors are mainly temperature and supporting rigidity. The experiments have been done under the condition of constant temperature. So supporting rigidity is the main influence factor. The experiments have been done by using the method of adding the non-equilibrium mass. The location adding non-equilibrium mass is determined by influence coefficient method that finding equilibrium. The non-equilibrium mass can change the system stiffness. But, according to m, when the m is very small, 1 will almost not change. If mm increases, 1 will be decreased. The first order critical rotator speed of the system will be decreased. These can be obtained from the experiments. The frequency spectrum figures show that the amplitude is obviously large when the frequency is one times (1X). It is monotonously increasing before the speed reaches critical rotator speed. The cases of frequency two times (2X), three times (3X) and four times (4X) all exist, but these amplitudes are very small and can be neglected. It is impossible that rotating table or system axis appear crosswise cracking when the experimental velocity is not high enough. The main reason causing vibration is that the mass of rotating table around axis is not uniform. Fig. 3 and Fig. 4 are the frequency spectrum diagrams that obtained through adding the different non-equilibrium mass to rotator system, respectively. Fig. 3 shows that the first order critical rotator speed is basically not changed, and is 2312 r/min, but the amplitude of frequency decreases. The horizontal amplitude of one times (1X) decreases to 145 m, and the vertical amplitude decreases to 134 m. For Fig. 4, they are 116 m and 87 m, respectively. The horizontal amplitude of 1X in Fig. 3 is decreases to 145 m, the vertical amplitude is decreased from 360 m to 134 m. The result that the second adding non-equilibrium mass is shown in Fig. 4. From this case, we can see: the 1X component is very obvious in the frequency spectrum. There are components for 2X, 3X, 4X, but their amplitudes are very small. They are not the main components of vibration. The 1X amplitude changes very small when change the system stiffness. This is determined by the location of adding non-equilibrium mass. The vibration amplitude can be effectively controlled only by calculation to find out the non-equilibrium point and non-equilibrium mass, then adding the same equilibrium mass at its opposite direction. Fig. 3 The frequence spectrum diagram of the first adding non-equilibrium mass Fig. 4 The frequence spectrum diagram of the second adding non-equilibrium mass 4 CONCLUSION (1) In the frequency spectrum figure, the one times frequency 1X components is too large. When the malfunction about bearing pedestal stiffness and axis joint join defect is not considered. The reason why the vibration is greater is the rotator non-equilibrium mass. (2) The one times frequency 1X amplitude is decreased by changing system stiffness. The decreasing amplitude is determined by the location of adding non-equilibrium mass. (3) The location of non-equilibrium mass is determined by the influence coefficient method. It is needed to find the non-equilibrium point and non-equilibrium mass by calculation. Then add the same equilibrium mass at the opposite direction. (4) The adding non-equilibrium mass is so small that it can not cause the large change of the system first order critical velocity. 頻譜分析在轉(zhuǎn)子動平衡中的應(yīng)用 摘 要 在模擬旋轉(zhuǎn)機械振動的實驗裝置上，通過不同的選擇來模擬機器的運行狀態(tài)，對單跨集中載荷情況下轉(zhuǎn)子由于不平衡質(zhì)量引起的振動進行了分析和討論，并用頻譜分析的方法找到了有效的解決辦法。 介紹 在傳統(tǒng)的島嶼核電站中， 汽 輪發(fā)電機組是一種非常重要 的核 熱能轉(zhuǎn)換成電能的設(shè)備。當 汽輪 發(fā)電機組 經(jīng)長時間運轉(zhuǎn)后 ,原來的 系統(tǒng) 平衡會因 金屬的 殘余變形、部件的 磨損或損壞 而遭到破環(huán) 。結(jié)果 ,系統(tǒng)的機械 振動將會 因此 增加。 所以 因此有必要 進行現(xiàn)場平衡 調(diào)整。另一方面 ,一個大型汽輪發(fā)電機組 在制作工序中 也需要調(diào)整平衡、調(diào)試、安裝和運行?，F(xiàn)場動平衡技術(shù)是消除 汽輪發(fā)電機組 劇烈振動的 一種重要的手段 。 我們可以 通過對汽輪發(fā)電機組的振動進行分析或做一些特殊的實驗 明確 了解振動的類型、振動動力源和振動特性 。 當 獲得 振動信號 之后， 頻譜可以用來分析振動信號，以便 迅速 診斷。利用頻譜分析 由振動感應(yīng)器獲得的 電機 的振動 信號，并 將廣泛的頻率范圍 分解為幾個主要的頻率成分。不同頻率成分 對汽輪發(fā)電機組有著不同的影響。頻譜分析是 研究汽輪發(fā)電機組振動的一個非常有用的方法 。 本文將對 由于 集中載荷引起的轉(zhuǎn)子質(zhì)量不平衡 進行 討論和分析 ， 并且給出了一個 利 用 頻譜分析 有效 防止振動 的方法。 1 實驗設(shè)備 實驗系統(tǒng)由電機、聯(lián)軸器及轉(zhuǎn)子等 構(gòu)成 ， 它的結(jié)構(gòu)是非常簡單的。轉(zhuǎn)子 由 電機直接驅(qū)動 ， 它的轉(zhuǎn)速可 進行大范圍 調(diào)節(jié)。該系統(tǒng)可以順利 、可靠的運作 。電機的額定電流為 2.5 A,輸出功率是 250 W。 電動機 的外部 勵磁 由 220 V 交流電源 經(jīng) 調(diào)節(jié)器整流后 提供 的 ,發(fā)動機的 電樞電流也 是 相同 的電源 提供 的 。它是由壓縮機 調(diào)節(jié)器進行 調(diào)整 控制 ， 通過調(diào)整壓縮機輸出電壓， 電動機 的速度 可逐步減少調(diào)節(jié) 至 范圍 0 10000 轉(zhuǎn) /分 ， 速度遞增可達 800 r /分。 實驗設(shè)備的長度是 1200 毫米 ,寬度為 108 毫米 ,質(zhì)量是大約 45 公斤 , 軸的 直徑是 9.5 毫米 ， 軸的 長度 500 毫米 ,最大撓度小于 0.005 0.015 毫米。 可以選擇沿軸向的 任何位置作為實驗的支承點。實驗 用 轉(zhuǎn) 盤的直徑和質(zhì)量分別 為 7619毫米 ,重 600 克。實驗設(shè)備的 安裝如 圖 1 所示 。 圖 1 實驗設(shè)備 的安裝 電渦流傳感器是用來測量 中軸 相對 軸承底座的 位移或振動的。 傳感器 分別 被安裝在 X 和 Y 方向 提供信號傳遞 。 它 們不接觸 軸 ,但 可以直接用于測量 轉(zhuǎn)動軸的振動信號。 閃動相位測量儀 是用來測量 轉(zhuǎn)子 速度 和傳動軸的相位 。 2 振動 信號 頻譜分析 輪機發(fā)電機組 真正的振動的 是 最簡單的諧波周期運動 。它的波型也是由許多簡單的諧波運動 構(gòu)成 。為了分析振動 ,我們應(yīng)該學(xué)習 振動的 的波型、頻率 組成 和振幅。 在頻域中 頻率 可以被看 作 X 軸來 描述振動。 在頻域中將 振動 分解 成 各種頻率成分的方法 叫做頻譜分析。頻譜分析的目的是為了 將振動信號 分解成不同的信號成分。所以振動 被分解 成為諧波運動包括不同的振幅、頻率和階段。 在轉(zhuǎn)子振動的頻譜 中 ,不同的頻率通常是對應(yīng)于不同的原因。如果我們能找到振動信號的 頻率成分 ，也 就 會發(fā)現(xiàn)引起 振動 的原因 。 在輪機發(fā)電機組工作現(xiàn)場 有大約 80%的 振動 事故 是由轉(zhuǎn)子不平衡 引起的 ,90%的事故 是 由 轉(zhuǎn)子質(zhì)量不平衡 引起的 。在本文中所描述的實驗 中 我們會采用頻譜分析的方法 研究轉(zhuǎn)子質(zhì)量平衡引起的振動 和 用 影響系數(shù)法 找出轉(zhuǎn)子質(zhì)量不平衡 的位置 。 3 實驗結(jié)果及分析 圖 2所示伯德 圖 是關(guān)于轉(zhuǎn)子的 橫向振動特性。 圖示轉(zhuǎn)子的 臨界速度 約是 2605 r /分鐘 ,轉(zhuǎn)子相對應(yīng)軸承殼 最大限度的 振幅為 371m,選擇頻率 (AMP-1X)是360m、相位差 (PHA-1X)是 -36。 在圖 2 中 標記位置是最大 幅度的振動 ,也就是振動相位變化的位置。當轉(zhuǎn)子速度小于 2152 r /分鐘 ,相對相位是