工業(yè)清灰裝置設(shè)計【雙聯(lián)旋風(fēng)除塵器】

工業(yè)清灰裝置設(shè)計【雙聯(lián)旋風(fēng)除塵器】,雙聯(lián)旋風(fēng)除塵器,工業(yè)清灰裝置設(shè)計【雙聯(lián)旋風(fēng)除塵器】,工業(yè),裝置,設(shè)計,旋風(fēng),除塵器 Rotor Brake A mechanical breaking system is besides the aerodynamic breaking function of the rotor an unavoidable component of a wind turbine. It is part of the mechanical drive train. The first task is to keep the rotor of a wind turbine in position when it is at a standstill. Locking the rotor is a must for servicing and repair work and is generally common practice during normal down times. Moreover, most turbines have locking bolts between rotor hub and nacelle for bridging extended periods of standstill and for servicing and repair work. The rotor can thus be secured in one or more positions. Rotor brakes are almost always disk brakes. Suitable disk brakes can frequently be adopted cost-effectively from existing production runs intended for other machines or vehicles .Against this background, the design of the rotor brake itself poses few problems. Nevertheless, the rotor brake presents the systems designer of a wind turbine with issues which have consequences for the entire system. The first and most important question is, which task the rotor brake is to fulfill within the operating concept. In the simplest case, its role is restricted to a mere holding function during rotor standstill. In this case, the brake must be dimensioned for the required holding torque of the rotor during standstill. This is determined in accordance with the aerodynamic forces calculated to occur at the assumed maximum wind speeds (Chapt.6.3.2). Apart from its function as a pure rotor parking brake, the rotor brake can also be dimensioned as a service brake. As long as the braking torque and braking power(thermal loading) can be absorbed, the mechanical rotor brake can be used as a second independent raking system in addition to aerodynamic rotor braking and the operational reliability of the wind turbine is considerably improved in this way. In small wind turbines, a mechanical rotor brake, which in cases of emergency prevents rotor runaway, has proved to be extraordinarily successful and is widely used today. With increasing turbine size, it becomes more and more difficult to meet this requirement. For a turbine with a rotor diameter of 60 to 80 m, the rotor brake takes on almost absurd dimensions if it is to brake the rotor torque and power during full-load operation. For this reason, the task of the rotor brake in large turbines is always restricted to the function of pure parking brake. Apart from the issue of the rotor brake’s task with respect to operations, there is the question of where in the drive train the rotor brake is best installed. The alternatives are for the rotor brake to be on the” low-speed” or on the” high-speed” side of the gearbox. In most turbines, efforts to keep the brake disk diameter as small as possible lead to the rotor brake being installed on the high-speed shaft, i. e. between gearbox and generator(Fig. 8.31). Owing to the higher rotational speed, the torque is one or even two orders of magnitude lower than at the slower rotor shaft, depending on the gear ratio. However, mounting the brake on the high-speed shaft has at least two disadvantages. It is inferior from the point of view of safety, since the braking function fails if the low-speed shaft or the gearbox break down. Moreover, the rotor must be held by the gears during a standstill. Gears react with increased wear of the tooth flanks to small oscillating movements, which are unavoidable in a stopped wind turbine due to air turbulence. In some turbines, it is attempted to solve this problem by no longer locking the rotor during standstill but by letting it” spin” at low speed. To avoid these disadvantages, the rotor brake was installed on the low-speed rotor shaft in some earlier systems. In small wind turbines a fully effective operating brake can be implemented with justifiable effort on the low-speed side, as long as design of the rotor shaft bearing assembly does not present an obstacle. The rotor brake on the low-speed side was a common feature of many earlier stall-controlled Danish wind turbines up to a power rating of about 100 kW in the Eighties. At that time it was considered to be an extra safety element even though the rotor brake was only designed as a parking brake. Installing the rotor brake on the slow side is much more problematic in large wind turbines, however. Even a parking brake already assumes a considerable size (Fig. 8.32).These disadvantages have led to the rotor brake being arranged on the high-speed side behind the gearbox in almost all new systems. 8.8 Gearbox The conversion of the greatly differing rotational speeds of the rotor and the electric generator has given the designers of the first wind turbines many headaches. This situation led to costly low-speed generator designs and to hydraulic or pneumatic transmission systems to the generator (Chapt.8.1).Aerodynamicists made efforts to drive the rotor speed as high as possible in order to lower the gear ratio. It was assumed that costs would also increase considerably with increasing gear ratios, so that the development of rotors with extremely high tip-speed ratios was pushed forward. This situation has changed with the progress which has been made in gearbox technology. Today, high-performance gearboxes with gear ratios of up to 1:100 and more are available. In many areas of mechanical engineering, gearboxes are used which are suitable for deployment in wind turbines, as regards their technical concept, their efficiency and their operating life. The gearbox for the wind turbine has become a” vendor-supplied component”, which, with certain adaptations, can be taken from the standard product range of the gearbox manufacturers. Regardless of this favorable situation, the gearbox has been and still is a source of failures and defects in many wind turbines. The cause of these “gearbox problems” is not so much the gearbox itself, rather the correct dimensioning of the gearbox with regard to the load spectrum. In wind turbines, it is easy to underestimate the high dynamic loads to which the gearbox is subjected. Thus, in the early phase, many turbines had gearboxes which were undersized. Having learned their lessons, successful manufacturers equipped their turbines with ever stronger gearboxes and thus, in the course of development, empirically arrived at the right dimension. 8.8.1Gearbox Configurations Toothed-wheel gearboxes are constructed in two different forms. One is the parallel shaft or spur-gear system, the other is the technically more elaborate planetary gearing. The gear ratio per single reduction is limited, so that the difference in diameter between the small and the large wheel does not become too unfavorable. Parallel-shaft-gear stages are built with a gear ratio of up to 1 :5, whereas planetary stages have a gear ratio of up to 1 : 12. Wind turbines generally require more than one stage. Fig. 8.33 shows what effects different designs have on gearbox size, mass and relative cost [11]. It is noteworthy that the three-stage planetary design has only a fraction of the overall mass of a comparable parallel shaft system. The relative costs are reduced to about one half. In the megawatt power class, the multi-stage planetary gearbox is, therefore, clearly superior. In smaller power classes, the comparison is not quite as unambiguous. In the range up to about 500 kW, parallel-shaft gear designs are often preferred for cost reasons. Small wind turbines are equipped with parallel-shaft gear systems.Theprevailingmodels are two-stage gearboxes which are commercially available from numerous manufacturers as modified universal transmissions (Fig. 8.34). In larger wind turbines, the planetary design definitely prevails. For outputs of several megawatts, two- or three-stage models are used (Fig. 8.35). Large gearboxes of this type are used, for example, in ship-building and several other fields of mechanical engineering, so that suitable gearboxes for large wind turbines can be derived from these production sources. Gearboxes with one planetary stage and two additional parallel-shaft stages are used in many late-model turbines (Fig. 8.36).With the additional parallel shaft, the primary and secondary shafts are no longer coaxial. This has the advantage that a hollow through shaft can be implemented more easily. In this way, power supply lines supplying power to the blade pitch drive, as well as measurement and control signals for the rotor, can be routed through the gearbox. In larger gearboxes, an auxiliary rotor drive is frequently flanged to the gearbox housing. Using this electric motor, the rotor can be turned slowly. Such an auxiliary unit is indispensable for assembly and maintenance work in large rotors. Gearbox lubrication is usually carried out via a central oil supply in the nacelle. As a rule, it also contains an oil cooler and a filter. In spite of indisputable advances having been achieved in the durability of the gearboxes, there is still “trouble with the gears” being experienced even in the latest wind turbines. Although it is possible to adapt gearboxes for wind turbines from other types of machine, they are subject to special demands which are often not encountered in other applications. Much negative experience in recent years has provided important insights into this issue: – Special attention must be devoted to the smooth running of the tooting. Particularly prominent gear meshing frequencies can cause resonances in the drive train.“Cheap” transmissions with simple tooting are unsuitable for use in wind turbines. – Oil leaks in the transmission are a particular problem. Labyrinth seals have proven more reliable than slipping type seals. In many cases, the housing flanges also showed leaks after some time. A box design with a top flange is apparently more advantageous than gearbox housings with flanges on the input and output side. – The quality of the lubrication has been found to be a decisive factor for the service life of the gearbox. Oil temperatures which are too high cause just as much damage as does contamination in the oil. Oil coolers and filters are indispensible for large gearboxes ands is the careful observance of oil change intervals. – The stiffness of the gearbox housing is an important criterion for its service life if the housing is integrated into the static design of the nacelle. Apart from these constructional measures, of course, the correct dimensioning has a decisive influence.  風(fēng)機剎車裝置 機械制動裝置作為主傳動鏈的一部分，同具有氣動剎車功能的轉(zhuǎn)子一樣，是風(fēng)力發(fā)電機一個不可或缺的組成部分。其首要任務(wù)是當(dāng)風(fēng)機停轉(zhuǎn)時，使風(fēng)機轉(zhuǎn)子處于適當(dāng)?shù)奈恢谩Ｔ诰S修工作的過程中，鎖定主軸是必須的，而且在正常的停工期間，鎖定主軸也是慣常的做法。更為重要的是，為了渡過持續(xù)的主軸停轉(zhuǎn)時期以及維修工作的進行，大多數(shù)風(fēng)機都會在輪轂和機艙之間安裝鎖緊螺栓，從而使主軸在一個或者多個位置得到保護。 主軸剎車通常采用盤式剎車，適當(dāng)?shù)谋P剎可以頻繁采用為其它的機器或裝置設(shè)計的現(xiàn)存的生產(chǎn)線，這樣就可以節(jié)約成本。在這樣的背景下，主軸剎車的設(shè)計本身不會產(chǎn)生很多問題。然而，主軸剎車也會給風(fēng)機系統(tǒng)的設(shè)計者帶來一些問題，而這些問題可能給整個系統(tǒng)帶來一些后果。 最為重要的問題是，在經(jīng)營理念之內(nèi)，剎車裝置要完成哪項任務(wù)。在最簡單的情況下，在主軸停轉(zhuǎn)的情況下，剎車裝置的尺寸必須滿足所需的支持轉(zhuǎn)矩。這是由假設(shè)風(fēng)速達(dá)到最大時所計算的空氣動力所決定的（第6.3.2節(jié)）。 剎車裝置除了單純地具有主軸駐車制動功能外，尺寸合適時，其也可以作為停車制動裝置。只要制動力矩和制動力（熱負(fù)載下）能夠被吸收，主軸機械剎車也可以用作除了氣動主軸剎車外的第二獨立剎車系統(tǒng)，這樣一來，風(fēng)機的運作可靠性得到大幅提高。對于小型風(fēng)機，在緊急情況下，主軸機械剎車會防止主軸失控，這一點被證實是非常成功的，并得到了廣泛的應(yīng)用。 隨著風(fēng)機尺寸的增大，這一需求會越來越難以得到滿足。對于主軸直徑為60米到80米的風(fēng)機來說，如果是在滿載運轉(zhuǎn)期間制動主軸轉(zhuǎn)矩和主軸功率，主軸剎車裝置幾乎承擔(dān)了離譜的規(guī)模。因此，在大型風(fēng)機中主軸剎車裝置的作用總是被限制在單純的進行駐車制動的上。 除了主軸剎車裝置的運轉(zhuǎn)任務(wù)的問題外的另一個問題是，剎車裝置安裝在主傳動鏈的哪個位置是最合適的。兩種可以選擇的方案分別是安裝在齒輪箱的低速軸和高速軸。在大多數(shù)的風(fēng)機中，為了使得剎車盤的直徑盡可能地小，常將剎車裝置安裝在高速軸上，比如安裝在如齒輪箱和發(fā)電機之間（圖8.31）。根據(jù)齒輪的傳動比，由于轉(zhuǎn)速較高，轉(zhuǎn)矩會比在低速軸上低一個甚至兩個數(shù)量級。 然而，將剎車裝置安裝在高速軸上至少有兩點缺點。由于如果低速軸或者齒輪箱出現(xiàn)故障時，剎車功能就會失效，所以從安全的角度上，這種方案是處于劣勢的。再者，在主軸停轉(zhuǎn)的過程中，其必須由齒輪支撐。隨著齒輪齒側(cè)的逐漸磨損，齒輪會產(chǎn)生一些小的震動，由于空氣擾動，對于一個一個停止工作的風(fēng)機來說，這一點是不可避免的。在有些風(fēng)機中，為了試圖解決這一問題，主軸停轉(zhuǎn)時，不再鎖定主軸，而是允許它在低速下轉(zhuǎn)動。為了避免以上缺點，早期的主軸剎車系統(tǒng)被安裝在低速軸上。在小型風(fēng)機中，只要主軸軸承裝置的設(shè)計不會出現(xiàn)問題，也可以通過無可非議的努力在低速端實施一個完全有效可操作的剎車方案。80年代，丹麥許多功率高達(dá)100kw的失速控制型風(fēng)機普遍采用主軸剎車安裝在低速軸的方案。那是，即使主軸剎車裝置僅僅被設(shè)計為駐車剎車裝置，這一方案也被看作是一個額外的安全因素。然而，在大型風(fēng)機中，將主軸剎車裝置安裝在低速軸是非常有爭議的。即使假設(shè)剎車裝置有一個相當(dāng)大的尺寸（圖8.32）。在幾乎所有新的系統(tǒng)中，這些缺點已經(jīng)引導(dǎo)主軸剎車裝置被安裝在齒輪箱后部的高速軸上。 圖8.31NORDEX N-80型風(fēng)機中的主軸制動裝置安裝在齒輪箱的高速軸上 圖8.32早期HOEDEN HWP-1000型風(fēng)機中 主軸制動裝置直接安裝在輪轂后的低速軸上 8.8齒輪箱 齒輪箱的布置 齒輪箱有兩種布置方式。一種是平行軸傳動或直齒輪傳動系統(tǒng)，另一種是技術(shù)上更為精確的行星輪傳動。單級減速的傳動比是有限制的，所以大小齒輪直徑的之間的差異并沒有十分不利。平行軸齒輪傳動級的傳動比最高達(dá)1:5，而行星輪傳動比可高達(dá)1:12。總的來說，風(fēng)機不只需要一個傳動級。圖8.33顯示不同的設(shè)計對齒輪箱的尺寸，質(zhì)量以及相應(yīng)的花費有什么影響。 圖8.33 不同齒輪箱設(shè)計的總體質(zhì)量和相對費用 顯然，三級行星輪系的的質(zhì)量只是相應(yīng)的平行軸傳動的總體質(zhì)量的一小部分。相應(yīng)的費用也縮小的接近一半左右。因此，在兆瓦級風(fēng)機中，多級行星輪系占有明顯的優(yōu)勢。而在小型分幾種，這種比較的結(jié)果則不是那么明顯。在500kw以下的風(fēng)機中，平行軸齒輪箱因其花費較少而多被采用。 小型風(fēng)機中采用平行軸齒輪傳動系統(tǒng)。普遍采用的模型是兩級傳動的齒輪箱，其是經(jīng)改造而成的一種通用的尺寸（圖8.34），由很多制造商提供。 圖8.34.200-500kw級風(fēng)機的兩級平行軸齒輪箱 而在大型風(fēng)機中，行星輪系的設(shè)計則非常普遍。對于產(chǎn)電量為幾兆瓦的風(fēng)機，常采用兩級或者三級行星輪系（圖8.35）。例如，這種大型齒輪箱在制船業(yè)以及機械工程的其它幾個領(lǐng)域都有采用，因此，大型風(fēng)機中的合適的齒輪箱都可以以這些產(chǎn)品為參照。在很多后現(xiàn)代的風(fēng)機中，多采用由一級行星兩級平行軸組成的齒輪箱（圖8.36）。由于有多加的平行軸，一級和二級的軸不再同軸。這樣布置的好處是可以很容易的應(yīng)用一個中空的軸。這樣，向葉片供電的供電線路以及主軸的控制和測量信號可以通過齒輪箱路由。 在大型風(fēng)機中，一個輔助的主軸裝置常通過法蘭和齒輪箱體連接。用這種點機，主軸轉(zhuǎn)速會變慢。這樣一個輔助裝置對大型主軸的組裝和維修工作是必不可少的。通常來說，它也帶有一個油液冷卻裝置和一個濾油器。 圖8.35 2-3MW級封系的三級行星輪系齒輪箱 盡管齒輪箱在耐久性上已經(jīng)取得了無可爭議的優(yōu)勢，即使在最新型的風(fēng)機中，齒輪的問題仍然存在。雖然從其它類型的機器中改編為機的齒輪箱是可以做到的，但是這些常用于特殊需求，在其它的應(yīng)用中并不常見。近幾年來，很多失敗的經(jīng)驗已經(jīng)對這個問題提供了重要的見解。我們一定要特別注意輪齒的平穩(wěn)運行。特別是齒輪嚙合的頻率可能造成主傳動鏈共振。制造簡單的變速器，其輪齒很簡單，并不適用于風(fēng)機中。箱體的漏油是個特殊的問題。經(jīng)證實，迷宮型密封比滑環(huán)型密封更可靠。很多情況下，有時箱體的法蘭也會造成漏油。將法蘭安裝在上部明顯比安裝在輸入或輸出端更有優(yōu)勢。人們發(fā)現(xiàn)，潤滑油的質(zhì)量是影響齒輪箱壽命的決定性因素。油溫過高和油被污染會造成同樣多的損害。對于大型風(fēng)機，油溫降溫裝置和濾油器是必不可少的，所以要小心遵守?fù)Q油期。如果箱體和機艙的靜態(tài)設(shè)計結(jié)合，箱體的硬度是其壽命的一個重要指標(biāo)。當(dāng)然，除了這些構(gòu)造措施，正確的尺寸標(biāo)注有這決定的影響。 14