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編號 無錫太湖學院 畢業(yè)設計（論文） 相關資料 題目： 80系列微型風冷活塞式壓縮機 W80II型 信機 系 機械工程及自動化專業(yè) 學 號： 　　　 0923105　　　 學生姓名： 肖秋紅 指導教師： 　 俞萍 （職稱：高級工程師 ） （職稱： ） 2013年5月25日 目 錄 一、畢業(yè)設計（論文）開題報告 二、畢業(yè)設計（論文）外文資料翻譯及原文 三、學生“畢業(yè)論文（論文）計劃、進度、檢查及落實表” 四、實習鑒定表 無錫太湖學院 畢業(yè)設計（論文） 開題報告 題目： 80系列微型風冷活塞式壓縮機 W80II型 機 電 系 機械工程及自動化 專業(yè) 學 號： 0923105 學生姓名： 肖秋紅 指導教師： 俞 萍（職稱：工程師 ） （職稱： ） 2012 年 11月 25日 課題來源 “80系列微型風冷活塞式壓縮機”的課題來源于企業(yè)； 結合所學知識，老師擬定題目； 綜合大學里所學知識，將理論與實踐相互結合。 科學依據(jù)（包括課題的科學意義；國內(nèi)外研究概況、水平和發(fā)展趨勢；應用前景等） 1、 化工、冶金、化肥、食品、醫(yī)療等眾多企業(yè)的生產(chǎn)過程需要用到氣體 壓縮機，而活塞式空氣壓縮機由于有較高的壓縮比，在高壓氣體生產(chǎn) 與輸送中尚不能被其它設備所替代，是許多工程項目中的關鍵設備。 2、 近幾十年來，我國壓縮機制造業(yè)在引進國外技術，消化吸收和自主開 發(fā)基礎上，克服不少難關，取得重大突破，其中活塞式壓縮機已達到 國際同類產(chǎn)品的水平。今后壓縮機的發(fā)展前景不僅僅在于努力提高技 術性能指標，更應著力于應用近代先進計算機技術進行性能模擬和優(yōu) 化設計，促成最佳性能的系列化、通用化、機組化和自動化，降低生 產(chǎn)成本，完善輔助成套設備，擴大應用領域，提高綜合技術經(jīng)濟指標。 3、 微型風冷活塞式壓縮機結構簡單，成本低，安裝方便，是當前活塞式 壓縮機的發(fā)展方向。 4、目前壓縮機制造業(yè)已經(jīng)發(fā)展成為機械制造工業(yè)的一個重要組成部分。 研究內(nèi)容 1、 80系列微型風冷活塞式壓縮機的工作原理以及工作形成； 2、 80系列微型風冷活塞式壓縮機參數(shù)與結構的設計； 3、 80系列微型風冷活塞式壓縮機設計圖紙的繪制。 研究計劃及預期成果 1、 首先對80系列微型風冷活塞式壓縮機整體結構進行分析，對傳動結構進行篩選，初步選擇達到設計要求的結構方案； 2、 對壓縮機的熱力部分及動力部分進行計算，通過壓縮機機構的分析計算可提高其自身的精度； 3、 對80系列微型風冷活塞式壓縮機的主要零件進行強度校核，提高機構穩(wěn)定性，穩(wěn)定性。 特色或創(chuàng)新之處 通過對80系列微型風冷活塞式壓縮機的設計及計算，形成一整套現(xiàn)代的設計方法，對理論和實踐的結合，起到整體的規(guī)劃的作用，達到降低損耗提高效率，優(yōu)化結構設計方便使用。 已具備的條件和尚需解決的問題 已具備的條件：擁有機械設計手冊等參考資料及文獻；到企業(yè)進行參觀， 對空氣壓縮機進行直觀的了解與認識，對所學的機械基礎 知識有較好的掌握；能熟練運用CAXA制圖軟件，提高 作圖效率。 尚需解決的問題：對于80系列微型風冷活塞式壓縮機的工作原理不是非常清楚和熟悉，缺乏設計經(jīng)驗。 指導教師意見 指導老師簽名： 年 月 日 教研室（學科組、研究所）意見 教研室主任簽名： 年 月 日 系意見 主管領導簽名： 年 月 日 英文原文 Efficiency And Operating Characteristics Of Centrifugal And Reciprocating Compressors Reciprocating compressors and centrifugal compressors have different operating characteristics and use different eificiency definitions. This article provides guidelines for an equitable comparison, resulting in a universal efficiency definition for both types of machines. The comparison is based on the requirements in which a user is ultimately interested. Further, the impact of actual pipeline operating conditions and the impact on efficiency at different load levels is evaluated. At first glance, calculating the efficiency for any type of compression seems to be straightforward: comparing the work required of an ideal compression process with the work required of an actual compression process. The difficulty is correctly defining appropriate system boundaries that include losses associated with the compression process. Unless these boundaries are appropriately defined, comparisons between centrifugal and reciprocating compressors become flawed. We also need to acknowledge that the efficiency definitions, even when evaluated equitably, still don't completely answer one of the operator's main concerns: What is the driver power required for the compression process?To accomplish this, mechanical losses in the compression systems need to be discussed. Trends in efficiency should also be considered over time, such as off-design conditions as they are imposed by typical pipeline operations, or the impact of operating hours and associated degradation on the compressors. The compression equipment used for pipelines involves either reciprocating compressors or centrifugal compressors. Centrifugal compressors are driven by gas turbines, or by electricmotors. The gas turbines used are, in general,two-shaft engines and the electric motor drives use either variable speed motors, or variable speed gearboxes. Reciprocating compressors are either low speed integral units, which combine the gas engine and the compressor in one crank casing,or separable "high-speed" units. The latter units operate in the 750-1,200 rpm range (1,800 rpm for smaller units) and are generally driven by electric motors, or four-stroke gas engines. Efficiency To determine the isentropic efficiency of any compression process based on total enthalpies (h), total pressures (p), temperatures (T)and entropies (s) at suction and discharge of the compressor are measured, and the isentropic efficiency r\^ then becomes: (Eq.1) and, with measuring the steady state mass flow m, the absorbed shaft power is: (Eq.2) considering the mechanical efficiency r\^. The theoretical (isentropic) power consumption (which is the lowest possible power consumption for an adiabatic system) follows from: (Eq.3) The flow into and out of a centrifugal compressor can be considered as "steady state."Heat exchange with the environment is usually negligible. System boundaries for the efficiency calculations are usually the suction and discharge nozzles. It needs to be assured that the system boundaries envelope all internal leakage paths,in particular recirculation paths fi^om balance piston or division wall leakages. The mechanical efficiency r)^.,, describing the friction losses in bearings and seals, as well as windage losses, is typically between 98 and 99%. For reciprocating compressors, theoretical gas horsepower is also given by Eq. 3,given the suction and discharge pressure are upstream of the suction pulsation dampeners and downstream of the discharge pulsation dampeners. Reciprocating compressors, by their very nature, require manifold systems to control pulsations and provide isolation from neighboring units (both reciprocating and centrifugal), as well as from pipeline flow meters and yard piping and can be extensive in nature.The design of manifold systems for either slow speed or high speed units uses a combination of volumes, piping lengths and pressure drop elements to create pulsation (acoustic) filters.These manifold systems (filters) cause a pressure drop, and thus must be considered in efficiency calculations. Potentially, additional pressure deductions from the suction pressure would have to made to include the effects of residual pulsations. Like centrifugal compressors, heat transfer is usually neglected. For integral machines, mechanical efficiency is generally taken as 95%. For separable machines a 97% mechanical efficiency is often used. These numbers seem to be somewhat optimistic, given the fact that a number of sources state that reciprocating engines incur between 8-15% mechanical losses and reciprocating compressors between 6-12%(Ref 1: Kurz , R., K. Brun, 2007). Operating Conditions For a situation where a compressor operates in a system with pipe of the length Lu upstream and a pipe of the length Ld downstream, and further where the pressure at the beginning of the upstream pipe pu and the end of the downstream pipe pe are known and constant, we have a simple model of a compressor station operating in a pipeline system (Figure 1). Figure 1: Conceptual model of a pipeline segment (Ref. 2: Kurz, R., M. Lubomirsky.2006). For a given, constant flow capacity Qstd the pipeline will then impose a pressure ps at the suction and pd at the discharge side of the compressor. For a given pipeline, the head (Hs)-flow (Q) relationship at the compressor station can be approximated by （Eq.4） where C3 and C4 are constants (for a given pipeline geometry) describing the pressure at either ends of the pipeline, and the friction losses, respectively(Ref 2: Kurz, R., M. Lubomirsky, 2006). Among other issues, this means that for a compressor station within a pipeline system, the head for a required flow is prescribed by the pipeline system (Figure 2). In particular, this characteristic requires the capability for the compressors to allow a reduction in head with reduced flow, and vice versa, in a prescribed fashion. The pipeline will therefore not require a change in flow at constant head (or pressure ratio). Figure 2: Stafion Head-Flow relationship based on Eq. 4. In transient situations (for example during line packing), the operating conditions follow initially a constant power distribution, i.e. the head flow relationship follows: （Eq.5） and will asymptotically approach the steady state relationship (Ref 3: Ohanian, S., R.Kurz, 2002). Based on the requirements above, the compressor output must be controlled to match the system demand. This system demand is characterized by a strong relationship between system flow and system head or pressure ratio.Given the large variations in operating conditions experienced by pipeline compressors, an important question is how to adjust the compressor to the varying conditions, and, in particular, how does this influence the efficiency. Centrinagal compressors tend to have rather flat head vs. flow characteristic. This means that changes in pressure ratio have a significant effect on the actual flow through the machine (Ref 4:Kurz, R., 2004). For a centrifugal compressor operating at a constant speed, the head or pressure ratio is reduced with increasing flow. Controlling the flow through the compressor can be accomplished by varying the operating speed of the compressor This is the preferred method of controlling centrifugal compressors. Two shaft gas turbines and variable speed electric motors allow for speed variations over a wide range (usually from 40-50% to 100% of maximum speed or more).It should be noted, that the controlled value is usually not speed, but the speed is indirectly the result of balancing the power generated by the power turbine (which is controlled by the fuel flow into the gas turbine) and the absorbed power of the compressor. Virtually any centrifugal compressor installed in the past 15 years in pipeline service is driven by a variable speed driver, usually a two-shaft gas turbine. Older installations and installations in other than pipeline service sometimes use single-shaft gas turbines (which allow a speed variation from about 90-100% speed) and constant speed electric motors. In these installations, suction throttling or variable inlet guide vanes are used to Drovide means of control. Figure 3: Typical pipeline operating points plotted into a typical centrifugal compressor performance map. The operating envelope of a centrifugal compressor is limited by the maximum allowable speed, the minimum flow (surge flow),and the maximum flow (choke or stonewall)(Figure 3). Another limiting factor may be the available driver power. Only the minimum flow requires special attention, because it is defined by an aerodynamic stability limit of the compressor Crossing this limit to lower flows will cause a flow reversal in the compressor, which can damage the compressor. Modem control systems prevent this situation by automatically opening a recycle valve. For this reason, virtually all modern compressor installations use a recycle line with control valve that allows the increase of the flow through the compressor if it comes near the stability limit. The control systems constantly monitor the operating point of the compressor in relation to its surge line,and automatically open or close the recycle valve if necessary. For most applications, the operating mode with an open, or partially open recycle valve is only used for start-up and shutdown, or for brief periods during upset operating conditions. Assuming the pipeline characteristic derived in Eq. 4, the compressor impellers will be selected to operate at or near its best efficiency for the entire range of head and flow conditions imposed by the pipeline. This is possible with a speed (N) controlled compressor, because the best efficiency points of a compressor are connected by a relationship that requires approximately (fan law equation): (Eq.6) For operating points that meet the above relationship, the absorbed gas power Pg is (due to the fact that the efficiency stays approximately constant): 　 (Eq.7) As it is, this power-speed relationship allows the power turbine to operate at, or very close to its optimum speed for the entire range.The typical operating scenarios in pipelines therefore allow the compressor and the power turbine to operate at its best efliciency for most of the time. The gas producer of the gas turbine will, however, lose some thermal efficiency when operated in part load. Figure 3 shows a typical real world example: Pipeline operating points for different flow requirements are plotted into the performance map of the speed controlled centrifugal compressor used in the compressor station. Reciprocating compressors will automatically comply with the system pressure ratio demands,as long as no mechanical limits (rod load power)are exceeded. Changes in system suction or discharge pressure will simply cause the valves to open earlier or later. The head is lowered automatically because the valves see lower pipeline pressures on the discharge side and/or higher pipeline pressures on the suction side. Therefore, without additional measures, the flow would stay roughly the same — except for the impact of changed volumetric efficiency which would increa.se, thus increasing the flow with reduced presstire ratio. The control challenge lies in the adjustment of the flow to the system demands. Without additional adjustments, the flow throughput of the compressor changes very little with changed pressure ratio. Historically, pipelines installed many small compressors and adjusted flow rate by changing the number of machines activated. This capacity and load could be fine-tuned by speed or by a number of small adjustments (load steps) made in the cylinder clearance of a single unit. As compressors have grown, the burden for capacity control has shifted to the individual compressors. Load control is a critical component to compressor operation. From a pipeline operation perspective, variation in station flow is required to meet pipeline delivery commitments, as well as implement company strategies for optimal operation (i.e., line packing, load anticipation).From a unit perspective, load control involves reducing unit flow (through unloaders or speed)to operate as close as possible to the design torque limit without overloading the compressor or driver The critical limits on any load map curve are rod load limits and HP/torque limits for any given station suction and discharge pressure.Gas control generally will establish the units within a station that must be operated to achieve pipeline flow targets. Local unit control will establish load step or speed requirements to limit rod loads or achieve torque control. The common methods of changing flow rate are to change speed, change clearance, or de-activate a cylinder-end (hold the suction valve open). Another method is an infinite-step unloader, which delays suction valve closure to reduce volumetric efficiency. Further, part of the flow can be recycled or the suction pressure can be throttled thus reducing the mass flow while keeping the volumetric flow into the compressor approximately constant. Control strategies for compressors should allow automation, and be adjusted easily during the operation of the compressor.In particular, strategies that require design modifications to the compres.sor (for example: re-wheeling of a centrifugal compressor, changing cylinder bore, or adding fixed clearances for a reciprocating compressor)are not considered here. It should be noted that with reciprocating compressors, a key control requirement is to not overload the driver or to exceed mechanical limits. Operation The typical steady state pipeline operation will yield an efliciency behavior as outlined in Figure 4. This figure is the result of evaluating the compressor efTiciency along a pipeline steady state operating characteristic. Both compressors would be sized to achieve their best efficiency at 100% flow, while allowing for 10% flow above the design flow. Different mechanical efficiencies have not been considered for this comparison. The reciprocating compressor efl'iciency is derived n-om valve efficiency measurements in Ref 5 (Noall, M., W. Couch, 2003) with compression efficiency and losses due to pulsation attenuation devices added. The efficiencies are achievable with low speed compressors. High speed reciprocating compressors may be lower in efficiency. Figure 4: Compressor Efficiency af different flow rates based on operation aiong a steady state pipeline characteristic. Figure 4 shows the impact of the increased valve losses at lower pressure ratio and lower flow for reciprocating machines, while the efficiency of the centrifugal compressor stays more or less constant. Conclusions Efficiency definitions and comparison between different types of compressors require close attention to the definition of the boundary conditions for which the efficiencies are defined as well as the operating scenario in which they are employed. The mechanical efficiency plays an important role when efficiency values are used to calculate power consumption. If these definitions are not considered, discussions of relative merits of different systems become inaccurate and misleading. 中文譯文 離心式和往復式壓縮機的工作效率特性 往復式壓縮機和離心式壓縮機具有不同的工作特性，而且關于效率的定義也不同。本文提供了一個公平的比較準則，得到了對于兩種類型機器普遍適用的效率定義。這個比較基于用戶最感興趣的要求提出的。此外，對于管道的工作環(huán)境影響和在不同負載水平的影響給出了評估。 乍一看，計算任何類型的壓縮效率看似是很簡單的：比較理想壓縮過程和實際壓縮過程的工作效率。難點在于正確定義適當?shù)南到y(tǒng)邊界，包括與之相關的壓縮過程的損失。除非這些邊界是恰好定義的，否則離心式和往復式壓縮機的比較就變得有缺陷了。 我們也需要承認，效率的定義，甚至是在評估公平的情況下，仍不能完全回應操作員的主要關心問題：壓縮過程所需的驅(qū)動力量是什么？要做到這一點，就需要討論在壓縮過程中的機械損失。 隨著時間的推移效率趨勢也應被考慮，如非設計條件，它們是由專業(yè)的流水線規(guī)定，或者是受壓縮機的工作時間和自身退化的影響。 　管道使用的壓縮設備涉及到往復式和離心式壓縮機。離心式壓縮機用燃氣輪機或者是電動馬達來驅(qū)動。所用的燃氣輪機，總的來說，是兩軸發(fā)動機，電動馬達使用的是變速馬達或者變速齒輪箱。往復壓縮機是低速整體單位或者是可分的“高速”單位，其中低速整體單位是燃氣發(fā)動機和壓縮機在一個曲柄套管內(nèi)。后者單位的運行在750-1,200rpm范圍內(nèi)（1,800rpm是更小的單位）并且通常都是由電動馬達或者四沖程燃氣發(fā)動機來驅(qū)動。 效率 　　要確定任何壓縮過程的等熵效率，就要基于測量的壓縮機吸入和排出的總焓(h)，總壓力(p)，溫度(T)和熵(s)，于是等熵效率變?yōu)椋? (Eq.1) 并且加上測量的穩(wěn)態(tài)質(zhì)量流m，吸收軸功率為： (Eq.2) 考慮機械效率。 理論（熵）功耗（這是絕熱系統(tǒng)可能出現(xiàn)的最低功耗）如下： (Eq.3) 流入和流出離心式壓縮機的流量可以視為“穩(wěn)態(tài)”。環(huán)境的熱交換通?？梢院雎?。系統(tǒng)邊界的效率計算通常是用吸入和排出的噴嘴。需要確定的是，系統(tǒng)邊界要包含所有內(nèi)部泄露途徑，尤其是從平衡活塞式或分裂墻滲漏的循環(huán)路徑。機械效率，在描述軸承和密封件的摩擦損失以及風阻損失時可以達到98%和99%。 對于往復式壓縮機，理論的氣體馬力也是由Eq.3給出的，鑒于吸力緩沖器上游和排力緩沖器下游的吸氣和排氣壓力脈動。往復壓縮機就其性質(zhì)而言，從臨近單位需要多方面的系統(tǒng)來控制脈動和提供隔離（包括往復式和離心式），以及可以自然存在的來自管線的管流量和面積管道。對于任何一個低速或高速單位的歧管系統(tǒng)設計，使用了卷相結合，管道長度和壓力降元素來創(chuàng)造脈動（聲波）濾波器。這些歧管系統(tǒng)（過濾器）引起壓力下降，因此必須在效率計算時考慮到。潛在的，從吸氣壓力扣除的額外壓力不得不包含進殘余脈動的影響。就像離心壓縮機一樣，傳熱就經(jīng)常被忽視。 對于積分的機器，機械效率一般取為95%。對于可分機機械效率一般使用97%。這些數(shù)字似乎有些樂觀，一系列數(shù)字顯示，往復式發(fā)動機機械損失在8-15%之間，往復壓縮機的在6-12%（參考1往復壓縮機招致號碼：庫爾茲，R.，K.，光布倫，2007）。 工作環(huán)境 在這樣的情況下，當壓縮機在一個系統(tǒng)中運行時，管道長度Lu上游和Ld下游，以及管道pu上游的初始壓力和管道pe下游的終止壓力均被視為常量，在管道系統(tǒng)中我們有一個壓縮機運行的簡單模型（圖1）。 圖1：管道段的概念模型（文獻2：庫爾茲.R，M.由羅穆斯基，2006年）。 對于給定的，標準管線定量流動能力將在吸入階段強加壓力，在壓縮機放電區(qū)強加壓力。對于給定的管線，壓縮機站頭部（）流（Q）關系可以近似表述為 （Eq.4） 其中和是常數(shù)（對于一個給定的管道幾何）分別描述了管道兩邊的壓力和摩擦損失（文獻2：庫爾茲.R，M.由羅穆斯基，2006年）。 除去其他問題，這意味著對于帶管道系統(tǒng)的壓縮機站，頭部所需流量揚程是由管道系統(tǒng)規(guī)定的（圖2）。特別地，這一特點對于壓縮機需要的能力允許頭部減量，按照規(guī)定的方式反之亦然。管道因此將不需要改變頭部的流量恒定（或壓力比）。 圖2：建立在4式上的機頭流量關系。 在短暫的情況下（如包裝其間），最初的操作條件遵循恒功率分布，如頭部流量關系如下： （Eq.5） 并將漸進地達到穩(wěn)定的關系（文獻3：奧海寧S.，R.庫爾茲，2002年） 在上述要求的