資源目錄里展示的全都有，所見(jiàn)即所得。下載后全都有,請(qǐng)放心下載。原稿可自行編輯修改=【QQ：401339828 或11970985 有疑問(wèn)可加】 編號(hào) 畢業(yè)設(shè)計(jì)（論文） 題目：80系列微型風(fēng)冷活塞式壓縮機(jī)的設(shè)計(jì) ——W80II型 信機(jī) 系 機(jī)械工程及自動(dòng)化 專業(yè) 學(xué) 號(hào)： 學(xué)生姓名： 指導(dǎo)教師： III 誠(chéng) 信 承 諾 書(shū) 本人鄭重聲明：所呈交的畢業(yè)設(shè)計(jì)（論文） 80系列微型風(fēng)冷活塞式壓縮機(jī)的設(shè)計(jì) 是本人在導(dǎo)師的指導(dǎo)下獨(dú)立進(jìn)行研究所取得的成果，其內(nèi)容除了在畢業(yè)設(shè)計(jì)（論文）中特別加以標(biāo)注引用，表示致謝的內(nèi)容外，本畢業(yè)設(shè)計(jì)（論文）不包含任何其他個(gè)人、集體已發(fā)表或撰寫(xiě)的成果作品。 班 級(jí)： 學(xué) 號(hào)： 作者姓名： III 　信 機(jī)　系 　機(jī)械工程及自動(dòng)化 　專業(yè) 畢 業(yè) 設(shè) 計(jì)論 文 任 務(wù) 書(shū) 一、題目及專題： 1、題目　 80系列微型風(fēng)冷活塞式壓縮機(jī)的設(shè)計(jì)　 　　 2、專題　 　 二、課題來(lái)源及選題依據(jù) 80系列微型風(fēng)冷活塞式壓縮機(jī)是風(fēng)冷單作用壓縮機(jī)，使用飛濺的有霧進(jìn)行潤(rùn)滑，在食品、醫(yī)療、儀表等行業(yè)廣泛應(yīng)用。壓縮機(jī)由三相異步電動(dòng)機(jī)作為原動(dòng)機(jī)，經(jīng)“V”型皮帶傳動(dòng)，使曲軸作旋轉(zhuǎn)運(yùn)動(dòng)，再通過(guò)連桿帶動(dòng)活塞在氣缸內(nèi)作往復(fù)運(yùn)動(dòng)?？諝庥蛇M(jìn)氣閥吸入一級(jí)氣缸，壓縮后經(jīng)排氣閥進(jìn)中間冷卻器后再經(jīng)二級(jí)氣缸壓縮后進(jìn)入儲(chǔ)氣罐。采用自動(dòng)停機(jī)方式控制排氣壓力，壓縮機(jī)的冷卻主要由兼作風(fēng) 扇的飛輪對(duì)氣缸及中間冷卻器進(jìn)行強(qiáng)制對(duì)流換熱來(lái)保證。 三、本設(shè)計(jì)（論文或其他）應(yīng)達(dá)到的要求： 1、 根據(jù)設(shè)計(jì)參數(shù)進(jìn)行壓縮機(jī)的熱、動(dòng)力計(jì)算（主要包括缸徑的 確定，電動(dòng)機(jī)功率計(jì)算及選型，壓縮機(jī)中的作用力的分析， 飛輪距的確定，慣性力和慣性力矩的平衡）； 2、繪制主機(jī)總圖及主要零件圖； 3、對(duì)壓縮機(jī)主要零件進(jìn)行強(qiáng)度校核； 4、根據(jù)計(jì)算結(jié)果，確定壓縮機(jī)結(jié)構(gòu)尺寸，完成總裝圖； 5、查閱相關(guān)資料，完成畢業(yè)設(shè)計(jì)說(shuō)明書(shū)一份，不少于30頁(yè)。 四、接受任務(wù)學(xué)生： 班 　　姓名 五、開(kāi)始及完成日期： 自2012年11月12日 至2013年5月25日 六、設(shè)計(jì)（論文）指導(dǎo)（或顧問(wèn)）： 指導(dǎo)教師　　　　　　　簽名 簽名 　　　　　　　簽名 教研室主任 　　　　　　〔學(xué)科組組長(zhǎng)研究所所長(zhǎng)〕　　　　　　　簽名 　　 　　系主任 　　　　　　簽名 2012年11月12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(fù)式壓縮機(jī)的工作效率特性 往復(fù)式壓縮機(jī)和離心式壓縮機(jī)具有不同的工作特性，而且關(guān)于效率的定義也不同。本文提供了一個(gè)公平的比較準(zhǔn)則，得到了對(duì)于兩種類型機(jī)器普遍適用的效率定義。這個(gè)比較基于用戶最感興趣的要求提出的。此外，對(duì)于管道的工作環(huán)境影響和在不同負(fù)載水平的影響給出了評(píng)估。 乍一看，計(jì)算任何類型的壓縮效率看似是很簡(jiǎn)單的：比較理想壓縮過(guò)程和實(shí)際壓縮過(guò)程的工作效率。難點(diǎn)在于正確定義適當(dāng)?shù)南到y(tǒng)邊界，包括與之相關(guān)的壓縮過(guò)程的損失。除非這些邊界是恰好定義的，否則離心式和往復(fù)式壓縮機(jī)的比較就變得有缺陷了。 我們也需要承認(rèn)，效率的定義，甚至是在評(píng)估公平的情況下，仍不能完全回應(yīng)操作員的主要關(guān)心問(wèn)題：壓縮過(guò)程所需的驅(qū)動(dòng)力量是什么？要做到這一點(diǎn)，就需要討論在壓縮過(guò)程中的機(jī)械損失。 隨著時(shí)間的推移效率趨勢(shì)也應(yīng)被考慮，如非設(shè)計(jì)條件，它們是由專業(yè)的流水線規(guī)定，或者是受壓縮機(jī)的工作時(shí)間和自身退化的影響。 　管道使用的壓縮設(shè)備涉及到往復(fù)式和離心式壓縮機(jī)。離心式壓縮機(jī)用燃?xì)廨啓C(jī)或者是電動(dòng)馬達(dá)來(lái)驅(qū)動(dòng)。所用的燃?xì)廨啓C(jī)，總的來(lái)說(shuō)，是兩軸發(fā)動(dòng)機(jī)，電動(dòng)馬達(dá)使用的是變速馬達(dá)或者變速齒輪箱。往復(fù)壓縮機(jī)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?xì)獍l(fā)動(dòng)機(jī)和壓縮機(jī)在一個(gè)曲柄套管內(nèi)。后者單位的運(yùn)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(dòng)馬達(dá)或者四沖程燃?xì)獍l(fā)動(dòng)機(jī)來(lái)驅(qū)動(dòng)。 效率 　　要確定任何壓縮過(guò)程的等熵效率，就要基于測(cè)量的壓縮機(jī)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(cè)量的穩(wěn)態(tài)質(zhì)量流m，吸收軸功率為： (Eq.2) 考慮機(jī)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(jī)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(jì)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(jī)械效率，在描述軸承和密封件的摩擦損失以及風(fēng)阻損失時(shí)可以達(dá)到98%和99%。 對(duì)于往復(fù)式壓縮機(jī)，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(dòng)。往復(fù)壓縮機(jī)就其性質(zhì)而言，從臨近單位需要多方面的系統(tǒng)來(lái)控制脈動(dòng)和提供隔離（包括往復(fù)式和離心式），以及可以自然存在的來(lái)自管線的管流量和面積管道。對(duì)于任何一個(gè)低速或高速單位的歧管系統(tǒng)設(shè)計(jì)，使用了卷相結(jié)合，管道長(zhǎng)度和壓力降元素來(lái)創(chuàng)造脈動(dòng)（聲波）濾波器。這些歧管系統(tǒng)（過(guò)濾器）引起壓力下降，因此必須在效率計(jì)算時(shí)考慮到。潛在的，從吸氣壓力扣除的額外壓力不得不包含進(jìn)殘余脈動(dòng)的影響。就像離心壓縮機(jī)一樣，傳熱就經(jīng)常被忽視。 對(duì)于積分的機(jī)器，機(jī)械效率一般取為95%。對(duì)于可分機(jī)機(jī)械效率一般使用97%。這些數(shù)字似乎有些樂(lè)觀，一系列數(shù)字顯示，往復(fù)式發(fā)動(dòng)機(jī)機(jī)械損失在8-15%之間，往復(fù)壓縮機(jī)的在6-12%（參考1往復(fù)壓縮機(jī)招致號(hào)碼：庫(kù)爾茲，R.，K.，光布倫，2007）。 工作環(huán)境 在這樣的情況下，當(dāng)壓縮機(jī)在一個(gè)系統(tǒng)中運(yùn)行時(shí)，管道長(zhǎng)度Lu上游和Ld下游，以及管道pu上游的初始?jí)毫凸艿纏e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(gè)壓縮機(jī)運(yùn)行的簡(jiǎn)單模型（圖1）。 圖1：管道段的概念模型（文獻(xiàn)2：庫(kù)爾茲.R，M.由羅穆斯基，2006年）。 對(duì)于給定的，標(biāo)準(zhǔn)管線定量流動(dòng)能力將在吸入階段強(qiáng)加壓力，在壓縮機(jī)放電區(qū)強(qiáng)加壓力。對(duì)于給定的管線，壓縮機(jī)站頭部（）流（Q）關(guān)系可以近似表述為 （Eq.4） 其中和是常數(shù)（對(duì)于一個(gè)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(xiàn)2：庫(kù)爾茲.R，M.由羅穆斯基，2006年）。 除去其他問(wèn)題，這意味著對(duì)于帶管道系統(tǒng)的壓縮機(jī)站，頭部所需流量揚(yáng)程是由管道系統(tǒng)規(guī)定的（圖2）。特別地，這一特點(diǎn)對(duì)于壓縮機(jī)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(jī)頭流量關(guān)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(guān)系如下： （Eq.5） 并將漸進(jìn)地達(dá)到穩(wěn)定的關(guān)系（文獻(xiàn)3：奧海寧S.，R.庫(kù)爾茲，2002年） 在上述要求的基礎(chǔ)上，必須控制壓縮機(jī)輸出與系統(tǒng)要求匹配。該系統(tǒng)需求的特點(diǎn)是系統(tǒng)流程和系統(tǒng)頭部或壓力比的強(qiáng)烈關(guān)系。管線壓縮機(jī)提供了在操作條件經(jīng)驗(yàn)下的大量變化，一個(gè)重要問(wèn)題就是如何使壓縮機(jī)適應(yīng)這樣變化的條件，具體的說(shuō)就是如何影響效率。 離心壓縮機(jī)具有相當(dāng)大的平頭部和流程特點(diǎn)。這意味著壓力比的改變對(duì)機(jī)器的實(shí)際流程有重大的影響（文獻(xiàn)4：庫(kù)爾茲R.，20004年）。對(duì)于一個(gè)恒速運(yùn)行的壓縮機(jī)，頭部或壓力比隨著流量的增加而減少?？刂茐嚎s機(jī)內(nèi)的流程可以實(shí)現(xiàn)壓縮機(jī)不同的運(yùn)行速度。這是控制離心壓縮機(jī)最便捷的方法。兩軸燃?xì)廨啓C(jī)和變速電機(jī)允許大范圍的速度變化（通常是最大速度或更多的40%或50%到100%）。應(yīng)當(dāng)指出，被控制的值通常不是速度，但速度是間接平衡由渦輪產(chǎn)生的動(dòng)力（受進(jìn)入燃?xì)廨啓C(jī)燃油流量控制）和壓縮機(jī)的吸收功率。 事實(shí)上，在過(guò)去15年安裝的任何離心壓縮機(jī)在管線服務(wù)方面是由調(diào)速器來(lái)驅(qū)使的，通常是兩軸燃?xì)廨啓C(jī)。年長(zhǎng)的設(shè)施和服務(wù)設(shè)施在其他管線服務(wù)有時(shí)使用單軸燃?xì)廨啓C(jī)（允許速度90%到100%的變化）和恒速電動(dòng)機(jī)。在這些裝置中，吸節(jié)流或可變進(jìn)氣導(dǎo)葉用來(lái)提供控制方法。 圖3：典型的管線運(yùn)行點(diǎn)繪制成的典型離心壓縮機(jī)性能圖。 離心壓縮機(jī)的運(yùn)行封套受最大允許速度限制，最小流量（涌）和最大流量（窒息或石墻）（圖3）。另一個(gè)限制因素可能是可用的驅(qū)動(dòng)電源。 只有最小流量需要特別注意，因?yàn)樗欢x為壓縮機(jī)的一種氣動(dòng)穩(wěn)定性的極限。跨越這個(gè)限制以降低流動(dòng)將導(dǎo)致壓縮機(jī)流動(dòng)逆轉(zhuǎn)，這可能會(huì)損壞壓縮機(jī)。調(diào)制解調(diào)器控制系統(tǒng)通過(guò)打開(kāi)一個(gè)循環(huán)閥來(lái)控制這種情況。出于這個(gè)原因，幾乎所有的現(xiàn)代壓縮機(jī)裝置都使用帶有控制閥的循環(huán)線，當(dāng)壓縮機(jī)內(nèi)的流量趨于穩(wěn)定極限時(shí)這種控制閥允許流量的增加?？刂葡到y(tǒng)不斷地監(jiān)測(cè)壓縮機(jī)關(guān)系喘振線的運(yùn)行點(diǎn)，并且有必要的話自動(dòng)地開(kāi)關(guān)循環(huán)閥。對(duì)于大多數(shù)應(yīng)用來(lái)說(shuō)，帶有開(kāi)放或部分開(kāi)放循環(huán)閥的運(yùn)行模式只被用于開(kāi)啟和關(guān)閉階段，或者是在混亂運(yùn)行條件時(shí)的短暫時(shí)期。 假設(shè)由公式4得到管線特點(diǎn)，壓縮機(jī)的葉輪將在達(dá)到或接近其最大效率時(shí)被選出來(lái)運(yùn)行，這個(gè)最大效率是由管線強(qiáng)加在整個(gè)系列的頭部和流量條件下的。這可能是有一個(gè)速度（N）控制的壓縮機(jī)，因?yàn)橐粋€(gè)壓縮機(jī)的最有效點(diǎn)是由一種關(guān)系而連接的，這種關(guān)系需要大約（風(fēng)扇法方程）： (Eq.6) 為滿足上述關(guān)系的操作點(diǎn)，吸入氣壓是（基于效率幾乎保持不變這個(gè)的事實(shí)）： 　 (Eq.7)　　正因?yàn)槿绱耍@種力-速度關(guān)系允許動(dòng)力渦輪運(yùn)行達(dá)到或非常接近其整個(gè)范 圍的理想速度。管線中典型的運(yùn)行方案允許壓縮機(jī)和動(dòng)力渦輪在大多數(shù)時(shí)間里在最有效點(diǎn)運(yùn)行。然而，燃?xì)廨啓C(jī)的燃?xì)馍a(chǎn)商將在部分負(fù)荷運(yùn)行時(shí)丟失一些熱效率。 圖3顯示了一個(gè)典型的實(shí)際例子：不同流動(dòng)要求的管線運(yùn)行點(diǎn)繪制成用于壓縮機(jī)站中的速度控制離心壓縮機(jī)性能圖。 往復(fù)壓縮機(jī)將自動(dòng)服從系統(tǒng)壓力比的需求，只要沒(méi)有超出機(jī)械的限制條件（桿負(fù)載功率）。系統(tǒng)吸排氣壓力的改變將僅能引起閥門或早或晚的開(kāi)啟。頭部可以自動(dòng)下降因?yàn)殚y門可以降低排氣端的管線壓力和/或吸入端更高的管線壓力。因此，如果沒(méi)有額外的措施，流量將大致恒定——除了容積效率將增加的變化，所以降低壓力比而增加流量。 控制的挑戰(zhàn)存在于系統(tǒng)要求的流量調(diào)整。如果沒(méi)有額外的調(diào)整，隨著壓力比的變化，壓縮機(jī)流量的改變微乎其微。從歷史上看，通過(guò)改變激活機(jī)器的數(shù)量使管線安裝許多小的壓縮機(jī)和調(diào)整流量。這個(gè)容量和負(fù)荷可通過(guò)速度調(diào)諧，或者通過(guò)一個(gè)單一單元的缸間隙中的許多小調(diào)整（加載步驟）來(lái)調(diào)諧。隨著壓縮機(jī)的發(fā)展，控制容量的負(fù)擔(dān)轉(zhuǎn)移到獨(dú)立壓縮機(jī)上。 負(fù)荷控制是壓縮機(jī)運(yùn)行的一個(gè)關(guān)鍵組成部分。從管線操作角度來(lái)看，在機(jī)組中流量變化要符合管線投出承諾，以及實(shí)施公司最佳操作（例如，線包裝，負(fù)載預(yù)期）。從一個(gè)單元的角度來(lái)看，負(fù)荷控制包含降低單元流量（通過(guò)卸載或速度）使操作盡可能的貼近設(shè)計(jì)扭矩限制，并在壓縮機(jī)或驅(qū)動(dòng)程序沒(méi)有超載的情況下進(jìn)行。對(duì)于任何給定的機(jī)組入口和出口壓力，在任何負(fù)荷圖曲線上的關(guān)鍵限制都是桿負(fù)荷限制和馬力/扭矩限制。瓦斯控制通常會(huì)建立在一個(gè)機(jī)組的單元上，而這個(gè)機(jī)組運(yùn)行必須達(dá)到管線流量目標(biāo)。地方單元控制將建立負(fù)載步驟或速度要求來(lái)限制桿負(fù)荷或達(dá)到扭矩控制。 改變流量的常用方法是改變速度，改變間隙，或取消激活缸頭（保持進(jìn)口閥開(kāi)啟）。另一種方法是卸載無(wú)限步驟，從而延緩吸氣閥封閉以減少容積效率。此外，流程的一部分可以回收或吸氣壓力可以節(jié)流從而降低質(zhì)量流量，同時(shí)保持進(jìn)入壓縮機(jī)的容積流量基本不間斷。 壓縮機(jī)控制策略應(yīng)該能夠?qū)崿F(xiàn)自動(dòng)化，并在壓縮機(jī)運(yùn)行期間能夠簡(jiǎn)便地調(diào)整。特別地，壓縮機(jī)設(shè)計(jì)修改的戰(zhàn)略需求（如：離心壓縮機(jī)重新旋轉(zhuǎn)，改變缸徑，或給往復(fù)壓縮機(jī)添加固定間隙）在這里不被考慮。需要指出的是，對(duì)于往復(fù)式壓縮機(jī)一個(gè)關(guān)鍵的控制要求是不超載驅(qū)動(dòng)或超過(guò)機(jī)械限制。 運(yùn)行 典型的穩(wěn)態(tài)管道運(yùn)行將產(chǎn)生圖4所示的一個(gè)有效行為。該圖是評(píng)估沿管道穩(wěn)定運(yùn)行特征狀態(tài)壓縮機(jī)效率的結(jié)果。大中型壓縮機(jī)都將達(dá)到100%流量的最佳效率，并允許超出設(shè)計(jì)流量的10%。不同的機(jī)械效率并沒(méi)有考慮這種對(duì)比。 往復(fù)壓縮機(jī)效率在文獻(xiàn)5中被推導(dǎo)出，從增加的閥門效率測(cè)量與壓縮效率和造成的損失脈動(dòng)衰減器。低速壓縮機(jī)的效率是可以實(shí)現(xiàn)的。高速往復(fù)壓縮機(jī)在效率上可能比較低。 圖4：以穩(wěn)態(tài)管線特性運(yùn)行為基礎(chǔ)的在不同流量率的壓縮機(jī)效率。 圖4顯示在較低壓力比下增加的閥門損失的影響和往復(fù)機(jī)器的較低流量，而離心壓縮機(jī)的效率幾乎保持常量。 結(jié)論 不同型號(hào)壓縮機(jī)間的效率定義和對(duì)比需要密切關(guān)注邊界條件的定義，對(duì)于這樣的邊界條件，效率和受用的運(yùn)行發(fā)展趨勢(shì)同時(shí)被定義。當(dāng)效率值用來(lái)計(jì)算功耗時(shí)機(jī)械效率具有重要作用。如果不考慮這些定義，不同系統(tǒng)的優(yōu)缺點(diǎn)討論將變得不準(zhǔn)