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存檔編碼：無無錫錫太太湖湖學(xué)學(xué)院院 2013 屆屆畢畢業(yè)業(yè)作作業(yè)業(yè)周周次次進進度度計計劃劃、檢檢查查落落實實表表 系別：信機系 班級：機械95班 學(xué)生姓名：顧佳慶 課題（設(shè)計）名稱：微型風(fēng)冷活塞式壓縮機(W-80)的設(shè)計 開始日期：2012.11 周次起止日期工作計劃、進度每周主要完成內(nèi)容存在問題、改進方法指導(dǎo)教師意見并簽字備 注1-32012年11月12日-2012年12月2日教師下達畢業(yè)設(shè)計任務(wù)，學(xué)生初步閱讀資料，完成畢業(yè)設(shè)計開題報告。按照任務(wù)書要求查閱論文相關(guān)參考資料，填寫畢業(yè)設(shè)計開題報告書問題：剛開始拿到資料，感覺無從下手，不知道從哪邊開始；改進方法：通過俞老師的耐心指導(dǎo)和手邊資料的查看，初步了解了壓縮機的構(gòu)成。4-10 2012年12月3日-2013年1月20日指導(dǎo)專業(yè)實訓(xùn)通過查找相關(guān)壓縮機的資料，了解畢業(yè)設(shè)計的內(nèi)容，對畢業(yè)設(shè)計怎么著手有一個概括性的認識。問題：自學(xué)參考資料的過程中，很多概念知識不是很理解；改進方法：通過多次跟指導(dǎo)老師的交流，解決了相關(guān)問題。11-122013年1月21日-2013年3月1日指導(dǎo)畢業(yè)實習(xí)了解壓縮機的總體結(jié)構(gòu)以及工作過程，特點及應(yīng)用，分類及形式等等。問題：剛開始接觸，對熱力過程，實際循環(huán)過程很陌生；改進方法：多跟指導(dǎo)老師溝通，多查閱資料手冊，得到俞老師的指點后，自己能夠很快的掌握。132013年3月4日 -3月8日開始著手壓縮機的熱力計算自己先查閱翻看相關(guān)的資料，了解相關(guān)概念，例如容積系數(shù)等等，進行熱力計算。問題：在熱力計算過程中，在確定缸徑和計算最大活塞力那塊有些問題。改進方法：在俞老師的指導(dǎo)下，對這些問題耐心講解，從而可以按照進度順利往下計算。142013年3月11日-3月15日熱力計算改進后，進行動力計算翻看壓縮機資料，了解壓縮機所受的作用力，進行動力計算。問題：在動力計算中往復(fù)慣性力和摩擦力飛輪矩以及難以確定；改進方法：俞老師給我很多的指導(dǎo)，從而讓我很快掌握其中的原理。152013年3月18日-3月22日動力計算改進后，開始著手零件圖紙的繪制自己讀懂壓縮機裝配圖，試著從裝配圖上拆分主要零件，進行繪制。問題：繪制零件圖的過程中，很多尺寸不確定，有些結(jié)構(gòu)也不是很確定；改進方法：俞老師給了我們一定的手冊和資料，同時在專業(yè)知識上給了我們相應(yīng)的輔導(dǎo)。162013年3月25日-3月29日繪制零件圖根據(jù)裝配圖，爆炸圖等繪制氣缸氣缸蓋零件圖。問題：在制圖過程中，有一些尺寸要從相關(guān)范圍內(nèi)選取，自己不知道那些經(jīng)驗值。改進方法：通過俞老師的悉心指導(dǎo)，我得以順利進行。172013年4月1日 -4月5日根據(jù)零件圖，繪制組件圖本周主要的任務(wù)就是繪制組件圖問題：在這個過程中，我發(fā)現(xiàn)之前的零件圖又有些不對的地方，不能組裝起來；改進方法：后來在俞老師的指點和相關(guān)資料的查詢下，我發(fā)現(xiàn)了錯誤，最終改正并且完成組件圖。182013年4月8日 -4月12日主要零部件的分析設(shè)計由于這個工作量相對比較大，所以本周主要的任務(wù)是進一步完善這項工作，并對前面做的進行復(fù)核問題：發(fā)現(xiàn)很多數(shù)據(jù)其實按照書上說的，算出來很不實際，達不到企業(yè)節(jié)省成本的目的；改進方法：俞老師給以指導(dǎo)，讓我明白設(shè)計和實際操作還是有一定差距的。192013年4月15日-4月19日曲軸的分析設(shè)計本周主要是分析曲軸上的平衡鐵與慣性力之間的關(guān)系問題：一開始對這些概念很模糊，不知道怎么著手，主要是不熟悉壓縮機方面的知識。改進方法：在俞老師的指點下，我認識到平衡鐵的作用。202013年4月22日-4月26日曲軸中的平衡計算本周主要是曲軸的平衡的計算，如何抵消往復(fù)慣性力和旋轉(zhuǎn)慣性力問題：不知道如何平衡往復(fù)慣性力和旋轉(zhuǎn)慣性力；改進方法：俞老師耐心的講解指導(dǎo)，最終順利完成。212012年4月29日-5月3日氣閥的分析了解氣閥的相關(guān)概念問題：一開始不理解氣閥的概念，也不知道選什么形式的氣閥；改進方法：俞老師給以一定的手冊，讓我能夠很順利的進行畢業(yè)設(shè)計。222013年5月6日 -5月10日活塞組件的設(shè)計筒形活塞的分析計算問題：筒形活塞材料和技術(shù)要求自己不敢確定；改進方法：俞老師給了一定的意見和資料，讓我很順利的完成了計算。232013年5月13日-5月17日完善說明書和相關(guān)資料本周完善好說明書很相關(guān)資料，并讓老師復(fù)核，并進行改進問題：發(fā)現(xiàn)某些排版方面的設(shè)置不太熟悉；改進方法：通過網(wǎng)絡(luò)和老師同學(xué)的幫助，很順利的解決。242013年5月20日-5月25日不足的地方改進，打印，裝訂，上交主要就是對不滿足要求的地方改進，然后進行打印，最終裝訂好了上交周次起止日期工作計劃、進度每周主要完成內(nèi)容存在問題、改進方法指導(dǎo)教師意見并簽字備 注
編號
無錫太湖學(xué)院
畢業(yè)設(shè)計（論文）
相關(guān)資料
題目：微型風(fēng)冷活塞式壓縮機(W-80)的設(shè)計
信機 系 機械工程及自動化專業(yè)
學(xué) 號： 0923208
學(xué)生姓名： 顧 佳 慶
指導(dǎo)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(shè)計（論文）開題報告
二、畢業(yè)設(shè)計（論文）外文資料翻譯及原文
三、學(xué)生“畢業(yè)論文（論文）計劃、進度、檢查及落實表”
四、實習(xí)鑒定表
無錫太湖學(xué)院
畢業(yè)設(shè)計（論文）
開題報告
題目：微型風(fēng)冷活塞式壓縮機(W-80)的設(shè)計
信機 系 機械工程及自動化 專業(yè)
學(xué) 號： 0923208
學(xué)生姓名： 顧佳慶
指導(dǎo)教師： 俞萍（職稱：高級工程師）
（職稱： ）
2012年11月20日
課題來源
本課題來源于企業(yè)；
結(jié)合所學(xué)知識，老師擬定題目；
綜合大學(xué)里所學(xué)知識，將理論與實踐相互結(jié)合。
科學(xué)依據(jù)（包括課題的科學(xué)意義；國內(nèi)外研究概況、水平和發(fā)展趨勢；應(yīng)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(shè)備所替代，是許多工程項目中的關(guān)鍵設(shè)備。
2、 活塞式壓縮機在圓筒形氣缸中具有一個可往復(fù)運動的活塞，氣缸上有控制進、排氣的閥門，當活塞作往復(fù)運動時，氣缸的容積便周期性的變化，借以實現(xiàn)氣體的吸進、壓縮、和排出。
3、 隨著經(jīng)濟的高速發(fā)展和科學(xué)技術(shù)的不斷進步，各種壓縮機在國民經(jīng)濟各大領(lǐng)域大顯身手，壓縮機是原基礎(chǔ)材料之一的冶金工業(yè)中極為重要的設(shè)備，又是石油化工流程中的心臟設(shè)備。車輛的制動、船用內(nèi)燃機啟動，航空發(fā)動機的運行都需要各種壓縮機，可以說壓縮機在陸?？战煌ㄟ\輸工具中都必不可少，與人民的日常生活更是休戚相關(guān)。
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、 微型風(fēng)冷活塞式壓縮機的工作原理以及工作形成；
2、 微型風(fēng)冷活塞式壓縮機參數(shù)與結(jié)構(gòu)的設(shè)計；
3、 微型風(fēng)冷活塞式壓縮機設(shè)計圖紙的繪制。
擬采取的研究方法、技術(shù)路線、實驗方案及可行性分析
研究方法：通過閱讀有關(guān)資料，文獻，收集篩選，整理課題研究所需的
有關(guān)數(shù)據(jù)，理論依據(jù)，綜合運用所學(xué)理論知識研究論文課題。
技術(shù)路線：分析微型風(fēng)冷活塞式壓縮機的各個參數(shù)的取值情況，包括結(jié)
構(gòu)參數(shù)、工藝參數(shù)、熱力學(xué)參數(shù)和動力學(xué)參數(shù)。確定各參數(shù)
的具體數(shù)值或取值區(qū)間。
可行性分析：通過對論文課題的學(xué)習(xí)研究，達到鞏固，擴大，深化已學(xué)
理論知識，提高思考分析解決實際問題等綜合素質(zhì)的目的。
研究計劃及預(yù)期成果
1、 首先對微型風(fēng)冷活塞式壓縮機整體結(jié)構(gòu)進行分析，對傳動結(jié)構(gòu)進行篩選，初步選擇達到設(shè)計要求的結(jié)構(gòu)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(gòu)的分析計算可提高其自身的精度；
3、 對微型風(fēng)冷活塞式壓縮機的主要零件進行強度校核，提高機構(gòu)穩(wěn)定性，穩(wěn)定性。
特色或創(chuàng)新之處
通過對微型風(fēng)冷活塞式壓縮機的設(shè)計及計算，形成一整套現(xiàn)代的設(shè)計方法，對理論和實踐的結(jié)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(jié)構(gòu)設(shè)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(shè)計手冊等參考資料及文獻；對活塞式壓縮機進行直觀的了解與認識，對所學(xué)的機械基礎(chǔ)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：對于微型風(fēng)冷活塞式壓縮機的工作原理不是非常清楚
和熟悉，缺乏設(shè)計經(jīng)驗。
指導(dǎo)教師意見
指導(dǎo)教師簽名：
年 月 日
教研室（學(xué)科組、研究所）意見
教研室主任簽名：
年 月 日
系意見
主管領(lǐng)導(dǎo)簽名：
年 月 日
無錫太湖學(xué)院
畢業(yè)設(shè)計（論文）外文資料翻譯
信機 系 機械工程及自動化 專業(yè)
院 （系）： 信 機 系
專 業(yè)： 機械工程及自動化
班 級： 機械95班
姓 名： 顧 佳 慶
學(xué) 號： 0923208
外文出處： 機械專業(yè)英語教程
附 件： 1.譯文；2.原文；3.評分表
2013年5月25日
英文原文
Efficiency And Operating Characteristics Of Centrifugal And Reciprocating Compressors
By Rainer Kurz, Bernhard Winkelmann, and Saeid iVIokhatab
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 from 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. Bun, 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 increase, 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 compressor (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 efficiency behavior as outlined in Figure 4. This figure is the result of evaluating the compressor efficiency 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 efficiency is derived n-on valve efficiency measurements in Ref 5 (Null, 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 at different flow rates based on operation along 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.
REFERENCES
1 Kurz . R.. K. Burn. 2007. " Efficiency Definition and Load Management for Reciprocating and Centrifugal Compressors," ASME Paper GT2OO7-27O81.
2 Kurz. R., M. Lubomirsky, 2006. "Asymttietric Solution for Compressor Station Spare Capacity."ASMt: Paper 2006-90069.
3 Ohanian. S.. R. Kurz. 2002, "Series or Parallel Arrangement in a Two-Unit Compressor Station." Trans.ASME Jeng for GT and Power. Vol.124.
4 Kurz. R.. 2004. "The Physies of Centrifugal Compressor Performance." Pipeline Simulation Interest Group. Palm Springs. CA.
5 Noral, M.. W. Couch. 2003, "Performance and Endurance Tests of Six Mainline Compressor Valves in Natural Gas Compression Service." Gas Machinery Conference. Salt Lake City. UT.
中文原文
離心式和往復(fù)式壓縮機的工作效率特性
Rainer Kurz , Bernhard Winkelmann , and Saeid Mokhatab
往復(fù)式壓縮機和離心式壓縮機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(guān)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(fù)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(yīng)操作員的主要關(guān)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(yīng)被考慮，如非設(shè)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(shè)備涉及到往復(fù)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(fù)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(fēng)阻損失時可以達到98%和99%。
對于往復(fù)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(fù)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(fù)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