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Exergy Analysis of the Revolving VaneCompressed Air Engine

Subiantoro Alison Wong Kin Keong Ooi Kim Tiow

2016

Subiantoro A Wong K K amp Ooi K T (2016) Exergy Analysis of the Revolving VaneCompressed Air Engine International Journal of Rotating Machinery 2016 5018467‑

httpshdlhandlenet1035682853

httpsdoiorg10115520165018467

copy 2016 Alison Subiantoro et al This is an open access article distributed under the CreativeCommons Attribution License which permits unrestricted use distribution andreproduction in any medium provided the original work is properly cited

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Research ArticleExergy Analysis of the Revolving Vane Compressed Air Engine

Alison Subiantoro12 Kin Keong Wong3 and Kim Tiow Ooi23

1Department of Industrial and Systems Engineering Krida Wacana Christian University Jalan Tanjung Duren Raya No 4Jakarta Barat 11470 Indonesia2Technical University of Munich-Campus for Research Excellence and Technological Enterprise (TUM CREATE)1 CREATEWay No 10-02 CREATE Tower Singapore 1386023School of Mechanical and Aerospace Engineering Nanyang Technological University 50 Nanyang Avenue Singapore 639798

Correspondence should be addressed to Alison Subiantoro alisonsubiantoroukridaacid

Received 3 December 2015 Revised 14 January 2016 Accepted 14 January 2016

Academic Editor Zuohua Huang

Copyright copy 2016 Alison Subiantoro et al This is an open access article distributed under the Creative Commons AttributionLicense which permits unrestricted use distribution and reproduction in any medium provided the original work is properlycited

Exergy analysis was applied to a revolving vane compressed air engineThe engine had a swept volume of 30 cm3 At the benchmarkconditions the suction pressure was 8 bar the discharge pressure was 1 bar and the operating speed was 3000 revsdotminminus1 It wasfound that the engine had a second-law efficiency of 296 at the benchmark conditions The contributors of exergy loss werefriction (49) throttling (38) heat transfer (12) and fluidmixing (1) A parametric study was also conductedThe parametersto be examined were suction reservoir pressure (4 to 12 bar) operating speed (2400 to 3600 revsdotminminus1) and rotational cylinderinertia (094 to 281 gsdotmm2) The study found that a higher suction reservoir pressure initially increased the second-law efficiencybut then plateaued at about 30 With a higher operating speed and a higher cylinder inertia second-law efficiency decreased Ascompared to suction pressure and operating speed cylinder inertia is the most practical and significant to be modified

1 Introduction

Since its inception in the 1800s reciprocating piston engineshave been widely used to drive motor vehicles Howevereven with all the progress and innovations in the reciprocat-ing engine technology there are still issues and challengesfacing the engine design To overcome these there havebeen attempts to come up with new engine designs Rotaryengines have been proposed as a possible design alternativeAs compared to a reciprocating engine rotary engines areusually more compact and have less number of componentswith better vibration and noise characteristics In responseWankel developed a rotary piston engine in the 1960s whichhas been adopted by Mazda to drive its sport vehiclesHowever leakage is one of the main issues with this enginedesign [1] More recently Sakita introduced a new rotarymachine design called the ldquocat andmouse typerdquo rotary engine[2] Shen and Hwang adopted the sliding vane machinedesign for an air-powered motorcycle [3]

In 2006 a new rotary machine design concept calledthe revolving vane (RV) mechanism was introduced by Teh

and Ooi [4] An improved version was later proposed toimprove the mechanical efficiency of the machine [5] Unlikethe more common rotary machines which use stationarycylinders the RV machine lets the cylinder move togetherwith the rotor This significantly reduces the frictions of themachine In addition it is also found that the RV machineexhibits better volumetric performance as compared to arolling piston machine [6] The machine was first introducedas a compressor but has subsequently been used in expanderand air engine applications too [7ndash9]

In its simplest form the RV mechanism consists of arotor a vane and a cylinder as shown in Figure 1(a) Howeverunlike other rotary machines the cylinder rotates togetherwith the rotor resulting in smaller relative velocities and lessfrictional losses at the rubbing surfaces The rotor and thecylinder are arranged eccentrically and they rotate at theirrespective axes The vane divides the space between the rotorand the cylinder inner wall into two that is the suction anddischarge chambers As the rotor rotates the volumes of thesechambers vary allowing for thermodynamics processes to

Hindawi Publishing CorporationInternational Journal of Rotating MachineryVolume 2016 Article ID 5018467 8 pageshttpdxdoiorg10115520165018467

2 International Journal of Rotating Machinery

Inlet

Discharge port

Vane slot

Vane

Suction port

Casing

Rotor

Cylinder

(a)

(A) (B)

(D) (C)

(b)

Figure 1 Revolving vane engine (a) schematics and (b) working principles [6]

occur A stationary casing is usually required to house theengine The working principles of an RV engine are shownin Figure 1(b)

Meanwhile it was reported that through the burning offossil fuels mainly gasoline and diesel the transportationsector contributed around 15 of the total global greenhouseemission [10] Therefore various cleaner alternative fuelshave been proposed in the recent years Among others oneof the most interesting is the compressed air engine [3 11 12]In this paper the RV machine is adopted for air engineapplication Various aspects of the mechanism have beencomprehensively studied before [7ndash9 13] However an exergyanalysis of themachine is still lacking It is the purpose of thispaper to describe the exergy analysis of the RV compressedair engine to quantify the irreversibilities involved in theoperationThis analysis will help to improve the future designprocess of the machine

Exergy also known as availability is the maximumamount of usefulwork that can be derived froma substance inbringing it to an equilibrium state with the environmentTheexergy of a system is the maximum useful work possible in aprocess that brings the system into thermal equilibrium withits surroundings After the system and surroundings reachequilibrium the exergy is zero Exergy also accounts for theirreversibility of a process

In addition to the first law exergy analysis also usesthe second law of thermodynamics which deals with theconcepts of quality of energy as well as entropy of a systemAs exergy is a property of the system the exergies of the fluidbefore and after the cycle can be quantifiedThe loss will thenbe the reduction in total exergy of the system The second-law efficiency is determined as well and it provides a morecomprehensive evaluation of the efficiency of the engine

Exergy analysis has been carried out for variousmachines including compressors and engines McGovernet al [14] developed an analysis technique that uses theexergy of a gas to quantify the shaft power wastage in scroll

and rotary compressors They summarized the losses ina compressor to be mainly contributed by throttling atvalves mechanical friction leakage heat transfers andreexpansion of compressed gases By employing equationsof irreversibilities for the different components they wereable to quantify that the major losses in the compressor aredue to throttling (12) leakage (40) and friction (24)In another paper McGovern and Harte [15] performedan exergy analysis for an open reciprocating compressorThey showed that results from computer simulationsof compressors can be readily utilised to provide anunderstanding for design optimization The breakdown ofshaft power wastage could be quantified using simulationsand instantaneous exergy destruction rates of the differentcomponents are presented Overall it was found that themajor contributions are due to throttling (537) internalconvection (187) friction (125) and fluid mixing (57)The exergy analysis in this paper used the method used inthe study

In other papers McGovern [16 17] outlined the methodsof exergy analysis and reported that it is common forcompressors to have second-law efficiencies of about 50Stecco and Manfrida [18] and McGovern [19] have proposedthe case for using exergy analysis to measure efficiency ofcompressors Both concluded that exergy analysis could beapplied to all types of compressors and could be applied tostudy of an entire plant or cycle

Reddy et al [20] performed an exergy analysis on internalcombustion engines using either diesel or biogas-diesel mix-ture as fuel Similar works on internal combustion engineshave also been carried out by other researchers [21ndash24]

2 Methodology

Mathematical models of various aspects of an RV enginehave been presented in detail and experimentally verifiedbefore [7ndash9 25] Therefore they will not be discussed again

International Journal of Rotating Machinery 3

here Interested readers are directed to the relevant literatureDeveloping on the foundation of those models the exergymodel of RV engine was created

In an RV engine exergy is destroyed mostly through heattransfer between the working fluid the solid components andthe surrounding dissipated heat from friction between therubbing surfaces fluid throttling and fluid mixing Duringoperation the working fluid undergoes thermodynamicsprocesses and its temperature varies This results in heattransfers between the fluid the engine components andthe surrounding This heat energy can potentially be usedto produce work However in reality some energy will bewasted resulting in exergy destruction Simultaneously someof the enginersquos moving components rub against one anotherduring operation This increases the local temperature that iseventually dissipated as heat resulting in exergy destructionThere are fluid flows into within and out from the enginetoo As a fluid enters a control volume its pressure dropsbecause downstream pressure is lower than the upstreamThis drop is called throttling and it destroys exergyMoreoverif the temperatures between upstream and downstream aredifferent there will be additional exergy destruction due tothe fluid mixing

Exergy destruction rates due to heat transfer frictionthrottling and mixing can be expressed according to (1) asproposed by McGovern and Harte [15]

119868ht = 1198761198951198790 (1

119879

minus

1

119879

119895

)

119868

119891=

119891(

119879

0

119879

119895

)

119868th = 1198790 (1199041015840

119894minus 119904

119894)

119868mx = 1198790((119904 minus 1199041015840

119894) minus

(ℎ

1015840

119894minus ℎ

119894)

119879

)

(1)

where 1199041015840119894is the specific entropy after throttling computed at

the pressure of the bulk fluid in the system and at enthalpyℎ

1015840

119894= ℎ

119894

A parameter called the second-law efficiency can beintroduced to quantify the exergy performance of the systemaccording to (2) The amount of exergy supplied is computedfrom the inflow of fluid into the suction chamber accordingto (3) as proposed before [26]

120578 = 1 minus

119868

119864

(2)

119864 = [(ℎ minus ℎ

0) minus 119879

0(119904 minus 119904

0)] (3)

In the exergy analysis of RV engine two control volumeswere in focus that is the enginersquos suction and dischargechambers In both chambers exergywas destroyed by frictionand heat transfer In addition at the suction chamber therewere inflow of fluid from the suction reservoir and outflowdue to internal leakage At the discharge chamber there wereinflow of fluid from suction chamber due to leakage and

Table 1 Benchmark operating conditions and major dimensions

Item ValueWorking fluid AirOperating speed 3000 revsdotminminus1

Suction reservoir temperature 25∘CSuction reservoir pressure 8 barDischarge reservoir temperature 25∘CDischarge reservoir pressure 1 barAtmospheric temperature 25∘CRadius of rotor 29mmInner radius of cylinder 35mmLength of engine chamber 25mmRotating inertia of rotor 021 gsdotm2

Rotating inertia of cylinder 281 gsdotm2

Friction coefficient of vane 015Assembly radial clearance 20120583mVane endface total gap width 180 120583mCylinder endface total gap width 40 120583m

outflow to the discharge reservoirThese flows contributed toexergy destruction through throttling and mixing

A simulation code to model the exergy behaviour ofan RV engine was developed in MATLAB programminglanguage This code was coupled to the existing RV enginemodelsmentioned at the beginning of this sectionWheneverrequired a general convergence criterion of 1 was usedthroughout the simulationThe benchmark operating condi-tions and major dimensions of the RV engine are shown inTable 1 The design configuration of the RV engine studied isshown in Figure 2 These follow those used in the validationof the RV engine models reported previously [25]

Following the benchmark analysis a parametric studywas conducted to analyse how different parameters affectthe exergy behaviours The parameters to be examined weresuction reservoir pressure operational speed of engine androtational cylinder inertia These parameters were selectedbecause they were found to affect the performance of theengine significantly in previous studies [7] Moreover theseparameters can be changed independently without affectingthe other parameters during the study

3 Results and Discussion

The benchmark results gathered from the simulation arepresented and discussed in this section Following the discus-sion a parametric study was presented

31 Benchmark Conditions Figure 3(a) shows the instanta-neous exergy destruction rate due to heat transfer The heattransfer was between the working fluid and the surroundingwalls The wall temperature was assumed constant andequal to the surrounding temperature The figure shows thatexergy destruction rate due to heat transfer in the suctionchamber was significantly higher than that of the dischargechamber throughout the cycle This is because of the greater


Discharge port

Suction portSuction hole Suction valve grooves

Rotor and vaneUpper bearing

Housing body

Lip seal

Lower bearing

Housing cover

Housing

Balancing holeVane slot

Cylinder and cover

Figure 2 The revolving vane engine components and sectional drawing [25]

0

50

100

150

200

0 60 120 180 240 300 360

Heat transfer

Suction chamberDischarge chamber

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)

(a)

0

20

40

60

80

100

0 60 120 180 240 300 360


Rotor angle (∘)

Tem

pera

ture

(∘C)

minus20

(b)

Figure 3 (a) Exergy destruction rate due to heat transfer and (b) temperature variations of the fluid

temperature fluctuation of the fluid in the suction chamberas compared to that in the discharge chamber as shownby Figure 3(b) In the suction chamber fluid temperatureincreased rapidly at the start of the cycle due to the rapidincrease in pressure at the beginning of the suction processTemperature then fell due to heat transfer to the surroundingwalls for the rest of the suction process It subsequentlydecreased rapidly when the expansion process started How-ever when the temperature was too low it bounced backbecause of heat transfer from the walls Temperature of fluidin the discharge control volumewas fairly constantTherewasa slight variation at the beginning of the cycle due to the rapiddrop of pressure in the chamber when it was first exposed tothe discharge reservoir

Figure 4 shows the instantaneous exergy destruction ratedue to friction Exergy destruction due to vane friction wasdominant followed by the cylinder bearing while the restwere almost negligible This was due to the high cylinderinertia of the engine prototype used in this study It is thecharacteristic of the RV engine that its vane contact force andthe corresponding frictional loss are linearly proportional

00

60 120 180 240 300 360

Friction

EndfaceRotor bearing

Cylinder bearingVane side

Rotor angle (∘)

400

300

200

100

Exer

gy d

estr

uctio

nra

te (W

)

Figure 4 Exergy destruction rate due to friction

to the cylinder inertia The results suggest that it is crucialto reduce the cylinder inertia to improve the exergy perfor-mance of the engine

Figure 5(a) shows the instantaneous exergy destructionrate due to throttling Average contributions of inflowoutflow and leakage were comparable The highest exergy


0

400

800

1200

1600

0 60 120 180 240 300 360Exer

gy d

estr

uctio

n ra

te (W

)

Throttling

InflowInternal leakage

Outflow

Rotor angle (∘)

(a)

01234567

98

0 60 120 180 240 300 360

Pres

sure

(bar

)

Rotor angle (∘)

SuctionDischarge

(b)

Figure 5 (a) Exergy destruction rate due to throttling and (b) pressure variations of the fluid

destruction of the inflow into the suction chamber occurredat a rotor angle of around 30∘ This was when inflow rateinto the suction chamber from the suction reservoir was thehighest It did not occur earlier because of the progressiveuncovering of the suction port For the outflow from thedischarge chamber exergy destruction peaked at a rotorangle of around 45∘ This was when the outflow rate wasthe highest Like the inflow it did not occur at an earlierrotor angle due to the progressive opening of the dischargeport However the uncovering process of the discharge portwas slower as compared to the suction port High exergydestruction rate was observed at rotor angles from 30∘ to180∘ This corresponds to the high leakage flow rate from thesuction to discharge chambers because of the large pressuredifference between the chambers as shown in Figure 5(b)From the findings it can be inferred that to reduce exergydestruction due to throttling the uncovering processes of theports should be made faster This can be done by placing theports closer to the vane and by increasing the diameter of theports For the leakage tightening the leakage gaps to reduceits flow rate will help to reduce the exergy destruction

Figure 6 shows the exergy destruction rate due to fluidmixing The exergy destruction due to inflow dominatedthe rest This was due to the significantly higher flow rateand fluid properties differences of the inflow as compared tothe outflow and leakage This is particularly true when thepressure in the suction chamber was different from that in thesuction reservoir Once the two pressures equated the exergydestruction rate dropped rapidly

Overall the average exergy supply rate into the enginewas 6616W while the average exergy destruction rate was4656W This equates to second-law efficiency of 296 Thebreakdown of the overall exergy destruction is shown inTable 2 It can be seen that the main contributor of exergyloss was friction accounting for almost half of the total lossTherefore it can be concluded that the first step to improvethe RV enginersquos performance is to reduce its frictional loss

32 Parametric Study The effects of different suction reser-voir pressure conditions to exergy are shown in Table 3 Inorder to avoid over- and underexpansions the suction valve

Table 2 Breakdown of average exergy destruction rate of the RVengine

Item Value (W) ContributionHeat transfer 553 12Friction 2262 49Throttling 1790 38Fluid mixing 51 1

Table 3 Data of variation of suction reservoir pressure

Suction pressure (bar) 4 6 8 10 12Suction valve opening angle 116∘ 94∘ 84∘ 67∘ 58∘

Second-law efficiency 30 300 296 316 302Exergy supplied (W) 359 551 662 737 803Exergy destroyed (W) 348 386 466 504 560Contribution of throttling (W) 100 119 179 206 249Contribution of friction (W) 224 225 226 226 227

020406080

100120140

0 60 120 180 240 300 360

Fluid mixing


Outflow

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)

Figure 6 Exergy destruction rate due to fluid mixing

opening angle was varied accordingly The results show thata higher suction reservoir pressure resulted in higher exergydestruction of the engine This is due to throttling and notfriction as can be seen in the table Friction is not a factor herebecause pressure does not affect the frictional loss of an RVengine significantly which is one of the unique characteristics


Table 4 Data of variation of operating speed

Operating speed (revsdotminminus1) 2400 3000 3600Second-law efficiency 386 296 221Exergy destroyed (W) 321 466 654Exergy supplied (W) 523 662 839Contribution of throttling (W) 149 175 179Contribution of friction (W) 123 226 374

of themechanismAt the same time a higher suction pressureresulted in higher exergy supplied as the fluid had a higherpotential to do useful work The exergy supply is increasedby the increase in suction mass flow rate too The total effectwas quantified by the second-order efficiency It is observedthat second-law efficiency climbed rapidly as suction pressureincreases from 4 to 6 bar but it stayed almost constant ataround 30 from 6 to 12 bar At 4 bar the exergy supplyrate was too low As suction pressure was increased boththe exergy supply and destruction rates increased at thesame pace Hence the second-law efficiency plateaus beyondsuction pressure of 6 bar

From this analysis the engine prototype was found to benot suitable for operation at suction pressures below 6 barThis is a specific behaviour of this particular engine not ageneral conclusion as the prototype was designed for suchapplications When the engine is needed for a low suctionpressure application some of the design parameters such asengine dimensions and bearing sizes should be modified toreduce the corresponding frictional losses

The effects of different operational speed to exergybehaviour of an RV engine are shown in Table 4 Gener-ally exergy destruction rate increased as rotational speedincreased The main losses came from higher fluid throttlinglosses and frictional losses The higher throttling loss wasbecause of the inability of the suction to catch up withthe increasing speed The higher frictional loss was becauseof the higher rotational acceleration which increased thecontact force and the corresponding frictional force at thevane contact line with the cylinder

At the same time a higher rotational speed also increasedthe mass flow rate resulting in a higher exergy supply rateHowever the exergy destruction rate increased faster over thesame range Hence the second-law efficiency decreased asspeed increased A low operating speed is hence favourableto maximise the work potential of the high-pressure fluidHowever a low operating speed will limit the output power ofthe engineTherefore a compromise has to be made betweenthe output power and the second-law efficiency in practice

The effects of different cylinder inertias to the exergycharacteristics of RV engine are tabulated in Table 5 It can beseen that exergy destruction increased as the cylinder inertiaincreased This is mainly because lower cylinder inertia gavea lower vane frictional loss Interestingly throttling loss wassmaller with a smaller inertia although not so significantlyThis was because of the increased internal leakage dueto larger radial clearance gap with a lighter cylinder Thebearings must be redesigned if a light cylinder was to be used

Table 5 Data of variation of cylinder inertia

Cylinder inertia (gsdotm2) 094 187 281Second-law efficiency 499 392 296Exergy destroyed (W) 332 403 466Exergy supplied (W) 662 662 662Contribution of throttling (W) 163 175 179Contribution of friction (W) 109 167 226

to keep this clearance gap small Meanwhile exergy suppliedby the fluid was maintained at a constant rate in this case asthe suction reservoir pressure and rotational speed were keptthe sameTherefore the second-law efficiencywas the highestwhen cylinder inertia is the smallest It can be seen that theeffect of inertia to second-law efficiency is significant Ascompared to suction pressure and operating speed cylinderinertia is the most practical to be modified as the othersare based on operating demands A smaller inertia can beachieved by removing the unnecessary mass of the cylinder(without compromising its dynamic balance) or by using alower density material (without compromising its strength)

4 Conclusion

Exergy analysis has been applied to the revolving vanecompressed air engine The air engine had a rotor radius of29mm cylinder radius of 35mm and length of 25mm Atthe benchmark conditions the suction pressure was 8 barthe discharge pressure was 1 bar and the operating speed was3000 revsdotminminus1 Temperatures of suction reservoir dischargereservoir and atmosphere were 25∘C Four main sourcesof exergy destruction were identified namely heat transferfriction throttling and fluid mixing From the analysis thefollowing was found

(i) Overall the air engine had a second-law efficiencyof 296 The main contributor of exergy loss wasfriction (49) followed by throttling (38) and heattransfer (12) The exergy loss due to fluid mixing isalmost negligible

(ii) The main contributor of the exergy destruction ratedue to heat transfer is in the suction chamber due tothe large temperature fluctuation of the fluid in it

(iii) The main contributor of exergy destruction due tofriction was that at the vane followed by the cylinderbearing while the rest were almost negligible Thiswas due to the high cylinder inertia of the engineprototype used The results suggest that it is crucialto reduce the cylinder inertia to improve the perfor-mance of the engine

(iv) The contributions of inflow outflow and internalleakage to the average exergy destruction due tothrottling were comparable From the findings itwas found that to reduce exergy destruction dueto throttling the uncovering processes of the ports


should be made faster This can be done by placingthe ports closer to the vane and by increasing thediameter of the ports For the leakage tightening theleakage gaps to reduce its flow rate will help to reducethe exergy destruction

(v) The exergy destruction due to mixing was dominatedby the inflow However the overall contribution offluid mixing to the exergy loss is very small (around1)

A parametric study was also conducted to analyse howdifferent parameters affect the exergy behavioursThe param-eters to be examined were suction reservoir pressure (from4 to 12 bar) operational speed of engine (from 2400 to3600 revsdotminminus1) and rotational cylinder inertia (from 094to 281 gsdotmm2) The study found the following

(i) A higher suction reservoir pressure resulted in bothhigher exergy destruction and supply of the engineHowever the rates of change were different so thesecond-law efficiency climbed rapidly as suction pres-sure increases from 4 to 6 bar but it stayed almostconstant at about 30 from 6 to 12 bar From thisanalysis the engine prototype was found to be notsuitable for operation at suction pressures below 6 barWhen the engine is needed for a low suction pressureapplication should be modified

(ii) Both exergy destruction and supply rates increased asthe rotational speed increased However the exergydestruction rate increased faster Hence the second-law efficiency decreased as speed increased A lowoperating speed is therefore favourable to maximisethe work potential of the high-pressure fluid How-ever it is noted that a low operating speed will limitthe output power of the engine

(iii) Exergy destruction increased as the cylinder inertiaincreased while the supply rate is constant Thesecond-law efficiency was the highest when cylinderinertia is the smallest As compared to suction pres-sure and operating speed cylinder inertia is the mostpractical and significant to be modified

Nomenclature

119864 Exergy supply rate (W)ℎ Specific enthalpy (Jsdotkgminus1)

119868 Exergy destruction rate (W) Mass flow rate (kgsdotsminus1)

119876 Heat transfer rate (W)119904 Specific entropy (Jsdotkgminus1)119879 Temperature (K) Power (W)

Greek Letters

120578 Efficiency

Subscripts

0 Ambient119891 Frictionht Heat transfer119894 Identifier for inlet flow stream119895 Identifier for thermal massth Throttlingmx Mixing

Conflict of Interests

The authors declare that there is no conflict of interestsregarding the publication of this paper

Acknowledgment

This work was financially supported in part by the Singa-pore National Research Foundation under its Campus forResearch Excellence and Technological Enterprise (CRE-ATE) programme

References

[1] R van Basshuysen and F Schaefer Internal Combustion EngineHandbook Basics Components Systems and Perspectives SAEInternational Warrendale Pa USA 2004

[2] M Sakita ldquoA cat-and-mouse type rotary engine engine designand performance evaluationrdquo Proceedings of the Institution ofMechanical Engineers D Journal of Automobile Engineering vol220 no 8 pp 1139ndash1151 2006

[3] Y-T Shen and Y-R Hwang ldquoDesign and implementation of anair-powered motorcyclesrdquo Applied Energy vol 86 no 7-8 pp1105ndash1110 2009

[4] Y L Teh and K T Ooi ldquoDesign and friction analysis of therevolving vane compressorrdquo in Proceedings of the InternationalCompressor Engineering Conference at Purdue C046 WestLafayette Ind USA July 2006

[5] Y L Teh and K T Ooi ldquoTheoretical study of a novel refrig-eration compressormdashpart I design of the revolving vane (RV)compressor and its frictional lossesrdquo International Journal ofRefrigeration vol 32 no 5 pp 1092ndash1102 2009

[6] Y L Teh and K T Ooi ldquoTheoretical study of a novel refrigera-tion compressormdashpart III leakage loss of the Revolving Vane(RV) compressor and a comparison with that of the rollingpiston typerdquo International Journal of Refrigeration vol 32 no5 pp 945ndash952 2009

[7] A Subiantoro and O K Tiow ldquoIntroduction of the revolvingvane expanderrdquoHVAC amp R Research vol 15 no 4 pp 801ndash8162009

[8] A Subiantoro and K T Ooi ldquoDesign analysis of the novelrevolving vane expander in a transcritical carbon dioxiderefrigeration systemrdquo International Journal of Refrigeration vol33 no 4 pp 675ndash685 2010

[9] A Subiantoro and K T Ooi ldquoComparison and performanceanalysis of the novel revolving vane expander design variantsin low and medium pressure applicationsrdquo Energy vol 78 pp747ndash757 2014

[10] International Transport ForumReducing Transport GreenhouseGas EmissionsmdashTrends amp Data Transport and InnovationUnleashing the Potential International Transport Forum 2010


[11] H Chen Y Ding Y Li X Zhang and C Tan ldquoAir fuelled zeroemission road transportation a comparative studyrdquo AppliedEnergy vol 88 no 1 pp 337ndash342 2011

[12] B R Singh and O Singh ldquoA study of performance output ofa multivane air engine applying optimal injection and vaneanglesrdquo International Journal of Rotating Machinery vol 2012Article ID 578745 10 pages 2012

[13] A Subiantoro and K T Ooi ldquoAnalytical study of the endfacefriction of the revolving vanemechanismrdquo International Journalof Refrigeration vol 34 no 5 pp 1276ndash1285 2011

[14] J A McGovern S Harte and G Strikis ldquoReal gas perfor-mance analysis of a scroll or rotary compressor using exergytechniquesrdquo in Proceedings of the International CompressorEngineering Conference (ICEC rsquo94) vol 979 pp 193ndash198 1994

[15] J AMcGovern and SHarte ldquoAn exergymethod for compressorperformance analysisrdquo International Journal of Refrigerationvol 18 no 6 pp 421ndash433 1995

[16] J A McGovern ldquoExergy analysismdasha different perspective onenergy part 1 the concept of exergyrdquo Proceedings of the Insti-tution of Mechanical Engineers Part A Journal of Power andEnergy vol 204 no 4 pp 253ndash262 1990

[17] J A McGovern ldquoExergy analysismdasha different perspective onenergy part 2 rational efficiency and some examples of exergyanalysisrdquo Proceedings of the Institution of Mechanical EngineersPart A Journal of Power and Energy vol 204 no 4 pp 263ndash2681990

[18] S S Stecco and G Manfrida ldquoExergy analysis of compressionand expansion processesrdquo Energy vol 11 no 6 pp 573ndash5771986

[19] J A McGovern ldquoCompressor rational efficienciesrdquo in Proceed-ings of the Institution of Mechanical Engineers Developments inIndustrial Compressors vol 10 pp 59ndash67 1989

[20] A V Reddy T S Kumar D K T Kumar B Dinesh and Y VS S Santosh ldquoEnergy and exergy analysis of IC enginesrdquo TheInternational Journal of Engineering and Science vol 3 no 5 pp7ndash26 2014

[21] M Ameri F Kiaahmadi M Khanaki and M NazoktabarldquoEnergy and exergy analyses of a spark-ignition enginerdquo Inter-national Journal of Exergy vol 7 no 5 pp 547ndash563 2010

[22] C Sayin M Hosoz M Canakci and I Kilicaslan ldquoEnergy andexergy analyses of a gasoline enginerdquo International Journal ofEnergy Research vol 31 no 3 pp 259ndash273 2007

[23] J A Caton ldquoExergy destruction during the combustion processas functions of operating and design parameters for a spark-ignition enginerdquo International Journal of Energy Research vol36 no 3 pp 368ndash384 2012

[24] S Jafarmadar R Tasoujiazar and B Jalilpour ldquoExergy analysisin a low heat rejection IDI diesel engine by three dimensionalmodelingrdquo International Journal of Energy Research vol 38 no6 pp 791ndash803 2014

[25] A Subiantoro K S Yap and K T Ooi ldquoExperimental investi-gations of the revolving vane (RV-I) expanderrdquoAppliedThermalEngineering vol 50 no 1 pp 393ndash400 2013

[26] A C Yunus and M A BolesThermodynamics An EngineeringApproach McGraw Hill New York NY USA 5th edition 2016

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Research ArticleExergy Analysis of the Revolving Vane Compressed Air Engine

Alison Subiantoro12 Kin Keong Wong3 and Kim Tiow Ooi23

1Department of Industrial and Systems Engineering Krida Wacana Christian University Jalan Tanjung Duren Raya No 4Jakarta Barat 11470 Indonesia2Technical University of Munich-Campus for Research Excellence and Technological Enterprise (TUM CREATE)1 CREATEWay No 10-02 CREATE Tower Singapore 1386023School of Mechanical and Aerospace Engineering Nanyang Technological University 50 Nanyang Avenue Singapore 639798

Correspondence should be addressed to Alison Subiantoro alisonsubiantoroukridaacid

Received 3 December 2015 Revised 14 January 2016 Accepted 14 January 2016

Academic Editor Zuohua Huang

Copyright copy 2016 Alison Subiantoro et al This is an open access article distributed under the Creative Commons AttributionLicense which permits unrestricted use distribution and reproduction in any medium provided the original work is properlycited

Exergy analysis was applied to a revolving vane compressed air engineThe engine had a swept volume of 30 cm3 At the benchmarkconditions the suction pressure was 8 bar the discharge pressure was 1 bar and the operating speed was 3000 revsdotminminus1 It wasfound that the engine had a second-law efficiency of 296 at the benchmark conditions The contributors of exergy loss werefriction (49) throttling (38) heat transfer (12) and fluidmixing (1) A parametric study was also conductedThe parametersto be examined were suction reservoir pressure (4 to 12 bar) operating speed (2400 to 3600 revsdotminminus1) and rotational cylinderinertia (094 to 281 gsdotmm2) The study found that a higher suction reservoir pressure initially increased the second-law efficiencybut then plateaued at about 30 With a higher operating speed and a higher cylinder inertia second-law efficiency decreased Ascompared to suction pressure and operating speed cylinder inertia is the most practical and significant to be modified

1 Introduction

Since its inception in the 1800s reciprocating piston engineshave been widely used to drive motor vehicles Howevereven with all the progress and innovations in the reciprocat-ing engine technology there are still issues and challengesfacing the engine design To overcome these there havebeen attempts to come up with new engine designs Rotaryengines have been proposed as a possible design alternativeAs compared to a reciprocating engine rotary engines areusually more compact and have less number of componentswith better vibration and noise characteristics In responseWankel developed a rotary piston engine in the 1960s whichhas been adopted by Mazda to drive its sport vehiclesHowever leakage is one of the main issues with this enginedesign [1] More recently Sakita introduced a new rotarymachine design called the ldquocat andmouse typerdquo rotary engine[2] Shen and Hwang adopted the sliding vane machinedesign for an air-powered motorcycle [3]

In 2006 a new rotary machine design concept calledthe revolving vane (RV) mechanism was introduced by Teh

and Ooi [4] An improved version was later proposed toimprove the mechanical efficiency of the machine [5] Unlikethe more common rotary machines which use stationarycylinders the RV machine lets the cylinder move togetherwith the rotor This significantly reduces the frictions of themachine In addition it is also found that the RV machineexhibits better volumetric performance as compared to arolling piston machine [6] The machine was first introducedas a compressor but has subsequently been used in expanderand air engine applications too [7ndash9]

In its simplest form the RV mechanism consists of arotor a vane and a cylinder as shown in Figure 1(a) Howeverunlike other rotary machines the cylinder rotates togetherwith the rotor resulting in smaller relative velocities and lessfrictional losses at the rubbing surfaces The rotor and thecylinder are arranged eccentrically and they rotate at theirrespective axes The vane divides the space between the rotorand the cylinder inner wall into two that is the suction anddischarge chambers As the rotor rotates the volumes of thesechambers vary allowing for thermodynamics processes to

Hindawi Publishing CorporationInternational Journal of Rotating MachineryVolume 2016 Article ID 5018467 8 pageshttpdxdoiorg10115520165018467


Inlet

Discharge port

Vane slot

Vane

Suction port

Casing

Rotor

Cylinder

(a)

(A) (B)

(D) (C)

(b)










2 Methodology






119868ht = 1198761198951198790 (1

119879

minus

1

119879

119895

)

119868

119891=

119891(

119879

0

119879

119895

)

119868th = 1198790 (1199041015840

119894minus 119904

119894)

119868mx = 1198790((119904 minus 1199041015840

119894) minus

(ℎ

1015840

119894minus ℎ

119894)

119879

)

(1)



1015840

119894= ℎ

119894


120578 = 1 minus

119868

119864

(2)

119864 = [(ℎ minus ℎ

0) minus 119879

0(119904 minus 119904

0)] (3)














Discharge port



Housing body

Lip seal

Lower bearing

Housing cover

Housing


Cylinder and cover


0

50

100

150

200

0 60 120 180 240 300 360

Heat transfer


Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)

(a)

0

20

40

60

80

100

0 60 120 180 240 300 360


Rotor angle (∘)

Tem

pera

ture

(∘C)

minus20

(b)




00

60 120 180 240 300 360

Friction



Rotor angle (∘)

400

300

200

100

Exer

gy d

estr

uctio

nra

te (W

)





0

400

800

1200

1600

0 60 120 180 240 300 360Exer

gy d

estr

uctio

n ra

te (W

)

Throttling


Outflow

Rotor angle (∘)

(a)

01234567

98

0 60 120 180 240 300 360

Pres

sure

(bar

)

Rotor angle (∘)

SuctionDischarge

(b)











020406080

100120140

0 60 120 180 240 300 360

Fluid mixing


Outflow

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)














4 Conclusion













Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation










Inlet

Discharge port

Vane slot

Vane

Suction port

Casing

Rotor

Cylinder

(a)

(A) (B)

(D) (C)

(b)










2 Methodology






119868ht = 1198761198951198790 (1

119879

minus

1

119879

119895

)

119868

119891=

119891(

119879

0

119879

119895

)

119868th = 1198790 (1199041015840

119894minus 119904

119894)

119868mx = 1198790((119904 minus 1199041015840

119894) minus

(ℎ

1015840

119894minus ℎ

119894)

119879

)

(1)



1015840

119894= ℎ

119894


120578 = 1 minus

119868

119864

(2)

119864 = [(ℎ minus ℎ

0) minus 119879

0(119904 minus 119904

0)] (3)














Discharge port



Housing body

Lip seal

Lower bearing

Housing cover

Housing


Cylinder and cover


0

50

100

150

200

0 60 120 180 240 300 360

Heat transfer


Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)

(a)

0

20

40

60

80

100

0 60 120 180 240 300 360


Rotor angle (∘)

Tem

pera

ture

(∘C)

minus20

(b)




00

60 120 180 240 300 360

Friction



Rotor angle (∘)

400

300

200

100

Exer

gy d

estr

uctio

nra

te (W

)





0

400

800

1200

1600

0 60 120 180 240 300 360Exer

gy d

estr

uctio

n ra

te (W

)

Throttling


Outflow

Rotor angle (∘)

(a)

01234567

98

0 60 120 180 240 300 360

Pres

sure

(bar

)

Rotor angle (∘)

SuctionDischarge

(b)











020406080

100120140

0 60 120 180 240 300 360

Fluid mixing


Outflow

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)














4 Conclusion













Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation













119868ht = 1198761198951198790 (1

119879

minus

1

119879

119895

)

119868

119891=

119891(

119879

0

119879

119895

)

119868th = 1198790 (1199041015840

119894minus 119904

119894)

119868mx = 1198790((119904 minus 1199041015840

119894) minus

(ℎ

1015840

119894minus ℎ

119894)

119879

)

(1)



1015840

119894= ℎ

119894


120578 = 1 minus

119868

119864

(2)

119864 = [(ℎ minus ℎ

0) minus 119879

0(119904 minus 119904

0)] (3)














Discharge port



Housing body

Lip seal

Lower bearing

Housing cover

Housing


Cylinder and cover


0

50

100

150

200

0 60 120 180 240 300 360

Heat transfer


Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)

(a)

0

20

40

60

80

100

0 60 120 180 240 300 360


Rotor angle (∘)

Tem

pera

ture

(∘C)

minus20

(b)




00

60 120 180 240 300 360

Friction



Rotor angle (∘)

400

300

200

100

Exer

gy d

estr

uctio

nra

te (W

)





0

400

800

1200

1600

0 60 120 180 240 300 360Exer

gy d

estr

uctio

n ra

te (W

)

Throttling


Outflow

Rotor angle (∘)

(a)

01234567

98

0 60 120 180 240 300 360

Pres

sure

(bar

)

Rotor angle (∘)

SuctionDischarge

(b)











020406080

100120140

0 60 120 180 240 300 360

Fluid mixing


Outflow

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)














4 Conclusion













Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation










Discharge port



Housing body

Lip seal

Lower bearing

Housing cover

Housing


Cylinder and cover


0

50

100

150

200

0 60 120 180 240 300 360

Heat transfer


Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)

(a)

0

20

40

60

80

100

0 60 120 180 240 300 360


Rotor angle (∘)

Tem

pera

ture

(∘C)

minus20

(b)




00

60 120 180 240 300 360

Friction



Rotor angle (∘)

400

300

200

100

Exer

gy d

estr

uctio

nra

te (W

)





0

400

800

1200

1600

0 60 120 180 240 300 360Exer

gy d

estr

uctio

n ra

te (W

)

Throttling


Outflow

Rotor angle (∘)

(a)

01234567

98

0 60 120 180 240 300 360

Pres

sure

(bar

)

Rotor angle (∘)

SuctionDischarge

(b)











020406080

100120140

0 60 120 180 240 300 360

Fluid mixing


Outflow

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)














4 Conclusion













Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation










0

400

800

1200

1600

0 60 120 180 240 300 360Exer

gy d

estr

uctio

n ra

te (W

)

Throttling


Outflow

Rotor angle (∘)

(a)

01234567

98

0 60 120 180 240 300 360

Pres

sure

(bar

)

Rotor angle (∘)

SuctionDischarge

(b)











020406080

100120140

0 60 120 180 240 300 360

Fluid mixing


Outflow

Rotor angle (∘)

Exer

gy d

estr

uctio

nra

te (W

)














4 Conclusion













Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation




















4 Conclusion













Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation
















Nomenclature




Greek Letters

120578 Efficiency

Subscripts




Acknowledgment


References






























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation




























RoboticsJournal of





Journal of



RotatingMachinery





VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation









research article exergy analysis of the revolving vane … · 2020. 3. 7. · research article...

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