A Micro Gas Turbine for UK Domestic Combined Heat and Power

Transcription

1 A Miro Gas Turbine for UK Domesti Combined Heat and Power 1 A Clay, and GD Tansley Shool of Engineering and Applied Siene, Aston University, Birmingham, United Kingdom Abstrat: Various miro radial ompressor onfigurations were investigated using 1D meanline and CFD tehniques for use in a Miro Gas Turbine (MGT) Domesti Combined Heat and Power (DCHP) appliation. Blade baksweep, shaft speed, and blade height were varied at a onstant pressure ratio. Shaft speeds were limited to 0,000 rev/min, to enable the use of a turboharger bearing platform. Off-design ompressor performane was established and used to determine the MGT performane envelope; this in turn was used to assess potential ost and environmental savings in a heat-led DCHP operating senario within the target market of a detahed family home. A low target stage pressure ratio provided an opportunity to redue diffusion within the impeller. Critially for DCHP, this produed very regular flow whih improved impeller performane for a wider operating envelope. The best performing impeller was a low speed, 170,000 rev/min, low baksweep, 15, onfiguration produing a 71.76% st age effiieny at a pressure ratio of.0. This produed a MGT design point system effiieny of 14.85% at 993, mathing prime movers in the latest ommerial DCHP units. Cost and CO savings were 10.7% and 6.3% respetively for annual power demands of 17.4 Mht and 6.1 Mhe ompared to a standard ondensing boiler (with grid) installation. The maximum ost saving (on design point) was 14.% for annual power demands of.6 Mht and 6.1 Mhe orresponding to an 8.1% CO saving. hen sizing, maximum savings were found with larger heat demands. hen sized, maximum savings ould be made by enouraging more eletriity export either by reduing household eletriity onsumption or inreasing mahine effiieny. Corresponding author: Mehanial Engineering and Design, Shool of Engineering and Applied Siene, Aston University, Aston Triangle, Birmingham, B4 7ET, UK. laya@aston.a.uk

2 A Clay, GD Tansley 1. Introdution The feasibility of a Miro Gas Turbine (MGT) Domesti Combined Heat and Power (DCHP) unit was previously assessed [1] wherein urrent tehnologial limits suggested a net power, & net, of 1 ke would produe a system effiieny, η s, of 15%, an improvement over existing (1%) [] and latest (14%) [3] ommerial DCHP prime movers. Higher system effiienies are required to provide better finanial and environmental inentives to the onsumer [4]. The use of a turboharger bearing platform is, at present, an aessible tehnology and remains a simple method for produing a low-ost unit within a marketable prie range. This paper investigates the potential for a MGT DCHP through the design, and performane analysis of a miro ompressor by Computational Fluid Dynamis (CFD).. Outline Compressor Design To allow the use of oil ooled journal bearings, shaft speeds less than 0,000 rev/min were required in a devie whih will deliver a target pressure ratio, r, of.15 at a ompressor effiieny, η, of 73 %. To avoid the use of parasiti devies or ompressor bleeding, other miro ompressors with similar MGT duties are looking to adopt non-ontat aerodynami air foil bearings [5] whih require a high temperature onformal oating [6] for shaft speeds of 500,000 rev/min with an impeller diameter of 0 mm. The stage effiieny disadvantages of small, high speed impellers are two fold: firstly, relative tip learane inreases due limitations in manufaturing toleraning, and seondly, a Reynolds number redution suggests a redued aerodynami effiieny [7]. In a bulkier set up, the speed limitation of the oil ooled journal bearing will require a 40 mm impeller diameter and restrits maximum attainable pressure ratio. However, and in spite of a system effiieny penalty [8], a lower pressure ratio an inrease ompressor stage operating range [9] whih is an important riterion for DCHP during periods of low power demand. The performane advantages of slightly larger turbomahinary omponents would seem to outweigh the size penalty in a DCHP appliation where MGTs have a natural advantage over existing prime movers suh as Stirling or Internal Combustion Engines (ICEs).

3 3. 1D Compressor design methodology A Clay, GD Tansley 3 Centrifugal stress was aommodated for by speifying the disharge tangential veloity omponent, U, < 470 m/s to permit the use of AlC355 T6 alloy [9], a material urrently used in the mass manufature of entrifugal impellers by investment asting for turbo hargers. An inlet shroud blade angle, β 1s, of around 61 was used to provide maximum flow apaity [10], a harateristi shared by most modern impeller inlets [11]. Mah numbers were limited to < 0.7 [10]. An absolute disharge flow angle, α m, of 65 was seleted to prevent reverse flow in th e vaneless diffuser [1] [13]. A mass flow, m&, of 0 g/s refleted the 1 k MGT net power, requirement and target pressure ratio, r. The 1D ompressor stage effiieny, η, was set to 75%. The remaining variables were alulated by ontinuity of mass, equation of state, Euler s turbomahinary equation and vetor diagrams. The remaining & net, variables: blade baksweep, β b, shaft speed, ω, impeller inlet hub radius, r 1 h, and impeller disharge radius, r, were varied in the optimization proess. It is generally onsidered that inreasing blade baksweep, β b, will provide a more stable and wider operating range [1] due to improved diffusion resulting from a uniform flow pattern at disharge [14]. From a geometri perspetive inreasing blade baksweep, β b, was shown to inrease required blade height, b, and redue pressure ratio, r, at onstant shaft speed, N ; or inrease blade height, b, and shaft speed, N, at onstant pressure ratio, r ; see Figure 1. Previous investigators [5] [15] have preferred to design miro impellers on the basis of speifi speed between for optimal effiieny [16]. Due to the lower speed restritions of the journal bearing platform, the speifi speed range for this design was , far below the optimum even with the largest blade baksweep.

4 A Clay, GD Tansley 4 Blade height, (m) ,000 rev/min 10,000 rev/min 00,000 rev/min 190,000 rev/min 180,000 rev/min 170,000 rev/min ,000 rev/min Blade baksweep, (deg) Figure 1 Relationship desribing the effet of inreased blade height, b, from inreasing blade baksweep, β b, and its impat of reduing pressure ratio, r, with limiting shaft speed, N. 4. 3D Compressor geometry definition A design ode was written in Matlab 1 whih produed.txt and.jou files needed to desribe the oordinates and geometry onstrution of the entrifugal impeller for use by Gambit (geometry modelling software). Bezier splines were used to desribe the meridional profile whilst polar oordinates were used to desribe the radial loation of eah parametri interval. The amber and blade angles were subsequently alulated following [13]. The Bezier equations were first represented in an Exel 3 spreadsheet where D plots representing the Meridional profile and Camber line were initially used to examine blade urvature. Flow area was alulated using a trapezoidal funtion at eah parametri interval, aross the hannel and between the 1 The Mathworks, In., 3 Apple Hill Drive, Natik, MA , USA. Version (R008a). ANSYS, In., Southpointe, 75 Tehnology Drive, Canonsburg, PA 15317, USA. Gambit version Mirosoft Corporation, One Mirosoft ay, Redmond, A , USA. Mirosoft Offie Exel 003 ( ) SP3.

5 A Clay, GD Tansley 5 hub and shroud ontours. The parametri intervals of the hub ontour were iterated to ensure the alulated flow area would be perpendiular to the mean flow path between the hub and shroud ontours. A 3D examination of the blade urvature was performed via Solidworks 4 using a design table linked to the Exel spreadsheet. In Exel, slight iterations were made to the meridional profile and polar oordinates of the parametri points whih instantly updated the Solidworks model. One satisfied with the onsisteny of urvature [9] and flow area to promote stable flow, the final oordinates were read from the spreadsheet by the Matlab ode to generate the.txt and.jou files for geometry generation in Gambit. The mesh was then exported into Fluent 5 (CFD software) where it was solved 3 dimensionally. No additional ommerial software was required. 5. CFD methodology The flow model onsisted of induer, hannel and diffuser fluid volumes along the axial diretion eah with individual tip learane volumes. A rotational periodi ondition was set up using a single hannel with interior faes between the induer/hannel, hannel/diffuser volumes and hannel/hannel tip learane. The hannel volume onsisted of the flow volume around the splitter between blade pressure side to blade sution side; see Figure. 4 DS Solidworks Headquarters, Dassault Systèmes Solidworks Corp., 300 Baker Avenue, Conord, MA Version 008 SP4.0 5 ANSYS, In., Southpointe, 75 Tehnology Drive, Canonsburg, PA 15317, USA. Fluent version

6 A Clay, GD Tansley 6 SS Tip learane volumes Diffuser PS Channel Induer Figure Impeller was modelled as a single rotating hannel volume with splitter, separated from stationary volumes; induer, diffuser and tip learane by interior faes. Periodi funtions were arranged at the pressure and sution sides (PS & SS) of induer and diffuser. An impliit, steady, pressure-based solver was used. The RNG k-ε visous turbulene model was used due to the urved surfaes with non-equilibrium wall funtions and visous heating to aount for ompressibility affets [17]. Based on Hydrauli diameter, other investigations saw Reynolds numbers less than 5000 for the smallest 3D miro ompressors [18] lose to the laminar/transition region used in pipe flow analogy. In this investigation alulated Reynolds number were 15,000 at design point following [19] suggestion that the use of turbulene modelling was suitable. The material was air, modelled as an ideal gas with a pieewise polynomial funtion for speifi heat apaity. Under-relaxation fators were onservative between 0.1 or 0.. Residual onvergene monitors were set to Convergene also used a fore monitor with a moment oeffiient on the blade surfaes and required mass flux imbalane to be less than 1-08 kg/s for a mass flow of -0 kg/s. The SIMPLE pressure-veloity oupling solution was invoked. Grid onvergene was ahieved at around 500,000 elements, but reasonable and onservative values for stage effiieny and pressure ratio were ahieved at 100,000 elements; at whih grid

7 A Clay, GD Tansley 7 densities stage effiieny, η, was under predited by % and pressure ratio, r, was under predited by 3.5%. All disretization was first order upwind apart from pressure whih was standard. In the interests of time, solutions with low mesh densities and first order disretization were used to produe onservative solutions suitable for omparison between different impeller geometries. A tip learane of 0.3 mm was measured and used from a Garrett turboharger 6 to reflet the manufaturing auray of mass produed turbomahinary omponents and radial growth at similar operating speeds. Splitters were positioned approximately /3 up the hannel. Splitter blades are used to give wider range during off-design [0] but in addition splitters brought a 1% effiieny inrease aross the stage at design point likely due to limiting slip effets. Splitter position sensitivity has been found to provide a stage effiieny inrease of between 1-% at design point [1]. No investigation into splitter position was performed in this study. All reported effiienies are based on total properties unless otherwise stated. 6. 3D Design point optimization 6.1 Blade Baksweep An initial two-zone impeller optimization ode following [1] was developed but later aborted when the impat of input variables ould not be assessed from the results. Instead, outline impeller geometry and stati pressure values were established from simple 1D meanline analysis. Several impeller geometries (denoted a to f) were reated by adjusting blade baksweep, β b, at different shaft speeds, N, to deliver a pressure ratio, r, of.15. Inlet hub blade angle, β 1h, mass flow, m&, and inlet hub radius, r 1 h, remained onstant; see Figure 3. Only impellers a (15 ) and b (31 ) are shown to have deelerating flow, DR >1 whilst the others had aelerating, DR <1 flow (see Table 1for details of eah impeller). 6 Honeywell International In., 101 Columbia Road, Morristown, NJ 0796, USA. Garrett Turbohargers by Honeywell, Small frame, GT1(41) family.

8 A Clay, GD Tansley 8 These impeller geometries were assessed in 3D using CFD, eah with a diffuser length of 5 mm; see Figure 4. Varying baksweep angle, β b, between 15 and 54 resulted in a gentle downward trend in impe ller effiieny, η i, and a flutuating downward trend in total stage pressure ratio, r, stage effiieny, η, showed little hange, and impeller pressure ratio dereased, again with a flutuation at impeller (44 ). In larger mahines, stage effiieny, η, should inrease by 1 points for every 10 of blade baksweep, β b [1]. Plus at miro sale, a redution in blade height, b, has shown to ause an effiieny penalty by reduing relative tip learane thus inreasing aerodynami loss []; neither of whih present themselves here. The trend for inreasing total stage pressure ratio, r, for dereasing blade baksweep, β b, is onfirmed by the 1D alulations of Figure 1 in terms of total pressure ratio. Aording to [14] inreased blade baksweep, β b, aelerates the flow, whih redues diffusion and blade loading within the impeller, minimizing seondary flow development to improve impeller effiieny but redue stati pressure generation. Reduing diffusion is learly seen in 1D from Figure 3 but not refleted in the 3D results. In this investigation, aelerated flow is attributed to dereasing disharge area or blade height, b, a onsequene of redued blade baksweep, β b. Inreased kineti energy explains inreasing total pressure ratio within the impeller and stage with dereasing blade baksweep, β b, and inreased impeller effiieny, η i. ithout aerodynami losses from redued blade height, b, the harateristis here are more like onventional sized impellers than the smaller diameter miro impellers seen in other investigations. Improved impeller effiieny, η i, from redued blade baksweep, β b, is a onsistent observation with [0] and numerial data using total properties from [14]. Slight improvements in impeller effiieny, η i, an also be attributed to redued frition from a smaller meridional hord or flow path length. In this investigation, blade wrap angle was redued from 95 on impeller e, to 55 on impeller a. Cross referening Figure 3, Figure 4, and previous work suggests ompressor performane is a trade off among diffusion ratio, DR, blade height, b, blade

9 A Clay, GD Tansley 9 baksweep, β b, and inlet onditions (reduing shaft speed, N, redued the meridional omponent, C m1, and area, A 1, at inlet by lowering the impeller tangential omponent, U 1 h ). Further investigation was performed and is presented below. Shaft speed, (rev/min) a = 15 b = 31 = 44 d = 54 e = 61 f = Blade baksweep, (deg) Shaft speed blade height Diffusion Ratio Blade height, (mm) & Diffusion ratio, Figure 3 Diffusion ratio, DR, and blade height, b, established from 1D meanline analysis of ompressor impellers with varying blade baksweep, β b, fored to run at different shaft speeds, N, by assigning onstant pressure ratio, r, of.15 and rotor diameter, r of 40 mm.

10 A Clay, GD Tansley 10 Compressor stage,, and impeller, effiieny (%) Pressure ratio, Blade baksweep, (deg) stage effiieny impeller effiieny stage pressure ratio (total) stage pressure ratio (stati) impeller pressure ratio (total) impeller pressure ratio (stati) Figure 4 Compressor stage effiieny, η, impeller stage effiieny, η i, and pressure ratio, r, established from CFD alulations using the impellers generated for Figure Blade height Figure 4, demonstrated that the slowest impellers with minimum blade baksweep, β b, produed the greatest diffusion ratios, DR, with the smallest blade heights, b. At 1D, reduing blade baksweep, β b, raised diffusion ratio, DR, by inreasing the disharge relative veloity omponent,, from an inrease in the disharge tangential veloity omponent, C θ. As a onsequene the disharge radial omponent, C m, must also inrease whih redues disharge area, A, and so blade height, b, for a onstant tip radius, r. To independently verify the interplay between baksweep, β b, and blade height, b, on ompressor performane, blade height, b, was redued on impellers b to e ompared with impeller a (the smallest) produing impellers b b,min to e b,min ; relative results with a 5 mm diffuser length are shown in Figure 5. For impeller e the redution in blade height, b, had a very positive impat all round. For the 3 remaining impellers, b b,min to d b,min, a redution in Euler head had an overall negative impat on impeller performane. A onsistent impeller effiieny inrease

11 A Clay, GD Tansley 11 with diffusion ratio, DR, redution is found for impellers b b,min to e b,min. ith less diffusion, the effiieny improvement learly ours as a result of aelerating flow from a redution in disharge area, A, to produe stable flow with less instability as outlined previously. The perentage differene in pressure ratio, r, between predited 1D mean flow and 3D CFD was very small, as shown in Table. This suggests the prodution of a minimal seondary or wake flow region sine 1D meanline analysis only aounts for the primary or jet flow. The following onstants were used for the yle analysis: turbine effiieny, η t, 75%, mehanial effiieny, η m, 90%, burner effiieny; η b, 98%, reuperator effetiveness, η HEX, 75%, pressure drop value, P, 90% Perentage hange from original, (%) Impeller name Blade Baksweep, stage effiieny impeller effiieny stage pressure ratio (total) stage pressure ratio (stati) impeller pressure ratio (total) impeller pressure ratio (stati) Diffusion ratio MGT effiieny MGT net power Figure 5 Perentage hange of ompressor and MGT performane with ompressor impellers of different blade baksweep, β b, and shaft speed, N, after using inlet onditions and small blade height, b, from impeller a. Compare with Figure 4. Table 1 Impeller summary Impeller Baksweep [ ] Speed [rev/min] Blade height [mm] DR a ,

12 A Clay, GD Tansley 1 b , , d 54 00, e 61 10, f 65 0, Table Differene between predited 1D meanline and 3D CFD results of pressure ratio, r, before and after hanging blade height. Impeller 1D predited 3D CFD result perentage differene(%) a b d e b b,min b,min d b,min e b,min Inlet geometry Impellers are traditionally designed with enough relative diffusion to provide a ontrolled maximum stati pressure rise for stable ombustion downstream (typially fluid veloity < 90 m/s [3]). Diffusion ratio, DR, an be inreased by inreasing the inlet meridional veloity omponent, C m1, to redue inlet area, A 1, and raise inlet relative shroud veloity, C m1 1 s.in this exerise the inlet meridional veloity omponent,, was inreased by inreasing the inlet hub radius, r 1 h, whilst maintaining the optimum inlet blade shroud angle, β 1s, of 61 with a 5 mm diffuser length. Impeller d was the baseline; relative results are shown in Figure 6. Reduing inlet area A 1, to raise the diffusion ratio, DR, by inreasing inlet relative shroud veloity, 1 s,

13 A Clay, GD Tansley 13 worsened ompressor performane most likely due to inreased blokage effets [0]. For a low stage pressure rise impeller, the magnitude of dynami pressure onversion or diffusion demand is less. Providing the dynami portion of the total pressure is small enough to maintain ombustion downstream, relative diffusion an be limited to provide maximum impeller effiieny. 0 Impeller name Perentage hange from original, (%) Hub radius,, [mm] stage effiieny impeller effiieny stage pressure ratio (total) stage pressure ratio (stati) impeller pressure ratio (total) impeller pressure ratio (stati) Diffusion ratio MGT effiieny MGT net power Figure 6 Perentage hange of ompressor and MGT performane of impeller d when inlet hub radius, r 1 h, is inreased to inrease diffusion ratio, DR, by reduing inlet area, A 1, against a onstant disharge area, A. 7. Compressor off design A 3D CFD off-design study was performed on impellers a and d see Figure 7. These impellers showed similar on-design performane but with opposing harateristis; low-speed, small blade baksweep vs. high-speed, large blade baksweep. A large blade baksweep an improve off-design performane, whilst lower speeds are benefiial for various mehanial reasons. ithout test data, suitable stati pressure values for the mass flow inlet, pressure outlet boundaries and operating pressure were found by onduting 1D

14 A Clay, GD Tansley 14 ompressor off-design analysis. Iterations on inlet stati temperature, T 1, and disharge radial veloity omponent, C m, to preserve mass ontinuity established 1D solutions from outline geometry. 1D effiieny was alulated following [4] based on iterating the pipe flow frition fator using the Colebrook-hite equation Pressure Ratio, = 15 = 54 Mass flow, (kg/s) 130, , , ,000 10, Shaft speed, (rev/min) Figure 7 3D off-design ompressor map from CFD for impeller a and impeller d 8. Off design MGT performane urve In order to use the ompressor map to establish gas turbine performane, an algorithm was written to establish the Turbine Inlet Temperature (TIT) and system effiieny, η s, from the pressure ratio, r, stage effiieny, η, and mass flow, m&, at eah ompressor off-design point. The ompressor off-design pressure ratio, r, was mathed with an optimum pressure ratio, r ', alulated by Brayton yle analysis, see Figure 8 for more details. The resulting gas turbine performane from eah ompressor off-design point produed the satter shown in Figure 9 and Figure 10 for impellers a and d respetively. From the satter, a urve based on distint gas turbine operating points operating on a least fuel operating strategy was fitted. Absene of satter indiates a least fuel off-design operating strategy is not possible in that region, hene the urves represent best possible off-design performane in

15 A Clay, GD Tansley 15 terms of least fuel. By definition the off-design strategy assumes variable speed for maximum system effiieny [5] whih will require inverter eletronis. The yle analysis onstants from Part 6 were again used here: mehanial effiieny, η m, 90%, burner effiieny; η b, 98%, reuperator effetiveness, η HEX, 75%, pressure drop value, P, 90%. To aount for off design turbine effiieny, η t, remained + % higher than the ompressor effiieny []. A 3 rd order polynomial funtion was fitted to eah performane urve to desribe the off-design relationship between net power, & net, and system effiieny, η s. The equations are shown below; β = 15, b β = 54, b η s = net net net η s = net net net Input ompressor data, perform Brayton yle analysis at different TIT Compressor example point 1 N r m& η Input ompressor data, perform Brayton yle analysis to establish r' at eah TIT 3 Seleting the TIT whih enables the ompressor point to run at its optimum η s net T 03 r ' η s net Eah ompressor point is analyzed and plotted, see Figure 9. Curve is fitted to example point not point 1 due to a lower fuel rate and higher effiieny Figure 8 Algorithm flow hart Off-design. This algorithm was used to alulate eah of the points in the satter plot of Figure 9 and Figure 10.

16 A Clay, GD Tansley ex ex System effiieny, (%) Correlation Coeffiients Data Curve Performane Fuel Fuel flow rate, (g/s) Net Power, Performane points Performane (least fuel) Performane 3rd poly fit Fuel onsumption points Fuel onsumption (least fuel) () Figure 9 Calulated MGT off design performane for impeller a, 15 blade baksweep, β b, impeller. Illustrated is the omparison between example point 1 shown expliitly in Figure 8 and example point whih, like every other satter point, was alulated in a similar way. The fitted urve passed through example point due to its lower fuel flow rate, m& f System effiieny, (%) Correlation Coeffiients Data Curve Performane Fuel Fuel flow rate, (g/s) Net Power, Performane data Performane (least fuel) Performane 3rd poly fit Fuel onsumption data Fuel onsumption (least fuel) () Figure 10 β b, impeller. Calulated MGT off design performane for impeller d, 54 blade baksweep,

17 9. Off design MGT DCHP performane A Clay, GD Tansley 17 Annual DCHP performane was predited using the MGT off design performane urves to aount for hanges in required power and MGT effiieny from varying seasonal loads. Average monthly power demands for a detahed house, the target market for DCHP units, were taken from [6]. For this building the annual loads were 17.4 Mht and 6.1 Mhe. Two other annual thermal loads were analysed at 70% and 130% (.6 Mht and 1.30 Mht) of the average (17.4 Mht) at a range of eletrial loads between 0% to 00% (. Mhe to 1. Mhe) of the average (6.1 Mhe). Using equations 1 and to represent a gas turbine with impellers a and d respetively, monthly DCHP analysis was performed. The analysis assumed a ontinuous operation strategy, the preferred operating regime for gas turbines and DCHP, utilising thermal storage. The omparative was a standard grid onnetion with a modern ondensing boiler; see Figure 11 In the three senarios, maximum monthly average DCHP power demands for eah annual heat demand were 435 e, 699 e, 1039 e (limited to 950e) and 345 e, 581 e, 817 e respetively for impellers a and d. Due to a superior offdesign performane, impeller a produed larger DCHP savings by generating more eletriity for export at speified heat demands. Generator effiieny, η GEN, was 85%, exhaust gas to water heat exhanger effetiveness, η HEX various ost and emission fators 0.194, E e = 0.396,, was 90%, ondensing boiler effiieny, η CB, was 90%, the C g = , C e = , C ex = 0.05 [7], E g = E ex = The emission fators used in this study were derived from Diretive 004/08/EC whih onsiders exported eletriity as arbon free and would displae entrally generated output. hen sizing, for a heat-led mahine of speified output power, maximum ost savings are shown to vary with an optimal annual eletrial load. The optimal eletrial load and maximum ost savings inrease with an inreasing annual heating load. This is due to the MGT being able to run at a higher output and produing better effiienies. Less annual heating demand, produed less ost savings but also redued the optimum height, see the optimum lous for impellers a and d in Figure 11. Reduing sensitivity with respet to the optimum eletrial load ould provide a more flexible appliation in terms of ost for lower heating loads. The differene in

18 A Clay, GD Tansley 18 ost savings between the off-design performanes of the two gas turbines is less pronouned with larger annual heat demands. Maximum CO savings are onsistent with the most effiient mahine and higher annual heat demands. hen sized, ost and CO savings are proportional to eletriity export. The magnitude of eletriity export is a funtion of redued eletrial loads, and better prime mover effiieny, as also found during DCHP field trials [4]. For a building with average annual heating demand (17.4 Mht), the average eletrial annual eletrial load (6.1 Mhe) must derease by approximately 1/3 for optimum ost savings. This provides users, installers and appliane makers with an inentive to ontinually redue eletriity onsumption with DCHP, whih importantly also yields maximum CO savings Perentage savings [%] Annual eletrial load R [Mh] Figure 11 Annual perentage ost (diamond points) and C0 (square points) savings from DCHP ompared to a standard grid onnetion with ondensing boiler at various annual eletrial loads for three annual heat loads of 1.8 Mht (dotted line), 17.4 Mht (dashed line),.6 Mht (solid line) using a MGT with impeller a (grey line) or d (blak line). 10. Conlusion CFD was used to investigate ompressors with varying blade baksweep, β b, shaft speed, N, and blade height, b. Compressors had low diffusion ratios, DR, due to the rotational speed restritions on the target pressure ratio from seleting an oil

19 A Clay, GD Tansley 19 ooled journal bearing platform. Inreasing the rotational speed would be haraterised by a higher speifi speed with smaller impeller diameters introduing additional aerodynami losses and mehanial hallenges. Two impellers were hosen for off design performane investigation, small blade baksweep low speed; β b = 15, N = 170,000 rev/min, η = 71.76%, r =.0 vs. large baksweep higher speed; β b = 54, N = 00,000 rev/min, η = %, r =.11. The low speed, small baksweep impeller demonstrated superior performane at off-design. The advantages of reduing seondary flow losses by reating more uniform flow from inreased blade baksweep were not evident as little or no diffusion took plae within the impeller. Instead impeller performane improved by reduing disharge area to aelerate the flow through the impeller. Restriting diffusion is permitted on gas turbine ompressors when ombustion inlet veloity is low enough to prevent flame blow out. This may be ahieved on low pressure ratio impellers sine the diffusion duty on an impeller is proportional to the total stage pressure rise. Design point MGT performanes of η s = 14.85%, & net = 993 and η s = 14.6%, & net = 906 were established for the small and large blade baksweep impeller ompressors respetively. ith better off-design performane the small baksweep impeller demonstrated maximal savings. Maximal ost and CO savings were found with larger heat demands. For a speified heat demand, maximal ost savings are found with an optimum eletrial load whih does not oinide with maximal CO savings (whih inrease with redued eletrial load). However, the optimal annual eletrial demand for a building with an average thermal load of 17.4 Mht was around 4.0 Mhe (1.9% ost and 7.5% CO saving), approximately 1/3 less than the orresponding average of 6.1 Mhe (10.7% ost and 6.3% CO saving) suggesting efforts to redue ost by reduing eletriity onsumption will also redue CO. Generally speaking, when sized, maximum savings would be made by enouraging more eletriity export either by reduing eletriity onsumption or inreasing mahine effiieny whih would likely our from an inreased eletrial output from greater heat demand.

20 A Clay, GD Tansley 0 APPENDIX 1 Notation β Impeller tip blade angle [ ] b b Impeller tip blade height [mm] C g Gas tariff [ 0.01/kh] C e Eletriity tariff [ 0.01/kh] C ex Eletriity export tariff [ 0.01/kh] D h Hydrauli Diameter [m] DR Diffusion Ratio, 1 s Η Total enthalpy rise aross ompressor [kj/kg] η b Burner effiieny [%] η ( T ) 0 01 Compressor effiieny, s T ( T T ) η CB Condensing boiler effiieny 0 01 [%] η GEN DCHP Generator effiieny, η HEX Heat exhanger effetiveness, D net ( T05 T0 ) ( T T ) 04 0 [%] (MGT) [%] Q D (DCHP) [%] Q out η m Mehanial effiieny, in out [%] η s ηb Gas turbine thermal effiieny, Q η ( T ) 03 T04 t Turbine effiieny, ( T T ) 03 04s in net [%] [%] E g CO Gas emission fator [kg/kh] E e CO Eletriity emission fator [kg/kh] E ex CO Eletriity export emission fator [kg/kh] N Shaft speed [rev/min]

21 N s Speifi Speed, A Clay, GD Tansley 1 ω m& ρ H m& Mass flow [kg/s] 0avg 3 4 P Pressure drop ratio, Pha Pb 1 P0 P0 Phg 1 + P01 P 01 Compressor inlet total pressure [bar] P 0 Compressor exit total pressure [bar] P b P 0 P ha P 0 Burner pressure drop Heat exhanger air side pressure drop P hg P 01 Heat exhanger gas side pressure drop Q R Required thermal energy (by user) [k] Q D Delivered thermal energy (from DCHP) [k] Q in Gas turbine thermal power input [k] Q out Gas turbine thermal energy output, in net Q [k] r Pressure ratio, P P O O1 r ' Pressure ratio, P P O O1, for optimum system effiieny Re Reynolds number, Re = avg D h υ avg r Impeller exit radius [m] r 1 h Impeller inlet hub radius [m] r 1 s Impeller inlet shroud radius [m] T O1 Compressor inlet total temperature [K] T 1 Impeller inlet stati temperature [K] T O Compressor exit total temperature [K]

22 A Clay, GD Tansley T p Impeller exit stati temperature [K] T O3 Turbine inlet total temperature (TIT) [K] T O4 Turbine exit temperature [K] T O5 Reuperator inlet total temperature [K] U Impeller exit blade speed [m/s] υ avg Average dynami visosity aross impeller [m /s] ω Shaft speed [rad/s] ( ) Average relative impeller speed, + 1 avg [m/s] 1 Impeller inlet relative flow speed (mean) [m/s] 1 s Impeller inlet relative flow veloity (shroud) [m/s] Impeller tip relative flow veloity [m/s] net Net eletrial power from gas turbine, out in [k] R Required eletrial power (by user) [k] D Delivered eletrial power (from DCHP) [k] EX Exported eletrial power, D R [k] Aronyms BOP CFD DCHP MGT Mhe Mht PS SS Best Operating Point Computational Fluid Dynamis Domesti Combined Heat and Power Miro Gas Turbine Mega att hours eletrial Mega att hours thermal Pressure Side Sution Side

23 A Clay, GD Tansley 3 REFERENCES 1. Clay, A. and G.D. Tansley, An Analysis of Miro Gas Turbines for UK Domesti Combined Heat and Power. Power Engineer, (3).. hispergen. 007 [ited 010 5th Marh]; Available from: 3. Baxi. Baxi Eogen. 009 [ited 010 5th Marh]; Available from: 4. CarbonTrust. Miro CHP Aelerator Interim Report. The Carbon Trust Publiations 007 [ited 010 5th Marh]; Available from: 5. Peirs, J., aumans, J., Vleugels, P., Al-Bender, F., Stevens, T., Verstaete, T., Stevens, S., D'hulst, R., Vertraete, D., Fiorini, P., Van den Brambusshe, R., Driesen, J., Puers., Hendrik, P., Baelmans, M., Reynarts, D., Miropower generation with mirogasturbines: a hallenge. Proeedings of the Institution of Mehanial Engineers, Part C: Journal of Mehanial Engineering Siene, (4): p MiTi, Low-Frition, ear-resistant Korolon Coatings for High-Temperature, High-Speed, Air Foil Bearings. Mohawk Innovative Tehnology In. Development Newsletter, : p Pampreen, R., C., Small turbomahinery ompressor and fan aerodynamis. ASME Journal of Engineering Power, Hodge, J., Cyles and Performane Estimation. 1st ed. Gas turbine series; vol : Butterworth's Sientifi Publiations Japikse, D., Introdution to Turbomahinery. nd ed. 1997, Vermont & Oxford: Conepts ETI. In & Oxford Univerisity Press. 10. Shepherd, D.G., Priniples of Turbomahinery. 1956, New York: The Mamillan Company. 11. Flaxington, D. and E. Swain, Turboharger aerodynami design. Proeedings of the Institution of Mehanial Engineers, Part C: Journal of Mehanial Engineering Siene, : p Japikse, D., Centrifugal Compressor Design and Performane. 1996, Vermont: Conepts ETI, In. 13. hitfield, A. and N.C. Bains, Design of Radial Turbomahines. 1990, Harlow: Longman Sientifi & Tehnial Hildebrandt, A. and M. Genrup, Numerial Investigation of the Effet of Different Bak Sweep Angle and Exduer idth on the Impellor Outlet Flow Pattern of a Centrifugal Compressor ith Vaneless Diffuser. Journal of Turbomahinery, : p Van Den Braembusshe, R.A. Miro Gas Turbines A Short Survey of Design Problems. In Miro Gas Turbines, Eduational Notes RTO-EN-AVT-131. In Miro Gas Turbines ( ) 005 [ited 010 5th Marh]; Available from: EN-AVT Rodgers, C. Speifi speed and effiieny of entrifugal impellers. in Performane predition of entrifugal pumps and ompressors; Proeedings of the Twenty-fifth Annual International Gas Turbine Conferene and Exhibit and Twenty-seond Annual Fluids Engineering Conferene New Orleans: Amerian Soiety of Mehanial Engineers. 17. FLUENT, Ansys Fluent 1.0 User's Guide 009: ANSYS In. 18. Verstraete, T., Z. Alsalihi, and R.A. Van den Braembusshe, Numerial study of the heat transfer in miro gas turbines. Journal of turbomahinery, (4): p Boye, M.P., The gas turbine engineering handbook. 3rd rev. ed. ed. 006, Boston; Oxford: Gulf Publishing. 0. Flaxington, D. and B. Mahbod. Turboharger ompressor developmens for broad range and high pressure ratio appliations. in 4th International onferene on turboharging and turbomahinary 1990: IMECHE. 1. Tang, J., Computational analysis and optimization of real gas flow in small entrifugal ompressors, in Department of Energy and Environmental Tehnology. 006, Lappeenranta University of Tehnology: Lappeenranta. p Isomura, K., Murayama, M., Teramoto, S., Hikihi, K., Endo, Y., Togo, S., Tanaka, S., Experimental verifiation of the feasibility of a 100 lass miro-sale gas turbine at an impeller diameter of 10 mm. Journal of Miromehanis and Miroengineering, 006(9): p. S Saravanamutto, H., G.F.C. Rogers, and H. Cohen, Gas Turbine Theory. 5th ed. 001, Harlow: Pearson Eduation Limited. 491.

24 A Clay, GD Tansley 4 4. Casey, M.V., A omputational Geometry for the Blades and Internal Flow Channels of Centrifugal Compressors. Transations - Amerian Soiety of Mehanial Engineers Journal of Engineering for Power, GT-155: p ang,., R. Cai, and N. Zhang, General harateristis of single shaft miroturbine set at variable speed operation and its optimization. Applied Thermal Engineering, (13): p Newborough, M., Assessing the benefits of implementing miro-chp systems in the UK. Proeedings of the I MECH E Part A Journal of Power and Energy, : p Costyn, J. Miro CHP Poliy Overview. in CHPA Annual Conferene London, England: CHPA.