Turbocharger undamentals
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Turbocharger Fundamentals Hann !""s#el"inen$ %ag&i '. 'hair
Abstract( Abstract ( )r*ochargers are centri+gal compressors &riven *y an e,hast gas tr*ine an& employe& in engines to *oost the charge air pressre. )r*ocharger per+ormance in+lences all important engine parameters$ sch as +el economy$ po-er$ an& emissions. It is important to n&erstan& a nm*er o+ +n&amental concepts *e+ore moving on to a more &etaile& &iscssion o+ tr*ocharger speci+ics. )r*ocharger Constrction )r*ocharger Compressor asic /rinciples o+ Compression /rocess Compressor %aps )r*ocharger )r*ine )r*ine Energy E,traction )r*ine /er+ormance
1 Turbocharger !onstruction A turbocharger consists of a compressor wheel and exhaust gas turbine wheel coupled together by a solid shaft and that is used to boost the the intake air pressure of an internal combustion combustion engine. The exhaust gas turbine extracts energy from the exhaust gas and uses it to drive the compressor and overcome friction. In most automotive-type applications, both the compressor and turbine wheel are of the radial flow type. Some applications, such as medium- and low- speed diesel engines, can use an axial flow turbine wheel instead of a radial flow turbine. The flow of gases through a typical turbocharger with radial flow compressor and turbine wheels is shown in igure ! "Schwit#er !$$!%.
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Figure 1. 1. )r*ocharger constrction an& +lo- o+ gases orce( ch-iter3
Center-Housing. The turbine-compressor common shaft is supported by a bearing system in the
center housing /bearing housing0 located between the compressor and turbine /igure (0. The shaft wheel assembly /S1A0 assembly /S1A0 refers to the shaft with the compressor and turbine wheels attached, i.e., the rotating assembly. The center housing rotating assembly /234A0 assembly /234A0 refers to S1A installed in the center-housing but without the compressor and turbine housings. The center housing is commonly cast from gray cast iron but aluminum can also be used in some applications. Seals help keep oil from passing through to the compressor and turbine. Turbochargers for high exhaust gas temperature applications, such a spark ignition engines, can also incorporate cooling passages in the center housing.
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Figure 1. 1. )r*ocharger constrction an& +lo- o+ gases orce( ch-iter3
Center-Housing. The turbine-compressor common shaft is supported by a bearing system in the
center housing /bearing housing0 located between the compressor and turbine /igure (0. The shaft wheel assembly /S1A0 assembly /S1A0 refers to the shaft with the compressor and turbine wheels attached, i.e., the rotating assembly. The center housing rotating assembly /234A0 assembly /234A0 refers to S1A installed in the center-housing but without the compressor and turbine housings. The center housing is commonly cast from gray cast iron but aluminum can also be used in some applications. Seals help keep oil from passing through to the compressor and turbine. Turbochargers for high exhaust gas temperature applications, such a spark ignition engines, can also incorporate cooling passages in the center housing.
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Figure ". ". ectional vie- o+ tr*ocharger ectional vie- o+ an e,hast gas tr*ocharger +or a gasoline engine sho-ing compressor -heel le+t3 an& tr*ine -heel right3. )he *earing system consists o+ a thrst *earing an& t-o +lly +loating 4ornal *earings. Note the cooling passages. orce( org5arner3
Bearings. The turbocharger bearing system appears simple in design but it plays a key role in a
number of critical functions. Some of the more important ones include the control of radial and axial motion of the shaft and wheels and the minimi#ation of friction losses in the bearing system. 5earing systems have systems have received considerable attention because of their influence on turbocharger friction and its impact on engine fuel efficiency. 1ith the exception of some large turbochargers for low-speed engines, the bearings that support the shaft are usually located between the wheels in an overhung position. This flexible rotor design ensures that the turbocharger will operate above its first, and possibly second, critical speeds and can therefore be sub6ect to rotor dynamic con ditions such as whirl and synchronous vibration. Seals. Seals are located at both ends of o f the bearing housing. These seals represent a difficult
design problem due to the need to keep frictional losses low, the relatively large movements of the shaft due to bearing clearance and adverse pressure gradients under some conditions. These seals primarily serve to keep intake air and exhaust gas out of the center housing. The pressures in the intake and exhaust systems systems are normally higher than in the turbocharger7s center housing which is typically at the pressure of the engine crankcase. As such, they would primarily be designed to seal the center housing when when the pressure in the center housing is lower than in the intake and exhaust exhau st systems. These seals are not intended to be the primary means of preventing oil from escaping from the center housing into the exhaust and air systems. 8il is is usually prevented from contacting these seals by other means such as oil deflectors and rotating flingers.
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Turbocharger seals are different from the soft lip seals normally found in rotating e9uipment operating at much lower speeds and temperatures. The piston ring type seal is one type that is often used. It consists of a metal ring, similar in appearance to a piston ring. The seal remains stationary when the shaft rotates. :abyrinth-type seals are another type sometimes used. ;enerally turbocharger shaft seals will not prevent oil leakage if the pressure differential reverses such that the pressure in the center housing is higher than in the intake or exhaust systems.
" Turbocharger !om#ressor A radial flow compressor stage is composed of two sections, the impeller or turn thus changing its flow from an axial to a radial direction. Air exits the compressor wheel at the exducer /(0, enters a narrow stationary diffuser and then passes through to a volute or scroll from which it is discharged from the compressor.
Volute Trailing edge 2"
2'
Diffuser
Shroud edge Hub edge Leading edge
2
Impeller 1
d1
d2
Figure $ Cross section o+ ra&ial +lo- compressor
Compressor Wheel. The critical components of the compressor wheel are the blades. These
blades have three regions a. the leading edge is a sharp pitch helix designed for scooping air in an d moving it axially? b. blades are curved to change the direction of the airflow from axial to radial and at the same time to accelerate it to a high velocity?
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c. blades terminate in a trailing edge which is designed to propel air radially out of the compressor wheeldefined by pitch, angle offset from radial andor back taper or back sweep. The shape of the compressor blade profile from the impeller inlet to the outlet can significantly impact performance and durability. ;enerating an appropriate shape can be a compromise between manufacturing cost and compressor performance. A significant development in manufacturing of compressor impellers was the extension of flank milling /milling with the side of the cutter0 to the production of arbitrary surfaces "1u !$*+%. Breviously, flank milling was limited to the production of ruled surfacessurfaces that were defined by a limited number of curves and arbitrary surfaces re9uired the use of point milling /milling with the tip of the cutter0 a very time consuming and costly process for volume production. Arbitrary surfaces allow the blade profile to be tailored more precisely to the demands of the flow to provide better performance and durability rather than the need to limit the surface to a few well defined mathematical curves as ruled surfaces do "3olset ())@%. The compressor wheel blade trailing edge or blade tip can either extend radially from the center of the wheel7s hub /as do the two impellers in igure $ be low0 or have a backward curvature, igure *. orward curved blades are also possible but are rarely used in turbochargers. 4adial impeller blades were very common in the past due to ease of manufacture and strength considerationsthe blade tip was relatively unaffected by centrifugal forces resulting from high rotational speeds. 3owever, efficiency is lower with radial blades and modern turbochargers will typically use a backward curved impeller to maximi#e compressor efficiency. 3owever, backward curvature can be limited by manufacturing considerations when impellers are produced by casting and by strength considerations. The backward curvature means that blade tips can be sub6ected to significant bending forces at high rotational speeds with the bending stress increasing with backward sweep angle. Selection of blade tip or trailing edge angle is a trade-off between p erformance and manufacturing considerations. igure @ shows some o f the performance trade-offs. The backward curved blade has the highest efficiency but tends to provide the lowest pressure ratio for a given tip speed that drops 9uickly "3anlon ())!%. 5ackward curved impeller blades re9uire higher tip speeds compared to radial blades to maintain a given pressure ratio.
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Backward
Radial
Forward
Flow Rate
Figure 4. Effect of blade tip angle on compressor performance
Turbocharger compressors often benefit from a higher number of blades to provide good flow guidance along the impeller. However, if full length blades are used, the optimum number of blades can lead to blockage of flow at the impeller inlet. To avoid this problem, splitter blades that start part way through the impeller can be used, Figure 5.
Figure 5. Compressor impeller with full length and splitter blades (Source: BorgWarner)
Blades have extremely close tolerances with the housing to minimie backflow. !roduction of cast impellers and housings to tight tolerances can be a significant challenge. "n some cases, the stationary housing can be coated with an abradable coating that is softer than the blade material. #pon initial start$up, the coating is worn away enough so that the impeller blade can pass freely and clearances are kept to a minimum. This can provide a significant benefit in turbocharger
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efficiency as evidenced by the improvement in engine fuel economy below +')) rpm shown in Figure & and by lower transient particulate emissions /not shown0 1%harp +2223.
Figure 6. Influence of turbocharger abradable coating on fuel economy
4ompressor wheel attachment to the shaft can be via a through$hole /through$bore wheel0 or a blind hole /boreless wheel0. "n boreless designs, attachment of the wheel to the shaft can be via a threaded connection or welding. "n thrubore designs, a nut at the end of the shaft is commonly used.
Figure 7. hrubore and boreless compressor wheels
Boreless compressor wheels can eliminate the area of high stress at the interface between the wheel and shaft bore at the axial location where wheel diameter is largest. However, the axial length of the wheel increases due to the need to attach the shaft to the wheel while avoiding shaft protrusion beyond the $plane /Figure 0. This can increase the footprint of the turbocharger and, because of the longer distance from the bearing support to the wheel6s center
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of mass, make the rotor dynamic effects more challenging. 7lso, alignment of the shaft and impeller as well as rotating assembly balancing can become more difficult and expensive. Diffuser. The
diffuser can be vaneless as shown in Figure ' or vaned, Figure * 1Heywood +2**3.
8arly compressor diffusers used a series of divergent noles to convert kinetic energy to pressure via volume expansion.
Figure 8. Segment of a radial compressor and !aned diffuser plate
9anes shorten the flow path through the diffuser, reducing frictional losses and controlling the radial velocity component of the gas. :ue to lower friction, head an d efficiency are enhanced. However, off$design operation rapidly changes the incidence angle to the v anes and flow separation occurs, resulting in a reduced operating range 1Hanlon ())+3. 9aned diffusers are common in marine and generator set applications where a high pressure ratio is often re;uired and a wide flow range is unnecessary. 9aneless diffusers are used when some efficiency can be sacrificed to reduce manufacturing costs, when sie is not a criterion, or if a wide range of flow rates at a given pressure ratio is re;uired. 9aneless diffusers are used in most automotive type applications. 7 vaneless diffuser provides a significantly lower$pressure recovery compared with a vaned diffuser of the same diameter. The diffuser walls can also be designed with a nonlinear area increase with radial distance. "n any design it is important to ensure smooth internal surfaces in the diffuser to reduce frictional losses 1%iuru ())'3. Volute. 7
spiral$shaped volute or scroll collects the flow from the diffuser and passes it to the
compressor outlet. Figure 2 shows two volutes and their corresponding compressor wheels. The cross$sectional area increase of volute from a, to b, to c helps convert kinetic ene rgy into potential energy, or air velocity into pressure 1%chwiter +22+3. Together with increased pressure, the compressed air experiences an increase in its temperatureote the different connection to the piping at the exit from the v olutes.
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Figure 9. wo !olute co!er designs for radial compressor
7n important ob?ective of volute design is to achieve a uniform flow at the volute exit. This is usually attained at the design flow$rate only, so that at off$design conditions the volute is either too small or too large and a pressure distortion develops circumferentially around the volute. 7t low flow$rates the pressure increases with aimuth angle, while at high flow$rates the pressure decreases. These circumferential pressure distortions can be transmitted back to the impeller discharge and even as far back as the impeller inlet. The pressure distortions reduce the compressor6s performance and have a direct impact on diffuser and impeller flow stability 14arter ())23.
3. Basic Principles of Compression Process Before discussing various performance aspects of turbocharger co mpressors, it is important to understand a few basic principles. The ideal compression process can be considered to be isentropic, Figure +). For an ideal gas, a good approximation for a turbocharger compressor, the change in enthalpy is a function only of temperature so this Figure +@lso reflects the enthalpy change. :ue to various inefficiencies, the actual process consumes more energy than the ideal.
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Figure 10. Ideal and actual compression process
"sentropic efficiency of a compressor can be defined as
ηisentropic = hisentropic/(hin-hout)
(1)
"nefficiency will show up as a higher than ideal compressor outlet temperature for a given pressure ratio, Figure ++. Turbocharger efficiency can have a direct impact on the capacity of the intercooler with less efficient compressors re;uiring higher intercooling capacity to maintain a given intake manifold temperature.
Figure 11. Effect of compressor efficiency on outlet t emperature
The energy behind the pressure increase in a ce ntrifugal compressor is supplied by the impeller via acceleration of the flow to very high velocities. Figure +( clarifies various components of these velocities for backward, radial and forward curved impeller blades 1Aolloch ())53.
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Figure 12. Velocity triangles for backward, radial and forward curved impeller blades
In Figure 12, C1 and C2 are the absolute velocity vectors (i.e., relative to the stationary frame of the compressor) of the air entering and exiting the impeller. 1 and 2 are the impeller blade tip speed vectors at these same locations. !1 and !2 are velocity vectors of the air flo" entering and leaving the impeller relative to the impeller#
W=C-U
(2)
$ressure generation inside the compressor consists of several steps# (1) %inetic energy is first supplied to the air by means of the rotor "hich acce lerates the air to a high speed, (2) as the air passes bet"een the blades of the impeller, the cross§ional flo" area bet"een the blades increases "hich causes some of the %inetic energy imparted to the flo" by the impeller to be converted into static pressure "ithin the impeller even "hile additional %inetic energy is being added, (') after the air exits the impeller, no more energy is transferred into the air flo". It is no" the role of the diffuser and volute to conve rt the remaining %inetic energy into static pressure as efficiently as possible. his process is further illustrated in Figure 1' "here the subscripts 1, 2 and 2 correspond to the impeller inlet, diffuser inlet and volute inlet as sho"n in Figure ' *+er%er 212-*anlon 21-.
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Figure 13. Centrifugal compression process in coordinates of specific enthalpy (h) and specific entropy (s)
he total enthalpy change through the compressor can be described by the 5uler e6uation#
∆ht = 1/2 [(C2 2 - C1 2) + (-W 2 2 + W1 2 ) + (U 2 2 - U1 2 )]
(3)
It can be sho"n that using 56uation (2), the p ressure increase across the impeller can be "ritten as
P 2 - P1 = [G/2gc] · (C2TU2 - C1T U1 )
(4)
"here C2 and C1 are the tangential components of C2 and C1 , Figure 12. Figure 17 sho"s ho" pressure ratio can vary "ith impeller tip speed, 2 *89: 2;-#
Figure 14. Effect of impeller tip speed on pressure ratio
9 parameter that is common in fan engineering and occasionally used for turbocharger compressors *5ngels 22- is the dimensionless pressure number , <. :ne "ay to "rite this is#
Ψ = 2∆ht/U 2 2
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he value of the pressure number relates the total enthalpy change in the compressor to the tip speed of the compressor impeller. =ifferent values of < reflect different compressor design philosophies and a trade&off can exist bet"een < and turbocharger efficiency.
4. Compressor Maps 9 common tool used in selecting a turbocharger compressor for a given application is a compressor map. 9 compressor map provides the operating envelope of the compressor over all stable operating points. For turbocharger compressors, the x&axis is normally in terms of flo" "hile the y&axis is in terms of pressure. In order to construct such a map, it is necessary to first determine the pressure characteristics of the compressor over a range of compressor speeds and flo"s. hese characteristics are then plotted on a single graph to create the compressor map. 9s sho"n in 56uation (7), the pressure developed by a radial flo" centrifugal compressor is proportional to the product of C2 and 2 . sing this relationship, 56uation 1 and Figure 12, one can deduce the shape of the ideal flo" versus pressure curves for various blade tip angles. In an impeller "ith radial blade tips, as flo" decreases (i.e., !2 decreases) at a constant impeller speed, the magnitude of C2 decreases and C2 is unaffected ma%ing the ideal pressure ratio versus flo" characteristic flat, Figure 10b. !ith bac%"ard curved blades, the magnitude of C2 increases and C2 also increases ma%ing the ideal pressure ratio highest at lo" flo"s Figure 10c *anlon 21-.
(a) Forward
(b) Radial
(c) Backward
Ideal
Flow Rate
Flow Rate
Flow Rate
Figure 15. General shape of pressure ratio versus flow characteristics (a) forward, (b) radial, (c) backward curved impeller blades
In addition to pressure ratio, flo" range is another important consideration that needs to be ta%en into consideration "hen generating a compressor map. he lo"est stable flo" possible from a centrifugal compressor at a particular rotational speed is determined by the surge limit "hile the highest flo" by choking . 9s flo" is reduced at a constant impeller speed, !2 decreases proportionately, causing the flo" angle >2 to decrease. In a vaneless diffuser, this creates a longer flo" path through the diffuser "hich increases frictional losses to the diffuser "alls. 9lso, on the inlet side of the impeller, !1
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decrease causing the flo" angle >1 to decrease. 9t lo" enough flo"s, the lo" angle of >1 can cause flo" separation to occur on the lo" pressure side of the b lade at the impeller inlet that can further reduce flo" into the impeller. his combined "ith the high frictional losses in the diffuser can lead to a situation "here the pressure generation in the diffuser falls belo" the delivery pressure and the flo" in the compressor can reverse. his reverse flo" continues until the resistance through diffuser drops enough for air discharge to be continued. his process is repeated in a cyclic fashion and is referred to as surge. he point on the pressure versus flo" curve at a given impeller speed "here surge occu rs is referred to as the surge limit . he surge limit can also be thought of as the maximum of the p ressure versus flo" curve at a given impeller speed. $rolonged operation during compressor surge can damage the compressor. ?anes in the diffuser can shorten the flo" path through the diffuser to reduce frictional losses and control the radial velocity component "ith the net result that pressure generation and efficiency are improved. o"ever, off&design operation changes the incidence angle to the vanes and flo" separation can occur, leading to a reduced operating range for the compressor. 9t the other end of the pressure versus flo" curve, as flo" increases, a point can be reached "here ! becomes too high for the flo" to remain attached to the impeller blades and separation can occur "ithin the passages bet"een the impellers blades. his separation reduces the effective flo" area and eventually a point is reached "here the +ach number reaches 1 and flo" through the impeller becomes choked . he point on the pressure versus flo" curve at a given impeller speed "here cho%ing occurs is referred to as the choke limit . In practice, it is common to define the cho%e limit in terms of compressor efficiency. For example, one manufacturer defines the cho%e limit is the point on the pressure versus pressure curve "here compressor efficiency drops belo" 0@. sing the above information, it is no" possible to understand a compressor map, an example of "hich is sho"n in Figure 1. Commonly, contours of constant compressor efficiency are included on the compressor map. In addition to the surge and cho%e limits, compressor maximum speed is fixedAusually by material limitations. hese three curves define the usable envelope of pressure ratio and flo" that the compressor can achieve.
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5.2 #aximum speed
116.0
4.6
0.68
Constant impeller speed, x1000 rpm
0.68 0.7 0.72 0.73 0.74 0.65 0.75
4.0 101.9
Surge line
3.4
0.76 0.77
2.8 Flo range
!.6
2.2 6!." 0.68
1.6
0.6
0.65
C!o"e limit
"#.
1.0 0
5
10
15
20
25
30
35
40
45
Corrected Mass Flow, kg/min Figure 16. Eample compressor map !org"arner E#$ %&' turbocharger
Corrected Parameters. In
order to account for differences in ambient conditions, the
compressor map can be expressed in terms of variables that have been corrected to standard temperature and pressure. Corrected mass flo" is normally expressed as
C!""#ct#$ %&'' !* = ct,& %&'' !* (.)/0
(6)
"here
= T&ct,&/T"##"#c#
(7)
0 = P&ct,&/P"##"#c#
(8)
his "as done for the compressor map of Figure 1. /ome manu facturers use corrected volume flo" instead#
C!""#ct#$ !,%# !* = ct,& !,%# !* / (.)
(9)
/everal other corrected parameters may also appear on compressor maps including#
C!""#ct#$ '##$ = ct,& '##$ / (.)
(1)
C!""#ct#$ t!",# = ct,& t!",# / 0
(11)
C!""#ct#$ !*#" = ct,& !*#" / (0.)
(12)
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The reference temperature and pressure used by different manufacturers is not en tirely consistent and the user should be careful to check what these conditions are. A compressor map can also be expressed in nondimensional variables. A dimensional analysis of a centrifugal compressor reveals that
(13)
For high Reynolds number, the Reynolds number effect tends to be weak. For a fixed geometry compressor and constant gas properties, this can be simplified to
(14)
The variable groupings in this simplification are no longer non-dimensional. n this representation, the compressor speed parameter could have the dimension of rps (revolutions
per second) / √K , while the mass flow could be expressed in units of (kg/s) √K / MPa. Figure !" shows a compressor map expressed in such simplified variable groupings derived from the dimensional analysis. Figure !" also plots compressor efficiency versus the flow parameter for a series of compressor speeds# a representation that is sometimes used. t should be noted that variables and units used in compressor maps from different manufacturers can vary.
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Figure 17. Compressor map plotted using simplified variable groupings from dimensional analysis (Source: Holset Turbochargers)
A parameter that is sometimes referred to in relation to compressors is the structural trim. For a compressor see Figure & for variable definitions/+
Trimstructural = (d1/d2)2
(15)
As can be seen in Figure !*, decreasing trim shifts the compressor map to the left. t is also important to note that a trim decrease lowers the flow rate at choking conditions for a given impeller speed.
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Figure 18. Effect of compressor trim on compressor map
01uation !'/ applies to compressor wheels where the trailing edge is parallel to the axis of rotation. n cases where it is not, the aerodynamic trim can be used 23oust ()!'4+
Trimaero = (d1/d2,R!) 2
(16)
And where d(,R:$ is defined as+
d 2,R! = √[(d2 2 + d 2,ti" 2)/2]
(17)
where+ d( 5 diameter of the wheel at root end of the exducer i.e., where the hub and trailing edges meet/ d(,tip 5 diameter of the shroud edge of the wheel at the exducer i.e., where the shroud and trailing edges meet/
The ratio of the area of the volute inlet to its distance from the compressor shaft, the so called AR ratio see Figure &"/ can also be referred to in relation to compressors. 6ompressor performance is relatively insensitive to AR ratio. 7irectionally, larger AR housings are used for low boost pressure and smaller AR housings for high boost pressure applications. The AR ratio is much more critical with respect to turbines see discussion later/. 8ther parameters that can be used to discuss compressor characteristics include the annulus
area and the vaneless diffuser annulus area. 7efinitions for these can be found elsewhere 23oust ()!'4.
Figure !9 illustrates how the airflow re1uirements at constant engine speed and load would appear on a typical compressor map.
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Figure 19. Compressor map overlayed with constant load and constant speed airflow characteristic from a !stro"e diesel engine
Figure () overlays the full load compressor performance curves for several different applications from the !9*)s on a compressor map with a dimensionless flow parameter 2;osch !9*%4. t should be noted that turbocharger control has changed c onsiderably and the
characteristics for more modern applications would be considerably different, especially for the truck and passenger car applications.
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4.0
3.5
Ship
3.0 Surge limit
2.5
Truck
2.0
u [m/s] 450 80
1.5
78
75
Passenger car 70 < [%]
300
150
1.0 0
0.1
0.2
0.3
0.4
0.5
Volumetric Flow Rate Parameter ɸ = 4V/D2a Figure 20. Compressor map showing e#ample engine performance lines
Figure (! shows the range of the compressor map that would be utili=ed between no load and full load for and older application where boost pressure is limited>possibly with a wastegate 23enein !9*'4. At very low speeds and loads, the compressor would produce no boost.
Figure 21. $erformance characteristics of turbocharger compressor
5. Turbocharger Turbine 5.1 Types of Turbine Geometry There are three types of turbines suitable for exhaust gas turbochargers+ radial-flow, axial-flow and mixed-flow, Figure (( 2:athis ())&4. The mixed flow turbine has characteristics between the
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radial and axial turbines. The high-pressure flow enters the turbine at the inlet (3”) and is guided to the stator inlet (3’) where vanes turn the flow in a direction tangential to the rotor. The flow leaves the stator vanes and enters the rotor blades (3), which turn the flow back in the opposite direction, extracting energ fro! the flow. The flow leaving the rotor blades ("), now at a lower pressure, passes through a diffuser where a controlled increase in flow area converts dna!ic head to static pressure. #fter the diffuser, the flow exits to the discharge conditions.
Figure 22. Three possible exhaust gas turbine geometries
The purpose of the inlet is to guide the flow fro! the suppl to the stator vanes with a !ini!u! loss in total pressure. # radial turbine will include a volute or scroll to help achieve this ob$ective. The stator or no%%le induces a swirl co!ponent to the flow so that a tor&ue can be i!parted to the rotor blades. 'tators can be e&uipped with nu!erous curved airfoils called vanes that turn the flow in the tangential direction. an radial-inflow turbines for turbochargers often have no vanes in the stator. To achieve good efficienc over a wide range of inlet flow conditions, variable-geo!etr stators can be used, tpicall with either pivoting stator vanes (igure *3) or a variable width no%%le (igure *") +arter */.
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Figure 23. Swing Vane Variable Geometry Turbine
Figure 24. Moving nozzle or moving wall type variable geometry turbine
The rotor extracts energ fro! the flow, converting it to shaft power. The rotor blades are attached to a rotating disk that transfers the tor&ue of the rotor blades to the turbine output shaft. 7ike the stator, the rotor has a nu!ber of individual curved airfoils called rotor or turbine blades. The blades are angled to accept the flow fro! the stator with !ini!u! disturbance when the turbine is operating at design conditions. The flow fro! the stator is then turned back in the opposite direction in a controlled !anner, causing a c hange in tangential !o!entu! and a force to be exerted on the blades. The flow leaving the turbine rotor can have a significant a!ount of kinetic energ. 8f this kinetic energ is converted to static pressure in an efficient !anner, the turbine can be operated with a rotor discharge static pressure lower than the static pressure at diffuser discharge. This increases the turbine power output for a given inlet and discharge conditions. Radial Flow Turbines. The
choice of turbine tpe depends on a variet of factors including
turbine efficienc, packaging re&uire!ents and !anufacturing cost. 9enerall, for engine applications of about 0 hp5turbocharger or less (i.e., turbine wheel dia!eter of about 03 !! or less), cost and si!plicit are critical factors and radial turbines are tpicall the !ost attractive option. 8n these applications, !axi!u! exhaust te!peratures tend to be higher and radial-flow turbines are the better choice. o!pared to the axial-flow turbines, there is a !uch larger difference between the rotor inlet relative and absolute velocities for the radial-flow turbine that results in a lower relative inlet total te!perature at the design point. #lso, due to the decrease in rotor speed with radius, the relative total te!perature decreases toward the root of
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the radial-flow turbine blades. This is a !a$or advantage for applications where turbine rotational speed can potentiall be li!ited b te!perature-dependent !aterial properties. :ue to bending stress considerations in the rotor blades, a radial-flow turbine would use a radial blade tip profile at the inlet to eli!inate bending loads. This co!bined with the decreased te!perature in the high-stress blade root areas allows the radial-inflow turbine to operate a t significantl higher wheel speeds than an axial-flow turbine to provide an appreciable increase in turbine efficienc for high-pressure-ratio applications +athis *3/. 8n this si%e range, radial turbines also offer better perfor!ance because of tip clearance considerations and there is no need for a separate no%%le ring which si!plifies !anufacturing and reduces cost when co!pared to axial turbines. Axial Flow Turbines. #xial
turbines, igure *1, are tpicall used in applications that exceed
about 0 hp5turbocharger (i.e., turbine wheel dia!eter of !ore than about 03 !!) where the offer an efficienc and si%e advantage over radial flow turbines. ;xa!ples of applications that can use an axial flow turbine include !ediu! and low speed diesel engines. or !an of these large engines, !axi!u! turbine inlet te!peratures and pressure ratios are lower than for s!aller high speed engines and the blade speed o f an axial wheel is not constrained b stress considerations. The radial-flow turbine would be at a si%e disadvantage. The use of radial blade tips on a radial flow turbine to eli!inate bending stress li!its the ratio of the tangential co!ponent of the absolute inlet velocit to blade tip speed (T 5<) to 0 or less while an axial flow turbine can have T 5< = 0 with onl a s!all i!pact on efficienc. This would result in a larger and heavier radial-inflow turbine than axial flow turbine when a given power and rotational speed is re&uired +athis *3/.
Figure 25. Turbocharger Featuring Radial Flow ompressor and !xial Flow Turbine M!" #iesel $ Turbo Series T!
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#xial turbines tpicall need a relativel large exhaust diffuser and collector for !axi!u! efficienc which can !ake the! !ore difficult to package for a single turbocharger installation where space is li!ited. >owever, if two turbines are used in a series arrange!ent such as for a turboco!pound or dual turbocharger installation, an axial turbine can bring both !anufacturing and packaging benefits. ?hen the exhaust fro! a radial turbine is directed straight into an axial turbine, a considerabl si!plified piping arrange!ent can be used, igure * +9res%ler *2/ +9obert *@/.
Figure 26. Turbo%compound utilizing axial &low turbine connected to a turbocharger with a radial &low turbine
#xial turbine !anufacture poses particular challenges. or large axial turbines, individual blades are often !ounted on the hub in what is known as the Afir tree’ root arrange!ent. >owever, such a techni&ue would be too costl for s!all wheels where one-piece, blade, disk and hub arrange!ents (Bblisk”) are used. >owever, this re&uires co!plex tooling and li!its the choice of blade shapes !ore so than with radial turbines. 7arge axial turbines also co!!onl use a Ada!ping wire’ to da!pen blade vibration. To incorporate this wire, a hole in each blade at about 1-@1C span is re&uired which would be i!practical on a s!all axial wheel being produced in high volu!es. Dotential vibration issues therefore have to be tackled through basic blade design that leads to co!pro!ises in perfor!ance. Mixed Flow Turbines. hoosing
between radial-flow and !ixed-flow is less straightforward.
The original !otivation for !ixed flow turbine wheels for turbochargers was the reduction of turbine !ass and inertia to i!prove engine transient response. # !ixed-flow turbine wheel can handle a higher gas flow than a radial turbine of e&uivalent dia!eter. 7ater it was discovered that !ixed-flow turbines can also allow both i!proved aerodna!ic blade loadingEleading to greater efficiencEand the abilit to Aad$ust’ the turbine’s peak efficienc point to a higher or
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lower engine speed that can be useful in !axi!i%ing pulse energ fro! the exhaust. >owever, the !erits of a !ixed-flow turbine !ust be weighed against the !ore co!plex tooling re&uired to produce it. urther!ore, the turbine housing re&uires !ore space which could be a n issue in a vehicle’s congested engine co!part!ent +>olset *"a/+7Fddecke *0*/ . #udi’s 3rd generation of T'8 engines (;#222) are an exa!ple of a !odern application that uses !ixed flow turbines. The 0.2 7 engine uses an 8>8 !ixed flow turbine while the *. 7 (9eneration 3G) uses a ontinental H##I !ixed flow turbine, igure *@.
Figure 27. ontinental's R!!( )R!dial !(ial* turbine &or !udi's +., - Generation / TFS0 engine
5.2 Constant Pressure and Pulse Turbines The volu!e of the piping between the exhaust !anifold and the turbine housing plas a ver i!portant role in the dna!ic operation of the turbocharger. G enlarging this volu!e as well as that of the !anifold, pulsations fro! the exhaust valves discharging into the !anifold are da!pened and the pressure at the turbine inlet is !uch !ore constant. This approach is referred to as constant pressure turbocharging . ?hile this !ethod ensures the continuous flow of exhaust gas over the turbine blades, it does not !axi!i%e the use of exhaust energ in the turbine because the pressure peaks are da!ped out. 8t can also result in higher exhaust back pressure which can reduce engine volu!etric efficienc. # second approach !akes use of the pulsating nature of exhaust e!anating fro! each of the exhaust valves and is called pulse turbocharging . # co!!on design for in-line engines fro! " to clinders couples the front clinders in a single !anifold that leads to one side of the connecting flange, and the rear clinders in another single !anifold leading to the other side of the sa!e connecting flange, igure *2a +Jbert 02/. This !ethod is preferred for applications where engine response is i!portant (!ost on-highwa applications), since exhaust energ utili%ation is !axi!i%ed and exhaust back pressure is lower.
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Figure 28. Pulse conservation and pulse converter systems
In modern engine designs utilizing pulse turbocharging and where one turbocharger is served by 4 or 6 cylinders, it is common to use a twin scroll turbocharger, Figure 30a. An eample !or a 4 cylinder engine is shown in Figure "#. $he ehaust inlet is split into two parts where one part carries ehaust !rom cylinders one and !our, while the other !rom cylinders two and three. %ulse turbocharging also prevents ehaust pulses eiting the cylinder !rom one cylinder !orcing ehaust gas bac& into another cylinder when eh aust valve open events overlap '(rosse "000). (reater part load e!!iciencies are possible and more air is available at !ull load. Fuel consumption reductions o! *+ to "0 are possible.
Figure 29. Twin scroll type turbocharger
Figure 30. Twin scroll turbine (a) and dual volute turbine (b) (Source: BorgWarner)
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haust gas recirculation in a pulse turbocharging system can be a cha llenge and pulse energy can be optimized i! ( is recirculated to the inta&e mani!old !rom only one side o! the two1 piece ehaust mani!old. $his would create a signi!icant di!!erence in the ehaust !lows reaching the turbine !rom the two ehaust mani!olds. 5ne solution is to use an asymmetrical twin scroll turbine, Figure 3*. $his approach is used by aimler in its engines intended !or 7- %A "0*0 and uro 8I applications.
Figure 31. Asymmetrical twin scroll turbine
5ne advantage o! the dual volute turbine is that it is much easier to combine the bene!its o! a variable geometry nozzle into a pulse turbine, Figure 3" '-auerstein "00#). 9hile a twin scroll turbine could in principle also incorporate a variable geometry nozzle, implementation is much more challenging. 5ne option that has been proposed is to use a moving divider wall, the wall separating the two scrolls 'Anschel "0**).
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Figure 32. ual volute turbine with variable geometry no!!le (Source: BorgWarner)
$he third and last method o! connecting the e haust mani!old to the turbine housing is called the pulse converter. Figure "b describes one embodiment o! the pulse converter method showing a venturi bo that converts ehaust &inetic energy to potential energy or p ressure, which is similar to that in the constant pressure system. In the past, constant pressure systems were common in two1stro&e engine applications and pulse conservation systems in !our1stro&e on1highway applications '5bert *#6).
6. Turbine Energy Extraction (as !low through a turbine is considered to be an adiabatic epansion process. Ideally, this process is isentropic as shown in the $1s diagram o! Figure 33. It should be noted that as with a compressor, assuming that the ehaust is an ideal gas is reasonable. For an ideal gas, en thalpy, h, is only a !unction o! temperature, h;$<, and the temperature di!!erences in Figure 33 also re!lect enthalpy di!!erences. $he actual energy trans!er in a turbine is smaller than the isentropic value due to irreversibilities in the !low. $he actual process is mar&ed by an increase in entropy and is also represented in Figure 33. $he enthalpy change associated with an entropy increase is less than that !or an isentropic process. $he ratio o! the actual to isentropic enthalpy change re!lects the degree o! entropy increase and is re!erred to as the isentropic ;sometimes adiabatic< efficiency/
ηisentropic = (hin - hout) / hisentropic
(18)
It can also be shown that the energy trans!er to the turbine wheel can be epressed as
∆hactual = U3 C T3 - U4 C T4
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(19)
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Figure 33. "deal and actual e#pansion process across a turbine
Losses. $he
di!!erence between the ideal and actual enthalpy change across the turbine is made
up o! losses !rom various sources. $hese sources include the inlet, stator, rotor, di!!user and eit '=athis "003).
Inlet losses primarily arise !rom !rictional and turning e!!ects. $he inlet should be made as large as the pac&aging restraints allow> reducing velocities and minimizing losses. Aial inlets are short and have relatively low velocities and as such have low !rictional losses but o!ten su!!er !rom turning losses due to !low separation along their outer diameter. ?onger aial inlets with more gradual changes in outer diameter tend to reduce the turning losses and prevent separation, but adversely impact turbine envelope. $angential entry inlets such as those !ound in radial !low turbines tend to have higher losses due to the tangential turning and acceleration o! the !low. $he spiral !low path also tends to be longer, increasing !rictional losses. -tator losses arise primarily !rom !riction within the vanes, secondary !lows caused by the !low turning, and eit losses due to bloc&age at the vane trailing edges. otor losses are similar to those !or stators but with a !ew additional components to account !or tip clearance and windage losses. $urbine rotors operate with a small clearance between the tips o! the blades and the turbine housing. Flow lea& s across this gap !rom the high1pressure side o! the blade to the low1pressure side, causing a reduction in the pressure di!!erence at the tip o ! the blade. $his reduces the tangential !orce on the blade, decreasing the tor@ue delivered to the sha!t. 9indage losses arise !rom the drag o! the turbine dis&. ?osses in the di!!user arise !rom sources similar to those in other !low passages, namely, !riction and !low turning. As !or eit losses, in a single stage turbocharger turbine they e@ual the eit &inetic energy o ! the !low unless used in an additional turbine stage or in some other manner. I! the &inetic energy o! the !low eiting the di!!user is used in later stages, the eit losses are zero.
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Dimensionless Parameters. $o
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ma&e comparisons between di!!erent turbines easier, a number
o! variable groupings arising !rom dimensional analysis are sometimes used '=athis "003). Specific Speed. $he speci!ic speed o! a turbine is de!ined a s
(20)
where " is the volumetric !low rate through the turbine at rotor eit conditions. $he speci!ic speed is used to relate the per!ormance o! geometrically similar turbines o! di!!erent size. In general, turbine e!!iciency !or two geometrically similar turbines at the same speci!ic speed will be approimately e@ual. Bowever, small di!!erences can eist due to clearance and eynolds number e!!ects. =aintaining speci!ic speed o! a turbine is a common approach to scaling a turbine to di!!erent !low rates. Specific Diameter. $he speci!ic diameter is de!ined as
(21)
where dtip is the tip diameter o! either a radial in1!low or aial !low turbine rotor. -peci!ic diameter and speci!ic speed are used to correlate turbine per!ormance. Blade-Jet Speed Ratio. $urbine per!ormance can also be correlated against the blade1Cet speed ratio, which is a measure o! the blade speed relative to the ideal stator eit velocity. $he ideal stator eit velocity, D0, is calculated assuming the entire ideal enthalpy drop across the turbine is converted into &inetic energy/
(22) $he blade1Cet speed ratio is then calculated !rom
(23)
$he value o! blade speed at the mean turbine blade radius is typically used !or aial turbines while !or radial1in!low turbines, the rotor tip speed is used.
7. Turbine Performance Figure 34 shows a typical per!ormance map !or a radial !low turbine. $he p ressure ratio represents the ratio o! inlet to outlet pressure. 5ne important !eature to note is that mass !low typically reaches a limiting value that is uni@ue to each rotational speed. $his is a result o! cho&ing. $he cho&ed !low limit increases as turbine speed decreases.
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Figure 34. Typical radial-flow turbine map
Rarely is an entire map such as that in Figure 34 shown. Rather, turbine performance around peak turbine efficiency is more commonly found, Figure 35. When the pressure ratio versus mass flow characteristics at a number of turbine speeds are plotted in this way, they often form a curve that is almost continuous. line drawn through these curves at the peak efficiency for each turbine speed is sometimes referred to as the turbine swallowing capacity and represents the desired operating curve of the turbine. !n order to better match the turbine swallowing capacity to the flow from the engine, it is important to select a suitable "R ratio for the turbine #see below$. Further ad%ustments can be made by controlling the swallowing efficiency with a wastegate or variable geometry turbine.
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Figure 35. Turbine swallowing capacity (lower curve) and efficiency (upper curve)
Figure 35 also shows measured turbine efficiency #the product of mechanical and isentropic efficiency$ at a number of different turbine speeds as a function of turbine pressure ratio. nother way to present turbine efficiency is to plot it against blade speed ratio, Figure 3. (he peak turbine efficiency is typically found near a blade speed ratio of about *.0 for radial flow turbines and around *.4 for a1ial flow turbines.
Figure 36. Turbine isentropic efficiency as a function of blade speed ratio
A/R Ratio. n
important parameter to match the swallowing capacity of the turbine to the flow
from the engine is the "R ratio. !t is defined as the crosssectional area of the inlet of the scroll over the distance of the centroid of that area to the center of the turbine shaft, Figure 30.
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R A
Cross-sectional area
Figure 37. Illustration and definition of A/ R ratio
Figure 3+ illustrates the effect of the "R ratio on the swallowing capacity of a radial flow turbine. small "R ratio increases the tangential velocity at the turbine wheel tip for a given e1haust flow and thus provides higher turbine and compressor rotational speeds. (his can improve air flow at low engine speeds that can be used to enhance low speed engine tor2ueespecially if tor2ue is smoke limited. 25 1.44 1.28 1.15
20
1.01
A/R 15
10 1.0
1.5
2.0
2.5
3.0
Pressure Ratio
Figure 38. Effect of A/R Ratio on wallowing !apacity of an E"#aust $as Turbine $arret $T%&'R turboc#arger wit# & turbine trim
Figures 3+ shows the effect of "R ratio on smoke emissions at engine speeds below peak tor2ue speed. (he turbocharger with the smaller "R ratio helped reduce smoke emissions at these two speeds by a substantial margin. (urbocharger lag is also reduced and thus e ngine transient response improved with a smaller "R ratio. owever, a small "R ratio turbine will choke at lower flow rates. (o ma1imie engine power, a larger "R ratio would be more beneficial.
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Figure 39. Effect of turbine A/R ratio on smoe from medium-duty *I diesel engine
For turbines, the trim is also an important parameter that influences flow capacity. larger trim will handle a higher flow rate and result in less backpressure but will recover less e1haust energy and increase turbocharger lag.
References nschel, 6., R. 7handramohanan, )*''. 8&implified 9ariable :eometry (urbocharger With 9ariable ;ole<, =& 6atent pplication )*''"*)3))+) ', http-""www.google.com"patents"=&)*''*)3))+) >osch, '?+. 8utomotive andbook<, &ociety of utomotive @ngineers, Warrendale, 6, )nd @dition 7arter, A., et al., )**?. 8(urbocharging technologies for heavyduty diesel engines<, !n- Bdvanced direct in%ection combustion engine technologies and development- Ciesel enginesB #9olume )$, @d. . Dhao, Woodhead 6ublishing, :reat bington, =E @ngels, >., )**). 8ifetime prediction for (urbocharger 7ompressor Wheels Why =se (itaniumG<, >orgWarner Enowledgeibrary, http-""www.3kwarner.de"tools"download.asp1GtHdocumentIrH'*0IdH'*? :obert, =., et al., )**0. 8(urbo compressor system for internal combustion engine comprising two serially placed turbo units with their rotation a1es essentially concentric<, =& 6atent 0,)+0,30? #9olvo astvagnar >$, http-""www.google.com"patents"=&0)+030? :olloch, R., )**5. 8Cownsiing bei 9erbrennungsmotoren<, &pringer :resler, ., )**+. 8Ciesel (urbocompound (echnology<, !77(";@&77F Workshop, !mproving the Fuel @conomy of eavyCuty Fleets !!, February )*, )**+, http-""www.nescaum.org"documents"improvingthefueleconomyofheavy dutyfleets'"gresler/volvo/session3.pdf :rosse, A., )***. 8;ew @ngines For Renault in Juest For :reen Kachine<, F( utomotive World, Aanuary )*** anlon, 6.7. #ed.$, )**'. 87ompressor andbook<, Kc:rawill enein, ;.., C.A. 6atterson, '?+5. 87ombustion @ngine @conomy 6erformance and @missions<, ecture ;otes to Ford (ractor Lperations, Lctober ', '?+5 eywood, A.>., '?++. 8!nternal 7ombustion @ngine Fundamentals<, Kc:rawill, ;ew Mork olset, )**4. 8Kachinedfromsolid compressor wheels<, (i Kagaine, ), 3 olset, )**4a. 87hoosing the right turbine for the automotive turbocharger<, (i Kagaine, ), 45 oust, 9., et al., )*'5. 8Functionally asymmetric twosided turbocharger wheel and diffuser<, =& 6atent application )*'5"**3?+* ' #oneywell$, http-""google.com"patents"=&)*'5**3?+* Nddecke, >., et al., )*'). 8Ln Ki1ed Flow (urbines for utomotive (urbocharger pplications<, !nternational Aournal of Rotating Kachinery, )*'), rticle !C 5+?0)*, doi-'*.''55")*')"5+?0)*, http-""downloads.hindawi.com"%ournals"i%rm")*')"5+?0)*.pdf Kathis, C.K., )**3. 8Fundamentals of (urbine Cesign<, !n- andbook of (urbomachinery, &econd @dition, @d. @. ogan, Ar. and R. Roy, 7R7 6ress, doi-'*.')*'"?0+*)*3?''??*.ch0
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