A comparative study of the performance of the mixed flow and radial flow variable geometry turbines for an automotive turbocharger

2018 ◽  
Vol 233 (8) ◽  
pp. 2696-2712 ◽  
Author(s):  
K Ramesh ◽  
BVSSS Prasad ◽  
K Sridhara
Keyword(s):  
Mass Flow ◽  
Flow Parameter ◽  
Radial Flow ◽  
Velocity Ratio ◽  
Variable Geometry ◽  
Flow Variable ◽  
Mixed Flow ◽  
Turbine Efficiency ◽  
Variable Geometry Turbine ◽  
Specific Speed

A new design of a mixed flow variable geometry turbine is developed for the turbocharger used in diesel engines having the cylinder capacity from 1.0 to 1.5 L. An equivalent size radial flow variable geometry turbine is considered as the reference for the purpose of bench-marking. For both the radial and mixed flow turbines, turbocharger components are manufactured and a test rig is developed with them to carry out performance analysis. Steady-state turbine experiments are conducted with various openings of the nozzle vanes, turbine speeds, and expansion ratios. Typical performance parameters like turbine mass flow parameter, combined turbine efficiency, velocity ratio, and specific speed are compared for both mixed flow variable geometry turbine and radial flow variable geometry turbine. The typical value of combined turbine efficiency (defined as the product of isentropic efficiency and the mechanical efficiency) of the mixed flow variable geometry turbine is found to be about 25% higher than the radial flow variable geometry turbine at the same mass flow parameter of 1425 kg/s √K/bar m2 at an expansion ratio of 1.5. The velocity ratios at which the maximum combined turbine efficiency occurs are 0.78 and 0.825 for the mixed flow variable geometry turbine and radial flow variable geometry turbine, respectively. The values of turbine specific speed for the mixed flow variable geometry turbine and radial flow variable geometry turbine respectively are 0.88 and 0.73.

Transient Response of Mixed Flow Variable Geometry Turbine for a Turbocharger

2015 ◽  
Author(s):  
Ramesh Kannan ◽  
Bhamidi Prasad ◽  
Sridhara Koppa
Keyword(s):  
Steady State ◽  
Transient Response ◽  
Velocity Ratio ◽  
Typical Result ◽  
Variable Geometry ◽  
Flow Variable ◽  
Data Logging ◽  
Mixed Flow ◽  
Response Behaviour ◽  
Variable Geometry Turbine

A specific design of mixed flow variable geometry turbine for an automotive sub 1.5 litre diesel engine turbocharger is proposed in this paper. An experimental set up is developed for measuring the steady state and transient response behaviour of the turbine at different nozzle vane opening positions. The rotor speed, pressure and temperature before and after the turbine are measured and recorded using high frequency data logging system. The steady state performance for mass flow, efficiency, velocity ratio, specific speed and the transient response behaviour of the mixed flow variable geometry turbine (MFVGT) are compared against the same parameters of a radial flow variable geometry turbine (RFVGT) of similar dimensions. Typical result indicates that the transient response of the MFVGT is faster by about 350 milliseconds than the radial at turbine inlet pressure of 0.2 bar (g).

Transient performance of the mixed flow and radial flow variable geometry turbines for an automotive turbocharger

2020 ◽  
Vol 234 (19) ◽  
pp. 3762-3775
Author(s):  
Ramesh Kannan ◽  
BVSSS Prasad ◽  
Sridhara Koppa
Keyword(s):  
Transient Response ◽  
Mass Flow ◽  
Radial Flow ◽  
Flow Variable ◽  
Mixed Flow ◽  
Flow Parameters ◽  
Actuation System ◽  
Turbine Inlet ◽  
Set Up ◽  
Swept Volume

In our previous paper, the steady-state test results of a mixed flow turbine with variable nozzle vanes for a turbocharger are reported. In this paper, the transient response of the same mixed flow turbine along with that of a similarly sized radial flow turbine is presented. The turbine size is suitable for handling the flow capacity of the diesel engines with swept volume up to 1.5 L. The previous experimental test set up is modified by adding a quick-release valve – actuation system before the turbine inlet to obtain a transient response. The radial and mixed flow turbines are tested for different turbine inlet pressures and for various opening positions of the nozzle vanes while matching the turbine mass flow parameters between radial and mixed flow turbines. Typically at nozzle vane openings corresponding to 50% mass flow parameter and 1.5 bar (abs) pressure at the inlet to the turbine, the transient response time for the turbine with mixed flow variable nozzle vanes configuration is about 0.770 s, as compared to 0.858 s for the turbine with radial flow variable nozzle vanes configuration.

scholarly journals Performance and Exit Flow Characteristics of Mixed-Flow Turbines

1997 ◽  
Vol 3 (4) ◽  
pp. 277-293 ◽  
Author(s):  
C. Arcoumanis ◽  
R. F. Martinez-Botas ◽  
J. M. Nouri ◽  
C. C. Su
Keyword(s):  
Steady State ◽  
Velocity Ratio ◽  
Tangential Velocity ◽  
Flow Characteristics ◽  
Mixed Flow ◽  
Cross Sectional ◽  
Flow Angle ◽  
Turbine Efficiency ◽  
And Performance ◽  
Inlet Angle

The performance and exit flow characteristics of two mixed-flow turbines have been investigated under steady-state conditions. The two rotors differ mainly in their inlet angle geometry, one has a nominal constant incidence (rotor B) and the other has a constant blade angle (rotor C), but also in the number of blades. The results showed that the overall peak efficiency of rotor C is higher than that of rotor B. Two different volutes were also used for the tests, differing in their cross-sectional area, which confirm that the new larger area volute turbine has a higher efficiency than the old one, particularly at lower speeds, and a fairly uniform variation with velocity ratio.The flow exiting the blades has been quantified by laser Doppler velocimetry. A difference in the exit flow velocity for rotors B and C with the new volute was observed which is expected given their variation in geometry and performance. The tangential velocities near the shroud resemble a forced vortex flow structure, while a uniform tangential velocity component was measured near the hub. The exit flow angles for both rotor cases decreased rapidly from the shroud to a minimum value in the annular core region before increasing gradually towards the hub. In addition, the exit flow angles with both rotors were reduced with increasing rotational speeds. The magnitude of the absolute flow angle was reduced in the case of rotor C, which may explain the improved steady state performance with this rotor. The results also revealed a correlation between the exit flow angle and the performance of the turbines; a reduction in flow angle resulted in an increase in the overall turbine efficiency.

A physics-based turbocharger model for automotive diesel engine control applications

2018 ◽  
Vol 233 (7) ◽  
pp. 1667-1686 ◽  
Author(s):  
Kang Song ◽  
Devesh Upadhyay ◽  
Hui Xie
Keyword(s):  
Mass Flow Rate ◽  
Mass Flow ◽  
Flow Rate ◽  
Turbine Rotor ◽  
Variable Geometry ◽  
Gas Velocity ◽  
Euler’S Equations ◽  
Variable Geometry Turbine ◽  
Euler's Equations ◽  
Compressor Power

Control-oriented models of turbocharger processes such as the compressor mass flow rate, the compressor power, and the variable geometry turbine power are presented. In a departure from approaches that rely on ad hoc empirical relationships and/or supplier provided performance maps, models based on turbomachinery physics and known geometries are attempted. The compressor power model is developed using Euler’s equations of turbomachinery, where the gas velocity exiting the rotor is estimated from an empirically identified correlation for the ratio between the radial and tangential components of the gas velocity. The compressor mass flow rate is modeled based on mass conservation, by approximating the compressor as an adiabatic converging-diverging nozzle with compressible fluid driven by external work input from the compressor wheel. The variable geometry turbine power is developed with Euler’s equations, where the turbine exit swirl and the gas acceleration in the vaneless space are neglected. The gas flow direction into the turbine rotor is assumed to align with the orientation of the variable geometry turbine vane. The gas exit velocity is calculated, similar to the compressor, based on an empirical model for the ratio between the turbine rotor inlet and exit velocities. A power loss model is also proposed that allows proper accounting of power transfer between the turbine and compressor. Model validation against experimental data is presented.

Case Study of Multi-Stage Impeller with Guidevane

2005 ◽  
Author(s):  
Takuji Tsugawa
Keyword(s):  
Velocity Ratio ◽  
Optimum Shape ◽  
Hydraulic Efficiency ◽  
Mixed Flow ◽  
Shape Factors ◽  
Flow Angle ◽  
Multi Stage ◽  
Absolute Flow ◽  
Specific Speed

In the previous paper, the optimum meridian profile of impeller was obtained for various specific speed by means of eight shape factors, that is, inlet relative flow angle β1, turning angle Δβ, axial velocity ratio kc = Cm2/Cm1, impeller diameter ratio kd = D1c/D2c, outlet hub-tip ratio ν2, tip solidity σtimp, mid span solidity σcimp and hub solidity σhimp. In this paper, the optimum meridian profile of multi-stage impeller with guidevane was obtained by means of twelve shape factors. The additional four shape factors are guidevane tip solidity σtgv, mid span solidity σcgv, hub solidity σhgv and coefficient of peripheral velocity at impeller inlet or guidevane outlet kCu1c. In the optimum method, the hydraulic efficiency and suction specific speed are calculated by diffusion factor. In the optimum condition, the best hydraulic efficiency or the best suction specific speed is obtained. In the cyclic flow condition of multi-stage impeller with guidevane, the absolute flow velocity of guidevane outlet is equal to that of impeller inlet, and the diameter of guidevane outlet is equal to that of impeller inlet. In this calculation, the diameter of impeller outlet is equal to that of guidevane inlet. The total calculation number of case study is very large, so the number of each parameter is about between four and seven. The best 1000 optimum meridian profiles and the best design parameter are selected for few kinds of specific speed using twelve dimensional optimum method. As the result of this calculation, the optimum meridian profile of multi-stage impeller and guidevane. The more detailed optimum multi-stage mixed flow impeller and guidevane profile is drawn. For, example, the 1000 specific speed is selected for case study of multi-stage mixed flow impeller. At first, the approximate optimum shape factors are present shape factors. And the optimum shape factors which have better efficiency are tried to find near the present shape factors. Then the study of shape factor changes is the objective of this paper.

Case Study of Impeller Profile Considering Additional Parameters

Volume 3 ◽  
2004 ◽  
Author(s):  
Takuji Tsugawa
Keyword(s):  
Velocity Ratio ◽  
Stream Line ◽  
Mixed Flow ◽  
Shape Factors ◽  
Flow Angle ◽  
Turning Angle ◽  
Calculation Time ◽  
Specific Speed ◽  
Stream Lines

In the previous paper, the optimum meridian profile of impeller was obtained for various specific speed by means of five shape factors. In this paper, the optimum meridian profile of impeller is obtained by means of eight shape factors. The basic five shape factors are inlet relative flow angle β1, turning angle Δβ, axial velocity ratio kc = Cm2/Cm1 impeller diameter ratio kd = D1c/D2c and outlet hub-tip ratio ν2 (β1 and Δβ are in mid span stream surface). The additional three parameters are three stream lines solidity (tip solidity σt, mid span solidity σc, and hub solidity σh). The blade length of impeller meridian profile is able to obtain by additional three parameters. The method of optimization is the calculation of hydraulic efficiency and suction specific speed in all combinations of eight shape parameters. The number of five shape factors are expressed by Nβ1, NΔβ, Nkc, Nkd, Nν2. The number of calculations is expressed by Nβ1 × NΔβ × Nkc × Nkd × Nν2. For example, Nβ1 = NΔβ = Nkc = Nkd = Nν2 = 40, the number of calculations is about 100000000. The calculation time is about 2 hours. The best parameters are selected in 100000000 cases. In case of eight shape factors, the number of calculation is Nβ1 × NΔβ × Nkc × Nkd × Nν2 × Nσt × Nσc × Nσh. Nβ1 = NΔβ = Nkc = Nkd = Nν2 = Nσt = Nσc = Nσh = 10, the number of calculation is 100000000. In this case, the calculation time of eight shape factors is as same as that of five shape factors. By means of this method, the more detailed optimum mixed flow impeller meridian shape is obtained. In case study, the best 1000 optimum meridian profiles and the best design parameter are selected for few kinds of specific speed using eight dimensional optimum method. In the previous paper, the mixed flow angle on tip meridian stream line isn’t able to be decided by this optimization using diffusion factor. But, in this paper, the mixed flow angle is able to be decided by the number of blade and optimum solidity. As the best solidity of three stream lines is obtained, the axial coordinates of impeller inlet and outlet are obtained. The more detailed optimum mixed flow impeller meridian shape is drawn.

Lean and Straight Nozzle Vanes in a Variable Geometry Turbine: A Steady and Pulsating Flow Investigation

2008 ◽  
Author(s):  
Srithar Rajoo ◽  
R. F. Martinez-Botas
Keyword(s):  
Pressure Ratio ◽  
Velocity Ratio ◽  
Pulsating Flow ◽  
Maximum Pressure ◽  
Variable Geometry ◽  
Flow Investigation ◽  
Variable Geometry Turbine ◽  
The Difference ◽  
Vane Angle ◽  
Ratio Range

Variable Geometry Turbines (VGT) are widely used to improve engine-turbocharger matching and currently common in diesel engines. VGT has proven to provide air boost for wide engine speed range as well as reduce turbo-lag. This paper presents the design and experimental evaluation of a variable geometry mixed flow turbocharger turbine. The tests have been carried out with a permanent magnet eddy current dynamometer within a velocity ratio range of 0.47 to 1.09. The peak efficiency of the variable geometry turbine corresponds to vane angle settings between 60° and 65°, for both the lean and straight vanes in the region of 80%. The variable geometry turbine was tested under pulsating flow with straight and lean nozzle vanes for different vane angle settings, 40Hz and 60Hz flow. In general, the range of mass flow parameter is higher in the straight nozzle vanes with an average of 66.4% and 69.7% for 40Hz and 60Hz flow respectively. The straight nozzle vanes also shows increasing pressure ratio range compared to the lean nozzle vanes, which is more apparent in the maximum pressure ratio experienced by the turbine in an unsteady cycle. In overall, the cycle averaged efficiency in the straight vane configuration is marginally higher than the lean vane. Furthermore, the difference to the equivalent quasi-steady is better in the straight vane configuration compared to the lean vane.

Case Study of High Specific Speed Radial Impeller

2002 ◽  
Author(s):  
Takuji Tsugawa
Keyword(s):  
Radial Flow ◽  
Velocity Ratio ◽  
Axial Flow ◽  
Optimum Shape ◽  
Design Parameters ◽  
Shape Factors ◽  
Flow Angle ◽  
Diffusion Factor ◽  
Cavitation Performance ◽  
Specific Speed

The optimum shape of high specific speed impeller is usually axial flow impeller. The radial impeller is often used without axial flow guidevane. Usually, the radial impeller is the high pressure and low specific speed impeller. The design parameters of radial high specific speed impeller have not been obtained yet. In the previous papers, the optimum meridian shape of axial flow impeller with axial flow guidevane is obtained for various specific speed. The optimum meridian shapes calculated by diffusion factor agree with meridian shapes of conventional impellers. In this paper, the design parameters of radial high specific speed impellers without guidevane are calculated by diffusion factor. And the optimum meridian shapes of radial high specific speed impellers are proposed. In case of the radial impeller, the hub diameter is equal to the tip diameter in impeller outlet. So, in radial impellers, the outlet hub-tip ratio is 1.0. The optimum meridian shapes of radial impellers for various specific speed are also obtained in this paper. The relative efficiency and cavitation performance of impellers in various shape factors were calculated. The calculation of radial meridian shape needs four kinds of shape factors as the previous papers. The four shape factors are inlet relative flow angle β1, turning angle Δβ, axial velocity ratio (meridian velocity ratio) kc = Cm2/Cm1 and impeller diameter ratio kd = D1c/D2c inmid span streamsurface. In initial step of impeller design, the result of the efficiency and cavitation performance of impeller calculated in optimum principal design parameters is important. The principal design parameters are hub-tip ratio, inlet-outlet diameter ratio, axial velocity ratio, solidity, inlet flow angle, turning angle and blade number. The author proposed the optimum meridian profile design method by diffusion factor for various condition of design parameters. There is a good correlation between the optimum hub-tip ratio and the specific speed considering cavitation performance. The optimum solidity is obtained for the specific speed considering efficiency and cavitation performance. It was found that the optimum meridian profile of high specific speed impeller with appropriate efficiency and cavitation performance has large inclination on hub and tip stream lines. The calculated data base is four dimensional using four various shape parameter β1, Δβ, kc and kd. Using the four shape factor, the optimum meridian shape of radial flow impeller is able to be obtained. The best 1000 optimum design parameters are selected using four dimensional calculated data. The aspect of optimization is recognized with 1000 plotted data on 6 planes. The result of radial flow impeller optimization is different from that of axial flow impeller. In case of axial flow impellers, the shape factors are optimized for each specific speed. But, in radial flow impellers, if both the specific speed and the total head coefficient are given, the optimum shape factors are optimized. The calculation results between profiles and specifications were very useful for the development of new type high specific speed radial impellers.

Case Study of Improved High Specific Speed Radial Impeller

2003 ◽  
Author(s):  
Takuji Tsugawa
Keyword(s):  
Design Method ◽  
Velocity Ratio ◽  
Flow Analysis ◽  
Design Parameters ◽  
Mixed Flow ◽  
Shape Factors ◽  
Flow Angle ◽  
Diffusion Factor ◽  
Cavitation Performance ◽  
Specific Speed

It is usually thought that the axial impeller is used for high specific speed impeller and the radial impeller is used for low specific speed impeller. In the previous paper, the optimum meridian profile of axial impeller and radial impeller were obtained for various specific speed by means of the optimization of four shape factors using diffusion factor. The four shape factors were inlet relative flow angle β1, turning angle Δβ, axial velocity ratio (meridian velocity ratio) kc = Cm2/Cm1 and impeller diameter ratio kd = D1c/D2c in mid span stream surface. In case of axial impeller, the optimum meridian profiles agreed with meridian profiles of conventional impellers. To develop the radial high specific speed impeller, the optimum four shape factors of radial high specific speed impellers were calculated by diffusion factor. And the optimum meridian profiles of radial high specific speed impellers were proposed. In case of the radial impeller, the hub diameter is equal to the tip diameter in impeller outlet (the outlet hub-tip ratio is 1.0). And in axial impeller, the outlet blade height depends on the outlet hub-tip ratio. On the other hand, in mixed flow impeller, the outlet hub-tip ratio is various and the outlet blade height is independent of the outlet hub-tip ratio. To obtain the optimum meridian profile of mixed flow impeller, the hub-tip ratio of impeller outlet ν2 is adopted new additional independent shape factor for optimization in this paper. The mixed flow angle on tip meridian stream line (= 0 degree in axial impeller, = 90 degrees in radial impeller) isn’t able to be decided by this optimization using diffusion factor. But, the mixed flow angle will be decided by the number of blade and solidity. And, it will be decided by meridian velocity distribution from hub to tip for each specific speed of impeller. So, in this paper the five shape factors are used for optimization by diffusion factor. (β1, Δβ, kc, kd, ν2) The optimum meridian profiles of mixed flow impellers for various specific speed are obtained. The relative efficiency or the cavitation performance of mixed flow impeller is better than that of radial or axial impeller. In this optimum method, the relative efficiency and the cavitation performance are calculated for all specified combinations of five shape factors. The number of five shape factors are expressed by Nβ1, NΔβ, Nkc, Nkd and Nν2. The number of calculations is expressed by Nβ1 × NΔβ × Nkc × Nkd × Nν2. The calculation time of five shape factors method is Nν2 times the calculation time of four shape factors method. Then, the best 1000 combinations of five shape factors are plotted on β1 - Δβ, kc - kd and kd - ν2 plane. The aspect of the best 1000 optimum conditions are found by these three figures. In initial step of impeller design, the result of the efficiency and cavitation performance of impeller calculated in optimum principal design parameters is important. The principal design parameters are hub-tip ratio, inlet-outlet diameter ratio, axial velocity ratio, solidity, inlet flow angle, turning angle and blade number. The author proposed the optimum meridian profile design method by diffusion factor for various condition of design parameters. There is a good correlation between the optimum hub-tip ratio and the specific speed considering cavitation performance. The optimum solidity is obtained for the specific speed considering efficiency and cavitation performance. It was found that the optimum meridian profile of high specific speed impeller with appropriate efficiency and cavitation performance had large inclination on hub and tip stream lines. The calculated data base is five dimensional using five shape factors β1, Δβ, kc, kd and ν2. Using the five shape factors in case of the best efficiency, the optimum meridian profile of improved radial flow impeller is able to be calculated. At first step of the case study, the best 1000 optimum meridian profiles and the best design parameter are selected using five dimensional optimum method. Next, the blade section shape of impeller is decided by the blade or cascade design method. Using impeller flow analysis, the cavitation performance decided by 3% head reduction is calculated. Finally, the relations between the many type of meridian profile and its impeller performance by flow analysis are obtained. These relations are very useful for new type of high specific speed impeller design. Consequently, radial impellers and axial impellers are improved by the consideration of the additional shape factor, that is, outlet hub-tip ratio ν2. This calculation shows that the improved radial high specific speed impeller considering outlet hub-tip ratio is used for high suction specific speed and high efficiency.

Mean Line Flow Model of Steady and Pulsating Flow of a Mixed-Flow Turbine Turbocharger

2010 ◽  
Author(s):  
Aman M. I. Bin Mamat ◽  
Ricardo F. Martinez-Botas
Keyword(s):  
Experimental Data ◽  
Steady Flow ◽  
Mass Flow ◽  
Flow Condition ◽  
Flow Parameter ◽  
Pulsating Flow ◽  
Mixed Flow ◽  
One Dimensional ◽  
Turbine Performance ◽  
Relative Standard

A one-dimensional investigation for a mixed-flow turbine turbocharger turbine is described in this paper. The main outcome of the research is to develop a validated procedure for turbine performance maps in both steady and pulsating flow. The approach is limited to a simple procedure that can be integrated in wave action codes and thus not requiring 2D or 3D calculations. The mass flow parameter map is used as an input for the investigation, thus requiring some knowledge of either from experiments or from mean line methods calculations. In this paper, the mass flow parameter is experimentally measured at the turbocharger facility in Imperial College. A realistic yet reduced-order model for turbine losses allows the prediction of the steady flow performance; the calibration of the model is performed at peak efficiency of the turbine for a given rotational speed. The loss model for steady flow is then extended for pulsating conditions and for presumed quasi-steady operation. Finally, the predicted turbine performance is compared with experimental data. The comparison between one-dimensional modeling and experimental data for steady flow condition has shown Relative Standard Deviation (RSD) range from 1.62% to 11.62%. Meanwhile, a good trend agreement has been achieved for pulsating flow condition.

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