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Injection performance and cavitation analysis of an advanced 250 MPa common rail diesel injector
International Journal of Heat and Mass Transfer 93 (2016) 388–397
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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Injection performance and cavitation analysis of an advanced 250 MPa common rail diesel injector Lian Duan a, Shou-qi Yuan a, Lin-feng Hu b, Wen-ming Yang c,⇑, Jian-da Yu b, Xing-lan Xia b a
Research Centre of Fluid Machinery Engineering & Technology, Jiangsu University, Zhenjiang 212013, China FAWÒ Wuxi Fuel Injection Equipment Research Institute, Wuxi 214063, China c Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore 117575, Singapore b
a r t i c l e
i n f o
Article history: Received 21 April 2015 Received in revised form 12 October 2015 Accepted 13 October 2015
Keywords: Common rail injection system Injection responsiveness Backflow fuel Multi-injection Valve cavitation
a b s t r a c t The present work is to demonstrate the working process and injection performance of a new-type diesel injector, and to make a contrastive analysis of cavitation which potentially occurs near control-valves of the new and another conventional injector. The advanced injector is developed for 250 MPa common rail injection systems of medium or heavy duty commercial vehicles. The originality is the development of a novel control-piston with the outward appearance of a poppet valve. The control-piston can automatically move downwards or upwards owing to pressure changes in related hydraulic chambers, and consequently appropriately activates or deactivates the function of additional inflowing control-orifices to improve system responsiveness. The injection performance demonstration was carried out by one-dimensional (1D) hydraulic mechanical simulation in combination with the experimental data input of actuators. On the other hand the cavitation evaluations, in both the conventional and new structures, were studied numerically by computational fluid dynamics (CFD) based on nonlinear cavitation model and dynamic grid method. Additionally the CFD model was validated against experimental result in terms of mass flow rate of backflow fuel in the conventional structure. The 1D simulation result shows that the new injector is capable of achieving multi-injection and ramp-shaped injection flow rate; closing-delay of injection can be reduced from 1.232 to 0.758 ms, and backflow fuel quantity index can be decreased to 14.1:182.6 which means a reduction of 15.5%. As regards to the contrastive analysis of cavitation, the CFD simulation result of the new structure shows several improvements as follows: a lower velocity field (below 311.88 m/s) near the control-valve, cavitation no longer occurring in the guide-hole, and a reduction of 70.22% in mass exchange rate between vapour–liquid phases near the sealing annulus of valve seat. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction To meet progressively strict emission regulations [1,2], performance and fuel-economy, diesel injection equipment [3] is expected to be capable of higher system pressure (SP), injection precision, robustness and rapid injection responsiveness. Common rail injection system (CRS) is the preferred diesel injection equipment due to its excellent flexibility. Increased SP has plenty of influence on injection performance, such as acquiring finer spray droplets, enhancing turbulent mixing rate to attain better fuel–air mixture [4], shortening ignition delay time [5,6], reducing soot emission [7,8] and improving specific power [9]. Currently as high as 250 MPa SP is important to obtain adequate spray characteristics for achieving stricter emission regu⇑ Corresponding author. Tel.: +65 6516 6481. E-mail address: [email protected] (W.-m. Yang). http://dx.doi.org/10.1016/j.ijheatmasstransfer.2015.10.028 0017-9310/Ó 2015 Elsevier Ltd. All rights reserved.
lations Euro-VI and more satisfactory engine performance. In such a high-pressure operating condition, several problems exist in common rail injector (CRI), such as insufficiently rapid responsiveness in closing fuel injection, a large quantity of backflow fuel, and worse cavitation damage (viz., cavitation erosion) on control-valve. Flow-induced cavitation frequently occurs in CRI owing to high pressure-drop or/and geometric change of flow path [10]. Cavitation can change flow-characteristic and damage wall surface, especially in the vicinity of control-valve. Thus it results in a deviation of fuel injection characteristic and a risk of fuel injection malfunction due to control-valve leakage. Therefore, it is necessary to evaluate the level of cavitation near control-valve in CRI’s. However, the flow condition near control-valve (sub-millimetre-scale, submillisecond-timescale, more than 100 MPa pressure difference and hundreds m/s flow velocity) is so complicate that experimental visualisation method is unfeasible, or the costs involved is prohibitive. Thereby numerical simulation method appears to be
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Nomenclature D25 OA OZ
code name of the new-type diesel injector outflowing control-orifice inflowing control-orifice
Abbreviations 1D one dimensional
a proper approach for disclosing the flow feature. Bianchi et al. [11] studied the cavitation influence on discharge coefficient in controlorifices and control-valve with STAR-CDÒ, nevertheless, the model was steady-state and laminar flow. Xia et al. [12] investigated the cavitation flow of the region by linear cavitation model with AVLFIREÒ, but correlative experimental validation was not presented. At present, several types of 250 MPa-level CRI products have been launched by manufacturers for diesel engines of commercial vehicles, such as BOSCHÒ-CRIN4.2 [13], CUMMINSÒ-XPI [14], DENSOÒ-G4S [15]. They have adopted different structural designs as a solution for aforementioned problems, but their sophisticated structures require advanced manufacturing technique, and their improved results in closing-delay are less than satisfactory. In order to obtain a better solution to aforementioned problems, a new type of diesel injector (code-named D25) is developed for 250 MPa CRS of medium or heavy duty commercial vehicles. In this paper, the structure and working process of D25 is described, and its injection performance is demonstrated by 1D hydraulic mechanical simulation in combination with the experimental data input of actuators. Moreover, in order to evaluate the effect of the new structure on cavitation in the neighbourhood of control-valve, a CFD simulation analysis is carried out contrastively in two types of structures (the new one and another conventional one). The CFD simulation is focused on three-dimensional unsteady two-phase flow, and is based on nonlinear cavitation model as well as dynamic grid method with AVL-FIREÒ v.2010. Additionally the CFD simulation is validated against experimental result in terms of the mass flow rate of backflow fuel in the conventional structure CRI.
BC CFD CRI CRS ET SP
boundary condition computational fluid dynamics common rail injector common rail injection system energising time system pressure
12.0), and the applications of the software for CRS modelling have been done by other authors [17–20]. With the software, the 1D model and validation of a conventional CRI have been performed in a previous work [21]. The trial-manufacture of D25 is still in preparation. Therefore injection performance experiment is not yet available for the validation of 1D model of D25; presented 1D simulation result serves solely for functionality demonstration. Nevertheless a part of
2. Development of new structure The contrast of hydraulic concepts between conventional CRI and D25 is illustrated in Fig. 1. There have been seven improvements in D25. r Clearance leakage is avoided; s the diameters of inflowing control-orifice (commonly known as OZ) and outflowing control-orifice (commonly known as OA) are decreased by about 46% and 41% respectively; t additional OZ and a controlpiston which has the outward appearance of a poppet valve are designed; u control-chamber is positioned directly above the nozzle needle instead of above a command-piston; v a fuel storage chamber (similarly introduced in [16]) is implanted; w three grooves with circular-segment shaped cross-section (Fig. 2), which are machined regularly around the nozzle needle, are adopted as fuel delivery channels in the nozzle; x an intermediate chamber is put between OA and control-valve. Based on approximate installing condition of an engine example (Table 1), the structure of D25 is shown in Fig. 3.
Fig. 1. Contrast of hydraulic concepts between conventional CRI and D25.
3. Working process and injection performance 3.1. One-dimensional model of D25 The injection performance simulation is carried out with AMESimÒ (Advanced Modelling Environment for Simulations, version
Fig. 2. Cross-section of needle and grooves.
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Table 1 Diesel engine characteristics shown as an applicable example for D25. Items
Units
Value
Engine type Prototyping injection system
/ /
Engine configuration Displacement Rated power (at speed) Max. torque (at speed)
/ L kW (rpm) N m (rpm)
DEUTZÒ-TCD12.0 or 16.0 BOSCHÒ-CRSN3.3 (Rated 200 MPa SP) 6 or 8 cylinder V-type 11.906 or 15.874 390 or 520 (2100) 2130 or 2890 (1400)
experimental data of component is utilised to partially simplify the 1D model, including the drive current and armature pin displacement of solenoid-actuator. The solenoid-actuators were developed by FAWÒ-WFIERI, and the measurements were carried out via a combination of a fibre-coupled vibrometer sensor head POLYTECÒ-OFV-534 and a vibrometer controller OFV-2510 with resolution 0.08 lm/count and maximum linearity error ±0.08 lm/count. 3.2. Description of working process Fig. 4 shows the working process of D25. As a representative operating condition at a SP of 250 MPa and an energising time (ET) of 2 ms, Fig. 5 shows a series of working curves of key components. When D25 is in the initial situation before energising the electro-magnet of solenoid-actuator, as shown in Fig. 4(a), the control-valve and the nozzle needle are closed, and the controlpiston clings to control-piston seat owing to the pre-force of control-piston spring. When the electro-magnet is energised, as shown in Fig. 4(b), the control-valve opens. The pressure in intermediate chamber is lower than that in control-piston chamber owing to the throttling action of OA, creating a pressure difference between the two end
Fig. 4. Working process sketch of D25.
Fig. 3. Structure drawing of D25 and approximate installing condition.
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Fig. 5. Working curves of D25 (operating condition: 2 ms ET, 250 MPa SP).
surfaces of the control-piston. This pressure difference exerts a strong force on the control-piston, therefore maintaining a good sealing between the control-piston and the control-piston seat. Thanks to a decrease of 46% in diameter of OZ, only a small quantity of fuel can flow into control-chamber by going across the OZ, and thus the pressure difference remains adequate during required ET at the expense of slight delay to the pressure reduction of control-chamber. Once the control-chamber pressure is decreased to a certain lower-critical value, the nozzle needle opens and then fuel injection starts; the lower-critical value is approximately equal to 181 MPa at the operating condition of 250 MPa SP, as shown in Fig. 5(b). When the electro-magnet is de-energised, as shown in Fig. 4(c), the control-valve is closed, and the pressures in intermediate chamber and in control-piston chamber tend to be equal. Thus the control-piston begins to move downward owing to the force exerted by high-pressure chamber, whereupon fuel goes through the additional OZ’s and flows into control-piston chamber as well as control-chamber. Due to a pressure drop of 15–20% occurring in nozzle sac chamber during the period of fuel injection, once the control-chamber pressure is increased to a certain uppercritical value, the nozzle needle is closed, and the fuel injection is ended; the upper-critical value is about 217 MPa at the operating condition of 250 MPa SP, as shown in Fig. 5(b). Shortly afterwards, the pressures in control-piston chamber and in high-pressure chamber revert to equilibrium, and the control-piston returns to the initial position owing to the force of control-piston spring.
Fig. 6. Simulation result of single-injection fuel quantity of D25 at different SP’s and ET’s.
3.3. Injection performance Thanks to the new structure, D25 demonstrates outstanding injection performance. Fig. 6 shows fuel injection gain-curve. Main parameters of adopted nozzle model are 8 injection orifices and injection orifice diameter of 0.18 mm. Fig. 7 shows injection delays. Table 2 gives the contrast of injection performance between D25 and other similar products at the operating condition of 1.5 ms ET and 200 MPa SP which is limited to the maximum allowable
Fig. 7. Fuel injection opening-delay and closing-delay of D25 (operating condition: 250 MPa SP, 2 ms ET).
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Table 2 Contrast of fuel injection performance (working condition: 200 MPa SP, 1.5 ms ET). CRI type
BoschÒ-CRSN3.3 (experimental data) CumminsÒ-XPI (experimental data) D25 (simulation data)
Stroke end value (unit: mm)
Opening-delay
Closing-delay
Backflow fuel quantity
Value (unit: ms)
Rate of change
Value (unit: ms)
Rate of change
Performance index
Rate of change
0.35 0.40 0.35
0.540 0.478 0.406
Reference value 11.5% 24.8%
1.232 1.371 0.758
Reference value +11.3% 38.5%
22.5:246.1 27.0:197.9 14.1:182.6
Reference value +49.2% 15.5%
4. Evaluation of cavitation near control-valve 4.1. Cavitation damage in CRI The photo of cavitation damage on control-valve seat is shown in Fig. 10 with ZEISS-EVOÒ scanning electron microscopes (SEM). The cavitation erosion is characterised by the micro-phenomenon of irregular tiny pits, which is distinguished from the behaviour of scratches caused by particulate matter in fuel. 4.2. CFD model of control-valves
Fig. 8. Simulation result of single-injection rate shaping of D25 at different SP’s.
pressure of test bench. The result in terms of the delays indicates that injection responsiveness is markedly improved by D25. Fig. 8 shows that D25 achieves ramp-shaped single-injection rate shaping at different SP’s. Thanks to the reductions in closingdelay, the shaping demonstrates a sharp shutting, which is helpful to ameliorate the injection pressure condition at injection end period. The ratio of backflow fuel quantity to fuel injection quantity under a given SP condition is regarded as a performance index of fuel injection. The contrast of the index shown in Table 2 demonstrates that a good improvement in minimizing backflow fuel quantity has been achieved by D25. As for multi-injection performance, D25 is able to produce more than five injection events in one cycle. The minimum interval between the end of pre-injection and the start of main-injection, as shown in Fig. 9, can be shortened to 0.05 ms at a SP of 100 MPa which represents a certain operating condition of low engine load and engine speed.
The simulation is focused on two types of structures contrastively. In the conventional structure, as shown in Fig. 11(a), guide-hole and OA are linked together direct; in the new structure, as shown in Fig. 11(b), guide-hole is connected to OA by the intermediate chamber which is an essential sub-structure in D25, and the diameter of OA is decreased by 41% in contrast with the conventional structure. However, both the computational dynamic girds vary according to the same experimental data of lift at the operating condition of 2 ms ET. Moreover, the identical settings of CFD solver shown in Table 3 are set for both of them. The medium and fine meshes of the conventional structure have 91,287 and 574,050 cells, and the mass flow rate curves based on the two meshes are compared in Fig. 12, showing negligible difference. The similar result is obtained with regard to the new structure, the medium and fine meshes of which have 152,550 and 1,124,850 cells. Therefore, the medium meshes are employed for all the simulations for decreasing computational time. The details of adopted nonlinear cavitation model and the conservation equations of Eulerian multi-fluid model can be found in [22]. More information of adopted k–f–f turbulence model and compound wall treatment can be found in [23]. In order to attain numerically stable solutions, fuel is treated as incompressible fluid, and energy equation is deactivated during the simulation. However the bulk temperature of fuel is defined to calculate bulk density and dynamic viscosity with the Roelands viscosity equation [24] and a modified density equation [25]. In order to obtain necessary reference values for the equations, an experimental measurement of test fuel was accomplished at constant normal pressure and varying temperature conditions with a set of fluid property analyser MEASUREMENT-SPECIALTIESÒ-FP A2400BST. The measurement errors of dynamic viscosity, density and temperature are less than ±2 10 4 Pa s, ±1% and ±0.1 K, respectively. 4.3. Experimental validation
Fig. 9. Ability in minimum interval of multi-injection of D25 (operating condition: 100 MPa SP).
The validation test for the CFD model was turned to a cumulative mass flow rate measurement on the fuel outlet of CRI, as shown in Fig. 13. The test used a conventional type of CRI manufactured by FAWÒ-WFIERI, and four modifications were carried out to the CRI. (1) The pre-tightening force for control-valve was decreased greatly. Thus the control-valve can open and keep the maximum displacement owing to fuel pressure when planned
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Fig. 10. Photos of cavitation damage on control-valve-seat with scanning electron microscopes (after 1000 h alternating fatigue test).
Fig. 11. The 22.5-degree sector grids.
experimental SP established. Furthermore, drive current was deactivated to avoid the damages of electro-magnet and electronic control unit (ECU) because safe ET is generally less than 30 ms. (2) In order to maintain constant pressure in control-chamber when the control-valve was left open for one minute, a 2 1.5 1 (mm3) channel was machined within orifice-plate where OZ had been located originally by electro-spark process. (3) A special fuel nozzle without any injection orifice was adopted. (4) A calibrated pressure sensor KISTLERÒ-type4067A2000 was installed into the CRI to monitor control-chamber pressure, and it was fitted with a piezoresistive amplifier type-4618A0.
Table 3 Main setting of CFD solver. Items
Value
Units
Inlet boundary condition Outlet boundary condition Pressure-correction eq. Turbulence model Space-discretization Time-discretization Time step (Max./Min.) Convergence criteria for each time step
Total pressure Total pressure, 0.2 SIMPLE/PISO k–f–f Central differencing Euler scheme 1 10 4/1 10 5 Normalised residuals, less than 10 4
MPa MPa / / / / ms 1
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Fig. 12. Mesh independence test of the conventional structure (inlet BC = 250 MPa, outlet BC = 0.2 MPa, Bulk temperature = 393 K).
A BOSCHÒ-MOEHWALD-CA4000 was chosen as the test bench for the measurement, and it was equipped with a pressure transducer HBMÒ-1-P3TCP/3000 bar to observe SP. The observation shows that the SP and the pressure in control-chamber are approximately equal and both of them remain constant when controlvalve is open. Moreover, the temperature of fuel tank within the
test bench can be monitored. Also the temperature of backflow fuel was overseen by a k-type thermocouple which was fittingly inserted into the backflow fuel line. A container was set above an electronic balance to collect the fuel and to measure the change of weight. A two-position three-port valve was introduced to direct the backflow fuel to the fuel tank or to the container. Another twoport valve was adopted to govern when the backflow fuel returned to the fuel tank from the container. A series of target SP’s were defined by a common difference of 20 MPa. In warm-up period, the backflow fuel flowed through the three-port valve and returned to the fuel tank. When a target SP established and the temperature of backflow fuel was stable, the three-port valve was switched and then the backflow fuel flowed into the container accurately for one minute. Weight change of the container was recorded to acquire correlative mass flow rate of backflow fuel. The experiments were repeated at each target SP condition to calculate respective average value. After every measurement, the backflow fuel returned to the fuel tank from the container, and the temperature of fuel tank was cooled down to 40 °C by the test bench. The CFD model is validated against the test result in terms of mass flow rate in the conventional structure CRI, as shown in
Fig. 13. Sketch of validation experiment for CFD numerical simulation.
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Fig. 14. Each CFD simulation data point is obtained from the instantaneous value of simulation result at one time-step before starting to close the control-valve. The validation test was limited to no more than 120 MPa SP condition for preventing fuel from being overheated as a result of throttling action. 4.4. Result and discussion
Fig. 14. CFD simulation validation in terms of mass flow rate in the conventional structure CRI.
The contrastive analysis of cavitation, which potentially occurs near the control-valves of the two types of injectors, is not only focused on the distributions of velocity (Fig. 15(a)) and void fraction (Fig. 15(b)), but also on the quantitative analysis of a specified region. The specified region is defined around the sealing annulus, as shown in Fig. 15(c). In order to facilitate the analysis, 14 moni-
Fig. 15. Contrast of two types of structures, based on CFD simulation of cavitation flow (outlet BC = 0.2 MPa).
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Table 4 Monitoring points in time sequence and the correlative lifts of control-valve. Sequence
Time (ms) Lift (mm)
1#
2#
3#
4#
5#
6#
7#
8#
9#
10#
11#
12#
13#
14#
0.034 0.02
0.059 0.03
0.077 0.04
0.093 0.05
0.400 0.05
0.700 0.05
1.000 0.05
1.300 0.05
1.600 0.05
1.870 0.05
1.922 0.04
1.945 0.03
1.960 0.02
1.974 0.01
toring points (Table 4) have been selected in time sequence, throughout the working process of the control-valves. Fig. 15(a) shows that the liquid-phase bulk velocity in the new structure is lower than that in the conventional structure, throughout whole time period. Fig. 15(b) shows that cavitation no longer occurs within the guide-hole of the new structure. Fuel-flow is separated from the inner wall surface of the guide-hole in the conventional structure, but the separating phenomenon does not appear in the new structure. The quantitative analysis of the specified region is shown in Fig. 15(d)–(f). Fig. 15(d) shows that the average velocity of fuelflow is up to a peak value of 900.25 m/s in the conventional structure at #14, but the corresponding value in the new structure is below 311.88 m/s. Fig. 15(e) shows that the void fraction values are approximate for the conventional structure under different inlet BC; it reveals that pressure-drop variance, among the present CRI working conditions (from 150 to 250 MPa), has little influence on void fraction. Although the similar void fraction levels are demonstrated in both the structures, a significant reduction in mass exchange rate in the new structure is shown in Fig. 15(f). The quantification of the reduction is equal to 70.22% calculated by means of interpolation and integral with respect to time. Less cavitation is observed in the specified region of the new structure as a result of the following factors: r lower flow rate, s nonexistence of cavitation bubbles in guide-hole, t lower mass exchange rate between vapour–liquid phases. Thanks to the 41% reduction in diameter of OA, more throttling action considerably reduces the flow rate in downstream region. Therefore the growths of cavitation bubbles induced by high-turbulent shear-flow are decelerated. The intermediate chamber acts as a buffer channel, and thus the bubbles caused by OA gradually collapse and cannot arrive at guide-hole. Moreover, the joint of flow path on the inlet of guide-hole becomes a sudden contraction instead of the conventional sudden expansion. The contraction has a dampening effect on eddy formation [26], and avoids pockets of eddying turbulence which occur in the sudden expansion owing to the flow separating from the wall of guide-hole. Thereby the growths of cavitation bubbles derived from eddies are also suppressed. Additionally, the lower mass exchange rate between vapour–liquid phases suggests that the frequency of collapses of cavitation bubbles near control-valve seat surface decreases. 5. Conclusion This study describes a new type of injector (code-named D25) for 250 MPa CRS diesel engines. The following conclusions can be drawn: 1. The 1D simulation result has demonstrated the application feasibility of D25 in a heavy duty diesel engine. D25 is capable of achieving ramp-shaped injection rate as well as multi-injection. 2. The injection responsiveness of D25 is significantly improved. Opening-delay can be reduced from 0.540 ms to 0.406 ms, and closing-delay can be lowered from 1.232 to 0.758 ms, which indicates considerable reductions of 24.8% and 38.5% respectively.
3. D25 achieves good progress in minimizing backflow fuel quantity. The performance index of backflow fuel quantity is ‘‘14.1 mm3/182.6 mm3” at the working condition of 200 MPa SP and 1.5 ms ET, corresponding to a decrease of 15.5%. 4. Around the control-valve of D25, a lower velocity field (below 311.88 m/s) is observed, and cavitation no longer occurs in the guide-hole. Moreover, the mass exchange rate between vapour–liquid phases is reduced by 70.22% near the sealing annulus of valve seat. 5. The result demonstrates that the cavitation intensity near control-valve can be reduced by two modifications in structure: one is a proper reduction in diameter of OA; the other is a welldesigned intermediate chamber which not only acts as a buffer channel positioned between the OA and the guide-hole, but also forms a sudden contraction on the inlet of the guide-hole instead of a conventional sudden expansion. Conflict of interest None declared. Acknowledgements This study is supported by the National Natural Science Foundation of China (Grant No. 51239005). Also this work is supported by the National Science and Technology Major Project of China (Grant No. 2011ZX04001-061). References [1] H. Busch, L. Henning, T. Körfer, et al., Diesel engine development for emission standards in emerging markets, MTZ 72 (12) (2011) 36–41. [2] P. Wuensche, F. Moser, R. Dreisbach, et al., Can the technology for heavy duty diesel engines be common for future emission regulations in USA, Japan and Europe? SAE Technical Paper 2003-01-0344, 2003. [3] J. Theobald, K. Schintzel, A. Krause, et al., Fuel injection system key component for future emission targets, MTZ 72 (4) (2011) 4–9. [4] K. Nishida, W. Zhang, T. Manabe, Effects of micro-hole and ultra-high injection pressure on mixture properties of D.I. Diesel spray, SAE Technical Paper 200701-1890, 2007, pp. 1353–1361. [5] F. Tao, P. Bergstrand, Effect of ultra-high injection pressure on diesel ignition and flame under high-boost conditions, SAE Technical Paper 2008-01-1603, 2008, pp. 1–22. [6] O.A. Kuti, W. Zhang, K. Nishida, et al., Effect of injection pressure on ignition, flame development and soot formation processes of biodiesel fuel spray, SAE Technical Paper 2010-32-0053, 2010. [7] P. Karra, S.C. Kong, Diesel emission characteristics using high injection pressure with converging nozzles in a medium-duty engine, SAE Technical Paper 2008-01-1085, 2008, pp. 578–592. [8] A.E. Catania, S. d’Ambrosio, R. Finesso, et al., Effects of rail pressure, pilot scheduling and egr rate on combustion and emissions in conventional and PCCI diesel engines, SAE Technical Paper 2010-01-1109, (3)1 (2010) 773–787. [9] M. Thirouard, S. Mendez, P. Pacaud, et al., Potential to improve specific power using very high injection pressure in HSDI diesel engines, SAE Technical Paper 2009-01-1524, 2009. [10] Ficarella, D. Laforgia, V. Landriscina, Evaluation of instability phenomena in a common rail injection system for high speed diesel engines, SAE technical paper 1999-01-0192, 1999, pp. 1–19. [11] G.M. Bianchi, S. Falfari, M. Parotto, et al., Advanced modeling of common rail injector dynamics and comparison with experiments, SAE Technical Paper 2003-01-0006, 2003. [12] S.H. Xia, J.B. Zhen, X.L. Miao, et al., Cavitation analysis on ball valve in solenoid injector, Trans. CSICE 30 (4) (2012) 354–358.
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Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Injection performance and cavitation analysis of an advanced 250 MPa common rail diesel injector Lian Duan a, Shou-qi Yuan a, Lin-feng Hu b, Wen-ming Yang c,⇑, Jian-da Yu b, Xing-lan Xia b a
Research Centre of Fluid Machinery Engineering & Technology, Jiangsu University, Zhenjiang 212013, China FAWÒ Wuxi Fuel Injection Equipment Research Institute, Wuxi 214063, China c Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore 117575, Singapore b
a r t i c l e
i n f o
Article history: Received 21 April 2015 Received in revised form 12 October 2015 Accepted 13 October 2015
Keywords: Common rail injection system Injection responsiveness Backflow fuel Multi-injection Valve cavitation
a b s t r a c t The present work is to demonstrate the working process and injection performance of a new-type diesel injector, and to make a contrastive analysis of cavitation which potentially occurs near control-valves of the new and another conventional injector. The advanced injector is developed for 250 MPa common rail injection systems of medium or heavy duty commercial vehicles. The originality is the development of a novel control-piston with the outward appearance of a poppet valve. The control-piston can automatically move downwards or upwards owing to pressure changes in related hydraulic chambers, and consequently appropriately activates or deactivates the function of additional inflowing control-orifices to improve system responsiveness. The injection performance demonstration was carried out by one-dimensional (1D) hydraulic mechanical simulation in combination with the experimental data input of actuators. On the other hand the cavitation evaluations, in both the conventional and new structures, were studied numerically by computational fluid dynamics (CFD) based on nonlinear cavitation model and dynamic grid method. Additionally the CFD model was validated against experimental result in terms of mass flow rate of backflow fuel in the conventional structure. The 1D simulation result shows that the new injector is capable of achieving multi-injection and ramp-shaped injection flow rate; closing-delay of injection can be reduced from 1.232 to 0.758 ms, and backflow fuel quantity index can be decreased to 14.1:182.6 which means a reduction of 15.5%. As regards to the contrastive analysis of cavitation, the CFD simulation result of the new structure shows several improvements as follows: a lower velocity field (below 311.88 m/s) near the control-valve, cavitation no longer occurring in the guide-hole, and a reduction of 70.22% in mass exchange rate between vapour–liquid phases near the sealing annulus of valve seat. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction To meet progressively strict emission regulations [1,2], performance and fuel-economy, diesel injection equipment [3] is expected to be capable of higher system pressure (SP), injection precision, robustness and rapid injection responsiveness. Common rail injection system (CRS) is the preferred diesel injection equipment due to its excellent flexibility. Increased SP has plenty of influence on injection performance, such as acquiring finer spray droplets, enhancing turbulent mixing rate to attain better fuel–air mixture [4], shortening ignition delay time [5,6], reducing soot emission [7,8] and improving specific power [9]. Currently as high as 250 MPa SP is important to obtain adequate spray characteristics for achieving stricter emission regu⇑ Corresponding author. Tel.: +65 6516 6481. E-mail address: [email protected] (W.-m. Yang). http://dx.doi.org/10.1016/j.ijheatmasstransfer.2015.10.028 0017-9310/Ó 2015 Elsevier Ltd. All rights reserved.
lations Euro-VI and more satisfactory engine performance. In such a high-pressure operating condition, several problems exist in common rail injector (CRI), such as insufficiently rapid responsiveness in closing fuel injection, a large quantity of backflow fuel, and worse cavitation damage (viz., cavitation erosion) on control-valve. Flow-induced cavitation frequently occurs in CRI owing to high pressure-drop or/and geometric change of flow path [10]. Cavitation can change flow-characteristic and damage wall surface, especially in the vicinity of control-valve. Thus it results in a deviation of fuel injection characteristic and a risk of fuel injection malfunction due to control-valve leakage. Therefore, it is necessary to evaluate the level of cavitation near control-valve in CRI’s. However, the flow condition near control-valve (sub-millimetre-scale, submillisecond-timescale, more than 100 MPa pressure difference and hundreds m/s flow velocity) is so complicate that experimental visualisation method is unfeasible, or the costs involved is prohibitive. Thereby numerical simulation method appears to be
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Nomenclature D25 OA OZ
code name of the new-type diesel injector outflowing control-orifice inflowing control-orifice
Abbreviations 1D one dimensional
a proper approach for disclosing the flow feature. Bianchi et al. [11] studied the cavitation influence on discharge coefficient in controlorifices and control-valve with STAR-CDÒ, nevertheless, the model was steady-state and laminar flow. Xia et al. [12] investigated the cavitation flow of the region by linear cavitation model with AVLFIREÒ, but correlative experimental validation was not presented. At present, several types of 250 MPa-level CRI products have been launched by manufacturers for diesel engines of commercial vehicles, such as BOSCHÒ-CRIN4.2 [13], CUMMINSÒ-XPI [14], DENSOÒ-G4S [15]. They have adopted different structural designs as a solution for aforementioned problems, but their sophisticated structures require advanced manufacturing technique, and their improved results in closing-delay are less than satisfactory. In order to obtain a better solution to aforementioned problems, a new type of diesel injector (code-named D25) is developed for 250 MPa CRS of medium or heavy duty commercial vehicles. In this paper, the structure and working process of D25 is described, and its injection performance is demonstrated by 1D hydraulic mechanical simulation in combination with the experimental data input of actuators. Moreover, in order to evaluate the effect of the new structure on cavitation in the neighbourhood of control-valve, a CFD simulation analysis is carried out contrastively in two types of structures (the new one and another conventional one). The CFD simulation is focused on three-dimensional unsteady two-phase flow, and is based on nonlinear cavitation model as well as dynamic grid method with AVL-FIREÒ v.2010. Additionally the CFD simulation is validated against experimental result in terms of the mass flow rate of backflow fuel in the conventional structure CRI.
BC CFD CRI CRS ET SP
boundary condition computational fluid dynamics common rail injector common rail injection system energising time system pressure
12.0), and the applications of the software for CRS modelling have been done by other authors [17–20]. With the software, the 1D model and validation of a conventional CRI have been performed in a previous work [21]. The trial-manufacture of D25 is still in preparation. Therefore injection performance experiment is not yet available for the validation of 1D model of D25; presented 1D simulation result serves solely for functionality demonstration. Nevertheless a part of
2. Development of new structure The contrast of hydraulic concepts between conventional CRI and D25 is illustrated in Fig. 1. There have been seven improvements in D25. r Clearance leakage is avoided; s the diameters of inflowing control-orifice (commonly known as OZ) and outflowing control-orifice (commonly known as OA) are decreased by about 46% and 41% respectively; t additional OZ and a controlpiston which has the outward appearance of a poppet valve are designed; u control-chamber is positioned directly above the nozzle needle instead of above a command-piston; v a fuel storage chamber (similarly introduced in [16]) is implanted; w three grooves with circular-segment shaped cross-section (Fig. 2), which are machined regularly around the nozzle needle, are adopted as fuel delivery channels in the nozzle; x an intermediate chamber is put between OA and control-valve. Based on approximate installing condition of an engine example (Table 1), the structure of D25 is shown in Fig. 3.
Fig. 1. Contrast of hydraulic concepts between conventional CRI and D25.
3. Working process and injection performance 3.1. One-dimensional model of D25 The injection performance simulation is carried out with AMESimÒ (Advanced Modelling Environment for Simulations, version
Fig. 2. Cross-section of needle and grooves.
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Table 1 Diesel engine characteristics shown as an applicable example for D25. Items
Units
Value
Engine type Prototyping injection system
/ /
Engine configuration Displacement Rated power (at speed) Max. torque (at speed)
/ L kW (rpm) N m (rpm)
DEUTZÒ-TCD12.0 or 16.0 BOSCHÒ-CRSN3.3 (Rated 200 MPa SP) 6 or 8 cylinder V-type 11.906 or 15.874 390 or 520 (2100) 2130 or 2890 (1400)
experimental data of component is utilised to partially simplify the 1D model, including the drive current and armature pin displacement of solenoid-actuator. The solenoid-actuators were developed by FAWÒ-WFIERI, and the measurements were carried out via a combination of a fibre-coupled vibrometer sensor head POLYTECÒ-OFV-534 and a vibrometer controller OFV-2510 with resolution 0.08 lm/count and maximum linearity error ±0.08 lm/count. 3.2. Description of working process Fig. 4 shows the working process of D25. As a representative operating condition at a SP of 250 MPa and an energising time (ET) of 2 ms, Fig. 5 shows a series of working curves of key components. When D25 is in the initial situation before energising the electro-magnet of solenoid-actuator, as shown in Fig. 4(a), the control-valve and the nozzle needle are closed, and the controlpiston clings to control-piston seat owing to the pre-force of control-piston spring. When the electro-magnet is energised, as shown in Fig. 4(b), the control-valve opens. The pressure in intermediate chamber is lower than that in control-piston chamber owing to the throttling action of OA, creating a pressure difference between the two end
Fig. 4. Working process sketch of D25.
Fig. 3. Structure drawing of D25 and approximate installing condition.
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Fig. 5. Working curves of D25 (operating condition: 2 ms ET, 250 MPa SP).
surfaces of the control-piston. This pressure difference exerts a strong force on the control-piston, therefore maintaining a good sealing between the control-piston and the control-piston seat. Thanks to a decrease of 46% in diameter of OZ, only a small quantity of fuel can flow into control-chamber by going across the OZ, and thus the pressure difference remains adequate during required ET at the expense of slight delay to the pressure reduction of control-chamber. Once the control-chamber pressure is decreased to a certain lower-critical value, the nozzle needle opens and then fuel injection starts; the lower-critical value is approximately equal to 181 MPa at the operating condition of 250 MPa SP, as shown in Fig. 5(b). When the electro-magnet is de-energised, as shown in Fig. 4(c), the control-valve is closed, and the pressures in intermediate chamber and in control-piston chamber tend to be equal. Thus the control-piston begins to move downward owing to the force exerted by high-pressure chamber, whereupon fuel goes through the additional OZ’s and flows into control-piston chamber as well as control-chamber. Due to a pressure drop of 15–20% occurring in nozzle sac chamber during the period of fuel injection, once the control-chamber pressure is increased to a certain uppercritical value, the nozzle needle is closed, and the fuel injection is ended; the upper-critical value is about 217 MPa at the operating condition of 250 MPa SP, as shown in Fig. 5(b). Shortly afterwards, the pressures in control-piston chamber and in high-pressure chamber revert to equilibrium, and the control-piston returns to the initial position owing to the force of control-piston spring.
Fig. 6. Simulation result of single-injection fuel quantity of D25 at different SP’s and ET’s.
3.3. Injection performance Thanks to the new structure, D25 demonstrates outstanding injection performance. Fig. 6 shows fuel injection gain-curve. Main parameters of adopted nozzle model are 8 injection orifices and injection orifice diameter of 0.18 mm. Fig. 7 shows injection delays. Table 2 gives the contrast of injection performance between D25 and other similar products at the operating condition of 1.5 ms ET and 200 MPa SP which is limited to the maximum allowable
Fig. 7. Fuel injection opening-delay and closing-delay of D25 (operating condition: 250 MPa SP, 2 ms ET).
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Table 2 Contrast of fuel injection performance (working condition: 200 MPa SP, 1.5 ms ET). CRI type
BoschÒ-CRSN3.3 (experimental data) CumminsÒ-XPI (experimental data) D25 (simulation data)
Stroke end value (unit: mm)
Opening-delay
Closing-delay
Backflow fuel quantity
Value (unit: ms)
Rate of change
Value (unit: ms)
Rate of change
Performance index
Rate of change
0.35 0.40 0.35
0.540 0.478 0.406
Reference value 11.5% 24.8%
1.232 1.371 0.758
Reference value +11.3% 38.5%
22.5:246.1 27.0:197.9 14.1:182.6
Reference value +49.2% 15.5%
4. Evaluation of cavitation near control-valve 4.1. Cavitation damage in CRI The photo of cavitation damage on control-valve seat is shown in Fig. 10 with ZEISS-EVOÒ scanning electron microscopes (SEM). The cavitation erosion is characterised by the micro-phenomenon of irregular tiny pits, which is distinguished from the behaviour of scratches caused by particulate matter in fuel. 4.2. CFD model of control-valves
Fig. 8. Simulation result of single-injection rate shaping of D25 at different SP’s.
pressure of test bench. The result in terms of the delays indicates that injection responsiveness is markedly improved by D25. Fig. 8 shows that D25 achieves ramp-shaped single-injection rate shaping at different SP’s. Thanks to the reductions in closingdelay, the shaping demonstrates a sharp shutting, which is helpful to ameliorate the injection pressure condition at injection end period. The ratio of backflow fuel quantity to fuel injection quantity under a given SP condition is regarded as a performance index of fuel injection. The contrast of the index shown in Table 2 demonstrates that a good improvement in minimizing backflow fuel quantity has been achieved by D25. As for multi-injection performance, D25 is able to produce more than five injection events in one cycle. The minimum interval between the end of pre-injection and the start of main-injection, as shown in Fig. 9, can be shortened to 0.05 ms at a SP of 100 MPa which represents a certain operating condition of low engine load and engine speed.
The simulation is focused on two types of structures contrastively. In the conventional structure, as shown in Fig. 11(a), guide-hole and OA are linked together direct; in the new structure, as shown in Fig. 11(b), guide-hole is connected to OA by the intermediate chamber which is an essential sub-structure in D25, and the diameter of OA is decreased by 41% in contrast with the conventional structure. However, both the computational dynamic girds vary according to the same experimental data of lift at the operating condition of 2 ms ET. Moreover, the identical settings of CFD solver shown in Table 3 are set for both of them. The medium and fine meshes of the conventional structure have 91,287 and 574,050 cells, and the mass flow rate curves based on the two meshes are compared in Fig. 12, showing negligible difference. The similar result is obtained with regard to the new structure, the medium and fine meshes of which have 152,550 and 1,124,850 cells. Therefore, the medium meshes are employed for all the simulations for decreasing computational time. The details of adopted nonlinear cavitation model and the conservation equations of Eulerian multi-fluid model can be found in [22]. More information of adopted k–f–f turbulence model and compound wall treatment can be found in [23]. In order to attain numerically stable solutions, fuel is treated as incompressible fluid, and energy equation is deactivated during the simulation. However the bulk temperature of fuel is defined to calculate bulk density and dynamic viscosity with the Roelands viscosity equation [24] and a modified density equation [25]. In order to obtain necessary reference values for the equations, an experimental measurement of test fuel was accomplished at constant normal pressure and varying temperature conditions with a set of fluid property analyser MEASUREMENT-SPECIALTIESÒ-FP A2400BST. The measurement errors of dynamic viscosity, density and temperature are less than ±2 10 4 Pa s, ±1% and ±0.1 K, respectively. 4.3. Experimental validation
Fig. 9. Ability in minimum interval of multi-injection of D25 (operating condition: 100 MPa SP).
The validation test for the CFD model was turned to a cumulative mass flow rate measurement on the fuel outlet of CRI, as shown in Fig. 13. The test used a conventional type of CRI manufactured by FAWÒ-WFIERI, and four modifications were carried out to the CRI. (1) The pre-tightening force for control-valve was decreased greatly. Thus the control-valve can open and keep the maximum displacement owing to fuel pressure when planned
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Fig. 10. Photos of cavitation damage on control-valve-seat with scanning electron microscopes (after 1000 h alternating fatigue test).
Fig. 11. The 22.5-degree sector grids.
experimental SP established. Furthermore, drive current was deactivated to avoid the damages of electro-magnet and electronic control unit (ECU) because safe ET is generally less than 30 ms. (2) In order to maintain constant pressure in control-chamber when the control-valve was left open for one minute, a 2 1.5 1 (mm3) channel was machined within orifice-plate where OZ had been located originally by electro-spark process. (3) A special fuel nozzle without any injection orifice was adopted. (4) A calibrated pressure sensor KISTLERÒ-type4067A2000 was installed into the CRI to monitor control-chamber pressure, and it was fitted with a piezoresistive amplifier type-4618A0.
Table 3 Main setting of CFD solver. Items
Value
Units
Inlet boundary condition Outlet boundary condition Pressure-correction eq. Turbulence model Space-discretization Time-discretization Time step (Max./Min.) Convergence criteria for each time step
Total pressure Total pressure, 0.2 SIMPLE/PISO k–f–f Central differencing Euler scheme 1 10 4/1 10 5 Normalised residuals, less than 10 4
MPa MPa / / / / ms 1
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Fig. 12. Mesh independence test of the conventional structure (inlet BC = 250 MPa, outlet BC = 0.2 MPa, Bulk temperature = 393 K).
A BOSCHÒ-MOEHWALD-CA4000 was chosen as the test bench for the measurement, and it was equipped with a pressure transducer HBMÒ-1-P3TCP/3000 bar to observe SP. The observation shows that the SP and the pressure in control-chamber are approximately equal and both of them remain constant when controlvalve is open. Moreover, the temperature of fuel tank within the
test bench can be monitored. Also the temperature of backflow fuel was overseen by a k-type thermocouple which was fittingly inserted into the backflow fuel line. A container was set above an electronic balance to collect the fuel and to measure the change of weight. A two-position three-port valve was introduced to direct the backflow fuel to the fuel tank or to the container. Another twoport valve was adopted to govern when the backflow fuel returned to the fuel tank from the container. A series of target SP’s were defined by a common difference of 20 MPa. In warm-up period, the backflow fuel flowed through the three-port valve and returned to the fuel tank. When a target SP established and the temperature of backflow fuel was stable, the three-port valve was switched and then the backflow fuel flowed into the container accurately for one minute. Weight change of the container was recorded to acquire correlative mass flow rate of backflow fuel. The experiments were repeated at each target SP condition to calculate respective average value. After every measurement, the backflow fuel returned to the fuel tank from the container, and the temperature of fuel tank was cooled down to 40 °C by the test bench. The CFD model is validated against the test result in terms of mass flow rate in the conventional structure CRI, as shown in
Fig. 13. Sketch of validation experiment for CFD numerical simulation.
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Fig. 14. Each CFD simulation data point is obtained from the instantaneous value of simulation result at one time-step before starting to close the control-valve. The validation test was limited to no more than 120 MPa SP condition for preventing fuel from being overheated as a result of throttling action. 4.4. Result and discussion
Fig. 14. CFD simulation validation in terms of mass flow rate in the conventional structure CRI.
The contrastive analysis of cavitation, which potentially occurs near the control-valves of the two types of injectors, is not only focused on the distributions of velocity (Fig. 15(a)) and void fraction (Fig. 15(b)), but also on the quantitative analysis of a specified region. The specified region is defined around the sealing annulus, as shown in Fig. 15(c). In order to facilitate the analysis, 14 moni-
Fig. 15. Contrast of two types of structures, based on CFD simulation of cavitation flow (outlet BC = 0.2 MPa).
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Table 4 Monitoring points in time sequence and the correlative lifts of control-valve. Sequence
Time (ms) Lift (mm)
1#
2#
3#
4#
5#
6#
7#
8#
9#
10#
11#
12#
13#
14#
0.034 0.02
0.059 0.03
0.077 0.04
0.093 0.05
0.400 0.05
0.700 0.05
1.000 0.05
1.300 0.05
1.600 0.05
1.870 0.05
1.922 0.04
1.945 0.03
1.960 0.02
1.974 0.01
toring points (Table 4) have been selected in time sequence, throughout the working process of the control-valves. Fig. 15(a) shows that the liquid-phase bulk velocity in the new structure is lower than that in the conventional structure, throughout whole time period. Fig. 15(b) shows that cavitation no longer occurs within the guide-hole of the new structure. Fuel-flow is separated from the inner wall surface of the guide-hole in the conventional structure, but the separating phenomenon does not appear in the new structure. The quantitative analysis of the specified region is shown in Fig. 15(d)–(f). Fig. 15(d) shows that the average velocity of fuelflow is up to a peak value of 900.25 m/s in the conventional structure at #14, but the corresponding value in the new structure is below 311.88 m/s. Fig. 15(e) shows that the void fraction values are approximate for the conventional structure under different inlet BC; it reveals that pressure-drop variance, among the present CRI working conditions (from 150 to 250 MPa), has little influence on void fraction. Although the similar void fraction levels are demonstrated in both the structures, a significant reduction in mass exchange rate in the new structure is shown in Fig. 15(f). The quantification of the reduction is equal to 70.22% calculated by means of interpolation and integral with respect to time. Less cavitation is observed in the specified region of the new structure as a result of the following factors: r lower flow rate, s nonexistence of cavitation bubbles in guide-hole, t lower mass exchange rate between vapour–liquid phases. Thanks to the 41% reduction in diameter of OA, more throttling action considerably reduces the flow rate in downstream region. Therefore the growths of cavitation bubbles induced by high-turbulent shear-flow are decelerated. The intermediate chamber acts as a buffer channel, and thus the bubbles caused by OA gradually collapse and cannot arrive at guide-hole. Moreover, the joint of flow path on the inlet of guide-hole becomes a sudden contraction instead of the conventional sudden expansion. The contraction has a dampening effect on eddy formation [26], and avoids pockets of eddying turbulence which occur in the sudden expansion owing to the flow separating from the wall of guide-hole. Thereby the growths of cavitation bubbles derived from eddies are also suppressed. Additionally, the lower mass exchange rate between vapour–liquid phases suggests that the frequency of collapses of cavitation bubbles near control-valve seat surface decreases. 5. Conclusion This study describes a new type of injector (code-named D25) for 250 MPa CRS diesel engines. The following conclusions can be drawn: 1. The 1D simulation result has demonstrated the application feasibility of D25 in a heavy duty diesel engine. D25 is capable of achieving ramp-shaped injection rate as well as multi-injection. 2. The injection responsiveness of D25 is significantly improved. Opening-delay can be reduced from 0.540 ms to 0.406 ms, and closing-delay can be lowered from 1.232 to 0.758 ms, which indicates considerable reductions of 24.8% and 38.5% respectively.
3. D25 achieves good progress in minimizing backflow fuel quantity. The performance index of backflow fuel quantity is ‘‘14.1 mm3/182.6 mm3” at the working condition of 200 MPa SP and 1.5 ms ET, corresponding to a decrease of 15.5%. 4. Around the control-valve of D25, a lower velocity field (below 311.88 m/s) is observed, and cavitation no longer occurs in the guide-hole. Moreover, the mass exchange rate between vapour–liquid phases is reduced by 70.22% near the sealing annulus of valve seat. 5. The result demonstrates that the cavitation intensity near control-valve can be reduced by two modifications in structure: one is a proper reduction in diameter of OA; the other is a welldesigned intermediate chamber which not only acts as a buffer channel positioned between the OA and the guide-hole, but also forms a sudden contraction on the inlet of the guide-hole instead of a conventional sudden expansion. Conflict of interest None declared. Acknowledgements This study is supported by the National Natural Science Foundation of China (Grant No. 51239005). Also this work is supported by the National Science and Technology Major Project of China (Grant No. 2011ZX04001-061). References [1] H. Busch, L. Henning, T. Körfer, et al., Diesel engine development for emission standards in emerging markets, MTZ 72 (12) (2011) 36–41. [2] P. Wuensche, F. Moser, R. Dreisbach, et al., Can the technology for heavy duty diesel engines be common for future emission regulations in USA, Japan and Europe? SAE Technical Paper 2003-01-0344, 2003. [3] J. Theobald, K. Schintzel, A. Krause, et al., Fuel injection system key component for future emission targets, MTZ 72 (4) (2011) 4–9. [4] K. Nishida, W. Zhang, T. Manabe, Effects of micro-hole and ultra-high injection pressure on mixture properties of D.I. Diesel spray, SAE Technical Paper 200701-1890, 2007, pp. 1353–1361. [5] F. Tao, P. Bergstrand, Effect of ultra-high injection pressure on diesel ignition and flame under high-boost conditions, SAE Technical Paper 2008-01-1603, 2008, pp. 1–22. [6] O.A. Kuti, W. Zhang, K. Nishida, et al., Effect of injection pressure on ignition, flame development and soot formation processes of biodiesel fuel spray, SAE Technical Paper 2010-32-0053, 2010. [7] P. Karra, S.C. Kong, Diesel emission characteristics using high injection pressure with converging nozzles in a medium-duty engine, SAE Technical Paper 2008-01-1085, 2008, pp. 578–592. [8] A.E. Catania, S. d’Ambrosio, R. Finesso, et al., Effects of rail pressure, pilot scheduling and egr rate on combustion and emissions in conventional and PCCI diesel engines, SAE Technical Paper 2010-01-1109, (3)1 (2010) 773–787. [9] M. Thirouard, S. Mendez, P. Pacaud, et al., Potential to improve specific power using very high injection pressure in HSDI diesel engines, SAE Technical Paper 2009-01-1524, 2009. [10] Ficarella, D. Laforgia, V. Landriscina, Evaluation of instability phenomena in a common rail injection system for high speed diesel engines, SAE technical paper 1999-01-0192, 1999, pp. 1–19. [11] G.M. Bianchi, S. Falfari, M. Parotto, et al., Advanced modeling of common rail injector dynamics and comparison with experiments, SAE Technical Paper 2003-01-0006, 2003. [12] S.H. Xia, J.B. Zhen, X.L. Miao, et al., Cavitation analysis on ball valve in solenoid injector, Trans. CSICE 30 (4) (2012) 354–358.
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