The Korean Society of Marine Engineering
[ Original Paper ]
Journal of Advanced Marine Engineering and Technology - Vol. 45, No. 4, pp.140-153
ISSN: 2234-7925 (Print) 2765-4796 (Online)
Print publication date 31 Aug 2021
Received 17 May 2021 Revised 14 Jun 2021 Accepted 04 Jul 2021
DOI: https://doi.org/10.5916/jamet.2021.45.4.140

Performance failure simulation and characteristic analysis of marine diesel engine turbocharging and gas-exchange system under different running conditions

Yihuai Hu ; Meng Wang1 ; Cun Zeng2 ; Jiawei Jiang3
1M. S., Department of Marine Engineering, Shanghai Maritime University, Tel: +86-18521012168 mengw725@foxmail.com
2Ph. D. Candidate, Department of Marine Engineering, Shanghai Maritime University, Tel: +86-18916271125 zc.shmtu@foxmail.com
3Researcher, Department of Marine Engineering, Shanghai Maritime University, Tel: +86-18635191739 oliver1992@vip.qq.com

Correspondence to: Professor, Department of Marine Engineering, Shanghai Maritime University, 1550 Haigang Avenue, Lingang New City, Pudong New Area, Shanghai, P. R. China, E-mail: yhhu@shmtu.edu.cn, Tel: +86-13651857174

Copyright © The Korean Society of Marine Engineering
This is an Open Access article distributed under the terms of the Creative Commons Attribution Non-Commercial License (http://creativecommons.org/licenses/by-nc/3.0), which permits unrestricted non-commercial use, distribution, and reproduction in any medium, provided the original work is properly cited.

Abstract

Taking a two-stroke marine diesel engine as an example, this study first builds up a thermodynamic simulation model of a turbocharging and gas-exchange system in MATLAB/Simulink, which is verified by the experimental results of a diesel engine under normal conditions. The boundary condition and corresponding model parameters are changed to simulate typical performance failures, including extremely high engine room temperature, intake filter blockage, intercooler fouling at the water side, overly high intercooler cooling water temperature, scavenging port fouling, clogged turbine nozzle, worn turbocharger bearing, and turbine exhaust passage fouling. The thermodynamic parameters of a diesel engine under different running conditions and performance failures are analyzed in terms of relative deviation, which demonstrates the relationship between the performance failures and thermodynamic parameters. This relationship could be mostly free of the engine running conditions. Finally, a normalization method is proposed to eliminate the influences of the engine room temperature and intercooler cooling water temperature. Thus, the performance failures could be detected according to the relative deviation of the proposed thermodynamic parameters under different running conditions throughout the entire engine working range. This method may provide a more practical foundation and possible breakthrough in the condition monitoring and failure diagnosis of marine diesel engines onboard ships.

Keywords:

Marine diesel engine, Turbocharging and gas‐exchange, Engine performance simulation, Failure diagnosis

1. Introduction

With the rapid development of ship technology, ocean-going ships are tending to be large scaled and intelligent. When abnormal phenomena occur during ship operation, it becomes increasingly difficult for engineers in the engine room to quickly analyze the causes of failure and respond accurately in the face of complicated puzzling information. The development of marine engine condition monitoring and failure diagnosis technology can help engineers address failures in time, save over 30% overhaul costs, reduce maintenance labor by 37%, and avoid accidents [1]. Additionally, the utilization of equipment can be extended, resulting in considerable economic benefits. The thermodynamic parameters of marine diesel engines, with significant failure information, have the characteristics of little external interference, good information quality, a wide diagnosis range, and strong availability, and have been the main method for the condition monitoring of modern ships. However, the main marine diesel engine of a ship is a complex system that integrates mechanical, electrical, thermal, and hydraulic features. It is composed of a fuel injection system, a chamber combustion system, a turbocharging and gas-exchange system, and starting and reversing systems [2]. There are many potential failure excitations and abundant characteristic parameters. No simple corresponding relationship exists between them, which is also influenced by the ship navigation conditions and engine operation boundary conditions [3]. The acquisition of a definite and invariable relationship between performance failures and thermodynamic parameters under different running conditions is the main bottleneck in the application of failure diagnosis to large marine diesel engines [4].

The failure simulation of a marine diesel engine by numerical calculation of the thermodynamic working process is a good way to reveal the intrinsic relationship between the thermodynamic parameters of the main engine and its performance failures under different ship navigation and engine running conditions, to solve the problem of parameter optimization, and provide training samples for automatic failure diagnosis by establishing a quantitative relationship between performance failures and characteristics under various engine operating conditions. In recent years, some scholars have carried out performance failure simulations and characteristic analyses of marine diesel engines. Jose Antonio Pagán Rubioa et al. introduced a four-stroke high-speed marine diesel engine simulator with a one-dimensional thermodynamic engine model developed in the AVLBoost. Numerous typical failures can be simulated, and a failure database for diagnosis can be built in this way [5]. Theotokatos et al. employed mean value engine models based on a zero-dimensional model. The simulation accuracy was validated against experimental data from ship trials, which can be used for engine performance prediction [6]. Matulić et al. simulated the failure of a two-stroke low-speed marine diesel engine based on an AVL multi-zone combustion model and other models. It was verified with onboard measurements under 40%, 60%, and 100% engine load with 2%–7% inaccuracy [7]. However, the thermodynamic parameters of diesel engines are not only related to performance failures, but also to their operating conditions, such as the engine room air temperature and cooling water inlet temperature [8]. To consider these factors, considerably more modeling should be carried out, which is nearly impossible for marine diesel engines. The acquisition of a definite and invariable relationship between thermodynamic parameters and performance failures under different running conditions is the main bottleneck in the application of failure diagnosis for large marine diesel engines, which has not been successfully solved by previous researchers.

This study considers a MAN B&W 6S35ME-B type two-stroke marine diesel engine as an example. The turbocharging and gas-exchange simulation system of the engine is built-up in MATLAB/Simulink with an average thermodynamic simulation model, which is verified by experimental results under normal conditions. The boundary condition and corresponding model parameters were changed to simulate typical performance failures. The thermodynamic parameters of a diesel engine under different working conditions and performance failures were analyzed in terms of the relative deviation, which demonstrated the relationship between the performance failures and thermodynamic parameters. This relationship could be mostly free of the engine running conditions. Finally, a normalization method is proposed to eliminate the influences of the engine room temperature and intercooler cooling water temperature. Future research will be carried out to verify this method, which lays a more practical foundation for the performance failure diagnosis of the diesel engine turbocharging ang gas-exchange system.


2. Numerical Calculation Under Normal Conditions

2.1 Simulation Models

The simulated engine is a two-stroke marine diesel engine with a MAN TCR 22 turbocharger equipped with an emergency blower of 18.5 kW. The rated speed was 142 r/min, and the rated power was 3570 kW with a 350/1550 mm diameter/stroke. The outlet diameter of the compressor end was 290 mm, and the scavenging tank volume was 167 L. The exhaust main pipe diameter is 796 mm and the length is 604.6 mm. The maximum compression ratio of the turbo charger was 5.0. The entire diesel engine turbocharging and gas-exchange system is divided into seven subsystems: air filter, compressor, intercooler, scavenging tank, chamber combustion, exhaust valve, exhaust manifold, and turbine, as shown in Figure 1 [9]. The thermodynamic model of the diesel engine system is shown in Figure 2. The mathematical simulation models are described as follows.

Figure 1:

Module composition of diesel engine turbocharging and gas-exchange system

1) Air Filter

It is simplified as a one-dimensional steady flow unit with main losses of local pressure drop ΔP1 and drag loss along the way ΔP2[10].

ΔP1=12ερv2(1) 
ΔP2=12λρ1dv2(2) 
ΔP0=ΔP1+ΔP2(3) 

Where ΔP1, ΔP2, ρ, l, d, P0 are local pressure loss at the inlet, resistance loss along the way, air density, pipe length, pipe diameter and compressor inlet pressure respectively. The loss coefficient ε is 0.5; The air viscous coefficient is taken as λ = 0.00001983 pa ∙ s. Air flow rate is v= mc/(ρAr), obtained from the compressor mass flow rate. Inlet equivalent area of compressor Ar is 0.039. The air filter simulation model in Matlab/Simulink is shown in Figure 3.

Figure 3:

Simulation model of air filter in Matlab/Simulink

2) Compressor

According to a one-dimensional adiabatic compression model with a compressor characteristic curve, the outlet temperature of the compressor [11]

Tcout=Tcin1+πcK-1K-11ηcs(4) 

The obtained torque of the compressor is

MC=30mc˙kRTcinπck-1k-1/πηcsntck-1(5) 

where Tcin and Tcout are the inlet and outlet temperatures of the compressor, respectively. ntc and πc are the supercharger speed and compressor pressure ratio respectively. Here, πc = Pclin/P0, and ηcs and mc˙ are the adiabatic efficiency and air flow rate of the compressor, respectively. The polytropic compression index and gas constant k and R can be obtained with imported ntc and πc. The simulation model of the compressor in the MATLAB/Simulink is shown in Figure 4.

Figure 4:

Simulation model of air compressor in Matlab/Simulink

3) Intercooler

Owing to its high heat exchange efficiency, the thermal inertia of the intercooler can be generally ignored, and only the cooling effect and pressure loss are considered, which are treated as a throttle cooling and pressure dropping unit [12]. With the cooling coefficient ε, the outlet temperature of the intercooler is

Tclout=Tclin+εTwi-Tclin(6) 

The pressure drop of the intercooler is related to the air flow rate, and the outlet air pressure of the intercooler can be expressed as [12]

Pclout=Pclin-ΔP0mc˙/mc0˙(7) 

where, Tclin, Tclout, and Pclin, Pclout are the air temperature and pressure at the intercooler inlet and outlet, respectively. Tw is the cooling water inlet temperature of the intercooler, ε is its cooling coefficient, and mc0˙ is the rated air flow rate of the intercooler. The initial air pressure loss of the intercooler P0 is 400 Pa [13]. The simulation model of the intercooler in MATLAB/Simulink is shown in Figure 5.

Figure 5:

Simulation model of inter-cooler in Matlab/Simulink

4) Scavenging Tank

The scavenging tank is modeled using the volumetric method calculated by the energy, mass, and ideal gas law conservation equations. The air pressure in the scavenging tank is [14]

Plm˙=kRmc˙Tclout-md˙Tim/Vim(8) 

The outlet air temperature in the scavenging tank is obtained from the ideal gas law equation as

Tim=PimVim/mgRg(9) 

where, Plm˙ and Tim are the air pressure rate and temperature in the scavenging tank, respectively. Here, the scavenging tank volume Vim of the MAN B&W 6S35ME-B9 diesel engine was calculated to be 0.167 m3. md is the intake airflow of the engine. The scavenging tank simulation model in MATLAB/Simulink is shown in Figure 6.

Figure 6:

Simulation model of scavenging tank in Matlab/Simulink

5) Chamber Combustion

The intake air flow of a diesel engine is obtained from the theoretical airflow of the engine with a filled cylinder

md˙=ηvPimVdneϕ/60RTim(10) 

where ηv is the charging efficiency of the engine, which is generally regarded as a single function of the running speed ne of the diesel engine. Here, ηv = a2 * ne2 + a1 * ne + a0,here a2 = −5.363 * 10−7, a1 = 2.04 * 10−3, a0 = 0.576107.

The exhaust gas temperature in the exhaust branch is expressed as

Te=Tim+K/1+L0α(11) 

where, K and α are the increase factors of the exhaust gas temperature and excess air coefficient, respectively. The required air per fuel L0 is 14.3. The exhaust gas temperature rise factor K is a function of the air-fuel ratio [13], which is interpolated with the excess air coefficient. In MATLAB/Simulink, the simulation model of the exhaust branch temperature of the diesel engine is shown in Figure 7.

Figure 7:

Simulation model of exhaust branch in Matlab/Simulink

Output torque of diesel engine

Ti=30Piπne=30ηiHumf˙πne(12) 

Its fuel oil consumption

gi=3600mf˙Pe=3600ηiHumf˙(13) 

Where mf, Ti, Te, gi are the fuel injection quantity per cylinder, output torque, exhaust gas temperature and fuel oil consumption rate of the diesel engine respectively. The lower fuel heat value Hu is 42700 kJ/(kgk).The evacuation volume per cylinder Vd is 0.87 m3. ηi is the indicated working efficiency of engine. The combustion dynamics and thermodynamic simulation model of combustion in the cylinder in Matlab/Simulink is shown in Figure 8.

Figure 8:

Thermodynamic simulation model of combustion in Matlab/Simulink

6) Exhaust Manifold

The exhaust gas in the manifold is also modeled with the "Volumetric method" and exhaust gas pressure

Pem˙=keRemout˙Te-mt˙Tem/Vem(14) 

The gas temperature in the exhaust manifold is obtained from the ideal gas equation law

Tem=PemVem/memRe(15) 

where Pem˙ and Tem are the gas pressure rate and temperature in the exhaust manifold, respectively. Here, the mixed gases in the exhaust manifold mout are md + mf, and the exhaust pipe volume Vem is 0.3 m3. The exhaust manifold simulation model in MATLAB/Simulink is shown in Figure 9.

Figure 9:

Simulation model of exhaust manifold in Matlab/Simulink

7) Gas Turbine

The turbine efficiency is fitted by the embedded RBF function in MATLAB by adopting the s-function from the turbine characteristic curve [14].

ntc'=ntcTem(16) 

The turbine outlet temperature is

Ttout=Tem1-ηts1-πt1-ktkt(17) 

The turbine output torque is

Mt=30mt˙ηtsktRTem1-πt1-ktkt/πntckt-1(18) 

The turbine is equivalent to a throttle nozzle for a simplified calculation. The gas flow, mt˙ is obtained as a one-dimensional isentropic expansion process:

mt˙=μTFresψPem/RTTem(19) 

When πt<kt+12ktkt-1 :

Ψ=2ktkt-1πt2kt-πt1+ktkt(20) 

When πtkt+12ktkt-1 :

ψ=2kt+1ktkt+12ktkt+1(21) 

where, Ttout, Mt, mt, ηts, and Fres (A0 in the following figure) are the gas outlet temperature, turbine output torque, gas flow, turbine efficiency, and gas flow area, respectively. The turbine expansion ratio πt was Pem/Pb. ψ, μT are the flow function and flow coefficient, respectively, where Fres is 0.039 and μT is 0.7. The turbine simulation model in MATLAB/Simulink is shown in Figure 10.

Figure 10:

Simulation model of turbine in Matlab/Simulink

The turbocharger rotating speed can be obtained from the D'Alembert's principle as

ntc˙=30ηmMT-Mc/πItc(22) 

where, ηm and Itc are the turbocharger mechanical working efficiency and turbocharger rotating inertia, respectively, where ηm is 0.89 and Itc is 16. The turbine rotor simulation model in MATLAB/Simulink is shown in Figure 11.

Figure 11:

Simulation model of rotor in Matlab/Simulink

2.2 Simulation Results

Thermodynamic parameters such as the running speed, power, and intake air temperature of the diesel engine were recorded in an engine laboratory at 25%, 50%, 75%, and 90% engine operation conditions. A comparison between the experimental and simulated results is presented in Table 1.

Comparison between experimental and simulated results

It is verified that the errors of most measured and calculated parameters are within 5%, except for a few parameters. These results indicate that the simulation models of the diesel engine turbocharging and gas-exchange system in MATLAB/Simulink are feasible, and the thermodynamic parameters from the diesel engine simulation calculation, such as the temperature, pressure, speed, and fuel consumption rate, can reflect the condition of the engine under different operating conditions.


3. Simulation and Analysis of Performance Failures

3.1 Performance Failure Simulation

Based on the diesel engine simulation model mentioned above, some performance failures of the super-charging and gas-exchange system were simulated by changing the model parameters to reveal the variation rules of each thermal parameter under different failure conditions. Additionally, the variation in the boundary conditions of a diesel engine also affect its operating performance and thermal parameters, such as the compressor intake air temperature and intercooler cooling water inlet temperature. For ease of representation, these are collectively referred to as performance failures; the performance failures for some failure states and the setting of the model parameters are listed in Table 2.

Performance failures and modeling parameters

In addition to the conventional thermodynamic parameters, some combined parameters are proposed here to indicate the performance failure of diesel engine, including the compressor coefficient Nc1, turbine coefficient Nt1, compressor coefficient Nc2, turbine coefficient Nt2, intercooler cooling coefficient Nc, intercooler flow coefficient Ncf, and turbine flow coefficient Ntf.

NC1=Tcout-TcinPclinntc(23) 
Nt1=ntcTem-TtoutPem(24) 
Nc2=Pclinntc(25) 
Nt2=ntcTem(26) 
Nc=Tclin-TcloutTclin-Twin(27) 
Ncf=Pclin-Pclout(28) 
Ntf=Pem-Pb(29) 

These thermodynamic parameters are represented in terms of the relative deviation under different running conditions and performance failures [15].

ε=x-x0x0(30) 

where x and x0 are the thermodynamic parameters of the diesel engine model under failure and normal conditions, respectively.

3.2 Performance Failures Analysis

The simulated results of turbocharging and gas-exchanging systems under different running conditions and various

performance failures were analyzed using 18 thermodynamic parameters, as shown in the following figures. The 18 calculated thermodynamic parameters are represented by different numbers, as listed in Table 3.

Thermodynamic parameters represented with numbers

1) Extremely High Inlet Air Temperature

Owing to the different shipping routes, the influence of the engine room temperature cannot be ignored, and the engine inlet air temperature could be different. The relative deviations of the thermodynamic parameters under different running speeds and engine room temperatures are shown in Figure 12. It can be observed that when the engine room temperature rises, the decrease in inlet air density worsens the chamber combustion inside the cylinder. The exhaust air temperature Te, exhaust gas temperature in manifold Tem, and turbine outlet gas temperature Ttout of the diesel engine increase significantly, whereas the intercooler inlet air pressure Pclin, air pressure in the scavenging tank Pim, and exhaust air pressure in the manifold Pem decrease. The most obvious change was the rise in the compressor outlet air temperature Tcout. With respect to the combined parameters, the turbine coefficient Nt1 increases the most obviously, whereas the compressor coefficient Nc1, compressor coefficient Nc2, turbine coefficient Nt2, intercooler flow coefficient Ncf, and turbine flow coefficient Ntf all decrease to varying degrees.

Figure 12:

Relative deviation under different engine room temperatures and running speeds

Comparing Figure 12 (a) and Figure 12 (b), it can be observed that the variations in the various parameters under different running conditions are quite similar. This indicates that the performance failure behavior of characteristic parameters in terms of the relative deviation is quite clear and immovable, even under different running conditions of the diesel engine, which is greatly convenient for failure diagnosis.

2) Intake Air Room Filter Blockage

The pressure drop due to resistance along the way of the filter mesh was simulated by changing the length diameter ratio of the filter (l/d) in Equation (2) to simulate the blockage of the air filter in the diesel engine.

Figure 13 shows the relative deviation of the characteristic parameters of the diesel engine when l/d is 1500 and 300 at different running speeds. It can be observed that when the air filter is blocked, the inlet air pressure of the intercooler Pclin, outlet air pressure of the intercooler Pclout, and exhaust gas pressure in manifold Pem decrease, whereas the exhaust gas temperature Te, exhaust gas temperature in the manifold Tem, and turbine outlet gas temperature Ttout of the diesel engine increase, and the outlet air temperature of the compressor Tcout, intercooler outlet air temperature Tclout, and scavenger air temperature Tim do not change significantly. Additionally, the compressor coefficient Nc2, turbine coefficient Nt2, intercooler flow coefficient Ncf, and turbine flow coefficient Ntf decrease significantly. It was found that when the air filter is blocked, the most affected characteristic parameter is the inlet pressure of the compressor, resulting in insufficient air into the compressor and poor combustion inside the cylinder; thus, the exhaust gas temperature increases in the exhaust manifold.

Figure 13:

Relative deviation under different air filter l/d and running speeds

3) Intercooler Fouling at Water Side

The fouling in the intercooler water passage also has a direct impact on the intake air of diesel engines.

As shown in Figure 14, owing to the fouling on the water side of the intercooler, the cooling coefficient decreases, and it is difficult for air to exchange heat with the cooling water. The intercooler outlet air temperature Tclout, scavenging air temperature Tim, exhaust gas temperature Te, exhaust gas temperature in manifold Tem, turbine outlet gas temperature Ttout, intercooler inlet air pressure Pclin, and air pressure in scavenging tank Pim all increase significantly. The compressor coefficient Nc1 and compressor coefficient Nc2 increase significantly, whereas the turbine coefficients Nt1, Nt2, and Nc decrease. The turbocharger speed increases slightly, which is similar to that of the extremely high intercooler cooling water temperature described below. Here, the intercooler cooling coefficient Nc decreases significantly, whereas it remains unchanged under other failures. Therefore, the intercooler cooling coefficient Nt could be used as a unique characteristic parameter for the intercooler water fouling failure at both the air and water sides.

Figure 14:

Relative deviation under different intercooler cooling coefficients and running speeds

4) Extremely High Intercooler Cooling Water Temperature

The intercooler has a direct impact on the intake air temperature of the diesel engine, especially on the intake air temperature of the cylinder. When the intercooler inlet cooling water temperature increases, the intake air flow slightly decreases, leading to an increase in the compressor outlet air temperature Tcout, intercooler outlet air temperature Tcout, scavenging air temperature Tim, exhaust gas temperature Te, exhaust gas temperature in manifold Tem, turbine outlet gas temperature Ttout, intercooler inlet air pressure Pclin, and scavenging air pressure Pim. The compressor coefficient Nc1 and compressor coefficient Nc2 increase significantly, whereas the turbine coefficient Nt1, turbine coefficient Nt2, and intercooler flow coefficient Ncf decrease, and the turbocharger rotating speed increases slightly, as shown in Figure 15.

Figure 15:

Relative deviation under different intercooler cooling water temperatures and running speeds

5) Scavenging Port Fouling

With the blockage of the scavenging port, the intake air flow into the cylinder decreases significantly, which worsens the fuel combustion inside the cylinder. The exhaust gas temperature Te, exhaust gas temperature in the manifold Tem, and turbine outlet gas temperature Ttout increase significantly, whereas the intercooler inlet air pressure Pclin, scavenging air pressure Pim, and exhaust gas pressure in the manifold Pem decrease. The most obvious increase is the compressor outlet air temperature Tcout, as shown in Figure 16. The turbine coefficient Nt1 increases most obviously, whereas the compressor coefficients Nc1, Nc2, Nt2, Ncf, and Ntf all decrease to varying degrees. The characteristic behavior is similar to that of air filter fouling, turbocharger bearing wear, and high inlet air temperature.

Figure 16:

Relative deviation under different scavenging port flowing coefficients and running speeds

6) Turbine Nozzle Blockage

Blocked turbine nozzles can block gas flow into the turbine, which significantly increases the exhaust gas pressure in the manifold Pem, whereas the exhaust temperature Te of the diesel engine, exhaust manifold temperature Tem, and turbine outlet temperature Ttout slightly decrease. When the turbine nozzle is blocked, its most direct influence is on the exhaust gas pressure in the manifold Pem and turbine flow coefficient Ntf, which could be used as two characteristic parameters for turbine nozzle blockage failure, as shown in Figure 17.

Figure 17:

Relative deviation under different nozzle flowing coefficients and running speeds

7) Worn Turbocharger Bearing

When the turbine rotor is mechanically worn out, its compressor air compression capacity and compressor air flow decrease, which cause an increase in the exhaust temperature Te, exhaust gas temperature in manifold Tem, and turbine outlet gas temperature Ttout, whereas the inlet air pressure Pclin, scavenging air pressure Pim, exhaust gas pressure in the manifold Pem, and turbocharger rotating speed ntc decrease significantly, similar to the case in the turbine outlet passage fouling failure, as shown in Figure 18 and 19.

Figure 18:

Relative deviation under different turbocharger mechanical efficiencies and running speeds

Figure 19:

Relative deviation under different turbine exhaust back pressures and running speeds

8) Turbine Outlet Passage Fouling

The turbine outlet passage fouling was simulated by increasing the back pressure during simulation. As shown in Figure 19, a higher back pressure of the turbine causes an increase in the exhaust gas temperature Te, exhaust gas temperature in manifold Tem, and turbine outlet gas temperature Ttout. The most direct impact is the decrease in the turbine working capacity, which reduces the turbocharger speed ntc, intercooler inlet air pressure Pclin, intercooler outlet air pressure Pclout, scavenging air pressure Pim, and exhaust gas pressure in manifold Pem. With respect to the combined parameters, the increase in the turbine coefficient Nt1 is the most obvious, whereas the compressor coefficient Nc1, compressor coefficient Nc2, turbine coefficient Nt2, intercooler flow coefficient Ncf, and turbine flow coefficient Ntf all decrease to some extent.

In conclusion, the characteristic parameters of extremely high inlet air temperature, intake air filter blockage, scavenging port fouling, worn turbocharger bearing, and turbine outlet passage fouling exhibit similar behavior even under different operating conditions of the diesel engine. However, when the engine room temperature is extremely high, the compressor outlet temperature Tcout increases significantly but decreases to a certain extent in the other failures, which provides a distinguishing parameter from other failures. The engine room temperature can be measured directly, the blocked air filter can be monitored with the differential pressure of the intake air filter, and the turbine back pressure can also be measured directly at the turbine outlet. Additionally, the worn turbocharger bearing can be judged from its vibration and scavenging port fouling can be detected by the pressure difference between the intercooler outlet air and scavenging air; thus, these different failures can be identified. The characteristic parameter behavior of the extremely high intercooler cooling water temperature is considerably similar to that of intercooler fouling at the water side, but the fouled intercooler can be determined directly by the intercooler cooling coefficient Nc, and the cooling water temperature in the intercooler can be measured directly. These two failures were the most easily distinguished. The most obvious characteristic of the turbine nozzle blockage is the obvious increase in the exhaust gas pressure in manifold Pem and turbine flow coefficient Ntf, which is not found in the other failures. In other words, the performance failure diagnosis of diesel engine turbocharging and gas-exchanging systems can be eventually realized with some other parameters and information.


4. Normalization of Parameters

In fact, extremely high inlet air temperatures and intercooler cooling water temperatures are not engine performance failures, but are different working boundary conditions, which change the thermodynamic parameters of the engine. Therefore, before performance failure diagnosis, the normalization of the engine boundary conditions should first be carried out to eliminate the influence of the different conditions. Figures 20 and 21 show the changing trends of the 18 characteristic parameters when the engine room temperature increases from 286 K to 400 K and the intercooler cooling water temperature increases from 315 K to 420 K at engine speeds of 129 r/min and 75% load, respectively.

Figure 20:

Relative deviations under different engine room temperatures

Figure 21:

Relative deviations under different intercooler cooling water temperatures

The same method can also be used to eliminate the influence of the intercooler cooling water temperature.

It can be observed that the relative deviations of these parameters change linearly with the change in the engine room temperature and intercooler cooling water temperature. If the 286 K engine room temperature and 300 K intercooler cooling water temperature are set as the reference boundary conditions, the thermodynamic parameters under different conditions are normalized to the reference condition according to the actual engine room temperature and intercooler cooling water temperature. The relative deviation analysis and failure diagnosis of the characteristic parameters can be carried out. The specific method uses the Newton interpolation formula to calculate the relative deviation value ε at the actual temperature of a certain engine compartment at z°K. According to Equation (32), the thermal parameter x is calculated after normalizing to the reference running condition x′.

Assuming the actual engine room temperature is z°K, the relative deviation ε of a thermodynamic parameter x could be normalized between 286 K and 400 K using the Newton interpolation method.

ε'=z-286ε114(31) 

Normalized thermodynamic parameter

x'=x1+ε'(32) 

Subsequently, the normalized parameter x′ can be compared to the parameter under the engine normal condition x0 to obtain the relative deviation necessary for failure diagnosis.

Although the characteristic parameters have similar behaviors under different running conditions for these failures, the deviation amplitudes are quite different, as shown in Figures 12-19. Therefore, before the performance failure diagnosis, it is necessary to normalize the relative deviation to eliminate the influence of different running conditions before failure diagnosis. All the 18 relative deviations under different running conditions will be unified to a 1.0 maximum scale. Suppose the maximum relative deviation among the 18 relative deviation values is ε1, which is normalized to 1.0- or -1.0, and the other relative deviation ε0 is normalized as ε:

ε=ε01.0ε1(33) 

Taking the intake air filter blockage as an example, as shown in Figure13, the relative deviations under the two running conditions are quite different and cannot be used directly for failure diagnosis. After normalization, the two sets of parameters have nearly the same scale and are comparable. In this way, the performance failures could be identified under any engine running condition with different failure severities. The severity of the failure can be directly determined using the relative deviation scale before normalization.

Figure 22:

Normalized relative deviations under different air filters l/d and running speeds


5. Conclusion

The proposed average thermodynamic simulation model for diesel engine turbocharging and gas-changing systems is available in MATLAB/Simulink. By changing the boundary conditions and relevant model parameters, performance failures including extremely high engine room temperature, intake filter blockage, intercooler fouling at the water side, extremely high intercooler cooling water temperature, scavenging port fouling, clogged turbine nozzle, worn turbocharger bearing, and turbine exhaust passage fouling can be simulated. The relative deviations of the thermodynamic parameters under different boundary conditions and different failures are analyzed, demonstrating the inherent relationship between the thermodynamic parameters and performance failures and could be mostly free of engine running conditions. By normalizing the parameters, the influence of different boundary conditions and running conditions could be eliminated, and the performance failure diagnosis of the diesel engine under different running conditions could be realized in practice. Further research will be carried out to verify the feasibility of this method in the application of the failure detection of diesel engines on board ships.

Nomenclature

ΔP1 : local pressure loss at air filter inlet
ΔP2 : resistance loss along air filter way
ρ : inlet air density
l : pipe length of the air filter
d : pipe diameter of the air filter
P0 : compressor inlet pressure
λ : inlet air viscous coefficient
v : inlet air flow rate
Ar : equivalent inlet area of compressor
Tcin : inlet air temperature of the compressor
Tcout : outlet air temperature of the compressor
ntc : turbocharger rotating speed
πc : compressor pressure ratio
k : polytropic compression index
R : gas constant
Tclin : air temperature at intercooler inlet
Tclout : air temperature at intercooler outlet
Pclin : air pressure at intercooler inlet
Pclout : air pressure at intercooler outlet
Tw : cooling water inlet temperature of intercooler
ε : cooling coefficient of cooling water
mc0 : rated air flow rate of intercooler
ΔP0 : initial air pressure loss of intercooler
Pim : air pressure in scavenging tank
Tim : air temperature in scavenging tank
Vim : scavenging tank volume
md : intake air flow of engine
ηv : charging efficiency of engine
ne : running speed of engine
K : increase factor of exhaust gas temperature
α : excess air coefficient
mf : injected fuel quantity per cylinder
Ti : output torque of engine
Te : exhaust gas temperature of engine
gi : fuel oil consumption rate of engine
Hu : fuel oil low heat value
Vd : evacuation volume per cylinder
ηi : indicated working efficient of engine
Pem : gas pressure in exhaust manifold
Tem : temperature in exhaust manifold
mout : mixed gases flow in exhaust manifold
Vem : exhaust manifold volume
ntc' : similar rotating speed of turbine
Ttout : turbine gas outlet temperature
Mt : turbine output torque
mt : gas flow of turbine
ηts : turbine working efficient
Fres : gas flow area of turbine
πt : turbine expansion ratio
μt : flow coefficient of turbine
ψ : flow function coefficient of turbine
ηm : turbocharger working efficiency
Itc : turbocharger rotating inertia
l/d : aspect ratio of compressor filter
μs : intake air flow coefficient
μT : turbine nozzle flow coefficient
Pb : turbine back pressure
Nc1 : compressor coefficient 1
Nc2 : compressor coefficient 2
Nc : intercooler cooling coefficient
Nt1 : turbine coefficient 1
Nt2 : turbine coefficient 2
Ncf : intercooler flow coefficient
Ntf : turbine flow coefficient

Acknowledgments

This research work was supported by the Science & Technology Commission of Shanghai Municipality and Shanghai Engineering Research Center of Ship Intelligent Maintenance and Energy Efficiency under Grant 20DZ2252300.

Author Contributions

Conceptualization, Y. Hu; Methodology, C. Zeng and M. Wang; Software, J. Jiang; Formal Analysis, Y. Hu and C. Zeng; Investigation, M. Wang; Resources, C. Zeng; Data Curation, M. Wang; Writing-Original Draft Preparation, Y. Hu and C. Zeng; Writing-Review & Editing, Y. Hu; Visualization, M. Wang; Supervision, Y. Hu; Project Administration, Y. Hu; Funding Acquisition, Y. Hu.

References

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Figure 1:

Figure 1:
Module composition of diesel engine turbocharging and gas-exchange system

Figure 3:

Figure 3:
Simulation model of air filter in Matlab/Simulink

Figure 4:

Figure 4:
Simulation model of air compressor in Matlab/Simulink

Figure 5:

Figure 5:
Simulation model of inter-cooler in Matlab/Simulink

Figure 6:

Figure 6:
Simulation model of scavenging tank in Matlab/Simulink

Figure 7:

Figure 7:
Simulation model of exhaust branch in Matlab/Simulink

Figure 8:

Figure 8:
Thermodynamic simulation model of combustion in Matlab/Simulink

Figure 9:

Figure 9:
Simulation model of exhaust manifold in Matlab/Simulink

Figure 10:

Figure 10:
Simulation model of turbine in Matlab/Simulink

Figure 11:

Figure 11:
Simulation model of rotor in Matlab/Simulink

Figure 12:

Figure 12:
Relative deviation under different engine room temperatures and running speeds

Figure 13:

Figure 13:
Relative deviation under different air filter l/d and running speeds

Figure 14:

Figure 14:
Relative deviation under different intercooler cooling coefficients and running speeds

Figure 15:

Figure 15:
Relative deviation under different intercooler cooling water temperatures and running speeds

Figure 16:

Figure 16:
Relative deviation under different scavenging port flowing coefficients and running speeds

Figure 17:

Figure 17:
Relative deviation under different nozzle flowing coefficients and running speeds

Figure 18:

Figure 18:
Relative deviation under different turbocharger mechanical efficiencies and running speeds

Figure 19:

Figure 19:
Relative deviation under different turbine exhaust back pressures and running speeds

Figure 20:

Figure 20:
Relative deviations under different engine room temperatures

Figure 21:

Figure 21:
Relative deviations under different intercooler cooling water temperatures

Figure 22:

Figure 22:
Normalized relative deviations under different air filters l/d and running speeds

Table 1:

Comparison between experimental and simulated results

Thermodynamic Parameters Data Type Engine Load
25% 50% 75% 90%
Engine Speed
ne(r/min)
Measured 89 114 129 140
Injected Fuel
Quantity mf(g)
Measured 6.11 8.49 10.28
5
12.58
Turbocharger Speed ntc(r/min) Measured 9679 14961 18300 19135
Calculation 9985 14890 18030 19370
Turbine Outlet Gas Temperature
Ttout(K)
Measured 540.7 541.1 523.6 542.0
Calculation 563.2 532.0 490.8 536.4
TurbineIinlet Gas Temperature
Tem(K)
Measured 590.7 636.4 649.0 685.3
Calculation 602.7 612.1 619.3 700.6
Exhaust Gas Temperature Te(K) Measured 592.0
3
635.4
7
601.1
3
675.9
Calculation 581.2 624.2 604.0 671.4
Compressor Inlet air Temperature
Tim(K)
Measured 313.4
6
315.5
3
317.7
4
320.8
2
Calculation 315.9 319.1 323.2 325.8
Compressor air Outlet Temperature Tcout(K) Measured 339.3
25
399.9
1
447.5
6
469.5
6
Calculation 342.8 396.2 443.8 474.5
Scavenging Air Pressure Pim(kpa) Measured 144 211 270 300
Calculation 153.2 229.4 305.2 340.3
Fuel Consumption Rate gi(kg/kw∙h) Measured 218.6 221.3 199.1 192.4
Calculation 209.9 207.4 196.1 194.8

Table 2:

Performance failures and modeling parameters

Performance Failure Modelling Parameter Normal Status Failure Status 1 Failure Status 2 Failure Status 3 Failure Status 4 Failure Status 5
Too High Engine Room Temperature Ambient Temperature
Tcin(K)
286 300 315 329 343 357
Intake Filter Blockage Aspect Ratio l/d 3 300 600 900 1200 1500
Intercooler Fouling at Water Side Cooling Coefficient of Intercooler ε 0.89 0.85 0.80 0.75 0.70 0.65
Too High Intercooler Cooling Water Temperature Cooling Water Temperature
TW(K)
300 315 330 345 360 375
Scavenging Port Fouling Intake Air Flow Coefficient uS 0.40 0.38 0.36 0.34 0.32 0.30
Clogged Turbine Nozzle Turbine nozzle Flow Coefficient uT 0.70 0.67 0.63 0.61 0.58 0.55
Worn Turbocharger Bearing Working Efficiency of Turbocharger ηm 0.89 0.85 0.80 0.75 0.70 0.65
Turbine Exhaust Passage Fouling Turbine Back Pressure Pb(Pa) 111400 116970 122540 128110 133680 139250

Table 3:

Thermodynamic parameters represented with numbers

Number Parameter Parameter Description
1 Tcout Outlet Air Temperature of Compressor
2 Tclout Intercooler Outlet Air Temperature
3 Tim Scavenger Air Temperature
4 Te Exhaust Gas Temperature
5 Tem Exhaust Gas Temperature in Manifold
6 Ttout Turbine Outlet Gas Temperature
7 Pclin Intercooler Inlet Air Pressure
8 Pim Intercooler Outlet Air Pressure
9 Pem Exhaust Gas Pressure in Manifold
10 ntc Turbocharger Rotating Speed
11 ge Engine Fuel Oil Consumption Rate
12 Nc1 Compressor Coefficient
13 Nt1 Turbine Coefficient
14 Nc2 Compressor Coefficient
15 Nt2 Turbine Coefficient
16 Nc Intercooler Cooling Coefficient
17 Ncf Intercooler Flow Coefficient
18 Ntf Turbine Flow Coefficient